Design Of An Axial Compressor Stage Engineering Essay

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Abstract

This research will focused on the axial compressor performance for contra rotating gas turbine engine. The aerodynamics term results will be compared between analytical method and computational method to verify compressor performances. On the other hand, one stage of axial compressor is modelled using NACA 65A010 of airfoil shape that depends on both aerodynamics and geometrical properties as the input values. Hence the generated blades for both rotor and stator are meshed using Turbo Grid application. Errors that showing the bad meshing in percentage may be ignored during the meshing process only after the mesh refinement. Simulation of both rotor-stators will be conducted using CFX Pre, CFX Solver and CFX Post to get the required profiles among the blades. For contra rotating stage, fan and first stage compressor are combined together in the simulation. The comparison of pressure, enthalpy and temperature ratio are main points to verify that counter rotating machine is more efficient compare to the same direction rotation of engine.

Introduction

An axial compressor is one of the major parts of gas turbine engine that is used widely in both industrial and aircraft applications. For industrial applications, it can be used as a generator drive when connected to an electrical motor or mechanical drive when it is connected to the compressor, pump and propeller. It is normally used at power generation plant, petrochemical plant, natural gas plant and even on the oil rigs for off shoring purposes. However for the aircraft application, a gas turbine is used to produce thrust force or propulsion hence it will move the aircraft during cruising. Now an axial compressor of a gas turbine is our main focus to improve its performance as one of the many improvement steps involved in the industries. It including by choosing or designing a blade profile for the rotor and stator of compressor stage through preliminary design analysis. Computational fluid dynamics (CFD) analysis will then be carried out on a contra rotating and unidirectional rotating fan-compressor system. Focus will be given on the performance of those two configurations (e.g. in terms of aerodynamic losses due to turbulence and separation and also pressure rise).

Previous study on contra rotating engine

According to D.S Pundhir and P.B Sharma in A Study of Aerodynamic Performance of Contra Rotating Axial Compressor Stage, the aerodynamics of contra stage is significantly affected by the speed ratio as well as the axial spacing between contra rotors. In their experiment the effect of speed rotational and axial spacing between contra rotors on the thermodynamics performance of contra rotor are investigated. The traverses flow of structure and pressure changes are determined at both upstream and downstream for first and second rotor. Results are analyzed to obtain relative total head loss and blade element efficiency of contra rotors.

According to Peter Schimming in Counter Rotating Fans-An Aircraft Propulsion for the Future?, 2003 , it is stated that the fuel consumption of jet engines has decreased about 40% since counter rotating engine has been introduced in civil aviation market. This is due to the increment of engine efficiencies and the change of turbojet to turbofan engine with increasing by part ratios of the engine itself. Additionally both numerical and experimental methods are conducted including designed and manufactured by MTU, Germany. Analysis of the upstream and downstream aerodynamic interaction effects of neighboring blade rows is one of area of interest during the study. The results of unsteady flow through both blade rows of 2D and 3D calculations and an aero elastic behavior of blades with respect to forced response are presented. Active control technique (ANC) is found to be a method to diminish fan noise in the future. Overall the results of investigation shows that counter rotating fans have a potential to be an alternative propulsor in the future.

Additionally, Y-Y Chen, B Liu, Y Xuan and X-Rxiang in A Study of Speed Ratio Affecting the Performance of a Contra-rotating Axial Compressor , 2008 , both numerical and experimental method are investigated to analyze the performance and the flow structure of a counter rotating compressor with different rotating speed and typical working condition.

Both results are agreed with each other in term of total pressure ratio when peak efficiency was a little bigger than the experimental one. At maximum speed ratio, the peak isentropic efficiency can be increased with minor reduction of total pressure ratio and safe margin. It shows that compressor will operate at stall point when flow reversal fist occurred at the start section of the outlet guide vane and it covered nearly 40% area of the whole flow at stall point with rotational speed ratio of rotor 1 and 2 were greater than or equal to 1. Else, the stator vanes could be cancelled from the compressor to decrease the engine size and increase the thrust weight ratio when using contra rotating technique.

Contra rotating turbofan engine

In contra rotating turbofan engine, it is normally consists of two (one internal shaft and other will be act as an external shaft) or more shaft that are connected with separate parts of compressor-turbine only. For example if the high pressure turbine is connected with high pressure compressor and rotates in a clockwise direction while the low pressure turbine is connected with the low pressure compressor but it will rotate in counter direction reflected to high pressure system. Additionally the counter rotation of fan and compressor in a turbofan engines are considered as contra rotating gas turbine engine as what we have in our current study.

Methodology

The preliminary design is conducted by determining the number of stages at which work done at a single stage is calculated and all stages are assumed to have an equal amount of work. The stage loading coefficient is taken from the previous experience of design study, 0.65 (Cohen, Rogers and Saravananmuttoo, (1978)) and notes that the maximum value of stage loading gives lowest number of stages. Other the blade speed, U is choosing due to the turbine design, referring to the common rotor speed. In this study we choose the value of ψ is 0.65, U= 247.877723 m/s, thus Δh0 = 39938.18763 J per stage. Note that power output of the compressor from the previous data is equal to 324589. 4957 J/kg. Thus the number of stages is 8.12729 ≈ 8 (it is manually change) stages of an axial compressor. The next step is to determine the velocity triangles at hub, mean and tip of the blade and the de Haller criterion must be satisfied to avoid stall and surge phenomena occur within the compressor. The value of an absolute velocity is assumed to be constant for all compressor stages and the degree of reaction is equal to 0.5, indicates a constant reaction for the following stages along the compressor. Additionally the pitch to chord ratio is determined based on number of blades required and the effect of friction and compressor weight are under consideration among the designers. Here, the blade is assumed in an aerofoil wrapped around a circular arc and for the purpose of preliminary design, incidence and deviation are neglected hence it can define the arc since blade angles are assumed to be equal to the flow angles.

The value of pitch to chord ratio is determined from the design chart and experimental experience. Then, blade profiles are generated using NACA 65A010 airfoil shape. Note that the airfoil's chamber line is curved and the symmetric shape of it profile is distributed about the chamber line to obtain the desired change in fluid flow direction. Additionally the blade is created based on the preliminary design results at which some properties such as air angle distributions to achieve the work per stage. The blade shape is constructed based on 50% constant reaction design. The output files are saved in '<file name>.curve' and '<file name> .inf' format.

Meshing is done by using TurboGrid application in ANSYS 12. The output files that are saved in '<file name>.curve' and '<file name> .inf' format will act as input for the blade geometry. Under topology, the method of using H/J/L-grid is selected for the compressor blade. The node counts for coarse, medium and fine mesh are 75000 nodes, 100000 nodes and 250000 nodes respectively.3D mesh is generated after doing the refinement of the grid topology by dragging the master control point closer to the blade leading edge and trailing edge respectively while the mesh statistics will show us the percentage of mesh bad is currently produced. The mesh file is then saved in '<file name>.gtm' format.

Then by using '<file name>.gtm' format, the CFX-Pre is used to setup the simulation by setting all the required information such as machine type, axis of rotation, components and the setup for physic condition- fluid properties, inflow and outflow boundary conditions and the solver parameters. Both interfaces and boundary conditions are checked carefully to ensure that the setup is matched to the requirements. After that the CFX Solver will run the iteration process of the setup model until the results are converged. Hence the CFX Post will display the results of the simulation by showing us the blade to blade and meridional view of pressure, temperature, velocity and other properties variation along the blade itself. The results of medium and fine mesh are compared with each other especially in term of pressure ratio.The combination of contra rotating stage is between the fan and first stage compressor at which both are rotated in an opposite direction due to their own rotational speed. The pressure ratio of same and contra rotation is to be compared and its specific fuel consumption is taken into consideration in term of economical analysis.

Results

Design points

Overall pressure ratio

15

Thrust [ kN ]

30

Flying level [ km ]

6.1

Mach number

0.6

Temperature exit from turbine [ K ]

1300

Fluid properties

Gas constant, R [ J/kg.K ]

287

Specific heat ratio, γ

1.4

Air inlet / atmospheric pressure, Pa [ N/m2 ]

48712.4

Air inlet / atmospheric temperature, Ta [ K ] (0≤h≤11 km)

248.35

Air inlet / atmospheric temperature, Ta [ K ] (11<h≤20 km)

216.5

Air axial velocity, Ca [m/s]

189.5345108

Specific heat of air, cpa [ J/ kg.K ]

1005

Specific heat of gas, cpg [ J/kg.K ]

1147

Table 5. 1 Design points of axial compressor design

Blade profile from analytical calculation

First stage compressor rotor blade = 89

Properties

Blade section

Pitch-chord , s/c

Chord, c [ m ]

Radius, r [m]

Hub

0.50

0.03771157

0.25808523

Between hub and mean

0.53

0.03947229

0.28634312

Mean

0.60

0.03830807

0.314601

Between mean and tip

0.80

0.03131171

0.34285888

Tip

1.00

0.0271139

0.37111676

First stage compressor stator blade = 86

Properties

Blade section

Pitch-chord , s/c

Chord, c [ m ]

Radius ,r [m]

Hub

0.50

0.03644039

0.25808523

Between hub and mean

0.53

0.03814176

0.28634312

Mean

0.60

0.03701679

0.314601

Between mean and tip

0.80

0.03025626

0.34285888

Tip

1.00

0.02619995

0.37111676

Table 5.2 Blade profile for first stage compressor rotor

Table 5.3 Blade profile for first stage compressor stator

Blade generation

(a) (b)

Figure 5.1 Rotor blade (a) stator blade (b)

Meshing

(a) (b)

Figure 5.2 Rotor (a) and stator (b) fine mesh with more than 250000 nodes count.

Simulation results

(a) (b)

Figure 5.3 The convergence results (a) and simulation setup for compressor stage (b).

(a) (b)

Figure 5.4 Pressure (a) and temperature (b) variation at 0.5 span locations.

(a) (b)

Figure 5.5 Velocity vectors at leading (a) and trailing edge (b) of rotor at 0.5 span locations

(a) (b)

Figure 5.5 Velocity vectors at leading (a) and trailing edge (b) of rotor at 0.5 span locations

Graphs of total pressure (a), temperature (b) and velocity (c) versus streamwise location at 0.5 span locations can be represented as below:

(b) (c)

Simulation results for an axial compressor stage

Mass averages

Quantity

Fine Mesh (>250000 nodes)

Medium Mesh (>100000 nodes)

Ratio (Out/In)

Ratio (Out/In)

Temperature

1.0696

1.0690

Total Temperature

1.1620

1.1615

Pressure

1.2693

1.2655

Total Pressure

1.5870

1.5840

Enthalpy

-1.6427

-1.6173

Total Enthalpy

8.2130

8.1872

Table 5.3 The ratio comparison between fine and medium mesh

Results

Properties

Fine Mesh

(>250000 nodes)

Medium Mesh (>100000 nodes)

Torque (one blade row) (kg m2 s-2)

-26.6590

-26.5972

Torque (all blades) (kg m2 s-2)

-26.6590

-26.5972

Power (all blades) (kg m2 s-3)

-21004.9

-20956.2

Dimensionless mass flow (theta)

0.1423

0.1424

Dimensionless impeller tip speed (Mu)

0.1125

0.112541

Flow Coefficient (phi)

1.2640

1.2654

Head Coefficient (psis)

55.6839

55.4376

Blade Loading Coefficient (psi)

-31.9775

-31.8685

Total pressure loss coefficient #1

-0.5870

-0.5840

Total pressure loss coefficient #2

-0.3699

-0.3687

Total pressure loss coefficient #3

-3.2216

-3.2016

Total pressure loss coefficient #4

-1.0693

-1.0631

Static Enthalpy Loss (zeta)

-0.1693

-0.1056

Total Enthalpy Loss (delta Q)

4.1356

4.1497

Total-to-total isen. efficiency

0.8707

0.8698

Total-to-static isen. efficiency

0.0662

0.0606

Total-to-total poly. efficiency

0.8788

0.8780

Total-to-static poly. efficiency

0.0710

0.0651

Table 5.4 The comparison of properties results between fine and medium mesh

Simulation results for same and contra direction engine

Mass averages

Rotation

Same rotation

Contra rotation

Fan

Compressor

Fan

Compressor

Quantity

Ratio (Out/In)

Ratio (Out/In)

Ratio (Out/In)

Ratio (Out/In)

Temperature

1.0799

1.0638

1.0715

1.1275

Total Temperature

1.1141

1.1399

1.1150

1.1905

Pressure

1.3577

1.2125

1.2820

1.4669

Total Pressure

1.4682

1.4741

1.4570

1.6156

Enthalpy

0.5421

0.2364

0.5909

-0.4171

Total Enthalpy

0.05593

-21.7276

0.0481

-36.5662

Table 5.5 The mass averages comparison in ratio between same and contra rotation engine

ii. Results

Rotation

Same rotation

Contra rotation

Fan

Compressor

Fan

Compressor

Torque (one blade row) (kg m2s-2)

-105.3210 kgm^2 s^-2

-89.6365 kgm2s-2

104.4950 kgm^2 s^-2

-122.7460 kgm2s-2

Torque (all blades)( kg m2s-2)

-105.3210 kgm^2 s^-2

-358.5460 kgm2s-2

104.4950 m^2 s^-2

-490.9830

Power (all blades)( kg m2s-3)

-66175 kgm^2 s^-3

-282502 kgm2s-3

65655.90 kgm^2 s^-3

-386851 kgm2s-3

Dimensionless mass flow (theta)

1.0521

2.4084

1.0518

2.4531

Dimensionless impeller tip speed (Mu)

0.09610

0.1142

0.0961

0.1141

Flow Coefficient (phi)

10.9480

21.0939

10.9443

21.4968

Head Coefficient (psis)

62.7896

44.9674

61.4674

56.4016

Blade Loading Coefficient (psi)

-30.8864

-26.8368

-31.1422

-36.5630

Total pressure loss coefficient #1

-0.4682

-0.4741

-0.4570

-0.6156

Total pressure loss coefficient #2

-0.3189

-0.3216

-0.3137

-0.3811

Total pressure loss coefficient #3

-1.9815

-1.6631

-1.9327

-1.9434

Total pressure loss coefficient #4

-1.0856

-0.7808

-0.9559

-1.0037

Static Enthalpy Loss (zeta)

-2.9301

1.2805

-0.5424

2.0841

Total Enthalpy Loss (delta Q)

-0.5084

4.3531

0.4085

8.3622

Total-to-total isen. efficiency

1.0164

0.8378

0.9868

0.7713

Total-to-static isen. efficiency

0.0912

-0.2859

-0.0528

0.0034

Total-to-total poly. efficiency

1.0156

0.8464

0.9876

0.7862

Total-to-static poly. efficiency

0.0959

-0.3116

-0.0560

0.0037

Table 5.6 The results comparison between same and contra rotation engine

Discussion

Based on the preliminary design, the thermodynamics and aerodynamics properties are taken into consideration such as pressure ratio, thrust, flying level and some fluid properties. Then the inlet and outlet pressure must be determined together with temperature and work done in the eight stages compressor. The velocity triangles of the hub, mean and tip for both rotor and stator are design based on the constant degree of reaction, ʌ= 0.5 (it is assumed that an air inlet angle at rotor is equal to relative air inlet angle at stator) thus make the design of rotor and stator is approximately in the same profile. Note that each blade profile is different for every stages of compressor which depends to the annulus area. The value of pitch to chord ratio is determined based on the previous experiment at design deflection curve (Figure 5.14, page 139, Cohen, Rogers and Saravanmuttoo).

The blade is then created based on NACA 65A010 airfoil shape. Some parameters such as stagger angle, inlet and outlet angle, chord length and radius of section are required to design the blade. NACA 65A010 is selected because its shape is specifically for compressor application (Mattingly J.D, (1996)). The blade height is calculated according to mean radius, mass flow rate, and the axial velocity along the compressor stage. In this study we can use both aerodynamic and thermodynamics analysis to design the compressor blades. Other than that the De Haller criterion (the value of W2/W1 >0.7) must be satisfied to avoid surge and stall in the compressor.

In meshing results the H/J/L/G topology is selected automatically but finally the hexahedra mesh is more preferable with the blade shape in turbo machinery application. The bad percentage of the 3D mesh with smallest value can be neglected because it brings a minor effect to the mesh creation and simulation. The fine mesh with 250000 node counts shows that the results are more accurate than the medium mesh with 100000 nodes count in term of ratio comparison. Note that the computational results are compared with the analytical method results in term of pressure, temperature and its velocity distribution along the compressor stage.

In CFX results, the selection of physic definition is based on steady state condition and it is found that the difference between the results and analytical calculation is definitely high especially pressure for an axial compressor stage. Some probable reasons for this case are listed as below:

The blades model generated based on too many assumptions (random selection of pitch to chord ratio) that may deviated with the actual engine data.

The mesh of blades consist of small values of bad percentage (was ignored in the meshing process) may affect the simulation results/ output data.

The setup of CFX simulation is selected based on very limited choices of turbulence model (only k-€ model and shear-stress transport model) will reduce the accuracy of results hence produces a higher value of percentage difference compare to the analytical results. Note that the flow direction of fluid in setup must be parallel with the meshes direction rather than direct in normal to the passage.

For the combine stage (fan and first stage of an axial compressor), it is found that the contra rotation produce greater pressure and temperature ratio compare to the same direction rotation between fan and first compressor stage even though the greater losses found in it. Here, it is stated that the contra rotating engine is better than the same direction of rotation engine hence it can help the commercial airplane to save cost in term of specific fuel consumption (sfc).

Conclusion

In a nutshell, the contra rotation engine is more efficient than the same direction of rotation engine due to the rises of pressure and temperature in such configuration. Though the difference is very small but it is meaningful from the economics of view consideration by both manufacturer and buyer of this engine especially its specific fuel consumption. Note that in contra rotating engine more turbulence and separation will produce hence make greater losses in the stages. However it efficiency is still higher than that the conventional existing engine. In order to validate this data, I would like to recommend that further experimental testing must be conducted in the future studies.

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