SIMULATING THE INTEGRATED SOLAR COMBINED CYCLE FOR POWER...
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CRANFIELD UNIVERSITY Gamal Elsaket SIMULATING THE INTEGRATED SOLAR COMBINED CYCLE FOR POWER PLANTS APPLICATION IN LIBYA SCHOOL OF ENGINEERING MSc THESIS
Transcript of SIMULATING THE INTEGRATED SOLAR COMBINED CYCLE FOR POWER...
CRANFIELD UNIVERSITY
Gamal Elsaket
SIMULATING THE INTEGRATED SOLAR COMBINEDCYCLE FOR POWER PLANTS APPLICATION IN LIBYA
SCHOOL OF ENGINEERING
MSc THESIS
CRANFIELD UNIVERSITY
SCHOOL OF ENGINEERING
MSc THESIS
Academic year 2006-2007
Gamal M. Elsaket
Simulating the Integrated Solar Combined Cycle forPower Plants Application in Libya
Supervisor Dr. Ossama Badr
September 2007
This thesis is submitted in partial fulfilment of the requirements for the degreeof Master of Science
© Cranfield University, 2007. All rights reserved. No part of this publicationmay be reproduced without the written permission of the copyright holder.
Abstract
The purpose of this research is to develop a mathematical code for the
integrated solar combined cycle (ISCC) power plant. The proposed design for
the ISCC includes using a solar field based on parabolic trough solar
collector. In addition, the direct steam generation (DSG) technology is used to
generate solar steam which is supplied to the steam turbine to increase the
power output during the sunny periods. The mathematical code will be used to
simulate the ISCC performance under Libyan climatic conditions. In this
research the mathematical code is used to predict the power output increase
of developing a simple gas turbine power plant to the ISCC. In addition, it is
used to evaluate the economical and the environmental benefits of this
modification. The proposed design does not include any extra fuel burning
where the main energy resource for driving the steam turbines is the waste
heat from the gas turbine and the parabolic trough solar field. The generated
electricity can be used locally to meet the annual increasing demand or can
be exported to the EU using the proposed high voltage direct current (HVDC)
network. The proposed design gives flexibility in the operation system where it
works as a conventional combined cycle during night time and it switches to
work as an ISCC during day time. The code results shows that modifying a
simple gas turbine unite to the ISCC has many advantages. It increases the
electricity output, reduces the fuel consumption per each produced MWe and
results in a significant carbon dioxide emissions reduction per each MWe.
TABLE OF CONTENTS
Page
1. Introduction ………………………………………………………………………….… (1)
1.1: Solar thermal power plants ……………………………………..…………….. (2)
1.1.1: Concentrating solar plants ………………………………….………….. (4)
1.1.1.1: Solar tower system ……………………………………...…… (5)
1.1.1.2: Parabolic dish ………………………………………………… (6)
1.1.1.3: liner Fresnel system …………………………………………. (7)
1.1.1.4: Parabolic trough system …………………………………..… (8)
1.2: Parabolic trough solar power plants ………………………………….…….. (10)
1.2.1: Sun tracking control system ……………………………….………..… (10)
1.2.2: Parabolic trough plants configurations …………………….………… (11)
1.2.2.1: Only solar mode ………………………………….……….… (12)
1.2.2.2: Hybrid systems …………………………………………....… (13)
1.2.2.3: Direct steam generation ………………………..…………… (13)
1.2.2.4: Solar desalination ……………………………….…………… (16)
1.2.2.5: Integrated solar combined cycle …………………………… (16)
1.3: Gas turbine for electricity generation ……………………………………..… (17)
1.4: Combined cycle power plants ……………………………………………..… (18)
1.5: The current situation of the Libyan power generation system …………… (21)
1.6: Drives to carry out this project …………………………………………….… (24)
1.6.1: Location advantages …………………………………………….…..… (24)
1.6.2: Electricity exporting potential ………………………………………….. (25)
1.6.3: CSP future trend and potential market ………………………….....… (26)
1.7: Why parabolic trough ……………………………………………...……….… (28)
2. The Methodology ………………………………………………………………….… (30)
2.1: The basic design………………………………………………………..… (30)
2.2: The proposed design ……………………………………………….….… (31)
2.3: The operation procedure ………………………………………………… (34)
2.4: Mathematical analysis………………………………………………….… (35)
2.4.1: Gas turbine analysis …………………………………………….… (36)
2.4.2: Solar radiation fundamentals ………………………….……….… (44)
2.4.3: Solar radiation estimation ……………………………...……….… (47)
2.4.4: Parabolic trough solar field analysis ………………….……….… (50)
2.4.5: Integrated solar combined cycle ………………………..……….. (59)
2.4.6. Economic and environmental analysis.…………….…..………... (68)
3. Results
3.1: Solution procedure and results of GTU performance.….………….…. (70)
3.2: Gas turbine subprogram validation ………………………………….…. (73)
3.3: Parabolic trough solar field analysis …………….……………….….…. (74)
3.3.1. The selected solar collector ……………………………….…. (76)
3.3.2. Solar field characteristics and operation conditions ………. (78)
3.4: Solar field performance ………..………………………………….….…. (79)
3.5: ISCC solution ………...………………….………………………….……. (83)
3.6: ISCC simulation results ………….. …………………….………………. (85)
3.6.1: The operation parameters for the ISCC ……….....……….. (85)
3.6.2: The simulation results for the ISCC ……………………....… (86)
4: Conclusions and Recommendations for further work …………………….. (89)
4.1: Conclusion ……………………………………….……………………….. (89)
4.2: the ISCC implementation ……………………………………………….. (90)
4.3 Recommendations for further work ……………………………….…. (92)
5: References ………………………………………………………………………..…. (93)
6: Appendices …………………………………………………………………..…...…. (98)
LIST OF FIGURE
Page
Figure 1.1 Solar chimney concept …………………………………………….....…..… (3)
Figure 1.2 Solar chimney prototype in Spain 50 KWe ………………...……….…..… (4)
Figure 1.3 CSP applications ……………………………………………………….....… (4)
Figure 1.4 Central solar tower ………………………………………………………..… (5)
Figure 1.5 Schematic of two types of solar thermal tower power plant ………….… (6)
Figure 1.6 Parabolic dish …………………………………………………………....…. (7)
Figure 1.7 Fresnel system elements …………………………………………….…..… (8)
Figure 1.8 Fresnel collector driving ammonia-water-chillier …………………....…… (8)
Figure 1.9 Parabolic trough system ……………………………………………….…… (9)
Figure 1.10 Sun tracking control system ……………………………………….….… (10)
Figure 1.11 Aerial view of 5 x 30 MW Solar SEGSs at California, USA …...…….. (11)
Figure 1.12 Solar thermal power plant with thermal storage system ………...…… (12)
Figure 1.13 Solar trough system with fossil fuel backup …………………………… (13)
Figure 1.14 trough plants operation systems ………………………………...…….. (15)
Figure 1.15 Direct steam generation in parabolic trough technology …………..… (15)
Figure 1.16 Parabolic trough desalination system ………………………………..… (16)
Figure 1.17 Schematic of ISCC ………………………………………...…………..… (17)
Figure 1.18 The actual and the ideal Brayton cycle ………………………...…….… (18)
Figure 1.19 Combined cycle power plant scheme ………………………………..… (19)
Figure 1.20 The heat recovery system in HRSG ………………………………..… (20)
Figure 1.21 The thermodynamic cycles of CC …………………………………..… (20)
Figure 1.22 Different power plants efficiencies ……………………………….......… (21)
Figure 1.23 Installed power plants in Libya …………………………………….….... (22)
Figure 1.24 Electricity production by type in Libya ………………………………..… (23)
Figure 1.25 The Potential of Direct Solar Radiation for the MENA ……………...... (25)
Figure 1.26 The proposed HVDC electricity network for the EU-MENA ………..… (26)
Figure 1.27 The future anticipation of energy generation measures ……...…..… (27)
Figure 1.28 Projected CSP plants ………………………………………...………..… (28)
Figure 2.1 The proposed design scheme ………………………………...………..… (33)
Figure 2.2 Gas turbine cycle ………………………………………………….....….… (36)
Figure 2.3 Gas turbine combustion chamber energy conservation ……………..… (39)
Figure 2.4 Beam and diffuse solar radiation …………………………………...….… (44)
Figure 2.5 Solar angles ……………………………………………...……………....… (46)
Figure 2.6 Thermal network for collector of solar field ……………………………… (53)
Figure 2.7 HRSG thermal analysis ……………………………………………...….… (61)
Figure 2.8 HRSG superheating section …………………………………………....… (63)
Figure 2.9 Re-feed water FV ……………………………………….………...……..… (63)
Figure 2.10 RFEH analysis …………………………………….……………....……… (64)
Figure 2.11 Deaerator thermal analysis …………………………………..……..…… (64)
Figure 2.12 Solar separator vessel thermal analysis ………………………….……. (66)
Figure 2.13 Fuel saving analysis ……………………………………………………… (69)
Figure 3.1 Gas turbine subprogram flowchart …………………………………….…. (71)
Figure 3.2 Solar field flow chart ……………………………………………………..… (75)
Figure 3.3 LS-3 collector ………………………………………………………………. (77)
Figure 3.4 Parabolic trough solar field performance ……………………………..… (80)
Figure 3.5 Solar field efficiency at selected dates ………………………………..… (81)
Figure 3.6 Solar field output at selected dates ………………………………. …….. (81)
Figure 3.7 ISCC flow chart …………………………………………………………….. (84)
Figure 3.8 HRSG steam capability GT8C2 ………………………………………….. (86)
Figure 3.9 Electricity generating during sunny periods at selected dates ………... (87)
Figure 3.10 Fuel saving and solar steam variation at 11th June …………………… (88)
Figure 3.11 Accumulated energy & fuel saving by solar field for each GTU …….. (88)
LIST OF TABLES
Table 1.1 Libyan power plants capacity ………………………………………..…….. (23)
Table 1.2 The maximum and minimum load (2006) …………..…………….……… (24)
Table 1.3 Market Potential Solar-Thermal Power Plants ………………….……….. (27)
Table 1.4 Performance data of various CSP technologies ……………….…………… (29)
Table 2.1 The design parameters of ABB GT8C at Azzwetenah ……………...….. (31)
Table 2.2 Correction factors for the Hottel method …………………………………. (49)
Table 3.1 Input data to gas turbine subprogram …………………………………….. (72)
Table 3.2 Results of gas turbine subprogram ……………………………………….. (72)
Table 3.3 Gas turbine subprogram validation ……………………………………….. (73)
Table 3.4 Solar collector's characteristics ……………………………………………. (75)
Table 3.5 Solar collector and solar field operation parameters ……………………. (78)
Table 3.6 Solar ISCC operation parameters ………………………………………… (85)
Table 4.1 The results of developing the gas turbine to ISCC ……………………… (91)
ABBREVIATIONS
AC Air compressor
ANU The Australian National University
CC Combined cycle
CHP Combined Heat and Power
CSES Center for Solar Energy Studies – Libya
CSP Concentrating Solar Power
DE The evaporator of the deaerator
DISS Direct solar steam European project
DLR German Aerospace Centre
DSG Direct steam Generation
EC European commission
ECC Equivalent combined cycleETB Engineering tool Book
EU-MENA Europe, Mediterranean North African region
FP Feed water pump
FV Flash vessel
G Electricity generator
GCC Gas Turbine Combustion chamber
GECOL General Electricity Company of Libya
GH Gas heater
GT Gas turbine
GTU Gas turbine unit
HPT High pressure turbine
HRSG Heat Recovery Steam Generation
HTF Heat Transfer Fluid
HVDC High Voltage Direct Current
ISCC Integrated Solar Combined Cycle Power Plant
LPT Low pressure turbine
LREC Libyan Renewable Energy Centre
MED Multi Effect Desalination Unit
MSF Multi Stage Flash Desalination UnitNREL National Renewable Energy LaboratoryRFWH Re-feed water heaterSEEN The Sustainable Energy and Economy NetworkSEGS Solar Electricity Generating Station
STU Steam turbine unit
SV Separator vessel
TRANS-CSP Trans-Mediterranean Interconnection for Concentrating Solar Power
TREC Trans-Mediterranean Renewable Energy Cooperation
NOMENCLATURE
A Altitude [m]
Ap Total outer area of the receiver tube [m2]
ASF
Total solar field aperture area [m2]
be The specific fuel consumption of the gas turbine unit [tonne/MWh]
BGT
Gas turbine fuel consumption [tonne/h]
C Solar collector concentration ratio [-]
Cp Specific heat [kJ/kg.°K]
Cpm h/T [kJ/kg.°K]
DB Fuel saving [tonne/h]
Dci Cover inner diameter [m]
Dco Cover outer diameter [m]
Devap Steam mass loss from the deaerator [kg/s]
DFW Mass flow of feed water [kg/s]
DK Water mass flow in plant condenser [kg/s]
DLoss Steam loss [kg/s]
Do Steam mass flow at the turbine inlet (reference point) [kg/s]
DRK Water mass flow in GH1 [kg/s]
DRK2 Water mass flow in GH2 [kg/s]
DRT The extracted steam to operate the plant deaerator [kg/s]
DSS The generated stem due to solar field contribution [kg/s]
Dti Receiver inner diameter [m]
Dto Receiver outer diameter [m]
h' Saturated water specific enthalpy [kJ/kg]
h'' Saturated steam specific enthalpy [kJ/kg]
Ib Beam solar radiation [W/m2]
Id Diffuse solar radiation [W/m2]
Isc Solar constant [W/m2]
Iso Extraterrestrial solar radiation [W/m2]
K Receiver thermal conductivity [W/m.°K]
kc Cover thermal conductivity [W/m.°K]
Ke Cover extinction coefficient [m-1
]
l Collector length [m]
M Number of collectors in each row [-]
m.
Water mass flow for each row in the solar field [kg/s]
m.SF
Water mass flow for whole solar field [kg/s]
mC Relative air mass flow for blades cooling in gas turbine unit
mf The relative fuel mass flow for gas turbine unit [kg fuel/kg air]
mgas Gases mass exhaust from gas turbine unit [kg/s]
mK Air mass flow in gas turbine compressor [kg/s]
mloss The relative air mass flow loss
n Day number of year [-]
N Number of rows of solar field [-]
n2 Cover refractive index [-]
NEGT
Gas turbine output [MW]
NEST
Steam turbine output [MW]
NFP Energy consumption by water feed pump [MW]
P Pressure [bar]
PD Deaerator pressure [bar]
PDE Deaerator's evaporator pressure [bar]
Pk Condenser pressure [bar]
PLPT Pressure at LPT inlet [bar]
PLPTO Pressure at LPT inlet for combined cycle operation [bar]
PSOSF
Design outlet pressure for solar field [bar]
Qc.v Fuel calorific value [kJ/kg]
QL Heat loss from solar collector [kW]
QSC Useful heat from solar field [kW]
QSF Nominal solar field output [kW]
Qu Useful heat gain in solar field (for each row) [kW]
R Gas constant [kJ/kg.°K]
S Specific entropy [kJ/kg.°K]
Sb Absorbed solar energy by receiver tube [W/m2.°K]
T Temperature [°C]
Ta Ambient temperature [°C]
Tbw The average temperature of gas turbine blades [°C]
Tex Exhaust Gases temperature after HRSG [°C]
Tfi Water temperature at solar field inlet [°C]
Tfo Water temperature at solar field outlet [°C]
TL Disposed water temperature [°C]
TRFW1 Re-feed water temperature [°C]
TS
Temperature for ideal process (isentropic) [°C]
Ua Wind Velocity [m/s]
UL Solar collector loss coefficient [W/m2.°K]
W Collector aperture width [m]
Wa Specific work done by gases in the GT [kJ/kg]
Wco Specific work done by cooling air in the GT [kJ/kg]
We Specific work for gas turbine unit [kJ/kg]
WK Compressor specific work [kJ/kg]
WT Total specific work of GT ( gases + air ) [kJ/kg]
XK Steam/water dryness factor [%]
Greek symbols
Efficiency [%]
ep Feed pump's electrical efficiency [%]
GH HRSG effectiveness [%]
GTU Gas turbine unite efficiency [%]
HPT High pressure steam turbine efficiency [%]
ISCC Integrated Solar Combined Cycle efficiency [%]
K Compressor efficiency (gas turbine unit) [%]
LPT Low pressure steam turbine efficiency [%]
mp Mechanical pump efficiency [%]
SF Solar field efficiency [%]
T Turbine efficiency (gas turbine) [%]
ηc.c Gas turbine combustion chamber efficiency [%]
ηG Generator efficiency [%]
ηm Mechanical efficiency [%]
ηnK Compressor polytropic efficiency [%]
D Deaerator efficiency [%]
Latitude [degree]
Angle of incidence [degree]
z Zenith angle [degree]
Declination [degree]
Slope [degree]
Surface azimuth angle [degree]
a The specific heat ratio for air
Hour angular representation [degree]
Energy loss coefficient in gas turbine due to using cooling system
Emissivity
Pressure loss coefficient
c.c Hydraulic losses coefficient in gas turbine combustion chamber
d Clear sky diffuse atmospheric transmittance
b Clear sky beam atmospheric transmittance
FP Heat gain by feed water pump [kJ/kg]
C Collector reflectance
c Cover emissivity
P Receiver emissivity,
C Cover thickness [m]
K Compression ratio in gas turbine compressor
T Expansion ratio in gas turbine
DE Relative mass flow for DE
Drain Drain water from the HRSG drum for re-feed water system
evap Relative steam mass loss from the deaerator.
FW Relative mass flow of feed water
Loss Relative steam loss in the stem boiler.
o Relative steam mass flow at the turbine inlet (reference point).
P Receiver absorbtivity
αG Excess air coefficient
ACKNOWLEGMENTS
I would like to thank Dr. Ossama Badr, my supervisor, for the time that
he has spent with me as I carried out this project. His kindness and help are
appreciated.
I express my sincere appreciation to Dr. Hussain Alrobaei from the
Higher Institution of Engineering, Libya for his guidance and insight
throughout the research. His worthy advice and his valuable academic
assistance are acknowledged.
Thanks go to the other SOE members, who have taught me throughout
my course, for their valuable suggestions and comments. The technical
assistance of Mr. Abdul Majeed Elgady and Mr. Wineas Wineas from the
General Electricity Company of Libya and Mr. Khalif Khalifa from Cranfield
University are gratefully acknowledged.
I express my thanks and appreciation to my family for their
understanding, motivation and patience.
Lastly, but in no sense the least, I am thankful to all colleagues and
friends who made my stay at the university a memorable and valuable
experience.
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1. Introduction
Parabolic trough solar power plants are the most proven system of
concentrating solar power (CSP) techniques. The nine parabolic trough solar
electricity generating system (SEGS) in California, USA illustrates the
capability of this technology to be a reliable, renewable energy resource. This
system has been operating commercially as large-scale thermal solar power
plants with a total output of 345 MW. CSP plants are promising technologies
to be the alternative clean energy resource to meet the increasing energy
demand and thus reduce the environmental impact. It is predicted that CSP
will play a significant role in providing the energy to meet the world’s energy
demands which are increasing rapidly in response to the growing economics
in both developed and developing countries. Electricity produced by CSP in
the Mediterranean and North African (MENA) region can be used to improve
the local energy production systems and can be exported to the EU. The
TRANS-CSP scheme has been introduced by the Trans-Mediterranean
Renewable Energy Cooperation. It aims at interconnecting the electricity grids
of Europe and the Mediterranean and North Africa regions, generating power
by employing CSP in MENA and exporting it to the EU using a high voltage
direct current HVDC network. The goal is to export about 700 TWh/year to the
EU by 2050. The anticipated cost is 0.05 €/kWh (DLR, 2006a).
Parabolic trough power plants can be operated in different configurations and
operating systems. They can be operated in only solar mode where the solar
collector’s array is the only energy resource for the thermal cycle.
Alternatively, they can be operated as a hybrid system, where a backup fossil
fuel boiler is used in parallel to the solar collector’s array. Most of the existing
trough plants use synthetic oils as a heat transfer medium to supply the heat
gained by the solar collectors to a Ranking cycle. However, a new concept of
direct steam generating (DSG) has been introduced, where the water is
evaporated and superheated in the solar collector tubes directly. This
operation technique results in a cost reduction of up to 26% and thermal
efficiency improvement (Zarza and Valenzuela, 2004).
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One of the most advanced operation systems is the integrated solar combined
cycle ISCC, where a solar field based on the parabolic trough technology is
coupled to a conventional combined cycle power plant. This system’s
advantages are cost reduction and operating flexibility because there is no
need to install a storage system of fossil fuel backup boilers. In this research
an investigation into the integrated solar combined cycle ISCC is carried out,
where a mathematical code has been developed to simulate the ISCC power
plant operating under Libyan climatic conditions. The mathematical analysis of
the integrated combine cycle components, and the results of the solar field
and electricity generation are outlined in this research. The aim of
implementing this research is to investigate the potential of improving the
electricity generation system locally and the potential of the available clean
energy resource.
1.1. Solar thermal power plants
The sun continuously supplies a massive amount of energy. Because of the
nature of this energy, which is spread out, it needs to be collected and
concentrated to be useable. There are many applications and techniques
where solar energy is utilised. In solar thermal power plants, solar energy is
absorbed as heat which is then transformed into electricity. Transforming the
thermal solar energy to electricity can be conducted by different approaches.
The most common techniques are concentrating solar power (CSP) plants
and the solar chimney. The CSP techniques are: solar tower, parabolic dish
and parabolic trough.
With the solar chimney, the solar radiation is converted to kinetic energy by
heating the air in an air solar collector (greenhouse). Then the heated air is
allowed to flow through a chimney located at the centre of the solar collector.
The buoyancy force of the air causes flow through the chimney. The flowing
air drives a turbine which is fixed at the entrance of the chimney to generate
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electricity. The solar chimney consists of a solar collector or greenhouse, high
constructed chimney and turbine. A storage system can be employed using
this technique to keep the plant working at night-time. The simple concept of
its storage system is to fabricate water storage beneath the absorber plate of
the solar collectors. Consequently the storage system will heat up the air and
this runs the chimney after sunset. Figure 1.1 shows the solar chimney
concept. This technology has been proven in the field by the Spanish
prototype which operated between 1986 and 1989 in Manzanares (see figure
1.2). The plant capacity was 50 kW, its chimney height 200 m, and it covers
about 46,600 m2.
This technology advantages are; it makes use of beam and diffuse radiation
so it is able to work during cloudy periods, it can work 24 hours if a storage
system is employed, the required materials to construct it are simple and
available in most regions of the world, and there is no need for cooling water
systems, so it is suitable for arid locations.
Figure 1.1 Solar chimney concept (Bernardes, 2003)
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Figure 1.2 Solar chimney prototype in Spain 50 KWe (Solar Millennium , 2007)
1.1.1. Concentrating Solar Power (CSP) plants
CSP plants provide energy with high temperatures which is used to run
conventional power cycles such as the steam turbine, gas turbine and Stirling
engine. Although CSP plants are used mostly for electricity generation, they
can, however, be used in many industrial applications. Figure 1.3 shows the
different applications for CSP systems. One of the most important boundaries
for choosing the most suitable technique for any proposed application is the
operating temperature. For example, in applications when the desired
operating temperature is above 600 °C, the suitable technique is the central
solar tower.
Figure 1.3 CSP applications (European Commission, 2004)
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1.1.1.1. Solar tower system
This technology provides a high ratio of solar radiation concentration of up to
600 which allows solar towers to achieve 1200 °C for air heating applications.
As shown in figure 1.4, the solar tower system consists of heliostat reflectors
located in circular array around the solar receiver. The reflectors track the
sun’s position to ensure directing the sunlight to a receiver. A heat transfer
medium is used in the receiver to absorb the concentrated solar energy. The
absorbed heat then is supplied to run a thermal power plant. The heat transfer
fluid in the central receiver can be water, air, molten salt or oils. Research
shows that this technique can be used to run a gas turbine where air is
pressurised first and then heated up in the receiver to 1000 °C (Alrobaei,
2006a). The solar tower is one of the proven CSP technologies in the field.
Examples of the operated solar towers are solar one and solar two in the
USA. Their capacity is 10 MWe each. Research has shown that the central
tower has a potential to be used in a wide range of applications of gas
turbines, combined cycles, CHP and some industrial processes (Schwarzbozl,
2006; and Rheinlander and Lippke, 1998). In addition projects are being
undertaking to investigate the technology potential in metal production and
hydrogen production.
Figure 1.4 Solar tower (Trieb, 2006)
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Figure 1.5 Schematic of two types of solar thermal tower power plant, showing (a) anopen volumetric receiver with steam turbine cycle and (b) a pressurized receiver with
combined gas and steam turbine cycle (Quaschning, 2003)
1.1.1.2. Parabolic dish-engine
The basic concept of this technique is to use a parabolic dish to concentrate
the solar radiation on an engine-generator set in the focal point of the
reflector. The engine can be a Stirling engine or a gas turbine. In terms of
efficiency, the parabolic dish is the most efficient technology of all solar
technologies, its peak efficiency can be as much as 29% (Trieb, 2006). The
typical diameter of the parabolic dish varies from 5 to 15 m with an output of 5
to 25 kW (DLR, 2002). This technology is suitable for decentralised power
supply and remote locations. The barriers to uptaking this technology are its
cost and proof of long term reliability. Figure 1.6 shows a parabolic dish solar
collector.
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Figure 1.6 Parabolic dish (European Commission, 2004)
1.1.1.3. Liner Fresnel system
This system consists of an array of liner reflectors to concentrate the solar
radiation on a central absorber. The absorber tube which is oriented along the
focal line of the reflectors receives the concentrated solar radiation and
converts the solar energy to heat. Figure 1.7 shows the Fresnel system
elements. Heat transfer fluid is used to absorb this energy to be used in the
proposed application. This type of collector offers good possibilities for solar
energy use and it is suitable for small- and large-scale applications. Some
prototypes have been tested. For example, in Germany a prototype of 50 kWe
was tested in 2005. Its operation temperature was 200 °C, its dimensions
were 16 m long × 4 m high and it consisted of 11 primary reflectors. Liner
Fresnel technology was used in the summer of 2006 for the first time in a real
industrial application to run an ammonia-water-chiller (see figure 1.8). One of
the advantages of this collector is that it does not need complex construction
materials.
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Figure 1.7 Fresnel system elements (DLR, 2002)
Figure 1.8 Fresnel collector driving an ammonia-water-chiller in Bergamo, Italy(PSE,2007)
1.1.1.4. Parabolic trough system
The difference between this technology and the Liner Fresnel system is that
parabolic trough system uses a parabolic shaped reflector. The concentration
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ratio can be 80 or more (Quaschning, 2003). The collected energy then
absorbed by heat transfer fluid runs inside the absorbed tube. Parabolic
trough technology supplies energy at a temperature of up to 400 °C. This
energy is supplied to run either a simple Rankin cycle or hybrid system. The
heat transfer fluid which is used to absorb the heat can be either water or
synthetic oils. Figure 1.9 shows the parabolic trough system elements.
The parabolic trough is the most proven technology in solar thermal power
plant applications thanks to the nine SEGS in the California desert, USA.
They have been running commercially for more than 20 years as large-scale
electric power plants. They are supplying 354 MWe to the southern
Californian grid and have shown that there is no doubt about the technology’s
reliability and its potential to be a competitive energy resource. Most of the
commercially proposed solar thermal power plants are planned to be operated
based on the parabolic trough system (Jones, 2007a).
Figure 1.9 Parabolic trough system (Greenpeace, 2003)
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1.2. Parabolic trough solar power plants
In this section the operation scenarios and the different installation
configuration for parabolic trough systems are explained. In addition, as this
technology uses a sun tracking control system, the used tracking system is
briefly discussed.
1.2.1. The sun tracking control system
Since only direct solar radiation can be concentrated (Jacobson, 2006)
parabolic trough systems use a sun tracking control system to ensure
maximum efficiency of the concentrating process. For parabolic trough
collectors the most appropriate control system is in a north-south oriented
rotation axis, where collectors are aligned on the north-south axis and
collectors rotate from east to west tracking the sun’s position. The control
system continuously drives the collectors from east at sunrise to west at
sunset. Small motors are used to drive this tracking system. Figure 1.10
shows the solar collector control system theory.
Figure 1.10 Sun tracking control system (Flagso, 2007)
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1.2.2. Parabolic trough plant configurations
Solar trough systems vary in configurations and operating systems. They can
be installed in solar mode only where only heat from the solar field is used to
operate the thermal cycle. However, these systems require a thermal storage
facility to ensure operation stability. Hybrid systems use different approaches.
Where the fossil fuel boiler (commonly natural gas fired) to supply the
required energy for the thermal power plant is used. Boilers are connected in
parallel to the solar field to heat up the feed water or to superheat the
generated steam in the thermal cycle. Other techniques have also been
introduced, such as solar desalination. The solar field consists of rows of
parabolic trough collectors each row consists of collectors. Figure 1.11 shows
an aerial view of a five parabolic trough power plant in the USA. The early
solar electricity power plants are shown in Appendix A.
Figure 1.11 Aerial View of 5 x 30 MW Solar SEGSs at California, USA (Solar
Millennium, 2007)
12
1.2.2.1. Solar only mode
In this configuration and operation system the only energy resource to run the
thermal plant is the solar field. There is no backup or assistance from fossil
fuels boilers. However, a thermal storage system is needed in this regime.
The average solar-operating hours are 10-12 hours during the summer. For
the remaining time the plant is operated by energy from thermal storage.
In solar only mode with storage the solar field starts running from sunrise to
supply heat to the Rankin cycle. For about 2-3 hours of solar radiation peak,
the solar field is operated to supply some energy to storage system in addition
to its primary task of running the steam turbine. When solar energy is not
sufficient to run the Rankin cycle, the storage system starts to supply some
energy to the thermal cycle. After sunset the plant runs completely on the
storage system (Herrmann, 2004). Two power plants with a capacity of 50
MW each are planned to be constructed in Spain with only solar mode. A
molten salt thermal storage system is planned to be employed at these plants
(European Commission, 2004). Figure 1.12 shows a solar thermal power
plant with a thermal storage system
Figure 1.12 Solar thermal power plant with thermal storage system(Herrmann, 2004)
13
Figure 1.13 Solar trough system with fossil fuel backup(Greenpeace, 2003)
1.2.2.2. Hybrid systems
The hybrid system solar power generation concept uses a backup fossil fuel
boiler which is used in parallel to the solar field to guarantee reliable operation
at night-time or when no solar radiation is available. Many configurations have
been introduced as hybrid systems. One fossil fuel boiler or more is used to
supply the required energy for the thermal cycle. Boilers can be used to
superheat the steam in the thermal cycle. Moreover in the hybrid systems one
solar field or more is allocated in different positions either to heat the feed
water or superheat the steam (Hosseini, 2005). Figure 1.13 shows hybrid
trough solar power plant.
1.2.2.3. Direct stem generation
The used heat transfer fluids in most of the existing parabolic trough solar
fields are synthetic oils. These oils are used as a medium to supply the
generated energy from the solar field to the thermal power plant. Heat
exchangers are used to supply this energy to water in the thermal cycle which
is usually a Rankin cycle. Figure 1.14 shows a comparison between a DSG
operation strategy and operation system with an oil HTF.
14
The concept of DSG is to use water as an HTF in the parabolic trough solar
field, so that the solar field preheats, evaporates and superheats the water
feed. Accordingly, steam can be expanded at a steam turbine directly. The
benefits of this operation strategy are cutting capital and operation costs.
Using water as an HTF results in eliminating the use of expensive synthetic
oils and eliminating the heat exchanger from the power plant. Furthermore,
the thermal efficiency of the thermal cycle is increased.
Three different operation regimes were tested by the European project DISS.
These experimental tests were carried out in southern Spain in real solar
radiation conditions and have proven the trough capability to generate steam
with good conditions for the Rankin cycle operation. The three operation
strategies are once-trough, injection system and recirculation system (Eck
and Hirsch, 2007). These operation systems are shown in figure 1.15 and
described below.
The once trough system: in this system the solar collector preheats,
evaporates and superheats the feed water as it passes along the collector. It
is the simplest system in terms of both construction and cost. However, it is
complex in its control and operation. In addition, the flow in the receiver tube
in this operation strategy involves problems with inhomogeneous
temperatures on the tube circumference, which lead to undesirable stress on
the receiver tube (Natan, 2003).
The injection system: the water is injected into several points along the
receiver tube. Experiments have shown that this regime has many problems
related to measurements and control operations system complexity.
The recirculation system: in this regime the solar collector line is divided into
two sections. The first section works to preheat and evaporate the feed water.
This section is followed by a water-steam separator. As the evaporator output
is a mixture of water and steam, the water is separated and sent back to the
15
solar collector inlet. The generated steam in the separator is fed to the second
section of the solar collector line to be superheated.
DISS results have shown that the recirculation strategy is the best system for
use in DSG operation systems. DSG operation system offers a cost reduction
of about 26% of electricity production (Valenzuela, 2005).
a. Solar system using HTF b. Solar system using DSGFigure 1.14 trough plants operation systems (Pitz-Paal, 2004)
Figure 1.15 Direct steam generation in parabolic trough technology (Valenzuela,2005)
16
1.2.2.4. Solar desalination
Research shows the potential of using solar parabolic trough systems in
seawater desalination, where the solar field is connected to seawater
desalination units such as multi-stage flash distillation (MSF) or multi effect
distillation (MED) units (Pitz-Paal, 2004). Figure 1.16 shows a parabolic
trough desalination system.
Figure 1.16 Parabolic trough desalination system (Trieb, 2006)
1.2.2.5. Integrated Solar Combined Cycle (ISCC)
The ISCC system is a combination of a solar field and gas turbine-combined
cycle. The waste heat from the gas turbine is used to generate some steam to
be expanded in a steam turbine. In addition, the solar field supplies extra heat
to the thermal cycle. The additional heat from the solar field results in
electricity generation increase during sunlight time. This combination results in
improving the overall thermal efficiency (SolarPaces, 2005). The benefits of
employing this technology are to overcome some problems related to startup
and shut down in solar power plants, reduce the capital cost and improve the
solar-to-electricity efficiency. Figure 1.17 shows a schematic of ISCC.
17
Figure 1.17 Schematic of ISCC (Greenpeace, 2003)
1.3. Gas Turbine for electricity generation
Gas turbine units have been used in many applications, e.g. electricity
generating, and operating compressors and pumps in oil industry. However,
the most common applications of gas turbines are in electricity power plants
and aircraft propulsion. In the electricity generation field, the gas turbine can
be employed as stand-alone units or with combined cycle power plants.
Electricity generating gas turbines are usually open cycle operated. The gas
turbine unit consists of air intake, compressor, combustion chamber, turbine
and gas turbine auxiliaries (Al-Hamdan, 2006). The gas turbine performance
depends on the performance of its components i.e. compressor, combustion
chamber and turbine (Lane, 2007).
With the compressor, the air is drawn at ambient conditions into the
compressor intake, where the compressor pressurises the air up to P2. Most
of the gas turbines in electricity generation use axial flow compressors.
18
In the combustion chamber, fuel is burned with air in the combustion chamber
at constant pressure. This added heat raises the temperature from T2 to the
turbine inlet temperature T3.
The hot gases are expanded from the gas turbine inlet pressure to the
ambient air pressure. The Compression ratio and turbine inlet temperature are
important parameters for gas turbine analysis. The thermodynamic cycle of
the gas turbine is known as the Brayton cycle. Four processes are employed
by the ideal Brayton cycle:
1. Isentropic compression
2. Constant pressure heat addition in the combustion chamber
3. Isentropic expansion
4. Constant pressure heat rejection
The actual Brayton cycle includes adiabatic compression, pressure drop
within heat adding process and adiabatic expansion. Figure 1.18 shows the
actual and the ideal Brayton cycle.
Figure 1.18 The actual and the ideal Brayton cycle (Huang and Gramoll, 2007)
1.4. Combined cycle power plant
Gas turbines reject gases with high temperatures; for a simple cycle gas
turbine the temperature of exhaust gases can be as high as 600 °C (Eastop
and McConkey, 1993). Moreover, the simple gas turbine (without heat
19
recovery) has a relatively low thermal efficiency. The design efficiency for
commercial advanced turbines can be 36% (ALSTOM, 2007). The average
efficiency for the whole operation life cycle is even worse. The exhaust gases
from the gas turbine unit can be used as an external boiler for a Rankin cycle,
where the heat recovery steam generator (HRSG) is used to generate and
superheat some steam which is driven to be expanded in a steam turbine
(Beasley, 1994). As a result, more electricity is generated and the overall
efficiency of the combined cycle (CC) is improved. Figure 1.19 shows the
combined cycle layout. HRSG is a heat exchanger which recovers the energy
from the hot gases stream and is used commonly in combined cycle power
plants. Figure 1.20 shows the heat recovery system in the CC where it is
divided to three main sections: heating the feed water to increase the water
temperature up to the saturated temperature, the evaporating process which
includes converting the water into steam, and the superheating section which
increases the steam temperature up to the desired state. Figure 1.21 shows
the thermodynamic cycles of the gas turbine and the steam turbine in the CC.
DKPCP
FP
ACGTG
G
ex
Condenser
HPT LPT
NEGT
NESTGCC
BGT
RFWH
HR
SG
FV
Figure 1.19 Combined cycle power plant scheme
20
Gases stream
tm
TGin
TGout
TSout
Twin
SuperheatingHeating Evaporating
TGin gases inlet temperatureTGout gases outlet temperatureTSout steam outlet temperatureTwin water inlet temperature
Heat recovery %
Te
mp
era
ture
Figure 1.20 The heat recovery system in CC (Najjar, 1996)
Figure 1.21 The thermodynamic cycles of CC (Al-Hamdan and Ebaid, 2006)
Figure 1.22 shows a comparison between efficiencies of the common power
plant systems. The combination of these two cycle gas turbines and steam
turbines improves the total cycle efficiency by up to 60% (Eastop and
McConkey, 1993).
21
Figure 1.22 Different power plants efficiencies (Spakovszky, 2007)
1.5. The current situation of the Libyanelectricity generation
Libya is an oil producing country located in North Africa. Its area is 1,750,000
km2 and most of this land is a desert. The majority of its population (6 million)
lives on the coast. Libya receives daily high amounts of solar radiation with a
daily average on a horizontal surface of 8.1 kWh/m2/day. Solar radiation
duration average in Libya is about 3500 hours/year (Saleh, 2006). The only
electricity supplier in Libya is the General Electricity Company of Libya
(GECOL) which is a nationalised company. The electricity demand is growing
rapidly (11% in 2005) due to economic growth and improving lifestyle. GECOL
has installed a number of power plants since it was established in 1984
(GECOL, 2006). Figure 1.23 shows the installed power plants in Libya. The
power sector in Libya currently relies on gas turbine and steam turbine power
plants to produce the required electricity. In previous years some small diesel
power plants used to contribute to the energy supply, especially in remote
regions. Thanks to the improvement in the network of electricity supply, diesel
power plants are no longer used. Table 1.1 shows the operating power plants
which supply electricity to the Libyan grid.
22
Figure 1.23 Installed power plants
As a first combined cycle in the Libyan power system, it is proposed to
develop the Azzawiyah gas turbine power plant (4×165 MW) to be operated
as a combined cycle. This will increase its output by about 50%. As the
existing power plant consists of four gas turbine units, each two units will
operate a new steam turbine. The first stage of this project is now ready to be
connected to the national grid (Elgady, 2007).
Libyan power generation analysis shows that about 60% of the electricity
generation is being generating by gas turbine units (GECOL, 2006). That
means that gas turbine units are being used to cover a large portion of the
base load. Figure 1.24 shows the Libyan electricity generation system by
type. The maximum and minimum loads are shown in table 1.2 for year 2006.
The maximum load was 4005 MW and the minimum load was 1691 MW.
23
Table 1.1 Libyan power plants capacity (GECOL, 2006)
Plant Fuel Type Units No. Unit Capacity MW Plant Capacity MW Operated
Steam Turbines
Al Khums Heavy/Gas 4 120 480 1982
West Tripoli Heavy 5 65 325 1976
Heavy 2 120 240 1980
Misratah Heavy/Gas 6 84.5 507 1990
Darnah Heavy 2 65 130 1985
Tubruq Heavy 2 65 130 1985
North Banghazi Heavy 4 40 160 1979
1972
Gas Turbines
Abukammash Light 3 15 45 1982
Al Khums Light /Gas 4 150 600 1995
South Tripoli Light /Gas 5 100 500 1994
North Bangazi Light /Gas 3 150 450 1995
Light /Gas 1 165 165 2002
Azzuwaytinah Light /Gas 4 50 200 1994
Al kufrah Light 2 25 50 1982
Azzawiyah Light /Gas 4 165 660 2000
West-mountain Light /Gas 4 156 624 2006
3294
Light : light Oil Heavy: heavy Oil
Energy Production by Type
0
5000
10000
15000
20000
25000
2001 2002 2003 2004 2005 Years
En
erg
yP
rod
ucti
on
GW
h
ST GT DU Total
a
Energy Production
40%
60%
ST GT
BGT gas turbine power plant, ST steam turbine power plant, DU Diesel Engine power plant
Figure 1.24 Electricity production by type (GECOL, 2006)
24
Table 1.2 the maximum and minimum load (2006) (GECOL, 2006)
Month Min load MW Peak load MW Min/Max load
January 2,260 4,005 0.56
February 1,995 3,937 0.51
March 1,773 3,778 0.47
April 1,723 3,237 0.53
May 1,702 3,535 0.48
June 1,819 3,758 0.48
July 2,283 3,738 0.61
August 2,255 3,949 0.57
September 2,218 3,783 0.59
October 1,840 3,386 0.54
November 1,691 3,385 0.50
December 1,960 3,943 0.50
It is obvious that there is a big different between the peak loads and the
minimum loads. The minimum load to the maximum load ratio varied from
47% to 61% in 2006. Due to this big difference the electricity supplier installed
a large capacity to supply electricity for peak periods. So that was the reason
for this large portion of gas turbine electricity generation. Because of the gas
turbines are suitable for peak demand, where they can be easily and quickly
connected to grid.
1.6. Drives to carry out this research on ISCC
1.6.1. Location advantages (High intensity of solar
radiation)
Libya is a sun-belt region country where a high intensity of solar radiation is
received. The direct solar radiation for flat unprotected land can be as high as
25
1800kW/m2.year. This available resource can be used by means of thermal
power plants to meet the annual increasing demand. Figure 1.25 shows the
direct solar radiation for the Mediterranean region.
Figure 1.25 Direct Solar Radiation for the Mediterranean Region (Pitz-Paal, 2004)
1.6.2. Electricity exporting potential
The Trans-Mediterranean Renewable Energy Cooperation (TREC) has a
scheme to cooperate in the field of generating electricity and desalinating
water by making use of thermal solar power plants and wind turbines. The aim
is to interconnect the electricity grids of Europe, the Middle East and North
Africa (EU-MENA) to secure energy, water and clean environment for this
region. One of their goals is to generate electricity in the sun-belt region in the
MENA and transmit this electricity to Europe. Installing a network of High
Voltage Direct Current (HVDC) will be the media used to transmit this energy
with a loss of about 10-15%. Appendix E shows the networks and
interconnection projects until 2010 in the Mediterranean region. The German
Aerospace Center (DLR) confirms the usefulness of establishing this network.
(TREC, 2007). Figure 1.26 shows the proposed electricity network for the EU-
MENA.
26
Figure 1.26 The proposed HVDC network for the EU-MENA (TREC, 2007)
1.6.3. CSP future trends and potential market
With the rapid increase in fossil fuel prices and the running out of some
conventional fuel's reservoirs, CSP is becoming more attractive. Researchers
anticipate that CSP will have the biggest share of energy production by 2050,
as shown in figure 1.27. This increasing interest in CSP has achieved
investors' confidence and governmental support. The World Bank for
instance, supports installing 2.0 GW per annum and anticipates that the solar
electricity cost will drop to 6 ¢/kWh by the year 2010 (Becker and Trieb,
2000). Table 1.3 shows some potential solar thermal power plants projects. It
is anticipated that the electricity production costs will come down to 14 €c
/kWh (in only solar mode). However, the electricity costs for the hybrid system
can be as low as 8 €c /KWh (Becker and Trieb, 2000).
27
Figure 1.27 The future anticipation of energy generation measures (DLR, 2002)
The German aerospace center study, Med-CSP, predicts that electricity
generation using CSP in 2050 will be twice as much electricity as, wind,
photovoltaic, biomass and geothermal together. Moreover, Trans-CSP study
shows that, in 2050, about 15% of the European electricity demand can be
accommodated by solar imports from the Middle East and North Africa (DLR,
2006a).
Table 1.3 Market Potential Solar-Thermal Power Plants (Solar Millennium, 2007)
Market Potential (global) for Solar-Thermal Power Plants
IEA (International Energy Agency) 20 –45 GW by 2020
Global Market Initiative
(for solar-thermal power plants)5 GW by 2015
World Bank 2 GW / year
Greenpeace/ ESTIA/ SolarPACES
(study of solar-thermal power plants)
100 GW by 2030; 600 GW by 2040
(200,000 new jobs by 2020)
US Department of Energy (DoE) 20 GW by 2020
28
1.7. Why a parabolic trough?
This technology is an appropriate technology to be used with the ISCC cycle.
The world’s largest commercial solar thermal power plants are based on
parabolic trough technology. The world’s largest nine commercial large-scale
thermal solar power plants are outlined in Appendix C. The parabolic trough
advantages over the other CSP technologies are shown in Appendix B.
Trough systems are the only ones proven in the field as large-scale
commercial units. Table 1.4 shows a comparison between the different CSP
performances. The reasons for choosing parabolic trough technology to be
used in this research are summarised in as:
- Proven commercially in the field for more than 20 years.
- Accepted technology by the World Bank.
- Reliable systems.
- Can be installed in large capacity units, i.e. 50 to 200 MW
Figure 1.28 shows the projected CSP plants.
Figure 1.28 Projected CSP plants (Becker and Trieb, 2000)
In addition, most of the solar thermal power plant projects under development
are proposed to be run by the ISCC operation system. Appendix D shows
some of the parabolic solar power plants currently under development.
29
Table 1.4 Performance data of various CSP technologies (DLR, 2006b)
CapacityUnit MW
Concen-tration
Peak SolarEfficiency
Annual SolarEfficiency
Thermal CycleEfficiency
CapacityFactor (solar)
Land Usem²/MWh/y
Trough 10-200 70-80 21% (d) 10 – 15% (d) 30 – 40 % ST 24% (d) 6-8
17 – 18% (p) 25 – 90% (p)
Fresnel 10-200 25-100 20% (p) 9 – 11% (p) 30 - 40 % ST 25 – 90% (p) 4-6
Power tower 10-150 300-1000 20% (d) 8 – 10% (d) 30 – 40 % ST 25 – 90% (p) 8-12
35% (p) 15 – 25% (p) 45 – 55 % CC
Dish-stirling 0.01-0.4 1000-3000 29% (d) 16 – 18% (d) 30 – 40 % Stirl. 25% (p) 8-12
18 – 23% (p) 20 – 30 % GT
d = demonstrated, p = projected, Solar efficiency = net power generation / incident beamradiation, Capacity factor = solar operating hours per year / 8760 hours per year
30
2. The methodology
The proposed implementation of this project is in Libya and the North African
region, an existing gas turbine power plant has been chosen and developed
to ISCC scheme. A FORTRAN code has been developed to analyse the
proposed design of the ISCC. As the ISCC is a combination of different
components; i.e. gas turbine, solar field, HRSG and steam turbine, the code
consists of some subprograms to solve each individual component of the
ISCC system. All of these components have been run together to investigate
the ISCC performance.
2.1. The basic design
In order to evaluate the benefit of developing gas turbine units to be ISCC for
the North African region, in particular the Libyan conditions, the Azzwetenah
gas turbine power plant has been chosen as a sample to be modified to an
ISCC power plant. The Azzwetenah electric power plant is a gas turbine
power plant which has been connected to the Libyan grid since 1997 and is
located in the North East region of Libya on the coastline. The gas turbine
power plant consists of 4 units each unit producing 51 MW. The used gas
turbine engine at the Azzwetenah power plant is GT8C, manufactured by the
Swiss company ABB (Elgady, 2007). The gas turbine (GT) unit is designed to
be capable for CHP and CC applications.
The axial flow compressor of GT8C has 12 compression stages. The
compression ratio is 15.7. The combustion chamber has 19 burners and it
increases the gases temperature at turbine inlet to 1100 °C, the combustion
chamber can run either on light oil or natural gas. The turbine consists of three
expansion stages. A turbine blades cooling system is employed where some
air is extracted from the compressor and directed to cool down the turbine
blades without entering to the combustion chamber.
31
Table 2.1 the design parameters of ABB GT8C at Azzwetenah (GECOL, 2007)
Manufacturer ABB
Model GT8C
Unit Output 51 MW
Total output 204 MW
Frequency 50 MHz
Electricity efficiency 32.3 %
Compressor pressure ratio 15.7
Turbine inlet temperature 1100 °C
Number of compressor stages 12
Number of turbine stages 3
Exhaust gas flow 200 kg/s
Exhaust gas temperature 497 °C
The electric generator is driven from the cold end of the gas turbine engine
(compressor side) to enable users to use the high temperature exhaust gases
in either CHP or CC. The generator frequency is 50 MHz and rotates at 6200
rpm. Table 2.1 shows the design parameters of the gas turbine engine GT8C
used at the Azzwetenah power plant.
2.2. The proposed design
The proposed design of the ISCC is shown in figure 2.1. It is an integration
between a conventional combined cycle power plant (gas & steam turbine)
and solar field, based on a parabolic trough solar collector. HRSG is one of
the CC components. It is used to recover the heat loss from the gas turbine
exhaust gases. Most advanced electricity generation gas turbines are capable
of being connected to heat recovery units. The main components of the
proposed ISCC are: gas turbine unit, HRSG, steam turbine unit, and solar
field based on parabolic trough technology.
The gas turbine unit is the major energy resource for the Rankin cycle. The
gas turbine components and the basic design parameters are explained in
Sections 1.3 and 2.1 respectively. The other ISCC components are as follows:
32
Figure 2.1 The proposed design scheme
33
HRSG, which is a heat exchanger used to recover heat from hot gases
streams (commonly used with gas turbines). The HRSG consists of three
main sections, i.e. superheating section, evaporator and economiser. The
economiser increases the feed water temperature to the saturation
temperature to recover as much heat as possible from the gases stream.
Then the steam generator (evaporator) converts the feed water to
saturated steam at the HRSG drum's pressure. The superheating section
increases the steam temperature to the desired temperature (HRSG,
2007). The proposed design includes using two gas heaters. The first gas
heater GH1 preheats water in the HRSG. In addition an evaporator is
used in the HRSG to generate some steam to supply some energy for the
deaerator operation. The aim of using DE is to minimise the steam
extraction from the steam turbine and maximise the heat recovery and
electricity production. In the proposed design the deaerator 's evaporator
DE converts the water to steam with a steam to water ratio of 65%:35%
to avoid problems related to the two phase flow. Another gas heater is
used in the HRSG for ISCC operation regime GH2.
Steam turbine unit. In combined cycles, steam turbines are the same as
the conventional steam turbines the only difference being they use HRSG
as an external boiler. The conventional steam turbine unit consists of
steam turbine, condenser and feed water system.
Solar field. The type of solar collector used in the proposed design is the
parabolic trough collector. The sun tracking control system drives the solar
collectors to track the sun position. The collectors are aligned on the
North-south axis. A separator vessel is used to circulate water in the solar
field, using the supplied energy by the solar field to generate steam.
34
2.3. The operation strategy
The power plant works as a conventional combined cycle during periods when
there is no solar radiation. When solar radiation is available the power cycle
works as ISCC. In the ISCC regime the solar field starts supplying energy to the
thermal cycle from sunrise to sunset. For the proposed design it is assumed that
the operation conditions for the HRSG (except GH2), HPT, and feed water
system are the same as those of the combined cycle operation.
The solar field operation system is assumed to be a recirculation system with a
slight difference. As has been described, in the recirculation operation system
the steam is separated by separator after the first section of the solar field, after
which the steam is sent to the solar superheating section. In this proposed
design there is no solar superheating. The generated steam is supplied to LPT.
The design working pressure of the SV is equal to the LPT pressure inlet in the
CC operation regime. The absorbed heat by solar field is supplied to the
separator vessel SV resulting in the generation of some steam in the separator
vessel. So, as the solar radiation is increased the generated steam in the SV is
also increased. As a result of this operation system, the affected parts by the
ISCC operation system in this proposal are LPT and condenser and GH2. The
electricity generation is consequently increased as solar radiation is increased.
The condensed water mass flow in the condenser DK at the CC operation regime
is equal to the remaining steam after extracting some steam to operate the
deaerator (DK =DRK =Do-DRT). During day time operation, the condensed water is
equal to the sum of the previous value and the generated steam in the SV is (DK
=DRK + DSS). In both operation regimes DRK is supplied to GH1. As DSS varies
with solar radiation intensity, the condensed water mass flow varies. So the
operation conditions of GH2 vary with time according to solar radiation intensity.
The solar field feed water is supplied from the SV, so its properties are equal to
the saturated water properties at the SV pressure. The solar field feed pump
35
increases the pressure up to the output pressure. An additional pressure is given
to overcome the pressure loss due to fluid flow.
Pressurising the water to high pressure ensures that the outlet temperature of
HTF is equal to the saturated water temperature at the outlet pressure or less.
Consequently stratification in the solar field tubes is avoided. The water mass
flow in the solar field tubes is determined from the solar field nominal capacity. If
the outlet temperature goes above the saturated water temperature of the outlet
pressure the mass flow increases to decrease the outlet temperature to the
desired value.
The reason for choosing this particular design is that it provides flexibility in
operation procedure. The plant is operated as a conventional combined cycle at
night-time. As the solar radiation is increased the solar field starts contributing in
energy supply to the thermal cycle, resulting in generating some steam at the SV.
This steam is supplied to LPT causing an electricity generation increase.
So the configuration advantages are cost reduction potential due to DSG
operation system use, and operating flexibility, combined cycle at night-time, and
ISCC when solar radiation is available. In addition no storage system is required
in this configuration.
2.4. Mathematical analysis of the integrated solarcombined cycle
The approach used to analyse the different components of the ISCC power plant
is explained in this section. It includes a thermodynamics analysis of the simple
cycle gas turbine, mathematical analysis for beam solar radiation estimation,
parabolic trough solar field analysis, and HRSG and steam turbine breakdown.
36
2.4.1. Gas-turbine thermodynamics analysis
In the ISCC cycle the main purpose of analysing the gas turbine unit is to
evaluate the waste energy within the exhaust gases. Estimating the exhaust
gases mass flow and its temperature is the main goal of the GT mathematical
solution. The procedure to achieve this goal is to evaluate the compressor, the
combustion chamber and turbine performances.
As shown in figure 2.2, subscripts 1, 2, 3 and 4 refer to the states of air and gas
at different stages of gas turbine cycle, while superscript s refers to the isentropic
states.
The axial compressor: as the air is passed through the compressor's intake some
pressure losses occur. So the pressure at the first stage is less than the pressure
at the compressor intake entrance. Experiments show that this loss can be
evaluated as (Alrobaei, 1998):
0.015):(0.01P1 bar
Figure 2.2 Gas turbine cycle
37
ΔP1: the hydraulic losses due to air flow through the compressor intake
Hence:
The pressure at the first compressor stage is equal to (Al-Hamdan, 2006):
P-PaP 11 ……………………………………………………………..………. (2.1)
The air temperature at the compressor entrance is assumed to be equal to the
ambient air temperature T1=Ta
The compressor pressure ratio is equal to:
1
2
P
PK ………………………………………………………………..…..….…. (2.2)
The ideal gas turbine thermodynamic cycle is known as the Brayton cycle which
is described by four processes: isentropic compression, constant pressure heat
addition, isentropic expansion and constant pressure heat release. So the first
step is to calculate the air conditions at the compressor exit.
a
a
P
P
T
TS
1
1
2
1
2
………………………………………………………….……..…… (2.3)
where:
a: specific heat ratio or isentropic expansion factor.
The compressor isentropic efficiency is equal to (Eastop and McConkey, 1993):
1
1
1
1
2
1
1
2
kna
a
a
a
P
P
P
P
K
………………………………………………………..….…… (2.4)
12
12
hh
hh S
K
………………………………………………………..………..…… (2.5)
ηK: compressor efficiency
ηnK: compressor polytropic efficiency
For gas turbine applications ηnK=0.9 to 0.91 (Alrobaei, 1998)
So the actual air condition after the compression process is evaluated.
38
K
S hhhh
12
12
aaCpmR /1
1*
…………………………………….……………………..…… (2.6)
Cpma= f (T1, T2S)
Where: Cpma= h/T for air (IWAI, 2003).
The first value of is assumed to be equal to the specific heat ratio of air , then
the solutions for equations (2.3) to (2.6) are carried out, after which a new value
for is calculated from equation (2.6). An iteration process is carried out until a
desired accuracy is met |*-|<0.0001. Consequently the actual conditions at the
end of the expansion process are achieved.
The compressor specific work is given by:
12 hhWK ………………………………………………………………….….… (2.7)
The combustion chamber:
After the compression process some air is extracted for the air cooling system.
The extracted air is used for the internal cooling system of turbine blades. An
experimental correlation is used to estimate the relative mass flow for cooling air
to the entire air mass flow in the gas turbine (Alrobaei, 1998).
1000/3600T-T0.000320.02m bw3C …………………………………… (2.8)
Where:
mC : relative air mass flow rate for blades cooling
Tbw : mean temperature of turbine blades, its typical value varies from 750 °C to
850 °C (Alrobaei, 1998)
Assuming that the air to fuel ratio for combustion Lo=15 kg air/kg fuel (Alrobaei,
1998), the energy conservation for the combustion chamber is carried out as
described in figure 2.3 (Alrobaei, 1998) .
39
Figure 2.3 Gas turbine combustion chamber energy conservation
h.)mm-m-(1..Qcvmh)mm-m-(1 3flosscc.cf2flossc ……………… (2.9)
where:
mf: fuel mass flow rate
mloss: relative air losses mass flow, its typical value 0.005 kg air/kg air
ηc.c: combustion chamber efficiency, its typical value 0.9 to 0.98 (Alrobaei, 1998)
The equation (2.9) is then reformed to:
3..
23 )()1(
hQ
hhmmm
ccvc
losscf
air
fuel
kg
kg
where:
h3= f (T3, G)
αG: excess air coefficient.
h3 is evaluated as a function of the turbine inlet temperature and the excess air
coefficient. The initial value of G is guessed. Then a new value for G is
calculated by (Alrobaei, 1998).
Lom
mm
f
losscG
)1(* …………………………………………..….…………… (2.10)
Finally an iteration process is carried out until a desired accuracy is met.
40
The ideal heat adding process is a pressure constant process. However, the
actual process includes some pressure drop. So in order to estimate the
pressure at the turbine inlet the pressure loss in the combustion chamber is
evaluated (Al-Hamdan, 2006):
P3=P2× (1-c.c) ………………………………………………….………………… (2.11)
where:
c.c is the hydraulic losses coefficient within the gas turbine combustion chamber,
its typical value 0.015 to 0.025 (Alrobaei, 1998)
The turbine:
The product’s gases from the combustion process are expanded in the gas
turbine. The cooling air is driven to cool the turbine blades and then expanded in
the turbine where it is mixed with the product’s gases. So the cooling air is
expanded at the gas turbine with a different expansion ratio. The total turbine
work is the total of work done by the gases’ expansion and air expansion. The
excess air coefficient is increased at the exit of the turbine as a result of mixing
the cooling air with the exhaust gases.
The gases are expanded from the turbine inlet pressure to the ambient air
pressure. Some hydraulic losses are taken into consideration:
PPaP 44 ……………………………………………………………..………. (2.12)
ΔP4: hydraulic resistance after the turbine, its typical value depends on the
conditions after the turbine exit (Alrobaei, 1998):
ΔP4 =0.02-0.03 bar if the turbine is connected to heat exchanger or HRSG.
ΔP4 =0.005-0.001 bar if exhaust gases are sent to stack.
The gases expansion ratio:
4
3
P
PT ………………………………………………………………….………. (2.13)
41
Estimating the gases conditions at the turbine exit (Eastop and McConkey,
1993):
G
G
Tn
G
G
P
P
P
P
T
1
4
3
1
4
3
1
1
………………………………………….……………… (2.14)
where G is the heat capacity ratio for product gases
SThh
hh
43
43
…………………………………………………………….……… (2.15)
ηT: turbine isentropic efficiency
ηnT: the turbine polytropic efficiency, its typical value 0.84 to 0.87 (Alrobaei, 1998)
So the actual air condition after the expansion process is evaluated.
STa hhhh 4334 .
CpmG= f (T3, T4S, αG)
Where:
h4a: gases’ specific enthalpy at the end of expansion process.
G
G
CpmR /1
1*
……………………………………………………………… (2.16)
CpmG =h/T for gases
Similar to the compression process G is obtained by first guess an iteration
process.
The relative turbine work for product gases, without taking into account the air
cooling system, is equal to (Alrobaei, 1998):
4a3flossc h-h.mm-m-1Wa ……………………………..………….…… (2.17)
Air cooling system analysis (Alrobaei, 1998):
3
4aT-1
T……………………………………………..……………….……… (2.18)
co2w2CO .T-TTT …………………………………………………….……… (2.19)
42
where:
T4a: gases temperature at the end of the expansion process without taking into
account the cooling system effect.
co: cooling system effectiveness. Typical value 0.42 (Alrobaei, 1998)
The relative quantity of extraction heat in the cooling system
)h-(hcomQco 2c …………………………………………………………… (2.20)
The relative work of the expansion process for product gases in the gas turbine
taking into account the effect of cooling air system
.-WaWac Qco ……………………………………………………………… (2.21)
The expansion ratio for cooling air within the turbine is then evaluated (Alrobaei,
1998):
Tco 1co ………………………………..………………………….…..… (2.22)
where:
co: expansion coefficient of cooling air
Then the previous equations of the expansion process are used to estimate the
air temperature at the end of expansion process Tcoa.
The relative work of the cooling air expansion in the gas turbine is calculated
(Alrobaei, 1998):
)h-(hmW CO4COCCO ……………………………………….……….……… (2.23)
where:
hCO: enthalpy of cooling air before the expansion process
hCO4: enthalpy of cooling air after the expansion process
Total relative work of the gas turbine:
coac WW WT ……………………………………………………………….… (2.24)
The net gas turbine output is equal to the difference between the turbine work
and compressor work:
43
We = WT-Wk ……………………………………………………..…………….… (2.25)
For a given capacity of the gas turbine the required air mass flow is estimated:
Gm
GT
KWe
NEm
..
1000 ………………………………………………………….…… (2.26)
NEGT =the gas turbine unit output
The gas turbine fuel consumption is calculated:
.mmB kfGT …………………………………………………….…………...… (2.27)
This fuel consumption based on natural gas.
The mass flow rate of exhaust gases from the gas turbine unit
)mm-(1mm flossKgas ………………………………………………….….… (2.28)
As the cooling air is mixed with the produced gases from the combustion
chamber within the expansion process, the final exhaust gases’ parameters must
be evaluated (Alrobaei, 1998):
)mm-(1
.hmh)mm-(1h
floss
co4C4afloss4
……………………….………………..….… (2.29)
Lo.m
)m-(1
f
lossG …………………………………………………………..……… (2.30)
The specific fuel consumption of the gas turbine unit
GTNE
GTBbe ………….. ………………….………………….……..……..… (2.31)
The gas turbine unit efficiency is equal to the net output divided by the energy
input to the thermal cycle.
vcGT
GT
GTUQB
NE
..
3600 ……………………………………………..…..….…….… (2.32)
44
2.4.2. Solar radiation fundamentals
Solar constant, the solar constant is the solar radiation intensity on a surface
normal to the sun ray's path at the mean sun-earth distance above the
atmosphere. Solar constant has some evaluations. The value used in this
research is Isc =1367 w/m2 (ASHRAE, 2003) which is the adopted value of the
World Radiation Center.
Extraterrestrial radiation, The earth rotates around the sun in an elliptical orbit.
This movement results in variation in an earth-sun distance by 1.7%. Therefore,
the extraterrestrial radiation varies in a range of ±3 w/m2. The extraterrestrial
radiation can be calculated as below (Duffie, 1991)
nIscIso
365
360cos033.01 …………………………………………………. (2.33)
Iso is the extraterrestrial radiation at n day number of the year.
Beam radiation, the solar radiation component which is received without being
scattered or absorbed has been described in figure 2.4.
Diffuse radiation, the solar radiation component which has been scattered by
the atmosphere.
Total solar radiation, the total amount of beam and diffuse radiation.
Figure 2.4 Beam and diffuse solar radiation (ANU, 2007)
45
Solar angles: It is important to introduce some definitions for beam radiation
angles. Figure 2.5 shows some of these angles. These angles describe the
relationship between the oncoming sun radiation from the sun and any plane on
the earth with a specific position.
Latitude (), represents the location, north or south the equator.
South -90≤≤90 North
Declination (), the sun position at solar noon. The axis of the earth (North-south
pole) is tilted related to the earth’s orbit around the sun at an angle of 23.45°.
This angle varies each day and can be calculated as below (ASHRAE, 2003)
365
248360sin45.23
n ……………………………………………..…………. (2.34)
Where n is the day of year, calculated starting from 1st January n=1 to 31st
December n=365.
Slope (), the angle between the horizontal surface and the inclined plane.
Surface azimuth angle (), the angle between the projection of the plane in
question and the south direction.
East -180≤≤180 West
Hour angle (), the angular presentation of hour for solar time (Duffie, 1991)
=0 at 1200 solar noon, before 1200 -180≤≤180 after 1200
The equation (2.35) is used to convert the solar hour to angular hour
= (Hour-12) × 15 ………………………………………………...……….……. (2.35)
Zenith angle (z), the angle between the oncoming beam radiation and the
normal on the horizontal surface
46
Angle of incidence (), the angle between the oncoming beam radiation and the
normal on an inclined surface
Figure 2.5 Solar angles (Duffie, 1991)
The relation between the angle of incidence, the solar position angles to the
studied plane is given by the equation (ASHRAE, 2003):
(2.36)....................................................................sinsinsincos
coscossinsincoscoscoscoscos
cossincossin-cossinsincos
For horizontal surface =0 the incidence angle equal to the Zenith angle:
sinsincoscoscoscos Z ….…………………………...……... (2.37)
Sunrise and sunset calculations:
For unprotected flat land sunrise and sunset times can be calculated from
equation (2.39) (Duffie, 1991),
The angular hour of sunrise and sunset is s:
47
tan.tancos 1 s ……………………………………...…………..………. (2.38)
Sunrise time and sunset time are given by equations (2.40a) and (2.40b):
1512 sHsr
…………………………………………………………….……… (2.39)
1512 sHst
……………………………………………………………….…… (2.40)
2.4.3. Solar radiation estimation
In order to evaluate the performance of the solar field and its contribution to the
combined cycle, it is necessary to estimate the solar radiation intensity from
sunrise to sunset. So the first step is to calculate the sunrise and sunset times at
the corresponding date and location. The design point is selected at period
where solar radiation intensity is high (in summer). The selected point is chosen
to be 1200 on the 17th of June. The proposed location is Azzuitenah, Libya
(Altitude 100 m, latitude 32).
Sunrise and sun times:
Declination is given by equation (2.34)
365
248360sin45.23
n
Where n is day of year, calculated starting from 1st January n=1 to 31st December
n=365.
For unprotected flat land the sunrise time is given by equation (2.39):
tan.tancos 1 s
1512 sHsr
Then the Zenith angle is given by equation (2.37).
48
The sun tracking system
The employed control system in this research tracks the sun’s position to achieve
the optimum slope angle for the collector's aperture. The optimum angle is given
by (Alrobaei, 1998). It is assumed that the aperture is turned towards the east
before noon (=-90) and turned towards the west after midday (=+90). At midday
(=0) the collector aperture is in horizontal position (=0); this process is shown
in figure 1.10. As result of the operation system, equation (2.36) can be written
as.
z
opt
cos
sin.costan 1 ……………………………………………….…... (2.41)
..............................................................................................sin.sin.cos
cos.cos.cos.cossin.sin.sincos
(2.42)
Solar radiation estimation: the adopted methodology to estimate the solar
radiation intensity in this research is the Hottel method (Hottel, 1976). Hottel has
presented correlations to estimate the atmospheric transmittance for four climate
types. The correlations take into account the zenith angle and altitude for
standard atmosphere.
z
bCos
KsEXPaa
10 ………………………………………...………... (2.43)
bd a 2939.0271.0 0 …………………………………………….…………. (2.44)
where:
b: atmosphere transmittance for clear sky beam radiation
d: atmosphere transmittance for clear sky diffuse radiation
a0, a1, and ks are constants for the standard atmosphere.
To calculate constants for different altitudes, corrections factors are used:
49
)45.......(......................................................................5.201858.02711.0
)45.........(......................................................................5.600595.05055.0
)45(................................................................................600821.04237.0
2*
2*
1
2*
0
cAKs
bAa
aAa
Then:
0
*
00 raa
1
*
11 raa
KSSS rKK *
The correction factors are given for different climate types in table 2.2.
Extraterrestrial solar radiation is given by equation (2.33). Clear sky beam and
diffuse radiation are given (Hottel, 1976).
zbIsoIb cos.. ……………………………………………………………..… (2.46)
zdIsoId cos.. …………………………………………………………..….. (2.47)
Table 2.2 Correction factors for the Hottel method (Hottel,1976)
Climate type r0 r1 rKs
Tropical 0.95 0.98 1.02
Midlatitude summer 0.97 0.99 1.02
Subarctic summer 0.99 0.99 1.01
Midlatitude winter 1.03 1.01 1.00
50
2.4.4. Solar collector and solar field mathematicalanalysis.
To determine the absorbed energy by receiver tube, it is necessary to calculate
the over all optical efficiency for the solar collector. A single glazed cover is used
to reduce heat losses from the receiver tube. The cover reflection, absorption
and transmittance calculation is given by (Duffie, 1991).
1
2
112
sinsin
nn
………………………………………………………..……. (2.48)
1: solar radiation incidence angle
2: solar radiation refraction angle though the glass cover.
n2,1: reflective indexes, for solar radiation calculations if one of the mediums is air
then n1=equal to unity (n2/n1=1.562) (Duffie, 1991).
The glass transmissivity based on the absorption of beam solar radiation
2cos
c
a
KeEXP ……………………………………………………...………. (2.49)
Where:
Ke: cover extinction coefficient, Ke values vary from 4 m-1 for good quality glass
to 32 m-1 for poor glass (Duffie, 1991).
The reflectivity of glass cover is given by (Duffie, 1991).
2
1
2
2
1
2
1
1
1
nn
nn
………………………………………………………….………. (2.50)
The transmissivity based on reflection-refraction of beam radiation (Duffie, 1991):
1
1
1
1
r ……………………………………………………………….….….. (2.51)
Transmissivity of glass cover is given by (Duffie, 1991):
ar . ……………………………………………………………………….…. (2.52)
51
The over all optical efficiency depends upon the reflectors (the mirrors) specular
reflectivity, cover transmission and receiver tube absorption.
rPCo …………………………………………………..…..…....…. (2.53)
where
P and C the receiver tube absorbivity and mirrors specular reflectivity
It is necessary to use the incident angle modifier which represents the error in the
concentration counter due to using the sun tracking system. Each collector has
its specific incident angle modifier. In the present research it is assumed that LS3
collector is used. A description about this collector's technical parameters is
followed. The modifier for this collector is given by (Jacobson, 2006).
432
3 069092.0950559.058047.0078043.01
LSK ….…………..... (2.54)
The modified optical efficiency is given by (Jacobson, 2006).
3mod, . LSoo K ……………………………………..…………………….…. (2.55)
The end effect correction for a receiver has the same length of a reflector is given
by (Jacobson, 2006).
tan..48
1.1
12
2
f
Wf ……………………………………..…………. (2.56)
The absorbed energy by solar collector receiver is given by (Jacobson, 2006):
cos... mod,oIbSb ………………………..………….………..….….…… (2.57)
As described in figure 2.1, the proposed configuration of the ISCC includes a
separation vessel which is used to feed the solar field. The feed water
temperature for the solar field is equal to the corresponding saturated water
temperature to separator pressure. The energy supply from the solar field is fed
to the separator vessel (SV) too. As this energy increases with the solar radiation
increase, the generated stem in the SV increases. The generated steam is sent
to the low pressure turbine (LPT) to increase the electricity generation. The inlet
pressure to LPT varies with solar radiation variation. The pressure of the SV is a
function of LPT, so the feed water temperature varies as Pssv varies.
The solar field SV pressure is (Alrobaei, 2004):
P1Pssv LPTSV …..………….……………………………………….……. (2.58)
52
where sv is the pressure loss coefficient (assumed 6%).
The SV supplies water to the solar field. The water properties are evaluated as
saturated water at the SV working pressure.
The water properties at the solar field outlet:
The solar field feed pump increases the feed water pressure to outer pressure
from the solar field PoSF and the pressure losses. To simplify, the pressure loss is
assumed to be a function of the PoSF:
Po1P SFSF
PO
SF
P ……………………………………………….…………... (2.59)
where:
PPOSF: the pressure at the exit of the solar field.
PSF: The hydraulic losses coefficient for the solar field.
The heat transfer properties of water at the solar field exit are evaluated as a
function of the outlet pressure.
Tfo= f (POSF), Tfo is the temperature of water at the solar field exit.
The inlet water to the solar field properties:
The solar field feed pump increases the feed water pressure to the desired
pressure which ensures no stratification will occur in the solar field receiver tubes
(the outlet temperature is equal or less than the saturated water of the outlet
pressure). As the water is pressurised by the solar field feed pump, some heat is
gained by feed water (Alrobaei, 2004):
f.10P 2SF
POFP P
H
SF Pssv
………………………...………………………...… (2.60)
where:
FPSF: heat gain by solar field feed water
HP: hydraulic efficiency of solar field's feed pump.
The water conditions after the compression process are given as
hSF
FP hfiSF
FP ……………………………………………..………..…… (2.61)
Tfi = f (hFPSF)
53
where: hFPSF and Tfi are the enthalpy and temperature of feed water at the solar
field entrance.
Solar field thermal performance and heat losses:
To evaluate the thermal performance of the solar collectors, the thermal network
of the energy balance and mass balance is carried out. Figure 2.6 shows the
thermal network of energy conservation between solar radiation, heat absorption
by water and heat losses form solar collectors. A single cover is used to minimise
the heat losses by radiation and convection from the receiver tube.
So
lar
Ra
dia
tion
hw
hfConvection
Useful heat
Reflection of the cover
Heat emissioncover-sky
Glass cover
Heat absorption ofthe cover
Reflection of the pipe
Heat emissionpipe-cover
Heat absorptionof the pipe
mf
Ta,Tsky, Ua
Absorber tube
Figure 2.6 Thermal network for collector of solar field
Heat transfer to fluid:
The mass flow rate is assumed to be constant and it is computed based on the
nominal solar field output.
54
TR
SF
FP
SFSF hfhfomQ ..
.
. …………………………………………..…..….. (2.62)
TR
SF
FP
SFSF
hfhfo
Qm
.
Where:
QSF: nominal output for solar field
SF
m
: Total mass flow for all solar field lines
TR: energy transportation efficiency
The mass flow rate for each line is given by:
N
mm
SF
Heat transfer properties of water are evaluated at the mean temperature:
Tfm= 0.5× (Tfi + Tfo)
νf: kinematics viscosity, Prf: Prandtl number, Kf: thermal conductivity, Cpf: heat
capacity f= density
The fluid flow area inside the tube
4
. 2DciAi
……………………………………………………….……..……… (2.63)
The mean velocity of the flow inside the tube is given by:
Ai
mU
f
f
f.
4
……………………………………………………………..……... (2.64)
To evaluate whether the flow is laminar or turbulent, Reynolds number is
calculated:
f
f DtiU
.fRe …………………………………………………………..………. (2.65)
As a first guess an assumption for tube wall average temperature is made. The
purpose is to evaluate the Prandtl number for water at this temperature:
TWm= Tfi+2
Prw= f ( TWm)
55
Nusselt number for laminar flow is given by equation (2.66) and for turbulent flow
by equation (2.67) (Jacobson, 2006):
If Ref<2300 then
Nuf=3.7 ……………………………………………………………………………. (2.66)
If Ref> 104 then
25.043.08.0 PrPrPrRe021.0 wfffNuf …………………………….……. (2.67)
Convection heat transfer coefficient to fluid is evaluated by:
Dti
KNuhf
ff . ………………………………………………………….……….…. (2.68)
Then the mean temperature of the tube wall is evaluated by employing the
energy balance between heat transfer from receiver tube to water and the useful
heat gain:
Qu=.
m ×CPf× ( Tfo – Tfi ) ……………………………………………..….…..……(2.69)
Qu=×Dti×l×M×hf× (TWm-Tfm) ………………………………………..….....…… (2.70)
TWm*=Tfm+Qu/×Dti×l×M×hf
Then an iteration process is carried out to evaluate TWm
Overall loss coefficient and cover temperature:
An assumption is made for UL in order to calculate the collector effectiveness.
The collector effectiveness (ASHRAE, 2003):
ft hDti
Dto
Dti
Dto
K
Dto
ULUL
F
ln2
1.
1' ……..…………………………..…… (2.71)
The receiver tube area for each line:
Ap=×Dto×l×M ………………………………………………………..………… (2.72)
Parabola concentration ratio:
56
areareceiver
areaapertureC
..Dto
DtoWC
……………………………………………………………..……… (2.73)
Collector heat removal (ASHRAE, 2003):
Cpfm
ApULFEXP
ApUL
CpfmFR
.
.
.'.1
.
.…………………………………..………… (2.74)
Qu: useful energy gain and energy loss are given by (Jacobson, 2006):
)(..)..( TaTfi
C
ULSMlDtoWFRQu …………………………….………… (2.75)
QL=S.(W-Dto).l.M-Qu ……………………………………..….…..…..….……… (2.76)
The average temperature of the receiver tube is then estimated:
TaTApUQ PMLL .. …………………………………………………..….……. (2.77)
TaApU
QT
L
LPM
.
Convection loss to ambient:
It is necessary to evaluate the heat transfer coefficient by convection between
receiver tube and cover and between cover and ambient. So an initial guess for
cover temperatures is made:
TCM=0.5 × (TPM+Ta)
TC1=TCM+2
TC2=TCM-2
where: TC1 and TC2 are the inside cover and outside cover temperatures
respectively.
57
The space between the absorber tube and the cover is evacuated so the
convection is not considered between them.
Convection from cover to ambient:
Depending on the ambient air velocity the Reynolds number is calculated and
then the flow boundary layer evaluated. Air properties are evaluated at the
ambient air temperature Ta:
a= f (Ta)
Ka= f (Ta)
Prc= f ( TC2)
a
ca
DcoUa
.Re, …………………………………………………………...…..…… (2.78)
Then Nusselt number calculated (Alrobaei, 2004):
If 5 < Re,ca < 1000 then
25.0
38.05.0
Pr
PrPrRe,5.0,
c
cacacacaNu ………………………………..…….….… (2.79)
If 1000 < Re, ca < 200,000 then
25.0
38.06.0
Pr
PrPrRe,26.0,
c
cacacacaNu ……………………………..……..….… (2.80)
If 200,000 < Re, ca < 10,000,000 then
25.0
4.08.0
Pr
PrPrRe,23.0,
c
cacacacaNu …………………………..…...…….….… (2.81)
Dco
KcaNuhc ca
ca
,, ……………………………………………………..……….… (2.82)
where hc,ca is the convection heat transfer coefficient from the cover to ambient.
Loss by radiation:
Heat transfer coefficient by radiation between the absorber tube and its cover
Dci
Dto
Ftc
TcTpmTcTpmtchr
c
c
t
t
.
.111., 1
2
12
……………………………………….…… (2.83)
58
where:
hr,tc: radiation heat transfer coefficient between tube and cover
TC1: The cover temperature
σ: Stefan Boltzman constant 5.67×10-8
εC: cover emissivity
εt: receiver tube emissivity
FTC: view factor between tube and cover, for two long concentric cylinders, view
factor is equal to 1.
Radiation heat transfer coefficient between the cover and the ambience:
The sky temperature (T in Kelvin) (ASHRAE, 2003):
5.10552.0 TaTsky ….…………………………………….………………..……… (2.84)
w
Where:
hr,ca: radiation heat transfer coefficient from cover to ambient
Qloss* = hr,tc ××Dto×l×M ×(T PM-TC1*) ……………………………......….….… (2.86)
Qloss* = 2××l×M ×KC× (T C1*-T C2*) …………….…………………..….……… (2.87)
Qloss* = (hr,ca +hc,ca) ××Dto×l×M ×(T C2*-Ta) …………………..………….… (2.88)
By solving these equations new values are obtained for TC1*, TC2*, TPM. Then an
iteration is carried out till a desired converge is met.
Then the overall loss coefficient is calculated by equation (2.89) and compared to
the assumed value, if the difference between the assumed value and the
calculated value of UL is greater than the desired accuracy (0.001 in this
research), iteration on UL value is done (Jacobson, 2006).
1
,
1ln
2
1
.,,*
tcCcaca hrDci
Dco
KDcohrhc
DtoUL ……………………..…… (2.89)
The last iteration is for Tfo value which is calculated by:
59
Cpfm
QuTfiTfo
.*
……………………………………………………….……… (2.90)
Solar field effective area:
NMlWASF ... …………………………………………………….……………… (2.91)
The solar field capacity is the total collected heat by each row:
TRNQuQsc .. ………………………………………………………..…….……. (2.92)
where: TR is the energy transportation efficiency.
The solar field efficiency is then given by (Alrobaei, 2004):
SFAIb
Qsc
z
SF
.cos
cos.
10. 3
…………………………………………………..…….…… (2.93)
2.4.5. Integrated solar combined cycle analysis
In this section the thermal breakdown of ISCC is given. The increase in electricity
derived from employing the integrated cycle is evaluated. The available heat in
the gas turbine exhaust is converted to electricity by generating and superheating
some steam and extending this steam in a steam turbine. The steam turbine is
coupled to an electricity generator. In addition the solar field supplies some extra
energy to this thermal cycle resulting in the generation of some extra electricity.
The mathematical analysis approach which is used in this section is taken form
two resources (Alrobaei, 2004 & 1998)
Energy and mass balance for water and steam in HRSG and steam turbine unit:
Steam and water mass flows at the different locations of ISCC are referred to the
reference point. The adopted reference point is the HPT inlet.
60
o : relative steam mass flow at the turbine inlet (reference point).
evap : relative steam mass loss from the deaerator.
Loss : relative steam loss in the stem boiler.
FW : relative mass flow of feed water
o = Do / Do ………………….…………………………………….…..……….. (2.94)
Loss = DLoss / Do ……………………………………………………….….……... (2.95)
Drain =DDrain / Do ………………………………………………………...…….... (2.96)
evap = Devap / Do ………………………………………………….……..…….... (2.97)
DSB = Do + DLOSS ……………………………………………………..…….….. (2.98)
SB = o + LOSS …………………………………………………….………….. (2.99)
DFW = DSB + DDrain ……………………………………………….…………….. (2.100)
FW = SB + Drain ……………………………………………….………….….. (2.101)
The exhaust gases mass flow from the gas turbine unit and the exhaust
temperature is evaluated by gas turbine mathematical analysis. The live steam
parameters (To, Po) are proposed based on the gases’ stream temperature and
the available energy in this stream. The technical operation parameters of the
proposed design are followed.
The energy balance for the HRSG section: Figure 2.7 shows HRSG thermal
analysis.
It is assumed that the steam boiler pressure is greater than the pressure at the
HPT inlet due to pressure loss:
P1P OSB SB …………………………………..…………………………… (2.102)
where SB is the pressure loss coefficient (assumed at 6%).
61
CP
ex
4
5
6
7
8
Do (reference point)
DFW
DDS
DRK
mgas
O = 1
Loss = 0.010
Drain= 0.015
evap = 0.001
SB = 1.010
FW = 1.025
DDRAIN
DRFW
FV
DRK2
GH2
GH1
DE
GH1: Gas heater 1GH2: Gas heater 2DE : Deareator’s evaporator
DFV
Figure 2.7 HRSG thermal analysis
The steam boiler temperature is greater than the temperature at HPT inlet due to
heat loss:
5T OSB T °C
The steam superheating section is analysed by equation (2.103), where the
steam mass flow in the super heating section is obtained (see figure 2.8).
'54 SBSBSBGHgas hhDhhm ……………………………………….……… (2.103)
62
where :
mgas : GT exhaust gasses mass flow
GH: heat exchanger effectiveness
DSB: steam boiler steam mass flow
SBh : steam specific enthalpy at superheating
section exit
'SBh : water specific enthalpy of at evaporating
section inlet.
4
5
Ggas
Ggas
h4
h5
hSB
h'SB
DSB
TSB
T5
min
Figure 2.8 HRSG superheating section
h5: is evaluated at the cold end of the heat exchanger, T5 is given assuming that
it is above the water temperature inlet by the minimum allowed temperature
difference at the heat exchanger.
min5 SBTT …………………………………………………………….…..… (2.104)
The second section of HRSG analysis is the feed water heater. The water leaves
this section as a saturated steam at the steam boiler pressure. The water enters
this section after increasing its pressure by a feed water pump. The feed water
pump pressurises the feed water pressure from the deaerator pressure to steam
boiler pressure. It is assumed that the water feed pump outlet pressure is above
the steam boiler pressure by a hydraulic losses coefficient. To simplify, this
pressure losses coefficient is assumed to be a function of the outlet pressure
Po1PPO P …….…………………….…………………………....……… (2.105)
where:
PPO: pressure at the exit of feed water pump
P: hydraulic losses coefficient for HRSG
Heat gained due to water pressurising by feed water pump:
10. 2
FP P
H
DPO fPP
………………………………………...………..…… (2.106)
63
where
FP : heat gain by main feed water pump
PD: deaerator pressure
HP: the hydraulic efficiency of water feed pump.
The water conditions after the compression process are given as
hFP =FP + h'D ……………………………………………………….………… (2.107)
TFP= f (hFP)
h'D : saturated water enthalpy which corresponds to deaerator pressure
TFP : the water temperature after the feed pump
Preheating section:
The energy balance for the preheating section is employed. So the gases
temperature after the feed water heater is obtained:
FPSBFwGHgas hhDhhm '65 ……………………………………………… (2.108)
T6 = f ( h6, G ) ……………………………………………….………….……… (2.109)
Re-feed water system analysis
In order to recover some heat from the drain water a heat recovery system is
used. This heat recovery system is shown in figure 2.9. It consists of a flash
vessel (FV) and re-feed water heater (RFWH). The flash vessel generates some
steam which is sent to the deaerator. The RFWH is a heat exchanger to recover
some heat from the water drain by supplying this heat to re-feed water.
The mass and energy balance for the re-feed water flash vessel:
Drain =FV +DS ……………….………...…. (2.110)
FV: flash vessel efficiency
P FV: flash vessel pressure
P FV:= 1.06 × P D
FVSBDRAINFVDSFVFV hhh ..... '''' ……….. (2.111)
Figure 2.9 Re-feed water FV
64
RFWH solution:TRFW1=30 C °TL= 60 °CRFW =FV +Loss +evap ………………...…. (2.112)
HLFVFVRFWRFWRFW hhhh ..'. 12 …… (2.113)
Figure 2.10 RFEH analysis
The solution of the expansion process in high pressure turbine (HPT):
Depending on the initial and final steam parameters the HPT internal efficiency is
evaluated. The design parameters give the pressure at the HPT exit.
So= f (Po, To)
So is the specific entropy at the HPT inlet.
SRTh = f (PRT, SO)
SRTh specific enthalpy at the end of the isentropic expansion of the steam in the
HPT.
SRT
HPTRT hhohoh ……………………………………………………….. (2.114)
Where HPT is the HPT efficiency. It is evaluated using equations (2.13-2.16).
The deaerator energy and mass balance:
It is proposed to use an evaporator for the deaerator (DE) to recover as much
heat as possible from the gases stream.
The mass and energy balance :
D: deaerator efficiency
P DE: deaerator evaporator pressure
TRK: = TD-5 °C
'Dh = f ( P D)
''Dh = f ( P D)
DE: relative mass flow for DE.
P DE:= 1.06 × P D
Figure 2.11 Deaerator thermal analysis
65
To avoid problems related to two phase flow, it is assumed that the deaerator
evaporator DE output is 65% steam and 35% saturated water. Then the mass
balance and energy balance of the plant deaerator are given by:
SevapFWSRFWDSDRK ………………….…….... (2.115)
''''
'''
2'' ...
35.0.65.0
..
..
DDSDevapDFWD
DEDES
RFWRFWFVDS
DDRKRK
hhh
hh
hh
hh
…………….….… (2.116)
The gases temperature after the DE:
'''''76 35.065.0 DDEDDESGHgas hhhhDhhm ……………..………… (2.117)
T7= f (h7) …………………………………………………..…………………….. (2.118)
where DS is the mass flow in the DE.
For the proposed design it is assumed that the operation conditions of the
superheater, steam boiler, water's heater, DE, deaerator, ED, GH1 and HPT are
constant. The solar field operation conditions vary as solar radiation intensity is
not stable. Solar field supplies energy to the separator vessel which supplies
steam to LPT, so the affected parts of the power plant by the ISCC operation
system are the LPT condenser and GH2.
Gas heater one analysis:
'87 KRKRKGHgas hhDhhm ………………………………….…….…. (2.119)
T8= f (h8) ……………………………………………………………..……….. (2.120)
DRK: water mass flow in GH1.
Dsc and Qsc are obtained from the solar field mathematical analysis. Then the
mass and energy balance of the separator vessel is carried out.
66
Dsc
,hfo
(DR
K2-D
ss),
h' S
V
Figure 2.12 solar separator vessel thermal analysis
By employing the energy balance for SV the generated steam DSS can be
calculated.
'2
'''22 ).(.... SVSSRKSVSCSVSSSVSCRKRK hDDhDhDhfoDhD ………….….. (2.121)
2
''
'22
' ..
RKSV
SVSVRKRKSVSVSCSS
hh
hhDhhfoDD
To obtain the design value of DRK2 it is assumed that the outlet temperature is
Tex =130 °C and then the second gas heater GH2 is evaluated:
mRKRKGHex hhDhhGgas 28 …………………………………..…..…. (2.122)
mRK
GHexRK
hh
hhGgasD
82 ………………………………………….……..….. (2.123)
67
Low pressure turbine LPT evaluation:
The steam mass flow in LPT is the sum of the remaining steam after extracting
some steam to operate the plant deaerator and the generated steam from the
separator vessel.
The pressure value at the LPT is calculated by:
22
2
2. KLPTO
KO
KKLPTN PP
D
DPP …………………………………….… (2.124)
where:
PLPTO: pressure at the LPT for CC regime. It is assumed as 97% PRT.
DKO: condensed water at the CC regime
Pk: condenser pressure
The expansion process in LPT is then evaluated:
Depending on the initial parameters of steam and the final parameters of
water/steam, the internal efficiency is evaluated LPTi . The same approach in
HPT analysis is used. However, if the conditions at the end of the expansion
process are a mixture of steam water, the efficiency is corrected by the dryness
fraction Xk.
KK
KKK
hh
hhX
'''
'
………………………………………………………..…..…….. (2.125)
2
11 KLPTiLPT X
……………………………………………..…………. (2.126)
The generated electricity from HPT and LPT can be estimated.
310.KLPTSSRTRToST hhDDDohhDoNE ………………..….... (2.127)
where:
ho : steam specific enthalpy at the HPT inlet
hRT: steam specific enthalpy at the HPT exit
DRT: extracted steam to operate the plant deaerator
DSS: generated stem due to solar field contribution
hLPT: specific enthalpy of steam at the LPT inlet
68
hK: specific enthalpy of mixture at the expansion process in LPT.
The energy consumption by water feed pump:
epmp
FPFWFP
DN
.
..1000 ……………….……………………….………..………... (2.128)
Where: mp, ep are the pump’s mechanical and electric efficiency.
The ISCC efficiency:
The integrated solar combined cycle power plant efficiency is then given by:
..10..
10.3
6
SFGT
FPSTGT
ISCCAIbQcvB
NNENE
…………………………..….…………….. (2.129)
At night-time the plant is working on combined cycle regime, so the steam at the
LPT will be the remaining steam after extracting the required steam to run the
deaerator. And the electricity generating for the CC regime is given by:
310.KLPToRTRTo
ccST hhDDohhDoNE ……………………….. (2.130)
Where:
(NEST)CC: electricity generating at night-time (combined cycle regime).
hLPTO: specific enthalpy of steam at LPT inlet on CC operating system.
2.4.6. Economic and environmental analysis
Fuel saving is the selected factor to represent the economic effectiveness of the
proposed design. The evaluation of fuel saving due to employing the proposed
design is carried out based on an assumption that the extra generated electricity
by the proposed design is being generated by another combined cycle with an
efficiency of 50%, see figure 2.13. i.e. a comparison between modification of the
existing GTU or using a CC to generate this amount of electricity .
69
GTBGT
NEGT
Qex
CCBCC
NECC
QexCC
GTBGT
NEGT
ST
NEST
QexCC
text
NECC=NEST
BGT+BCC>BGTOption: 2
Option: 1
Figure 2.13 Fuel saving analysis
ECC
ECC
Qcvbe
.
3600 …………………………………………..………..………….. (2.131)
where:
beECC: is the specific fuel consumption for the equivalent combined cycle
ECC: is the equivalent combined cycle efficiency.
The fuel saving for the ISCC at time (t) then is given by equation (2.132):
ECCFP
CCSTST beNNENEDB . …………………………………………... (2.132)
Fuel saving for the CC operation:
ECCFP
CCSTCC beNNEDB . ………………………………………………... (2.133)
This fuel saving will result in carbon dioxide release avoidance which is assumed
to be 3.1 tonne of carbon dioxide per tonne of oil. The adapted carbon dioxide
emission factor is chosen from the Stockholm Institute (SEEN, 1997).
70
3. Solution procedure and results
3.1. Gas turbine unit
The flowchart of the gas turbine subprogram which is used to predict the gas
turbine unit performance is outlined in figure 3.1. Based on the mathematical
analysis of the gas turbine unit presented in Section 2.4.1, the model
evaluates the performances of GTU components. It is assumed that the GTU
is operated at an annual average temperature of air at the power plant
location of 25 °C (CSES, 2007). Alternatively, the program is capable of
reading the temperature data of ambient air all day long – if they are available.
The input data for the computer program are shown in table 3.1.
The model starts by evaluating the compression process in order to estimate
the final conditions of air at the end of the compression process. A first guess
is given for a to obtain the actual process of compression in the compressor.
Then an iteration process is carried out to calculate the actual process.
Solving the combustion chamber energy balance is the second step. Taking
into account the extraction of some air of the turbine blades cooling system,
the relative fuel mass flow and relative mass flow of cooling air are obtained
from experimental correlations. This allows us to obtain the gases state at the
turbine inlet. A guessed value for specific heat ratio G for gas is given to
solve the actual expansion process, and similarly to the compression process,
by employing an iteration process the actual condition of gases at the GTU
outlet is evaluated. By analysing the cooling air system the net output of the
turbine is evaluated. Lastly, the gas turbine efficiency, gases mass flow and
gases temperature are obtained.
The exhaust gas temperature and mass flow are then exported to the main
program to evaluate the available heat within the exhaust gas. This heat is the
main energy source for the steam turbine unit. Table 3.2 shows the results of
the gas turbine subprogram.
71
Figure 3.1 Gas turbine subprogram flowchart
72
Table 3.1 Input data to gas turbine subprogram
Ta Ambient temperature 25 °C
Pa Ambient pressure 1.013 bar
P1 Pressure loss at compressor intake 1% P1 bar
P2 Pressure loss after GTU 1% P4 bar
K Compression ratio 15.7
T3 Turbine inlet temperature 1100 °C
Tbw Average temperature of turbine blades 850 °C
R Gas constant for air 0.28669 kJ/kg.K°
NEGT Gas turbine output 51 MW
Qcv Fuel calorific value (natural gas) 44.30 MJ/kg
Qcv Fuel calorific value (Oil) 29.31 MJ/kg
co Blades cooling effectiveness 0.42 %
co Expansion coefficient of cooling air 0.35 %
ηG Electricity generator efficiency 0.98 %
ηm Mechanical efficiency 0.98 %
Table 3.2 Results of gas turbine subprogram.
We Net specific work of the GTU 247 kJ/kg
WT Total turbine work 662 kJ/kg
Wa Gas work in the turbine 622 kJ/kg
Wco Cooling air work in the turbine 40 kJ/kg
GT Gas turbine efficiency 32.2 %
BGT Gas turbine fuel consumption 19.4 tonne / h
beGT Specific fuel consumption 0.38 tonne /MWh
mC Relative mass flow of cooling air 0.105 -
mgas Exhaust gas mass flow 208 kg/s
mK Mass flow at GT compressor 210.5 kg/s
G Excess air coefficient 3.9 -
Qex Rejected heat 102.51 MW
T4 Exhaust gas temperature 492 °C
73
3.2. Gas turbine subprogram validation
To validate the obtained results from the gas turbine subprogram, the
program is run to simulate two existing gas turbine power plants which are
working to supply electricity to the Libyan grid. These are the Azzwetenah and
the Azzawayah gas turbine power plants.
The simulation results are compared to the design data of the two gas turbine
units. The design data were obtained from the local operator GECOL (Elgady,
2007). Table 3.3 shows a comparison between the model results and the
design data for the Azzwetenah and Azzawayah gas turbine units.
Three factors of the gas turbine model's results were compared to the design
data of the existing gas turbine power plants. Theses factors are; the GTU
efficiency, exhaust mass flow and exhaust temperature. The achieved
accuracies for the GTU efficiencies were 99.6% and 99.7% for GT8C and
GT13E2 respectively. The exhaust mass flow results were also accurate,
where the inaccuracies for GT8C and GT13E2 simulation were 4% and 2.3%
for and GT13E2 respectively. The obtained exhaust temperatures from the
GTU program were less than the design temperatures. The exhaust
temperature for GT8C was 99.4% accurate and for the GT13E2 the exhaust
temperature was 99% accurate. So the GTU subprogram can be used
reasonably to simulate these units and to evaluate their performance.
Table 3.3 Gas turbine subprogram validation
Plant's location Azzwetenah Azzawayah
Gas turbine model GT8C GT13E2
Manufacturer ABB ALSTOM
Efficiency [%] Program 32.2 35.6
Data 32.3 35.7
Exhaust mass flow [kg/s] Program 208 535.3
Data 200 532.0
Exhaust temperature [°C] Program 492 523.3
Data 495 525.0
74
3.3. Solution of solar field procedure
The flow chart of the solar field subprogram is shown in figure 3.2. It starts by
calculating the sunrise and sunset times to simulate the solar field operation
during this period. Solar radiation is estimated by the Hottel method at each
time step. Clear sky beam radiation is obtained based on the location
information: latitude, altitude and climate type. The solution procedure
assumes that for the solar field, the ambient temperature is uniform for the
whole solar field and it is equal to the annual average temperature at the plant
location. Then the optical efficiency of the solar collector and the absorbed
heat by the receiver tube are calculated. The water mass flow is obtained
based on the nominal solar field capacity and inlet and outlet temperatures of
water. The inlet temperature of water is specified as the saturated water of the
SV pressure. The outlet temperature initially is assumed as the saturated
water of the outlet pressure from the solar field. After solving the solar field a
new value for the water outlet temperature is obtained, if the obtained outlet
temperature is greater than the saturated water of the outlet pressure, the
solar field mass flow is increased to reduce the outlet temperature. Heat
transfer coefficients are evaluated according to initial guessed values for
cover and tube wall temperatures. After employing the energy balance
between the incoming solar radiation, heat loss by radiation and convection
and the absorbed heat by the heat transfer fluid, an iteration process is
created to evaluate these temperatures. Consequently the solar field
performance is obtained. Useful heat, heat loss and solar field efficiency are
then evaluated. Then the values of the useful heat, the outlet temperature of
water, are exported to the main program. The purpose of evaluating the solar
field is to estimate the absorbed heat which will be used in the thermal cycle
in order to increase electricity generation and improve the cycle efficiency.
75
Start
Input : Date, Ta,Dti, Dto, Dci, Dco, L,M, N,
Calculate :
Z, Ib, Id, Tsky, Sb
Let: Tfi= f (Psv)Tfo= f (Pso)UL=5 w/m2
Calculate:Tfm, m'
No
Calculate:time= sunshine
If|UL -UL|<0.001
Yes
If
|Tc*-Tc|<0.001
Let:Tc1 = Tc1*,Tc2= Tc2*
Calculate:hf inside the pipe
Calculate:F’, FR, QL, QU, Tpm
Calculate hr, hc
If|Tpm*-Tpm|
<0.001
Let :Tpm= Tpm*
No
Let: Tc1, Tc2
Calculate:Tc1*, Tc2*
Calculate :UL*
Let :UL= UL*
01
03
01
No
Yes
Yes
Calculate:Tfo, hfo
Calculate:Tfo, Qu, SF
Out put:Tfo, Qu, SF
Stop
If|Tfo -Tfo|<0.001
No
Yes
Let: Tfo=Tfo*
03
If hfo>Tfo=f(Po)
Calculate:new mass flow
03
Yes
No
Figure 3.2 Solar field flow chart
76
3.3.1. The selected solar collector
The solar collector is composed of reflectors (mirrors), supporting structure,
liner receiver and tracking system. Efforts are being made to improve the
optical efficiency of the solar collector and reduce the manufacturing cost.
Parabolic trough collector performance improved gradually, as shown in table
3.4.
The selected collector in this research is the LS-3 (LUZ system 3) which is
one of the most advanced parabolic trough collectors. This collector is the
most recent collector of the SEGS series and is used in the largest SEGS
solar power plant with a capacity of 80 MW. LS-3 is one of the three known
collector systems which have the potential to be the main three parabolic
trough suppliers (Badran and Eck, 2006). Figure 3.3 shows the LS-3
components. A glass cover is used to reduce heat loss from the receiver
tube. Its outer diameter is 0.09 mm and its thickness is 0.025 mm (Jacobson,
2006)
Table 3.5 shows the proposed solar field characteristics, where solar collector
parameters, the solar field layout and the operating conditions are explained.
The solar collector properties are obtained from the LS-3 characteristics.
Table 3.4 Solar collector's characteristics (Mills, 2004)
CollectorAcurex
3001
M.A.N
M480
Luz
LS-1
Luz
LS-3
Luz
LS-3
Luz
LS-3
Year 1981 1984 1984 1985 1988 1989
Area [m2] 34 80 128 235 545
Aperture [m] 1.8 2.4 2.5 5 5.7
Length [m] 20 38 50 48 99
Receiver diameter [m] 0.051 0.058 0.042 0.07 0.07
Concentration ratio 36:1 41:1 61:1 71:1 82:1
Optical efficiency [%] 0.77 0.77 0.734 0.737 0.764 0.8
Receiver absorptivity 0.96 0.96 0.94 0.94 0.99 0.96
Mirror reflectivity 0.93 0.93 0.94 0.94 0.94 0.94
Receiver emittance 0.27 0.17 0.3 0.24 0.19 0.19
77
Fig
ure
3.3
LS
-3colle
cto
r(P
rice,
2006)
78
Table 3.5 Solar collector and solar field operation parameters
Solar Trough Reflector Value Unit
W Collector aperture width 5.7 m
l Collector length 99 m
C Collector reflectance 0.9 -
M Number of collectors in each row 5 -
N Number of lines 13 -
Solar Trough Receiver
Dti Receiver outer diameter 0.07 m
Dto Receiver inner diameter 0.05 m
Dco Cover outer diameter 0.09 m
Dci Cover thickness 0.0025 m
K Receiver thermal conductivity (Steel) 45 W/m. °C
kc Cover thermal conductivity (Glass) 0.78 W/m. °C
P, P Receiver emittance, absorbance 0.19, 0.94 -
c Cover emittance 0.88 -
Ke Cover extinction coefficient 12.5 m-1
N2 Cover refractive index 1.526 -
HTF Heat transfer fluid water -
Location information
Latitude 32 Degree
A Altitude 100 m
- Climate type Tropical -
Solar field
QSF Nominal solar field output 25 MW
PSOSF
Design outlet pressure 65 bar
Ambient Conditions
Ta Ambient temperature 25 °C
Ua Wind Velocity 6 m/s
3.3.2. Solar field characteristics and operation conditions
The solar field capacity depends upon the CC characteristics and the
operation strategy. The constraints to increasing the solar field capacity are
the available energy in the exhaust gases after the first stage of the high
79
pressure economiser, the steam parameters at the inlet of the HPT, and the
design characteristics of the LPT, where increasing the solar field capacity
above a certain point leads to decreased heat recovery in the HRSG. For the
present design, the solar field supplies solar steam to the LPT, so the mass
flow rate of high and low pressure fossil-steam and the steam parameters at
the inlet and the outlet of LPT are important factors. Another factor to be taken
into account is to specify the solar field capacity for the proposed design is the
projected ISCC schemes (see Appendix D), where the ratio of solar capacity
to fossil fuel capacity varies from 4% to 25%. The solar capacity of the
proposed ISCC in Algeria is 35 MW where the total capacity is 140 MW and
solar to fossil fuel capacity of Yazd projected in Iran is 17:467 MW
(Greenpeace, 2003). For this particular design with no more burning of fossil
fuel, the nominal solar field capacity for the proposed design is chosen to be
25 MW (Alrobaei, 2004). The optimal solar field capacity to be used in this
configuration is not discussed in this research.
The outlet pressure from the solar field is chosen to be 65 bar. This pressure
is chosen based on experimental results by DISS where the DSG facility was
tested with three operational pressures; 30, 60, and 100 bar. Good results for
30 and 60 bar operation pressures were achieved, but using 100 bar caused
some problems, leak at flanges for instance (Zarza, 2006).
As shown in table 3.5 ambient temperature and wind velocity are chosen to be
the annual average values at the existing power plant location which is in the
north-east region of Libya (El-Osta, 2003). Although the program results were
obtained at the average values for ambient temperature and wind velocity, the
program is capable of simulating the solar field performance at a daily
variation of temperature and wind if a file of temperature and wind velocity
daily variation is provided.
3.4. Solar field performance
The subprogram of the parabolic trough solar field is operated to simulate the
solar field performance at the given parameters in table 3.5. The subprogram
80
predicts the heat output by evaluating the outlet temperature and the water
mass flow rate. As described in the operation strategy, the mass flow
remained constant unless the outlet temperature went above the saturated
water of the outlet pressure from the solar field.
Figure 3.4 shows the inlet and outlet water temperatures (Tfi and Tfo) and the
solar field heat output QSC on 17th of July. The mass flow rate remained
constant due to the operation strategy. The inlet temperature fluctuates within
a very small range as the pressure of SV varies due to the solar radiation
fluctuating. The amount of the generated steam affects the pressure at the
LPT inlet, which consequently affects the SV pressure. The outlet temperature
and the absorbed heat are increased and dropped from sunrise to sunset
according to the solar radiation fluctuation.
Figures 3.5 and 3.6 show the solar-thermal efficiency, and the useful heat
which is absorbed by the solar field for the representative days of March,
June, September and December. The thermal efficiency and the useful heat
both increase according to the increase in solar radiation from sunrise till
sunset of each day, where the operation duration varies for each day.
0
5
10
15
20
25
30
5.25 7.25 9.25 11.25 13.25 15.25 17.25
Solar Time
So
lar
Fie
ldO
utp
ut
MW
0
50
100
150
200
250
300
350
400
Te
mp
era
ture
C°
Qsc Tfo Tfi
Date: 17 July
f=32 Degree
Mass flow =2.24 kg/s each row
Figure 3.4 Parabolic trough solar field performance
81
0.50
0.52
0.54
0.56
0.58
0.60
0.62
5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20
Solar Time
Eff
icie
nc
y%
16/Mar 11/Jun 15/Sep 10/Dec
Figure 3.5 Solar field efficiency at selected dates
0
5
10
15
20
25
5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20
Solar Time
So
lar
Fie
ldO
utp
ut
MW
16/Mar 11/Jun 15/Sep 10/Dec
Figure 3.6 Solar field output at selected dates
82
3.5. Integrated Solar combined cycle solution
The ISCC program is compiled of the different subprograms, gas turbine,
solar radiation estimating and the combined cycle. The combined cycle
solution is included in the main ISCC program. Figure 3.7 shows the flowchart
of the main program of the ISCC. The solution procedure of the ISCC is
explained in the following steps:
The program starts by calculating the sunrise and sunset times to
simulate the ISCC during the daytime.
The gas turbine solver is then run to provide the exhaust temperature
and gases mass flow rate to evaluate the rejected heat from GTU.
These gases are the major energy resource for the ISCC.
Then the relative mass flow rate for water and steam in the steam
turbine, HRSG and feed water system is balanced as has been
described in equations (2.100) to (2.107).
Solving the superheating section and the evaporating section in the
HRSG: The energy balance for the superheating-evaporating section is
carried out based on the minimum allowed temperature difference on
the cold end of the heat exchanger (60 °C) (i.e the gases temperature
after the evaporating section is greater than the steam boiler
temperature by 60 °C), the gases mass flow rate and the steam boiler
temperature. The steam boiler temperature is assumed to be above the
temperature of the supplied steam to HPT (To) by 5 °C. By analysing
the superheating section in the HRSG, the generated steam mass flow
rate in the boiler's drum is obtained. as a result the feed water mass
flow rate is evaluated
The steam expansion process in the HPT is carried out to evaluate the
steam condition at the end of the expansion process. At this point some
steam is extracted to operate the deaerator and the remaining steam is
sent to the LPT. In the CC operation regime, this amount of steam is
the only expanded steam in the LPT. In the ISCC the expanded steam
in the LPT is equal to the generated steam in the boiler minus the
83
extracted steam for the deaerator plus the generated steam in the
separator vessel of the solar field.
The feed water system analysis: the feed water system solution
includes solving the re-feed water system, the deaerator's evaporator
and the deaerator. The re-feed water system includes using heat
exchanger and flash steam generator to recover some heat from the
drained water. Solving the feed water system results in obtaining the
gases' temperature after the GH1.
The proposed design is operated as a combined cycle at night-time, so
the plant characteristics for this operation system are calculated. The
expansion of steam in the LPT is carried out and consequently the
electricity being generated for the CC regime is obtained.
Solving the solar field: The first guessed value for the operating
pressure of the SV is assumed to be above the LPT inlet pressure by
6% to overcome pressure loss. So the temperature of the inlet water to
the solar field is obtained as saturated water at the SV pressure. Then
the solar field performance is evaluated to obtain the heat gain.
The SV and the second gas heater are analysed to obtain the
generated steam in the SV.
A new value of pressure at the LPT is evaluated due to the mixing
process of the solar steam from the SV and the fossil-fuel steam from
the HPT exit. The new pressure at the LPT inlet is compared to the
initial value. Then an iteration procedure is carried out to obtain the
operating pressure of the SV and LPT pressure inlet.
The expansion process at the LPT is conducted to predict the electricity
generation, fuel saving and thermal efficiency.
A new time step is chosen from sunrise till sunset to evaluate the ISCC
performance during the daylight.
The generated electricity for the CC regime is constant and the generated
electricity for the ISCC varies with the solar radiation variation. So the fuel
saving and cycle efficiency vary from sunrise to sunset.
84
Figure 3.7 ISCC flow chart
85
3.6. Integrated Solar combined cycle modeling
results
3.6.1. The operation parameters for the ISCC
The data related to solar field operation are presented in table 3.5. The
operation parameters of the HRSG and steam turbine unit are presented in
table 3.6. From a thermodynamics point of view, for the Rankin cycle, it is
desirable to increase the steam superheating as much as possible because it
improves the thermal efficiency and improves the steam quality at the steam
turbine outlet. Increasing the steam pressure improves the thermal efficiency
of the Rankin cycle. However, it decreases the dryness fraction on the last
stages of the steam turbine which causes turbine blade wear (Yunus, 1997).
Therefore the limit for increasing the steam temperature at the steam boiler is
the exhaust gases' temperature. The exhaust gas temperature is about 495
°C and to allow a minimum temperature difference at the hot end of the heat
exchanger the steam temperature is chosen to be 440 °C. The steam
temperature and pressure can be obtained also from the chart provided by the
gas turbine supplier for each unit capability for steam generation, as shown in
figure 3.8.
Table 3.6 Solar ISCC operation parameters
Symbol Description Value Unit
Po Pressure at HPT inlet 45 bar
To Temperature at HPT inlet 440 °C
PRT Pressure at HPT exit 6.0 bar
PK Condenser pressure 0.08 bar
TL Drain water temperature 60 °C
TRFW1 Re-feed water temperature 30 °C
D deaerator efficiency 0.98 %
GH HRSG efficiency 0.98 %
Loss Relative steam loss in the steam boiler. 0.010 -
Drain Drain water from the HRSG drum 0.020 -
evap Relative steam mass loss from the deaerator. 0.001 -
86
Figure 3.8 HRSG steam capability as a function of pressure and temperature for
GT8C2 (ALSTOM, 2007)
3.6.2. The simulation results for the ISCC
The results from the computer program of solving the ISCC are presented in
figures 3.9 to 3.11. Figure 3.9 shows the solar electricity production for each
gas turbine unit on representative days of four different months. These
months are chosen to illustrate the ISCC performance at different seasons of
the year. The selected days are 16th March, 11th June, 15th September and
the 10th December.
The electricity generation for the combined cycle operation is 20.53 MW for
each 51 MW gas turbine unit. So the total output is 71.53 MW per unit. This
amount of electricity is the electricity generation at night-time when the plant is
working on the CC regime. This electricity increase improves the performance
of the plant cycle and reduces the specific carbon dioxide emissions per MW
where the fuel consumption remains the same. As a result of this
improvement the specific fuel consumption for the power plant is improved
beCC = o.275 tonne /MWh, relative to the basic design beGT = 0.38 tonne
/MWh.
87
Then from sunrise to sunset the amount of electricity varies according to the
solar radiation intensity variation, as shown in figure 3.9. The amount of
generated energy during the summer is greater than for the other seasons
due to the higher solar radiation intensity and longer solar radiation duration.
Figure 3.10 shows the variation of solar steam generation and the fuel savings
according to the solar radiation variation. Figure 3.11 presents the generated
energy and the fuel saving for each month of the year for each gas turbine
unit.
At the design point (12:00 17 July) the total steam turbine output is equal to
22.96 MW which means an increase in electricity generation by about 2.43
MW for each gas turbine unit. As a result of the electricity generation increase
the specific fuel consumption drops. The specific fuel consumption at this
point is equal to 0.262 tonne /MWh and the fuel saving, because of employing
the solar field, is DBSF = 0.55 tonne /h at the design point.
The results of integrating the solar field into the thermal cycle for each gas
turbine unit are: the annual generated solar electricity is 98.55 TWh, the
annual solar fuel saving is 1.86 k tonne and the avoided carbon dioxide
emissions are 5.76 k tonne per year.
21.6
21.8
22.0
22.2
22.4
22.6
22.8
23.0
23.2
5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20
Solar Time
So
lar
Ele
ctr
icit
yM
W
16-Mar 11-Jun 15-Sep 11-Dec
Figure 3.9 Electricity generating during sunny periods at selected dates
88
0
100
200
300
400
500
600
700
800
900
5.25 7.25 9.25 11.25 13.25 15.25 17.25
Solar Time
Beam
rad
iati
on
W/m
2
0.2
0.25
0.3
0.35
0.4
0.45
0.5
0.55
0.6
Fu
el
Savin
gto
nn
e/h
so
lar
ste
am
*10
kg
/s
Ib Solar Steam Fuel saving
Date: 11 June
f=32 Degree
Figure 3.10 Fuel saving and solar steam variation at 11th June
0
100
200
300
400
500
600
700
1 2 3 4 5 6 7 8 9 10 11 12
Month
Fu
el
Sa
vin
gto
nn
e
0
2
4
6
8
10
12
Ele
ctr
icit
yG
en
era
tin
gM
Wh
Fuel Saving CO2 Saving Electricity Generating
Figure 3.11 Accumulated energy & fuel saving by solar field for each GTU
89
4. Conclusions and Recommendations for
further work
4.1. Conclusions
A mathematical code to simulate the ISCC performance has been developed,
where by the different components of the ISCC were evaluated. The gas
turbine analytic solution is capable to predict the gas turbine output with good
accuracy. The gas turbine subprogram is used to evaluate the rejected heat
from the gas turbine unit. This rejected energy can be used in CC units or
CHP applications. This program is capable to be used in further work in many
different applications. For example, it can be used in cogeneration power
plants to predict the electricity generation and the drinking water desalination
output.
The solar field computer code predicts the solar field performance. In the
present work, it is used to estimate the solar field contribution to the thermal
cycle simultaneously from sunrise to sunshine. The program results of the
outlet temperature from the solar field were good compare to the existing
trough power plants. The program achieves 350 °C as outlet temperature of
water, where this value for the existing trough power plants varies from 307
°C to 390 °C (NREL, 2007). An efficiency of 78% for the solar field is achieved
at design point, which represents good results for the solar field efficiency.
The solar field program can be used also in any further research which
includes parabolic trough solar field. For example, it can be used for solar
desalination simulation or in hybrid systems simulation. The energy from a
parabolic trough solar field can be used to supply the required energy duty for
a MSF or a MED desalination unit in order to produce drinking water.
Two advanced techniques are used in this research, the ISCC operation
system and the DSG technique. They represent the most recent advanced
techniques in solar trough applications. The ISCC simulation results show
90
that the model is capable to predict the output of the ISCC with DSG. The
system responses sensibly to the solar radiation increase, where the
electricity out put increases accordingly to the solar radiation increase.
The simulation results show that developing the existing gas turbine power
plant to an ISCC power plant will bring a package of benefits; electricity
generation increase, fuel savings and carbon dioxide release avoidance.
Developing the gas turbine into CC power plant results in electricity increase,
fuel saving and carbon dioxide emissions avoidance without burning any extra
fossil fuel. The CC regime operation provides about 40% electricity increase
causing a fuel saving of 143.83 k tonne annually and avoiding emitting 445.86
k tonne of carbon dioxide per year.
Solar energy can be converted to electricity by integrating a parabolic trough
solar field and a combined cycle power plant. This integration provides an
operation system flexibility and reliability. This flexible design leads to a
reduction in capital cost where there is no need for heat exchangers to supply
the solar heat to the Rankin cycle.
The ISCC operation increases the plant capacity to 286.12 MW at the design
point. The total fuel saving is 151.26 k tonne of oil annually. This fuel saving
avoids releasing 468.91 k tonne of carbon dioxide per year.
4.2. The ISCC result implementation
The existing gas turbine power plant consists of four units of 51 MW each and
the total capacity is 204 MW. Employing the suggested design will result in the
following benefits:
For the combined cycle operation system the electricity generation
increases by about 40% where the total capacity will be 71.53 MW for
each unit, the total capacity will be 286.12 MW.
91
The fuel saving for each unit due to this electricity increase will be
4,994 tonne /h, and the annual fuel saving will be 35.96 k tonne of oil
per unit, the total annual fuel saving is 143.83 k tonne.
The avoided carbon dioxide emissions as a result of employing the
combined cycle is 111.47 k tonne / year for each unit, the total amount
for the four units is 445.86 k tonne /year.
The results of developing the basic design are explained in table 4.1 where
the fuel consumption remains the same and the electricity generation, the fuel
saving and the carbon dioxide avoidance are increased
Table 4.1 The results of developing the gas turbine to ISCC.
Fuel consumptiontonne/h
Electricity productionMW
Fuel savingk tonne / year
CO2 savingk tonne / year
Basic design 77.6 204 - -
CC 77.6 286 143.83 445.86
ISCC 77.6 295.72* 151.26 468.91
* this amount is at the design point of the 17th
July.
92
4.3. Recommendations for further work
Securing the drinking water is an important issue for countries both
north and south of the Mediterranean, so an investigation about the
solar-cogeneration power plants in this region is recommended, where
a CHP application can be employed to generate electricity and
desalinate drinking water.
Building a database of solar radiation and ambient conditions in this
region for a period of years is recommended also to simulate the solar
thermal power plants with real data.
A study about the optimum solar field capacity share in the ISCC is
recommended.
An economic assessment for the cost of exporting the electricity to the
countries north of the Mediterranean from countries located within the
sunbelt south of the Mediterranean is also recommended.
More comprehensive research about ISCC based on real solar
radiation data and includes the transit operation and start-up and shut
down times.
93
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Appendix A Early solar thermal power plants (Trieb, 2006)
Name Location Size (MWe)Type, Heat Transfer fluid &
storage system
Start-up
dateFunding
Aurelios Adrano, Sicily 1 Tower, Water-Steam 1981 European Community
SSP/CRS Almeria, Spain 0.5 Tower, Sodium 1981 8 European countries & USA
Almeria, Spain 0.5 Trough, Oil 1981 8 European countries & USA
Sunshine Nio, Japan 1 Tower, Water-Steam 1981 Japan
Solar one California, USA 10 Tower, Water-Steam 1982 US Dept. of Energy& utilities
Themis Targasonne, France 2.5 Tower, Molten Salt 1982 France
CESA-1 Almeria, Spain 1 Tower, Water-Steam 1983 Spain
MEGS-1 Albuquerque, USA 0.75 Tower, Molten Salt 1984 US Dept. of Energy & utilities
SEGS-1 California, USA 14 Trough, Oil, Oil Storage 1984 Luz (private company)
Vanguard-1 USA 0.025 Dish 1984 Advanco Corp.
MDA USA 0.025 Dish 1984 McDonnell-Douglas
C3C-5 Crimea, Russia 5 Tower, Water-Steam 1985 Russia
99
Appendix B Cmparison between the different CSPs (Greenpeace, 2003)
Parabolic trough Central power Parabolic Dish
ApplicationsGrid-connected plants, process heat (Highest solar unit sizebuilt to date: 80 MWe)
Grid-connected plants, high temperature processheat (Highest solar unit size built to date: 10MWe)
Stand-alone applications or small off-grid power systems (Highest solar unitsize built to date: 25 kWe)
Advantages
• Commercially available – over 10 billion kWh operationalexperience; operating temperature potential up to 500°C (400°Ccommercially proven)• Commercially proven annual performance of 14% solar to netelectrical output• Commercially proven investment and operating costs
• Lowest materials demand• Best land use• Modularity
• Hybrid concept proven• Storage capability
• Good mid-term prospects for high conversionefficiencies, with solar collection; operatingtemperature potential up to1000°C (565°C proven at10MW scale)• Storage at high temperaturesHybrid operation possible
• Very high conversion efficiencies– peak solar to electricconversion of about 30%• Modularity• Hybrid operation possible• Operational experience of firstprototypes
Disadvantages
• The use of oil based heat transfer media restricts operatingtemperatures to 400°C, resulting in moderate steam qualities• Land availability, water demand
Projected annual performancevalues, investment and operating costs still needto be proved in commercial operation
• Reliability needs to be improved• Projected cost goals of massproduction still need to be achieved
100
Appendix C The commercial parabolic trough power plants SEGS 1-9 (Greenpeace, 2003)
Plant operated System output Operational Dispatch Status Status
SEGS I 1985Solar steam generation with natural gas superheating, includingthree hours of thermal storage
13.8Solar operation during sunny hours, thermal storageused to dispatch to peak period
Daily Operation without thermalstorage system (thermal storagedamaged in 1999 fire)
SEGS II 1986Solar operation during sunny hours. Natural gas backup operatedto augment solar during summer peak from 12 noon to 6:00 PMas necessary
30 Solar operation. Natural gas backup Daily Operation
SEGS III 1987Solar steam generation and solar superheating. Auxiliary naturalgas boiler to provide backup capability during low and non-solarhours
30 Solar operation. Natural gas backup Daily Operation
SEGS IV 1988Solar steam generation and solar superheating. Auxiliary naturalgas boiler to provide backup capability during low and non-solarhours
30 Solar operation. Natural gas backup Daily Operation
SEGS V 1988Solar steam generation and solar superheating. Auxiliary naturalgas boiler to provide backup capability during low and non-solarhours
30 Solar operation. Natural gas backup Daily Operation
SEGS VI 1989Solar steam generation and solar superheating. Auxiliary naturalgas boiler to provide backup capability during low and non-solarhours
30 Solar operation. Natural gas backup Daily Operation
SEGS VII 1989Solar steam generation and solar superheating. Auxiliary naturalgas boiler to provide backup capability during low and non-solarhours
30 Solar operation. Natural gas backup Daily Operation
SEGS VIII 1990Solar steam generation and solar superheating. Auxiliary naturalgas HTF heater to provide backup capability during low and non-solar hours
80 Solar operation. Natural gas backup Daily Operation
SEGS IX 1991Solar steam generation and solar superheating. Auxiliary naturalgas HTF heater to provide backup capability during low and non-solar hours
80 Solar operation. Natural gas backup Daily Operation
101
Appendix D Parabolic trough projects under development (Greenpeace, 2003)
Name/location Total capacity MWe Solar Capacity MWe Cycle
Algeria 140 35 ISCC
Kuraymat, Egypt 150 30 ISCC
Theseus-crete Greece 50 50 Steam cycle
Mathnania, Idia 140 30 ISCC
Yazd / Iran 467 17 ISCC
Israel 100 100 Steam with hybrid fossil fuel firing
Italy 40 40 Steam cycle
Baja California 291 30 ISCC
Ain Beni Mathar , Morocco 220 30 ISCC
Spain 12x50 12x50Steam with 0.5 to 12 hours storage for solar-
only operation with 12-15% hybrid firing
Nevada* 50 50 SG-1SEGS
Nevada one went on online in June 2007 (Jones, 2007b)
102
Appendix E Networks and interconnection projects until 2010 in the Mediterranean region (DLR, 2006)
Networks and interconnection projects until 2010 in the Mediterranean region
year 2020 2030 2040 2050Transfer capacityGW
2×5 8×5 14×5 20×5
Electricity TransferTWh/y
60 230 470 700
Capacity Factor 0.6 0.67 0.75 0.80Turnover Billion €/y 3.8 12.5 24 35Land Area CSPkm×km HVDC
15×153100×0.1
30×303600×0.4
40×403600×0.7
50×50
Investment CSPBillion € HVDC
425
14320
24531
35045
Electricity cost CSP€/kWh HVDC
0.0500.014
0.0450.010
0.0400.010
0.0400.010
Main indicators of the total EUMENA HVDC interconnection and CSPplants from 2020-2050 according to the TRANS-CSP scenario. In the
final stage in 2050, lines with a capacity of 5GW each will transmitabout 700TWh/y of electricity from 20 different locations in MENA to
the main centers of demand in EU
103
Appendix F ISCC program results
symbol description value Unit
T5 Gases temperature after the superheating section 308.1 °C
T6 Gases temperature after the feed water heater 251.7 °C
T7 Gases temperature after the DE 248.9 °C
T8 Gases temperature after GH1 206.1 °C
Tex Exhaust gases temperature in ISCC regime 130.0 °C
Do Steam mass flow at HPT inlet 20.5 kg/s
DFW Water mass flow at the feed water inlet 21.0 kg/s
DRT The extracted steam for deaerator operating 0.045 kg/s
DDE Water mass flow in the DE 0.45 kg/s
DSS The generated steam in SV mass flow 20.50 kg/s
DRK2 Water mass flow in GH2 28.50 kg/s
Dsc Water mass flow in the solar field 29.17. kg/s
TK The condenced water temperature 41.31 °C
TFP Water temperature after the feed pump 157.88 °C
PLPTO Pressure at LPT inlet at the CC regime 5.82 bar
PLPT Pressure at LPT inlet 5.61 bar
PSV Separator operating pressure 5.94 bar
TSV Saturated water temperature at SV pressure 158.42 °C
NFP Energy consumption by feed pump 0.2 MW
TRK2 Water temperature after GH2 194.41 °C
NECC
steam turbine output for the CC regime 20.53 MW
NEST
steam turbine output for the ISCC regime 22.96 MW
DBCC
Fuel saving for combined cycle regime 4.99 tonne /h
DBISCC
Fuel saving for ISCC 0.55 tonne /h
beCC
Specific fuel consumption for equivalent CC 0.246 tonne /MWh
beISCC
Specific fuel consumption for ISCC 0.262 tonne /MWh
Ib Beam solar radiation 812.41 w/m2
QSc The solar field heat output 25.7 MW
SF solar field efficiency (Solar to heat) 0.78 %
ISCC The ISCC efficiency 0.39 %
Tfi Water inlet temperature to the solar field 161.7 °C
Tfo Water outlet temperature from the solar field 344.5 °C
Design point: 17th July 12:00 O'clock
