i

ISTANBUL TECHNICAL UNIVERSITY FACULTY OF AERONAUTICS AND ASTRONAUTICS

Radial Compressor Analysis Using CFD for a Micro-Jet

GRADUATION PROJECT

Koray KURT

Department of Astronautical Engineering

ii

iii

ISTANBUL TECHNICAL UNIVERSITY FACULTY OF AERONAUTICS AND ASTRONAUTICS

Radial Compressor Analysis Using CFD for a Micro-Jet

GRADUATION PROJECT

Koray KURT

110150156

Department of Astronautical Engineering

iv

Koray KURT, student of ITU of Aeronautical and Astronautics student ID

110150156, successfully defended the graduation entitled “RADIAL

COMPRESSOR ANALYSIS USING CFD FOR A MICRO-JET” which he

prepared after fulfilling the requirements specified in the associated legislations,

before the jury whose signature are below.

Thesis Advisor: Prof. Dr. Aydın MISIRLIOĞLU …………………..

Istanbul Technical University

Jury Members: Prof. Dr. Fırat Oğuz EDİS …………………..

Istanbul Technical University

Assoc. Prof. Dr. Bayram ÇELİK …………………..

Istanbul Technical University

Date of Submission: 14 June 2021

Date of Defense: 28 June 2021

v

To all my family and my friends,

vi

FOREWORD

I would like to thank all my professors at Istanbul Technical University. Also,

I would like to thank Prof. Dr. Aydın Mısırlıoğlu for his contribution to the

realization of my thesis.

I would also like to thank my family and friends who supported me. I

especially thank the my close friend Rıdvan Kağan Altunkıran who support me in

this study.

June,2021 Koray KURT

vii

TABLE OF CONTENTS FOREWORD .............................................................................................................. vi

NOMENCLATURE .................................................................................................. viii

LIST OF TABLES ...................................................................................................... ix

LIST OF FIGURE ........................................................................................................ x

SUMMARY ................................................................................................................. 1

ÖZET............................................................................................................................ 2

1.INTRODUCTION .................................................................................................... 3

1.1 Radial (Centrifugal) and Axial Compressor ...................................................... 4

1.2 Radial (Centrifugal) Compressor ....................................................................... 5

1.3 Velocity of Triangle ........................................................................................... 8

1.4 Concepts of Computational Fluid Dynamics ..................................................... 9

1.5 Process of Computational Fluid Dynamics ...................................................... 10

2.DESIGN OF RADIAL COMPRESSOR ................................................................ 11

2.1. Calculations of Pressure and Mach Number ................................................... 11

2.1. Calculation of Velocity Triangle ..................................................................... 13

3. COMPUTATIONAL FLOW DYNAMICS OF THE COMPRESSOR ................ 15

3.1. Pre-Processing ................................................................................................. 16

3.1.1. Mesh Setup ............................................................................................... 17

3.2 Results .............................................................................................................. 21

3.2.1 Vista TF (Meridional) Results .................................................................. 21

3.2.2 CFX (Blade to Blade) Results ................................................................... 24

4. CONCLUSION ...................................................................................................... 26

5.REFERENCES ........................................................................................................ 27

viii

NOMENCLATURE

CFD : Computational Fluid Analysis

PR : Pressure Ratio

APU : Auxilary Power Units

SST : Shear Stress Model

RPM : Revolution Per Minute

ix

LIST OF TABLES

Table 1. Input Parameters ......................................................................................... 11

Table 2. Output Parameter ........................................................................................ 12

Table 3.Velocity Triangle ......................................................................................... 13

Table 4 Comparison of CFD Results ........................................................................ 14

x

LIST OF FIGURE

Figure 1. A) Axial Compressor B) Radial Compressor [3] .......................................... 5

Figure 2. Radial Compressor Flow Dimension[4] ....................................................... 5

Figure 3. Working Principle of Centrifugal Compressor [4] ....................................... 6

Figure 4. Configuration of Radial Compressor (Top View) [2] .................................. 7

Figure 5. Configuration of Radial Compressor (Side View)[2] ................................... 7

Figure 6. Velocity Triangle[1] ..................................................................................... 8

Figure 7. Process of CFD ........................................................................................... 10

Figure 8 Vista TF Project Schematic ......................................................................... 15

Figure 9. CFX Project Schematic ............................................................................... 15

Figure 10. Meridional View Figure 11. Auxiliary View ............ 16

Figure 12. 3D Geometric Model ................................................................................ 16

Figure 13 TurboMesh- Meshing ................................................................................ 18

Figure 14 Boundary Condition ................................................................................... 19

Figure 15 Boundary Condition Data Figure 16 Domain Physics for CFX 20

Figure 17 Cθ Meridional Flow Graph ......................................................................... 21

Figure 18 Cx Meridional Plot Vista TF ..................................................................... 22

Figure 19. Temperature Meridional Plot Vista TF ..................................................... 23

Figure 20. Static Pressure Meridional Plot Vista TF ................................................. 23

Figure 21. Temperature Blade to Blade Flow ............................................................ 24

Figure 22. Mach Number Blade to Blade Flow ......................................................... 24

Figure 23. Pressure Blade to Blade Flow ................................................................... 25

Figure 24. Mach Number Meridional Flow ............................................................... 25

xi

1

RADIAL COMPRESSOR ANALYSIS USING CFD FOR A

MICRO-JET

SUMMARY

Radial (centrifugal) compressors are now widely utilized variety of sectors

for a variety of reasons. In the aviation and defense industry, in various engine types

such as turbojet, turbofan, turboshaft, it compresses the fluid to entrance the engine

and increases its pressure. In this thesis, the differences between the radial

compressor and the axial compressor, the advantages and disadvantages of the radial

compressor are discussed. In addition, a radial compressor design has been made for

a small jet engine, and the design has been analyzed with computational fluid

dynamics (CFD), and data such as density, Mach number, temperature have been

examined. The analysis was first examined as a two-dimensional and then a three-

dimensional flow. The results obtained at the end of the analysis were compared with

the calculated results of the designed compressor. It also contributed to the study by

considering the reference articles on design and analysis for the radial compressor.

2

RADIAL COMPRESSOR ANALYSIS USING CFD FOR A

MICRO-JET

ÖZET

Radyal (santrifüj) kompresörler günümüzde birçok sektörde kullanılmaktadır.

Havacılık ve savunma alanında ise; turbojet, turbofan, turboşaft gibi çeşitli motor

türlerinde motora giren akışkanı sıkıştırarak basıncının artırılmasını sağlar. Bu

çalışmada, radyal kompresörle eksenel kompresör arasındaki farklar, radyal

kompresörün avantaj ve dezavantajları ele alınmıştır. Ayrıca küçük boyutlardaki bir

jet motoru için radyal kompresör tasarımı yapılmış olup, tasarımın hesaplamalı

akışkanlar dinamiği (HAD) ile analizi yapılarak yoğunluk, mach sayısı, sıcaklık gibi

verilerin incelenmesi gerçekleştirilmiştir. Analiz öncelikle iki boyutlu ardından üç

boyutlu akış olarak incelenmiştir. Analiz sonunda elde edilen sonuçlar tasarlanan

kompresörün hesaplanan sonuçlarıyla karşılaştırılmıştır. Ayrıca radyal kompresör

için tasarım ve analiz yapan referans makaleler de göz önüne alınarak çalışmaya

katkı sağlamıştır.

3

1.INTRODUCTION

During the Second World War, tremendous progress was achieved in the

construction of gas turbines, with particular emphasis on the basic turbojet engine.

When it became evident that smaller gas turbines would require centrifugal

compressors, major research and development activity resumed. Simple turboprops,

turboshafts, and auxiliary power units (APUs) have been manufactured in

considerable quantities and almost all have utilized centrifugal compressors such as

those manufactured by Pratt & Whitney Canada, Honeywell, and others. They are

also employed as the high-pressure spools in small turbofans, which is one of the

earliest applications of a centrifugal compressor as the high-pressure spool in a small

turbofan. Centrifugal were originally chosen for their ability to handle small-volume

flows, but they also have several other advantages, including a shorter length than an

equivalent axial compressor, greater resistance to foreign object damage, less

susceptibility to performance loss due to deposit build-up on the blade surfaces, and

the ability to operate over a wider range of mass flow at a given rotational speed.

(Saravanamutto,1972) [1]

4

1.1 Radial (Centrifugal) and Axial Compressor

Radial and axial compressors are used in gas turbine engines to compress the

air that enters the engine, allowing the fluid to enter the combustion chamber at a

higher pressure. If it is necessary to compare radial and axial flow compressors in

terms of quality, there are some applications where each has an advantage.

Axial compressors are lighter in weight because they have smaller engine

diameter. Assuming the same mass flow for both types of compressors, the frontal

area required to achieve a given pressure ratio in an axial compressor is half that of a

centrifugal compressor. Due to manufacturing constraints, the diameter of the

centrifugal impeller, and therefore the mass flow and pressure ratio capabilities, has a

realistic maximum limit of around 0.8 m. Axial compressors have a higher isentropic

efficiency at mass flow rates larger than around 5 kg/s; the degree of this advantage

grows with mass flow rate. (Walsh and Fletcher,2004) [2] Therefore, for this reason,

an axial compressor is preferred in gas turbine engines exposed to high mass flow.

The radial compressor provides a higher compression ratio than the axial

compressor in a single stage. Assuming the same mass flow for both types of

compressors, Centrifugal compressor is lower than axial compressor in cost. In radial

compressor, isentropic efficiency is improved for mass flow rates substantially less

than 5 kg/s. It is because the efficiency of an axial flow compressor quickly

decreases when the size of the compressor is decreased owing to the growing relative

levels of tip clearance, blade leading and trailing edge thicknesses, and roughness of

the blade’s surface with fixed manufacturing tolerances. (Walsh and Fletcher, 2004)

Centrifugal compressor surge lines, defined as a line along which there is a

discontinuity in flow speed, are less susceptible to high tip clearance than axial

compressor surge lines because the pressure rise does not entirely show up

differential pressure across each blade.

5

Figure 1. A) Axial Compressor B) Radial Compressor [3]

1.2 Radial (Centrifugal) Compressor

Flow is drawn into a centrifugal compressor with radial blades and pushed

around the compressor by centrifugal force. This is known as the radial discharge

flow, and it is a distinguishing characteristic of centrifugal compressors. The

centrifugal impeller changes the direction of flow from axial to radial, and this is

accompanied by a radial diffuser. The increasing diameter results in a much higher

area ratio and thus greater diffusion in both than is possible with an axial flow stage.

Figure 2. Radial Compressor Flow Dimension[4]

6

Centrifugal compressors are composed mostly of a stationary casing

containing a rotating impeller that imparts high velocity to the air and a series of

fixed diverging tubes through which the air is decelerated, resulting in an increase in

static pressure. Due to the fact that the latter operation is a diffusion one, the

compressor component holding the diverging passageways is referred to as the

diffuser.

Figure 3. Working Principle of Centrifugal Compressor [4]

Figure 4 and Figure 5 show the parts of the radial compressor. The inducer

refers to the point at which flow first makes contact the radial compressor. The part

where the air from the compressor passes to the diffuser is also called the exducer. Impeller is the name given to the rotating body part of the compressor. The point

where the blades intersect with the impeller is called the hub, and the farthest point

from the impeller is called the shroud.

The radial diffuser is the part of the exducer that receives the flow from the

exducer region. The use of an axial diffuser ensures that the flow enters the

combustion chamber as steeply and laminarly as possible when approaching the

chamber. (Figure 5)

7

Figure 4. Configuration of Radial Compressor (Top View) [2]

Figure 5. Configuration of Radial Compressor (Side View) [2]

8

1.3 Velocity of Triangle

While designing the aerodynamics of a radial compressor, first of all, velocity

triangles are created. Velocity triangles provide answers to questions such as how the

calculated input and output velocities will enter the impeller and how large the

absolute and relative velocities should be, allowing for blade design. The velocity

vector of a fluid particle flowing through a turbomachine is most readily represented

in cylindrical dimensions, with the z-coordinate corresponding to the machine's

rotational axis.

Compressor inlet gas arrives with a velocity of C1 when it enters the impeller

eye. C refers to tangential velocity. The beta and alpha angles are used to calculate

the relative (W) and absolute (U) velocity. When determining these elements, it is

preferable to utilize the best practice values that have been found in previous studies.

The relative velocity is vectorially in the direction that the fluid enters the blade.

Figure 6. Velocity Triangle [1]

9

1.4 Concepts of Computational Fluid Dynamics

Computational fluid dynamics (CFD) is an analysis method that provides the

numerical solution of flow, heat and mass transfer problems. This method analyzes

with the help of differential equations that describe the motion of the fluid, and by

solving these equations numerically, pressure, velocity and temperature distributions

in the flow are obtained. In line with these results, performance data such as

efficiency and pressure ratio can be obtained.

CFD is focused with utilizing computers to provide numerical solutions to

fluid flow issues. CFD can now solve a wide variety of flow problems, including

those that are compressible or incompressible, laminar, or turbulent, chemically

reactive or non-reacting. (Zawawi and Saleha, 2018) [5]

To study the object under investigation, control volumes are generated. On

top of these volumes, a velocity and pressure field are generated. In the x, y, and z

coordinate systems, velocity is represented as the letters u, v, and w. Pressure and

velocity are the four unknowns that are being solved using four separate equations.

The first of these equations is derived from the conservation of mass, and the

remaining three equations are derived from the conservation of momentum.

First principle is continuity equation:

𝜕𝜌

𝜕𝑡+

𝜕

𝜕𝑥(𝜌𝑢) +

𝜕

𝜕𝑦(𝜌𝑣) +

𝜕

𝜕𝑧(𝜌𝑤) = 0 1.1

Due to the fact that momentum is a vector quantity, it will have three

components. Three separate equations will be generated by this method:

𝜌 [𝜕𝑢

𝜕𝑡+

𝜕𝑢

𝜕𝑥𝑢 +

𝜕𝑢

𝜕𝑦𝑣 +

𝜕𝑢

𝜕𝑧𝑤] = −

𝜕𝜌

𝜕𝑥+ 𝜇 (

𝜕2𝑢

𝜕𝑥2+

𝜕2𝑢

𝜕𝑦2+

𝜕2𝑢

𝜕𝑧2) + 𝜌𝑔𝑥 1.2

𝜌 [𝜕𝑣

𝜕𝑡+

𝜕𝑣

𝜕𝑥𝑢 +

𝜕𝑣

𝜕𝑦𝑣 +

𝜕𝑣

𝜕𝑧𝑤] = −

𝜕𝜌

𝜕𝑦+ 𝜇 (

𝜕2𝑣

𝜕𝑥2+

𝜕2𝑣

𝜕𝑦2+

𝜕2𝑣

𝜕𝑧2) + 𝜌𝑔𝑦 1.3

𝜌 [𝜕𝑤

𝜕𝑡+

𝜕𝑤

𝜕𝑥𝑢 +

𝜕𝑤

𝜕𝑦𝑣 +

𝜕𝑤

𝜕𝑧𝑤] = −

𝜕𝜌

𝜕𝑧+ 𝜇 (

𝜕2𝑤

𝜕𝑥2+

𝜕2𝑤

𝜕𝑦2+

𝜕2𝑤

𝜕𝑧2) + 𝜌𝑔𝑧 1.4

10

The unknowns u, v, w, and p can be numerically solved using these four

equations. These equations are known as Navier-Stokes equations. These equations

are known as Navier-Stokes equations.

1.5 Process of Computational Fluid Dynamics

• In the pre-processing part, the geometry of the model is generated. It is

created computational domain. İt is made the definition of flow physics and

created mesh.

• In the computation part, İt is defined solver setting and generated compute

solution.

• In the post-processing part, it is performed convergence studies. The results

are examined and the validate result is compared with measured data.

Figure 7. Process of CFD

Pre-Processing

Computation Post-

Processing

11

2.DESIGN OF RADIAL COMPRESSOR

2.1. Calculations of Pressure and Mach Number

When designing the radial compressor, it is critical to consider aspects such

as mass flow rate, pressure ratio, and mach number. Values such as inlet temperature,

inlet pressure, compression ratio, mass flow rate are determined as input values

according to the desired thrust. In the calculation of the Mach number, the rule that

the mass flow rate remains constant at the inlet and outlet is applied and the

appropriate Mach numbers are calculated. In this way, the speed at the entrance and

exit is determined. The revolutions per minute and distances of the compressor have

been determined by the best examples and benchmark analysis. Table 1. shows the

inlet parameters of the radial compressor.

Pt,1 101325 Pa

Tt,1 288.15 K

İnlet Diameter 0.1 m

Outlet Diameter 0.16 m

Pressure Ratio (PR) 4.8

Mass Flow Rate 1 kg/s

RPM 55500 rev/m

Table 1. Input Parameters

The compression ratio is found by dividing the total outlet pressure by the

total inlet pressure.

𝑃𝑅 =𝑃0,1

𝑃0,2 2.1

Static pressure (Ps) formula is:

𝑃𝑠 = 𝑃𝑡(1 +𝛾 − 1

2𝑀2)

−𝛾𝛾−1 2.2

Flow rate remains constant at the inlet and outlet is applied and the

appropriate Mach numbers are calculated. Heat capacity ratio is determined as 1.4.

For an ideal compressible gas flow rate formula:

12

�̇� =𝐴𝑃𝑡

√𝑇𝑡

√𝛾

𝑅𝑀(1 +

𝛾 − 1

2𝑀2)

−𝑦+1

2(𝑦−1) 2.3

The ratio of radial compressor outlet radius to inlet radius should be between

1.3 and 1.6 in turbojet and turbofan engines according to best practice. (Walsh and

Fletcher,2004) This range has increased to 1.8 in new design. This ratio is

determined as 1.6, so the outlet diameter was calculated as 0.16 m.

The previously known temperature, pressure, and mass flow information are

also used to determine the mach number at the inlet and outlet of compressor. The

values calculated here will be compared to the Ansys results in the following sections

of the thesis.

Pt,2 486365 Pa

Ps,1 88179.84 Pa

Ps,2 465666 Pa

M1 0.45

M2 0.25

Outlet Diameter 0.16 m

Table 2. Output Parameter

13

2.1. Calculation of Velocity Triangle

The impeller's velocity values at the inlet and outlet, as well as the entry

angles of the speeds into the blade, are critical considerations for designing the

impeller. These calculations are carried out using the velocity triangles, which are

described in detail in section 1.3.

The tangential velocity (C) is calculated with the following formula for inlet

and outlet.

𝐶 = 𝑀√𝛾𝑅𝑇𝑠

(1 +𝛾 − 1

2 𝑀2) 2.4

Absolute velocity depends on the revolution per minute and is found by

formula 2.5. N refers to revolution per minute (RPM) and d is the compressor

diameter. The inlet diameter is used for inlet absolute velocity U1, The outlet

diameter (d2) is used for the outlet absolute velocity U2.

𝑈 =𝜋𝑁𝑑

60 2.5

At the compressor inlet, the alpha angle(α) can be taken as 0, so it is accepted

as:

𝐶𝑥,1 = 𝐶 𝑎𝑛𝑑 𝐶𝜃,1 = 0 2.6

It is possible to determine the alpha and beta angles at the entrance and exit

by using geometrical calculations. After that, the triangle is used to calculate the

relative velocity(W).

Table 3.Velocity Triangle

Compressor Inlet Compressor Outlet

Alpha (α) 0 73.4

Beta (β) 63.7 47

Cθ 0 352.19

Cx 147.16 105.12

C 147.16 367.54

U 297.57 464.96

W 331.97 154.56

14

2.2 Aerodynamic Design of The Compressor

ANSYS-VISTA CCD (Centrifugal Compressor Design) is used for radial

(centrifugal compressor). Aerodynamic data, geometry input values, gas information

were entered in Vista CCD interface. In this way, a suitable blade design was made

through the program. The design is transferred to the BladeGen to perform mesh

operations and three-dimensional analysis and to edit the blade design.

The Vista CCD program generates a result page that contains the values that

were entered as input. The goal of the design is to reduce as much as possible the

ratio between the calculated values and the results found in the program. In this way,

the design can be controlled. Calculated values and ANSYS results are given in

Table 4. In addition, margins of error are calculated.

In the program, pressure ratio, mass flow rate rotation speed is entered. In

addition, the incidence angle is 1.5 degrees, the isentropic efficiency is 0.85, and the

calculated relative velocity ratio is 0.48, added to the aerodynamic data part. The

dimensions of the hub and shroud are entered in the geometry section. The number of

blades has been determined to be nine in number. It is selected ideal gas as gas

properties model and air is selected as material.

Table 4 Comparison of CFD Results

Calculated Values Vista CCD Results Deviation

Outlet

Diameter(D2)

16 17.2 %7.2

Alpha (α)(outlet) 73,4 76,7 %4.49

Absolute Velocity 464.96 504.68 %8.6

15

3. COMPUTATIONAL FLOW DYNAMICS OF THE

COMPRESSOR

In compressor design, VistaCCD for blade design, VistaTF for two-

dimensional flow calculations, BladeGen for blade design arrangements and

detailing, TurboGrid for mesh, and CFX module for solution. On Vista TF, pressure

and velocity analyses are demonstrated first in section 3.1, followed by three-

dimensional analyses, which are demonstrated in section 3.3 on CFX module. The

concept of meshing over TurboGrid is explained in Section 3.2.

Figure 8 Vista TF Project Schematic

Figure 9. CFX Project Schematic

16

3.1. Pre-Processing

The pre-processing section generates the model's geometry. It is a

computational domain that has been created. It is defined the flow physics and

created the mesh. Below is a geometric representation of the compressor as generated

by BladeGen.

Figure 10. Meridional View Figure 11. Auxiliary View

Figure 12. 3D Geometric Model

17

Another step in the pre-processing stage is the determination of the boundary

conditions. The boundary conditions are required to solve the equation system. The

mathematical model must include boundary conditions. Boundaries control the

direction of flow. Cell zones denote the fluid and solid regions. Cell zones are

assigned to material and source terms. Face zones are used to represent boundaries

and internal surfaces. Face zones are assigned boundary data. At the inlet, a velocity

inlet boundary condition was used. Typical boundary conditions in CFD are no-slip,

axisymmetric, inlet, outlet, and periodic. (Zawawi and Saleha, 2018)[5] In the mesh

setup section (3.1.1) below, the operations applied for the boundary layer will be

explained.

3.1.1. Mesh Setup

The compressor, the details of which are provided in BladeGen, must be

covered with fabric in order to conduct CFD analysis. Turbogrid was used to remove

the mesh. To begin, in the Turbogrid, the tip-fan section was selected using the

normal distance in the shroud tip section and a value of 0.25 mm was entered. The

topology set was split using the single splitter method. By selecting the global size

factor for mesh facades, a value of 1 was entered for the size factor. The maximum

expansion ratio that should be used is 1.3, and the first element of the set method

should be used to specify mesh qualities. The first element of the first set method

creates 5 micron-sized elements on the boundary layer surfaces. The mesh detail and

image of the mesh are shown in Figure 13. According to the mesh statistics, there are

677198 nodes and 634845 elements in the system. Following the completion of the

TurboGrid mesh process, CFX was applied.

18

Figure 13 TurboMesh- Meshing

When it comes to turbulent flow, Ansys CFX offers two different models to

choose from: the k- ω Shear Stress model and the k- ε Model. Full-turbulent non-

separated flows are well represented by the k-ε model. It is not capable of calculating

extremely accurate flow fields that display a reverse pressure gradient and a high

degree of curvature. With the k- ε model, it is possible to simulate problems

involving an external body with more accuracy. In order to provide accurate

modeling in near-wall treatment and internal flow, such as turbomachinery, the k- ω

shear stress model (SST) is recommended. (Chaudhary et all,2018)[6] In this study,

SST model is used to simulate steady-state flow simulation by referencing the

information in the mentioned article.

The total pressure at the inlet and the static pressure at the output are defined

for the solution. While no wall velocity value is entered for the hub part, the counter

rotating wall is selected for the shroud, and the shroud part is rotated in the opposite

direction compared to the rotor part. CFX was also modified to include a steady state

solution, which was achieved by selecting a centrifugal compressor from the tools

section.

19

Figure 14 Boundary Condition

20

Figure 15 Boundary Condition Data Figure 16 Domain Physics for CFX

21

3.2 Results

3.2.1 Vista TF (Meridional) Results

Below are the meridional flow diagrams from Vista TF. The figures of Cx and

Cθ are depicted in the first two figures. Because there was only axial velocity at the

input, the value 0 was accepted in the Cθ value of inlet that were calculated. In

accordance with the velocity triangles(Table 3) , the calculated Cθ value at the outlet

of the radial compressor was discovered to be approximately 350 m/s.

The Cx value was accepted as equal to the Cθ value in velocity triangles

because the input alpha angle is equal to zero in this case. As a result, the calculated

Cx value is 147 meters per second. The calculated value for the radial compressor

output Cx was approximately 105 m/s, which was found to be accurate. It was

calculated that the value of Cx had decreased slightly.

The meridional graphs below show the change in Cx and Cθ values from the

input to the output, respectively. This observation is in accordance with the

calculated results, as shown by the fact that the Cθ graph begins at zero meters per

second and accelerates to 350 meters per second at the exit.

Figure 17 Cθ Meridional Flow Graph

22

The calculated tangential velocity of input of radial compressor is 147.16 m/s.

In the meridional graph of the tangential velocity, it is seen that the input velocity

produces a result parallel to the calculated velocity. Upon closer inspection of the

meridian graph, it can be seen that the speed reaches its highest levels in the shroud

part and its lowest levels in the hub part. According to expectations, the Cx value

decreased slightly at the compressor's outlet. When we take a look at the color

contour, we can see that the speed at the output is in the range of 100-110 m/sec.

Figure 18 Cx Meridional Plot Vista TF

The centrifugal compressor's characteristic radial discharge flow draws air

into a rotating compressor with radial blades and pushes it toward the compressor's

periphery via the effect of centrifugal force. Pressure is increased and kinetic energy

is produced as a result of the outlet area being smaller than the compressor inlet area

and the radial movement of the air. Due to the fact that the adiabatic solution is

formed, the temperature should increase proportionately to the increase in pressure.In

addition, Since high pressure changes on the impeller will put the compressor in

23

surge/stall state, it can create an unacceptable situation. The pressure diagram shown

in the meridional flow indicates no such problem from the engine.

Figure 19. Temperature Meridional Plot Vista TF

Figure 20. Static Pressure Meridional Plot Vista TF

24

3.2.2 CFX (Blade to Blade) Results

The CFX Mach number results of the 3D analysis, whose mesh adjustment

was made and setup was explained in the Turbo Mesh program in the previous

sections, are given below. In addition, Blade to blade plots of temperature pressure

and mach number.

Figure 21. Temperature Blade to Blade Flow

Figure 22. Mach Number Blade to Blade Flow

25

Figure 23. Pressure Blade to Blade Flow

Figure 24. Mach Number Meridional Flow

26

4. CONCLUSION

In this thesis, A small-scale turbojet engine was investigated, and the

compressor design and flow analysis were carried out. In order to determine whether

to use an axial or radial compressor, it was determined that the radial compressor

would be more appropriate due to the sizing and compression ratio parameters. The

values for velocity and pressure were calculated using the appropriate compression

ratio. When performing these calculations, it was assumed that the mass flow rate

would remain constant. With the help of the velocity values that were discovered,

velocity triangles were formed, and the aerodynamic design of the radial compressor

was successfully completed.

When comparing the calculated values to those found in the Vista CCD

program, there are some margins of error to be considered. This comparison is given

as a table in the study. The outlet diameter was compared with the outlet flow angle

and absolute velocity CFD results and approximately 7%, 4% and 8% deviation were

detected, respectively. Comparing similar articles [6], it has been discovered that

similar percentages of error are also included in those articles when comparing

similar articles. One of the parameters that should be improved in the study is the

reduction of these deviations, which can be considered as one of the improvement

targets.

Upon examination of the flow charts, it was determined that the observed

velocity values corresponded to the calculated velocity values. The Mach number

was also analyzed three-dimensionally, and no issues with blade-to-blade design

were discovered. The static pressure diagram is identified as an additional

component of the thesis that requires development.

As a result, the positive and negative aspects of the radial compressor

designed in this study were observed thanks to the flow analysis.

27

5.REFERENCES

[1] Cohen, H., Rogers, G. F. C., & Saravanamuttoo, H. I. H. (1972). Gas turbine

theory. London: Longman.

[2] Walsh, P., Fletcher, P. (2004). Gas turbine Performance. Malden, MA.

[3] Gaurav,G. (2019).Preventing Choke and Surge in a Compressor. Retrieved

from https://blog.softinway.com/preventing-choke-and-surge-in-a-

compressor/

[4] Tun, K.N.Z., Zaw, C.(2014). International Journal of Scientific Engineering

and Technology Research Volume.03, IssueNo.11. Pages: 2554-2558.

[5] Zawawi, M. H., Saleha, A., Salwa, A., Hassan, N. H., Zahari, N. M., Ramli,

M. Z., & Muda, Z. C. (2018). A review: Fundamentals of computational fluid

dynamics (CFD). doi:10.1063/1.5066893

[6] Chaudhary.A, et al. (2018). The 6th International Symposium-Supercritical

CO2 Power Cycles, Pittsburgh, PA