Gas Turbine,Finned Heat Exchanger

Republic of Iraq

Ministry of Higher Education and Scientific Research

Al- Mustansiriya University

College of Engineering

Mechanical Engineering Department

The Enhancement of Two- Shaft Gas Turbine Performance Using Improved Air Temperature

A THESIS SUBMMITTED TO THE COLLEGE OF ENGINEERING OF AL- MUSTANSIRIYA UNIVERSITY AS PARTIAL FULLFILMENT OF

THE REQUIREMENTS FOR THE DEGREE OF MASTER IN MECHANICAL ENGINEERING SCIENECE

BY Ali Ahmed Abdulrasool

Supervised by

Dr.Ali Hussain Tarrad Asst.Prof Dr. Fouad Alwan Saleh

2009

بسم الله الرحمن الرحيم

الكتاب والميزان لقد أرسلنا رسلنا بالبينات وأنزلنا معهم

ليقوم الناس بالقسط وأنزلنا الحديد فيه بأس شديد ومنافع

للناس وليعلم الله من ينصره ورسله بالغيب إن الله قوي

عزيز

صدق اهللا العظيم

)25(ية د اآلسورة الحدي

Supervisor Certificate

We certify that the preparation of this thesis entitled (THE

ENHANCEMENT OF TWO SHAFT GAS TURBINE PERFORMANCE

USING IMPROVED AIR TEMPERATURE) was made under our

supervision at the Mechanical Engineering Department , College of

Engineering , Al-Mustansiriya University in partial fulfillment of the

requirements for the degree of MASTER OF SCIENCE IN MECHANICAL

ENGINEERING.

(SUPERVISORS) Signature:

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In view of the available recommendation I forward this thesis for

debate by the examining committee

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(CHAIRMAN OF MECHANICAL ENGINEEING DEPARTMENT)

EXAMINATION COMMITTEE CERTIFICATE

We certify that we have read the thesis entitled (THE ENHANCEMENT OF

TWO SHAFT GAS TURBINE PERFORMANCE USING IMPROVED

AIR TEMPERTURE) and as an examining committee, examined the student

(Ali Ahmed Abdulrasool) in its context and that in our opinion it is adequate

as a thesis for the degree of MASTER OF SCIENCE IN MECHANICAL

ENGINEERING.

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DEDICATION

TO MY DEAR FAMILY

WITH

LOVE AND RESPECT

ALI

Acknowledgement

All praise is due to Allah, the lord of the worlds

I would like to express my deep sense of appreciation and thank to my

supervisors (Dr.Ali Hussain Tarrad and Dr. Fouad Alwan Saleh)

for their continuous encouragement and support during the study.

Finally, I would like to thank for all of people who helped me and introduce

their opinion and companion to end this thesis including: Mr. Ahmed Muneer

Dein, Mr. zyaad Talal ,Mr. Wa'al Najam and Mr. Nehaad Hashim.

Ali

ABSTRACT In the present research, a study of the performance improvement of a two-

shaft gas turbine engine is conducted. This has been accomplished by utilizing

sensible load cooling system. It is related to the effect of reducing compressor

inlet temperature on gas turbine performance. An experimental study has been

done on an existing a two-shaft gas turbine with a (5 kW) power, taking into

consideration the effect of reducing compressor inlet temperature. An

instrumented experimental rig was built for this object by adding an air-

cooled heat exchanger in series with water supplying system. The

experimental results for gas turbine performance showed that the percent of

design increases by (15%) for the power output increasing overall efficiency

by (25%).Moreover, heat consumption has been reduced by (10%) when

reducing compressor inlet temperature from (30°C to 15 °C).The percent of

design is defined by the ratio between the parameter difference at both

temperatures to the design point.

A simplified new numerical model based on the step by step technique has

been developed for the design and predicting the air cooled heat exchanger

performance. The numerical model was designed in a new form so that

variation of all design parameters can be calculated. The model has been

checked and validated using the experimental laboratory data. The maximum

discrepancy between the experimental data and model predicted values of the

overall heat transfer coefficient and heat load were about (5%). This

percentage value was obtained for the given range of the simulated conditions.

Furthermore, a computational model program to investigate theoretically the

effect of compressor inlet temperature on the performance of the gas turbine

has been developed. The model utilizes a non-dimensional approach. The

discrepancy percentage between the experimental and theoretical predictions

was about (18 %).

TABLE OF CONTENTS

Title

Pages

List of Figures VIII

List of Tables XII

Nomenclature XIII

CHAPTER ONE: Introduction 1

1.1 General 2

1.2 Two Shaft Engine 2

1.3 Influence of External Factors on the Gas Turbine Performance 3

1.4 Gas Turbine Inlet Air Cooling 5

1.4.1 Evaporative Cooling 5

1.4.2 Cooling with Absorption Chiller 5

1.5 Compact Heat Exchanger 6

1.6 Aims of the Present Work 7

CHAPTER TWO : Literature Survey 8

2.1 General 9

2.2 The Previous Work 9

2.2.1 Modeling and Simulation of Air Cooled Heat Exchanger 9

2.2.2 Performance Improvement of the Gas Turbine Engine 11

2.3 The Present Work 16

CHAPTER THREE: Experimental Work 18

3.1 Test Rig for Heat Exchanger 19

3.1.1 Heat Exchanger 19

3.1.2 Water Supplying System 19

3.1.3 Measuring Instrumentation 21

3.1.4 The Electrical Board 21

3.1.5 Air Circulation System 22

3.2 Experimental Setup 26

3.3 Gas Turbine 26

3.3.1 Design Information 28

3.3.2 Performance design 28

3.4 Operating Principles 28

3.5 Main Component Parts of Gas Turbine 29

3.5.1 Gas Generator: 29

3.5.1.1 Centrifugal Compressor 29

3.5.1.2 Combustion Chamber 30

3.5.1.3 Gas Generator Turbine 31

3.5.2 Power Turbine 31

3.5.3 Fuel System 32

3.5.4 Oil Lubricating System 32

3.5.5 Starting System 33

3.6 Dimensionless and Parameter Groups 33

3.6.1 Corrected Compressor Data 34

3.6.2 Corrected Basic Data 35

3.6. 3 Corrected Derived Data 36

3.7 Experimental Work Procedure 40

3.8 Test Procedure 40

CHAPTER FOURE : Mathematical Model 43

4.1 General 44

4.2 Heat Exchanger Thermal Design 44

4.2.1 A Comprehensive Design Procedure 44

4.2.2 Numerical Modeling of Cross Flow Compact Heat Exchanger 46

4.2.2.1 Grid System 46

4.2.2.2 Physical Characteristics of Heat Exchanger 49

4.2.2.3 Mass Conservation 49

4.2.2.4 Log-Mean Temperature Difference 50

4.2.2.5 Heat Load 51

4.2.2.6 Overall Heat – Transfer Coefficient (Uo) 51

4.2.2.7 Forced Convection Heat Transfer coefficient Inside Tube 52

4.2.2.8 Forced Convection Heat Transfer coefficient for Air Side 52

4.2.2.9 Power of Fan 54

4-2-2-10 The Computer Program 55

4.3 Gas Turbine 57

4.4 Basic Gas Turbine Cycles 58

4.5 A non-Dimensional Analyses of Gas Turbine

Performance

58

4.5.1 Component Performance 59

4.5.2 Graphical Plot 61

4.6 Computer Calculations for Two-Shaft Gas Turbine 61

4.7 Results and Discussion of Theoretical Calculations 62

4.7.1 The effect of compressor inlet temperature on maximum to

minimum ratio (Ø)

62

4.7.2 The effect of compressor inlet temperature on the expansion

ratio

64

CHAPTER FIVE : Results & Discussion 70

5.1 General 71

5.2 Computational Model Results for Heat Exchanger 71

5.2.1 Heat Load for heat exchanger 72

5.2.2 Heat Transfer Coefficient for air side 72

5.2.3 Overall Heat Transfer Coefficient 73

5.2.4 Air Temperature Distribution 74

5.2.5 The Effect of Core Aspect Ratio (H/L) and Size (LXDXH) 74

5.3 Experimental Results for Gas Turbine 75

5.3.1 Effect of Compressor Inlet Temperature on Power Output 76

5.3.2 Effect of Compressor Inlet Temperature on Fuel Mass Flow

Rate

76

5.3.3 Effect of Compressor Inlet Temperature on Specific Fuel

Consumption

77

5.3.4 Effect of Compressor Inlet Temperature on Heat

Consumption

77

5.3.5 Effect of Compressor Inlet Temperature on Heat Rate 78

5.3.6 Effect of Compressor Inlet Temperature on Overall

Efficiency

78

5.3.7 Effect of Compressor Inlet Temperature on Air Mass Flow

Rate

78

5.3.8 Effect of Compressor Inlet Temperature on Pressure Ratio 78

5.3.9 Effect of Compressor Inlet Temperature on Power Input to

Compressor

79

5.3.10 Effect of Compressor Inlet Temperature on Turbine Inlet

Temperature

79

5.4 Percent of Design 80

5-5 Comparison between the Experimental and Theoretical

Predictions of the gas turbine engine

81

5.6 Conclusion 81

CHAPTER SIX : Conclusions & Recommendations 114

6.1 Conclusions 115

- 6.2 Recommendations 116

References 117

Appendix (A)Experimental Work ,Data Tables 122

Appendix (B) Flow Charts and Computer Program 134

Appendix (C)Gas Turbine parameter Groups 146

List of Figures Title

Pages

Figure (1.1) Two Shaft Gas Turbine Diagram 3

Figure (1.2) Influences of External Factors on the Gas Turbine Performance 4

Figure (1.3) Evaporative Cooler 5

Figure (1.4 )Air "Chilling Cooling" System, Based on Absorption 6

Figure (2.1)Effect of Compressor Inlet Temperature on Gas Turbine Performance 13

Figure (2.2) Temperature-Power-Speed Interrelationships 15

Figure 2.3 Net Output Power Versus Inlet Temperature for Gas Cycle 16

Figure (3.1a) Schematic Diagram of the Built Rig (Heat Exchanger) 23

Figure (3.1b) Configuration of the Built Rig (Heat Exchanger) 24

Figure (3.1c) Configuration of the Test Rig (Inlet Cooling System) Preparing to Gas Turbine

24

Figure (3.3.a) Top View of Heat Exchanger Geometry 25

Figure (3.3.b) Front View of Heat Exchanger Geometry 25

Figure (3.3) Schematic Diagram of the Test Rig (Overall) 27

Figure.(3.4) Schematic Arrangement for Gas Turbine (GT-85) 32

Figure (4.1) Methodology of Heat Exchanger Design 45

Figure (4.2) Step by Step method with two Directions 47

Figure (4.3a) Slice for inlet single Tube 48

Figure (4.3b) Exit of one row inlet to next row 48

Figure (4.3c) Nodal Points Distributions with two directions 49

Figure (4.4) The Mean Temperature Difference Along a Single Pass 50

Figure (4.5) Basic Gas Turbine Engine 59

Figure (4.6 )(T-s) Diagram for Irreversible Two-Shaft Circuit Simple Plant 56

Figure (4.7)Specific Heats and Their Ratios for ‘Real’ Gases-Air and Products of Combustion

62

Figure (4.8) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=699 0C

65

Figure (4.9) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=628.5 0C

65

Figure (4.10) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=588 0C

66

Figure (4.11) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=547.5 0C

66

Figure (4.12)The Effect of Turbine Inlet Temperature (High Pressure Turbine)on

the Theoretical Overall Efficiency ,Expansion Ratio=1.12

67

Figure(4.13) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on

the Theoretical Overall Efficiency ,Expansion Ratio=1.08

67

Figure (4.14) Theoretical Overall Efficiency as a Function of Expansion Ratio with Isentropic Efficiency (ηt,ηc =0.9) andT1=150C

68

Figure (4.15) Theoretical Overall Efficiency as a Function of Expansion Ratio with Isentropic Efficiency (ηηt,ηc =0.9) andT1=300C

68

Figure (4.16) The Effect of Compressor Inlet Temperature (with Variable Expansion Ratio)on the Theoretical Overall Efficiency, T3= 669 0C

69

Figure(4.17)The Effect of Compressor Inlet Temperature (with Variable Expansion Ratio) on the Theoretical Overall Efficiency, T3= 628.5 0C

69

Figure (5.1) Comparison between the Experimental and Present Model for the Effect of Air Velocity on Heat Load at Water Flow Rate 2000 (L/h)

93

Figure (5. 2) Comparison between the Experimental and Present Model for the Effect of Air Velocity on Heat Transfer Coefficient.ha (w/m2.c) at Water Flow Rate 2000 (L/h)

94

Figure (5. 3) Variation Heat Transfer Coefficient ha.(w/m2.c) a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate 2000 cfm

95

Figure(5. 4) Variation Heat Transfer Coefficient ha.(w/m2.c) a long Heat Exchanger Height at Water Flow Rate 2000 (L/h),Air Flow Rate2000 cfm

96

Figure (5. 5) Variation Heat Transfer Coefficient. ha ( w/m2. c) a long Heat Exchanger Depth at Water Flow Rate 2000 (L/h), Water Entering Temp. 10 0C, Air Flow Rate 2000 cfm

97

Figure (5. 6) Comparison between the Experimental and Present Model for the Effect of Air Velocity on Overall Heat Transfer Coefficient (w/m2. c) at Water Flow Rate 2000 (L/h)

98

Figure(5. 7) Variation Overall Heat Transfer Coefficient (w/m2.c) a long Heat Exchanger Height with Water Flow Rate 2000

(L/h), Air Flow Rate 2000 cfm

99

Figure(5. 8) Variation Overall Heat Transfer Coefficient (w/m2.c) a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate 500 cfm

100

Figure(5. 9) Variation Overall Heat Transfer Coefficient ( w/m2. c) a long Heat Exchanger Depth at Water Flow Rate 2000 (L/h),

Water Entering Temp. 10 0C, Air Flow Rate 2000 cfm

101

Figure (5. 10) Comparison between the Experimental and Present Model for the Effect of Air Velocity on Air Exit Temperature (0C) At Water Flow Rate 2000 (L/h)

102

Figure (5. 11) Variation Exit Air Temperature (0C) a long Heat Exchanger Height with Water Flow Rate 2000 (L/h), Air Flow Rate 2000 cfm

103

Figure(5. 12)Variation Exit Air Temperature (0C) a long Heat Exchanger Height atWater Flow Rate 2000 (L/h), Air Flow Rate 500 cfm

104

Figure(5. 13) Variation Air Exit Temperature ( 0C) a long Heat Exchanger Depth at Water Flow Rate 2000 (L/h), Water Entering Temp. 10 0C,2000 cfm

105

Figure (5. 14) The Effect of Aspect Ratio (H/L) with Different Core Size (L × D × H) on the Pressure Drop in Air Side Water Flow Rate 2000 (L/h) ,Air Flow Rate 500 cfm

105

Figure (5.15) The Effect of Turbine Inlet Temperature on the Power Output with Variable Compressor Inlet Temperature at Rang Gas Generator Speed (50000-65000 RPM)

106

Figure (5.16) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed ) on the Fuel Mass Flow Rate

106

Figure (5.17) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed ) on the Specific Fuel Consumption

107

Figure (5.18) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Heat Consumption

107

Figure (5.19) The Effect of Turbine Inlet Temperature (with Variable Compressor Inlet Temperature) on the Heat Rate, at Rang Gas Generator Speed (50000-65000 RPM)

108

Figure (5.20) The Effect of Turbine Inlet Temperature (with Variable Compressor Inlet Temperature) on the Overall Efficiency (%),at Rang Gas Generator Speed (50000-65000 RPM)

108

Figure (5.21) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the on Air Mass Flow Rate

109

Figure (5.22) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the on Air Flow Rate

109

Figure (5.23) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the on Compression Ratio

110

Figure (5.24) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the on Compressor Work

110

Figure (5.25) The Effect of Turbine Inlet Temperature (°C )on the HP.Turbine Work (kW) with variable Compressor Inlet Temperature (°C) At Rang Gas Generator Speed (50000-65000 RPM)

111

Figure (5.26) The Influence Compressor Inlet Temperature (°C )on Gas Turbine Performance at (Gas Generator Speed 45000 RPM)

112

Figure (5.27) The Influence Compressor Inlet Temperature (°C )on Gas Turbine Performance at(Gas Generator Speed 55000 RPM)

112

Figure (5.28) Comparison between the Experimental and Theoretical Predictions of the gas turbine engine (GT-85)

113

List of Tables Title Pages

Table (2-1) GE-Design Parameters 12

Table (2-2) Turbine Inlet Cooling Options 14

Table (3.1) Heat Exchanger Geometry 20

Table (5.1) Air Temperature Distribution Along H.EX.Depth,Vw=2000

(l/h),Va=2000 (cfm) , Tw (in)=10 0 C 83


(l/h),Va=2000 (cfm), Tw (in)=50 0 C 84


(l/h),Va=500 (cfm), Tw (in)=10 0 C 85


(l/h),Va=500 (cfm), Tw (in)=50 0 C 86

Table (5-5) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37

87

Table (5-6) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37

88


89


90


91


91


91


92

Nomenclature

Symbol Description Units A Total Heat Transfer Surface Area m²

Af Fin Area m²

Acw Crosse Sectional Area of Water Side m2

Aca Crosse Sectional Area of Air Side m2

Aexp Exposed Area of the Bare Tube

Cpa Specific Heat Capacity of Dry Air J/kg..k

Cpw Specific Heat Capacity of Water J/kg..k

Cpg Specific Heat Capacity of Gas J/kg..k

dh Hydraulic Diameter mm

Dt Tube Depth mm

Df Fin Depth mm

D Heat Exchanger Depth cm

ha Convection Heat Transfer Coefficient of Air Side W/m².k

hw Convection Heat Transfer Coefficient of Water Side W/m².k

Ht Tube Height cm

H Heat Exchanger Height cm

H.V Heating Value kJ/kg

Lf Fin Length mm

L Heat Exchanger Length cm

am& Air Mass Flow Rate Kg/s

fm& Fuel Mass Flow Rate Kg/s

Nu Nusselt Number khd

= -

Nt Number of Tubes -

Nr Number of Rows -

N Number of Slices -

P Pressure bar

Q& Rate of Heat Transfer Loss or Gain W

Re Reynolds Number ρμ

hud= -

R Gas Cosatant J/kg.K

rC Compression Ratio -

rT Expansion Ratio -

sg Specific Gravity of Fuel -

T Temperature °C −

T Gauge Measurement Temperature °C

tf Fin Thickness mm

tt Tube Thickness mm

Uo Overall Heat Transfer Ccoefficient W/m².˚K .

V Volumetric Flow m3/s

v Velocity m/s .

W Power Watt

XT Transverse Space mm

XL Longitudinal Space mm

xc Isentropic Temperature Ratio for Compressor = 1

2

TT s -

xt Isentropic Temperature Ratio for Turbine =sT

T

4

3 -

Greek Letters

oη ( )ff

AA

η−−= 11 Overall Surface Efficiency

fη Fin Efficiency

ρ Density (kg/m³)

μ Viscosity (kg/m.s)

Ø Maximum to Minimum Temperature Ratio through the gas turbine engine

α t* Ø η c* η

oη Overall Efficiency of Gas Turbine

ƒ Fuel / Air Ratio

tλ Perimeter of Tube Side (m3)

fλ Perimeter of Fin (m3)

γ Specific Heat Ratio

Ŧ Torque (N.m)

Subscripts a Air

C Corrected Value

c Compressor

f Fin

g Gas

i Input

o Output

s Isentropic Process

t Tube

T turbine

w Water

1 Compressor Inlet

2 Compressor Discharge 3 1stTurbine Inlet

4 1st Turbine Outlet

5 2nd Turbine Outlet

Abbreviations NDCW Non-dimensional Compressor Work

NDTW Non-dimensional Turbine Work

c.c Combustion Chamber

ISO International Standard Organization

HVAC Heating, Ventilating and Air Conditioning

CBT Compressor Burner Turbine simple cycle

LMTD Logarithmic Mean Temperature Difference

HP High Pressure

LP Low Pressure

CHAPTER ONE

Introduction

Chapter One…………………………………………………………………….Introduction

1.1 General

Gas turbines are used in a wide range of services; they power aircraft of all

types and drive mechanical equipment such as pumps, compressors, and

generators in electric utilities. They also generate power for peak loads and

base-load duties. Recently, the interest in gas turbines has grown significantly

in combined-cycle plants. These plants use combinations of gas and steam

turbines in various configurations of turbines, heat recovery steam generators,

and regenerators.

Gas turbines have many advantages over steam plants. These are as followed:

1. They are smaller in size, mass, and initial cost per unit output.

2. Their delivery time is relatively short and they can be installed quickly.

3. Their starting is quicker (as low as 10 s) Philip (2002) [1], often by

remote control.

4. Their running are smooth and have a capacity factor (percent of time

the unit is operating at full power) of 96 to 98 percent.

5. They can be used in a wide variety of liquid and gaseous fuels

including gasified coal and synthetic fuels.

6. They can be subjected to fewer environmental restrictions other than

prime movers.

1.2 Two Shaft Engine

A two shaft gas turbine Figure (1.1) consists of an air compressor, a

combustor, a gas generator turbine, and a power turbine. The air compressor

generates air at a high pressure, which is fed into to the combustor where the

fuel is burned. The combustion products and excess air leave the combustor at

high pressure and high temperature. This gas is expanded in the gas generator

turbine, which has the sole task of providing power to turn the air compressor.

After leaving the gas generator turbine, the gas still has a high pressure and a

high temperature. It is now further expanded in the power turbine. The power


turbine is connected to the driven equipment. It must be noted at this point

that the power turbine (together with the driven equipment) can run at a speed

which is independent of the speed of the gas generator portion of the gas

turbine. The gas generator is controlled by the amount of fuel which was

supplied to the combustor. Its two operating constraints: the firing

temperature and the maximum gas generator speed. If the fuel flow is

increased, both firing temperature and gas generator speed will also increased

until one of the two operating limits is reached.

Exhaust

Gas Generator

Inlet Air

Figure 1.1 Two Shaft Gas Turbine Engine

1.3 Influence of External Factors on Gas Turbine Performance

A gas turbine uses ambient air; therefore, its performance is greatly

affected by all factors that influence the flow rate of air delivered to the

compressor, in terms of weight and its physical conditions. These factors are:

1. Ambient Temperature

2. Ambient Pressure

3. Relative humidity

In this regard, the reference conditions for the three above-mentioned factors

are (15 °C, 1013 mbar, and 60 %) respectively, ISO (1973) [2].When the

compressor inlet temperature increases, the specific work needed to compress

Combustion Chamber Pow

er Turbine

Turbine C

ompressor

Driven Equipment


air will also be increased. However, the weight of the air delivered will be

decreased (because of a decrease in specific weight). Consequently, the

turbine efficiency and useful work (and, therefore, power) diminish as well. If

compressor inlet temperature decreases, the reverse process occurs. This

temperature depends on the air aspirated by the compressor. The power and

efficiency varies from turbine to turbine, according to cycle parameters,

compression and expansion output and air delivery rate…etc. And as a result,

the variation ratio of gas turbine performance parameters is taken proportional

to design point (manufactured levels). Figure (1.2) shows an example of how

power, heat consumption, heat rate and the delivery rate of exhaust gases

depend on ambient temperature. Design point performance is a central to the

engine concept design process. The engine configuration, cycle parameters,

component performance levels and sizes are selected to meet the given

specification.

Figure 1.2 Influences of External Factors on the Gas Turbine

Performance, Frank (2002) [3]


1.4 Gas Turbine Inlet Air Cooling

The turbine inlet air cooling methods can be divided into two categories: 1.4.1 Evaporative Cooling

In this process water is distributed over pads of fibers through which the

passing air should be humidified. Spray intercoolers or fogging systems were

also used to cool the inlet air. When the power and efficiency can be

increased by decreasing compressor inlet temperature ,the latter can be

reduced artificially by using an evaporative cooler located upstream of the

suction filter. Water, fractioned into drops or in the form of a liquid film,

cools the air by evaporating in the cooler as it flows in contrary direction.

Figure 1.3 Evaporative Cooler

1.4.2 Cooling with Absorption Chiller

The absorption chiller works on the principle of vapor absorption

refrigeration cycle. The main advantage of this chiller lies in the fact that the

inlet air can be cooled down to a specific temperature for a wide range of

ambient air temperatures. Thus, the power output of a gas turbine remains

more or less constant, independent of ambient air conditions.


The low grade exhaust energy can be used to drive the chiller. The chilled

water produced by the absorption system, is passed through the inlet air

cooler, which is an indirect type air to water heat exchanger.

Figure 1.4 Air "Chilling Cooling" System, Based on

Absorption, GE [4]

1.5 Compact Heat Exchanger

Compact heat exchangers have been widely used in various applications in

thermal fluid systems including automotive thermal management systems.

Radiators for engine cooling systems, evaporators and condensers for HVAC

systems, oil coolers, and intercoolers are typical examples of the compact heat

exchangers which can be found in ground vehicles. Among the different types

of heat exchangers for engine cooling applications, cross flow compact heat

exchangers with plain fins are of a special interest. This is because of their

higher heat performance capability with the lower flow resistance.


1.6 Aims of the present Work

The drop in overall performance of gas turbine engines is believed to be due

to the increasing of compressor inlet temperature above ISO condition

(15°C).The present work is an attempt to improve the gas turbine

performance and bringing it near to ISO condition. This is accomplished by

applying sensible cooling technique consisting of air cooled heat exchanger

and water supplying system. In the present work a theoretical and

experimental study has been developed for both, air cooled heat exchanger

and gas turbine engine. A predictive numerical model for the air cooled heat

exchanger has been developed which is based on step by step technique

method so that a design tool for the heat exchanger can be developed.

Such work requires developing a versatile experimental facility to examine

the air cooling effect on gas turbine performance. This aim can not be

achieved without some important requirements which can be summarized as

follows:

1- Studying the performance of an existing gas turbine system available in the

laboratory. This was achieved by controlling the inlet air temperature to the

compressor.

2- An experimental rig was built up to conditioned air at different

temperatures controlled by the water supply temperature and then studying

this effect on the gas turbine performance.

3- Making theoretical assessments for the heat exchanger performance and the

effect of compressor inlet temperature on the gas turbine performance.

4-Developing a computational program for simulate the heat exchanger

design and estimating its performance with variable inlet conditions. This is

done by feeding with the experimental data to validate this simulation. A

computational program to predict the effect of air inlet temperature on the gas

turbine performance was built for this object depending on the engine design

information only.

CHAPTER TWO

Literature Survey

Chapter Two……………………………………………………………….Literature Survey

2.1 General

The gas turbines are generally used for large scale power applications. The

basic gas turbine cycle has low thermal efficiency. So it is important to look

for improved gas turbine based cycles. The inlet air cooling helps to increase

the performance of gas turbines. The demand of energy in the developing

regions of the world, particularly in Asia, has witnessed pronounced increase

in the recent past. According to a report of International Energy Outlook

(2004)[5], the world net power consumption is expected to be doubles nearly

over the next two decades. Much of the growth in new electricity demand is

expected to come from countries of the developing world. Therefore, it is

important to find improved technologies for power generations that have a

high efficiency and specific power output, low emissions of pollutants, low

investment, and low operating and maintenance cost for a sustainable use of

available fuels. 2.2 Pervious Work

The previous work can be classified into two categories: The first one is

concerned with modeling and simulation of air cooled heat exchanger. The

other related with the performance improvement of the gas turbine engine.

The improvement of two-shaft gas turbine performance by making

modifications or addition of some parts to the main components. This will

lead to decrease in the compressor inlet temperature and enhance the power

output and thermal efficiency of gas turbine.

2.2.1 Modeling and Simulation of Air Cooled Heat Exchanger

Ganapathy (1979) [6] concluded that for air- cooled condensers, the ambient

air is the most important variable in the design. Since ambient temperature in

a location varies throughout the year. Using higher value, would result in over

sizing the unit. Where as A lower value would give poor performance.

Current practice is to use a design temperature that exceeds (2 to 5%) of the

annual period.


Hedderich and Kellehere (1982)[7] developed a computer code for the

analysis of air cooled heat exchangers and was coupled with a numerical

optimization program to produce an automated air cooled heat exchanger

design leading to optimization procedure.

A general iteration free approximation method was used for the analysis

which calculates the mean overall heat transfer coefficient and the overall

pressure drop for many arrangements.

The analysis takes into account the variation of heat transfer coefficients

and the pressure drop with temperature and length of flow path .The

capability is demonstrated by the design of an air cooled finned tube heat

exchanger and is shown to be useful tool for the heat exchanger design.

Zhang (1994) [8] proposed three dimension numerical model predicting the

performance for large power plant condensers. He compared his predicted

results with the experimental data. The prediction was achieved by solving the

governing mass, momentum and air concentration using semi implicit

consistent control-volume for simulation model with different conditions in

work of condenser.

Matthew and Joseph (2002) [9] developed a conceptual designs for wet and

dry cooling systems applied to a new ,gas –fired, combined cycle 500-MW

plant at four sites chosen to represent conditions in California. The

requirements for cooling dry systems are four to six times those for wet

systems. Dry systems, which are limited by the ambient dry bulb temperature,

cannot be achieved as low a turbine back pressure as wet systems, which are

limited by the ambient wet bulb temperature.

Dohoy and Dennis (2006)[10] divided the heat exchanger core into small

control volumes along the tube. Finite Difference Method (FDM) with

staggered grid system was utilized in study. FDM can take into account the

significant air temperature increase as well as the local variations of the

properties and the heat transfer coefficient. The maximum difference between


the experimental data and calculated results was noticed to be 5% for the

given range of the simulated conditions.

Tarrad, et al. (2008) [11] investigated the performance prediction of the

cross flow air-cooled heat exchanger. They developed a new simplified

correlation for the air side heat transfer coefficient which depends on the

dimensional analysis with Buckingham-pi theorem. The discrepancy between

the predicted and their own experimental values of the overall heat transfer

coefficient and heat duty were within 2% and 4% respectively for both of the

tested tube banks.

2.2.2 Performance Improvement of the Gas Turbine Engine

There are two basic methods available for inlet air-cooling evaporative

cooling and chilling cooling. The most widely accepted system is evaporative

air-cooling. Evaporative coolers make use of the evaporation of water, and are

the most cost-effective way to improve machine capacity during warm

weather. Mostly percent of design concept is used to examine gas turbine

performance. Of the two cooling methods of inlet air, namely, evaporative

cooling and the absorption cooling, the absorption cooling technique

demonstrated a higher gain in power output and efficiency than evaporative

cooling for a simple cycle gas turbine.

De Lucia, et al.(1995) [12] reported that evaporative inlet-cooling is

economical and simple, but only suitable for dry hot climates.

they concluded that evaporative inlet cooling could enhance power out put by

(2–4)% depending on the weather.

Saleh (1996)[13] presented that water can be injected in a simple two shaft

gas turbine (GT-85) to improve the performance. The studied cases were

water injection prior to the combustion chamber and water injection in intake

of the compressor. Maximum increasing in performance data was obtained

when water injection prior to the combustion chamber where, the increase of


the power output up to (40.4٪) and thermal efficiency up to (34.8٪) .In

addition, specific fuel consumption was reduced by (32.12 ٪ ).

Ait-Ali (2001) [14] presented the concept of inlet air refrigeration to boost

the power output from the gas turbine. Chillers can increase the gas turbine

power output by 15-20%.The absorption chiller works on the principle of

vapor absorption refrigeration cycle. The main advantage of this chiller lies in

the fact that the inlet air can be cooled down to a specific temperature for a

wide range of ambient air temperatures. Therefore, the power output of a gas

turbine remains more or less constant, independent of ambient air conditions.

A typical absorption chiller with a capacity of 3000 refrigeration tons and a

COP of 0.70. This absorption system uses the waste heat to produce required

steam by the chiller.

Nuovo Pignone (2002)[15] publishing curves of the compressor inlet

temperature effect on gas turbine performance as shown in fig.(2.1).It is

obvious that power output and heat rate are improved as compressor inlet

temperature was decreased. Lowering the compressor inlet temperature can be

accomplished by installing an evaporative cooler or inlet chiller in the inlet

ducting downstream of the inlet filters. General electric estimates theses

performance curves with respect to the design point and maximum speed as

shown below in table (2-1).

Table (2-1) GE-Design Parameters


Figure 2.1 Effect of Compressor Inlet Temperature on Gas Turbine Performance [15]

Donald and Icksoo (2003)[16] presented the various types of turbine inlet

cooling applicable to small to mid-size turbines .These have been described

along with their comparative benefits. The greatest benefit was shown to be

obtainable from an exhaust heat-powered ammonia absorption cycle. An

ammonia absorption cycle was especially designed for this application. A

300-refrigeration ton aqua ammonia refrigeration unit is required to cool the

inlet of a (5 MW) gas turbine from (35°C to 5°C). This cooling will increase


the power output by 1 MW. The added power was at a marginal efficiency of

39%, compared to 29% for the base turbine power. The cooling option is

listed in table (2-2). Table (2-2) Turbine Inlet Cooling Options

4.7 MWe Simple Brayton Cycle - 30% efficiency at ISO

Hameed (2004)[17] concluded that water injection in the air intake is

strongly effecting the performance parameters of the two-shaft gas turbine

cycle (GT-85) . the power output has been increased up to (23.15%) for

simple cycle and the thermal efficiency is higher than that of normal cycle by

(29%).

Benjalool (2006) [18] concluded that, in September, the range of ambient

temperature in the Nafoora oil field varies typically between (29-36 °C). The

temperature variation leads to change the maximum engine power output

from (5.1 MW) to (4.85 MW).

Tony Giampaolo (2006) [19] concluded that at a constant gas generator

speed, and ambient temperature decreases, turbine inlet temperature will be

decreased slightly, and power output will be increased significantly, fig.

(2.2).This increase in gas horsepower results from the increase in compressor

pressure ratio and aerodynamic loading. Therefore, the control must protect


the gas turbine on cold days from overloading the compressor airfoils and

over-pressurizing the compressor cases. Sensing ambient inlet temperature

helps insure that engine internal pressures are not exceeded, and sensing

turbine inlet temperature insures that the maximum allowable turbine

temperatures are not exceeded. Sensing gas generator speed enables the

control to accelerate through any critical speed points (gas turbines are

typically flexible shaft machines and, therefore, have a low critical speed).

Figure 2.2 Temperature-Power-Speed Interrelationships

Kuamit (2006) [20] concluded that the effect of compressor inlet temperature

has an important role on the power output as shown in fig.(2.3). It may be

seen that the power out put is influenced by compressor inlet temperature due

to the change of air density and compressor work .Since a lower compressor

inlet temperature leads to a higher air density and a lower compressor work

that in turn gives a higher gas turbine output.


270 290 310 330260 280 300 320 340

Inlet Temp, (K)

110

130

150

100

120

140

160

Out

put P

ower

, (MW

)

Figure 2.3 Net Output Power Versus Inlet Temperature for Gas Cycle 2.3 Summary and Motivation

From the review of previous works it is obvious that industrial gas turbines

performance is different from turbine to turbine. It depends on the type and

what full performance at ISO conditions .Also, what equipment that used to

improve the performance. Clearly, the performance is not only affected by

compressor inlet temperature, but also by other parameter such as relative

humidity, inlet pressure, maximum temperature in the cycle, and speed of the

shaft. However, operating at constant speed has constant volumetric flow rate.

Since the specific volume of air is directly proportional to temperature, cooler

air has a higher mass flow rate. It generates more power in the turbine. Cooler

air also requires less energy to be compressed to the same pressure as warmer

air. Thus, gas turbines generate higher power output when incoming air is

cooler. A gas turbine inlet air cooling system is a good option for applications

where electricity prices increase during the warm months. The power output

increases by decreasing the compressor inlet temperature of the incoming air.

The aim of this work is to study the approach used in the enhancement of

gas turbine performance by making the compressor inlet temperature supplied

to the gas turbine closed to ISO conditions. This was achieved by building a


cooling system for entering air to the compressor consisting of air cooled heat

exchanger and water supply system. In this regard, a theoretical and

experimental study for both air cooled heat exchanger and gas turbine engine

will be conducted.

A supporting computer program to simulate a new technique of heat

exchanger design has been developed. In this model, heat exchanger was

described in two directions, the height and depth to form horizontal slices

which will described later. A computer program prepared for this purpose has

the ability to analyze heat exchanger performance for any slice to be located

in two heat exchanger dimension. The heat exchanger type used was a type of

air cooled heat exchanger (finned-tube surfaces, flat tubes, continuous fins). It

is part of cooling system accommodating different components like valves,

pipe fittings, supply pump, and two reservoirs (hot reservoir and cold

reservoir).

CHAPTER THREE

EXPERIMENTALWORK

Chapter Three…………………………………………………………Experimental Work

3.1 Test Rig for Heat Exchanger

The test rig layout was built as shown in Figure (3.1). It consists of inlet air

cooling system having air cooled heat exchanger, tank, and water circulation

system, pump and control panel. The water circulation system is a

modification of an existing laboratory system, which had already built by

Tarrad and Mohammed (2006) [21].

3.1.1 Heat Exchanger

The test section is made of a compact heat exchanger which is of 19

80Chevette radiator type using, Cross-flow exchanger with one flow mixed.

Edges of flat vertical tubes, having dimensions of (55 cm) length,(3.5 cm)

depth and (37 cm) height, heat exchanger geometry is illustrated in table

(3.1). The Compact heat exchanger configuration is shown in Fig. (3.2). The

water flows in the tubes in cross direction to the air flowing normal to the

tubes. Thermometers are connected to the heat exchanger. The gauges are

fixed in the specially prepared pockets mounted on the required locations.

3.1.2 Water Supplying System

The cold water is supplied by a constant tank head of (200 liters) capacity.

The water is pumped by a single stage centrifugal pump from the tank

through the test section, and then it returns back to the tank. The water in the

hot tank is heated by four electrical heaters, which have a total heating

electrical power of (12 kW). The four heaters are fixed in the same level at

150 mm from the bottom of the hot tank separated at (90◦) apart. The hot tank

is opened in cold tank by gating valve which be found in the pipe that linked

the two tanks, as shown in Fig. (3.1). This arrangement of cold and hot tanks

enable as to control manually the temperature in the tank .It is will help in

providing a good control of the water temperature during tests. The mixing of

water has been accomplished in the cold tank to obtain the required


temperature. A piece of ice was immersed into cold tank so that the mixture

becomes homogenous. Table (3.1) Heat Exchanger Geometry

Parameter Dimension

Core Length (L) 550 mm

Depth (D) 35 mm

Height (H) 370 mm

No.of Tubes 110

No.of tube/row 55

No.of fins/ tube 256

Normal Distance (XT) 9.92 mm Longitudinal Distance (XL) 20.46 mm

Fin Pitch (Pf) 1.46 mm

Length (Lf) 7.52 mm

Depth (Df) 15.88 mm

Thick (Tf) 0.24 mm

Tube Height (Ht) 2.4 mm

Depth (Dt) 15.88 mm

Thick (Tt) 0.28 mm

A special transparent glass tube level is fixed on the outer shell of the tanks

in order to monitor the water level. Temperature gauge (thermometer) is

connected to the shell of tank to monitor the water temperature in the tank

during tests. Both hot and cold water flow is controlled by gate valves and the

flow is measured by using a vertical variable area rotameters. Piping system

was made of carbon steel metal, insulated by glass wool to minimize the heat

loss.


3.1.3 Measuring Instrumentation

The parameters which are to be measured during the test as follow:

(1) The inlet and outlet temperatures of the heat exchanger.

(2) The air flow rate across the heat exchanger.

(3) The water flow rate of the tube side.

Thermometers:

Thermocouples having a range of (0 ◦C – 120 ◦C), are used to measure the

temperatures at inlet and outlet of heat exchanger. The accuracy rang is about

(0.02).

Pressure Gauges:

Two pointer pressure gauges installed on both sides of heat exchanger to

measure the pressure of the water and have a range of (0 – 2.5 bar), which are

connected to the heat exchanger the accuracy rang about (0.03).

Rotameters :

Two vertical variable area rotameters are used to measure the flow rates of

the water. A rotameter of (200 – 3000) L/h range is used the accuracy rang

about (0.02).

Tube:Pitot

Pitot tube is located at end of air flow duct is used to measure air flow rates

by calculating pressure difference for two points.

3.1.4 The Electrical Board

The electrical board contains the main circuit breaker and other secondary

switches, supplying power to the whole system components of the test rig.

The board consists of an electrical contactor of (4×16 Amp), which is

connected to a thermostat and four heaters. This contactor is controlled and

receives the electrical signal from the thermostat switching on and off the

electricity to the four heaters. There are four separate switches each of which

controls one heater. In addition for safety, there is a switch controling the


operation of the contactor (i.e., controls the operation of the four heaters and

the thermometer), Tarrad and Mohmmed [21].

3.1.5 Air Circulation System

The air was supplied to the test heat exchanger through a fan. A forced

draught arrangement was selected for the test object by variable fan speed.

Three volumetric flow rate was prepared of capacity of (2000) cfm ,(1000

cfm) and (500 cfm) .The fan was close enough to the test section avoiding

leakage of air to the surrounding. The air volumetric flow rate is measured by

using (Pitot-Tube) .The pressure drop was calculated for two points through

air flow duct.


Figure (3.1a) Schematic Diagram of the Built Rig (Heat Exchanger)

Valve

Valve

Cold Tank

Ice

Pump

By Pass

Valve

Rot

amet

er R

otam

eter

Hot Tank

Heaters

Fan

Heat Exchanger

Duct T

PP T


Fan

Cold Tank

Figure (3.1b) Configuration of the Built Rig (Heat Exchanger)

Pitot-Tube

Rotameter

Heat Exchanger Gas

Turbine Hot

Tank

Figure (3.1c) Configuration of the Test Rig (Inlet Cooling System) Preparing to Gas Turbine


Figuer (3.2.a) Top View of Heat Exchanger Geometry

Figure (3.2.b) Front View of Heat Exchanger Geometry

H

L

D

Pf

Lf

XL Dt XT Ht


3.2 Experimental Setup

A schematic diagram of the experimental setup is shown in Figure (3.3).

The test loop was designed to allow easy control parameters such as upstream

air flow rate to heat exchanger, and compressor inlet temperature to gas

turbine. As mentioned in the previous section, the loop consists of the heat

exchanger set prior to gas turbine apposite inlet box. To achieve this it is

required building up the main experimental rig with measuring

instrumentations obtaining the suitable circumstances regarding air dry bulb

temperatures, and flow rate. The air leaving the heat exchanger is fed directly

to the compressor intake of the gas turbine engine.

3.3 Gas Turbine

The chosen test engine for this present work is a two-shaft gas turbine type

(GT-85) of 5 kW two shaft machine for industrial drive (Pumps,

Compressors).The design strategy of a gas turbine unit is divided into two

fields. These are; design information and performance information. Design

information such as the compressor whether it is a centrifugal or axial flow,

pressure ratio, the number of fuel nozzle in combustion chamber, a single

shaft or two shaft. The performance information such as output power, heat

rate, exhaust flow, exhaust temperatures. GT-85 performance parameter

measurements have been obtained over variable gas generator speed. The

speed variation is accomplished by manual fuel flow control valve. The

measuring parameters obtained for a particular gas generator speed (N1) are,

Vf,P2,T1,T2,T3,P3,T4,P4,T5,N2.


Dynomometer

Duct

Fig

ure

(3.3

) Sch

emat

ic D

iagr

am o

f the

Tes

t Rig

(Ove

rall)

Col

dTa

nk

Am

bien

t

Air

Hot

Tan

k

C

old

Tan

k

H

ot

Tan

k

Pum

p

Heat Exchanger

Compressor

HP.Turbine

LP.Turbine C

ombu

stor

Rot

amet

ers

Val

ve

Val

ve


3.3.1 Design Information

Compressor:

• Centrifugal flow

• Pressure ratio , 2:1

Combustion:

• Annular combustion, single chamber

• Liquid fuel capability

Turbine:

• High Pressure turbine 1 stage , 1.69:1 expansion ratio • Low Pressure turbine 1 stage 1.18:1 expansion ratio

Package: • Gas Generator, Power Turbine and auxiliary system mounted on a

single base plate

• Control system (instrumentation, sensors ,electronic panel,

mechanical regulation) to monitor temperature and pressure

3.3. 2 Performance design

• Power turbine output: 5 KW

• Maximum cycle temperature: 750 (°C)

• Maximum Gas generator speed : 90000 RPM

• Maximum Power turbine speed: 35000 RPM

The performance power, fuel consumption, temperatures, shaft speeds... etc.

of a gas turbine engine is crucially dependent upon its inlet and exit

conditions. The environmental envelop is impartment item for any gas turbine

plant performance. 3.4 Operating Principles

A gas turbine works in the following way:

• It aspirates air from the surrounding environment

• It compresses it to a higher pressure

• It increases the energy level of compressed air by the addition of fuel gas


which undergoes combustion in a combustion chamber

• It directs high pressure and high temperature air to a turbine section, which

converts thermal energy into mechanical energy allowing shaft to revolve.

This serves on one hand, to supply useful energy to the driven machine,

coupled to the machine by means of a coupling and, on the other hand, to

supply energy necessary for air compression, which takes place in a

compressor connected directly with the turbine section itself.

• The remaining energy is supplying through power turbine, finally • Gas turbine sensors are designed to withstand the vibration and high

temperatures found in these engines ,Watlow (2001) [22]. 3.5 Main Component Parts of Gas Turbine

3.5.1 Gas Generator:

The gas generator consists of a centrifugal compressor; a combustion

chamber and a radial flow turbine as shown in Fig. (3.3).Air enters the

centrifugal compressor through an intake silencer and a bell mouth air flow-

meter. The outlet of the compressor is directed towards a vertically oriented

combustion chamber which in turn is connected to the compressor turbine.

The compressor and turbine run at a maximum speed of 90000 RPM. Such a

compressor speed can achieve a compression ratio of 2:1 for this particular

compressor design. The speed of the compressor and turbine is measured by

means of an opto-electronic technique.

3.5.1.1 Centrifugal Compressor :

The centrifugal compressor draws in air at the center or the eye of the

impeller and accelerates it around and outward. It consists of an impeller, a

diffuser and compressor manifold. The diffuser is bolted to the manifold, and

often the entire assembly is referred to as the diffuser. The impeller may be

either single entry or dual entry. The principal differences between a single

entry and the dual are the size of the impeller and the ducting arrangement.

The single entry impeller permits ducting directly to the inducer vanes, as


opposed to the more complicated ducting needed to reach the rear side of the

dual-entry type .Although it is slightly more efficient in receiving air, the

single-entry impellers must be of a greater diameter to provide sufficient air.

The compressor draws in air at the hub of the impeller and accelerates it

radially outward by centrifugal force through the impeller. It leaves the

impeller at high speed and low pressure flowing through the diffuser.

The diffuser converts the high speed, low–pressure air to low-speed, high-

pressure air. The compressor manifold diverts the low-speed, high-pressure

air from the diffuser into the combustion chamber. In this design, the

manifold has one outlet port for each combustion chamber. The outlet ports

are bolted to an outlet elbow on the manifold. The outlet ports ensure that the

same amount of air is delivered to each combustion chamber. The outlet

elbows change the airflow from radial to axial flow. The diffusion process is

completed after the turn .Each elbow contains from two to four turning vanes

that perform the turning process and reduce air pressure losses by providing a

smooth turning surface.

3.5.1.2 Combustion Chamber:

In a gas turbine, the hydrocarbon fuel is burnt in the combustion chamber

as a continuous process at essentially constant pressure. The combustion

system of the (GT- 85) consists mainly of two components: the swirl unit

incorporating a spray nozzle and the combustion chamber. The swirl unit

produces a swirling air flow which mixes with the finely atomized fuel

sprayed from the nozzle. The air/fuel mixture issuing from the swirl generator

is fed into the combustion chamber as an expanding swirling flow. Schematic

drawing Fig.(3.4), shows the flow issuing from swirl generator into the

combustion chamber and the subsequent flow patterns. As it will be seen from

the figure, the flow initially on entering the combustion chamber should form

a toroidar vortex. This vortex being formed as a result of the expanding swirl

flow and the air entry through the primary holes. Combustion takes place here


under an approximately stoichiometric air/fuel ratio, i.e., A/F =15:1. The

second row of holes, defined as the secondary zone admits more air to cool

the combustion products and gives an overall A/F at this point of about 35:1

The dilution zone contains larger holes than any of the preceding sections and

admits a larger quantity of air producing an exit gas flow with an overall a/f

ratio of approximately 70:1 .For all sections in the combustion chamber,

series of small holes is provided around the circumference of the combustion

chamber. These holes provide a cooling flow for the walls and maintain the

wall temperature within the safe operating temperature.

3.5.1.3 Gas Generator Turbine:

A radial flow turbine consists essentially of a stationary casing containing

a rotating impeller which is rotated as a result of the high velocity flow

leaving the stationary nozzles. The function of the casing is to accelerate the

flow smoothly producing uniform high velocity to the impeller tip. The

turbine impeller is of radial vane type. Both the compressor and generator

turbine run at the maximum speed of 90000 RPM. 3.5.2 Power Turbine:

The hot gas issuing from the exhaust of the gas generator turbine is passed

through a flexible circular duct to the inlet of the power turbine. The power

turbine is larger than the gas generator turbine and is of inward radial flow

design. The turbine operates at a maximum speed of 35000 RPM .The

maximum power output is (5 kW) with an inlet temperature of 700C and a

speed of 35000 RPM .The power turbine is directly coupled to the eddy-

current brake by means of a special high speed coupling .The speed of the

power turbine is measured by electromagnetic sensor and a digital frequency

meter.


Fig.(3.4) Schematic Arrangement for Gas Turbine (GT-85)

3.5.3 Fuel System:

The gas turbine utilizes kerosene as the operating fuel system. During the

starting sequence, methylated alcohol (starting fuel) is ignited by the ignition

system to provide a large flame in front of for the initiation of the kerosene

combustion process. The fuel is pumped by an electrically driven gear pump

at a pressure of 6 bars. The fuel flow is regulated by a manuall valve located

on the control panel. Fuel flow rate is measured by a ‘Rotometer’. For safety

operating process, a solenoid valve is incorporated in the fuel feed piping

system to the combustion chamber. This will automatically stop the supply of

fuel in the event of malfunction. 3.5.4 Oil Lubricating System:

As mentioned before, the gas generator turbine and the power turbine run

at very high speed. Therefore, a lubricating system, Fig.(3.4) is incorporated

to ensure a safe running to turbines . Oil pressure of minimum 3 bars should


be provided to lubricate the journal bearings by the gear type oil pump. Oil

temperature and pressure are automatically monitored and the gas turbine will

be shutdown in case of a non-safe running. 3.5.5 Starting System:

In order to start the gas turbine, it is required to drive the compressor to a

certain speed to achieve the required air flow rate to the combustion chamber,

so that the starting fuel (methylated alcohol) can be ignited. This is done by

blowing air through the compressor inlet duct using three electrical fans.

The flow rate of air supplied by these fans is regulated by a gate valve. High

energy spark ignition system is used to ignite the starting fuel. The sustained

flame generated by starting fuel burning will raise the flow gases temperature

to the required temperature to ignite the kerosene fuel. Consequently, self

sustained gas generator running is obtained and the air blowing fans will be

switched off Gilbert and Gordon LTD (1978) [23].

3.6 Dimensionless and Parameter Groups

The importance of dimensionless, referred and scaling parameter groups to

all aspects of gas turbine performance cannot be over emphasized.

Understanding and remembering the form of the parameter group

relationships allows judgments concerning the performance effects of

changing ambient conditions, scaling an engine, a change of working fluid,

see Appendix C (Engine Parameter Groups)

The parameter group for mass flow is then a function of , Philip (2004) [24]:

1. Ambient temperature

2. Ambient pressure

3. Engine rotational speed

4. Engine diameter (scale factor)

5. Gas constant of working fluid

7. Gamma for working fluid

8. Viscosity of working fluid


in order to rationalize the performance data obtained from the gas turbine

engine operating with a wide range of inlet conditions, it is necessary to

reduce the data to known standard intake conditions. The correction formula

used to achieve this rationalization can be derived by a non-dimensional

analysis of the components in the gas turbine cycle. Non-dimensional analysis

leads to various dimensionless parameters which are based on the dimension's

mass (M) , length (L) , and time (T) . Based on these elements, one can obtain

various independent parameters. These parameters will lead to form various

dimensionless groups.

By using non-dimensional groups as applied to the following basic equation

for compressor non-dimensional analysis, Cohen,et al. (1996) [25]:

f ( Nc, ma , P1 , P2 , R× T1, R× T2 ) =constant ……………….…….…..(3-1)

3.6.1 Corrected Compressor Data

From the equations are illustrated in Appendix (C), and that referring to

equation (3-1) for compressor non-dimensional analysis, and since :-

1TNc

TsNcc

= ……………………………………………………………...…(3-2)

s

sac

a

aa

PTM

PTM

..

= ……………………………………………………..…(3-3)

TsT

TaT c22 = …………………………………………………………………(3-4)

where:

Ts= 15 (°C) =288.16 (k)

Ps= 1.0133 bar

For other values of Ta different from the standard value a correction should be

made as follows:


accc T

NN 16.288= ………………………………………………….….….(3-5)

a

aaac P

TMM

..

16.2880133.1

×= ……………………………………………..…..(3-6)

2216.288 T

TT

ac ⎟⎟

⎠

⎞⎜⎜⎝

⎛= ………………………………………………………….(3-7)

3.6.2 Corrected Basic Data

Corrected basic data for the effects of variation in both, ambient pressure

and temperature should be done. In the case of the actual test data, the

pressure measurements taken are all static values, whilst the temperature

measurements will be essentially total values. For the formula below, no

differentiation between total and static is made because of the small

differences will be small between the total and the static value of pressure or

temperature. The following formulas are the corrected basic data, which are

suggested by the manufacturer.

1. K 16.273)( +=−

CTaTa o

2. 1000

)(Re mbaradingBarometerPa = bar

3. )16.288)(16.273( 11a

C TTT +=−

K

4. bar )(109.97 61 PPP a ΔΧ−= −

5. bar PaPP c +=−

22

6. )16.288)(16.273( 22 TaTT c +=−

K

7. bar PaPP C +=−

33

8. )16.288](16.273[ 33a

c TTT +=−

K

9. bar PaPP c +=−

44


10 )16.288)(16.273( 44 TaTT c +=−

K

11. )16.288)(16.273( 55 TaTT c +=−

K

12. )(001333.0 5445 PPP cc Δ−= bar

3.6.3 Corrected Derived Data

The measured variables during the tests on the (GT-85) are presented in

tables shown in Appendix (A).When the temperature and pressure at each

point around the cycle, the following formulas may be used for the

performance calculation in the experiments.

The following represents the data reduction method applied in the present

work.

1. Compression Ratio ( ) cr

c

cc P

Pr

1

2= …………………………………………………………………..(3-8)

2. Compressor Isentropic Efficiency ( cη )

00)1(

12

1 100)1( ×−−

= − γγη ccc

cc r

TTT ……………………………….………….(3-9)

3. Air Mass Flow Rate (kg/s)

)16.288

(

3005.011

1

Taamam

PP

PT

am

c &&

&

=

Δ=

............................................................................(3-10) where : in mm wg PΔ

1T in K


4. Fuel Flow Rate (kg/s)

Tav

m fffc

16.2883600

10 3−×= ρ& …………………………………..……….…(3-11)

where: sgwf ∗= ρρ The specific gravity of the fuel (sg) is 0.774 5. Air/Fuel Ratio (A/F) A/F = fcac mm && ……………………………………………………………(3-12) 6. Combustion Chamber Temperature Rise (K)

6.60078.11043740 2

+−

=Δ

−

FAT

T ccct +2 ………………………………………..…….(3-13)

Where: 16.27322 −=−

cC TT 7. Combustion Chamber Pressure Loss (٪)

100)(2

32 ×−

=Δc

cccc P

PPP ……………………………………………….…..(3-14)

8. Gas Generator Turbine Expansion Ratio

c

ct P

Pr

4

31 = …………………………………………………………………(3-15)

9. Theoretical Power Input to Compressor ( ) w&

)( 12 ccC TTaCpamW −= && ……………………………………………………(3-16) Where: 1000=acp kgKJ


10. Theoretical Power Output Of Gas Generator Turbine ( ) w&

)()( 431 ccgfcact TTCpmmW −+= &&& ……………………………………………(3-17)

where: 1150=gcp kgKJ

11. Compressor Mechanical Efficiency ( mcη )

100)2

(1

×′+

=tc

cmc ww

w&

&η ………………………………………………………(3-18)

12. Gas Generator Turbine Mechanical Efficiency ( 1mtη )

100)2

(1

11 ×

+=

t

tCmt W

WW&

&&η …………………………………………….….……(3-19)

13. Compressor Overall Efficiency ( ocη )

ocη = 100

mcc

ηη × ……………………………………………………..………(3-20)

14. Power Turbine Expansion Ratio

c

c

PP

r5

412 = …………………………………………………………………(3-21)

15. Power Turbine Isentropic Efficiency ( 12η )

12η = 100])1(1[ )1(

124

54 ×−

−− γγ

rT

TT

C

CC

=

………………………………………………(3-22)

where: γ 1.33 16. Power Turbine Power Output ( ) w&

16.288602 2

2TaN

W ctτπ

=& ……………………………………………………(3-23)

Where: N 2 power turbine speed (RPM) and τ is the torque in (N.m)


17. Power Turbine Theoretical Power Output ( ) w&

)()( 542 ccgfcact TTCpmmW −+= &&& ………………………………….………….(3-24) 18. Power Turbine Mechanical Efficiency ( 2mtη )

2mtη = 100)(2

2 ×t

ct

WW&

&…………………………………………………………..(3-25)

19. Power Turbine Overall Efficiency ( 2otη )

2otη = 100

22

mtt

ηη × …………………………………………………………….(3-26)

20. Overall Thermal Efficiency ( thη )

thη = 100).

( 2 ×× fc

ct

mVHW

&

& …………………………………………………..…(3-27)

Where: H.V= L.C.V 4.1868 × kgkJ for kerosene used in the tests (GT-85), and

L.C.V= 10300 kgKcal the lower calorific value of the fuel used, Gilbert and

Gordon LTD [23].

21. Specific Fuel Consumption ( sKwkg . )

1000)(2

×=ct

fc

Wm

sfc&

& ……………………………………………………..…..(3-28)

22. Heat Consumption (kW)

Heat Consumption = fuel mass flow * heating value

= mf (kg) ×H.V (kW/kg)……………………………...(3-29)

23. Heat Rate

Heat Rate = specific fuel consumption × heating value

= sfc (kg/kW) H.V (kW/kg)………………………………….(3-30) ×


3.7 Experimental Work Procedure:

Since the main object of present work is the enhancement of gas turbine

performance by improved intake air temperature therefore, the test procedure

has been separated to classify for heat exchanger once and gas turbine again

another time. This was because of the following reasons :

1. Parameters effected on the design and performance calculations for heat

exchanger are needed high air flow rate values, which are negatively effect on

gas turbine performance.

2. The high inlet velocity into compressor may be increasing pressure drop for

what of which negatively effect on gas turbine performance.

3. Difficulty to make sure of compressor inlet temperature at desired value for

present work requirements.

3.8 Test Procedure:

Two types of experiments were performed:

1. Measurements for the heat exchanger:

On commencing the tests, all valves around the constructed rig are closed

without water circulation through the heat exchanger. For the chilled water

circulation tests, a piece of ice was added to the water tank continuously to

keep a constant feeding temperature to the test heat exchanger section. The

hot water tests were conducted by switching on the immersion heaters of the

hot tank and controlling the temperature by setting the thermostat at the

required water temperature. A check should be made to the air supplying fan

prior to the experiments.

After completing checking above steps, the test process begins by

switching on the circuit breaker that supplies power to the system. The water

pump will start and open the gate valve for the cold tank that controls the flow

rate of water which circulates on the tube side of the heat exchanger. At

particular time, when the temperature inside the cold tank is reaching (10°C),

the operation conditions for heat exchanger were recorded when the air flow


rate was fixed at (2000 cfm).Since the air flow passes through the duct that

connected the heat exchanger to gas turbine, the air flow rate (cfm) can be

measured by using (pitot-tube) fixed at compressor inlet box.

After a few minutes, the gate valve of the hot tank is opened, so that the hot

water is mixed with cold water to increase the temperature gradually until

reaching particular values such as (20°C) , (30°C) , (40°C) and (50°C) .The

process was repeated for other air flow rate such as (1000 cfm) and (500 cfm)

for different circulated water flow rates.The following operating conditions

were measured during the tests for each air flow rate:

• The inlet and exit temperature of water side across the heat exchanger.

• The circulated water flow rate.

• The air temperature on both sides of the heat exchanger.

The experimental data collected for the heat exchanger are listed in

Appendix (A).

2. Measurements for the gas turbine:

The Kerosene fuel is supplying in fuel tank about (20 Lit.), for lubricating

shafts oil (SAE-10) is supplying in oil tank. The gas turbine engine is

connected to a water supplying source to cooling purpose. Starting operation

begins by firing of about (0.3) liter of (methylated alcohol) to ensuring the

appropriate flame for kerosene combustion.

The gas turbine was allowed to operate until it reaches the required

conditions at a particular speed; at this time the engine takes the appropriate

air flow rate .The compressor inlet temperature is fixed at indicated value

such as (15°C) ,(20°C), (25°C) and (30°C) by utilizing from:

• Variation of ambient temperature along the day and season.

• Variation of air exit temperature from heat exchanger.

The time was too long to take the measurement variations of the ambient

temperature (during the season) due to the fact that the temperature gradient

across heat exchanger (Single-Pass Elliptical Tube, two Row Radiator) is


small. It was decided to run the rig for (5 to 10) times at each specified

ambient temperature with different times along the year. When the

compressor inlet temperature is fixed at (15°C),the tests were conducted at

different gas generator speed starting from (45000 RPM) . This speed was

increased to (55000 RPM) by burning more fuel. The data were collected for

each specified speed and compressor inlet temperature. The same procedure

was repeated for other air intake temperature such as (20°C),( 25°C)and

(30°C). At all tests, the ambient temperature was changing from (18°C) to

(30°C).To show the relationship between the gas generator speed ,compressor

inlet temperature and turbine inlet temperature, the turbine inlet temperature

was fixed at a particular value such as (586 °C) and compressor inlet

temperature was taken at different time such as (15 °C) and (22°C).The data

collected during the tests are shown in Appendix (A).

CHAPTER FOURE

Chapter four…………………………………………………………………………Theory

4.1 General

This chapter deals with the mathematical modeling of the core design of

heat exchanger (Length, Depth, Height) corresponding to the intake air box of

gas turbine and mathematical modeling of gas turbine.

The step by step technique will be used to simulate core design of the air

cooled heat exchanger. In this method, the heat exchanger is divided into two

dimensions (Depth, Height) ,as it will be described later.

On other hand, Modeling of the computational program to predicate the

effect of compressor inlet temperature on gas turbine performance by using

the non-dimension method was also considered. 4.2 Heat Exchanger Thermal Design:

4.2.1 A Comprehensive Design Procedure

The methodology of arriving at an optimum heat exchanger design is a

complex one. Not only because of the arithmetic involved, but it is more

particularly because of the many qualitative judgments that must be

introduced. The design procedure in a schematic presentation is shown in Fig

(4.1) Kays and London (1984) [26].

The design theory procedure can be set-up on a computer program. The

inputs to the design theory procedure include:

1. Surface Characteristics: flattened tubes, surfaces with flow normal to banks

of smooth tubes, Finned-tube surfaces, normal distance (XT), longitudinal

distance (XL) …etc.

2. Problem Specifications: The problem statement may specify a

consideration of different exchangers. For instance, periodic-flow and direct-

transfer types. Like cross flow, inline tubes, both fluids unmixed flow …etc.

3. Physical Properties: Some options may be allowed in the physical

properties the matrix material to be used in a periodic flow type exchanger.


Figure 4.1 Methodology of Heat Exchanger Design

4. Optional Solutions: optional solutions may represent an estimate of what a

competitor may offer, others may represent customer's suggestions. For

example, what are exit parameters from heat exchanger, heat transfer

coefficient, heat load …etc.

Surface Characteristics

Problem Specifications

Physical Properties

Design Theory

Procedure

Optional Solutions

Optimum Solution

Evaluation Procedure

Evaluation Criteria


5. Evaluation Procedure: The theoretical design must be furnished with

evolutions criteria to obtain optimum solutions for design theory.

4.2.2 Numerical Modeling of Cross Flow Compact Heat Exchanger

In order to develop a numerical model with the predictive capability for

various design parameters of the heat exchanger, step by step method with

two dimensions was employed in this study. Each tube row was divided into a

number of horizontal slices occupying the total length of the heat exchanger

and for air side the exit of one row considering inlet to next row as shown in

Fig. (4.2).Step by step method enables us to take into account the significant

air temperature increase as well as the local variations of the properties and

the heat transfer coefficient. Forms of the mass and the energy conservation

equations were derived for two dimensional grid systems.

4.2.2.1 Grid System

In the present study the following assumptions were assumed :

1. Homogenous temperature distribution of air all over the frontal face area of

the heat exchanger and hence for each slice.

2. Uniform mass flow distribution for both stream sides of heat exchanger.

3. The exit air condition for each row represents a mean value for all of the

slices of the considered row. This will be the inlet condition for the next row.

4. The inlet air velocity for each row was assumed to be uniform represented

by a mean representative value.

5. The water temperature variations between the rows were also assumed to

be negligible.

6. The air velocity stream and maximum air velocity difference assumed to

be negligible.

In the present study the design requirements for the theoretical model are:

1. The velocity in tube side (vw)

2. The velocity of air side (va.)


Figure 4.2 Step by Step method with two Directions


3. The air temperature difference (∆Ta) per row

4. The air mass flow for each slice.

Therefore only a single fin tube of one horizontal slice assembly was

explained as shown in Fig. (4.3a) and the heat exchanger performance for any

locate can be calculated the performance for each slice and calculating

performance for each row Fig.(4.3b). The nodal points for the calculation of

the variables of the air and the coolant were defined as illustrated in Fig.

(4.3c). Nodal points for air temperature were assigned on the east and the

west sides of the control volume and the nodal points for the coolant were

assigned on the north and the south sides of the control volume. The nodes

were evenly distributed in two directions Ni × Nj. The heat transfer

coefficient and heat load were calculated at each control volume.

Figure (4.3a) Slice for Figure (4.3b) Exit of one row

inlet single Tube inlet to next row

Two

Twi

Slice 2

Slice 1 Lf

Water Flow Tao

Dt

Tar2

Tar1

Tai

Air Flow


J=1 2 ………………Nr

N-1

.……

……

……

….…

…2

I=1

Figure (4.3c) Nodal Points Distributions

with two directions

4.2.2.2 Physical Characteristics of Heat Exchanger

From the selected velocity of the tube side of the water stream and the

known tube cross section dimension, the total number of tubes of the heat

exchanger can be calculated from:

(Nt) = [cwwr

w

AvNV

××

.

] ………………………………………………….(4-1)

This value will be used for the estimation of the length and depth of the heat

exchanger as:

)1( +×= tT NXL ………………………………………………….(4-2)

tL DNrXD +−×= )]1([ ………………………………………………….(4-3)

4.2.2.3 Mass Conservation

Mass conservation of the water flow through the tube is simply

Σ in −Σ out = 0 ………………………………………………….(4-4) .m

.m

and the mass conservation equation for the water at each slice can be written

as:


.m

.m

.m

..

)()( jVj aa ×ρ

w(i+1,j)= w(i,j) ………………………..…..………………………….(4-5) In case of the heat exchangers with plain continuous fins, the mass flow rate

in the air flow direction can be calculated by the following equation.

a(j)= ……………………………………………………...(4-6) 4.2.2.4 Log-Mean Temperature Difference

To estimate the true mean temperature difference ( mTΔ ) between the two

fluids Fig. (4.4) shows the possible flow direction of both streams in compact

heat exchanger.The following relations may be used for the estimation of the

logarithmic mean temperature difference according to counter flow directions,

Smith(1997) [26]:

)),(),1(())1,(),((

ln

)),(),1(())1,(),((

jiTjiTjiTjiT

jiTjiTjiTjiTLMTD

wa

wa

wawa

−++−

−+−+−= ………………………….(4-7)

Ta,(i,j)

∆Tm

Tw,(i,j+1) Ta,(i+1,j)

∆Tm

Tw,(i,j)

0 Atotal

area Figure 4.4 The Mean Temperature Difference Along a Single Pass


The actual temperature difference of a cross flow compact heat exchanger is

obtained by applying a correction factor (F) to the (LMTD) value as, Hewitt

(1998) [27]:

LMTDFTm ×=Δ …………………………………………………….(4-8)

where:

])1(1[2

])1(1[2ln[)1(

)1()1(ln)1(

2

2

2

+++−

+−+−−

−−+=

RRS

RRSR

RSSRF ……………………………………….(4-8a)

)),()1.(()),1(),((

jiTjiTjiTjiT

Rcc

hh

−++−

= ……………………………………….(4-8b)

)),(),(()),()1,((

jiTjiTjiTjiT

Sch

cc

−−+

= ……………………………………….(4-8c)

4.2.2.5 Heat Load

The heat load passes through a control volume on the water side is: Q (i,j)= (i,j) ×Cp (i,j) ×ΔT (i,j) …………………………………..…….(4-9)

.m

It can be expressed with the overall heat transfer coefficient, U(i,j), as follows:

Q(i,j)= Uo (i,j) × A × (i,j) × ΔTm (i,j) ……………………………...…….(4-10)

The above equation is applying for each slices per one row.

4.2.2.6 Overall Heat – Transfer Coefficient (Uo):

The local overall heat transfer coefficient for each slice based on the outside tube

area can be written as follows , Holman(1989) [28]:

)114....()]

)(1()

)(),(1()

)(()

),(1[(

1),( −

×+

×++

×

=

utonifutoniwoutint

t

oa

o

AAhAAjihAAKt

jih

jiU

η


4.2.2.7 Forced Convection Heat Transfer Coefficient Inside Tube:

Numerous relations have been proposed for predicting fully developed

turbulent flow in tubes. The popular Dittus-Boelter equation, Dittus-Boelter

(1930) [30] is usually given in the form:

)( Pr Re 023.0 8.0

hww d

kh n=

]

………………………………………... (4-12)

and cross flow area for water side (rectangular tube with semi circular ends ):

[ ] [ π×−+−×−= 2)()()( ttttttwc THHDTHA ……………………………..(4-13)

also

μρ whud

=Re ……………………………………..…………………….(4-14)

t

cwh

Ad

λ4

= ……………………………………..………………….….(4-15)

kCpμ

=Pr ……………………………………..…………………....….(4-16)

where the ranges of Re and Pr are:

120Pr5.010Re6000 7

≤≤≤≤

where the coefficient (0.023) is recommended by McAdams (1954) [31] in

place of (0.0243) originally given by Dittus-Boelter.

Also,

n= 0.4 for heating

n= 0.3 for cooling

4.2.2.8 Forced Convection Heat Transfer Coefficient for Air Side:

The entrance region for the development of the longitudinal velocity profile

and the temperature profile is about 10 times the hydraulic diameter. This

criterion is particularly valid for calculating the time-averaged coefficient for

fluids (air and water) Adrian [32] . There are several empirical relationships for

heat transfer between the duct surface and the fully developed flow, the


analytical form of these relationships is based on exploiting the analogy

between momentum and heat transfer. The popular Dittus-Boelter equation is

also used for fully developed turbulent flow (air side):

)( Pr Re 023.0 8.0

haa d

kh n= ………………………………………….. (4-17)

For laminar flow, the Sieder and Tate (1930) [33] correlation can be used.

( ) )(PrRe86.13.0

3.0

haf

haa d

kD

dh ⎟⎠⎞

⎜⎝⎛= ………………………………….……..….….(4-18)

where cross flow area for air side:

))(2

( tTf

ac HXP

A −= ………………………………………………..….(4-19)

The hydraulic diameter for the air side:

f

caah

Ad

λ4

= ……………………………..………………….….(4-20)

fλ =2 Lf + pf

The number of slices (N) depends on the water temperature difference along

the tube bare; by assuming the air mass flow rate across frontal area is divided

equally on the slices (N), therefore the Reynolds number for air side is

represented by:

Njmajima )(),(

..

= …………………………………………………………..(4-21)

)(),(

.

ca

aa A

jimG = ……………………………………..………………..……. (4-22)

),()(

),(ji

GdjiR

a

aaha μ

= ………………………………..………………….… (4-23)

By assuming that there is no temperature gradient between the tubes that

share the fins then the fin can be considered to be insulated at the center.

Therefore fin efficiency (ηf) can be calculated as that for the case with an

adiabatic tip, Briggs and Young (1963) [34]:


))2

((

))]2

([tanh(

f

f

f Lm

Lm

×

×=η ………………………………..………………… (4-24)

where:

)()(),((

[cat

fa

AKPjih

m×

×= ………………………………..………………..… (4-25)

The overall surface fin efficiency is:

)]1()[(1 ff

o AA

ηη −×−= ………………………………..………………… (4-26)

where:

)1)(4(f

fff

PDL

HA

××= ………………………………..……………….. (4-27)

)1)](2())(2()[((exp

ffttt P

THDHH

A×−−×+×= π ………..………………… (4-28)

)()( exp

HA

HA

HA f += ………..…………………………………. ………….. (4-29)

The log- mean temperature difference and overall heat transfer coefficient is

calculated to obtain the height for one slice as follow:

)),(),((),(),(

jiLMTDFAjiUjiQjiH

o ×××= …………………………………….(4-30)

4.2.2.9 Power of Fan:

The pressure drop across heat exchanger can be calculated from the general

pressure drop relationship, London(1983)[35] is most often written in terms of

hydraulic diameter,

ha

fa

dD

fv

P ××=Δ )2

(2

ρ …………………………………………………..(4-31)


To calculate ΔP, the friction factor ƒ must be known and it can be derived

from the flow solution.The friction factors derived from the Colburn [35] flows

described by eqs. (3.32) for fully developed laminar flow:

ƒ=Ra64 …………………………………………………………………...(4-32)

For isosceles triangular ducts, Bhatti and Shah(1987)[36] recommend, for

fully-developed turbulent flow:

ƒ= 25.0)(078.0

Ra……………………………………………………………….(4-33)

For the power of the fan the equation is:

fan

airfan

pVPη

Δ=

.

……………………..………………………………….….(4-34)

And,

air

airair

mV

ρ&

=.

……………………..………………………………..…….(3-35)

Thus,

fanmotor

fanfanmotor

PP

,, η

= ……………………..…………………………….….(3-36)

4.2.2.10 The Computer Program

The computer program was built in this study to establish the thermal

design of the compact heat exchanger incorporating the quick-basic computer

language. The flow chart of the program (CPHE) is shown in appendix

(B).The following procedure describing the calculation technique of the

present model :

1. Input surface characteristics and inlet operation conditions for both fluids.

2. Calculating the cross sectional area of tube side.

3. Choosing the velocity in tubes and input water flow rate wV&

4. Calculating (No. of Tubes / Row) ,Eq.(4-1)


5. Calculating the length and depth of heat exchanger Eq.(4-2) and Eq.(4-3).

6. Calculating the No. of slices from water temperature conditions.

7. Loop (j=1…….j=no. of row)

8. Assume the air exit temperature for first row.

9. Calculating the air mass flow rate for first row, Eq (4-6).

10. Assume the air velocity over tubes.

11. Calculating the height of the heat exchanger.

12. Loop (I=.1……i=no. of slices).

13. Choosing the water temperature difference for each slice.

14. Calculating the heat load for first slice and first row, Eq.(4-9).

15. Assume the air mass flow rate for each slice, Eq.(4-21).

16. Calculating the air exit temperature for first slice.

17. Correct the fluids properties.

18. Calculating the correct heat load for first slice.

19. Calculating the correct air exit temperature for first slice.

20. Calculating the cross sectional area, hydraulic diameter, Reynolds No., Nusselt No, Eq.(4-13) to Eq.(4-20).

k

21. Calculating the overall heat transfer coefficient, Eq.(4-11).

22. Calculating the height of heat exchanger, Eq.(4-30).

23. Repeat the calculation with the iterated value of air exit temperature until

the error percent calculated from:

..).(

%calVariable

assuVariblecalVarible −=ξ is converged to a value within 1 .hence,

the variable here is air exit temperature or air velocity over tubes.

310−×

24. Repeat process for all slices (Ni).

25. Calculating the mean air exit temperature and check it with assumed value.

26. Repeat the process for all rows (Nj).

27. Calculating heat exchanger performance (heat transfer coefficient for air

side (ha), overall heat transfer coefficient (Uo), heat load)


28. Calculating height of heat exchanger (H).

The step by step method is formulated in this program to establish the

following characteristics and operating conditions of the heat exchanger:

1. Air exit temperature

2. Air mass flow

3. Heat transfer coefficient for both sides

4. Overall heat transfer coefficient

5. Heat load

6. The height of heat exchanger (H)

7. The length of heat exchanger (L)

8. The depth of heat exchanger (D)

9. Number of fins per tube

10. Number of tubes per row

11.In addition, heat exchanger performance can be predicted for any required

point in heat exchanger in two direction (Ni X Nj).

4.3 Gas Turbine

A conventional power plant receiving fuel energy (Qadd), producing work ( )

and rejecting heat to a sink at low temperature (Qrej) is shown in Fig.( 4.5) as

a block diagram.

.w

Figure. 4.5 Basic Gas Turbine Engine


The objective is to achieve the least fuel input for a given work output as this

will be economically beneficial in the operation of the power plant.

Usually, a gas turbine plant operates on ‘open circuit’, with internal

combustion Horlock(1987) [37]. Air and fuel pass into the compressor and

combustion chamber, respectively, and the combustion products leave the gas

turbine after expansion through the turbine.

4.4 Basic Gas Turbine Cycles

In power plant thermodynamics for high thermal efficiency led us to

emphasis on raising the maximum temperature T3 and lowering the minimum

Temperature T1. Thus, the efficiency will be increased with the ratio (T3/T1).

In a gas turbine plant, this search for high maximum temperatures is limited

by material considerations and cooling of the turbine is required. This is

usually achieved in ‘open’ cooling systems, using some compressor air to

cool the turbine blades and then mixing it with the mainstream flow. In

practical open circuit gas turbine plants with combustion, real gas effects are

present (in particular the changes in specific heats, and their ratio, with

temperature), together with combustion and duct pressure losses.

4.5 A non- Dimensional Analyses of Gas Turbine Performance:

In practical open circuit gas turbine plants with combustion, real gas effects

are present (in particular the changes in specific heats, and their ratio, with

temperature), together with combustion and duct pressure losses. Now some

modifications of the air standard analyses and their graphical presentations for

such open gas turbine plants must be developed, as an introduction to more

complex computational approaches.


The Hawthorne and Davis(1956) [38] analysis is first generalized for the

[CBT] open circuit plant, with fuel addition for combustion, (ƒ) per unit air

flow, changing the working fluid from air in the compressor to gas products

in the turbine, as indicated in Fig.(4.6).

T 3

Combustion Fuel

4

2

2s 4s (1+ƒ)

5 Gases Products

5s

1

S Figure 4.6 (T-s) Diagram for Irreversible Two-Shaft Circuit Simple Plant

Real gas effects are present in this open gas turbine plant; specific heats and

their ratio are functions of ƒ and T, and allowance is also made for pressure

losses.

4.5.1 Component Performance

Before moving on to the air standard analyses of irreversible gas turbine

cycle, need to be define various criteria for the performance of some

components. In addition, to the irreversibilities associated with these

components, pressure losses (Δp) may occur in various parts of the plant (in

the entry and exit ducting, the combustion chamber, and the heat exchanger).


These are usually expressed in terms of non-dimensional pressure loss

coefficients,

Ѕ= Δp/ (p)in ……………………………………………………………(4-37)

where:

(p)in=is the pressure at entry to the duct.

ΔP23= 2

32

PPP − ………………………………………………………..(4-37a)

ΔP5a =a

a

PPP −5 (Pa is ambient pressure)………………………….…(4-37b)

As alternatives to the isentropic efficiencies for the turbomachinery

components, ηT, ηC, which relate the overall enthalpy changes, small-stage or

polytropic efficiencies (ηpT and ηpC) are often used.

The pressure-temperature relationship along an expansion line Fig.(3.6) is

then

P/ Tz = constant ………………………………………………………(4.38)

where:

Zg=γg / [(γg-1) ×ηpT ] …………………………………………………..…(4.39)

and the entry and exit temperatures are related by

T4 / T5s =( rT)1/Zg = xt …………………………………………………..(4.40)

Along a compression line,

P/ Tz = constant …………………………………………………….….(4.41)

where:

Za=[γa×ηpC]/ (γa-1) ……………………………………………………...(4.42)

and exit and entry temperatures are related by

T2s / T1 =( rC)1/Za = xc ………………………………………………….(4.43)


4.5.2 Graphical Plot

In graphical interpretation, using isentropic rather than polytropic

efficiencies,Hawthorne and Davis [38] plotted the following non-dimensional

quantities, all against the parameter x = r (γ-1)/γ as follows:-

(i) Non-dimensional compressor work,

Wc= cpa×T1× (xc-1)/ηc …………………………………………………(4.44)

(ii) NDCW = Wc/cp(T3 – T1) = =(Xc-1)/(ηc× (Ø-1)) ………………….....(4.45)

Non-dimensional turbine work,

WT=(1+ƒ) × cpg ×T3×ηt(1-1/xt) …………………………………………(4.46) (iii) NDTW = WT/cp (T3 – T1) ==((1+f) ×ηt× (1-1/Xt))/((1-1/Ø) ×n) .…(4.47)

Non-dimensional overall efficiency,

ηo=NDTW /(H. V× (1+ƒ)) …………………………………...…………(4.48)

where: Ø=T3/T1 ratio ……………………..……………………………………(4.49) n= cpa/ cpg ……………………………………………………………(4.50)

4.6 Computer Calculations for Two-Shaft Gas Turbine

The analytical approach outlined above for the two-shaft gas turbine plants

is that used in modem computer codes. However, gas properties, taken from

tables such as those of Keenan and Kaye(1945) [39] as shown in fig. (4.7).The

flow chart of the program (CPTGTP) is illustrated in Appendix [B].


1

Figure 4.7 Specific Heats and Their Ratios for ‘Real’ Gases-Air and

Products of Combustion

The purpose of the computer calculations is to examine the possibility of

increasing the overall performance of two-shaft gas turbine depending on inlet

conditions only. 4.7 Results and Discussion of Theoretical Calculations

The dimensionless parameter such as overall efficiency, expansion ratio,

and maximum to minimum ratio (Ø) can be used for predicting the effect of

compressor inlet temperature on gas turbine performance. The prediction is

focused on the following concepts:

4.7.1 The effect of compressor inlet temperature on maximum to

minimum ratio (Ø):

One of two operating limits for gas turbine it is the turbine inlet

temperature. It is not often when the engine runs at topping temperature will

produce higher overall efficiency, but it depends on the minimum

temperature. Therefore, wherever the minimum temperature is reduced, the

engine will be at higher efficiency. This is due to the fact that the equilibrium

condition between the compressor power requirement (which increases at

high minimum temperatures) and the power produced by gas generator


turbine (which is not directly influenced by the minimum temperature) will be

satisfied at a lower value. Figures (4.8) to (4.11) show the effect of the Ø on

the overall efficiency for inlet high pressure turbine temperature of

(699°C),(628.5°C),(588°C) and (547.5°C) respectively at expansion ratios of

1.08 and 1.12.It is clear from these figures that for a given expansion ratio the

overall efficiency shows a gradual increase with Ø .Further, the trend shows

a raise in the overall efficiency as the inlet turbine temperature increases for a

given expansion ratio. It is obvious that the increasing percentage varies with

expansion ratio or turbine inlet temperature. When turbine inlet temperature

was (699°C) the increasing percentage of overall efficiency was about (18%)

at expansion ratio (1.08), while the increasing percentage of overall efficiency

fell to (15%) at expansion ratio (1.12).Also, for the turbine inlet temperature

(547.5°C) the increasing percentage of overall efficiency was about (20%) at

expansion ratio (1.08),while the increasing percentage of overall efficiency

was reduced to about (17%) at expansion ratio (1.12). However, there is

relationship between (T3), (T1), and engine speed, at full load. Two shaft

engines will run either at temperature topping or at speed topping. At speed

topping, the engine will not reach its full firing temperature, while at

temperature topping; the engine will not reach its maximum speed.

The effect of the compressor inlet temperature on the overall efficiency is

shown in figures (4.12) and (4.13) for expansion ratios of 1.12 and 1.08

respectively. When reducing the compressor inlet temperature from (30°C) to

(15°C), the overall efficiency shows an increase by (17%) at expansion ratio

(1.12).Also, the overall efficiency shows an increase by (19%) at expansion

ratio (1.08).Noting that the higher the expansion ratio, the higher the overall

efficiency is obtained as shown above.


4.7.2 The effect of compressor inlet temperature on the expansion ratio:

Changes in the minimum temperature have less impact the component

efficiencies than the overall cycle output.Therefor, the impact of minimum

temperature is usually less pronounced for the expansion ratio than for the

overall efficiency. It is known from gas turbine cycle when expansion ratio

increases, the overall efficiency will be increased as fig.(4.14),and fig.(4.15).

But as both turbine inlet temperature (T3 ) and Ø (maximum to minimum

ratio) increased it was found that the overall efficiency will be increased too.

From the mentioned conditions, it can be concluded that any gas turbine

cycle shows an improvement in thermal efficiency as long as increasing the

turbine inlet temperature (T3) or lowering compressor inlet temperature. As,

the expansion ratio increases, the work consumed by the compressor is

reduced. Figure (4.16) and figure (4.17) showed the effect of reduced

compressor inlet temperature on the overall efficiency. It is obvious that this

effect can be clearly seen when the engine runs at high speed. So, when

reducing compressor inlet temperature from (30°C) to (15°C), the increasing

percentage in overall efficiency is fluctuated among (15 %), (17%) and (20%)

at expansion ratio (1.12), (1.1) and (1.08), respectively.


2

3

4

5

6

7

8

3.08 3.1 3.12 3.14 3.16 3.18 3.2 3.22 3.24 3.26 3.28

T3/T1 ( )

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)

Expansion Ratio (rT) =1.12

Fig (4.8) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=699 °C

Fig (4.9) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=628.5 °C

2

2.5

3

3.5

4

4.5

5

5.5

6

6.5

7

3.08 3.1 3.12 3.14 3.16 3.18 3.2 3.22 3.24 3.26 3.28

T3/T1 ( )

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)



Ø


Expansion Ratio rT=1. 12


Ø

rT=1. 12 Expansion Ratio rT=1. 12

rT=1. 08 Expansion Ratio rT=1. 08


2

2.5

3

3.5

4

4.5

5

5.5

6

6.5

3.08 3.1 3.12 3.14 3.16 3.18 3.2 3.22 3.24 3.26 3.28

T3/T1 ( )

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)

Expansion Ratio (rT) =1

Fig (4.10) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=588 °C

Fig (4.11) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio)

with Isentropic Efficiency (ηt,ηc =0.9) ,T3=547.5 °C

.12

Expansion Rat .08io (rT) =1

rT=1. 12

rT=1. 08

Ø

2

2.5

3

3.5

4

4.5

5

5.5

6

6.5

3.08 3.1 3.12 3.14 3.16 3.18 3.2 3.22 3.24 3.26 3.28

T3/T1 ( )

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)


Expansion Ratio ( 8rT) =1.0

rT=1. 12

rT=1. 08

Ø


3

3.5

4

4.5

5

5.5

6

6.5

7

7.5

500 520 540 560 580 600 620 640 660 680

Turbine Inlet Temperature (T3)

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)

T1= 15 C

T1= 30 C

Fig (4.12) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on the

Theoretical Overall Efficiency ,Expansion Ratio=1.12

2

2.5

3

3.5

4

4.5

5

5.5

500 520 540 560 580 600 620 640 660 680

Turbine Inlet Temperature (T3)

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)

T1= 15 C

T1=30 C

Fig (4.13) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on the

Theoretical Overall Efficiency ,Expansion Ratio=1.08


3

3.5

4

4.5

5

5.5

6

6.5

7

7.5

1.075 1.08 1.085 1.09 1.095 1.1 1.105 1.11 1.115 1.12 1.125

Expansion Ratio (rT)

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)

T3=669 C , =3.2

T3=628.5 C, =3.1

T3=588 C, =2.9

Ø

Ø

Ø

Fig (4.14) Theoretical Overall Efficiency as a Function of Expansion Ratio with Isentropic Efficiency (ηt,ηc =0.9) and T1=15°C

3

3.5

4

4.5

5

5.5

6

1.075 1.08 1.085 1.09 1.095 1.1 1.105 1.11 1.115 1.12 1.125

Expansion Ratio (rT)

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)

T3= 699 C, =3.10

T3=628.5 C, =2.97

T3=588 C, =2.84

Ø

Ø

Ø

Fig (4.15) Theoretical Overall Efficiency as a Function of Expansion Ratio with Isentropic Efficiency (ηηt,ηc =0.9) and T1=30°C


2

3

4

5

6

7

8

10 15 20 25 30 35

Compressor Inlet Temperature (C)

Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)

Expansion Ratio (rT =1.12 )

Fig (4.16) The Effect of Compressor Inlet Temperature (with Variable Expansion Ratio)

on the Theoretical Overall Efficiency, T3= 669 °C

Fig (4.17) The Effect of Compressor Inlet Temperature (with Variable Expansion

Ratio) on the Theoretical Overall Efficiency, T3= 628.5 °C


Expansion Ratio (rT) = 1.08

2

2.5

3

3.5

4

4.5

5

5.5

6

6.5

7

10 15 20 25 30 35


Theo

retic

al O

vera

ll Ef

ficie

ncy

(%)


Expansion Ratio =1.1 (rT)

Expansion Ratio 1.08 (rT) =rT= rT=

rT=

rT=1. 08

Expansion Ratio rT=1. 1 rT=1. 1 2Expansion Ratio rT=1. 12


CHAPTER FIVE

Results & Discussion

Chapter Five .......................................................................................Results &Discussion

5.1 General

Since gas turbine is an air-breathing engine, its performance is changed by

anything that affects the density and/or mass flow of the air intake to the

compressor. The volumetric flow is constant with any shaft speed, it is

possible to increase mass flow rate by increasing air density. Air-cooled heat

exchanger has been sitting prior to intake of compressor .In order to find out

the effect of compressor inlet temperature on the gas turbine performance.

These important variables are: power output, fuel mass flow rate, heat

consumption, heat rate, overall efficiency, air mass flow. The design and

thermal performance of air-cooled heat exchanger are also studied. A new

design technique is suggested which divided the heat exchanger in two

directions to get more accurate thermal performance: heat load, heat transfer

coefficient (air side), overall heat transfer coefficient, air mass flow, air exit

temperature from heat exchanger. The present study concentrates on getting

variable air exit temperature, air mass flow rate, size (aspect ratio), with

known inlet operation conditions for both fluids only.

5.2 Computational Model Results for Heat Exchanger

The computational model results of the developed program shall be

discussed, including heat load, heat transfer coefficient for air side (ha), and

overall heat transfer coefficient for each row of the heat exchanger. Air exit

temperature and air mass flow rate discharged from heat exchanger are to be

entered to gas turbine which must be controlled according to the requirements

of present study. The computational model has been fed with the same

operating conditions as those of the experimental test rig. For the object of

validity of the theoretical prediction it was decided to use two set of

experimental tests for the present heat exchanger in this work. These were

conducted at entering water temperature of (10 °C) and (50 °C) for water flow

rate of (2000 l/h) at two different air flow rates of (500 cfm) and (2000 cfm),

as shown in tables (5-1) to (5-4).For the performance simulation, these tests


were achieved at the same entering air temperature to the heat exchanger of

(32 °C) .In these tables, three different heat exchanger core sizes were

obtained for the same operating conditions on both sides, air and tube, sides

of the test section.

5.2.1 Heat Load for heat exchanger

Figures (5.1a) and (5.1b) show the variation of heat exchanger load with

the air flow velocity through the tube bank for water entering temperature of

(10°C) and (50°C) respectively. The air velocity was ranged between (1.2

m/s) and (4.6 m/s) corresponding to (500 cfm) and (2000 cfm) air flow rates

respectively. It is obvious that the heat load experience an increase as the air

flow rate increase .This is due to the improvement of the overall heat transfer

coefficient of the heat exchanger by increasing the air side heat transfer

coefficient. The heat load was increased by 3-4 times when the velocity was

raised from (1.2 m/s) to (4.6 m/s) respectively, at entering water temperature

of 10°C (Cooling Mode) and 50°C (Heating Mode). The heat exchanger load

was ranged between (1 kW) and (4.5 kW) for the whole range of air flow rate.

The step by step simulation model shows a good agreement with the

experimental data as shown in figure (5.1) for the whole range of air velocity.

The predicted heat exchanger performance (Heat Load, Overall Heat Transfer

Coefficient ...etc) fell within 5 % for most of the simulation range. The trend

of the lines of heat load is the same for upper point for both cases, but the

magnitude of heat load is having different values at low air velocities. For

example, in figure (5.1a) at air velocity (1.5 m/s), the heat load is (1 kW) at

entering water temperature of (10 °C) and figure (5. 1b) for entering water

(50 °C), the heat load is (1.5 kW).Which explains the heat load behavior is

affected by the fluid properties at high temperature.

5.2.2 Heat Transfer Coefficient for air side

Since the heat transfer performance of the heat exchanger for (gas-liquid)

type is dependent on heat transfer coefficient of air side .Figures (5.2a), and


(5.2b), show a comparison for the heat transfer coefficient for air side

between the theoretical prediction and the experimental data. These figures

represent the behavior of the heat transfer coefficient variation with water

flow rates of (2000 l/hr). It is obvious from these data that, the heat transfer

coefficient values decrease at low air velocity. And as a matter of fact, it has

lower value when water entering at (10 °C).The prediction of the present

model shows that it is possible to obtain the higher temperature difference for

the air side at the same heat exchanger load when using a fat heat exchanger.

This is due to the higher overall heat transfer produced at the same air flow

rate as shown in tables (5-1) to (5-4). The predicted values from the

computational model for the heat transfer coefficient for air side along its

depth (D) and along heat exchanger height (H) are shown in figures(5-3) and

(5-4) for air flow rates of (2000) and (500) cfm at water entering temperature

(10 0C) and (50 °C) respectively. It is obvious that the heat transfer coefficient

of the air side nearly stayed unchanged ,and it is essentially a constant value.

However, it is more pronounced when the heat exchanger geometry was

deeper in the depth direction as shown in figure (5.5).This is for a heat

exchanger having an overall dimension of (18 × 11 × 18) cm3, table (5-1).It

shows that the air heat transfer coefficient is in the range between (94) to (96)

W/m2.°C.

5.2.3 Overall Heat Transfer Coefficient

Figures (5.6a), (5.6b) show the variation of the overall heat transfer

coefficient with air flow velocity through heat exchanger at water entering

temperature of (10 °C) and (50 °C) respectively. The predicted values from

computational model for the overall heat transfer coefficient along its depth

(D) and along heat exchanger height (H) are shown in figures(5-7) and (5-8)

for air flow rates of (2000) and (500) cfm at water entering temperature

(10 °C) and (50 °C) respectively. It is obvious that the overall heat transfer

coefficient of the air side did not change much and it is essentially a constant


value. However, it is more pronounced when the heat exchanger geometry

was deeper in the depth direction as shown in figure (5.9).This is for a heat

exchanger having an overall dimension of (18 × 11 × 18) cm3, table (5-1).It

shows that the air overall heat transfer coefficient is in the range between

(84) to (85) W/m2.°C.The overall heat transfer coefficient of the heat

exchanger approaches (47) and (45) W/m2.°C at the maximum air velocity

falling to a minimum at (1.2 m/s) where Uo is (15 W/m2.°C) and (30 W/m2.°C)

for cooling and heating modes respectively. The simulation prediction and

experimental data showed a good agreement with maximum descripancy of

(5%) for the whole range of test conditions.

5.2.4 Air Temperature Distribution

Figure (5.10) shows the variation of the experimental and present model of

the exit air temperature out of the heat exchanger and the air velocity. For

both of the cooling and heating modes of the air passing through the heat

exchanger, the exit air temperature shows a trend of decreasing as the air

velocity increases. This is due to the fact that when the air velocity increases

causes an increase in the air side heat transfer coefficient (ha ) and the

(Uo )value which in turn produces higher heat exchanger load. However, the

exit air temperature per any row and slice is one of the dominant parameters

which were focused in the present work. Figures (5.11) and (5.12) show the

air temperature distribution along its depth (D) and along heat exchanger

height (H) for air flow rates of (2000 cfm) and (500 cfm) at water entering

temperature (10 °C) and (50 °C). Figure (5.13) shows the temperature

variation with row number with the leaving side for each row for a heat

exchanger of 6 rows having (18 × 11 × 18) cm3.

5.2.5 The Effect of Core Aspect Ratio (H/L) and Size (L×D×H)

The objective of this study is to develop a heat exchanger model which has

the capability to predict the heat exchanger performance depending on the

design parameters without relying on experimental data. The predictive


capability of the model was demonstrated by studying two different cases. For

the first case, the aspect ratio of the heat exchanger core was changed from

(0.67 - 1) when air velocity (1.2 m/s). The water flow rate was kept constant

at (2000 l/h). The aspect ratio is defined as the ratio of the height (H) to the

length (L) of the core for the heat exchanger. Since the model is based on the

down flow type with the water tanks on top and bottom of the core, larger

aspect ratio means relatively longer tubes. For a given range of the aspect

ratio, the results showed that a heat exchanger with smaller aspect ratio can

perform better than that of larger aspect ratio case as shown in Figure (5.14).

Pressure drop in air side (Δpa ) was increased by changing the aspect ratio

from (0.67 - 1) with the different core size (L×D×H). This is because at

higher aspect ratio, the frontal area decrease over air side experiencing more

abstraction in flow. For the second case, pressure drop variation with the

change of the core size was examined. The pressure drop was calculated for a

given water flow rate (2000 L/h) and air velocity (1.2 m/s) . The effect of core

downsizing on the pressure drop is presented in tables (5.1-5.4). As the core

size is reduced the pressure drop rate is increased. This is another evidence of

the importance of the predictive model which can properly reveal the effect of

the core size (L× D × H) variation on the pressure drop in air side.

5.3 Experimental Results for Gas Turbine

Changes in ambient temperature have an impact on full-load power and heat

rate, and, also on part-load performance and optimum power turbine speed.

Manufacturers typically provide performance maps that describe these

relationships for International Organization for Standardization (ISO)

conditions. The Design Point of the gas turbine engine is concerned with the

following concepts:

• Standard ambient conditions

• Improved fuel type

• Full gas turbine shaft speed


• Full power output

• Minimum heat rate • Minimum heat consumption

The percent of design considers the change ratio of performance information

to gas turbine by changing location or operation conditions with respect to

original design point, as shown in figure (1.1).

Detailed discussion of the effect of compressor inlet temperature will be

carried out in the following sections:

5.3.1 Effect of Compressor Inlet Temperature on Power Output

The power output can be increased by increasing either air mass flow or

fuel mass flow. However, increasing fuel mass flow is used according to the

application but the power output can be increased by improving ambient

conditions which attempts to make it near standard conditions. Power output

versus turbine inlet temperature at various compressor inlet temperature is

shown in fig.(5-15).The increasing of power output is due to increasing

turbine inlet temperature and also because of reducing the compressor inlet

temperature ,which increases the mass flow rate across the power turbine. An

increase in power output about (22 %) was obtained due to reducing

compressor inlet temperature from (30 - 15 °C) at gas generator speed of

(65000 RPM).

5.3.2 Effect of Compressor Inlet Temperature on Fuel Mass Flow Rate

Burning hydrocarbons (Kerosene-RT 10) with air leads to combustion gases

that have practically the same gas constant as dry air. Thus the assumption

that R = 287 J/kg.K is valid for both compressor and turbine. But, the

isentropic exponents γC and γT are in reality not constant and they change

significantly with temperature. Fuel mass flow rate versus compressor inlet

temperature at various gas generator speed is shown in fig.(5-16).The

increasing of fuel mass flow rate is due to increasing both ,gas generator

speed and compressor inlet temperature. A decrease in fuel mass flow rate


about (10 %) was obtained due to reducing compressor inlet temperature from

(30 - 15 °C) at gas generator speed of (55000 RPM).

5.3.3 Effect of Compressor Inlet Temperature on Specific Fuel

Consumption

The fuel consumption is depended not only on the combustion chamber

design, but also on required power output. The high pressure turbine work is

accelerated by forced gas mass on it's blades. That means in decreasing

compressor inlet temperature will increase gas generator speed without

increasing in fuel mass flow. Also, specific fuel consumption is down to any

power required. Specific fuel consumption versus compressor inlet

temperature at various gas generator speed is shown in fig.(5-17).The increase

in specific fuel consumption was due to the increase of both gas generator

speed and compressor inlet temperature. A decrease in specific fuel

consumption about (44 %) was obtained due to reducing compressor inlet

temperature from (30 - 15 °C) at gas generator speed of (55000 RPM).

5.3.4 Effect of Compressor Inlet Temperature on Heat Consumption

Work from a gas turbine can be defined as the product of mass flow, heat

energy in the combusted gas (Cp), and temperature differential across the

turbine. The mass flow in this equation is the sum of compressor airflow and

fuel flow. The heat energy is a function of the elements in the fuel and the

products of combustion. The power output from gas turbine can be increases

by burning more fuel which means more heat consumption. Heat

consumption versus compressor inlet temperature at various gas generator

speed is shown in fig.(5-18).The increasing of heat consumption is due to

increasing gas generator speed and compressor inlet temperature. A decrease

in heat consumption of about (8 %) was obtained due to reducing compressor

inlet temperature from (30 - 15 °C) at gas generator speed of (55000 RPM).


5.3.5 Effect of Compressor Inlet Temperature on Heat Rate

Heat Rate is the inverse of efficiency, which is an indication of the ratio

between thermal energy, resulting from the combustion process, and

mechanical energy, obtained on the power shaft. Heat rate versus turbine inlet

temperature at various compressor inlet temperature is shown in fig.(5-

19).The increasing of heat rate is due to increasing of both turbine inlet

temperature and compressor inlet temperature. A decrease in heat rate of


(30 - 15 °C) at gas generator speed of (65000 RPM).

5.3.6 Effect of Compressor Inlet Temperature on Overall Efficiency

Overall efficiency versus turbine inlet temperature at various compressor

inlet temperature are shown in fig.(5-20).The increase in overall efficiency is

due to the increase in turbine inlet temperature and also because of reduction

of compressor inlet temperature. An increase in overall efficiency about

(40 %) was obtained due to reducing compressor inlet temperature from (30 -

15 °C) at gas generator speed of (65000 RPM).

5.3.7 Effect of Compressor Inlet Temperature on Air Mass Flow Rate

In general, for lowering compressor inlet temperature the air mass flow rate

will be increased because of increasing air specific weight at low temperature.

Air mass flow rate versus compressor inlet temperature at various gas

generator speed are shown in fig.(5-21) and fig.(5-22).The increasing of air

mass flow rate is due to the increase of gas generator speed and also because

of reduction of compressor inlet temperature. An increase in air mass flow

rate and air volumetric flow rate about (15 %) and (6 %) respectively were

obtained due to reducing compressor inlet temperature from (30 - 15 °C) at

gas generator speed of (55000 RPM).

5.3.8 Effect of Compressor Inlet Temperature on Pressure Ratio

The pressure ratio of the compressor at constant speed becomes smaller

with increasing compressor inlet temperature. There will be more work (or


head) required to achieve a certain pressure rise. The increased work has to be

provided by the gas generator turbine, and is thus a loss for the power turbine.

Pressure ratio versus compressor inlet temperature at various gas generator

speed are shown in fig.(5-23).The increasing of pressure ratio is due to the

increase gas generator speed and also because of reduction of compressor

inlet temperature. An increase in pressure ratio about (8 %) was obtained due

to reducing compressor inlet temperature from (30 - 15 °C) at gas generator

speed of (55000 RPM).

5.3.9 Effect of Compressor Inlet Temperature on Power Input to

Compressor

It is known the theoretical power input to compressor is equal to the

theoretical power to high pressure Turbine. The compressor overall efficiency

from isentropic efficiency and mechanical efficiency is increasing by

decreasing compressor inlet temperature together with compression ratio is

also increasing. Therefore, by reducing compressor inlet temperature the

energy consumed to compressor decreasing and theoretical power work (same

spool with high pressure gas turbine) increasing at the same gas generator

speed as showed in Fig (5.24). On other hand the high air weight is needed

little work to compress. Power input to compressor versus gas generator speed

at various compressor inlet temperature is shown in fig.(5-24).The increasing

of power input to compressor is due to the increase gas generator speed and

also because of reduction of compressor inlet temperature. An increase in

power input to compressor about (30 %) was obtained due to reducing

compressor inlet temperature from (30 - 15 °C) for at gas generator speed

(55000 RPM).

5.3.10 Effect of Compressor Inlet Temperature on Turbine Inlet

Temperature

The two parameters that play a main role in gas turbine design are pressure

ratio and turbine inlet temperature for both gas generator and power generator


engines. When decreasing compressor inlet temperature, both the

compression ratio will be increased and air mass flow increasing too. On

other hand turbine inlet temperature will be decreased. Therefore, the power

output is increased by burning more fuel with increasing gas turbine speed as

shown in the figure (5.25). Also an increase in high pressure turbine work of


(30 - 15 °C) for constant turbine inlet temperature at (650 °C) . 5.4 Percent of Design

Design point performance is vital to the engine concept design process. The

engine configuration, cycle parameters, component performance levels and

sizes are selected to meet a given specification. This section describes this

performance input, which cannot be divorced from component design. Design

point performance must be defined before analysing of any of the other

possible operating conditions. A number of key parameters that define overall

engine performance are utilised to assess the suitability of a given engine

design to the application. These engine performance parameters are described

below to the present study which was taken from design performance

calculations for gas turbine (GT-85) :

• The full power output is 5 (kW)

• The maximum overall efficiency is 3.5 (%) • The minimum heat consumption

The maximum heat consumption can be add for gas turbine (GT-85) as

follow:

heat consumption =035.5

0=142.8 (kW)

• The minimum heat rate The minimum heat rate can be calculated for (GT-85) as follow:

heat rate=035.01 =28.56


%100)30()15( 11×

=−==

ValueDesign

TatValueCalculatedTatValueCalculatedDesignofPercent

One main concept from present study is to attempt to work or operate near

the above engine performance parameters of the design point as shown in

figures (5.26) and (5.27). The change in ratio relation to original design

(Percent of Design) is about about (15 %) increase in power output, (25 %)

increase in overall efficiency and (10 %) reduction in heat consumption, these

results were obtained with reducing compressor inlet temperature from (30 -

15 °C) at gas generator speed (55000 RPM). 5.5 Comparison between the Experimental and Theoretical Predictions of

the gas turbine engine

Figure (5-28) shows a comparison between experimental and theoretical

predictions for overall efficiency of gas turbine engine (GT-85) with variable

turbine inlet temperature. Both of the experimental data and the predicated

values have the same trend of variation showering an increase in overall

efficiency with raising turbine inlet temperature .The predicated values were

higher than that the experimental data for the whole temperature range in the

field between (550-675°C) about (18%) for the expansion ratio of (1.03).

When the expansion ratio is raised then, the discrepancy percentage will be

increased .However, the effect of irreversibility and operating with different

conditions makes this discrepancy between the experimental and theoretical

predictions for overall efficiency with the same turbine inlet temperature

increasing. 5.6 Conclusion

From literature survey and pervious discussion , it can be concluded that

there is a relationship among compressor inlet temperature, turbine inlet

temperature and gas generator speed. It is obvious that the effect of

compressor inlet temperature on gas turbine performance is reduced at high


gas generator speed. Therefore, it can be noticed that for a particular turbine

inlet temperature such as (650 °C) fig. (5.15), the power output is increased of

about (22%) due to reduction the compressor inlet temperature from (30-

15°C) at gas generator speed up to (65000 RPM). It can be also that an

increase of about (44%) when reduction compressor inlet temperature from

(30-15°C) at gas generator speed (55000 RPM).It is worth while mentioning

that the industrial gas turbine is carried out by two types: power output with

constant speed and power output with variable speed. The first type is popular

using in the power plant (Power Generation, Co-Generation), the speed is

important parameter which rely on it the current frequency, therefore the plant

worked at full speed for all time. The second type is used to feed power for

pumps and compressors (Oil, Gas). Therefore, the volumetric flow rate is

incorporated with the speed of device which is operating at higher speeds.


×

××

×

Tabl

e (5

-1) A

ir Te

mpe

ratu

re D

istri

butio

n A

long

H.E

X.D

epth

,Vw

=200

0 (L

/h),V

a=20

00 (c

fm)

Cor

e Si

ze

(LD

H)

Hea

t Ex

chan

ger

(55

3.6

3

7 ) c

m3

Hea

t Ex

chan

ger

(28.

5×

7.7

×

27)

cm

3

Hea

t Ex

chan

ger

(18

×

11

×

18

) cm

3

No.

of

R

OW

1 2 1 2 3 4 1 2 3 4 5 6

∆pa

(mba

r)

.242

.241

.242

.241

.239

.238

.242

.241

.239

.238

.237

.235

Tw (o

ut) °

C

12

12

12

12

12

12

12

12

12

12

12

12

Tw (i

n) °

C

10

10

10

10

10

10

10

10

10

10

10

10

Ta (o

ut)

°C

29.7

9

27.6

9

29.7

9

27.6

9

25.6

9

23.5

9

29.7

9

27.6

9

25.6

9

23.5

9

21.4

9

19.3

9

Ta (i

n) °C

32

29. 7

9

32

29.7

9

27.6

9

25.6

9

32

29.7

9

27.6

9

25.6

9

23.5

9

21.4

9

ha w

/m2 . °

C)

50.6

0406

50.8

1485

68.3

6579

68.6

5057

68.9

3851

69.2

2968

94.1

8447

94.5

7677

94.9

7345

95.3

7463

95.7

8031

96.1

9064

Uo(

w/m

2 . °C

)

47.1

0219

47.2

8972

62.5

5761

62.8

0377

63.0

5259

63.3

0411

84.2

4339

84.5

6969

84.8

9945

85.5

6966

85.2

3277

85.9

1026

Q

(KW

)

2.32

8

2.32

8

1.16

4

1.16

4

1.16

4

1.16

4

.776

.776

.776

.776

.776

.776


Tabl

e (5

-2) A

ir Te

mpe

ratu

re D

istri

butio

n A

long

H.E

X.D

epth

,Vw

=200

0 (L

/h),V

a=20

00 (c

fm)

Cor

e Si

ze

(LD

H)

×

××

×

Hea

t Ex

chan

ger

(55

3.6

3

7) c

m3

Hea

t Ex

chan

ger

(28.

5×

7.7

×

27)

cm

3

Hea

t Ex

chan

ger

(18×

11 ×

18

) cm

3

No.

of

R

OW

1 2 1 2 3 4 1 2 3 4 5 6

∆pa

(mba

r)

.243

.245

.243

.245

.246

.247

.243

.245

.246

.247

.248

.250

Tw (o

ut) °

C

48

48

48

48

48

48

48

48

48

48

48

48

Tw (i

n) °

C

50

50

50

50

50

50

50

50

50

50

50

50

Ta (o

ut) °

C

34.0

9

36.1

9

32.0

9

34.1

9

36.2

9

38.3

9

32.0

9

34.1

9

36.2

9

38.3

9

40.4

9

42.5

9

Ta (i

n) °C

32

34.0

9

32

32.0

9

34.1

9

36.2

9

32

32.0

9

34.1

9

36.2

9

38.3

9

40.4

9

ha (w

/m2 . °

C)

47.7

9097

48.5

9657

65.0

0219

65.5

1846

65.2

5092

65.7

6105

89.9

0861

90.2

618

89.8

9323

90.2

4413

90.5

9176

90.2

2915

Uo (

w/m

2 . °C

)

45.0

4536

45.3

1548

60.2

4699

60.6

9621

60.4

6152

60.4

6152

81.5

5405

81.8

4826

81.5

3701

81.8

2931

82.1

1871

81.8

1271

Q

(KW

)

2.28

4

2.28

4

1.14

2

1.14

2

1.14

2

1.14

2

.761

.761

.761

.761

.761

.761


Tabl

e (5

-3) A

ir Te

mpe

ratu

re D

istri

butio

n A

long

H.X

.Dep

th ,V

w=2

000

(L/h

),Va=

500

(cfm

)

Cor

e Si

ze

(LD

H)

×

××

×

Hea

t Ex

chan

ger

(55

3.6

3

7) c

m3

Hea

t Ex

chan

ger

(28.

5×

7.7

×

27)

cm

3

Hea

t Ex

chan

ger

(18

×

11 ×

18

) cm

3

No.

of

RO

W

1 2 1 2 3 4 1 2 3 4 5 6

∆pa

(m

bar)

E-02

6.32

6.29

6.32

6.29

6.25

6.22

6.32

6.29

6.25

6.22

6.18

6.15

Tw (o

ut) °

C

10.4

10.4

10.4

10.4

10.4

10.4

10.4

10.4

10.4

10.4

10.4

10.4

Tw (i

n) °

C

10

10

10

10

10

10

10

10

10

10

10

10

Ta (o

ut) °

C

29.9

9

27.8

9

29.9

9

27.8

9

25.7

8

23.6

7

29.9

9

27.8

9

25.7

8

23.6

7

21.5

7

19.4

6

Ta (i

n) °

C

32

29.9

9

32

29.9

9

27.8

9

25.7

8

32

29.9

9

27.8

9

25.7

8

23.6

7

21.5

7

ha (w

/m2 . °

C)

20.6

4679

20.6

8979

18.9

7783

19.0

5688

19.1

368

19.2

1763

26.5

0834

26.6

1876

26.7

304

26.8

433

26.9

5747

27.0

7296

Uo (

w/m

2 . °C

)

20.0

9592

20.1

3704

18.5

0125

18.5

7706

18.6

5369

18.7

3119

25.6

5959

25.7

6418

25.8

6992

25.9

7685

26.0

8497

26.1

943

Q

(KW

) .456

.456

.232

9

.232

9

.232

9

.232

9

.155

.155

.155

.155

.155

.155


Tabl

e (5

-4) A

ir Te

mpe

ratu

re D

istri

butio

n A

long

H.X

.Dep

th ,V

w=2

000

(L/h

),Va=

500

(cfm

)

Cor

e Si

ze

(LD

H)

×

××

×

Hea

t Ex

chan

ger

(55

3.6

3

7) c

m3

Hea

t Ex

chan

ger

(28.

5×

7.7

×

27)

cm

3

Hea

t Ex

chan

ger

(18

×

11

×

18

) cm

3

No.

of

R

OW

1 2 1 2 3 4 1 2 3 4 5 6

∆pa

(m

bar)

E-02

6.41

6.57

6.41

6.57

6.78

6.95

6.41

6.57

6.75

6.98

7.25

7.58

Tw (o

ut) 0 C

49.5

49.5

49.5

49.5

49.5

49.5

49.5

49.5

49.5

49.5

49.5

49.5

Tw (i

n) °

C

50

50

50

50

50

50

50

50

50

50

50

50

Ta (o

ut)

°C

34.1

2

43.0

2

34.1

2

43.0

2

54.4

6

69.8

8

34.1

2

43.0

2

54.4

6

69.8

8

78.3

7

88.1

2

Ta (i

n) °

C

32

34.1

2

32

34.1

2

43.0

2

54.4

6

32

34.1

2

43.0

2

54.4

6

69.8

8

78.3

7

ha (w

/m2 . °

C)

23.8

4225

23.4

331

25.7

3128

26.0

863

25.9

8009

26.5

3511

36.4

5073

35.8

2522

36.2

082

36.4

3759

36.5

0051

37.3

6954

Uo (

w/m

2 . °C

)

23.1

8965

22.7

9861

25.2

6978

26.0

1909

25.5

097

26.0

4461

35.5

3004

34.9

3548

35.2

9956

35.5

1754

35.5

7734

36.4

0247

Q

(KW

) .684

.684

.342

.342

.342

.342

.228

.228

.228

.228

.228

.228


Table (5-5) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37 Tw. in = 10(°C) Tw.out =12(°C) Ta .in =32(°C) Vw = 2000 l/h Va=2000cfm

ha (w/m2.°C) Q (watt) Uo (w/m2.°C) Ta out (°C) A (m2)

50.60411

50.6041

50.6041

50.6041

50.6041

50.60409

50.60408

50.60408

50.60407

50.60406

50.60406

50.60405

50.60405

50.60405

50.60404

50.60403

50.60403

50.60403

50.60402

50.60402

116.4624

116.4564

116.4504

116.4474

116.4474

116.4385

116.4326

116.4267

116.4208

116.4149

116.409

116.4032

116.3974

116.3945

116.3858

116.3801

116.3743

116.3686

116.3629

116.3572

47.09801

47.0985

47.09899

47.09924

47.09924

47.09997

47.10045

47.10094

47.10142

47.10191

47.0239

47.10287

47.10334

47.10359

47.10431

47.10478

47.10526

47.10573

47.1062

47.10667

29.89904

29.89915

29.89925

29.89931

29.89931

29.89947

29.89958

29.89968

29.89979

29.89989

29.9

29.9001

29.90021

29.90026

29.90042

29.90052

29.90063

29.90073

29.90083

29.90093

.1187113

.1192767

.1198476

.1201254

.1201254

.1210062

.1215941

.1221879

.1227876

.1233933

.1240051

.1246231

.1252474

.1255509

.1265153

.1271591

.1278096

.1284669

.1291311

.1298023


Table (5-6) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37 Tw. in= 10(°C) Tw.out =12(°C) Ta .in =32(°C) Vw = 2000 l/h Va=2000cfm

ha (w/m2. °C) Q (watt) Uo (w/m2. °C) Ta out (°C) A (m2)

50.81489

50.81488

50.81488

50.81488

50.81488

50.81487

50.81486

50.81486

50.81486

50.81485

50.81485

50.81484

50.81483

50.81483

50.81482

50.81482

50.81482

50.81481

50.8148

50.8148

116.4624

116.4564

116.4504

116.4474

116.4474

116.4385

116.4326

116.4267

116.4208

116.4149

116.409

116.4032

116.3974

116.3945

116.3858

116.3801

116.3743

116.3686

116.3629

116.3572

47.28552

47.28601

47.2865

47.28675

47.28675

47.28749

47.28798

47.28846

47.28896

47.28944

47.28992

47.29041

47.29089

47.29113

47.29185

47.29234

47.29281

47.29329

47.29376

47.29423

27.799

27.7991

27.79921

27.79926

27.79926

27.79943

27.79953

27.79964

27.79975

27.79985

27.79996

27.80006

27.80017

27.80022

27.80038

27.80048

27.80058

27.80069

27.80079

27.80089

. 129798

.131507

.132205

.1329107

.133253

.133253

.1343451

.1350742

.1358114

.1365567

.1373105

.1380727

.1388436

.1396233

.140001

.1412096

.1420166

.142833

.1436589

.1444946


Table (5-7) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37 Tw. in= 50(°C) Tw.out =48(°C) Ta .in =32(°C) Vw = 2000 l/h Va=2000cfm


48.69674

48.69672

48.69672

48.69672

48.69672

48.6967

48.69669

48.69669

48.69668

48.69667

48.69666

48.69665

48.69664

48.69664

48.69663

48.69662

48.69662

48.69661

48.6966

48.69659

114.1267

114.1362

114.1457

114.1504

114.1504

114.1645

114.1739

114.1833

114.1926

114.2018

114.2111

114.2203

114.2295

114.234

114.2477

114.2567

114.2657

114.2747

114.2837

114.2926

45.37363

45.37796

45.38214

45.38418

45.38418

45.39007

45.39383

45.39747

45.40099

45.40438

45.0767

45.41086

45.41394

45.41545

45.41983

45.42263

45.42537

45.428

45.43056

45.43305

34.09845

34.09862

34.0988

34.09888

34.09888

34.09914

34.09932

34.09949

34.09966

34.09983

34.1

34.10017

34.10034

4.10042

34.10067

34.10084

34.10101

34.10117

34.10133

34.1015

.1494603

.1503539

.1512587

.1516965

.1516965

.1531029

.1540425

.1549942

.1559581

.1569344

.1579234

.1589252

.1599401

.1604302

.1620101

.1630657

.1641354

.1652195

.1663181

.1674317


Table (5-8) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37 Tw. in = 50(°C) Tw.out =48(°C) Ta .in =32(°C) Vw = 2000 l/h Va=2000cfm


48.49659

48.49654

48.49653

48.49652

48.49652

48.49652

48.4965

48.4965

48.49649

48.49648

48.49648

48.49647

48.49646

48.49645

48.49644

48.49644

48.49643

48.49642

48.49641

48.49641

48.4964

114.1226

114.1267

114.1362

114.1457

114.1504

114.1504

14.1645

114.1739

114.1833

114.1926

114.2018

114.2111

114.2203

114.2295

114.234

114.2477

114.2567

114.2657

114.2747

114.2837

114.2926

45.19305

45.19515

45.19946

45.20361

45.20564

45.20564

45.21149

45.21524

45.21885

45.22234

45.22573

45.229

45.23216

45.23523

45.23672

45.24109

45.24387

45.24658

45.2492

45.25175

45.25422

36.19837

36.19855

36.19872

36.19881

36.19881

36.19907

36.19924

36.19941

36.19958

36.19976

36.19992

36.2001

36.20026

36.20035

36.2006

36.20076

36.20093

36.2011

36.20126

36.20142

36.20128

.1674317

.1714795

.172653

.1738432

.1744162

.1744162

.1762752

.1775177

.1787783

.1800573

.1813552

.1826723

.1840092

.1853661

.1860176

.1881417

.1895614

.191003

.192467

.1939538

.195464


Table (5-9) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37 Tw. in = 10(°C) Tw.out=10.4(°C) Ta .in =32(°C) Vw =2000 l/h Va=500cfm


20.64677

20.64677

20.64677

20.64676

116.4624

116.4564

116.4504

116.4474

20.09579

20.09587

20.09595

20.09598

29.89989

29.9

29.90011

29.90016

.2781911

.2795178

.2808575

.2815096

Table (5-10) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37 Tw. in = 10(°C) Tw.out=10.4(°C) Ta .in =32(°C) Vw =2000 l/h Va=500cfm


20.73277

20.73277

20.73277

20.73276

116.4624

116.4564

116.4504

116.4474

20.17804

20.17811

20.17819

20.17823

27.79996

27.80007

27.80018

27.80023

.3081484

.3097861

.3114415

.3122447

Table (5-11) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37 Tw. in = 50(°C) Tw.out=49.5(°C) Ta .in =32(°C) Vw =2000 l/h Va=500cfm


19.86854

19.86854

19.86854

19.86854

114.1267

114.1362

114.1457

114.1504

19.32309

19.32381

19.32451

19.32485

34.09974

34.09991

34.10009

34.10017

.3509701

.353089

.3552339

.356272


Table (5-12) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55×3.6×37 Tw. in = 50(°C) Tw.out=49.5(°C) Ta .in =32(°C) Vw =2000 l/h Va=500cfm


19.52759

19.52758

19.52758

19.52758

114.1267

114.1362

114.1457

114.1504

18.99886

18.99919

18.99919

19.00015

43.02019

43.02028

43.02028

43.02053

.3620837

.3668006

.3668006

.3632253


• Entering water at 10 °C Fig (5. 1a)

• Entering water at 50 °C Fig (5. 1b)

Fig (5.1) Comparison between the Experimental and Present Model for the Effect of Air Velocity on Heat Load at Water Flow Rate 2000 (L/h)

0

1

2

3

4

5

6

0 0.5 1 1.5 2 2.5 3 3.5 4 4

Air Velocity (m/s)

Hea

t Loa

d (k

W)

.5 5

Simulation

Experiment

0

1

2

3

4

5

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5

Air Velocity (m/s)

Hea

t Loa

d (k

W)

Simulation

Experiment




Fig (5. 2) Comparison between the Experimental and Present Model for the Effect of Air Velocity on Heat Transfer Coefficient ha, at Water Flow Rate 2000 (L/h)

0

10

20

30

40

50

60

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5

Air Velocity (m/s)

Hea

t Tra

nsfe

r Coe

ffici

ent. h

a (W

/m2.

C)

SimulationExperiment

0

10

20

30

40

50

60

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5

Air Velocity (m/s)

Hea

t Tra

nsfe

r Coe

ffici

ent .h

a (W

/m2.

C)

5

SimulationExperiment


50.45 50.5 50.55 50.6 50.65 50.7 50.75 50.8 50.85

1

2

3

4

Slic

e Nu

mbe

r (B

ar T

ube)

Heat Transfer Coefficient ha.(w/m2.c)

ROW 1ROW 2

• Entering water at 10 °C

Fig (5. 3a)

48.35 48.4 48.45 48.5 48.55 48.6 48.65 48.7 48.75

1

2

3

4

Slic

e N

umbe

r (B

ar T

ube)


ROW 1ROW 2


Fig (5. 3b)

Fig (5. 3) Variation Heat Transfer Coefficient (ha) a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate 2000 cfm


20.6 20.62 20.64 20.66 20.68 20.7 20.72 20.74

1

2

3

4

Slic

e Nu

mbe

r (B

ar T

ube)


ROW 1ROW 2


Fig (5. 4a)

19.4 19.4 19.5 19.5 19.6 19.6 19.7 19.7 19.8 19.8 19.9 19.9

1

2

3

4

Slic

e N

umbe

r


ROW 1ROW 2


Fig (5. 4b)

Fig(5. 4) Variation Heat Transfer Coefficient( ha )a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate 500 cfm


93

93.5

94

94.5

95

95.5

96

96.5

1 2 3 4 5 6

Row Number

Hea

t Tra

nsfe

r Coe

ffici

ent (

W/m

2.c)

Fig(5. 5) Variation Heat Transfer Coefficient (ha) a long Heat Exchanger Depth at Water Flow Rate 2000 (L/h), Water Entering Temp.

10 °C, Air Flow Rate 2000 cfm




Fig (5. 6) Comparison between the Experimental and Present Model for the Effect of Air Velocity on Overall Heat Transfer Coefficient at Water Flow Rate 2000 (L/h)

0

5

10

15

20

25

30

35

40

45

50

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5

Air Velocity (m/s)

Ove

rall

Hea

t Tra

nsfe

r Coe

ffici

ent (W

/m2.

C) Simulation

Experiment

0

5

10

15

20

25

30

35

40

45

50

0 1 2 3 4

Air Velocity (m/s)

Ove

rall

Hea

t Tra

nsfe

r Coe

ffici

ent (

W/m

2.C

)

5

Simulation

Experiment




Fig(5. 7) Variation Overall Heat Transfer Coefficient a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate

2000 cfm

45.1 45.15 45.2 45.25 45.3 45.35 45.4

1

2

3

4

Slic

e N

umbe

r

Overall Heat Transfer Coefficient (w/m2.c)

ROW 1ROW 2

47 47.05 47.1 47.15

1

2

3

4

Slic

e Nu

mbe

r (B

ar T

ube)

Overall Heat Transfer

47.2 47.25 47.3 47.35

Coefficient (w/m2.c)

ROW 1ROW 2



Fig (5. 8a)


Fig(5. 8) Variation Overall Heat Transfer Coefficient a long Heat Exchanger

Height at Water Flow Rate 2000 (L/h), Air Flow Rate 500 cfm

18.8 18.9 19 19.1 19.2 19.3 19.4

1

2

3

4

Slic

e N

umbe

r

Overall Heat Transfer Coefficient (w/m2.c)

ROW 1ROW 2

20.04 20.06 20.

1

2

3

4

Slic

e Nu

mbe

r

Overall

08 20.1 20.12 20.14 20.16 20.18 20.2

Heat Transfer Coefficient (w/m2.c)

ROW 1ROW 2


Fig(5. 9) Variation Overall Heat Transfer Coefficient a long Heat

Exchanger Depth at Water Flow Rate 2000 (L/h),Water Entering Temp. 10 °C, Air Flow Rate 2000 cfm

83

83.5

84

84.5

85

85.5

86

86.5

1

Ove

rall

Hea

t Tra

nsfe

r Coe

ffici

ent (W

/m2.

c)

2 3 4 5 6

Row Number




Fig (5. 10b) Fig (5. 10) Comparison between the Experimental and Present Model for the Effect

of Air Velocity on Air Exit Temperature at Water Flow Rate 2000 (L/h)

27.55

27.6

27.65

27.7

27.75

27.8

27.85

27.9

27.95

0 0.5

Air

Exit

Tem

pera

ture

(C

1 1.5 2 2.5 3 3.5 4 4.5 5

Air Velocity (m/s)

)

Simulation

Experiment

35

36

37

38

39

40

41

42

43

44

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5

Air Velocity (m/s)

Air

Exit

Tem

pera

ture

(C )

Simulation

Experiment




Fig (5. 11) Variation Exit Air Temperature (°C) a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate

2000 cfm

26.5 27 27.5 28

1

2

3

4

Slic

e N

umbe

r (B

ar T

ube)

Exit Air T

28.5 29 29.5 30 30.5

emperature (c)

ROW 1

ROW 2

33 33.5 34 34.5 35 35.5 36 36.5

1

2

3

4

Slic

e N

umbe

r (B

ar T

ube)

Exit Air Temperature (c)

ROW 1

ROW 2




Fig(5. 12) Variation Exit Air Temperature) a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate

500 cfm

30 32.5 35 37.5 40 42.5 45

1

2

3

4

Slic

e N

umbe

r (B

ar T

ube)

Exit Air Temperature (c)

Row 1

Row 2

26.5 27 27.5 28 28

1

2

3

4

Slic

e N

umbe

r (B

ar T

ube)

Exit Air Temp

.5 29 29.5 30 30.5

erature (c)

ROW 1

ROW 2


0

5

10

15

20

25

30

35

1 2

Air

Exit

Tem

pera

ture

(c)

3 4 5 6

No.of Row Row Number

Fig(5. 13) Variation Air Exit Temperature a long Heat Exchanger Depth atWater Flow Rate 2000 (L/h), Water Entering Temp. 10 °C,2000 cfm

10

13

1619

22

25

28

31

3437

40

43

0.6 0.7 0.8 0.9 1 1.1Aspect Ratio (H/L)

Pres

sure

Dro

p in

Air

Side

(mba

r)*10

-2

Water Inlet Temperature 10 C

Water Inlet Temperature 50 C

Fig (5. 14) The Effect of Aspect Ratio (H/L) with Different Core Size (L × D × H) on the Pressure Drop in Air Side,

Water Flow Rate 2000 (L/h) Air Flow Rate 500 cfm


0

0.5

1

1.5

2

2.5

610 620 630 640 650 660

Turbine Inlet Temperature (C )

Pow

er O

utpu

t (kW

)

T1=15 C

T1=30 C

Fig (5.15) The Effect of Turbine Inlet Temperature on the Power Output With Variable Compressor Inlet Temperature, at Rang gas Generator Speed (50000-65000 RPM)

2

2.1

2.2

2.3

2.4

2.5

2.6

10 15 20 25 30 35

Compressor Inlet Temperature (C )

Fuel

Mas

s Fl

ow R

ate

(g/s

)

45000 RPM

55000 RPM

Fig (5.16) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Fuel Mass Flow Rate


Fig (5.17) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed ) on the Specific Fuel Consumption

Fig (5.18) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Heat Consumption

96

98

100

102

104

106

108

110

10 15 20 25 30 35


Hea

t Con

sum

ptio

n (k

W)

45000 RPM

55000 RPM

0.6

0.8

1

1.2

1.4

1.6

1.8

2

10 15 20

Compressor

Spec

ific

Fuel

Con

sum

ptio

n (g

/kW

.s)

25 30 35

Inlet Temperature (C )

45000 RPM

55000 RPM


0

10

20

30

40

50

60

70

80

90

100

610 620 630 640 650 660

Turbine Inlet Temperature (C)

Heat

Rat

e (k

j/kW

.s)

T1=15 C

T1=30 C

Fig (5.19) The Effect of Turbine Inlet Temperature (with Variable Compressor Inlet Temperature) on the Heat Rate, at Rang Gas Generator (50000-65000 RPM)

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

610 620 630 640 650 660

Turbine Inlet Temperature (C )

Ove

rall

Effic

ienc

y (%

)

Ti=15 C T1=30 C

Fig (5.20) The Effect of Turbine Inlet Temperature (with Variable Compressor Inlet Temperature) on the Overall Efficiency (%),at Rang Gas Generator Speed

(50000-65000 RPM)


0.04

0.05

0.06

0.07

0.08

0.09

0.1

0.11

0.12

0.13

0.14

10 15 20 25 30 35


Air

Mas

s Fl

ow R

ate

(kg/

s)

45000 RPM

55000 RPM

Fig (5.21) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Air Mass Flow Rate

200

250

300

350

400

450

10 15 20 25 30 35Compressor Inlet Temperature (C )

cfm

45000 RPM 55000 RPM

Fig (5.22) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Air Flow Rate


1.14

1.19

1.24

1.29

1.34

1.39

1.44

10 15 20 25 30 3


Com

pres

sion

Rat

io

5

45000 RPM

55000 RPM

Fig (5.23) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Compression Ratio

0

1

2

3

4

5

6

10 15 20 25 30 35


Com

pres

sor W

ork

(kW

)

45000 RPM

55000 RPM

Fig (5.24) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Compressor Work


Fig (5.25) The Effect of Turbine Inlet Temperature (°C) on the HP.Turbine Work with variable Compressor Inlet Temperature (°C) ,

at Rang Gas Generator Speed (50000-65000 RPM)

0

1

2

3

4

5

6

7

8

9

610 620 630

Turbine Inlet

HP.

Turb

ine

Wor

k (k

W)

640 650 660

Temperature (C )

T1=15 C

T1=30 C


0

10

20

30

40

50

60

70

80

10 15 20 25 30 35


Perc

ent o

f Des

ign

POWER OUTPUT

OVERALL EFFICIENCY

HEAT CONSUMPTION

Fig (5.26) The Influence Compressor Inlet Temperature (°C )on Gas Turbine Performance at (Gas Generator Speed 45000 RPM)

0

10

20

30

40

50

60

70

80

10 15 20 25 30 35


Perc

ent o

f Des

ign

POWER OUTPUT OVERALL EFFICIENCY

HEAT CONSUMPTION

Fig (5.27) The Influence Compressor Inlet Temperature (°C )on Gas Turbine Performance at(Gas Generator Speed 55000 RPM)


1.2

1.4

1.6

1.8

2

2.2

2.4

2.6

2.8

500 525 550 575 600 625 650 675 700

HP.Turbine Inlet Temperature (C)

Ove

rall

Effic

ienc

y (%

)

Theoretical,ExpansionRatio=1.03

Experimental,ExpansionRatio=1.03

Fig (5.28) Comparison between the Experimental and Theoretical Predictions of the gas turbine engine (GT-85)

CHAPTER SIX

Conclusions & Recommendations

Chapter Six ..................................................................... Conclusions &Recommendations

6.1 Conclusions:

The main goal of this work was to examine the possibility of increasing the

power output and overall efficiency of gas turbine engine. Also, it was aimed

to reduce heat rate. This was accomplished by reducing compressor inlet

temperature.

The following major conclusions are drawn from this work:

1. A computational model for design of air-cooled heat exchanger has been

developed. For validation of the model, the heat load performance of a typical

air-cooled heat exchanger was simulated over wide ranges of the air velocity

and water flow rate .These were compared with the experimental data

provided by experimental work. The specifications of the test heat exchanger

are given in [Table (3.1)] shows a good agreement with the predicted

operating conditions of the present model. The maximum discrepancy

between the experimental data and calculated results for overall heat transfer

coefficient and heat load was about (5%) for the given range of the simulated

conditions. The model evaluation is obtained by checking its validity against

experimental results obtained in the present work.

2. Increasing the ambient temperature lowers the density of the compressor

inlet air. Thus reducing the mass flow through the turbine, and therefore will

be reduced the power output. When the volume flow remains approximately

constant, the mass flow will increase with decreasing temperature and will be

decreased with increasing temperature. Therefore, when the compressor inlet

temperature reduced from (30 °C) to (15 °C), the percent of design increase

up to (15%) of the power output ,and ,the overall efficiency increased up to

(25%) ,while the heat consumption reduction was about (10%).

3. The prediction of the gas turbine performance by using non-dimensional

analysis is a good method which has been used in modern computer codes.

The results revealed that the discrepancy percentage between theoretical and

experimental prediction for gas turbine performance was about (18%).

Chapter Six ..................................................................... Conclusions &Recommendations

4. There is a relationship among the compressor inlet temperature, gas

generator speed and turbine inlet temperature. At the higher speeds the

increase in percentage of power output is lower than that of lower speeds

when reducing the compressor inlet temperature.

6.2 Recommendations:

1. At extreme high climate temperatures, the cooling of the ambient

temperature technique is preferable and worth doing where the benefits are

high. On the other hand, the benefits are low in regions with mild or low-

round temperatures, and that what has been detected through the experimental

data.

2. Using the computational model program to design and study the effect of

varying different design parameters: such as (number of rows, fin spacing, air

velocity, velocity in tube, number tubes. For air-cooled heat exchanger and

entering these parameters as data to intake of gas turbine is good idea,

therefore by using different of inlet cooling system like absorption

refrigeration cycle, evaporative cooling especially in low relative humidity

regions is benefits.

3. Modifying the computational model program to investigate the gas turbine

performance and study the effect of compressor inlet temperature on the

performance using non-dimensional analysis, this requires further future

investigation by other researchers.

References

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SAE.Paper , 2006-01-0726, The University of Michigan.

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the Prediction of the Thermal Performance for Cross Flow Air Cooled Heat

Exchangers with a New Air Side Thermal Correlation ", Journal of

Engineering and Development, vol. 12,No.3, 2008.

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inlet air cooling for gas-turbine cogeneration plants”, Proceedings of the

International Gas Turbine and Aero-engine Congress and Exposition,

Houston Texas, 5–8 June 1995.

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Performance", M.Sc.Thesis, AL-Mustansiria University College of

Engineering, 1996.

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compressor inlet air refrigeration.”, Journal of Engineering for Gas Turbines

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21,2002.

16- Donald C. Erickson and Icksoo Kyung, “Aqua Absorption Turbine Inlet

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Exposition, Draft IMECE2003-42870,pp.113-115.

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Performance with Regeneration", M.Sc.Thesis, AL-Mustansiria University

College of Engineering, 2004.

18-Benjalool,A.,"Evaluation of Performance Deterioration on Gas Turbines

due to Compressor Fouling",MSc Thesis,School of Engineering,Academic

Year, pp.46-48,2006,.

19-Tony Giampaolo, " Gas Turbine Handbook: Principles and Practices" 3rd

Edition, Published by the Fairmont Press, Inc.700 Indian Trail, 2006.

20-Kuamit, A.A.," Design of Combined Cycle Power Plant and Air Cooling

System", Ph.D,Thesis, AL-Mustansiria University College of Engineering,

2006.

21-Tarrad, A. H., and Mohammed, A. G., " A Mathematical Model for

Thermo-Hydraulic Design of Shell and Tube Heat Exchanger Using a Step by

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Gas Turbine Temperatures sensors" Watllow electric Manufacturing

Company,2001.

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Turbine " , Gilkes, Kendal, England, 1978.

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Edition, 143 pp, 2004.

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Theory", 4th ed, 1996.

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McGraw-Hill Book Company, 1984.

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Approach", John Wiley and Sons, New York, 1997.

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Yourk,1989.

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,

Appendix (A) Experimental Work

Data Tables

The experimental Data for gas turbine: Table (A-1) Ambient Temperature: 19.5 (°C) Ambient Pressure: 1.0133 bar Compressor Inlet Temperature: 15 (°C)

Engine Parameters

Gas Generator Speed

(45000 RPM)

Gas Generator Speed

(55000 RPM)

T1 (°C) 15 15

Δp (mmwg) 32 43

Vf (L/h) 5 6

P2 (bar) 0.23 0.32

T 2 (°C) 46 52

T3 (°C) 561 577

P3 (bar) 0.22 0.31

T4 (°C) 519 531

P4 (mmbar) 35 45

T5 (°C) 485 492

Δp 4/5 (mmHg) 25 30

Torque (N.m) 0.92 0.93

Power Turbine

Speed (RPM)

14400 16300

Table (A-2) Ambient Temperature: 25 (°C) Ambient Pressure: 1.0133 bar Compressor Inlet Temperature: 20 (°C)

Engine Parameters

Gas Generator Speed

(45000 RPM)

Gas Generator Speed

(55000 RPM)

T1 (°C) 20 20

Δp (mmwg) 30 40

Vf (L/h) 4.9 5.7

P2 (bar) 0.25 0.33

T 2 (°C) 57 63

T3 (°C) 588 604

P3 (bar) 0.23 0.31

T4 (°C) 546 558

P4 (mmbar) 37 50

T5 (°C) 506 517

Δp 4/5 (mmHg) 25 35

Torque (N.m) 1.3 1.36

Power Turbine

Speed (RPM)

17100 21900

Table (A-3) Ambient Temperature: 29 (°C) Ambient Pressure: 1.0133 bar Compressor Inlet Temperature:25 (°C)

Engine Parameters Gas Generator Speed

(45000 RPM)

Gas Turbine Speed

(55000 RPM)

T1 (°C) 25 25

Δp (mmwg) 40 50

Vf (L/h) 4.7 5

P2 (bar) 0.29 0.34

T 2 (°C) 56 63

T3 (°C) 618 635

P3 (bar) 0.25 0.31

T4 (°C) 551 565

P4 (mmbar) 40 52

T5 (°C) 508 523

Δp 4/5 (mmHg) 25 33

Torque (N.m) 0.66 0.85

Power Turbine

Speed (RPM)

17900 22000

Table (A-4) Ambient Temperature: 29(°C) Ambient Pressure: 1.0133 bar Compressor Inlet Tempertaur:30 (°C)

Engine Parameters Gas Turbine Speed

(45000 RPM)

Gas Turbine Speed

(55000 RPM)

T1 (°C) 30 30

Δp (mmwg) 39 59

Vf (L/h) 5 6

P2 (bar) 0.31 0.45

T 2 (°C) 64 73

T3 (°C) 631 650

P3 (bar) 0.28 0.35

T4 (°C) 552 578

P4 (mmbar) 38 59

T5 (°C) 511 531

Δp 4/5 (mmHg) 42 57

Torque (N.m) 0.71 1.9

Power Turbine

Speed (RPM)

11800 21600

Table (A-5) Ambient Pressure:1.0133 bar Turbine Inlet Temperature: 586 (°C)

Engine Parameters

Running with

Compressor Inlet

Temperature 15 (°C)

Running with

Compressor Inlet


T3 (°C) 586 586

Gas Generator ( RPM) 55000 50000

Δp (mmwg) 43 47

Vf (L/h) 6.2 5.8

P2 (bar) 0.32 0.35

T1 (°C) 15 22

T2 (°C) 52 60

P3 (bar) 0.3 0.32

T4 (°C) 539 534

P4 (mbar) 50 46

T5 (°C) 499 496

Δp 4/5 (mmHg) 32 30

Torque (N.m) 1.09 0.74

Power Turbine(RPM) 18800 16500


Engine Parameters

Running with

Compressor Inlet


Running with

Compressor Inlet


T3 (°C) 590 590


Δp (mmwg) 39 33

Vf (L/h) 5.5 5

P2 (bar) 0.3 0.25

T1 (°C) 18 22

T2 (°C) 52 53

P3 (bar) 0.27 0.21

T4 (°C) 544 549

P4 (mbar) 45 40

T5 (°C) 504 509

Δp 4/5 (mmHg) 30 27

Torque (N.m) 1.16 1.33



Engine Parameters

Running with

Compressor Inlet


Running with

Compressor Inlet


T3 (°C) 620 620


Δp (mmwg) 44 40

Vf (L/h) 5.2 4.7

P2 (bar) 0.34 0.29

T1 (°C) 20 30

T2 (°C) 58 60

P3 (bar) 0.31 0.25

T4 (°C) 550 554

P4 (mbar) 50 40

T5 (°C) 500 510

Δp 4/5 (mmHg) 32 25

Torque (N.m) 0.69 0.66


Table (A-8) Ambient Pressure: 1.0133 bar Turbine Inlet Temperature: 625 (°C)

Engine Parameters

Running with

Compressor Inlet


Running with

Compressor Inlet

Temperature 30(°C)

T3 (°C) 625 625


Δp (mmwg) 53 38

Vf (L/h) 5.6 5.2

P2 (bar) 0.41 0.28

T1 (°C) 15 30

T2 (°C) 61 64

P3 (bar) 0.34 0.26

T4 (°C) 550 549

P4 (mbar) 55 35

T5 (°C) 508 500

Δp 4/5 (mmHg) 36 22

Torque (N.m) 0.59 0.36


Table (A-9) Ambient Pressure: 1.0133 bar Turbine Inlet Temperature: 635 (°C)

Engine Parameters

Running with

Compressor Inlet


Running with

Compressor Inlet


T3 (°C) 635 635


Δp (mmwg) 41 39

Vf (L/h) 5.2 4.7

P2 (bar) 0.34 0.31

T1 (°C) 25 29

T2 (°C) 62 61

P3 (bar) 0.31 0.26

T4 (°C) 564 566

P4 (mbar) 52 45

T5 (°C) 522 523

Δp 4/5 (mmHg) 33 29

Torque (N.m) 0.85 0.79


The experimental Data for heat exchanger Table (A-10)

Vw=2000 L/h , Va=2000 Cfm

Twi (°C) Two (°C) Tai (°C) Tao (°C) 50 48 32 36.7 40 38 32 35 20 22 32 27.7 15 17 32 27.5 10 12 32 28 Table (A-11)

Vw=2000 L/h , Va=1000 Cfm

Twi (°C) Two (°C) Tai (°C) Tao (°C) 50 49 32 36.2 40 38.9 32 35.9 20 21 32 28.2 15 15.9 32 28 10 10.9 32 27 Table (A-12) Vw=2000 L/h , Va=500 Cfm

Twi (°C) Two (°C) Tai (°C) Tao (°C) 50 49.5 32 38.4 40 39.4 32 36.5 20 21.4 32 28 15 15.4 32 27.9 10 10.4 32 26.9

Table (A-13)

Vw=3000 L/h , Va=2000 Cfm

Twi (co) Two (co) Tai (co) Tao (co) 50 48.5 32 36.3 40 39.7 32 33 20 21.5 32 27.9 15 16.5 32 27.7 10 11.5 32 27 Table (A-14) Vw=3000 L/h , Va=1000 Cfm

Twi (co) Two (co) Tai (co) Tao (co) 50 49.4 32 36 40 39.3 32 35 20 20.8 32 27 15 15.9 32 26.5 10 11.7 32 26.8 Table (A-15)

Vw=3000 L/h , Va=500 Cfm

Twi (co) Two (co) Tai (co) Tao (co) 50 49.8 32 35.8 40 39.7 32 35.1 20 20.4 32 27 15 15.4 32 26.9 10 10.4 32 26.4

Appendix (B) Flow Charts and

Computer Programs

Flow Chart for Theoretical heat exchanger Performance Prediction(CPHE)

Start

Input Surface Characteristics & Inlet Operation Conditions

for Both Fluids

Choose the Velocity in Tubes Calculate the No. of Slices from Water

Temperature conditions

Calculate the air mass flow rate for first row, Eq (4-6)

Calculate the Heat Load for First Slice & First Row,EQ.(3-9)

Assume the Air Exit Temperature for First Row

Calculate (No. of Tubes / Row) ,EQ.(3-1) & Calculate Length of H.EX, EQ.(3-2) & Calculate Depth of H.EX, EQ.(3-3)

IF (Twi >Two)

YES ∆Twater = 0.1, nw=0.3,na=0.4,F=1

NO

∆Twater= 0.1, nw=0.4 ,na=0.3,F=-1

B A

Calculate the Heat Load for First Slice

Correct the Fluids Properties

Calculate the Air Exit Temperature for First Slice

Calculate the Correct Air Exit Temperature for First Slice

Assume the Air Velocity Over Tubes.

Calculate the Height of Heat Exchanger

Calculate the Air Mass Flow Rate for First Row

J=1, Calculate the No. of Slices from Water Temp.

Conditions

I=.1, Divided the Air Mass Flow Rate per No. of Slices

Calculate the Fluids Properties from Inlet Conditions

B A

D C E

Calculate the Cross Sectional Area Temperature, Hydraulic Diameter, Reynolds No., Nusslat No. .

Calculate the Overall Heat Transfer Coffi.,

Calculate the Height of Heat Exchanger Calculate Error from Calculating Value &

Assuming Value.

Repeat Process for all Slices (Ni)

Calculate the Mean Air Exit Temp. Calculate the Error from Calculating Value &

Assuming Value

Repeat the process for all Rows (Nj)

Calculate Heat Exchanger Performance

Calculate Heat Exchanger Physical Characteristics

NO YES NO YES

IF Error <=0.001

IF Error <=0.001

C D E

End

Flow Chart for Theoretical Gas Turbine Performance Prediction(CPTGTP)

STRAT

INPUT DESIGN INFORMATION Tmax,T1,Rc,Rt,f,H.V,ηc,ηt

Ømax=Tmax/T1,M=0

For Ø=Ømax to 1 step -1

T3=Ø*T1

Za=γa/(γa-1), Xcm=(Rc)^(1/Za)

For Xc=Xcm to 1.1 step -.05 NDCW=(Xc-1)/(ηc*(Ø-1))

Zg=γg/(γg-1), Xtm=(Rt)^(1/Zg) ,n=cpa/cpg

NDTW=((1+f)*ηt*(1-1/Xt))/((1-1/Ø)*n) ηo=NDTW/(H .v*(1+ƒ))

PRINT T3,Ø,Xc,Xt,ηo,NDCW NDTW

NEXT Xc , NEXT Xt , NEXT Ø

END

Appendix (C) Gas Turbine parameter

Groups

(C-1

)

(C-2

)

Gas Turbine,Finned Heat Exchanger

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