Gas Turbine,finned Heat Exchanger
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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
ﻦ اﻟ َّﺮﺣِﻴ ِﻢ ِﺑﺴْ ِﻢ اﻟَّﻠ ِﻪ اﻟ َّﺮﺣْ َﻤ ِ
ن ب وَاﻟْﻤِﻴﺰَا َ ت وَأَﻧْﺰَﻟْﻨَﺎ َﻣ َﻌ ُﻬ ُﻢ اﻟْ ِﻜﺘَﺎ َ ﺳ َﻠﻨَﺎ ﺑِﺎﻟْ َﺒ ِّﻴﻨَﺎ ِ َﻟ َﻘﺪْ أَرْﺳَﻠْﻨَﺎ ُر ُ ﺤﺪِﻳ َﺪ ﻓِﻴ ِﻪ ﺑَﺄْسٌ ﺷَﺪِﻳﺪٌ َو َﻣﻨَﺎ ِﻓ ُﻊ ﻂ وَأَﻧْﺰَﻟْﻨَﺎ اﻟْ َ س ﺑِﺎﻟْ ِﻘﺴْ ِ ِﻟ َﻴﻘُﻮ َم اﻟ َﻨّﺎ ُ ن اﻟَّﻠ َﻪ َﻗ ِﻮيٌّ ﺐ ِإ َّ ﺳ َﻠ ُﻪ ﺑِﺎﻟْ َﻐﻴْ ِ ﺼ ُﺮ ُﻩ َو ُر ُ س َو ِﻟ َﻴﻌْ َﻠ َﻢ اﻟَّﻠ ُﻪ َﻣﻦْ َﻳﻨْ ُ ﻟِﻠ َﻨّﺎ ِ ﻋَﺰِﻳﺰٌ ﺻﺪق اﷲ اﻟﻌﻈﻴﻢ
ﺳﻮرة اﻟﺤﺪﻳﺪ اﻵﻳﺔ )(25
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:
Signature:
Name:
Name:
Date:
Date:
In view of the available recommendation I forward this thesis for debate by the examining committee Signature: Name: Date: (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. Signature: Name: Date: (Chairman) Signature:
Signature:
Name:
Name:
Date:
Date:
(Member)
(Member)
Signature:
Signature:
Name:
Name:
Date:
Date:
(Supervisor)
(Supervisor)
Approved by the Dean of College of Engineering Signature: Name: Date:
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 twoshaft 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 aircooled 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 :
8
Literature Survey
2.1 General
9
2.2 The Previous Work
9
2.2.1 Modeling and Simulation of Air Cooled Heat Exchanger 2.2.2 Performance Improvement of the Gas Turbine Engine
9 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
58
Performance 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
62
minimum ratio (Ø) 4.7.2 The effect of compressor inlet temperature on the expansion
64
ratio 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
76
Rate 5.3.3 Effect of Compressor Inlet Temperature on Specific Fuel
77
Consumption 5.3.4 Effect of Compressor Inlet Temperature on Heat
77
Consumption 5.3.5 Effect of Compressor Inlet Temperature on Heat Rate
78
5.3.6 Effect of Compressor Inlet Temperature on Overall
78
Efficiency 5.3.7 Effect of Compressor Inlet Temperature on Air Mass Flow
78
Rate 5.3.8 Effect of Compressor Inlet Temperature on Pressure Ratio
78
5.3.9 Effect of Compressor Inlet Temperature on Power Input to
79
Compressor 5.3.10 Effect of Compressor Inlet Temperature on Turbine Inlet
79
Temperature 5.4 Percent of Design
80
5-5 Comparison between the Experimental and Theoretical
81
Predictions of the gas turbine engine 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 Figure (3.3.a) Top View of Heat Exchanger Geometry
24
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
25
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
67
the Theoretical Overall Efficiency ,Expansion Ratio=1.12 Figure(4.13) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on
67
the Theoretical Overall Efficiency ,Expansion Ratio=1.08 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) 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
94
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
95
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
83
0
(l/h),Va=2000 (cfm) , Tw (in)=10 C Table (5.2) Air Temperature Distribution Along H.EX.Depth,Vw=2000 0
(l/h),Va=2000 (cfm), Tw (in)=50
84
C
Table (5.3) Air Temperature Distribution Along H.EX.Depth,Vw=2000
85
0
(l/h),Va=500 (cfm), Tw (in)=10 C Table (5.4) Air Temperature Distribution Along H.EX.Depth,Vw=2000 (l/h),Va=500 (cfm), Tw (in)=50
0
86
C
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
Table (5-7) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
89
Table (5-8) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
90
Table (5-9) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
91
Table (5-10) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
91
Table (5-11) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
91
Table (5-12) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
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
Lf
Fin Length
mm
L
Heat Exchanger Length
cm
m& a
Air Mass Flow Rate
Kg/s
m& f
Fuel Mass Flow Rate
Kg/s
Nu
Nusselt Number =
Nt
Number of Tubes
-
Nr
Number of Rows
-
N
Number of Slices
-
P
Pressure
bar
Q&
Rate of Heat Transfer Loss or Gain
W
kJ/kg
hd k
-
ρud h μ
Re
Reynolds Number =
R
Gas Cosatant
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
−
.
J/kg.K
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 =
xt
Isentropic Temperature Ratio for Turbine =
.
T3 T4 s
T2 s T1
-
Greek Letters
ηo
= 1−
ηf
Fin Efficiency
ρ
Density (kg/m³)
μ
Viscosity (kg/m.s)
Ø
Maximum to Minimum Temperature Ratio through the gas turbine engine
α
t*
ηo
Overall Efficiency of Gas Turbine
ƒ
Fuel / Air Ratio
λt
Perimeter of Tube Side (m3)
λf
Perimeter of Fin (m3)
γ
Specific Heat Ratio
Ŧ
Torque (N.m)
Af A
(1 − η ) Overall Surface f
Efficiency
Ø η c* η
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 3
Compressor Discharge 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
Chapter One…………………………………………………………………….Introduction
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 Combustion Chamber
Power Turbine
Turbine
Compressor
Driven Equipment
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
Chapter One…………………………………………………………………….Introduction
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]
Chapter One…………………………………………………………………….Introduction
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.
Chapter One…………………………………………………………………….Introduction
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.
Chapter One…………………………………………………………………….Introduction
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.
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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.
Chapter Two……………………………………………………………….Literature Survey 160
Output Power, (M W )
150
140
130
120
110
100 260
270
280
290
300
310
320
330
340
Inlet Temp, (K)
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
Chapter Two……………………………………………………………….Literature Survey
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
EXPERIMENTAL WORK
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
Chapter Three…………………………………………………………Experimental Work
temperature. A piece of ice was immersed into cold tank so that the mixture becomes homogenous. Table (3.1) Heat Exchanger Geometry
Parameter Core
Fin
Tube
Dimension 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
Pitch (Pf)
1.46 mm
Length (Lf)
7.52 mm
Depth (Df)
15.88 mm
Thick
0.24 mm
(Tf)
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.
Chapter Three…………………………………………………………Experimental Work
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). Pitot Tube: 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
Chapter Three…………………………………………………………Experimental Work
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.
Chapter Three…………………………………………………………Experimental Work
Heat Exchanger Duct
T
Fan
P
By Pass
Rotameter
T
Rotameter
P
Valve
Hot Tank
Cold Tank
Heaters
Ice
Valve
Pump
Valve
Figure (3.1a) Schematic Diagram of the Built Rig (Heat Exchanger)
Chapter Three…………………………………………………………Experimental Work
Fan
Cold Tank
Figure (3.1b) Configuration of the Built Rig (Heat Exchanger)
Pitot-Tube Rotameter
Heat Exchanger Hot Tank
Gas Turbine
Figure (3.1c) Configuration of the Test Rig (Inlet Cooling System) Preparing to Gas Turbine
Chapter Three…………………………………………………………Experimental Work
XL Dt XT Ht
Figuer (3.2.a) Top View of Heat Exchanger Geometry
Pf Lf H
D
L Figure (3.2.b) Front View of Heat Exchanger Geometry
Chapter Three…………………………………………………………Experimental Work
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.
Pump
Valve
Rotameters
Cold Tank
Cold Tank
Valve
Hot Tank Hot Tank
Compressor
Duct
Heat Exchanger
Figure (3.3) Schematic Diagram of the Test Rig (Overall)
Ambient Air
Combustor
Chapter Three…………………………………………………………Experimental Work
Dynomometer
LP.Turbine
HP.Turbine
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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 flowmeter. 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
Chapter Three…………………………………………………………Experimental Work
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, highpressure 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
Chapter Three…………………………………………………………Experimental Work
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 eddycurrent 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.
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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 :-
Ncc Ts
=
Nc T1
.
……………………………………………………………...…(3-2) .
M a Ta Pa
=
M ac Ts Ps
……………………………………………………..…(3-3)
T2 T2c = …………………………………………………………………(3-4) Ta Ts
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:
Chapter Three…………………………………………………………Experimental Work
N cc = N c
288.16 ………………………………………………….….….(3-5) Ta .
.
M ac =
1.0133 288.16
×
M a Ta
……………………………………………..…..(3-6)
Pa
⎛ 288.16 ⎞ ⎟⎟T2 ………………………………………………………….(3-7) T2c = ⎜⎜ ⎝ Ta ⎠
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.
Ta = Ta ( o C ) + 273.16
2.
Pa =
3.
T1C = (T1 + 273.16)(
4.
P1 = Pa − 97.9 Χ10 −6 (ΔP)
5.
P2c = P2 + Pa
Barometer Re ading (mbar ) 1000 −
288.16 ) Ta
−
−
T2 c = (T2 + 273.16)(
7.
P3C = P3 + Pa
288.16 ) Ta
−
−
T3c = [T3 + 273.16](
9.
P4 c = P4 + Pa
−
bar K bar bar
6.
8.
K
K bar
288.16 ) Ta
K bar
Chapter Three…………………………………………………………Experimental Work −
288.16 ) Ta
K
−
288.16 ) Ta
K
10
T4 c = (T4 + 273.16)(
11.
T5c = (T5 + 273.16)(
12.
P5c = P4 c − 0.001333(ΔP4 5 )
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 ( rc ) rc =
P2c …………………………………………………………………..(3-8) P1c
2. Compressor Isentropic Efficiency (η c ) ηc =
T1c ( γ −1) γ (rc − 1) × 100 0 0 ……………………………….………….(3-9) T2 c − T1c
3. Air Mass Flow Rate (kg/s) m& a
T1 P1
= 0.3005
m& a c = m& a(
ΔP P1
Ta ) 288.16
............................................................................(3-10)
where : ΔP in mm wg T1 in K
Chapter Three…………………………………………………………Experimental Work
4. Fuel Flow Rate (kg/s) m& fc = ρ f
v f × 10 −3
288.16 Ta
3600
…………………………………..……….…(3-11)
where:
ρ f = ρ w ∗ sg
The specific gravity of the fuel (sg) is 0.774 5. Air/Fuel Ratio (A/F) A/F = m& ac m& fc ……………………………………………………………(3-12) 6. Combustion Chamber Temperature Rise (K) −
ΔTcct
43740 − 10 T2 c = +2 ………………………………………..…….(3-13) 1.0078 A F + 6.6
Where:
−
T2C = T2 c − 273.16
7. Combustion Chamber Pressure Loss (٪) ΔPcc = (
P2 c − P3c ) × 100 ……………………………………………….…..(3-14) P2 c
8. Gas Generator Turbine Expansion Ratio rt1 =
P3c …………………………………………………………………(3-15) P4c
9. Theoretical Power Input to Compressor ( w& ) W& C = m& aCpa(T2 c − T1c ) ……………………………………………………(3-16)
Where: cpa = 1000 J kgK
Chapter Three…………………………………………………………Experimental Work
10. Theoretical Power Output Of Gas Generator Turbine ( w& ) W& t 1 = (m& ac + m& fc )Cp g (T3c − T4c ) ……………………………………………(3-17)
where: cp g = 1150 J kgK 11. Compressor Mechanical Efficiency (η mc ) η mc = (
2 w& c ) × 100 ………………………………………………………(3-18) w& c + wt′1
12. Gas Generator Turbine Mechanical Efficiency (η mt1 ) η mt1 = (
W& C + W& t1 ) × 100 …………………………………………….….……(3-19) 2W& t1
13. Compressor Overall Efficiency (η oc ) η oc = η c ×
η mc 100
……………………………………………………..………(3-20)
14. Power Turbine Expansion Ratio P4 c …………………………………………………………………(3-21) P5c
r12 =
15. Power Turbine Isentropic Efficiency (η12 )
η12 =
T4C − T5C × 100 ………………………………………………(3-22) 1 (γ −1) γ ] T4C [1 − ( ) r12
where: γ = 1.33 16. Power Turbine Power Output ( w& ) 2πN 2τ W& t 2c = 60
Ta ……………………………………………………(3-23) 288.16
Where: N 2 power turbine speed (RPM) and τ is the torque in (N.m)
Chapter Three…………………………………………………………Experimental Work
17. Power Turbine Theoretical Power Output ( w& ) W& t 2 = (m& ac + m& fc )Cp g (T4c − T5c ) ………………………………….………….(3-24)
18. Power Turbine Mechanical Efficiency (η mt 2 )
η mt 2 = (
W& t 2 c ) × 100 …………………………………………………………..(3-25) W& t 2
19. Power Turbine Overall Efficiency (η ot 2 )
η ot 2 = η t 2 ×
η mt 2 100
…………………………………………………………….(3-26)
20. Overall Thermal Efficiency (η th ) η th = (
W& t 2 c ) × 100 …………………………………………………..…(3-27) H .V × m& fc
Where: H.V= L.C.V × 4.1868 kJ kg for kerosene used in the tests (GT-85), and L.C.V= 10300 Kcal kg the lower calorific value of the fuel used, Gilbert and Gordon LTD [23]. 21. Specific Fuel Consumption ( kg Kw.s ) sfc = (
m& fc ) × 1000 ……………………………………………………..…..(3-28) W& t 2c
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)
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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 directtransfer 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.
Chapter four…………………………………………………………………………Theory
Surface Characteristics
Problem Specifications
Design Theory Procedure
Physical Properties
Optional Solutions
Evaluation Criteria
Evaluation Procedure
Optimum Solution
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.
Chapter four…………………………………………………………………………Theory
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.)
Chapter four…………………………………………………………………………Theory
Figure 4.2 Step by Step method with two Directions
Chapter four…………………………………………………………………………Theory
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.
Twi
Water Flow Tao
Lf Slice 1
Dt Tar2
Tar1
Slice 2
Tai
Two Air Flow
Figure (4.3a) Slice for inlet single Tube
Figure (4.3b) Exit of one row inlet to next row
Chapter four…………………………………………………………………………Theory
2 ………………Nr
N-1.………………….……2
I=1
J=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: .
Vw (Nt) = [ N r × v w × Acw
] ………………………………………………….(4-1)
This value will be used for the estimation of the length and depth of the heat exchanger as: L = X T × ( N t + 1) ………………………………………………….(4-2) D = [ X L × ( Nr − 1)] + Dt ………………………………………………….(4-3)
4.2.2.3 Mass Conservation Mass conservation of the water flow through the tube is simply .
.
Σ m in −Σ m out = 0 ………………………………………………….(4-4) and the mass conservation equation for the water at each slice can be written as:
Chapter four…………………………………………………………………………Theory .
.
m w(i+1,j)= m 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. .
.
.
m a(j)= ρ a ( j ) × V a ( j ) ……………………………………………………...(4-6)
4.2.2.4 Log-Mean Temperature Difference To estimate the true mean temperature difference ( ΔTm ) 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]: LMTD =
(Ta (i, j ) − Tw (i, j + 1)) − (Ta (i + 1, j ) − Tw (i, j )) ………………………….(4-7) (Ta (i, j ) − Tw (i, j + 1)) ln (Ta (i + 1, j ) − Tw (i, j ))
Ta,(i,j) ∆Tm Ta,(i+1,j)
Tw,(i,j+1)
∆Tm Tw,(i,j) 0
Atotal area
Figure 4.4 The Mean Temperature Difference Along a Single Pass
Chapter four…………………………………………………………………………Theory
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]: ΔTm = F × LMTD
…………………………………………………….(4-8)
where: F=
( R 2 + 1) ln (1 − S ) (1 − RS ) ( R − 1) ln[
2 − S[ R + 1 − ( R 2 + 1) ]
……………………………………….(4-8a)
2 − S [ R + 1 + ( R 2 + 1) ]
R=
(Th (i, j ) − Th (i + 1, j )) ……………………………………….(4-8b) (Tc (i. j + 1) − Tc (i, j ))
S=
(Tc (i, j + 1) − Tc (i, j )) ……………………………………….(4-8c) (Th (i, j ) − Tc (i, j ))
4.2.2.5 Heat Load The heat load passes through a control volume on the water side is: .
Q (i,j)= m (i,j) × Cp (i,j) × ΔT (i,j) …………………………………..…….(4-9) 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]:
U o (i, j ) =
1
tt 1 1 1 [( )+( )+( )+( )] ha (i, j ) × η o K t ( Ain Aout ) hw (i, j ) × ( Ai n Ao ut ) h f × ( Ai n Ao ut )
....(4 − 11)
Chapter four…………………………………………………………………………Theory
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: h w = 0 . 023 Re 0.8 Pr n (
k d hw
………………………………………... (4-12)
)
and cross flow area for water side (rectangular tube with semi circular ends ):
[
]
Ac w = [( H t − Tt ) × ( Dt − H t )] + ( H t − Tt ) 2 × π ……………………………..(4-13)
also Re =
dh =
Pr =
ρud h w ……………………………………..…………………….(4-14) μ 4 Acw
λt
μCp k
……………………………………..………………….….(4-15) ……………………………………..…………………....….(4-16)
where the ranges of Re and Pr are: 6000 ≤ Re ≤ 10 7 0.5 ≤ Pr ≤ 120
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
Chapter four…………………………………………………………………………Theory
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): h a = 0 . 023 Re 0 .8 Pr n (
k d ha
………………………………………….. (4-17)
)
For laminar flow, the Sieder and Tate (1930) [33] correlation can be used. ha = 1.86(Re Pr )
0 .3
0.3
k ⎛ d ha ⎞ ( ) ⎟ ⎜ D f ⎠ d ………………………………….……..….….(4-18) ⎝ ha
where cross flow area for air side: Pf
Ac a = (
2
)( X T − H t ) ………………………………………………..….(4-19)
The hydraulic diameter for the air side: dha =
4 Aca
……………………………..………………….….(4-20)
λf
λ f =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: .
ma ( j ) ma(i, j ) = …………………………………………………………..(4-21) N .
.
m (i, j ) ……………………………………..………………..……. (4-22) Ga = a ( Aca ) Ra (i, j ) =
(d h a Ga ) ………………………………..………………….… (4-23) μ a (i, j )
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]:
Chapter four…………………………………………………………………………Theory
ηf =
[tanh(m × ( (m × (
Lf 2
Lf 2
))]
………………………………..………………… (4-24)
))
where:
m= [
(ha (i, j ) × ( Pf ) ( K t × Aca )
………………………………..………………..… (4-25)
The overall surface fin efficiency is: η o = 1 − [(
Af A
) × (1 − η f )] ………………………………..………………… (4-26)
where: Af H
Aexp H
= (4 × L f × D f )(
1 ) ………………………………..……………….. (4-27) Pf
= [((π × H t ) + (2 × ( Dt − H t )) − (2 × T f )](
1 ) ………..………………… (4-28) Pf
Af Aexp A =( )+( ) ………..…………………………………. ………….. (4-29) H H H
The log- mean temperature difference and overall heat transfer coefficient is calculated to obtain the height for one slice as follow:
H (i, j ) =
Q(i, j ) (U o (i, j ) × A × F × LMTD (i, j ))
…………………………………….(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, 2 Df va ΔP = ( ρ )× f × 2 d ha
…………………………………………………..(4-31)
Chapter four…………………………………………………………………………Theory
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=
64 …………………………………………………………………...(4-32) Ra
For isosceles triangular ducts, Bhatti and Shah(1987)[36] recommend, for fully-developed turbulent flow: ƒ=
0.078 ……………………………………………………………….(4-33) ( Ra) 0.25
For the power of the fan the equation is: .
Pfan =
V air Δp
η fan
……………………..………………………………….….(4-34)
And, .
V air =
m& air
ρ air
……………………..………………………………..…….(3-35)
Thus,
Pmotor , fan =
Pfan
η motor , fan
……………………..…………………………….….(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 V&w 4. Calculating (No. of Tubes / Row) ,Eq.(4-1)
Chapter four…………………………………………………………………………Theory
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 k No., Nusselt No, Eq.(4-13) to Eq.(4-20). 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: ξ% =
(Varible cal. − Varible assu.)
Variable cal.
is converged to a value within 1 × 10 −3 .hence,
the variable here is air exit temperature or air velocity over tubes. 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)
Chapter four…………………………………………………………………………Theory
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 ( w ) and rejecting heat to a sink at low temperature (Qrej) is shown in Fig.( 4.5) as a block diagram.
Figure. 4.5 Basic Gas Turbine Engine
Chapter four…………………………………………………………………………Theory
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.
Chapter four…………………………………………………………………………Theory
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).
Chapter four…………………………………………………………………………Theory
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=
P2 − P3 ………………………………………………………..(4-37a) P2
ΔP5a =
P5 − Pa Pa
(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)
Chapter four…………………………………………………………………………Theory
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].
Chapter four…………………………………………………………………………Theory
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
Chapter four…………………………………………………………………………Theory
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.
Chapter four…………………………………………………………………………Theory
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.
Chapter four…………………………………………………………………………Theory
Theoretical Overall Efficiency (%)
8
7
6
5
4 Expansion Ratio (rT)rT=1. =1.12 Expansion Ratio 12 3 Expansion Ratio (rT)rT=1. =1.08 Expansion Ratio 08 2 3.08
3.1
3.12
3.14
3.16
3.18 T3/T1 (
Ø
3.2
3.22
3.24
3.26
3.28
)
Fig (4.8) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=699 °C
7
Theoretical Overall Efficiency (%)
6.5 6 5.5 5 4.5 4 3.5 Expansion Ratio =1.12 Expansion Ratio(rT) rrT=1. T=1. 12 12
3
Expansion Ratio(rT) rrT=1. T=1. 08 08 Expansion Ratio =1.08
2.5 2 3.08
3.1
3.12
3.14
3.16
3.18 T3/T1 (
3.2
3.22
3.24
3.26
Ø)
Fig (4.9) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=628.5 °C
3.28
Chapter four…………………………………………………………………………Theory
Theoretical Overall Efficiency (%)
6.5 6 5.5 5 4.5 4 3.5 3
Expansion Ratio r(rT) =1.12 T=1. 12
2.5
T=1. 08 Expansion Ratio r(rT) =1.08
2 3.08
3.1
3.12
3.14
3.16
3.18
T3/T1 (
3.2
3.22
3.24
3.26
3.28
Ø)
Fig (4.10) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=588 °C 6.5
Theoretical Overall Efficiency (%)
6 5.5 5 4.5 4 3.5 3 Expansion Ratio (rT) rT=1.=1.12 12 2.5
Expansion Ratio (rT) =1.0 rT=1. 08 8
2 3.08
3.1
3.12
3.14
3.16
3.18 T3/T1 (
3.2
3.22
3.24
3.26
3.28
Ø)
Fig (4.11) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=547.5 °C
Chapter four…………………………………………………………………………Theory
7.5
Theoretical Overall Efficiency (%)
7 6.5 6 5.5 5 4.5 T1= 15 C
4 3.5
T1= 30 C
3 500
520
540
560
580
600
620
640
660
680
Turbine Inlet Temperature (T3)
Fig (4.12) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on the Theoretical Overall Efficiency ,Expansion Ratio=1.12
Theoretical Overall Efficiency (%)
5.5 5 4.5 4 3.5 3 T1= 15 C 2.5 T1=30 C 2 500
520
540
560
580
600
620
640
660
680
Turbine Inlet Temperature (T3)
Fig (4.13) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on the Theoretical Overall Efficiency ,Expansion Ratio=1.08
Chapter four…………………………………………………………………………Theory
7.5
Theoretical Overall Efficiency (%)
7 6.5 6 5.5 5 4.5 T3=669 C ,
Ø=3.2
4 T3=628.5 C,Ø =3.1 3.5
T3=588 C, Ø =2.9
3 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)
Fig (4.14) Theoretical Overall Efficiency as a Function of Expansion Ratio with Isentropic Efficiency (ηt,ηc =0.9) and T1=15°C
Theoretical Overall Efficiency (%)
6
5.5
5
4.5
4 T3= 699 C, T3=628.5 C,
3.5
T3=588 C,
Ø =3.10 Ø=2.97 Ø=2.84
3 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)
Fig (4.15) Theoretical Overall Efficiency as a Function of Expansion Ratio with Isentropic Efficiency (ηηt,ηc =0.9) and T1=30°C
Chapter four…………………………………………………………………………Theory
Theoretical Overall Efficiency (%)
8
7
6
5
4 Expansion Ratio (rT) =1.12 Expansion Ratio rT=1. rT=1.1212
Expansion Ratio rT=1. 1 Expansion Ratio (rT) =1.1
3
Expansion Ratio rrT=1. Expansion Ratio (rT) = 1.08 T=1. 08 08 2 10
15
20
25
30
35
Compressor Inlet Temperature (C)
Fig (4.16) The Effect of Compressor Inlet Temperature (with Variable Expansion Ratio) on the Theoretical Overall Efficiency, T3= 669 °C 7
Theoretical Overall Efficiency (%)
6.5 6 5.5 5 4.5 4 Expansion Ratio (rT) rT= =1.12
3.5 3
Expansion Ratio (rT) r =1.1 T=
2.5
rT= =1.08 Expansion Ratio (rT)
2 10
15
20
25
30
35
Compressor Inlet Temperature (C)
Fig (4.17) The Effect of Compressor Inlet Temperature (with Variable Expansion Ratio) on the Theoretical Overall Efficiency, T3= 628.5 °C
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
Chapter Five .......................................................................................Results &Discussion
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
Chapter Five .......................................................................................Results &Discussion
(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
Chapter Five .......................................................................................Results &Discussion
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
Chapter Five .......................................................................................Results &Discussion
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
Chapter Five .......................................................................................Results &Discussion
• 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
Chapter Five .......................................................................................Results &Discussion
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).
Chapter Five .......................................................................................Results &Discussion
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.(519).The increasing of heat rate is due to increasing of both turbine inlet temperature and compressor inlet temperature. A decrease in heat rate of about (37 %) was obtained due to reducing compressor inlet temperature from (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
Chapter Five .......................................................................................Results &Discussion
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
Chapter Five .......................................................................................Results &Discussion
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 about (22 %) was obtained due to reducing compressor inlet temperature from (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 =
5 =142.8 (kW) 0.035
• The minimum heat rate The minimum heat rate can be calculated for (GT-85) as follow: heat rate=
1 =28.56 0.035
Chapter Five .......................................................................................Results &Discussion
Percent of Design =
Calculated Value at (T1 = 15 ) − Calculated Value at (T1 = 30 ) Design Value
× 100%
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
Chapter Five .......................................................................................Results &Discussion
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 (3015°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.
47.28972
62.55761
62.80377
63.05259
63.30411
84.24339
84.56969
84.89945
85.56966
85.23277
85.91026
2.328
1.164
1.164
1.164
1.164
.776
.776
.776
.776
.776
.776
96.19064
95.78031
95.37463
94.97345
94.57677
21.49
23.59
25.69
27.69
29.79
19.39
21.49
23.59
25.69
27.69
29.79
23.59
10
10
10
10
10
10
10
10
12
12
12
12
12
12
12
12
12
12
12
12
.235
.237
.238
.239
.241
.242
.238
.239
.241
.242
.241
.242
6
5
4
3
2
1
4
3
2
1
2
1
Heat Exchanger (55 3.6 37) cm3 Heat Exchanger (28.5 7.7 27) cm3
Heat Exchanger (18 11 18) cm3
×
32
25.69
25.69
10
10
10
10
×
94.18447
69.22968
27.69
27.69
29.79
27.69
29.79
Core Size (L D H)
×
68.93851
29.79
32
29. 79
32
No. of ROW
×
68.65057
68.36579
50.81485
50.60406
(mbar)
∆pa
×
47.10219
Uo(w/m2. °C) ha w/m2. °C) Ta (in) °C Ta (out) °C Tw (in) °C Tw (out) °C
×
2.328
Q (KW)
Table (5-1) Air Temperature Distribution Along H.EX.Depth ,Vw=2000 (L/h),Va=2000 (cfm)
Chapter Five .......................................................................................Results &Discussion ×
×
45.31548
60.24699
60.69621
60.46152
60.46152
81.55405
81.84826
81.53701
81.82931
82.11871
81.81271
2.284
1.142
1.142
1.142
1.142
.761
.761
.761
.761
.761
.761
90.22915
90.59176
90.24413
89.89323
90.2618
40.49
38.39
36.29
34.19
32.09
42.59
40.49
38.39
36.29
34.19
32.09
38.39
50
50
50
50
50
50
50
50
48
48
48
48
48
48
48
48
48
48
48
48
.250
.248
.247
.246
.245
.243
.247
.246
.245
.243
.245
.243
6
5
4
3
2
1
4
3
2
1
2
1
Heat Exchanger (55 3.6 37) cm3 Heat Exchanger (28.5 7.7 27) cm3
Heat Exchanger (18 11 18) cm3
×
32
36.29
36.29
50
50
50
50
×
89.90861
65.76105
34.19
34.19
32.09
36.19
34.09
×
65.25092
32.09
32
34.09
32
Core Size (L D H)
×
65.51846
65.00219
48.59657
47.79097
No. of (mbar) ROW
∆pa
×
45.04536
Ta (out) °C Tw (in) °C Tw (out) °C
×
2.284
Q (KW) Uo (w/m2. °C) ha (w/m2. °C) Ta (in) °C
Table (5-2) Air Temperature Distribution Along H.EX.Depth ,Vw=2000 (L/h),Va=2000 (cfm)
Chapter Five .......................................................................................Results &Discussion ×
×
20.13704
18.50125
18.57706
18.65369
18.73119
25.65959
25.76418
25.86992
25.97685
26.08497
26.1943
.456
.2329
.2329
.2329
.2329
.155
.155
.155
.155
.155
.155
27.07296
26.95747
26.8433
26.7304
26.61876
21.57
23.67
25.78
27.89
29.99
19.46
21.57
23.67
25.78
27.89
29.99
23.67
10
10
10
10
10
10
10
10
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
6.15
6.18
6.22
6.25
6.29
6.32
6.22
6.25
6.29
6.32
6.29
6.32
6
5
4
3
2
1
4
3
2
1
2
1
Heat Exchanger (55 3.6 37) cm3 Heat Exchanger (28.5 7.7 27) cm3
Heat Exchanger (18 11 18) cm3
×
32
25.78
25.78
10
10
10
10
×
26.50834
19.21763
27.89
27.89
29.99
27.89
29.99
×
19.1368
29.99
32
29.99
32
×
19.05688
18.97783
20.68979
20.64679
Core Size (L D H) ×
20.09592
E-02
No. of (mbar) ROW
∆pa
×
.456
Q (KW) Uo (w/m2. °C) ha (w/m2. °C) Ta (in) °C Ta (out) °C Tw (in) °C Tw (out) °C
Table (5-3) Air Temperature Distribution Along H.X.Depth ,Vw=2000 (L/h),Va=500 (cfm)
Chapter Five .......................................................................................Results &Discussion
×
×
22.79861
25.26978
26.01909
25.5097
26.04461
35.53004
34.93548
35.29956
35.51754
35.57734
36.40247
.684
.342
.342
.342
.342
.228
.228
.228
.228
.228
.228
37.36954
36.50051
36.43759
36.2082
35.82522
78.37
69.88
54.46
43.02
34.12
88.12
78.37
69.88
54.46
43.02
34.12
69.88
50
50
50
50
50
50
50
50
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
7.58
7.25
6.98
6.75
6.57
6.41
6.95
6.78
6.57
6.41
6.57
6.41
6
5
4
3
2
1
4
3
2
1
2
1
Heat Exchanger (55 3.6 37) cm3 Heat Exchanger (28.5 7.7 27) cm3
Heat Exchanger (18 11 18) cm3
×
32
54.46
54.46
50
50
50
50
×
36.45073
26.53511
43.02
43.02
34.12
43.02
34.12
×
25.98009
34.12
32
34.12
32
×
26.0863
25.73128
23.4331
23.84225
Core Size (L D H) ×
23.18965
E-02
No. of (mbar) ROW
∆pa
×
.684
Q (KW) Uo (w/m2. °C) ha (w/m2. °C) Ta (in) °C Ta (out) °C Tw (in) °C Tw (out) 0C
Table (5-4) Air Temperature Distribution Along H.X.Depth ,Vw=2000 (L/h),Va=500 (cfm)
Chapter Five .......................................................................................Results &Discussion ×
×
Chapter Five .......................................................................................Results &Discussion
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)
ha (w/m2.°C)
Q (watt)
Uo (w/m2.°C)
Ta out (°C)
A (m2)
50.60411
116.4624
47.09801
29.89904
.1187113
50.6041
116.4564
47.0985
29.89915
.1192767
50.6041
116.4504
47.09899
29.89925
.1198476
50.6041
116.4474
47.09924
29.89931
.1201254
50.6041
116.4474
47.09924
29.89931
.1201254
50.60409
116.4385
47.09997
29.89947
.1210062
50.60408
116.4326
47.10045
29.89958
.1215941
50.60408
116.4267
47.10094
29.89968
.1221879
50.60407
116.4208
47.10142
29.89979
.1227876
50.60406
116.4149
47.10191
29.89989
.1233933
50.60406
116.409
47.0239
29.9
.1240051
50.60405
116.4032
47.10287
29.9001
.1246231
50.60405
116.3974
47.10334
29.90021
.1252474
50.60405
116.3945
47.10359
29.90026
.1255509
50.60404
116.3858
47.10431
29.90042
.1265153
50.60403
116.3801
47.10478
29.90052
.1271591
50.60403
116.3743
47.10526
29.90063
.1278096
50.60403
116.3686
47.10573
29.90073
.1284669
50.60402
116.3629
47.1062
29.90083
.1291311
50.60402
116.3572
47.10667
29.90093
.1298023
Vw = 2000 l/h Va=2000cfm
Chapter Five .......................................................................................Results &Discussion
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)
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
50.81489
116.4624
47.28552
27.799
. 129798
50.81488
116.4564
47.28601
27.7991
.131507
50.81488
116.4504
47.2865
27.79921
.132205
50.81488
116.4474
47.28675
27.79926
.1329107
50.81488
116.4474
47.28675
27.79926
.133253
50.81487
116.4385
47.28749
27.79943
.133253
50.81486
116.4326
47.28798
27.79953
.1343451
50.81486
116.4267
47.28846
27.79964
.1350742
50.81486
116.4208
47.28896
27.79975
.1358114
50.81485
116.4149
47.28944
27.79985
.1365567
50.81485
116.409
47.28992
27.79996
.1373105
50.81484
116.4032
47.29041
27.80006
.1380727
50.81483
116.3974
47.29089
27.80017
.1388436
50.81483
116.3945
47.29113
27.80022
.1396233
50.81482
116.3858
47.29185
27.80038
.140001
50.81482
116.3801
47.29234
27.80048
.1412096
50.81482
116.3743
47.29281
27.80058
.1420166
50.81481
116.3686
47.29329
27.80069
.142833
50.8148
116.3629
47.29376
27.80079
.1436589
50.8148
116.3572
47.29423
27.80089
.1444946
Vw = 2000 l/h Va=2000cfm
Chapter Five .......................................................................................Results &Discussion
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
ha (w/m2. °C)
Uo (w/m2. °C)
Ta out (°C)
Q (watt)
A (m2)
48.69674
114.1267
45.37363
34.09845
.1494603
48.69672
114.1362
45.37796
34.09862
.1503539
48.69672
114.1457
45.38214
34.0988
.1512587
48.69672
114.1504
45.38418
34.09888
.1516965
48.69672
114.1504
45.38418
34.09888
.1516965
48.6967
114.1645
45.39007
34.09914
.1531029
48.69669
114.1739
45.39383
34.09932
.1540425
48.69669
114.1833
45.39747
34.09949
.1549942
48.69668
114.1926
45.40099
34.09966
.1559581
48.69667
114.2018
45.40438
34.09983
.1569344
48.69666
114.2111
45.0767
34.1
.1579234
48.69665
114.2203
45.41086
34.10017
.1589252
48.69664
114.2295
45.41394
34.10034
.1599401
48.69664
114.234
45.41545
4.10042
.1604302
48.69663
114.2477
45.41983
34.10067
.1620101
48.69662
114.2567
45.42263
34.10084
.1630657
48.69662
114.2657
45.42537
34.10101
.1641354
48.69661
114.2747
45.428
34.10117
.1652195
48.6966
114.2837
45.43056
34.10133
.1663181
48.69659
114.2926
45.43305
34.1015
.1674317
Chapter Five .......................................................................................Results &Discussion
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)
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
48.49659
114.1226
45.19305
36.19837
.1674317
48.49654
114.1267
45.19515
36.19855
.1714795
48.49653
114.1362
45.19946
36.19872
.172653
48.49652
114.1457
45.20361
36.19881
.1738432
48.49652
114.1504
45.20564
36.19881
.1744162
48.49652
114.1504
45.20564
36.19907
.1744162
48.4965
14.1645
45.21149
36.19924
.1762752
48.4965
114.1739
45.21524
36.19941
.1775177
48.49649
114.1833
45.21885
36.19958
.1787783
48.49648
114.1926
45.22234
36.19976
.1800573
48.49648
114.2018
45.22573
36.19992
.1813552
48.49647
114.2111
45.229
36.2001
.1826723
48.49646
114.2203
45.23216
36.20026
.1840092
48.49645
114.2295
45.23523
36.20035
.1853661
48.49644
114.234
45.23672
36.2006
.1860176
48.49644
114.2477
45.24109
36.20076
.1881417
48.49643
114.2567
45.24387
36.20093
.1895614
48.49642
114.2657
45.24658
36.2011
.191003
48.49641
114.2747
45.2492
36.20126
.192467
48.49641
114.2837
45.25175
36.20142
.1939538
48.4964
114.2926
45.25422
36.20128
.195464
Vw = 2000 l/h Va=2000cfm
Chapter Five .......................................................................................Results &Discussion
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
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
20.64677
116.4624
20.09579
29.89989
.2781911
20.64677
116.4564
20.09587
29.9
.2795178
20.64677
116.4504
20.09595
29.90011
.2808575
20.64676
116.4474
20.09598
29.90016
.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
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
20.73277
116.4624
20.17804
27.79996
.3081484
20.73277
116.4564
20.17811
27.80007
.3097861
20.73277
116.4504
20.17819
27.80018
.3114415
20.73276
116.4474
20.17823
27.80023
.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
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
19.86854
114.1267
19.32309
34.09974
.3509701
19.86854
114.1362
19.32381
34.09991
.353089
19.86854
114.1457
19.32451
34.10009
.3552339
19.86854
114.1504
19.32485
34.10017
.356272
Chapter Five .......................................................................................Results &Discussion
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
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
19.52759
114.1267
18.99886
43.02019
.3620837
19.52758
114.1362
18.99919
43.02028
.3668006
19.52758
114.1457
18.99919
43.02028
.3668006
19.52758
114.1504
19.00015
43.02053
.3632253
Chapter Five .......................................................................................Results &Discussion
6 Simulation Experiment
Heat Load (kW)
5
4
3
2
1
0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
5
Entering water at 10 °C Fig (5. 1a)
Simulation Experiment
Heat Load (kW)
4
3
2
1
0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
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)
Chapter Five .......................................................................................Results &Discussion
Heat Transfer Coefficient .ha (W/m 2.C)
60
Simulation Experiment
50
40
30
20
10
0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Heat Transfer Coefficient. ha (W/m 2.C)
60
Entering water at 10 °C Fig (5. 2a)
Simulation Experiment
50 40 30 20 10 0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Entering water at 50 °C Fig (5. 2b) 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)
Chapter Five .......................................................................................Results &Discussion
Heat Transfer Coefficient ha.(w/m2.c) 50.45
50.5
50.55
50.6
50.65
50.7
50.75
50.8
50.85
Slice Number (Bar Tube)
1
2 ROW 1 ROW 2 3
4
•
Entering water at 10 °C Fig (5. 3a)
Heat Transfer Coefficient ha.(w/m2.c) 48.35
48.4
48.45
48.5
48.55
48.6
48.65
48.7
48.75
Slice Number (Bar Tube)
1
2 ROW 1 ROW 2 3
4
•
Entering water at 50 °C 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
Chapter Five .......................................................................................Results &Discussion
Heat Transfer Coefficient ha.(w/m2.c) 20.6
20.62
20.64
20.66
20.68
20.7
20.72
20.74
Slice Number (Bar Tube)
1
2 ROW 1 ROW 2 3
4
•
Entering water at 10 °C Fig (5. 4a)
Heat Transfer Coefficient ha.(w/m2.c) 19.4 19.4 19.5 19.5 19.6 19.6 19.7 19.7 19.8 19.8 19.9 19.9
Slice Number
1
2
ROW 1 ROW 2
3
4
•
Entering water at 50 °C 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
Chapter Five .......................................................................................Results &Discussion
Heat Transfer Coefficient (W/m2.c)
96.5 96 95.5 95 94.5 94 93.5 93 1
2
3
4
5
6
Row Number
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
Chapter Five .......................................................................................Results &Discussion
Overall Heat Transfer Coefficient (W/m 2.C)
50 Simulation 45
Experiment
40 35 30 25 20 15 10 5 0 0
1
2
3
5
4
Air Velocity (m/s)
Overall Heat Transfer Coefficient(W/m 2.C)
•
Entering water at 10 °C Fig (5. 6a)
50 Simulation 45
Experiment
40 35 30 25 20 15 10 5 0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Entering water at 50 °C Fig (5. 6b) 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)
Chapter Five .......................................................................................Results &Discussion
Overall Heat Transfer Coefficient (w/m2.c) 47
47.05
47.1
47.15
47.2
47.25
47.3
47.35
Slice Number (Bar Tube)
1
2
ROW 1 ROW 2
3
4
•
Entering water at 10 °C Fig (5. 7a)
Overall Heat Transfer Coefficient (w/m2.c) 45.1
45.15
45.2
45.25
45.3
45.35
45.4
Slice Number
1
2
ROW 1 ROW 2
3
4
•
Entering water at 50 °C Fig (5. 7b)
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
Chapter Five .......................................................................................Results &Discussion
Overall Heat Transfer Coefficient (w/m2.c) 20.04
20.06 20. 08
20.1
20.12
20.14
20.16 20.18
20.2
Slice Number
1
2
ROW 1 ROW 2
3
4
•
Entering water at 10 °C Fig (5. 8a)
Overall Heat Transfer Coefficient (w/m2.c) 18.8
18.9
19
19.1
19.2
19.3
19.4
Slice Number
1
2 ROW 1 ROW 2 3
4
•
Entering water at 50 °C Fig (5. 8b)
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
Chapter Five .......................................................................................Results &Discussion
Overall Heat Transfer Coefficient(W/m2.c)
86.5 86 85.5 85 84.5 84 83.5 83 1
2
3
4
5
6
Row Number
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
Chapter Five .......................................................................................Results &Discussion
27.95
Air Exit Temperature (C )
27.9 27.85 27.8 27.75 27.7 27.65
Simulation 27.6
Experiment 27.55 0
1
0.5
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Entering water at 10 °C Fig (5. 10a)
44
Air Exit Temperature (C )
43 42 41 40 39 38 37 Simulation 36 Experiment 35 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Entering water at 50 °C 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)
Chapter Five .......................................................................................Results &Discussion
Exit Air Temperature (c) 26.5
27
27.5
28
28.5
29
29.5
30
30.5
Slice Number (Bar Tube)
1
2
ROW 1 ROW 2
3
4
•
Entering water at 10 °C Fig (5. 11a) Exit Air Temperature (c)
33
33.5
34
34.5
35
35.5
36
36.5
Slice Number (Bar Tube)
1
2 ROW 1 ROW 2 3
4
Entering water at 50 °C Fig (5. 11b) 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 •
Chapter Five .......................................................................................Results &Discussion
Exit Air Temperature (c) 26.5
27
27.5
28
28 .5
29
29.5
30
30.5
2 ROW 1 ROW 2 3
4
•
Entering water at 10 °C Fig (5. 12a)
Exit Air Tem perature (c) 30
32.5
35
37.5
40
42.5
45
1 Slice Number (Bar Tube)
Slice Number (Bar Tube)
1
2
Row 1 Row 2
3
4
•
Entering water at 50 °C Fig (5. 12b)
Fig(5. 12) Variation Exit Air Temperature) a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate 500 cfm
Chapter Five .......................................................................................Results &Discussion
35
25
20
15
10
5
0 1
2
3
4
5
6
Row No.ofNumber Row 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
43 Water Inlet Temperature 10 C
40 Pressure Drop in Air Side (mbar)*10-2
Air Exit Temperature (c)
30
37 Water Inlet Temperature 50 C
34 31 28 25 22 19 16 13 10 0.6
0.7
0.8
0.9
1
Aspect Ratio (H/L)
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
1.1
Chapter Five .......................................................................................Results &Discussion
2.5
Power Output (kW)
2
1.5
1
0.5
T1=15 C T1=30 C
0 610
620
630
640
650
660
Turbine Inlet Temperature (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)
2.6
Fuel Mass Flow Rate (g/s)
2.5
2.4
2.3
2.2
45000 RPM
2.1
55000 RPM 2 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.16) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Fuel Mass Flow Rate
Chapter Five .......................................................................................Results &Discussion
Specific Fuel Consumption (g/kW.s)
2 1.8 1.6 1.4 1.2 1 45000 RPM
0.8 55000 RPM
0.6 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.17) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed ) on the Specific Fuel Consumption
110
Heat Consumption (kW)
108 106 104 102 100 45000 RPM 98 55000 RPM 96 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.18) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Heat Consumption
Chapter Five .......................................................................................Results &Discussion
100 90
Heat Rate (kj/kW.s)
80 70 60 50 40 30 20
T1=15 C
10 0 610
T1=30 C 620
630
640
650
660
Turbine Inlet Temperature (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)
2 1.8
Overall Efficiency (%)
1.6 1.4 1.2 1 0.8 0.6 0.4
Ti=15 C
0.2 0 610
T1=30 C 620
630
640
650
660
Turbine Inlet Temperature (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)
Chapter Five .......................................................................................Results &Discussion
0.14 0.13 Air Mass Flow Rate (kg/s)
0.12 0.11 0.1 0.09 0.08 0.07 0.06
45000 RPM
0.05
55000 RPM
0.04 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.21) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Air Mass Flow Rate
450
400
cfm
350
300
45000 RPM
250
55000 RPM
200 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.22) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Air Flow Rate
Chapter Five .......................................................................................Results &Discussion
1.44
Compression Ratio
1.39 1.34 1.29 1.24 45000 RPM
1.19 55000 RPM
1.14 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.23) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Compression Ratio
6
Compressor Work (kW)
5 4 3 2 45000 RPM
1
55000 RPM
0 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.24) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Compressor Work
Chapter Five .......................................................................................Results &Discussion
9
HP.Turbine Work (kW)
8 7 6 5 4 3 2 T1=15 C 1 0 610
T1=30 C 620
630
640
650
660
Turbine Inlet Temperature (C )
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)
Chapter Five .......................................................................................Results &Discussion
80 70
Percent of Design
60 50 40 30 POWER OUTPUT 20 OVERALL EFFICIENCY 10
HEAT CONSUMPTION
0 10
15
20
25
30
35
Compressor Inlet Temperature (C)
Fig (5.26) The Influence Compressor Inlet Temperature (°C )on Gas Turbine Performance at (Gas Generator Speed 45000 RPM)
80 70
Percent of Design
60 50 40 30 POWER OUTPUT
20 OVERALL EFFICIENCY
10
HEAT CONSUMPTION
0 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.27) The Influence Compressor Inlet Temperature (°C )on Gas Turbine Performance at(Gas Generator Speed 55000 RPM)
Chapter Five .......................................................................................Results &Discussion
2.8 2.6
Overall Efficiency (%)
2.4 2.2 2 1.8
Theoretical,Expansion Ratio=1.03
1.6
Experimental,Expansion Ratio=1.03
1.4 1.2 500
525
550
575
600
625
650
675
700
HP.Turbine Inlet Tem perature (C)
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 lowround 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
1-Philip, K., "Power Generation Hand Book", 1st edition, McGraw-Hill Professional, ISBN: 0071396047, pp.115-117, 2002. http:// www.digitalengineeringlibrary.com 2-ISO , Gas Turbine Acceptance Tests, ISO 2314, International Organisation for Standardisation, Geneva,(1973). 3-Frank ,J. B.," GE Gas Turbine Performance Characteristics", GE Power Systems Schenectady, NY, GER-3567H, pp.6-8, 2002. 4-General Electric Energy (Oil & Gas)," MS 5001 Gas Turbine" Technical Training. 5-International Energy Outlook, DOE/EIA-0484, 2004. 6-Ganapathy,V., "Process-Design Criteria of Air Cooled Heat Exchangers" ,Chemical Engineering,pp.418-425,McGraw-Hill Publication Book,Co.,Newyork,1979. 7-Hedderich, C.P., and Kellerher, M.D. "Design and Optimization of Air Cooled Heat Exchangers", Journal of Heat Transfer, vol. 104 pp683-890, Nov. (1982). 8-Zhang, C., "Numerical Modeling Using a Quasi-Tree Dimensional Procedure for Large Power Plant", Trans. of ASME Journal of Heat Transfer, v.116, p.841, 1994. 9-Matthew, S., Layton and Joseph O'Hagan, "Comparison of Alternate Cooling Technologies for California Power Plants", Electric power research institute (2002). 10-Dohoy, J. , Dennis, N.," Numerical Modeling of Cross Flow Compact Heat Exchanger with Louvered Fins using Thermal Resistance Concept" SAE.Paper , 2006-01-0726, The University of Michigan. 11-Tarrad, A. H., Khudor, D. S., and Wahed, M.A.," A Simplified Model for 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.
12-De Lucia ,M., Lanfranchi, C., and Boggio, V., “Benefits of compressor 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. 13-Saleh, F. A., "The Influence of Water Injection on Two Shaft Gas Turbine Performance", M.Sc.Thesis, AL-Mustansiria University College of Engineering, 1996. 14-Ait-Ali,M. A., “Optimum power boosting of gas-turbine cycles with compressor inlet air refrigeration.”, Journal of Engineering for Gas Turbines and Power, Vol. 119, pp. 124–129, 2001. 15-GE Nuovo Pignone Internal DT-'N',"Gas Turbine Performance Curves",Hyundi Engineering and Construction CO, Commessa-Job 170552021,2002. 16- Donald C. Erickson and Icksoo Kyung, “Aqua Absorption Turbine Inlet Cooler”, ASME International Mechanical Engineering Congress and Exposition, Draft IMECE2003-42870,pp.113-115. 17- Hameed,N.,"The Influence of Water Injection on Two-Shaft Gas Turbine 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 Step Technique" Engineering and Development Journal, Vol. 10, No.4, December 2006. 22-Watlow products and technical support delivered worldwide," Diesel and Gas Turbine Temperatures sensors" Watllow electric Manufacturing Company,2001. 23-Gilbert Gilkes and Gordon LTD, "Technical Handbook of Two-Shaft Gas Turbine " , Gilkes, Kendal, England, 1978. 24-Philip P., Walsh," Gas Turbine Performance", Blackwell Publishing, 2nd Edition, 143 pp, 2004. 25-H., Cohen, G., F., C., Rogers and H., I., H., Saravanamuttoo," Gas Turbine Theory", 4th ed, 1996. 26-Kays, W.M., and London, A.L., "Compact Heat Exchangers", 3rdedition, McGraw-Hill Book Company, 1984. 27-Smith, E.M.," Thermal Design of Heat Exchangers ,a Numerical Approach", John Wiley and Sons, New York, 1997. 28-G.,F.,Hewitt, " Heat Exchanger Design Hand book",Begell House, New York, 1998. 29- Holman, J.,P., "Heat Transfer" 7th edition McGraw-Hill, New Yourk,1989. 30- Dittus, F.W., and Boelter, L.M.K., Univ. Calif. (Berkeley) Pub. Eng., Vol.2, pp.443, 1930. 31-McAdams, W.H., "Heat Transmission", 3rd, Ed., McGraw-Hill, New York, 1954. 32- Adrian, B.," Forced Convection: Internal Flows", Mechanical Engineering and Materials Science, Chap. 5,420 pp. 33- Sieder, E.N., and Tate, C.E., "Heat Transfer and Pressure Drop of Liquids in Tubes", Ind. Eng. Chem., Vol.28, pp.1429, 1930.
34- Briggs D.E. and Young, E.H., "Convection Heat Transfer and Pressure Drop of Air Flowing Across Triangular Pitch Banks of Finned Tubes", Chem. Eng. Prog. Symp. Ser. Vol.59, No.41, pp1-10, (1963). 35- London,A. L.,"Compact Heat Exchanger-Design Methodology", ed.Kakac,S.,Shah,R.K.and Bergles,A.E.,Hemisphere,New York,(1983). 36-Bhatti,M.S. and Shah,R.K.,"Turbulant and Transition Flow Forced Convective Heat Transfer in Ducts",Single Phase Convective Heat Transfer,Wiley,1987. 37-Horlock,J.H.,"Co-generation:Combined Heat and Power", Pergamon Press, Oxford, See also 2nd edn, Krieger, Melbourne, FL, 1987. 38-Hawthorne,W.R., and Davis, G.de V."Calculating gas turbine performance", Engineering 181, 361 -367, 1956. 39- Keenan, J.H. and Kaye, J.," Gas Tables", Wiley, New York, 1945.
,
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 15 (°C) Compressor Inlet Temperature:
Engine Parameters
Gas Generator Speed
Gas Generator Speed
(45000 RPM)
(55000 RPM)
T1 (°C)
15
15
Δp (mmwg)
32
43
Vf
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
14400
16300
(L/h)
Speed (RPM)
Table (A-2) Ambient Temperature: 25 (°C) Ambient Pressure: 1.0133 bar Compressor Inlet Temperature: 20 (°C)
Engine Parameters
Gas Generator Speed
Gas Generator Speed
(45000 RPM)
(55000 RPM)
T1 (°C)
20
20
Δp (mmwg)
30
40
Vf
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
17100
21900
(L/h)
Speed (RPM)
Table (A-3) Ambient Temperature: 29 (°C) Ambient Pressure: 1.0133 bar Compressor Inlet Temperature:25 (°C) Engine Parameters
Gas Generator Speed
Gas Turbine Speed
(45000 RPM)
(55000 RPM)
T1 (°C)
25
25
Δp (mmwg)
40
50
Vf
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
17900
22000
(L/h)
Speed (RPM)
Table (A-4) Ambient Temperature: 29(°C) Ambient Pressure: 1.0133 bar Compressor Inlet Tempertaur:30 (°C) Engine Parameters
Gas Turbine Speed
Gas Turbine Speed
(45000 RPM)
(55000 RPM)
T1 (°C)
30
30
Δp (mmwg)
39
59
Vf
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
11800
21600
(L/h)
Speed (RPM)
Table (A-5) Ambient Pressure:1.0133 bar Turbine Inlet Temperature: 586 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 15 (°C)
Temperature 22 (°C)
Engine Parameters
T3 (°C)
586
586
Gas Generator ( RPM) 55000
50000
Δp (mmwg)
43
47
Vf
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
(L/h)
Table (A-6) Ambient Pressure:1.0133 bar Turbine Inlet Temperature: 590 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 18 (°C)
Temperature 22 (°C)
Engine Parameters
T3 (°C)
590
590
Gas Generator ( RPM) 50000
45000
Δp (mmwg)
39
33
Vf
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
Power Turbine(RPM)
19800
17900
(L/h)
Table (A-7) Ambient Pressure:1.0133 bar Turbine Inlet Temperature: 620 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 20 (°C)
Temperature 30 (°C)
Engine Parameters
T3 (°C)
620
620
Gas Generator ( RPM) 55000
50000
Δp (mmwg)
44
40
Vf
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
Power Turbine(RPM)
21200
17900
(L/h)
Table (A-8) Ambient Pressure: 1.0133 bar Turbine Inlet Temperature: 625 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 15 (°C)
Temperature 30(°C)
Engine Parameters
T3 (°C)
625
625
Gas Generator ( RPM) 65000
50000
Δp (mmwg)
53
38
Vf
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
Power Turbine(RPM)
21500
13100
(L/h)
Table (A-9) Ambient Pressure: 1.0133 bar Turbine Inlet Temperature: 635 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 25 (°C)
Temperature 29 (°C)
Engine Parameters T3 (°C)
635
635
Gas Generator ( RPM) 57200
50000
Δp (mmwg)
41
39
Vf
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
Power Turbine(RPM)
22200
19500
(L/h)
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 (No. of Tubes / Row) ,EQ.(3-1) & Calculate Length of H.EX, EQ.(3-2) & Calculate Depth of H.EX, EQ.(3-3)
YES ∆Twater = 0.1, nw=0.3,na=0.4,F=1
IF (Twi >Two)
NO ∆Twater= 0.1, nw=0.4 ,na=0.3,F=-1
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
B
A
B
A 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
Calculate the Heat Load for First Slice
Calculate the Air Exit Temperature for First Slice
Correct the Fluids Properties
Calculate the Correct Air Exit Temperature for First Slice
Assume the Air Velocity Over Tubes.
Calculate the Height of Heat Exchanger
D
C
E
C
D
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) NO IF Error <=0.001 YES Calculate the Mean Air Exit Temp. Calculate the Error from Calculating Value & Assuming Value
NO
IF Error <=0.001 YES Repeat the process for all Rows (Nj)
Calculate Heat Exchanger Performance
Calculate Heat Exchanger Physical Characteristics
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)
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
ﻦ اﻟ َّﺮﺣِﻴ ِﻢ ِﺑﺴْ ِﻢ اﻟَّﻠ ِﻪ اﻟ َّﺮﺣْ َﻤ ِ
ن ب وَاﻟْﻤِﻴﺰَا َ ت وَأَﻧْﺰَﻟْﻨَﺎ َﻣ َﻌ ُﻬ ُﻢ اﻟْ ِﻜﺘَﺎ َ ﺳ َﻠﻨَﺎ ﺑِﺎﻟْ َﺒ ِّﻴﻨَﺎ ِ َﻟ َﻘﺪْ أَرْﺳَﻠْﻨَﺎ ُر ُ ﺤﺪِﻳ َﺪ ﻓِﻴ ِﻪ ﺑَﺄْسٌ ﺷَﺪِﻳﺪٌ َو َﻣﻨَﺎ ِﻓ ُﻊ ﻂ وَأَﻧْﺰَﻟْﻨَﺎ اﻟْ َ س ﺑِﺎﻟْ ِﻘﺴْ ِ ِﻟ َﻴﻘُﻮ َم اﻟ َﻨّﺎ ُ ن اﻟَّﻠ َﻪ َﻗ ِﻮيٌّ ﺐ ِإ َّ ﺳ َﻠ ُﻪ ﺑِﺎﻟْ َﻐﻴْ ِ ﺼ ُﺮ ُﻩ َو ُر ُ س َو ِﻟ َﻴﻌْ َﻠ َﻢ اﻟَّﻠ ُﻪ َﻣﻦْ َﻳﻨْ ُ ﻟِﻠ َﻨّﺎ ِ ﻋَﺰِﻳﺰٌ ﺻﺪق اﷲ اﻟﻌﻈﻴﻢ
ﺳﻮرة اﻟﺤﺪﻳﺪ اﻵﻳﺔ )(25
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:
Signature:
Name:
Name:
Date:
Date:
In view of the available recommendation I forward this thesis for debate by the examining committee Signature: Name: Date: (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. Signature: Name: Date: (Chairman) Signature:
Signature:
Name:
Name:
Date:
Date:
(Member)
(Member)
Signature:
Signature:
Name:
Name:
Date:
Date:
(Supervisor)
(Supervisor)
Approved by the Dean of College of Engineering Signature: Name: Date:
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 twoshaft 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 aircooled 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 :
8
Literature Survey
2.1 General
9
2.2 The Previous Work
9
2.2.1 Modeling and Simulation of Air Cooled Heat Exchanger 2.2.2 Performance Improvement of the Gas Turbine Engine
9 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
58
Performance 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
62
minimum ratio (Ø) 4.7.2 The effect of compressor inlet temperature on the expansion
64
ratio 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
76
Rate 5.3.3 Effect of Compressor Inlet Temperature on Specific Fuel
77
Consumption 5.3.4 Effect of Compressor Inlet Temperature on Heat
77
Consumption 5.3.5 Effect of Compressor Inlet Temperature on Heat Rate
78
5.3.6 Effect of Compressor Inlet Temperature on Overall
78
Efficiency 5.3.7 Effect of Compressor Inlet Temperature on Air Mass Flow
78
Rate 5.3.8 Effect of Compressor Inlet Temperature on Pressure Ratio
78
5.3.9 Effect of Compressor Inlet Temperature on Power Input to
79
Compressor 5.3.10 Effect of Compressor Inlet Temperature on Turbine Inlet
79
Temperature 5.4 Percent of Design
80
5-5 Comparison between the Experimental and Theoretical
81
Predictions of the gas turbine engine 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 Figure (3.3.a) Top View of Heat Exchanger Geometry
24
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
25
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
67
the Theoretical Overall Efficiency ,Expansion Ratio=1.12 Figure(4.13) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on
67
the Theoretical Overall Efficiency ,Expansion Ratio=1.08 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) 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
94
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
95
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
83
0
(l/h),Va=2000 (cfm) , Tw (in)=10 C Table (5.2) Air Temperature Distribution Along H.EX.Depth,Vw=2000 0
(l/h),Va=2000 (cfm), Tw (in)=50
84
C
Table (5.3) Air Temperature Distribution Along H.EX.Depth,Vw=2000
85
0
(l/h),Va=500 (cfm), Tw (in)=10 C Table (5.4) Air Temperature Distribution Along H.EX.Depth,Vw=2000 (l/h),Va=500 (cfm), Tw (in)=50
0
86
C
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
Table (5-7) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
89
Table (5-8) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
90
Table (5-9) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
91
Table (5-10) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
91
Table (5-11) Air Temperature Distribution a long First Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
91
Table (5-12) Air Temperature Distribution a long Second Row Height (No. of Slice in i-Direction), Core Size=55 × 3.6 × 37
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
Lf
Fin Length
mm
L
Heat Exchanger Length
cm
m& a
Air Mass Flow Rate
Kg/s
m& f
Fuel Mass Flow Rate
Kg/s
Nu
Nusselt Number =
Nt
Number of Tubes
-
Nr
Number of Rows
-
N
Number of Slices
-
P
Pressure
bar
Q&
Rate of Heat Transfer Loss or Gain
W
kJ/kg
hd k
-
ρud h μ
Re
Reynolds Number =
R
Gas Cosatant
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
−
.
J/kg.K
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 =
xt
Isentropic Temperature Ratio for Turbine =
.
T3 T4 s
T2 s T1
-
Greek Letters
ηo
= 1−
ηf
Fin Efficiency
ρ
Density (kg/m³)
μ
Viscosity (kg/m.s)
Ø
Maximum to Minimum Temperature Ratio through the gas turbine engine
α
t*
ηo
Overall Efficiency of Gas Turbine
ƒ
Fuel / Air Ratio
λt
Perimeter of Tube Side (m3)
λf
Perimeter of Fin (m3)
γ
Specific Heat Ratio
Ŧ
Torque (N.m)
Af A
(1 − η ) Overall Surface f
Efficiency
Ø η c* η
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 3
Compressor Discharge 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
Chapter One…………………………………………………………………….Introduction
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 Combustion Chamber
Power Turbine
Turbine
Compressor
Driven Equipment
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
Chapter One…………………………………………………………………….Introduction
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]
Chapter One…………………………………………………………………….Introduction
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.
Chapter One…………………………………………………………………….Introduction
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.
Chapter One…………………………………………………………………….Introduction
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.
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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
Chapter Two……………………………………………………………….Literature Survey
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.
Chapter Two……………………………………………………………….Literature Survey 160
Output Power, (M W )
150
140
130
120
110
100 260
270
280
290
300
310
320
330
340
Inlet Temp, (K)
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
Chapter Two……………………………………………………………….Literature Survey
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
EXPERIMENTAL WORK
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
Chapter Three…………………………………………………………Experimental Work
temperature. A piece of ice was immersed into cold tank so that the mixture becomes homogenous. Table (3.1) Heat Exchanger Geometry
Parameter Core
Fin
Tube
Dimension 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
Pitch (Pf)
1.46 mm
Length (Lf)
7.52 mm
Depth (Df)
15.88 mm
Thick
0.24 mm
(Tf)
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.
Chapter Three…………………………………………………………Experimental Work
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). Pitot Tube: 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
Chapter Three…………………………………………………………Experimental Work
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.
Chapter Three…………………………………………………………Experimental Work
Heat Exchanger Duct
T
Fan
P
By Pass
Rotameter
T
Rotameter
P
Valve
Hot Tank
Cold Tank
Heaters
Ice
Valve
Pump
Valve
Figure (3.1a) Schematic Diagram of the Built Rig (Heat Exchanger)
Chapter Three…………………………………………………………Experimental Work
Fan
Cold Tank
Figure (3.1b) Configuration of the Built Rig (Heat Exchanger)
Pitot-Tube Rotameter
Heat Exchanger Hot Tank
Gas Turbine
Figure (3.1c) Configuration of the Test Rig (Inlet Cooling System) Preparing to Gas Turbine
Chapter Three…………………………………………………………Experimental Work
XL Dt XT Ht
Figuer (3.2.a) Top View of Heat Exchanger Geometry
Pf Lf H
D
L Figure (3.2.b) Front View of Heat Exchanger Geometry
Chapter Three…………………………………………………………Experimental Work
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.
Pump
Valve
Rotameters
Cold Tank
Cold Tank
Valve
Hot Tank Hot Tank
Compressor
Duct
Heat Exchanger
Figure (3.3) Schematic Diagram of the Test Rig (Overall)
Ambient Air
Combustor
Chapter Three…………………………………………………………Experimental Work
Dynomometer
LP.Turbine
HP.Turbine
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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 flowmeter. 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
Chapter Three…………………………………………………………Experimental Work
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, highpressure 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
Chapter Three…………………………………………………………Experimental Work
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 eddycurrent 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.
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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 :-
Ncc Ts
=
Nc T1
.
……………………………………………………………...…(3-2) .
M a Ta Pa
=
M ac Ts Ps
……………………………………………………..…(3-3)
T2 T2c = …………………………………………………………………(3-4) Ta Ts
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:
Chapter Three…………………………………………………………Experimental Work
N cc = N c
288.16 ………………………………………………….….….(3-5) Ta .
.
M ac =
1.0133 288.16
×
M a Ta
……………………………………………..…..(3-6)
Pa
⎛ 288.16 ⎞ ⎟⎟T2 ………………………………………………………….(3-7) T2c = ⎜⎜ ⎝ Ta ⎠
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.
Ta = Ta ( o C ) + 273.16
2.
Pa =
3.
T1C = (T1 + 273.16)(
4.
P1 = Pa − 97.9 Χ10 −6 (ΔP)
5.
P2c = P2 + Pa
Barometer Re ading (mbar ) 1000 −
288.16 ) Ta
−
−
T2 c = (T2 + 273.16)(
7.
P3C = P3 + Pa
288.16 ) Ta
−
−
T3c = [T3 + 273.16](
9.
P4 c = P4 + Pa
−
bar K bar bar
6.
8.
K
K bar
288.16 ) Ta
K bar
Chapter Three…………………………………………………………Experimental Work −
288.16 ) Ta
K
−
288.16 ) Ta
K
10
T4 c = (T4 + 273.16)(
11.
T5c = (T5 + 273.16)(
12.
P5c = P4 c − 0.001333(ΔP4 5 )
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 ( rc ) rc =
P2c …………………………………………………………………..(3-8) P1c
2. Compressor Isentropic Efficiency (η c ) ηc =
T1c ( γ −1) γ (rc − 1) × 100 0 0 ……………………………….………….(3-9) T2 c − T1c
3. Air Mass Flow Rate (kg/s) m& a
T1 P1
= 0.3005
m& a c = m& a(
ΔP P1
Ta ) 288.16
............................................................................(3-10)
where : ΔP in mm wg T1 in K
Chapter Three…………………………………………………………Experimental Work
4. Fuel Flow Rate (kg/s) m& fc = ρ f
v f × 10 −3
288.16 Ta
3600
…………………………………..……….…(3-11)
where:
ρ f = ρ w ∗ sg
The specific gravity of the fuel (sg) is 0.774 5. Air/Fuel Ratio (A/F) A/F = m& ac m& fc ……………………………………………………………(3-12) 6. Combustion Chamber Temperature Rise (K) −
ΔTcct
43740 − 10 T2 c = +2 ………………………………………..…….(3-13) 1.0078 A F + 6.6
Where:
−
T2C = T2 c − 273.16
7. Combustion Chamber Pressure Loss (٪) ΔPcc = (
P2 c − P3c ) × 100 ……………………………………………….…..(3-14) P2 c
8. Gas Generator Turbine Expansion Ratio rt1 =
P3c …………………………………………………………………(3-15) P4c
9. Theoretical Power Input to Compressor ( w& ) W& C = m& aCpa(T2 c − T1c ) ……………………………………………………(3-16)
Where: cpa = 1000 J kgK
Chapter Three…………………………………………………………Experimental Work
10. Theoretical Power Output Of Gas Generator Turbine ( w& ) W& t 1 = (m& ac + m& fc )Cp g (T3c − T4c ) ……………………………………………(3-17)
where: cp g = 1150 J kgK 11. Compressor Mechanical Efficiency (η mc ) η mc = (
2 w& c ) × 100 ………………………………………………………(3-18) w& c + wt′1
12. Gas Generator Turbine Mechanical Efficiency (η mt1 ) η mt1 = (
W& C + W& t1 ) × 100 …………………………………………….….……(3-19) 2W& t1
13. Compressor Overall Efficiency (η oc ) η oc = η c ×
η mc 100
……………………………………………………..………(3-20)
14. Power Turbine Expansion Ratio P4 c …………………………………………………………………(3-21) P5c
r12 =
15. Power Turbine Isentropic Efficiency (η12 )
η12 =
T4C − T5C × 100 ………………………………………………(3-22) 1 (γ −1) γ ] T4C [1 − ( ) r12
where: γ = 1.33 16. Power Turbine Power Output ( w& ) 2πN 2τ W& t 2c = 60
Ta ……………………………………………………(3-23) 288.16
Where: N 2 power turbine speed (RPM) and τ is the torque in (N.m)
Chapter Three…………………………………………………………Experimental Work
17. Power Turbine Theoretical Power Output ( w& ) W& t 2 = (m& ac + m& fc )Cp g (T4c − T5c ) ………………………………….………….(3-24)
18. Power Turbine Mechanical Efficiency (η mt 2 )
η mt 2 = (
W& t 2 c ) × 100 …………………………………………………………..(3-25) W& t 2
19. Power Turbine Overall Efficiency (η ot 2 )
η ot 2 = η t 2 ×
η mt 2 100
…………………………………………………………….(3-26)
20. Overall Thermal Efficiency (η th ) η th = (
W& t 2 c ) × 100 …………………………………………………..…(3-27) H .V × m& fc
Where: H.V= L.C.V × 4.1868 kJ kg for kerosene used in the tests (GT-85), and L.C.V= 10300 Kcal kg the lower calorific value of the fuel used, Gilbert and Gordon LTD [23]. 21. Specific Fuel Consumption ( kg Kw.s ) sfc = (
m& fc ) × 1000 ……………………………………………………..…..(3-28) W& t 2c
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)
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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
Chapter Three…………………………………………………………Experimental Work
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 directtransfer 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.
Chapter four…………………………………………………………………………Theory
Surface Characteristics
Problem Specifications
Design Theory Procedure
Physical Properties
Optional Solutions
Evaluation Criteria
Evaluation Procedure
Optimum Solution
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.
Chapter four…………………………………………………………………………Theory
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.)
Chapter four…………………………………………………………………………Theory
Figure 4.2 Step by Step method with two Directions
Chapter four…………………………………………………………………………Theory
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.
Twi
Water Flow Tao
Lf Slice 1
Dt Tar2
Tar1
Slice 2
Tai
Two Air Flow
Figure (4.3a) Slice for inlet single Tube
Figure (4.3b) Exit of one row inlet to next row
Chapter four…………………………………………………………………………Theory
2 ………………Nr
N-1.………………….……2
I=1
J=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: .
Vw (Nt) = [ N r × v w × Acw
] ………………………………………………….(4-1)
This value will be used for the estimation of the length and depth of the heat exchanger as: L = X T × ( N t + 1) ………………………………………………….(4-2) D = [ X L × ( Nr − 1)] + Dt ………………………………………………….(4-3)
4.2.2.3 Mass Conservation Mass conservation of the water flow through the tube is simply .
.
Σ m in −Σ m out = 0 ………………………………………………….(4-4) and the mass conservation equation for the water at each slice can be written as:
Chapter four…………………………………………………………………………Theory .
.
m w(i+1,j)= m 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. .
.
.
m a(j)= ρ a ( j ) × V a ( j ) ……………………………………………………...(4-6)
4.2.2.4 Log-Mean Temperature Difference To estimate the true mean temperature difference ( ΔTm ) 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]: LMTD =
(Ta (i, j ) − Tw (i, j + 1)) − (Ta (i + 1, j ) − Tw (i, j )) ………………………….(4-7) (Ta (i, j ) − Tw (i, j + 1)) ln (Ta (i + 1, j ) − Tw (i, j ))
Ta,(i,j) ∆Tm Ta,(i+1,j)
Tw,(i,j+1)
∆Tm Tw,(i,j) 0
Atotal area
Figure 4.4 The Mean Temperature Difference Along a Single Pass
Chapter four…………………………………………………………………………Theory
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]: ΔTm = F × LMTD
…………………………………………………….(4-8)
where: F=
( R 2 + 1) ln (1 − S ) (1 − RS ) ( R − 1) ln[
2 − S[ R + 1 − ( R 2 + 1) ]
……………………………………….(4-8a)
2 − S [ R + 1 + ( R 2 + 1) ]
R=
(Th (i, j ) − Th (i + 1, j )) ……………………………………….(4-8b) (Tc (i. j + 1) − Tc (i, j ))
S=
(Tc (i, j + 1) − Tc (i, j )) ……………………………………….(4-8c) (Th (i, j ) − Tc (i, j ))
4.2.2.5 Heat Load The heat load passes through a control volume on the water side is: .
Q (i,j)= m (i,j) × Cp (i,j) × ΔT (i,j) …………………………………..…….(4-9) 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]:
U o (i, j ) =
1
tt 1 1 1 [( )+( )+( )+( )] ha (i, j ) × η o K t ( Ain Aout ) hw (i, j ) × ( Ai n Ao ut ) h f × ( Ai n Ao ut )
....(4 − 11)
Chapter four…………………………………………………………………………Theory
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: h w = 0 . 023 Re 0.8 Pr n (
k d hw
………………………………………... (4-12)
)
and cross flow area for water side (rectangular tube with semi circular ends ):
[
]
Ac w = [( H t − Tt ) × ( Dt − H t )] + ( H t − Tt ) 2 × π ……………………………..(4-13)
also Re =
dh =
Pr =
ρud h w ……………………………………..…………………….(4-14) μ 4 Acw
λt
μCp k
……………………………………..………………….….(4-15) ……………………………………..…………………....….(4-16)
where the ranges of Re and Pr are: 6000 ≤ Re ≤ 10 7 0.5 ≤ Pr ≤ 120
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
Chapter four…………………………………………………………………………Theory
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): h a = 0 . 023 Re 0 .8 Pr n (
k d ha
………………………………………….. (4-17)
)
For laminar flow, the Sieder and Tate (1930) [33] correlation can be used. ha = 1.86(Re Pr )
0 .3
0.3
k ⎛ d ha ⎞ ( ) ⎟ ⎜ D f ⎠ d ………………………………….……..….….(4-18) ⎝ ha
where cross flow area for air side: Pf
Ac a = (
2
)( X T − H t ) ………………………………………………..….(4-19)
The hydraulic diameter for the air side: dha =
4 Aca
……………………………..………………….….(4-20)
λf
λ f =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: .
ma ( j ) ma(i, j ) = …………………………………………………………..(4-21) N .
.
m (i, j ) ……………………………………..………………..……. (4-22) Ga = a ( Aca ) Ra (i, j ) =
(d h a Ga ) ………………………………..………………….… (4-23) μ a (i, j )
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]:
Chapter four…………………………………………………………………………Theory
ηf =
[tanh(m × ( (m × (
Lf 2
Lf 2
))]
………………………………..………………… (4-24)
))
where:
m= [
(ha (i, j ) × ( Pf ) ( K t × Aca )
………………………………..………………..… (4-25)
The overall surface fin efficiency is: η o = 1 − [(
Af A
) × (1 − η f )] ………………………………..………………… (4-26)
where: Af H
Aexp H
= (4 × L f × D f )(
1 ) ………………………………..……………….. (4-27) Pf
= [((π × H t ) + (2 × ( Dt − H t )) − (2 × T f )](
1 ) ………..………………… (4-28) Pf
Af Aexp A =( )+( ) ………..…………………………………. ………….. (4-29) H H H
The log- mean temperature difference and overall heat transfer coefficient is calculated to obtain the height for one slice as follow:
H (i, j ) =
Q(i, j ) (U o (i, j ) × A × F × LMTD (i, j ))
…………………………………….(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, 2 Df va ΔP = ( ρ )× f × 2 d ha
…………………………………………………..(4-31)
Chapter four…………………………………………………………………………Theory
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=
64 …………………………………………………………………...(4-32) Ra
For isosceles triangular ducts, Bhatti and Shah(1987)[36] recommend, for fully-developed turbulent flow: ƒ=
0.078 ……………………………………………………………….(4-33) ( Ra) 0.25
For the power of the fan the equation is: .
Pfan =
V air Δp
η fan
……………………..………………………………….….(4-34)
And, .
V air =
m& air
ρ air
……………………..………………………………..…….(3-35)
Thus,
Pmotor , fan =
Pfan
η motor , fan
……………………..…………………………….….(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 V&w 4. Calculating (No. of Tubes / Row) ,Eq.(4-1)
Chapter four…………………………………………………………………………Theory
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 k No., Nusselt No, Eq.(4-13) to Eq.(4-20). 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: ξ% =
(Varible cal. − Varible assu.)
Variable cal.
is converged to a value within 1 × 10 −3 .hence,
the variable here is air exit temperature or air velocity over tubes. 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)
Chapter four…………………………………………………………………………Theory
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 ( w ) and rejecting heat to a sink at low temperature (Qrej) is shown in Fig.( 4.5) as a block diagram.
Figure. 4.5 Basic Gas Turbine Engine
Chapter four…………………………………………………………………………Theory
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.
Chapter four…………………………………………………………………………Theory
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).
Chapter four…………………………………………………………………………Theory
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=
P2 − P3 ………………………………………………………..(4-37a) P2
ΔP5a =
P5 − Pa Pa
(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)
Chapter four…………………………………………………………………………Theory
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].
Chapter four…………………………………………………………………………Theory
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
Chapter four…………………………………………………………………………Theory
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.
Chapter four…………………………………………………………………………Theory
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.
Chapter four…………………………………………………………………………Theory
Theoretical Overall Efficiency (%)
8
7
6
5
4 Expansion Ratio (rT)rT=1. =1.12 Expansion Ratio 12 3 Expansion Ratio (rT)rT=1. =1.08 Expansion Ratio 08 2 3.08
3.1
3.12
3.14
3.16
3.18 T3/T1 (
Ø
3.2
3.22
3.24
3.26
3.28
)
Fig (4.8) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=699 °C
7
Theoretical Overall Efficiency (%)
6.5 6 5.5 5 4.5 4 3.5 Expansion Ratio =1.12 Expansion Ratio(rT) rrT=1. T=1. 12 12
3
Expansion Ratio(rT) rrT=1. T=1. 08 08 Expansion Ratio =1.08
2.5 2 3.08
3.1
3.12
3.14
3.16
3.18 T3/T1 (
3.2
3.22
3.24
3.26
Ø)
Fig (4.9) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=628.5 °C
3.28
Chapter four…………………………………………………………………………Theory
Theoretical Overall Efficiency (%)
6.5 6 5.5 5 4.5 4 3.5 3
Expansion Ratio r(rT) =1.12 T=1. 12
2.5
T=1. 08 Expansion Ratio r(rT) =1.08
2 3.08
3.1
3.12
3.14
3.16
3.18
T3/T1 (
3.2
3.22
3.24
3.26
3.28
Ø)
Fig (4.10) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=588 °C 6.5
Theoretical Overall Efficiency (%)
6 5.5 5 4.5 4 3.5 3 Expansion Ratio (rT) rT=1.=1.12 12 2.5
Expansion Ratio (rT) =1.0 rT=1. 08 8
2 3.08
3.1
3.12
3.14
3.16
3.18 T3/T1 (
3.2
3.22
3.24
3.26
3.28
Ø)
Fig (4.11) Theoretical Overall Efficiency as a Function of (T3/T1 Ratio) with Isentropic Efficiency (ηt,ηc =0.9) ,T3=547.5 °C
Chapter four…………………………………………………………………………Theory
7.5
Theoretical Overall Efficiency (%)
7 6.5 6 5.5 5 4.5 T1= 15 C
4 3.5
T1= 30 C
3 500
520
540
560
580
600
620
640
660
680
Turbine Inlet Temperature (T3)
Fig (4.12) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on the Theoretical Overall Efficiency ,Expansion Ratio=1.12
Theoretical Overall Efficiency (%)
5.5 5 4.5 4 3.5 3 T1= 15 C 2.5 T1=30 C 2 500
520
540
560
580
600
620
640
660
680
Turbine Inlet Temperature (T3)
Fig (4.13) The Effect of Turbine Inlet Temperature (High Pressure Turbine)on the Theoretical Overall Efficiency ,Expansion Ratio=1.08
Chapter four…………………………………………………………………………Theory
7.5
Theoretical Overall Efficiency (%)
7 6.5 6 5.5 5 4.5 T3=669 C ,
Ø=3.2
4 T3=628.5 C,Ø =3.1 3.5
T3=588 C, Ø =2.9
3 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)
Fig (4.14) Theoretical Overall Efficiency as a Function of Expansion Ratio with Isentropic Efficiency (ηt,ηc =0.9) and T1=15°C
Theoretical Overall Efficiency (%)
6
5.5
5
4.5
4 T3= 699 C, T3=628.5 C,
3.5
T3=588 C,
Ø =3.10 Ø=2.97 Ø=2.84
3 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)
Fig (4.15) Theoretical Overall Efficiency as a Function of Expansion Ratio with Isentropic Efficiency (ηηt,ηc =0.9) and T1=30°C
Chapter four…………………………………………………………………………Theory
Theoretical Overall Efficiency (%)
8
7
6
5
4 Expansion Ratio (rT) =1.12 Expansion Ratio rT=1. rT=1.1212
Expansion Ratio rT=1. 1 Expansion Ratio (rT) =1.1
3
Expansion Ratio rrT=1. Expansion Ratio (rT) = 1.08 T=1. 08 08 2 10
15
20
25
30
35
Compressor Inlet Temperature (C)
Fig (4.16) The Effect of Compressor Inlet Temperature (with Variable Expansion Ratio) on the Theoretical Overall Efficiency, T3= 669 °C 7
Theoretical Overall Efficiency (%)
6.5 6 5.5 5 4.5 4 Expansion Ratio (rT) rT= =1.12
3.5 3
Expansion Ratio (rT) r =1.1 T=
2.5
rT= =1.08 Expansion Ratio (rT)
2 10
15
20
25
30
35
Compressor Inlet Temperature (C)
Fig (4.17) The Effect of Compressor Inlet Temperature (with Variable Expansion Ratio) on the Theoretical Overall Efficiency, T3= 628.5 °C
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
Chapter Five .......................................................................................Results &Discussion
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
Chapter Five .......................................................................................Results &Discussion
(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
Chapter Five .......................................................................................Results &Discussion
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
Chapter Five .......................................................................................Results &Discussion
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
Chapter Five .......................................................................................Results &Discussion
• 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
Chapter Five .......................................................................................Results &Discussion
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).
Chapter Five .......................................................................................Results &Discussion
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.(519).The increasing of heat rate is due to increasing of both turbine inlet temperature and compressor inlet temperature. A decrease in heat rate of about (37 %) was obtained due to reducing compressor inlet temperature from (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
Chapter Five .......................................................................................Results &Discussion
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
Chapter Five .......................................................................................Results &Discussion
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 about (22 %) was obtained due to reducing compressor inlet temperature from (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 =
5 =142.8 (kW) 0.035
• The minimum heat rate The minimum heat rate can be calculated for (GT-85) as follow: heat rate=
1 =28.56 0.035
Chapter Five .......................................................................................Results &Discussion
Percent of Design =
Calculated Value at (T1 = 15 ) − Calculated Value at (T1 = 30 ) Design Value
× 100%
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
Chapter Five .......................................................................................Results &Discussion
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 (3015°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.
47.28972
62.55761
62.80377
63.05259
63.30411
84.24339
84.56969
84.89945
85.56966
85.23277
85.91026
2.328
1.164
1.164
1.164
1.164
.776
.776
.776
.776
.776
.776
96.19064
95.78031
95.37463
94.97345
94.57677
21.49
23.59
25.69
27.69
29.79
19.39
21.49
23.59
25.69
27.69
29.79
23.59
10
10
10
10
10
10
10
10
12
12
12
12
12
12
12
12
12
12
12
12
.235
.237
.238
.239
.241
.242
.238
.239
.241
.242
.241
.242
6
5
4
3
2
1
4
3
2
1
2
1
Heat Exchanger (55 3.6 37) cm3 Heat Exchanger (28.5 7.7 27) cm3
Heat Exchanger (18 11 18) cm3
×
32
25.69
25.69
10
10
10
10
×
94.18447
69.22968
27.69
27.69
29.79
27.69
29.79
Core Size (L D H)
×
68.93851
29.79
32
29. 79
32
No. of ROW
×
68.65057
68.36579
50.81485
50.60406
(mbar)
∆pa
×
47.10219
Uo(w/m2. °C) ha w/m2. °C) Ta (in) °C Ta (out) °C Tw (in) °C Tw (out) °C
×
2.328
Q (KW)
Table (5-1) Air Temperature Distribution Along H.EX.Depth ,Vw=2000 (L/h),Va=2000 (cfm)
Chapter Five .......................................................................................Results &Discussion ×
×
45.31548
60.24699
60.69621
60.46152
60.46152
81.55405
81.84826
81.53701
81.82931
82.11871
81.81271
2.284
1.142
1.142
1.142
1.142
.761
.761
.761
.761
.761
.761
90.22915
90.59176
90.24413
89.89323
90.2618
40.49
38.39
36.29
34.19
32.09
42.59
40.49
38.39
36.29
34.19
32.09
38.39
50
50
50
50
50
50
50
50
48
48
48
48
48
48
48
48
48
48
48
48
.250
.248
.247
.246
.245
.243
.247
.246
.245
.243
.245
.243
6
5
4
3
2
1
4
3
2
1
2
1
Heat Exchanger (55 3.6 37) cm3 Heat Exchanger (28.5 7.7 27) cm3
Heat Exchanger (18 11 18) cm3
×
32
36.29
36.29
50
50
50
50
×
89.90861
65.76105
34.19
34.19
32.09
36.19
34.09
×
65.25092
32.09
32
34.09
32
Core Size (L D H)
×
65.51846
65.00219
48.59657
47.79097
No. of (mbar) ROW
∆pa
×
45.04536
Ta (out) °C Tw (in) °C Tw (out) °C
×
2.284
Q (KW) Uo (w/m2. °C) ha (w/m2. °C) Ta (in) °C
Table (5-2) Air Temperature Distribution Along H.EX.Depth ,Vw=2000 (L/h),Va=2000 (cfm)
Chapter Five .......................................................................................Results &Discussion ×
×
20.13704
18.50125
18.57706
18.65369
18.73119
25.65959
25.76418
25.86992
25.97685
26.08497
26.1943
.456
.2329
.2329
.2329
.2329
.155
.155
.155
.155
.155
.155
27.07296
26.95747
26.8433
26.7304
26.61876
21.57
23.67
25.78
27.89
29.99
19.46
21.57
23.67
25.78
27.89
29.99
23.67
10
10
10
10
10
10
10
10
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
10.4
6.15
6.18
6.22
6.25
6.29
6.32
6.22
6.25
6.29
6.32
6.29
6.32
6
5
4
3
2
1
4
3
2
1
2
1
Heat Exchanger (55 3.6 37) cm3 Heat Exchanger (28.5 7.7 27) cm3
Heat Exchanger (18 11 18) cm3
×
32
25.78
25.78
10
10
10
10
×
26.50834
19.21763
27.89
27.89
29.99
27.89
29.99
×
19.1368
29.99
32
29.99
32
×
19.05688
18.97783
20.68979
20.64679
Core Size (L D H) ×
20.09592
E-02
No. of (mbar) ROW
∆pa
×
.456
Q (KW) Uo (w/m2. °C) ha (w/m2. °C) Ta (in) °C Ta (out) °C Tw (in) °C Tw (out) °C
Table (5-3) Air Temperature Distribution Along H.X.Depth ,Vw=2000 (L/h),Va=500 (cfm)
Chapter Five .......................................................................................Results &Discussion
×
×
22.79861
25.26978
26.01909
25.5097
26.04461
35.53004
34.93548
35.29956
35.51754
35.57734
36.40247
.684
.342
.342
.342
.342
.228
.228
.228
.228
.228
.228
37.36954
36.50051
36.43759
36.2082
35.82522
78.37
69.88
54.46
43.02
34.12
88.12
78.37
69.88
54.46
43.02
34.12
69.88
50
50
50
50
50
50
50
50
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
49.5
7.58
7.25
6.98
6.75
6.57
6.41
6.95
6.78
6.57
6.41
6.57
6.41
6
5
4
3
2
1
4
3
2
1
2
1
Heat Exchanger (55 3.6 37) cm3 Heat Exchanger (28.5 7.7 27) cm3
Heat Exchanger (18 11 18) cm3
×
32
54.46
54.46
50
50
50
50
×
36.45073
26.53511
43.02
43.02
34.12
43.02
34.12
×
25.98009
34.12
32
34.12
32
×
26.0863
25.73128
23.4331
23.84225
Core Size (L D H) ×
23.18965
E-02
No. of (mbar) ROW
∆pa
×
.684
Q (KW) Uo (w/m2. °C) ha (w/m2. °C) Ta (in) °C Ta (out) °C Tw (in) °C Tw (out) 0C
Table (5-4) Air Temperature Distribution Along H.X.Depth ,Vw=2000 (L/h),Va=500 (cfm)
Chapter Five .......................................................................................Results &Discussion ×
×
Chapter Five .......................................................................................Results &Discussion
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)
ha (w/m2.°C)
Q (watt)
Uo (w/m2.°C)
Ta out (°C)
A (m2)
50.60411
116.4624
47.09801
29.89904
.1187113
50.6041
116.4564
47.0985
29.89915
.1192767
50.6041
116.4504
47.09899
29.89925
.1198476
50.6041
116.4474
47.09924
29.89931
.1201254
50.6041
116.4474
47.09924
29.89931
.1201254
50.60409
116.4385
47.09997
29.89947
.1210062
50.60408
116.4326
47.10045
29.89958
.1215941
50.60408
116.4267
47.10094
29.89968
.1221879
50.60407
116.4208
47.10142
29.89979
.1227876
50.60406
116.4149
47.10191
29.89989
.1233933
50.60406
116.409
47.0239
29.9
.1240051
50.60405
116.4032
47.10287
29.9001
.1246231
50.60405
116.3974
47.10334
29.90021
.1252474
50.60405
116.3945
47.10359
29.90026
.1255509
50.60404
116.3858
47.10431
29.90042
.1265153
50.60403
116.3801
47.10478
29.90052
.1271591
50.60403
116.3743
47.10526
29.90063
.1278096
50.60403
116.3686
47.10573
29.90073
.1284669
50.60402
116.3629
47.1062
29.90083
.1291311
50.60402
116.3572
47.10667
29.90093
.1298023
Vw = 2000 l/h Va=2000cfm
Chapter Five .......................................................................................Results &Discussion
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)
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
50.81489
116.4624
47.28552
27.799
. 129798
50.81488
116.4564
47.28601
27.7991
.131507
50.81488
116.4504
47.2865
27.79921
.132205
50.81488
116.4474
47.28675
27.79926
.1329107
50.81488
116.4474
47.28675
27.79926
.133253
50.81487
116.4385
47.28749
27.79943
.133253
50.81486
116.4326
47.28798
27.79953
.1343451
50.81486
116.4267
47.28846
27.79964
.1350742
50.81486
116.4208
47.28896
27.79975
.1358114
50.81485
116.4149
47.28944
27.79985
.1365567
50.81485
116.409
47.28992
27.79996
.1373105
50.81484
116.4032
47.29041
27.80006
.1380727
50.81483
116.3974
47.29089
27.80017
.1388436
50.81483
116.3945
47.29113
27.80022
.1396233
50.81482
116.3858
47.29185
27.80038
.140001
50.81482
116.3801
47.29234
27.80048
.1412096
50.81482
116.3743
47.29281
27.80058
.1420166
50.81481
116.3686
47.29329
27.80069
.142833
50.8148
116.3629
47.29376
27.80079
.1436589
50.8148
116.3572
47.29423
27.80089
.1444946
Vw = 2000 l/h Va=2000cfm
Chapter Five .......................................................................................Results &Discussion
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
ha (w/m2. °C)
Uo (w/m2. °C)
Ta out (°C)
Q (watt)
A (m2)
48.69674
114.1267
45.37363
34.09845
.1494603
48.69672
114.1362
45.37796
34.09862
.1503539
48.69672
114.1457
45.38214
34.0988
.1512587
48.69672
114.1504
45.38418
34.09888
.1516965
48.69672
114.1504
45.38418
34.09888
.1516965
48.6967
114.1645
45.39007
34.09914
.1531029
48.69669
114.1739
45.39383
34.09932
.1540425
48.69669
114.1833
45.39747
34.09949
.1549942
48.69668
114.1926
45.40099
34.09966
.1559581
48.69667
114.2018
45.40438
34.09983
.1569344
48.69666
114.2111
45.0767
34.1
.1579234
48.69665
114.2203
45.41086
34.10017
.1589252
48.69664
114.2295
45.41394
34.10034
.1599401
48.69664
114.234
45.41545
4.10042
.1604302
48.69663
114.2477
45.41983
34.10067
.1620101
48.69662
114.2567
45.42263
34.10084
.1630657
48.69662
114.2657
45.42537
34.10101
.1641354
48.69661
114.2747
45.428
34.10117
.1652195
48.6966
114.2837
45.43056
34.10133
.1663181
48.69659
114.2926
45.43305
34.1015
.1674317
Chapter Five .......................................................................................Results &Discussion
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)
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
48.49659
114.1226
45.19305
36.19837
.1674317
48.49654
114.1267
45.19515
36.19855
.1714795
48.49653
114.1362
45.19946
36.19872
.172653
48.49652
114.1457
45.20361
36.19881
.1738432
48.49652
114.1504
45.20564
36.19881
.1744162
48.49652
114.1504
45.20564
36.19907
.1744162
48.4965
14.1645
45.21149
36.19924
.1762752
48.4965
114.1739
45.21524
36.19941
.1775177
48.49649
114.1833
45.21885
36.19958
.1787783
48.49648
114.1926
45.22234
36.19976
.1800573
48.49648
114.2018
45.22573
36.19992
.1813552
48.49647
114.2111
45.229
36.2001
.1826723
48.49646
114.2203
45.23216
36.20026
.1840092
48.49645
114.2295
45.23523
36.20035
.1853661
48.49644
114.234
45.23672
36.2006
.1860176
48.49644
114.2477
45.24109
36.20076
.1881417
48.49643
114.2567
45.24387
36.20093
.1895614
48.49642
114.2657
45.24658
36.2011
.191003
48.49641
114.2747
45.2492
36.20126
.192467
48.49641
114.2837
45.25175
36.20142
.1939538
48.4964
114.2926
45.25422
36.20128
.195464
Vw = 2000 l/h Va=2000cfm
Chapter Five .......................................................................................Results &Discussion
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
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
20.64677
116.4624
20.09579
29.89989
.2781911
20.64677
116.4564
20.09587
29.9
.2795178
20.64677
116.4504
20.09595
29.90011
.2808575
20.64676
116.4474
20.09598
29.90016
.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
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
20.73277
116.4624
20.17804
27.79996
.3081484
20.73277
116.4564
20.17811
27.80007
.3097861
20.73277
116.4504
20.17819
27.80018
.3114415
20.73276
116.4474
20.17823
27.80023
.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
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
19.86854
114.1267
19.32309
34.09974
.3509701
19.86854
114.1362
19.32381
34.09991
.353089
19.86854
114.1457
19.32451
34.10009
.3552339
19.86854
114.1504
19.32485
34.10017
.356272
Chapter Five .......................................................................................Results &Discussion
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
ha (w/m2. °C)
Q (watt)
Uo (w/m2. °C)
Ta out (°C)
A (m2)
19.52759
114.1267
18.99886
43.02019
.3620837
19.52758
114.1362
18.99919
43.02028
.3668006
19.52758
114.1457
18.99919
43.02028
.3668006
19.52758
114.1504
19.00015
43.02053
.3632253
Chapter Five .......................................................................................Results &Discussion
6 Simulation Experiment
Heat Load (kW)
5
4
3
2
1
0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
5
Entering water at 10 °C Fig (5. 1a)
Simulation Experiment
Heat Load (kW)
4
3
2
1
0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
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)
Chapter Five .......................................................................................Results &Discussion
Heat Transfer Coefficient .ha (W/m 2.C)
60
Simulation Experiment
50
40
30
20
10
0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Heat Transfer Coefficient. ha (W/m 2.C)
60
Entering water at 10 °C Fig (5. 2a)
Simulation Experiment
50 40 30 20 10 0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Entering water at 50 °C Fig (5. 2b) 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)
Chapter Five .......................................................................................Results &Discussion
Heat Transfer Coefficient ha.(w/m2.c) 50.45
50.5
50.55
50.6
50.65
50.7
50.75
50.8
50.85
Slice Number (Bar Tube)
1
2 ROW 1 ROW 2 3
4
•
Entering water at 10 °C Fig (5. 3a)
Heat Transfer Coefficient ha.(w/m2.c) 48.35
48.4
48.45
48.5
48.55
48.6
48.65
48.7
48.75
Slice Number (Bar Tube)
1
2 ROW 1 ROW 2 3
4
•
Entering water at 50 °C 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
Chapter Five .......................................................................................Results &Discussion
Heat Transfer Coefficient ha.(w/m2.c) 20.6
20.62
20.64
20.66
20.68
20.7
20.72
20.74
Slice Number (Bar Tube)
1
2 ROW 1 ROW 2 3
4
•
Entering water at 10 °C Fig (5. 4a)
Heat Transfer Coefficient ha.(w/m2.c) 19.4 19.4 19.5 19.5 19.6 19.6 19.7 19.7 19.8 19.8 19.9 19.9
Slice Number
1
2
ROW 1 ROW 2
3
4
•
Entering water at 50 °C 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
Chapter Five .......................................................................................Results &Discussion
Heat Transfer Coefficient (W/m2.c)
96.5 96 95.5 95 94.5 94 93.5 93 1
2
3
4
5
6
Row Number
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
Chapter Five .......................................................................................Results &Discussion
Overall Heat Transfer Coefficient (W/m 2.C)
50 Simulation 45
Experiment
40 35 30 25 20 15 10 5 0 0
1
2
3
5
4
Air Velocity (m/s)
Overall Heat Transfer Coefficient(W/m 2.C)
•
Entering water at 10 °C Fig (5. 6a)
50 Simulation 45
Experiment
40 35 30 25 20 15 10 5 0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Entering water at 50 °C Fig (5. 6b) 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)
Chapter Five .......................................................................................Results &Discussion
Overall Heat Transfer Coefficient (w/m2.c) 47
47.05
47.1
47.15
47.2
47.25
47.3
47.35
Slice Number (Bar Tube)
1
2
ROW 1 ROW 2
3
4
•
Entering water at 10 °C Fig (5. 7a)
Overall Heat Transfer Coefficient (w/m2.c) 45.1
45.15
45.2
45.25
45.3
45.35
45.4
Slice Number
1
2
ROW 1 ROW 2
3
4
•
Entering water at 50 °C Fig (5. 7b)
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
Chapter Five .......................................................................................Results &Discussion
Overall Heat Transfer Coefficient (w/m2.c) 20.04
20.06 20. 08
20.1
20.12
20.14
20.16 20.18
20.2
Slice Number
1
2
ROW 1 ROW 2
3
4
•
Entering water at 10 °C Fig (5. 8a)
Overall Heat Transfer Coefficient (w/m2.c) 18.8
18.9
19
19.1
19.2
19.3
19.4
Slice Number
1
2 ROW 1 ROW 2 3
4
•
Entering water at 50 °C Fig (5. 8b)
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
Chapter Five .......................................................................................Results &Discussion
Overall Heat Transfer Coefficient(W/m2.c)
86.5 86 85.5 85 84.5 84 83.5 83 1
2
3
4
5
6
Row Number
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
Chapter Five .......................................................................................Results &Discussion
27.95
Air Exit Temperature (C )
27.9 27.85 27.8 27.75 27.7 27.65
Simulation 27.6
Experiment 27.55 0
1
0.5
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Entering water at 10 °C Fig (5. 10a)
44
Air Exit Temperature (C )
43 42 41 40 39 38 37 Simulation 36 Experiment 35 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
Air Velocity (m/s)
•
Entering water at 50 °C 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)
Chapter Five .......................................................................................Results &Discussion
Exit Air Temperature (c) 26.5
27
27.5
28
28.5
29
29.5
30
30.5
Slice Number (Bar Tube)
1
2
ROW 1 ROW 2
3
4
•
Entering water at 10 °C Fig (5. 11a) Exit Air Temperature (c)
33
33.5
34
34.5
35
35.5
36
36.5
Slice Number (Bar Tube)
1
2 ROW 1 ROW 2 3
4
Entering water at 50 °C Fig (5. 11b) 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 •
Chapter Five .......................................................................................Results &Discussion
Exit Air Temperature (c) 26.5
27
27.5
28
28 .5
29
29.5
30
30.5
2 ROW 1 ROW 2 3
4
•
Entering water at 10 °C Fig (5. 12a)
Exit Air Tem perature (c) 30
32.5
35
37.5
40
42.5
45
1 Slice Number (Bar Tube)
Slice Number (Bar Tube)
1
2
Row 1 Row 2
3
4
•
Entering water at 50 °C Fig (5. 12b)
Fig(5. 12) Variation Exit Air Temperature) a long Heat Exchanger Height at Water Flow Rate 2000 (L/h), Air Flow Rate 500 cfm
Chapter Five .......................................................................................Results &Discussion
35
25
20
15
10
5
0 1
2
3
4
5
6
Row No.ofNumber Row 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
43 Water Inlet Temperature 10 C
40 Pressure Drop in Air Side (mbar)*10-2
Air Exit Temperature (c)
30
37 Water Inlet Temperature 50 C
34 31 28 25 22 19 16 13 10 0.6
0.7
0.8
0.9
1
Aspect Ratio (H/L)
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
1.1
Chapter Five .......................................................................................Results &Discussion
2.5
Power Output (kW)
2
1.5
1
0.5
T1=15 C T1=30 C
0 610
620
630
640
650
660
Turbine Inlet Temperature (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)
2.6
Fuel Mass Flow Rate (g/s)
2.5
2.4
2.3
2.2
45000 RPM
2.1
55000 RPM 2 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.16) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Fuel Mass Flow Rate
Chapter Five .......................................................................................Results &Discussion
Specific Fuel Consumption (g/kW.s)
2 1.8 1.6 1.4 1.2 1 45000 RPM
0.8 55000 RPM
0.6 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.17) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed ) on the Specific Fuel Consumption
110
Heat Consumption (kW)
108 106 104 102 100 45000 RPM 98 55000 RPM 96 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.18) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Heat Consumption
Chapter Five .......................................................................................Results &Discussion
100 90
Heat Rate (kj/kW.s)
80 70 60 50 40 30 20
T1=15 C
10 0 610
T1=30 C 620
630
640
650
660
Turbine Inlet Temperature (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)
2 1.8
Overall Efficiency (%)
1.6 1.4 1.2 1 0.8 0.6 0.4
Ti=15 C
0.2 0 610
T1=30 C 620
630
640
650
660
Turbine Inlet Temperature (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)
Chapter Five .......................................................................................Results &Discussion
0.14 0.13 Air Mass Flow Rate (kg/s)
0.12 0.11 0.1 0.09 0.08 0.07 0.06
45000 RPM
0.05
55000 RPM
0.04 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.21) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Air Mass Flow Rate
450
400
cfm
350
300
45000 RPM
250
55000 RPM
200 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.22) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Air Flow Rate
Chapter Five .......................................................................................Results &Discussion
1.44
Compression Ratio
1.39 1.34 1.29 1.24 45000 RPM
1.19 55000 RPM
1.14 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.23) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Compression Ratio
6
Compressor Work (kW)
5 4 3 2 45000 RPM
1
55000 RPM
0 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.24) The Effect of Compressor Inlet Temperature (with Variable Gas Generator Speed) on the Compressor Work
Chapter Five .......................................................................................Results &Discussion
9
HP.Turbine Work (kW)
8 7 6 5 4 3 2 T1=15 C 1 0 610
T1=30 C 620
630
640
650
660
Turbine Inlet Temperature (C )
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)
Chapter Five .......................................................................................Results &Discussion
80 70
Percent of Design
60 50 40 30 POWER OUTPUT 20 OVERALL EFFICIENCY 10
HEAT CONSUMPTION
0 10
15
20
25
30
35
Compressor Inlet Temperature (C)
Fig (5.26) The Influence Compressor Inlet Temperature (°C )on Gas Turbine Performance at (Gas Generator Speed 45000 RPM)
80 70
Percent of Design
60 50 40 30 POWER OUTPUT
20 OVERALL EFFICIENCY
10
HEAT CONSUMPTION
0 10
15
20
25
30
35
Compressor Inlet Temperature (C )
Fig (5.27) The Influence Compressor Inlet Temperature (°C )on Gas Turbine Performance at(Gas Generator Speed 55000 RPM)
Chapter Five .......................................................................................Results &Discussion
2.8 2.6
Overall Efficiency (%)
2.4 2.2 2 1.8
Theoretical,Expansion Ratio=1.03
1.6
Experimental,Expansion Ratio=1.03
1.4 1.2 500
525
550
575
600
625
650
675
700
HP.Turbine Inlet Tem perature (C)
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 lowround 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
1-Philip, K., "Power Generation Hand Book", 1st edition, McGraw-Hill Professional, ISBN: 0071396047, pp.115-117, 2002. http:// www.digitalengineeringlibrary.com 2-ISO , Gas Turbine Acceptance Tests, ISO 2314, International Organisation for Standardisation, Geneva,(1973). 3-Frank ,J. B.," GE Gas Turbine Performance Characteristics", GE Power Systems Schenectady, NY, GER-3567H, pp.6-8, 2002. 4-General Electric Energy (Oil & Gas)," MS 5001 Gas Turbine" Technical Training. 5-International Energy Outlook, DOE/EIA-0484, 2004. 6-Ganapathy,V., "Process-Design Criteria of Air Cooled Heat Exchangers" ,Chemical Engineering,pp.418-425,McGraw-Hill Publication Book,Co.,Newyork,1979. 7-Hedderich, C.P., and Kellerher, M.D. "Design and Optimization of Air Cooled Heat Exchangers", Journal of Heat Transfer, vol. 104 pp683-890, Nov. (1982). 8-Zhang, C., "Numerical Modeling Using a Quasi-Tree Dimensional Procedure for Large Power Plant", Trans. of ASME Journal of Heat Transfer, v.116, p.841, 1994. 9-Matthew, S., Layton and Joseph O'Hagan, "Comparison of Alternate Cooling Technologies for California Power Plants", Electric power research institute (2002). 10-Dohoy, J. , Dennis, N.," Numerical Modeling of Cross Flow Compact Heat Exchanger with Louvered Fins using Thermal Resistance Concept" SAE.Paper , 2006-01-0726, The University of Michigan. 11-Tarrad, A. H., Khudor, D. S., and Wahed, M.A.," A Simplified Model for 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.
12-De Lucia ,M., Lanfranchi, C., and Boggio, V., “Benefits of compressor 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. 13-Saleh, F. A., "The Influence of Water Injection on Two Shaft Gas Turbine Performance", M.Sc.Thesis, AL-Mustansiria University College of Engineering, 1996. 14-Ait-Ali,M. A., “Optimum power boosting of gas-turbine cycles with compressor inlet air refrigeration.”, Journal of Engineering for Gas Turbines and Power, Vol. 119, pp. 124–129, 2001. 15-GE Nuovo Pignone Internal DT-'N',"Gas Turbine Performance Curves",Hyundi Engineering and Construction CO, Commessa-Job 170552021,2002. 16- Donald C. Erickson and Icksoo Kyung, “Aqua Absorption Turbine Inlet Cooler”, ASME International Mechanical Engineering Congress and Exposition, Draft IMECE2003-42870,pp.113-115. 17- Hameed,N.,"The Influence of Water Injection on Two-Shaft Gas Turbine 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 Step Technique" Engineering and Development Journal, Vol. 10, No.4, December 2006. 22-Watlow products and technical support delivered worldwide," Diesel and Gas Turbine Temperatures sensors" Watllow electric Manufacturing Company,2001. 23-Gilbert Gilkes and Gordon LTD, "Technical Handbook of Two-Shaft Gas Turbine " , Gilkes, Kendal, England, 1978. 24-Philip P., Walsh," Gas Turbine Performance", Blackwell Publishing, 2nd Edition, 143 pp, 2004. 25-H., Cohen, G., F., C., Rogers and H., I., H., Saravanamuttoo," Gas Turbine Theory", 4th ed, 1996. 26-Kays, W.M., and London, A.L., "Compact Heat Exchangers", 3rdedition, McGraw-Hill Book Company, 1984. 27-Smith, E.M.," Thermal Design of Heat Exchangers ,a Numerical Approach", John Wiley and Sons, New York, 1997. 28-G.,F.,Hewitt, " Heat Exchanger Design Hand book",Begell House, New York, 1998. 29- Holman, J.,P., "Heat Transfer" 7th edition McGraw-Hill, New Yourk,1989. 30- Dittus, F.W., and Boelter, L.M.K., Univ. Calif. (Berkeley) Pub. Eng., Vol.2, pp.443, 1930. 31-McAdams, W.H., "Heat Transmission", 3rd, Ed., McGraw-Hill, New York, 1954. 32- Adrian, B.," Forced Convection: Internal Flows", Mechanical Engineering and Materials Science, Chap. 5,420 pp. 33- Sieder, E.N., and Tate, C.E., "Heat Transfer and Pressure Drop of Liquids in Tubes", Ind. Eng. Chem., Vol.28, pp.1429, 1930.
34- Briggs D.E. and Young, E.H., "Convection Heat Transfer and Pressure Drop of Air Flowing Across Triangular Pitch Banks of Finned Tubes", Chem. Eng. Prog. Symp. Ser. Vol.59, No.41, pp1-10, (1963). 35- London,A. L.,"Compact Heat Exchanger-Design Methodology", ed.Kakac,S.,Shah,R.K.and Bergles,A.E.,Hemisphere,New York,(1983). 36-Bhatti,M.S. and Shah,R.K.,"Turbulant and Transition Flow Forced Convective Heat Transfer in Ducts",Single Phase Convective Heat Transfer,Wiley,1987. 37-Horlock,J.H.,"Co-generation:Combined Heat and Power", Pergamon Press, Oxford, See also 2nd edn, Krieger, Melbourne, FL, 1987. 38-Hawthorne,W.R., and Davis, G.de V."Calculating gas turbine performance", Engineering 181, 361 -367, 1956. 39- Keenan, J.H. and Kaye, J.," Gas Tables", Wiley, New York, 1945.
,
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 15 (°C) Compressor Inlet Temperature:
Engine Parameters
Gas Generator Speed
Gas Generator Speed
(45000 RPM)
(55000 RPM)
T1 (°C)
15
15
Δp (mmwg)
32
43
Vf
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
14400
16300
(L/h)
Speed (RPM)
Table (A-2) Ambient Temperature: 25 (°C) Ambient Pressure: 1.0133 bar Compressor Inlet Temperature: 20 (°C)
Engine Parameters
Gas Generator Speed
Gas Generator Speed
(45000 RPM)
(55000 RPM)
T1 (°C)
20
20
Δp (mmwg)
30
40
Vf
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
17100
21900
(L/h)
Speed (RPM)
Table (A-3) Ambient Temperature: 29 (°C) Ambient Pressure: 1.0133 bar Compressor Inlet Temperature:25 (°C) Engine Parameters
Gas Generator Speed
Gas Turbine Speed
(45000 RPM)
(55000 RPM)
T1 (°C)
25
25
Δp (mmwg)
40
50
Vf
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
17900
22000
(L/h)
Speed (RPM)
Table (A-4) Ambient Temperature: 29(°C) Ambient Pressure: 1.0133 bar Compressor Inlet Tempertaur:30 (°C) Engine Parameters
Gas Turbine Speed
Gas Turbine Speed
(45000 RPM)
(55000 RPM)
T1 (°C)
30
30
Δp (mmwg)
39
59
Vf
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
11800
21600
(L/h)
Speed (RPM)
Table (A-5) Ambient Pressure:1.0133 bar Turbine Inlet Temperature: 586 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 15 (°C)
Temperature 22 (°C)
Engine Parameters
T3 (°C)
586
586
Gas Generator ( RPM) 55000
50000
Δp (mmwg)
43
47
Vf
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
(L/h)
Table (A-6) Ambient Pressure:1.0133 bar Turbine Inlet Temperature: 590 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 18 (°C)
Temperature 22 (°C)
Engine Parameters
T3 (°C)
590
590
Gas Generator ( RPM) 50000
45000
Δp (mmwg)
39
33
Vf
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
Power Turbine(RPM)
19800
17900
(L/h)
Table (A-7) Ambient Pressure:1.0133 bar Turbine Inlet Temperature: 620 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 20 (°C)
Temperature 30 (°C)
Engine Parameters
T3 (°C)
620
620
Gas Generator ( RPM) 55000
50000
Δp (mmwg)
44
40
Vf
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
Power Turbine(RPM)
21200
17900
(L/h)
Table (A-8) Ambient Pressure: 1.0133 bar Turbine Inlet Temperature: 625 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 15 (°C)
Temperature 30(°C)
Engine Parameters
T3 (°C)
625
625
Gas Generator ( RPM) 65000
50000
Δp (mmwg)
53
38
Vf
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
Power Turbine(RPM)
21500
13100
(L/h)
Table (A-9) Ambient Pressure: 1.0133 bar Turbine Inlet Temperature: 635 (°C) Running with
Running with
Compressor Inlet
Compressor Inlet
Temperature 25 (°C)
Temperature 29 (°C)
Engine Parameters T3 (°C)
635
635
Gas Generator ( RPM) 57200
50000
Δp (mmwg)
41
39
Vf
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
Power Turbine(RPM)
22200
19500
(L/h)
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 (No. of Tubes / Row) ,EQ.(3-1) & Calculate Length of H.EX, EQ.(3-2) & Calculate Depth of H.EX, EQ.(3-3)
YES ∆Twater = 0.1, nw=0.3,na=0.4,F=1
IF (Twi >Two)
NO ∆Twater= 0.1, nw=0.4 ,na=0.3,F=-1
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
B
A
B
A 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
Calculate the Heat Load for First Slice
Calculate the Air Exit Temperature for First Slice
Correct the Fluids Properties
Calculate the Correct Air Exit Temperature for First Slice
Assume the Air Velocity Over Tubes.
Calculate the Height of Heat Exchanger
D
C
E
C
D
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) NO IF Error <=0.001 YES Calculate the Mean Air Exit Temp. Calculate the Error from Calculating Value & Assuming Value
NO
IF Error <=0.001 YES Repeat the process for all Rows (Nj)
Calculate Heat Exchanger Performance
Calculate Heat Exchanger Physical Characteristics
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)
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