xxvi cycle igcc combined cycle section: gas turbine...
TRANSCRIPT
DOCTORAL SCHOOL of ENGINEERING
SECTION OF MECHANICAL AND INDUSTRIAL ENGINEERING
XXVI CYCLE
IGCC COMBINED CYCLE SECTION:
GAS TURBINE AND STEAM CYCLE MODELS AND SIMULATORS
PhD Student
Stefano Mazzoni
Tutor Coordinator
Prof. Ing. G. Cerri Prof. Ing. E. Bemporad
Pag. 2 of 202
Index
Index ........................................................................................................................................... 2
Index of Figures ......................................................................................................................... 7
Nomenclature ........................................................................................................................... 14
Introduction .............................................................................................................................. 20
Chapter I: IGCC Power Plants ................................................................................................. 22
1.0 Introduction .................................................................................................................... 22
1.1 Introduction on IGCC Power Plants............................................................................... 22
1.1.1 Existing IGCC Plants .............................................................................................. 23
1.1.1.1 Wabash River IGCC Repowering Project ........................................................ 23
1.1.1.2 Tampa Electric Company IGCC Plant ............................................................. 24
1.1.1.3 Puertollano IGCC plant .................................................................................... 25
1.1.1.4 Buggenum IGCC Plant..................................................................................... 26
1.1.1.5 Nakoso IGCC Plant .......................................................................................... 27
1.2 H2-IGCC Power Plant .................................................................................................... 29
1.2.1 Gasification Island ................................................................................................... 31
1.2.1.1 Coal Milling and Drying .................................................................................. 32
1.2.1.2 Air Separation Unit (ASU) ............................................................................... 33
1.2.1.3 Gasifier, Sygnas Cooler and Scrubber ............................................................. 33
1.2.1.4 Water Gas-Shift ................................................................................................ 34
1.2.1.5 Acid Gas Removal Unit (AGR) ....................................................................... 35
1.2.1.5.1 H2S removal unit ....................................................................................... 36
1.2.1.5.2 CO2 removal unit and CCS ....................................................................... 36
1.2.2 Power Island ............................................................................................................ 37
1.2.2.1 Gas Turbine ...................................................................................................... 38
1.2.2.2 Steam Cycle ...................................................................................................... 38
1.3 Technical Background of H2-IGCC Power Island ........................................................ 39
1.3.1 Gas Turbine ............................................................................................................. 39
1.3.1.1 Compressor ....................................................................................................... 41
1.3.1.2 Combustion Chamber ....................................................................................... 42
Pag. 3 of 202
1.3.1.3 Expander ........................................................................................................... 43
1.3.1.4 Cooling System ................................................................................................ 44
1.3.2 Steam Cycle ............................................................................................................. 45
1.3.2.1 HRSG ............................................................................................................... 47
1.3.2.2 Steam Turbine .................................................................................................. 49
1.3.2.3 Condenser ......................................................................................................... 50
1.4 Reference ........................................................................................................................ 51
Chapter II: Modelling Approach and Solution Strategy .......................................................... 53
2.0 Introduction .................................................................................................................... 53
2.1 Thermo-mechanical Systems and modular approach ..................................................... 54
2.1.1 Modular Approach .................................................................................................. 55
2.2 Modelling Approach ...................................................................................................... 56
2.3 Methodological Approach .............................................................................................. 59
2.4 Solution Strategy ............................................................................................................ 60
2.4.1 Plant Unbalance Definition ..................................................................................... 61
2.4.2 Objective Function Definition ................................................................................. 62
2.5 Solution Methods ........................................................................................................... 63
2.5.1 Sequential ................................................................................................................ 63
2.5.2 Simultaneous ........................................................................................................... 63
2.5.3 Hybrid ...................................................................................................................... 66
2.6 Reference ........................................................................................................................ 69
Chapter III: IGCC Component Models .................................................................................... 70
3.0 Introduction .................................................................................................................... 70
3.1 Fluid Properties .............................................................................................................. 70
3.1.1 Gas Properties ......................................................................................................... 70
3.1.2 Steam properties ...................................................................................................... 72
3.1.3 Working fluid properties ......................................................................................... 73
3.2 Gas Turbine Component Models ................................................................................... 77
3.2.1 300MW F Class GT Brayton Cycle Evaluation Model .......................................... 77
3.2.1.1 Compressor Section .......................................................................................... 78
Pag. 4 of 202
3.2.1.2 Combustion Chamber Section .......................................................................... 79
3.2.1.3 Expander Section .............................................................................................. 79
3.2.1.4 Gas Turbine Equations ..................................................................................... 80
3.2.1.5 GT Global Model for the evaluation of the overall cooling mass flow ........... 82
3.2.2 Compressor .............................................................................................................. 83
3.2.3 Combustion Chamber .............................................................................................. 84
3.2.4 Expander Model ...................................................................................................... 87
3.2.5 GT Cooling Model .................................................................................................. 91
3.2.5.1 Heat transfer scheme and cooling scheme ...................................................... 91
3.2.5.1.1 Flow in the expander stages ...................................................................... 93
3.2.5.2 Blade Cooling Model ..................................................................................... 100
3.2.5.2.1 Cooling Effectiveness ............................................................................. 106
3.2.5.2.2 Effectiveness – Number of heat Transfer Unit ........................................ 107
3.3 Steam Cycle Component Models ................................................................................. 112
3.3.1 Heat Transfer Devices ........................................................................................... 113
3.3.2 Condenser .............................................................................................................. 119
3.3.3 Steam Turbine ....................................................................................................... 120
3.3.4 Deaerator ............................................................................................................... 123
3.3.5 Pump ...................................................................................................................... 124
3.3.6 Pressure Loss Devices ........................................................................................... 127
3.3.7 Junctions ................................................................................................................ 127
3.3.7.1 Water/Steam Mixer ........................................................................................ 128
3.3.7.1 Gas Mixer ....................................................................................................... 128
3.3.8 Splitter ................................................................................................................... 130
3.4 Gasification Island Simulator ....................................................................................... 131
3.4.1 Gasification Block ................................................................................................. 131
3.4.2 Water Gas Shift Block ........................................................................................... 134
3.4.3 Carbon Capture and Sequestriation Block ............................................................ 134
3.5 Reference ...................................................................................................................... 135
Chapter IV: Gas Turbine and Steam Cycle Simulators .......................................................... 137
Pag. 5 of 202
4.0 Introduction .................................................................................................................. 137
4.1 Gas Turbine Component Simulators ............................................................................ 137
4.1.1 Reference GT Brayton Cycle Evaluation and overall coolant flows .................... 137
4.1.2 Compressor ............................................................................................................ 139
4.1.3 Combustion Chamber ............................................................................................ 144
4.1.4 Expander ................................................................................................................ 146
4.2 Cooling System ............................................................................................................ 155
4.3 Gas Turbine Simulator ................................................................................................. 156
4.3.1 CH4 Gas Turbine .................................................................................................. 157
4.3.1.1 Nominal Running Point .................................................................................. 158
4.3.1.2 Part Load Analysis ......................................................................................... 159
4.3.1.3 Simulator Validation ...................................................................................... 160
4.3.2 Hydrogen Rich Syngas Gas Turbine ..................................................................... 161
4.3.2.1 33H2R Base Load Map ................................................................................. 164
4.3.3 Gas Turbine Control Rules .................................................................................... 165
4.4 Steam Cycle Component Simulator ............................................................................. 167
4.4.1 HRSG .................................................................................................................... 167
4.4.2 Steam Turbine ....................................................................................................... 171
4.4.3 Condenser .............................................................................................................. 173
4.5 Steam Cycle Simulator ................................................................................................ 174
4.6 Power Island Simulator ................................................................................................ 176
4.7 Reference ...................................................................................................................... 181
Chapter V: IGCC Plant Simulator .......................................................................................... 182
5.0 Introduction .................................................................................................................. 182
5.1 Gasification Island Simulator ....................................................................................... 183
5.2 Control policies for optimum, safe and stable operating conditions ............................ 185
5.2.2 Plant Control Philosophy ...................................................................................... 185
5.3 H2-IGCC Plant Simulator ............................................................................................ 188
5.4 H2-IGCC Plant Mapping ............................................................................................. 188
5.4.1 GT Load Changes .................................................................................................. 189
Pag. 6 of 202
5.4.2 Ambient condition changes ................................................................................... 195
5.4.3 Discussion and Concluding Remarks .................................................................... 200
5.5 Reference ...................................................................................................................... 202
Pag. 7 of 202
Index of Figures
Fig. 1.1: Block scheme of an IGCC plant without CCS (from [1]) ......................................... 22
Fig. 1.2: Block scheme of an IGCC plant with CCS (from [1]) .............................................. 22
Fig. 1.3: Schematic view of the IGCC Wabash river power plant [2] ..................................... 23
Fig. 1.4: Schematic view of the IGCC Tampa power plant [3] ............................................... 24
Fig. 1.5: Schematic view of the Puertollano Tampa power plant [4] ....................................... 25
Fig. 1.6: Schematic view of the Buggenum IGCC power plant [5] ......................................... 26
Fig. 1.6b: Schematic view of the Nakoso IGCC power plant [6] ............................................ 27
Table 1.1: Design features of coal fed IGCC power plants ..................................................... 28
Table 1.2: Performance of coal fed IGCC power plants .......................................................... 29
Fig. 1.7: H2-IGCC & CCS Reference Plant Layout [10] ......................................................... 30
Fig. 1.8: H2-IGCC plant block scheme .................................................................................... 31
Table 1.3: Mass Composition and heating values of reference IGCC Coal [10] ..................... 32
Fig. 1. 9: Sketch of H2-IGCC coal input, milling and drying system ...................................... 32
Fig. 1.10: Sketch of H2-IGCC ASU sub-system ..................................................................... 33
Fig. 1.11: Sketch of H2-IGCC Gasification, Syngas Cooling and Scrubber Sub-System ....... 34
Fig. 1.12: Sketch of H2-IGCC WGS Sub-System ................................................................... 35
Fig. 1.13: Sketch of H2-IGCC AGR Sub-System .................................................................... 36
Fig. 1.14: Sketch of the H2-IGCC Gas- Steam Combined Cycle Layout ............................... 37
Table 1.4: Generic 250-300MW Class Gas Turbines .............................................................. 40
Table 1.6: Siemens and Ansaldo GT - Characteristic Quantities ............................................. 40
Fig. 1.15: Cross Section of the SGT5 – 4000F (94.3A) ........................................................... 40
Fig. 1.16: Schematic View of the Compressor Bleed Sections (courtesy of Siemens) ............ 41
Fig. 1.17: SGT5 – 8000H – Siemens AG 2012. ....................................................................... 42
Fig. 1.18: Main Flow path in the Combustor (Ansaldo) – As Example .................................. 42
Fig. 1.19: Cross Section of the Cooling Paths (SIEMENS) ..................................................... 43
Fig. 1.20: Temperature distribution between combustor outlet and 1st Nozzle vane inlet [16] 44
Fig. 1.21: Scheme of SGT6-5000F three pressure level with drum type evaporator combined
cycle [23] .................................................................................................................................. 45
Fig. 1.22: Existing Plant Specification ..................................................................................... 46
Fig. 1.23: Specifications of under construction plant............................................................... 46
Fig. 1.24: Isometric View of 3PL-Drum Type HRSG – Horizontal and Vertical Type .......... 47
Pag. 8 of 202
Fig. 1.25: Typical 3 pressure level HRSG arrangement for combined plant (DRUM Type
EVA) ........................................................................................................................................ 47
Fig. 1.26: Scheme of Conventional Drum VS Benson Once Through Boiler [23] .................. 48
Fig. 1.27: Sketch of a finned tube bundle ................................................................................ 48
Fig. 1.28: Cross Section of SST5-3000 Steam Turbine [26] ................................................... 49
Fig. 1.29: SGT5-4000F and SST5-5000 electric generator connection [27] ........................... 49
Fig. 1.30: Water Cooling Condenser [24] ................................................................................ 50
Fig. 2.1: Sketch of IGCC plant Diagram .................................................................................. 53
Fig. 2.2: Sketch of the module input, output and attributes ..................................................... 55
Fig. 2.3: Finned Tube Heat Transfer Device - Stations and central node ................................ 56
Fig.2.4: Tube Bundle – Stations and central Node .................................................................. 57
Fig.2.5: Axial Compressor – Stations and central Node .......................................................... 57
Fig.2.6: Finite Volume Row – Stations and central Node ....................................................... 58
Fig.2.7: Condenser – Multi-zone heat transfer device ............................................................. 58
Fig. 2.8: Sketch of k-th Module ................................................................................................ 61
Fig. 2.9 : modular structure calculation method – ECRQP ...................................................... 64
Fig. 2.10: Solution Path along the Locus of P(z,r) Minima ..................................................... 65
Fig. 2.11: Hybrid methodology – Genetic Algorithm/ECRQP ................................................ 67
Fig. 2.12: complex modular structure calculation method – Hybrid Algoritm GA-ECRQP ... 68
Fig. 3.1: Block Scheme of the ENGA 5 Subroutine ................................................................ 71
Fig. 3.2: Block Scheme of the COGAS 5 Subroutine .............................................................. 71
Fig. 3.3: Block Scheme of the SYGPROP Subroutine ............................................................ 72
Table 3.0a: Wet Air – RH60% mass fraction composition ...................................................... 73
Table 3.0b: Gas Mass Fraction Composition of CH4 combustion with an 45 AFR ................ 73
Table 3.0c: ISO Air Properties ................................................................................................. 74
Table 3.0d:Gas Properties of CH4 combustion with an 45 AFR ............................................. 75
Table 3.0e: Steam Properties for different pressure ................................................................. 76
Blue – Water ; Red - Steam ...................................................................................................... 76
Fig. 3.4-a: Scheme of a Generic 300MW F Class GT ............................................................. 77
Fig.3.4-b: Scheme of a GT Brayton Cycle ............................................................................... 77
Fig. 3.5: Turbine Inlet Temperature Nomenclature ................................................................. 82
Fig. 3.6: Sketch of compressor through Flow Section ............................................................. 83
Fig. 3.7: Compressor sub-components to account the bleed extraction ................................... 84
Fig. 3.8: Sketch of combustion chamber component model .................................................... 84
Pag. 9 of 202
Fig. 3.9: Combustion Chamber Off-Design Curves ................................................................. 86
Fig. 3.10: Sketch of the Expander through Flow Section ........................................................ 87
Fig. 3.11: Schematic representation of a expander cooled row ............................................... 88
Fig. 3.12 Schematic Representation of the Mixing: ................................................................. 89
a) Momentum Conservation – b) Thermal Equilibrium ........................................................... 89
Fig. 3.13: Cooling and main flow expansion on h-s chart ....................................................... 90
Fig. 3.14: Cross Section of the Cooling Paths (SIEMENS) ..................................................... 92
Fig. 3.15: Schematic View of the main stream and coolant streams along the combustor ...... 92
and of the heat fluxes moving through the GT to the casing and to the inner components
(shaft, disk, etc.) ....................................................................................................................... 92
Fig. 3.16: Schematic view of the cooling paths along the disks – As example ....................... 94
Fig. 3.17: Example of a Generic Gas Turbine Cooling Path along Stator and Rotor Row ...... 95
Fig. 3.18: Schematic View of the Cooled components of the Stator Row – As Example ....... 95
Fig. 3.19: Typical Temperature Distribution along a 1st Stage Aeronautic Rotor Disk – As
Example .................................................................................................................................... 96
Fig. 3.20: Schematic View of a 1st Nozzle Vane Cooling Components – As Example ........... 97
Fig. 3.21: Schematic View of a 1st Rotor Blade Cooling Components – As Example ........... 97
Fig. 3.22: Comparison between cooled blade and uncooled blade coolant flow ..................... 98
Table 3.1: Fractions of the overall mass flow for each row (in percentage %) ....................... 98
Fig. 3.23: Schematic View of the cooling path ........................................................................ 99
from the compressor bleeding station to the expander row injection station ........................... 99
Fig. 3.24: Sketch of a Rotor Blade temperature distribution along the layers ....................... 100
Fig. 3.25: Simplified view of the thermal resistance for a generic blade ............................... 101
Fig 3.26: Schematic view of the enhance system of the internal heat transfer coefficient .... 102
Fig. 3.27 a-b: a) rib distribution – b) Influence of Turbulent promoter on the NU number .. 103
Fig 3.28 : Influence of jet impingement architecture on internal heat transfer coefficient .... 103
Fig 3.29: Schematic view of the depression of the external heat transfer coefficient ............ 104
owing to the film cooling ....................................................................................................... 104
Fig. 3.30: Typical heat transfer distribution among the blade row surface ............................ 105
Fig. 3.31: External heat transfer coefficient depressed by the film cooling ........................... 105
Fig. 3.32: Influence of the Thickness TBC layer on the coolant flows .................................. 106
Fig. 3.33: Temperature profile along the various blade layers ............................................... 107
Fig. 3.34: schematically main stream temperature decrease – Not to scale ........................... 109
Fig. 3.35: RO3 Cooling Design Curve – Stator Row and Rotor Row ................................... 111
Pag. 10 of 202
Fig. 3.36: Sketch of H2-IGCC Steam Cycle .......................................................................... 112
Fig. 3.37: Heat Transfer Device scheme ................................................................................ 113
Fig. 3.38: Sketch illustrating nomenclature for in-line tube arrangements [15] .................... 115
Fig. 3.39: NU vs Re max for in-line tube arrangement [15] .................................................. 116
Fig.3.40: Correction Factor to account the number of the Row [15] ..................................... 116
Fig.3.41: Heat Flux VS Temperature Difference .................................................................. 117
Table 3.1: coefficient exponents of the heat transfer coefficient calculation ......................... 119
Fig. 3.42: Multi-Zone Condenser ........................................................................................... 119
Fig. 3.43: Scheme of a generic steam expander ..................................................................... 120
Fig. 3.44: Stodola Ellipse Sketch and steam turbine body with governing valve .................. 121
Fig. 3.45: Deaerator scheme ................................................................................................... 123
Fig. 3.46: scheme of a generic pump ..................................................................................... 124
Fig. 3.47: Pumps Characteristic Non-Dimensional Curves ................................................... 126
Fig 3.48: Mixer scheme .......................................................................................................... 128
Fig 3.49: Gas Mixer scheme .................................................................................................. 129
Fig. 3.50: Splitter Scheme ...................................................................................................... 130
Fig. .3.51: Gasification Island Block Scheme ........................................................................ 131
Fig. 3.52: Gasifier Reactor Model Scheme ............................................................................ 132
Fig. 3.53: Syngas Cooler Model Scheme ............................................................................... 133
Fig. 3.54: WGS Block Scheme .............................................................................................. 134
Table 4.0a: Input Data for Cycle Calculation ........................................................................ 138
Table 4.0b: Cycle Mass Flows and Outlet Quantites ............................................................. 138
Table 4.0c: Evaluation of the overall coolant mass flow ....................................................... 139
for various coolant and blade temperature, respectively ........................................................ 139
Table 4.1: Compressor Sizing Quantities ............................................................................... 140
4.1: H2-IGCC Compressor Through Flow Shape .................................................................. 140
Fig. 4.3: Pressure ratio versus corrected mass flow curves at different ................................. 142
compressor inlet temperatures ................................................................................................ 142
Fig. 4.4: Compressor isentropic efficiency versus pressure ................................................... 142
ratio curves at different compressor inlet temperatures ......................................................... 142
Fig. 4.5: Pressure ratio versus corrected mass flow curves at different VIGV openings ....... 143
Fig. 4.6: Compressor isentropic efficiency versus pressure ................................................... 143
ratio curves at different IGV openings ................................................................................... 143
Table 4.2: Combustion Chamber Input Data ......................................................................... 144
Pag. 11 of 202
Table 4.3: Combustion Chamber output quantities ................................................................ 145
Fig. 4.7: Combustion Chamber Off-Design Curves data ....................................................... 145
Table 4.4: Blade Cooling input .............................................................................................. 146
Table 4.5: Expander Sizing Input Data .................................................................................. 147
Table 4.6: Expander output quantities .................................................................................... 147
Table 4.7: Row by Row geometric quantities ........................................................................ 147
Fig. 4.8: Expander through Flow Section including the rear frame ...................................... 148
Fig. 4.9: 1st Rotor Velocity Diagrams .................................................................................... 149
Fig. 4.10: 2nd
Rotor Velocity Diagrams ................................................................................ 150
Fig. 4.11: 3rd
Rotor Velocity Diagrams ................................................................................. 151
Fig. 4.12: 4th
Rotor Velocity Diagrams ................................................................................. 152
Fig. 4.13: Expander blade to blade overview ......................................................................... 153
Fig. 4.14: Pressure Ratio vs Corrected mass flow.................................................................. 154
for different firing temperature .............................................................................................. 154
Fig. 4.15: Total to Static Efficiency vs Pressure Ratio .......................................................... 154
for different firing temperature .............................................................................................. 154
Fig. 4.15b: Off-Design cooling effectiveness VS TCR ......................................................... 155
Fig. 4.16: ECRQP block scheme of the gas turbine matching ............................................... 156
Fig. 4.17: Sketch of the Generic 300MW F Class GT Simulator .......................................... 157
Fig. 4.18: Gas Turbine Through flow shape .......................................................................... 157
Table 4.9: RO3 Simulator - Nominal Running Point CH4 Fed ............................................. 158
Table 4.10: Results of the Lumped Model for cooling requirement - CH4 .......................... 159
Fig. 4.19: CH4 fed GT part load behaviour ........................................................................... 159
Fig. 4.20: RO3 Power Output and Efficiency at Generator Terminals .................................. 160
Fig. 4.21: Siemens SGT5-4000F Power Output and Efficiency at Generator Terminals [3] 160
Fig. 4.22: CH4 Gas Turbine Simulator Running Point for different fuel feeding ................. 161
Table 4.11: GT Simulator and Cooling System Performance Results for the various Re-
Staggering Steps ..................................................................................................................... 163
Fig. 4.23: 33H2R GT –Load and efficiency non dimensional value versus ambient
temperature ............................................................................................................................. 164
Fig. 4.24: 33H2R GT –Tex and VIGV non dimensional data versus ambient temperature .. 164
Fig. 4.: 33H2R Gas Turbine Behaviour versus Ambient Temperature .................................. 165
Fig.4.: 33H2R Gas Turbine Behaviour versus Ambient Temperature ................................... 166
Table 4.12: Gas Side quantities for HRSG calculation .......................................................... 167
Pag. 12 of 202
Fig. 4.25: Gas Steam Combined Cycle plant layout .............................................................. 168
Table 4.13: Temperature Differences of HRSG .................................................................... 169
Fig. 4.26: Gas Side and Steam Side Temperature Profile along the HRSG stations. ............ 169
Table 4.15: HRSG Sizing Results .......................................................................................... 170
Fig. 4.27: Sketch of the three turbine bodies and the HRSG interactions.............................. 171
Table 4.16: Steam Turbine Sizing Quantities ........................................................................ 172
Fig. 4.28: Steam turbine bodies (HP, IP, LP) ......................................................................... 172
Off-Design behaviour for various steam mass flowand fixed condensing pressure. ............. 172
Table 4.17: Condenser relevant sizing quantities ................................................................... 173
Fig. 4.29: Condenser Off-Design behaviour for different steam flows ................................. 173
Condensing pressure and cooling water temperature VS steam mass flow ........................... 173
Fig. 4.30: ECRQP block scheme of the steam cycle matching .............................................. 174
Table 4.18: Steam Cycle Simulator – Nominal Running Point ............................................. 175
Fig. 4.31: Schematic view of the H2-IGCC Power Island ..................................................... 176
Table 4.19: Power Island Nominal Running Point – ISO Conditions ................................... 177
Fig. 4.32: Non dimensional values of GT relevant quantities ................................................ 178
for ISO conditions and changing GT load ............................................................................ 178
Fig. 4.33: Non dimensional values of the steam side relevant quantities for various ISO
conditions loads ...................................................................................................................... 180
Fig. 5.1a: Sketch of IGCC Plant ............................................................................................. 182
Power, Heat and mass flow interactions between the various plant sections ........................ 182
Fig. 5.1b: IGCC Layout Block Scheme ................................................................................. 183
Fig. 5.2a: pressures trends and valve opening versus plant load ............................................ 187
Fig. 5.2b: Sketch of the control system of the GT fuel admission valve ............................... 187
Table 5.1: Whole System Map – Test Case ........................................................................... 188
Fig. 5.3: ISO Conditions – GT Exhaust Mass Flow and Temperature VS GT Load ............. 189
Fig. 5.4: ISO Conditions – GT Nozzle Vane and Rotor Blade life consumption rates VS GT
Load ........................................................................................................................................ 190
Fig. 5.5: ISO Conditions –Superheating temperature (HP, IP, LP) VS GT Load .................. 190
Fig. 5.6: ISO Conditions – Boiler Outlet Steam Mass Flow (HP, IP, LP) VS GT Load ....... 191
Fig. 5.7: ISO Conditions – Whole System power VS GT Load ............................................ 192
Fig. 5.8: ISO Conditions – Power Ratio VS GT Load ........................................................... 193
Fig. 5.9: ISO Conditions –33H2R Syngas and primary coal mass flow VS GT Load .......... 193
Table 5.2: Coal mass fraction composition [6] ...................................................................... 194
Pag. 13 of 202
Fig. 5.10: ISO Conditions –IGCC Power and Efficiency VS GT Load ................................. 194
Fig. 5.11: GT Exhaust Mass Flow and Temperature VS Ambient Temperature ................... 195
Fig. 5.12: ISO Conditions –Superheating temperature (HP, IP, LP) VS Ambient Temperature
................................................................................................................................................ 196
Fig. 5.13: ISO Conditions – Boiler Outlet Steam Mass Flow (HP, IP, LP) VS GT Load ..... 196
Fig. 5.14: ISO Conditions – Whole System power VS Ambient Temperature ..................... 197
Fig. 5.15: ISO Conditions – Boiler Outlet Steam Mass Flow (HP, IP, LP) VS GT Load ..... 198
Fig. 5.16: ISO Conditions –33H2R Syngas and primary coal mass flow VS GT Load ........ 198
Fig. 5.17: ISO Conditions –IGCC Power and Efficiency VS GT Load ................................. 199
Pag. 14 of 202
Nomenclature
33H2R-GT 33MJ/kg H2 Rich Fuel fed Re-Staggered Gas Turbine
af Vector of Actuality Function
AF Actuality Function
AFR Air Fuel Ratio
ANN Artificial Neural Network
ASU Air Separation Unit
b Vector of Boundary Conditions
BAT Best Available Technologies
BM Bulk Material
C Mass flow times heat capacity
CC Combustion Chamber, Combined Cycle
CCS Carbon Capture and Storage
CEM cooled expander model
CFD Computational Fluid Dynamics
CHP Combined Heat and Power
CHV Coal Heating Value
COND Condenser
cp specific constant pressure heat
cv specific constant volume heat
D Vector of Plant Model Inequalities
d Vector of Boundary Conditions, data
DB Data Base
DEGA Deaerator /Degasser
EBC Equivalent Brayton Cycle
ECLM Expander Cooling Lumped Model
ECO Economizer
ECRQP Equality Constraint Recursive Quadratic Programming
Eff Effectiveness
EVA Evaporator / Boiler
EX Extraction
f Life Consumption Rates (lcr)
F Vector of Plant Model Equations
Pag. 15 of 202
FOB Objective Function
fob Partial objective functions
FV Finite Volume
g Vector of Plant Geometric Data
GA Genetic Algorithms
GGT Generic Gas Turbine
GI Gasification Island
GT gas turbine
GTNM Gas Turbine Neural Model
GTS Gas Turbine Simulator
h enthalpy
H2R Hydrogen Rich
H2RS Hydrogen Rich Syngas
HDGT Heavy Duty Gas Turbine
HFGTS high-fidelity gas turbine simulator
HGTCR Hot Gas Thermal Capacity Rate
HP High Pressure
HRSG Heat Recovery Steam Generator
HTD Heat Transfer Device
IGCC Integrated Gasification Combined Cycle
IP Intermediate Pressure
J j-th Station
k heat ratio
k Coefficient
L Lagrangian Function, Load
LASM Lowest Allowable Stall Margin
LCS last compressor stage
LFV lumped finite volume
LHV Low Heating Value
LP Low Pressure
m mass flow
mb Bleeding Mass Flow
mc Coolant Mass Flow
Pag. 16 of 202
mc Coolant Mass Flow
mg Gas Mass Flow
n Rotational Speed
N&C New and Clean
NC Nominal Condition
NG Natural Gas
NN Neural Network
NTU Number of Transfer Units
NV Nozzle Vane
OEM Original Equipment Manufacturer
P Power, Penalty Function
p Pressure, price
PI Power Island
PR Pressure Ratio
PRE Pre – Heater (Primary Economizer)
Q heat
QTH Thermal Power
R gas constant
R Rotor Blade
rf Vector of Reality Function
RF Reality Function
RGT Reference Gas Turbine
RH Relative Humidity
RHDGT Reference Heavy Duty Gas Turbine
RMSE Root Mean Square Error
RO3 Roma Tre University
s Thickness
S surface
S Heat exchanger Surface
SC Steam Cycle
SH Super Heater
SoA State of the Art
SS Steam Section
Pag. 17 of 202
ST Steam Turbine
T temperature
Tb Blade Temperature
TBC Thermal Barrier Coating
Tc Coolant temperature
Tc Coolant Temperature
Tc Coolant Temperature
TCR, χ Thermal Capacity Ratio
Tf Firing Temperature
Tg Hot gas temperature
Tg Gas Temperature
TIT Turbine Inlet Temperature
TW Blade Temperature
U Global Heat transfer coefficient
Uc Coolant Heat Transfer Coefficient
UEBC Uncooled Equivalent Brayton Cycle
Ug Gas Heat Transfer Coefficient
UJ Heat Transfer Coefficient of j-th flow
VIGV Variable Inlet Guide Vane
W Work
WGS Water Gas Shift
x Vector of Unknown Variables
xx mass fraction composition
Y Generic Reference Variable
z Vector of unknown variables
Greek Symbol
Air fuel ratio
β Pressure ratio
∆ Unbalance
ɛ Effectiveness
λ Thermal Conductivity, Lagrangian Multpliers
μ Dynamic Viscosity
Density
Pag. 18 of 202
Vector of Degree of Freedom
ηc Cooling Effectiveness
ν Array of the Active Constraint
η Efficiency
p pressure loss
efficiency
Subscript
0 Reference Condition / Standard Condition
- negative
# number
* Reference
+ positive
+/- sub-set
1,2…. Station Order
amb ambient
b Blade
bJ j-th bleed
C Cold Stream, Compressor, Coolant
E Expander
el electric
ex Exhaust
f fuel
g Gas
GT Gas Turbine
H Hot Stream
i Inlet
is Isentropic
min minimum
N Nominal
o Outlet
p politropic
r rotor blade
Pag. 19 of 202
RJ j-th rotor
s steam / stator vane
s,i isentropic
SJ j-th stator
w water
Operators
Included into a set
Union of Set
~ Complement of sub set
Pag. 20 of 202
Introduction
Greenhouse gas carbon dioxide emitted during fossil fuel combustion leads to global climate
warming, it influences human’s life more and more. Nowadays, people call for environmental
friendly and higher efficient electric power production technologies, like integrated
gasification combined cycle (IGCC) which was developed since 1970s. Due to the rising price
of natural gas, depletion of petroleum and availability of coal, people pay more attention to
coal energy. Accordingly, Europe, USA, China focus their interest on IGCC power plants
equipped with Carbon Capture & Storage (CCS) technologies to meet the energy and
environment requirements. An IGCC power plant is a combination of a chemical plant (coal
gasification) that converts coal into a Synthetic Gaseous Fuel (SGF) and a gas-steam
combined power plant that converts the chemical energy of SGF into electricity.
Roma Tre University has been partner of the H2-IGCC Project Under the EU's 7th
Framework Programme for R&D . I’ve been involved in the H2-IGCC project as a PhD
Student in the Professor Cerri research group. Roma Tre University has been interested in two
sub-projects: Turbomachinery and System Analysis. Accordingly, I’ve been dealt with gas
turbine, steam cycle and whole system topics.
Aim of this work is the development of an IGCC Power Plant Simulator that adopts a Lumped
Performance (LP) methodology employing a Finite Volume (FV) approach based on detailed
Architecture, Geometry, Lumped Physics and Chemistry including all the empirically known
phenomena characterizing the specific behaviour of the plant components (i.e. GT, ST, Heat
transfer devices, etc.). Such a simulator has been built up taking the available technologies
and the state of the art of the existing F, G and H Class Gas Turbines and Steam Cycle
specifications of many Manufacturers into account. Features of such a simulator have been
developed as to be close to those of the existing machines of some European O&M’s.
Adoption of reality and actuality functions allows the simulator to be tailored ad hoc to the
H2-IGCC plant layout and to be a replica of the reference plant. The simulator can be seen as
a test bench of infinite sensors able to replicate and reproduce the whole system behaviour
and to forecast the power production owing to the operating conditions change (i.e. prices,
taxes, temperatures, etc.). To allow the simulator to give a real time response, neural network
modules of some plant components have been carried out.
Modular approach of elementary component models (i.e. compressor, heat transfer device,
pump, steam turbine, etc.) have been employed to perform the whole system simulator. Sizing
and off-design analyses of each modelled IGCC power plant component have been performed
Pag. 21 of 202
and H2-IGCC plant simulator has been achieved by matching the various component maps
together. Accordingly, plant part load behaviour have been investigated by means of such a
simulator tool under the adoption of proper plant control policies that takes various aspects
such as thermal and mechanical stresses as well as costs and life consumption rates of
components into consideration.
As a result of such analyses, IGCC maps have been obtained for different ambient conditions
and power demands.
Pag. 22 of 202
Chapter I
IGCC Power Plants
1.0 Introduction
Integrated Gasification Combined Cycle (IGCC) power plants are one of the most innovative
clean coal technology that puts together modern coal gasification systems, GTs and steam
cycle for electric power production. In this chapter, State of the Art (SoA) and Technical
Background (TB) of IGCC power plants and of the main components (i.e. gas turbine, heat
recovery steam generator, steam turbine, etc.) constituting such plants are reported,
respectively.
1.1 Introduction on IGCC Power Plants
IGCC plants are based on gasification that is one of the most flexible and clean process to
generate synthetic fuels from solid and liquid heavy fuels. Emissions into the environment are
lower in comparison with traditional coal plants, moreover gasification has the possibility to
capture CO2 relatively efficiently. Two alternatives are given in Fig.1.1 and Fig. 1.2 where
block schemes of an IGCC without Carbon Capture and Storage (CCS) and with CCS are
depicted.
Fig. 1.1: Block scheme of an IGCC plant without CCS (from [1])
Fig. 1.2: Block scheme of an IGCC plant with CCS (from [1])
Pag. 23 of 202
Coal is converted into a Synthetic Gaseous Fuel (SGF) by a partial combustion (oxidation)
gasification process. The raw gas contaminant substances such as sulfur and nitrogen
compounds, mercury, coal ash, particulate matter, etc. and CO2 for CCS as well as for other
uses, may be removed from the Raw Syngas (RS) by established techniques. The Clean Raw
Syngas (CRS) is a clean, transportable gaseous energy carrier. Such a CRS is used to feed the
Gas Turbine (GT) being the GT cycle the top one of the whole combined section. The Bottom
Cycle is the steam one. Heat contained in the GT exhaust stream is recovered to produce
steam in a Heat Recovery Steam Generator (HRSG). Additional steam is generated by the
gasification and purification processes. Bottoming steam turbine is fed by the above steam to
produce power.
1.1.1 Existing IGCC Plants
This section gives an outline of the existing coal based IGCC plants equipped with entrained
flow gasifiers with a brief description of their main features.
1.1.1.1 Wabash River IGCC Repowering Project
Fig. 1.3: Schematic view of the IGCC Wabash river power plant [2]
In the Wabash River power plant (Fig. 1.3) the produced syngas is fed to a GE 7FA Gas
Turbine. The gasification island is constituted by a low pressure ASU (6 bar), a slurry fed,
Pag. 24 of 202
oxygen blown two stage E-Gas gasifier, a firetube syngas cooler, and finally, by gas and
water cleaning systems. Slag is removed from the slag bath at the base of the gasifier by a
proprietary continuous letdown system.
The gas cleaning system is composed by a candle filter for hot gas filtration, a COS
hydrolysis unit, a heat fransfer device to cool the gas, an acid gas removal section based on
MDEA and a syngas saturation unit. To avoid COS catalyst degradation, a water scrubbing
unit was added downstream the candle filter to remove chlorides. The H2S loaded gas exiting
from MDEA regenerator stripper is sent to a Claus unit where elemental sulfur is produced.
The tail gas is recycled to the gasifier. The power island is equipped with a GE 7FA GT and a
HRSG that generates steam for a pre-existing 105 MW steam turbine. The NOx control is
achieved by clean syngas saturation and by injection of intermediate pressure steam into the
GT combustion chamber.
1.1.1.2 Tampa Electric Company IGCC Plant
Fig. 1.4: Schematic view of the IGCC Tampa power plant [3]
The Tampa Electric IGCC plant (Fig. 1.4) is constituted by a high pressure ASU, an oxygen
blown, down flow, single stage Texaco gasifier including heat transfer devices for syngas
cooling (a wall radiant cooler located below the gasifier, two parallel firetube convection
Pag. 25 of 202
coolers, two gas/gas heat exchangers), gas and water clean up sections. The power island is
made of a GE 7FA GT based combined cycle.
The gas clean up comprises a particulate scrubber, a raw syngas gas cooling device, a COS
hydrolysis unit to remove sulfur species (mainly H2S) and a MDEA based AGR system. The
peculiarity of Tampa IGCC in respect to other IGCC existing plant consists in the production
of sulfuric acid rather than elemental sulfur.
Nitrogen from the ASU is used for NOx formation control. In order to further reduce NOx
emissions an additional syngas saturator was included in 2002. The project demonstration
phase started in late 1996, and since then the plant has been successfully operated at design
load. Occasional part load operations have been carried out with any particular problem.
1.1.1.3 Puertollano IGCC plant
Fig. 1.5: Schematic view of the Puertollano Tampa power plant [4]
The Puertollano IGCC plant (fig. 1.5) adopts the Prenflo pressurized entrained flow, oxygen
blown gasification technology. The produced raw syngas is cleaned and supplied to a Siemens
Pag. 26 of 202
94.3 based gas-steam combined cycle. The oxygen is produced in an integrated ASU, which
also produces nitrogen for feedstock drying and transport.
The raw syngas undergoes a complete cleaning process to eliminate the pollutants. Then it is
saturated and sent to the GT combustion chamber. The GT is able of operating with both
syngas and natural gas. Steam is generated in a HRSG fed by the GT exhaust, in heat transfer
devices imbedded in the gasifier, and in heat exchangers for raw syngas cooling. The plant's
target energy efficiency is 45% in ISO conditions. The heat recovery system arrangement for
steam production is really effective. Other then power production, steam is used to
accomplish several duties concerning coal preparation, gasification, desulphurization
processes.
1.1.1.4 Buggenum IGCC Plant
Fig. 1.6: Schematic view of the Buggenum IGCC power plant [5]
The NUON (formerly Demkolec) plant at Buggenum (fig. 1.6) has been the first IGCC
European demonstration project (1994). The plant is arranged with a high pressure ASU, a
dry fed, oxygen blown Shell entrained flow gasifier, a first raw syngas cooling step to about
800°C (operated by recycling the fuel gas taken downstream the de-pulverisation section)
followed by a water tube syngas cooler for saturated steam production. Gas cleaning
apparatuses consist in a fly ash cyclone followed by a ceramic candles filter operating at
250°C, a water scrubbing unit, a COS hydrolysis unit, a Sulfinol based AGR section for H2S
Pag. 27 of 202
removal. The plant scheme is shown in Fig. 6.1. The power island is based on a Siemens
V94.2 GT combined cycle with a turbine inlet temperature of 1100°C. Both saturation and
nitrogen dilution of the syngas are adopted for NOx emission control. According to the design
features of the Siemens GT and to the desire to achieve a high plant efficiency, the full
integration between ASU and GT has been adopted.
1.1.1.5 Nakoso IGCC Plant
Fig. 1.6b: Schematic view of the Nakoso IGCC power plant [6]
The Nakoso IGCC demonstration project is owned by Japan’s Clean Coal Power R&D Co
Ltd, a consortium of Japanese power utilities and research organizations. It is based on a two-
stage, air blown MHI gasifier followed by cold syngas cleaning. The power island is arranged
with a modified M701DA GT allowing an air extraction at compressor discharge to feed the
air blown gasifier. A stand-alone ASU is included to produce nitrogen used as inert
pressurized gas to feed the coal to the gasifier. The oxygen exiting the ASU is fed to the
gasifier to enrich the gasification air [8]. The plant started demonstration operations during
Pag. 28 of 202
2007, after an extensive research and pilot testing program mainly carried out by CRIEPI
(Japan’s Central Research Institute for the Electric Power Industry). A 42.4% net efficiency
(LHV) based on Chinese bituminous coal has been achieved. The future use of US and
Indonesian sub-bituminous coals is foreseen [7].
Main features and open literature available data regarding the existing coal fed IGCC plant
are gathered in Table 1.1 and Table 1.2.
Table 1.1: Design features of coal fed IGCC power plants
Wabash
River
Tampa El.
Company Puertollano Buggenum Nakoso
Gasifier.
- Gasifier tech. GE Gas Texaco Prenflo Shell MHI
- gasifier Type
Two stage, O2,
upflow,
entrained
Single stage, O2,
downflow,
entrained
Single stage, O2,
upflow, entrained
Single stage, O2,
upflow, entrained
Two stage,
enriched air,
upflow, entrained
-feed system Slurry Slurry Dry coal Dry coal Dry coal
-recycle gas quench To second
stage no Large recycle Large recycle no
-Syngas Cooling Downflow
firetube
Downflow radiant
and convective
Two pass radiant
and convective
Downflow,
watertube
Downflow,
watertube
ASU
-pressure Low pressure High pressure High pressure High pressure
-air supply compr. Dedicated Dedicated 100% from GT 100% from GT Dedicated
- nitrogen use Vented GT NOx control GT NOx Control GT NOx Control Coal transport
Gas Clen-up
- part. removal Candle filter Water scrubbing Candle filter Candle filter Ceramic filter
- COS hydrolysis Yes Yes Yes Yes Yes
- AGR solvent MDEA MDEA MDEA Sulfinol MDEA
- sulfur recovery Claus plant Sulfuric acid Claus plant Claus plant Gypsium
- Gas saturation Yes Yes Yes Yes No
Gas Turbine
-Type GE7 FA GE7 FA Siemens 94.3 Siemens 94.2 M701DA
-Combustor Can annular Can annular Horizontal silos Vertical silos Can annular
- Firing temp. 1260 °C 1260 °C 1260°C 1100°C 1200°C
- NOx control Saturation and
steam inj.
Saturation and N2
dilution
Saturation and N2
dilution
Saturation and N2
dilution SCR
Pag. 29 of 202
Table 1.2: Performance of coal fed IGCC power plants
1.2 H2-IGCC Power Plant
In the scenario of the clean coal energy, the H2-IGCC Project under the EU’s 7th
Framework
Programme for R&D [9] is aimed to provide and demonstrate technical solution and related to
the use of state-of-the-art Gas Turbines suitable to be fed with undiluted H2-rich syngas
obtained from a pre-combustion CO2 capture process. Accordingly, a reference plant has
been studied, according with project partners [10, 11]. Layout of such a H2-IGCC power
plant is given in figure 7.1.
The plant combines a very complex fuel processing unit and a power production section based
on a gas steam combined cycle. A macro-blocks view of the plant is given in Figure 8. The
IGCC power plant with carbon sequestration and capture is a combination of a chemical plant
(the gasification island) that converts coal into a Synthetic Gaseous Fuel (SGF) and a gas-
steam combined power plant that converts the chemical energy of SGF into electricity.
The Gas turbine (GT) is a generic 300 MW one developed within the H2-IGCC project
including all the features related to the updated best available technology. Shell gasification
technology, low pressure ASU with no integration with the GT, two-stage sour gas shift and
combined H2S and CO2 removal by a Selexol process have been adopted. Combined-cycle
steam section is based on a three-pressure level HRSG cycle highly integrated with the
gasification island. The plant produces a net power o some 400MW with an efficiency of
36.2%.
In this paragraph, description of the macro islands, gasification and power island, constituting
the H2-IGCC power plant is given.
Wabash RiverTampa El.
CompanyPuertollano Buggenum Nakoso
GT - P [MW] 192 192 196 155 130
ST - P [MW] 98 125 144 128 n.a
Auxiliary P [MW] 36 66 37 31 n.a
Net P [MW] 252 250 291 252 250
LHV Net η [%] 41.2 39.8 42.4 43 42.5
HHV Net η [%] 39.7 37.5 41.7 41.4 40.5
Pag. 30 of 202
Fig. 1.7: H2-IGCC & CCS Reference Plant Layout [10]
Pag. 31 of 202
Fig. 1.8: H2-IGCC plant block scheme
1.2.1 Gasification Island
The Gasification Island (GI) is made of various macro sub-components such as Coal Milling
and Drying (CMD), Air Separation Unit (ASU), Gasifier, Syngas Cooler, Scrubber, Water
Gas Shift (WGS), Acid Gas Removal (AGR) and others. Such components are integrated by
means of mass, power, heat interactions with the Power Island (PI).
In the following paragraphs description of the H2-IGCC reference GI sections is given.
Pag. 32 of 202
1.2.1.1 Coal Milling and Drying
The coal used in the reference IGCC plant is a mixture of various trade coals available on the
world market characterized by a certain composition. Such a mixture mass composition is
shown in Table 1.3.
Table 1.3: Mass Composition and heating values of reference IGCC Coal [10]
The coal is milled and dried. The milling process leads to a fine particulate coal powder ready
for the gasification. The dried process leads to a moisture level of 2% wt. by burning
approximately 0.9% of the shifted syngas. The transport and the injection of the coal is made
by the pressurized N2 from the ASU. The amount of coal input depends on the power of GT.
The sketch of the coal input subsystem is shown in Fig. 1.9.
Fig. 1. 9: Sketch of H2-IGCC coal input, milling and drying system
Pag. 33 of 202
1.2.1.2 Air Separation Unit (ASU)
Ambient air fed into a three-stage intercooled compressor is discharged at the pressure of 5.5
bara. Then the compressed air is separated to oxygen with a purity of 95 mol% (with 2% N2,
and 3% Ar) and pure N2 by ASU. At the end, the gaseous O2 is compressed to 55 bara in a
nine stage intercooled compressor and fed to gasifier, while gaseous N2 is compressed to 80
bara in a multi-stage intercooled compressor and used for coal input system and fuel feeding
to gasifier. Some excess N2 is exhausted from ASU. The ASU subsystem with air, O2 and N2
compressors is shown in Figure 1.10.
Fig. 1.10: Sketch of H2-IGCC ASU sub-system
1.2.1.3 Gasifier, Sygnas Cooler and Scrubber
Fine powder coal is pneumatically transported from CMD system into the gasifier by means
of compressed 80 bara pure N2. Such a N2 is taken from the ASU subsystem together with
the O2. A compressed O2 stream from the ASU is fed into the gasifier to react with the coal.
Coal gasification takes place in the Shell gasifier at 45 bara and 1600°C. The single pass
gasifier converts 99.3% carbon into raw syngas that contains CO, CO2, H2, COS. The melted
ash leaves the bottom of the gasifier while the flying ash are captured by the ceramic filters.
The rest of fine particular ash stayed in the raw syngas will be got rid of by after the cleaning
processes. The tube membrane wall of the Shell gasifier receives part of the heat to generate
steam.
Pag. 34 of 202
The raw syngas at the exit of the gasifier being at 1500-1600 °C, is cooled to 900°C by adding
a stream of recycled cool syngas taken after the ash separation. The aim is to lower the raw
syngas temperature below the ash melting point. Then the raw syngas is cooled to 340°C
passing through the syngas cooler where High Pressure (HP) and Intermediate Pressure (IP)
steam is produced. The HP steam is fed into the bottom steam cycle while the IP steam is fed
into the gas shift.
The cooled raw syngas passes through the dry particulate filter where fly ash are removed and
then through the wet scrubber where the water soluble species are removed together with the
trace particulate matters such as unconverted carbon, slag and metals. Part of the raw syngas
is recirculated while the excess syngas is delivered to the WSG sub-system being the pressure
43 bara and the temperature 165 °C. The whole gasification, syngas cooling, and scrubber
subsystems are shown in Fig. 1.11.
Fig. 1.11: Sketch of H2-IGCC Gasification, Syngas Cooling and Scrubber Sub-System
1.2.1.4 Water Gas-Shift
The main species of the raw syngas after the scrubber process are: H2, COS, CO2, CO, H2O.
Sour gas (CO and COS) is harmful for the GT. So, the reference IGCC plant uses two stage
sour gas-water shift subsystem to convert CO and COS to CO2, H2, and H2S. There are two
key reactions during the sour gas-water shift, reaction (A) and reaction (B).
Pag. 35 of 202
2 2 2SteamCO H O CO H (A)
2 2 2SteamCOS H O CO H S (B)
Reaction (A) is exothermic (44 kJ/mole), thermodynamically favoured at low temperatures
where reaction rates are comparatively slow. The two stage sour gas-water shift subsystem
has high temperature stage to convert the sour gas quickly, and low temperature stage to
convert the sour gas thoroughly.
Before entering the HT (High Temperature)-SWGS, the syngas should be preheated to 250°C
by mixing with steam. The syngas temperature increases to 463°C after HT-SWGS process, it
is cooled to 250°C for LT (Low Temperature)-SWGS process. The syngas temperature
increases from 250°C to 278°C during LT-SWGS process, it has to be cooled to 25°C before
entering into the AGR subsystem. There hot syngas exiting from LT-SWGS process can be
used to preheat the raw syngas entering HT-SWGS and HP boiler feed water to cool down.
The WGS section is depicted in figure 1.12.
Fig. 1.12: Sketch of H2-IGCC WGS Sub-System
1.2.1.5 Acid Gas Removal Unit (AGR)
The reference IGCC power plant is integrated with Carbon Capture and Storage (CSS) [11].
So the AGR subsystem of the reference IGCC plant includes two stages:
Sulfide Hydrogen removal;
Carbon Dioxide removal stage (CCS).
Pag. 36 of 202
1.2.1.5.1 H2S removal unit
The syngas containing acid gas (H2S and CO2) from SWGS subsystem passes through the first
stage of the AGR subsystem to get rid of H2S. The syngas enters in the first absorption
column where the H2S is removed by a counter current flow of the Selexol solution. The H2S
gas rich solution exits the bottom of the absorber column, then is flashed and stripped off in a
regenerator. The regenerated solvent is cooled and recycled back to the top of the absorber,
while H2S is sent to a sulphur recovery unit including a Claus plant for oxidizing H2S to
elemental sulphur and a Shell Claus off gas treating (SCOT) plant for tail gas clean-up.
1.2.1.5.2 CO2 removal unit and CCS
The syngas from H2S absorber enters in the second absorber to remove CO2. The syngas
enters into the first absorption column where the H2S is removed by a counter current flow of
the Selexol solution. The CO2 rich solution exits from the bottom of the absorber column, then
it is flashed and stripped off in a regenerator. The regenerated solvent is cooled and recycled
back to the top of the absorber, while H2S is sent to a seven-stage intercooled compressor to
60 bara, liquefied and then pumped up to final pressure of 150 bara. After two-stage AGR
process, the H2 rich syngas will be ready for the Gas Turbine (GT).
The whole AGR subsystem is shown in Fig. 1.13.
Fig. 1.13: Sketch of H2-IGCC AGR Sub-System
Pag. 37 of 202
1.2.2 Power Island
Power Island is made of a Generic 300MW F Class Gas Turbine (i.e. Siemens SGT5-4000F
and Ansaldo 94.3AE) and by a 3 pressure level steam cycle (high pressure, intermediate
pressure and low pressure). In figure 1.14 a block scheme of the H2-IGCC Power Island is
given.
Fig. 1.14: Sketch of the H2-IGCC Gas- Steam Combined Cycle Layout
Pag. 38 of 202
1.2.2.1 Gas Turbine
The Generic 300MW F Class Turbine adopted to fed the bottomed HRSG has been chosen to
be similar to Siemens SGT5-4000F and Ansaldo 94.3AE GT’s. Such a GT is mainly
constituted by a compressor, made of 15 stages and equipped with Variable Inlet Guide Vane
(VIGV), by a non-sequential combustor and by a 4 stage expander. First three stages (Nozzle
Vane and Rotor Blade) are cooled. The Gas Turbine, originally fuelled by CH4 and driving a
power of some 300MW, has been re-designed to be operated with H2R Syngas. GT details
are given in the next paragraphs and chapters.
1.2.2.2 Steam Cycle
The Steam Cycle (SC) is made of a Heat Recovery Steam Generator (HRSG), a Steam
Turbine (ST) with extractions and admissions of steam, a condenser, a deaerator, and other
devices such as pumps, valves and junctions. Steam is produced in the HRSG at three fixed
pressure levels. According with the sketch of figure 14, a briefly description of the three
pressure lines is given:
High Pressure Steam Section (HPSS): in this line HP steam is produced by one
Super Heater (HP-SH), one Boiler (HP-EV) and three Economizers (HP-EC1,
HP-EC2, HP-EC3). In the evaporation section a fraction of the overall HP mass
flow is split to the gasification section and the other one is sent to the boiler tube
bundles. After the boiler, a significant mass flow is get from the Gasification
Section and mixed with the HP Steam Line. The sum of the two streams is super-
heated in the HP-SH and sent to the High Pressure Steam Turbine (HP-ST). The
steam mass flow entering the Steam Turbine is of some 144.0 kg/s at the
conditions of 140 bar and 530°C.
Intermediate Pressure Steam Section (IPSS): the IPSS is made of one Super
Heater (IP-SH), one Evaporator/Boiler (IP–EV) and one Economizer (IP – EC).
IP steam mass flow, taken from the drum, is mixed with a fraction of the HP-ST
outlet mass flow. This stream is sent to the Water Shift Gas (WSG) while the
other fraction of the HP-ST steam mass flow is re-heated in the IP-SH and sent to
the Intermediate Pressure Steam Turbine (IP-ST). The steam mass flow entering
the IP Steam Turbine is of some 100 kg/s at the conditions of 43 bar and 530°C.
Pag. 39 of 202
Low Pressure Steam Section (LPSS): the LPSS is made of one Super Heater (LP–
SH) that provides some 20kg/s at 4bar and 300°C to the main steam flux exiting
the IP – ST, one Evaporator/Boiler (LP–EV) and one Economizer (LP–EC). LP
steam mass flow is mixed together with the IP-ST mass flow and sent to the Low
Pressure Steam Turbine (LP-ST). Some Flash Steam is taken before LP-ST inlet
and addressed to the deaerator (DEGA). Flash Steam for the DEGA is taken from
the LP steam stream entering the LP – ST at 1.2bar and 240°C. Moreover, a Pre-
heater (PRE) allows to heat the feeding water mass flow from the condenser
temperature to some 15°C under the saturation temperature of the DEGA system.
H2-IGCC Steam Turbine is made of three bodies: High Pressure, Intermediate Pressure and
Low Pressure Steam Turbine (HP_ST, IP_ST, LP_ST). Steam Turbine inlet mass flow,
temperature and pressure are strictly connected with the other IGCC plant islands. Indeed the
steam turbine has many interactions with the whole plant (HRSG, WGS, etc.) that are taken
into account by considering some steam mass flows entering and exiting the boundary of the
Steam Turbine sub-system. Steam turbine and Gas Turbine are connected to the same Electric
Generator. In order to ease the HRSG integration with the rest of the plant, high and
intermediate steam production pressures are controlled by acting on the governing valves
admitting steam to high and intermediate pressure turbines. The Steam Turbine back pressure
is assumed according to that of the condenser. Anyway it should be noticed that such a
pressure is a little bit higher than the condensing one pressure because of the non-
condensable. According with the plant specification, the condenser is a surface water cooled
fed by sea.
1.3 Technical Background of H2-IGCC Power Island
Analyses of the specification concerning the H2-IGCC power island components has been
carried out. In the following paragraphs, Gas Turbine and Steam Cycle components are
described, according with the Best Available Technologies (BAT) of the State of the Art
(SoA).
1.3.1 Gas Turbine
Among the various GT’s Manufacturers (Alstom, Ansaldo, GE, Mitsubishi, Siemens, etc.) an
analysis of the 250-300MW Gas Turbines specifications has been performed. In table 1.4, GT
having output power from some 256MW to 340MW have been reported together with GT
Pag. 40 of 202
efficiencies, compressor pressure ratios, speed, number of the stages and other relevant
quantities characterizing each GT.
Table 1.4: Generic 250-300MW Class Gas Turbines
Among the various above listed GT and according with H2-IGCC power section
specifications, the Siemens SGT5-4000F and the Ansaldo AE94.3A GTs have been assumed
as reference for the development of the Generic machine that incorporates the BAT of all the
O&M’s. The Relevant quantities (mex, Tex, Power, etc.) describing gas turbines are given in
Table 5, according with [12,13]. A cross Section view of SGT5 – 4000F (94.3A) is depicted
in Fig. 1.15.
Table 1.6: Siemens and Ansaldo GT - Characteristic Quantities
Power
[MW]
Efficiency
[%]
Exhaust Temp
[°C]
Exhaust Mass
[kg/s]
Pressure Ratio
[#]
Siemens
SGT5 - 4000F 292 39.8 577 692 18.2
Ansaldo
AE 94.3A 294 39.7 580 702 18.2
Fig. 1.15: Cross Section of the SGT5 – 4000F (94.3A)
Pag. 41 of 202
H2-IGCC Generic 300MW F Class Gas Turbine simulator has been developed taking the best
available technologies of F, H and G Class GT’s into consideration. Data and information
concerning compressor, combustor, expander and cooling system performance and
arrangements have been found in various documents such as manufacturer brochures, papers
and technical report. A description of the components is given in the following paragraphs.
1.3.1.1 Compressor
Compressor looking like the Siemens and Ansaldo GT’s is an axial 15-stage high-efficiency
compressor [13] with four extractions for cooling and services purposes (i.e. sealant, piston
balance, etc.). Extractions take place at the exit of the 5th
, 9th
, 13th
and 15th
stages. First of
them, is addressed to GT services and the others to cooling purposes. Scheme of bleed
extractions is given in figure 1.16, where orange circle highlights the extractions sections.
Fig. 1.16: Schematic View of the Compressor Bleed Sections (courtesy of Siemens)
Improvements in airfoil design and in compressor off-design operating conditions (Variable
Nozzle Vane) lead to increase the compressor performance in terms of power consumption
and pressure ratio. In [13,14] many comments about the optimized flow paths and control
diffusion airfoil and about the upgrades that make better the combined plant operating
conditions are given.
Pag. 42 of 202
Moreover, figure 1.17 shows a sketch of the SGT5 -8000H in which the manufacturer
describes the peculiarities of such a machine and how the last improvements allow to better
operate the gas turbine.
Data concerning polytropic efficiency and pressure ratio are reported in two papers [17,18].
H2-IGCC compressor is characterized by a pressure ratio of 18.2 and by a polytropic
efficiency of some 93%.
Fig. 1.17: SGT5 – 8000H – Siemens AG 2012.
1.3.1.2 Combustion Chamber
Combustion chamber of such a GT is Annular combustion chamber with 24 hybrid burners
for uniform flow and temperature distribution [13]. In figure 1.18 main flow path along the
combustor is shown [12].
Fig. 1.18: Main Flow path in the Combustor (Ansaldo) – As Example
Pag. 43 of 202
1.3.1.3 Expander
H2-IGCC Gas Expander is made of four stages and a diffuser. Nozzle vanes and rotor blades
of the first three stages are cooled by means of cooling mass flows extracted from the 15th
,13th
and 9th
compressor stages. Last stage nozzle vane and rotor blade are not cooled internally, by
cooling takes places by means of the 5th
extractions bleed that re-enters the expander and
mixed together with the main flow. Scheme of expander and cooling system is depicted in
figure 1.19.
Fig. 1.19: Cross Section of the Cooling Paths (SIEMENS)
Specification concerning blade design and materials are reported in [12]. The document
states:
‘’The blades of the first and second turbine stages have to withstand thermal stresses and are therefore
fabricated from a heat-resistant alloy which is allowed to solidify as a single-crystal structure. They also have
an additional ceramic coating. They are cooled internally through a complex array of air channels and
externally by film cooling. These measures combine to ensure a long blade service life. High-efficiency vortex
and convection cooling in the blade interior with film cooling of the blade surface. Single-crystal blades made of
high-grade alloys with additional ceramic coating’’
Typical cooled polytropic efficiency value are of some 85-87% as also reported in [17,20].
Such values a pretty lower than the uncooled ones because of the mixing between cooling
flows and main flow.
Pag. 44 of 202
1.3.1.4 Cooling System
Gas Turbine cooling allows to maintain the hot components temperatures under the limit that
ensure a certain life consumption rate of the machine, as reported in [15,16].
Briefly description of cooling path along the machine is now given, taking figs. 16 and 19 into
consideration. Moving from the 1st vane of the compressor to the last rotor row of the gas
expander, the main flow path is split in various stations for various purposes, as schematically
represented in fig. 19. Some fractions of the compressor inlet mass flow are extracted at
different compressor stages and move to the expander stages mixing with the hot gas main
flow. Main flow at the compressor exit is split in various fraction. One is directed to the 1st
Nozzle Row, a second one is addressed to the 1st Rotor Row while the major of them is used
for the combustion process. All the fluxes are also adopted to cool the combustion chamber
externally and internally, respectively. Indeed, combustor is also taken in the complex cooling
path into consideration because of the high temperature of the combustion process
Accordingly, in figure 1.20 [16] is shown that the leading edge is partly cooled by purging air
which exits the gap between the combustor exit and turbine vane 1. Such a solution allows to
lower the temperature in correspondence of the stagnation point.
Fig. 1.20: Temperature distribution between combustor outlet and 1st Nozzle vane inlet [16]
In such Gas Turbine Classes, amount of cooling air in respect of the compressor inlet mass
flow is about 24-26%. Values similar to that are given in [17, 19, 20, 21].
Pag. 45 of 202
1.3.2 Steam Cycle
H2-IGCC Steam Cycle is similar to that given in figure 1.21. 300MW F Class Gas Turbine
fed a horizontal three pressure level Heat Recovery Steam Generator. Produced Steam is sent
to three turbine bodies (HP, IP, LP) and re-superheating takes place between HP steam
turbine and IP steam Turbine.
Fig. 1.21: Scheme of SGT6-5000F three pressure level with drum type evaporator combined cycle [23]
An analysis of the present of the combined cycle based on the Best Available Technology of
the F, G, H Gas Turbine has been carried out [25]. In figure 1.22 and 1.23 are reported some
specification about the steam mass flow, the steam properties (temperature, pressure), the
number of the pressure levels, the circulation system, the overall power and the installation
year of the plant. (Blue HRSG is vertical type, Red HRSG is horizontal type).
It can be notice that to improve the efficiency of the plant, according to the plant
specifications (power and heat demands), the number of pressure level is usually set to be
equal to 3: High Pressure (some 120-160bar), Intermediate Pressure (some 20-50 bar) and
Low pressure (some 3-7 bar). Steam mass flows and temperatures are different in relation
with the integration level of the plant
Pag. 46 of 202
Fig. 1.22: Existing Plant Specification
Fig. 1.23: Specifications of under construction plant
Pag. 47 of 202
1.3.2.1 HRSG
Typical layout Heat Recovery Steam Generator adopted in the combined sections of IGCC
power plants are schematically given in figure 1.24. Horizontal and Vertical HRSG are
depicted.
Fig. 1.24: Isometric View of 3PL-Drum Type HRSG – Horizontal and Vertical Type
H2-IGCC heat recovery steam generator is a horizontal one equipped with drum type
evaporator and with finned tube banks. Non-condensable are extracted by the adoption of a
Tray-type deaerator. In figure 1.25, typical horizontal three pressure level HRSG is given.
Fig. 1.25: Typical 3 pressure level HRSG arrangement for combined plant (DRUM Type EVA)
Pag. 48 of 202
Comparison between Benson and Drum evaporator has been carried out and according with
the plant specifications the adoption of conventional drum type boiler has been selected. In
figure 1.26, the two options are given [23].
Fig. 1.26: Scheme of Conventional Drum VS Benson Once Through Boiler [23]
Super-heater, economizer and boiler tube bundles are finned tube type. Such a solution is
typically adopted in these kinds of power plants. Adoption of finned tube banks leads to
increase the external heat transfer coefficients and to improve the heat transfer phenomena.
Typical HRSG finned tube banks layout is given in figure 1.27.
Fig. 1.27: Sketch of a finned tube bundle
Pag. 49 of 202
1.3.2.2 Steam Turbine
Steam Turbine adopted in such IGCC power plants is similar to the SST5-5000 and to the
SST5-3000. Three turbine bodies with a Re-heating between HP and IP ST are employed in
the plant to increase the power driven by the steam turbine. A cross section of the SST-3000
is given in figure 1.28.
Fig. 1.28: Cross Section of SST5-3000 Steam Turbine [26]
According with [27], adoption of SST-5000 and SST-3000 allows to connect the steam
turbine and gas turbine to the same electric generator. Such a solution is schematically plot in
figure 29.
Fig. 1.29: SGT5-4000F and SST5-5000 electric generator connection [27]
Pag. 50 of 202
1.3.2.3 Condenser
‘Cold cooling water results in a low condenser back pressure. Therefore a steam turbine with
a huge exhaust area is needed. For these locations, Siemens can provide a single shaft RPP
with a double flow LP steam turbine.’[27].
According with the plant specification and with the H2-IGCC project context the condenser is
a surface cooling water system fed by sea water. Condensing pressure has been assumed,
according also with some manufacturer declaration at 2.5mbar. In figure 1.30 a schematic
view of the condenser is given.
Fig. 1.30: Water Cooling Condenser [24]
Pag. 51 of 202
1.4 Reference
[1] - Xu Zhaofeng, Jens Hetland, Hanne M. Kvamsdal, Li Zheng, Liu Lianbo, “Economic
evaluation of an IGCC cogeneration power plant with CCS for application in China”, Energy
Procedia 4 (2011) 1933–1940
[2] - “Wabash River Coal Gasification Repowering Project: A DOE Assessment”, Report N.
DOE/NETL-2002/1164, 2002.
[3] - “Tampa Electric Integrated Gasification Combined-Cycle Project. A DOE Assessment”,
Report N. DOE/NETL – 2004/1207, 2004.
[4] - “IGCC Puertollano. A clean coal gasification power plant”, published by ELCOGAS
[5] - NETL Gasifipedia - Gasification in Detail, available on 28th
of July, 2011, at:
http://www.netl.doe.gov/technologies/coalpower/gasification/gasifipedia/6-apps/6-2-6-4_nuon.html
[6] - Energy for sustainable future, on 28th
of July, 2011, at:
http://energy-21.blogspot.com/2010/11/nakoso-igcc-plant.html
[7] - Ishibashi, Y., Shinada, O., “First year operation results of CCP’s Nakoso 250 MW air-
blown IGCC demonstration plant”, Gasification Technologies Conference, Washington DC,
USA, 2008.
[8] - Higman, C., van der Burgt, M.,,”Gasification”, Gulf Professional Publishing, Elsevier,
2nd
Edition, 2008.
[9] - http://www.h2-igcc.eu/default.aspx
[10] - Nikolett Sipöcz, Mohammad Mansouri, Peter Breuhaus & Mohsen Assadi, “Plant
specification and detailed thermodynamic performance analysis of selected IGCC cycle”, H2-
IGCC Report, October 2010
[11] - Department of Mech. & Structural Eng. & Material Science, University of Stavanger ,
“IGCC State of the art report, a part of EU-FP7 Low Emission Gas Turbine Technology for
Hydrogen-rich Syngas ”, H2-IGCC Report, April 2010
[12] - Ansaldo Energia Brochure AE94.3A GAS TURBINE; Genoa, Italy; May 2012.
[13] - SIEMENS AG, Siemens Gas Turbine SGT5-4000F. Answers for energy, 2008.
[14] - SIEMENS AG, Compressor Mass Flow Increase Upgrade for SGT5 – 4000F Gas
Turbines, 2008.
[15] - SIEMENS AG, Siemens Gas Turbine SGT6-5000F, Answer for Energy, 2008
[16] - SIEMENS AG, Latest performance upgrade of the Siemens gas turbine SGT5 – 4000F,
Answer for energy, 2008
Pag. 52 of 202
[17] - Jonsson M., Bolland O., Bucker D., Rost M. (Siemens), 2005, ‘Gas Turbine Cooling
Model for Evaluation of Novel Cycles’. Proceedings of ECOS 2005, Trondheim, Norway,
June 20-22, 2005.
[18] - Giuffrida A., Romano M. C., Lozza G. G., 2010, ‘Thermodynamic assessment of IGCC
power plants with hot fuel gas desulfurization’. Elsevier, Applied Energy 87 (2010), ppg.
3374 – 3383.
[19] - Kim Y.S., Lee J. L., Kim T.S, Sohn J.L., Joo Y. J., 2010: ‘Performance analysis of a
syngas-fed gas turbine considering the operating limitations of its components’, Elsevier,
Applied Energy 87 (2010), ppg. 1602-1611.
[20] - Final Report of the RTO Applied Vehicle Technology, 2007: ‘Performance Prediction
and Simulation of Gas Turbine Engine Operation for Aircraft, Marine, Vehicular, and Power
Generation’
[21] - Ashok Rao., 2010, ‘1.3.2 Advanced Bryton Cycles’
[22] – Walter H., Hofmann R., 2010: ‘How can the heat transfer correlations for finned-tubes
influence the numerical simulation of the dynamic behavior of a heat recovery steam
generator?’, Accepted Manuscript, Applied Thermal Engineering.
[23] - SIEMENS AG, Siemens Gas Turbine SGT6-5000F, Application Overview, 2008
[24] – Noordermeer J., Gryphon International Engineering Service Inc.
[25] – CMI Energy, Horizontal &Vertical HRSGs Reference List, Cockerill Maintenance &
Ingénierie.
[26] – Siemesn AG 2010: ‘Siemens Steam Turbine SST-3000 Series for combined cycle
application’.
[27] – Emberg H., Alf M., SCC5-4000F Single Shaft (SST5-5000): ‘A single shaft concept
for cold cooling water conditions’.
[28] - Xu Zhaofeng, Jens Hetland, Hanne M. Kvamsdal, Li Zheng, Liu Lianbo, “Economic
evaluation of an IGCC cogeneration power plant with CCS for application in China”, Energy
Procedia 4 (2011) 1933–1940.
Pag. 53 of 202
Chapter II
Modelling Approach and Solution Strategy
2.0 Introduction
In order to evaluate the IGCC plant operating maps and to establish appropriate control
policies, a steady state plant simulator has been set up. The H2-IGCC plant has been
developed taking two macro island into consideration: Gasification Island (GI) and Power
Island (PI).
The Gasification Island (GI) is made of many components such as the Gasifier, the Syngas
Cooler and so on. GI simulator has been developed by the assumption of component models
based on empirical correlations taken from the State of the Art (SoA), connecting the inputs to
the outputs. Connections between the Power Island (PI) and the Gasification Island have been
established taking the above empirical correlations into consideration. Chemical reactions
have been considered at equilibrium.
On the other hand, looking at the PI, detailed models have been adopted in order to described
Gas Turbine (GT) and Steam Cycle (SC) macro components. Using such a modelling
approach a simulator has been established. The simulator is a detailed replica of the various
machines and equipment’s and it has been adopted to map the plant performance, evidencing
dangerous behaviour (i.e. GT over-pressures, over-temperatures, shaft over-load, etc.) under
various operating conditions and loading.
In the following paragraphs, description of the modelling approach and the solution
techniques is given.
Fig. 2.1: Sketch of IGCC plant Diagram
Pag. 54 of 202
2.1 Thermo-mechanical Systems and modular approach
An IGCC power plant is a thermo-mechanical system made of many components. Each of
them is devoted to one process transfer of heat, work combustion and so on. Such a power
plant can be sketched as the figure 2.1.
Blocks 1,2,…,N represent components or group of components that describe the real plant
layout. In such a figure connections between components, input and output streams are
schematically depicted ( U
InE being the vector of the useful inlet quantity fluxes, U
OE being the
vector of fluxes of useful quantities, while R
OE being the vector of fluxes of rejected
quantities).
Behaviour of a generic plant can be described by an equation set:
( ) 0F z (2.1)
and by an inequalities set:
( ) 0D z (2.2)
z being the overall variable set:
z y b u g (2.3)
y being the variable set:
y x (2.4)
made by independent variables (DOFs) and by unknown variables x.
b being the boundary conditions set (ambient, etc.)
g being architecture and geometric data.
u being the status of the system set made of rf and af
f fu r a (2.5)
Moreover, , , , , , ,n d q n s b g F D ξ x u b g
In general equations F are highly non linear and express conservation of mass, momentum,
energy and entropy1, and other phenomena such as work and heat transfer, combustion,
pressure loss, etc. F includes also fluid properties, auxiliary equations, machine and
equipment specifications. Equations can also be expressed in terms of graphs or tables. D
1 The conservation of entropy for a steady state open thermodynamic system bounded by a fixed border states:
the entropy of the system does not change along the time, thus the entropy convected into the system by the
entering flows, plus the entropy increase due to external heat fluxes, plus the entropy produced by the internal
irreversibility’s is equal to the entropy extracted by the exiting flows. For a non steady state thermodynamic
process the conservation law leads to a time differential equation that takes the rate of entropy accumulation
inside the system equal to the above entropy fluxes (inlet flows, outlet flows, heat fluxes and internal entropy
production).
Pag. 55 of 202
represents a set of physical, thermal, chemical and geometrical conditions, as well as other
constrains which determine the domain where the problem (2.1) solution exists.
The values associate to the vector components usually is establish K according to suitable
criteria one of which can be the search of an appropriate objective function optimum value. x
is the solution of (2.1) and (2.2). Of course quantities can be exchanged between and x
rf and af and are the vectors of realty functions and actuality ones respectively. Reality
functions rfs are introduced to accommodate the model to reproduce the existing component
behaviour in a reference situation (New & Clean). Since during operations the component
features behaviour change continuously due to various phenomena leading to performance
modification, the model of each component has to be tailored to the new situation. Therefore
the models of the major components include suitable actuality functions afs that can represent
the actual status of the component. af accounts for the deviation of the actual component
performance from a condition assumed as the reference.
2.1.1 Modular Approach
The plant simulator has been developed taking a modular description at level of components.
A library of component models suitable to arrange the various sections of the H2-IGCC
power plant has been implemented. Each module represents a plant component. Module
structure may be defined by one or more subroutines. Each subroutine contains the model of
the corresponding elementary unit (i.e. a compressor row, heat exchanger elementary section
and so on). Complex components are built up linking the subroutines as macro-components.
Modules take various aspects such as emissions, costs, and others into account. Schematic
representation of a generic module is given in figure 2.2.
Fig. 2.2: Sketch of the module input, output and attributes
Pag. 56 of 202
2.2 Modelling Approach
The modelling approach is based on a Finite Volume (FV) discretization of plant components.
Each FV is defined by boundary surfaces J and J+1. The approach allows the introduction
into the component model of information concerning spatial and (in case) time distribution of
relevant quantities. Preliminary detailed 2D, 3D or CFD calculations can be performed by
using suitable codes. Results constitute a Data Base (DB) used to lump on the boundary J and
J+1 of each FV the distributions of quantities of interest (i.e. temperatures, pressures, wok,
losses, etc.) by means of an averaging procedure on surface and time.
The approach is addressed to model any kind of machines and apparatuses made of
elementary components such as: compressor rows, expander rows, combustion chambers,
heat exchangers, pumps, etc.
Fig. 2.3: Finned Tube Heat Transfer Device - Stations and central node
A lumping procedure is adopted also for the quantities involved in performance calculation.
The lumped features are then reduced to the FV central node JN. As an example, a HRSG tube
bundle can be sub-divided into FVs, each of them comprising a tube row, according to Fig.
2.3. J and J+1 represent the FV boundary surfaces and JN the central node. Spatial distribution
of temperature, pressure, velocity, etc. resulting from detailed calculation are averaged on the
boundary surfaces. Heat transferred from a fluid to the other (the performance) is related to
the lumped flow features and to the geometric features of the row by adapting classical heat
Pag. 57 of 202
transfer model. The connection between row features and heat transfer model is established
according to the amount of data available by detailed simulations.
Similar approaches are adopted for other machines and apparatuses such as shell & tube heat
and axial compressor. Accordingly, such systems can be modelled according to the FV
elementary device given in figure 2.4 and 2.5.
Fig.2.4: Tube Bundle – Stations and central Node
Fig.2.5: Axial Compressor – Stations and central Node
Pag. 58 of 202
The compressor is divided in row section each representing a stator or rotor cascade. Each
row is included in finite volumes FV’s delimited by a boundary, as figure 2.6 shows. The
inlet station and the central node are described by the same number J
Fig.2.6: Finite Volume Row – Stations and central Node
Moreover, adoption of this modelling approach allows to model a condenser by means of
multi-zone heat transfer device modules, each of them being characterized by FV’s. In figure
2.7, a multi-zone condenser is schematically represented.
Fig.2.7: Condenser – Multi-zone heat transfer device
Pag. 59 of 202
2.3 Methodological Approach
Equations and inequalities describing machines and plant behaviour are addressed to solve
different kind of problems and lead to obtain the whole plant simulator by following various
steps:
Cycle Calculation: this procedure is related to preliminary cycle calculation whet the
cycle potentials are going to be investigated with only few constraints concerning
thermodynamic quantities. Data are usually related to the state of the art machinery
and equipment’s (i.e efficiency, heat transfer effectiveness and so on). If related to
such above quantities cost specifications are available an optimization procedure can
take place. Thermodynamic optimisation is always possible. Indeed overall plant
efficiency and specific of work or a combination of these quantities may be chosen as
objective function. Results of this calculation are thermodynamic quantities at some
plant stations, mass flows, value of powers crossing component boundaries and
overall performances.
Sizing: this phase is preliminary to the next component off design component and
plant part load analyses. It consists in the calculation of size of machines and
equipment’s and alternative global parameters to describe off-design behaviour of
components. Input data are from the previous cycle calculations or may come from
data base DB related to the commercially available machines and equipment’s whose
design features are close to that of required cycle calculation. In this phase
specifications concerning costs of machines and equipments are used for optimized
design. Results of this inverse calculation phase may be devoted to equipment and
machine preliminary designs, but at present the are mainly addressed to the next plant
off-design investigation.
Off-Design Analysis: this direct phase investigation requires the knowledge of
geometric data, architecture and some global parameters related to the plant
components. Maps of the machine and equipment are obtained and how they match in
the plant is studied. Changes in the independent quantities (DOF’s) may be
investigated according to control policies the related rules may be implemented as
specifications. In this casa the component state quantities (u) may be used to optimize
Pag. 60 of 202
operations according with load requirements (electric and thermal power) which are
implemented as time dependent constraints.
Matching: this step consist in assembly together the component maps to perform the
whole plant simulator. Two sub-steps have to be described. First of them consists in
the sizing of the matched component connections (i.e. establish the equivalent opening
of the cooling ducts in a gas turbine cooling system). The second one, once the
connection have been set, consists in the off-design plant simulator development. By
this second step is possible to investigate whole plant part load operating conditions.
Calculations at the various steps have been performed by means of the various optimization
techniques based on the quasi Newton algorithms and on Genetic Algorithms (GA). A
comprehensive and detailed description of the solution strategy and techniques is given in the
following paragraphs.
2.4 Solution Strategy
With reference to section 2.1, the formulation of the overall set of equations F and
inequalities D includes the sub-set of equations φ and of inequalities δ related to each module.
1 2 ... kF (2.6)
1 2 ... kD (2.7)
k being the k-th module.
Scheme of a generic module with input, attributes and outputs is given in figure 2.8.
Each module takes some input quantities and gives some outcomes. Referring to figure 2.8, a
sub-set zk of the whole plant variables is given as input to the module. Module outputs are the
sub-set of equalities, inequalities, unbalance and partial objective function. Such outcomes
contribute to define the whole simulator unbalance and objective function that have to be
minimized to achieve the best solution. In the following paragraphs definition of unbalance
and of objective function is given.
Pag. 61 of 202
Fig. 2.8: Sketch of k-th Module
2.4.1 Plant Unbalance Definition
Once the parameters u and the DOFs y have been given, the search of the unknown z may be
performed by minimization of the plant unbalance function (2.x)
( ) ( ) ( )Tz F z F z (2.8)
When the solution of F is achieved, the unbalance is zero.
( ) 0z (2.9)
The necessary condition ∆(z) minimum is achieved, the following n equations have to be
satisfied.
1
0 [1, ]N
j
j
j i
ff i n
z
(2.10)
Of course this occurs when:
*( ) 0 [1, ]jf z j n (2.11)
In this case, the Hessian matrix of ∆(z) is definite not negative for *z z
Pag. 62 of 202
*( )0 [1, ]
j
i
f zi n
z
(2.12)
In this case z* is a stationary point for function fj(z), therefore, it may not be the searched
solution point.
The above suggests the idea that stating the following minimization problem
Search * *:min ( ) ( , , ) 0; ( , , ) 0;z z F u z y D u z y u u y y (2.13)
The solution of ( ) with ( ) is assumed because constraints of the
minimization problem are both equations ( ) and inequalities ( ).
At any k-th step, ( ) represents the plant unbalance that vanishes when the solution is
achieved.
2.4.2 Objective Function Definition
In order to solve problems of sizing, optimization, matching an appropriate algorithm
(ECRQP) for the search of the minimum of an objective function has been adopted. In
relation to the issues addressed, the objective function takes on different expressions. Indeed a
set of objective functions nfob R may be established. The global objective function Fob is
1
N
j j
j
Fob w fob
(2.14)
The first element, for 1j , represent the unbalance ( 11 fob ), the other elements may
express a special objectives (like initial cost, operating cost, volume, weight, etc.) and weight
vector elements Tw can take the value zero or one. Of course 1w always must be 1.
Adopting the suitable formulation of the objective function Fob and the vector of unknown
quantities z the following problem may be solved:
Search :min ( , , , , , , ) 0; ( , , , , , , ) 0f f f fz Fob F x b u g r a D x b u g r a (2.15)
Matching constraints and therefore plant unbalance are still taking into consideration.
Pag. 63 of 202
2.5 Solution Methods
Various optimisation techniques based on Equality Constraint Recursive Quadratic
Programming (ECRQP), Genetic Algorithms (GA) and Simulated Annealing (SA) as well as
hybrid GA-ECRQP and SA-ECRQP have been applied and compared [7, 8, 9]. The choice of
the most suitable one depends on the peculiar problem to be solved.
2.5.1 Sequential
The most widely adopted method is the sequential one, by this method the plant is divided
into modules corresponding to the plant components. For each module subsets of equations
and inequalities are established. Each module is analysed sequentially, module outputs are
solved from input quantities. Two major aspects related to the computing time have to be
pointed out. The first is connected with the non linearity of the module equations which
require internal iterations to get outputs. The second is related to closed loops and recycling
streams (i.e. when the module under analysis needs other not yet analysed module outputs
means that those variables have to be given as tentative ones, therefore external iteration
levels in order to have balanced solutions of subsystem process groups). From given data,
usually the solution starts from one module and continues following one fluid streams.
Due to the component equations being non-linear and really numerous for complex plants,
various level (nested) iterative loops are needed. This method requires a big computation
effort and a long CPU time.
2.5.2 Simultaneous
Simultaneous means that all the unknown variables are foreseen (i.e. each assume a proper
value) at the beginning of any step (iteration). Since all the unknown quantities are assumed
in the iteration (see fig. 2.9) the contributions of all the component to the objective function
(components unbalance, costs, etc.) and to the constraint structure may be calculated.
Therefore the plant performance (when it is under an unbalanced condition), costs, emissions
of pollutants and the objective function are evaluated. Components are described by algebraic
relationships and by differential equations which are reduced to algebraic ones by adopting a
finite difference procedure. Performance of a plant component is related to its load level. This
relationship is influenced by its history (ageing, deterioration, fouling, maintenance and so
on).
Pag. 64 of 202
Fig. 2.9 : modular structure calculation method – ECRQP
Problem (2.1, 2.2) could be solved adopting an optimisation technique developed by Cerri et.
Al. [6,7,8,9] based on ECRQP that provides to introduce two merit functions:
the penalty function:
1( , ) ( ) TP z r Fob z v v
r (2.16)
r being the penalty parameter and v being the vector of active constraints.
the Lagrange function:
( , ) ( ) TL z Fob z v (2.17)
being the set of Lagrange multipliers related to the constraints.
Pag. 65 of 202
The parameter r must be positive and when it tends to zero the minimum of ( , )P z r tends to the
minimum of Fob . The minimum of ( , )L z also coincides with the minimum of Fob .
The solution is found starting from an initial tentative solution 0x . At the generic kth iteration the step
kd (which moves the tentative solution from kz to 1k k kz z d ) is searched by solving a quadratic-
programming problem. The objective function is a quadratic approximation of Fob :
1
2
T
q k k k k kF f d d d H
(2.18)
kf being the gradient of the Fob and kH its Hessian matrix, both evaluated at point kz . Secondo
order Taylor’s series expansion around kz lead to approximate expression of the penalty function
gradient:
2
( , ) ( )T T
k k k k k k k k k k
k
P z r f H d A v A A dr
(2.19)
and the Lagrange function gradient:
( , ) T
k k k k k k kL z f H d A (2.20)
kA being the Jacobian matrix of active constraints calculated for kz z .
The search of kd is performed by imposing the condition of minimum ( ( , ) 0)k kP P z r and using
further conditions resulting by equating the right terms of Eqs. (2.16) and (2.17). Therefore the steps
towards the minimum of ( )Fob z are performed along the locus of penalty function minima, as shown
in figure 2.10.
Fig. 2.10: Solution Path along the Locus of P(z,r) Minima
Z
Pag. 66 of 202
2.5.3 Hybrid
The hybrid process consists in the division of the variables into different sets: one is the
Dependent Variables DV that are the unknowns of the independent equation set; the second
variable set consists in the Independent Variables IV that have to be given a priori and do not
change during the calculations. The IV set is made of the degree of freedom DOF’s and of the
Boundary Variables BV or β such as ambient conditions and similar ones. The hybrid
approach consists in dividing the calculation environment into two zone. In the first zone the
IV set is established and the final outputs are saved. The second zone consists in the
calculation of the DV set using the Non Linear Equation Solution NLES that can be performed
by a simultaneous or sequential approach. This hybrid methodology is suitable also for the
solution of optimization problems. In this case the DOF set is divided in two sets. One is ξ that
consists in the DOF’s to be optimized and the remaining IV’s consists in the β set whose
components k remain constant during the calculations.
Accordingly there are three zones:
the first zone inputs inside the calculation process a suitable j
ξ and calculates the
related objective function.
the second zone inputs into the calculation procedure the β set.
the third zone provides the calculation of the unknowns by a simultaneous or a
sequential procedure
Maps of the plant can be calculated by suitably changing the point inside the β domain.
The above procedure is implemented by adopting Genetic Algorithm GA, Simulated
Annealing SA and ECRQP. The GA-ECRQP hybrid algorithm is schematically represented in
figure 2.11.
Since direct application of physical and empirical models to a problem that requires iterative
calculations can lead to a quite long calculation time, alternative simulation procedures must
be considered. In order to perform low CPU occupancy and to get the solution in short time,
ANN techniques have been chosen and applied. The purpose has been to approximate a stated
input-output map that represents the behaviour of the plant. The plant model has been utilized
to generate the database needed for ANN training and testing. Then, single-layer feed forward
networks have been trained with backpropagation algorithm and a parametric simulator of the
Pag. 67 of 202
plant has been produced. Cerri et al. [4,5] extensively discussed neural methodologies to
speed up calculations related to heat and power cogeneration plants.
Fig. 2.11: Hybrid methodology – Genetic Algorithm/ECRQP
The broader and more complex block diagram is shown in figure 2.12 and includes modules,
neural modules and solved through the hybrid algorithm GA-ECRQP. By adoption of
simultaneous – GA Hybrid Algorithm any kind of plant layout can be easy simulated and
optimized.
Pag. 68 of 202
Fig. 2.12: complex modular structure calculation method – Hybrid Algoritm GA-ECRQP
Pag. 69 of 202
2.6 Reference
[1] - Cerri. G., Gazzino M., Borghetti S., 2006: “Hot Section Life Assessment by a Creep
Model to Plan Gas Turbine Based Power Plant Electricity Production”, Proceedings of The
Future of Gas Turbine Technology 3rd International Conference, Bruxelles (B), 11 – 12
October 2006
[2] - Cerri G., Gazzino M., Botta F., Salvini C., (2007): “Production Planning with Hot
Section Life Prediction for Optimum Gas Turbine Management”, Proceedings of the
International Gas Turbine Congress, Tokyo, December 2 – 7, 2007
[3] - Cerri G., Gazzino M., Iacobone F.A., Giovannelli A., (2009): “Optimum Planning of
Electricity Production”, Journal of Engineering for Gas Turbines and Power (Vol.131, Iss.6).
November 2009.
[4] - Cerri, G., Khatri, D. S.,1998, "A Neural Network Approach in Thermodynamic Process
Evaluation," International Conference on Engineering Application of Neural Network
(EANN-98). Gibraltar, Great Britain, June 10-12, 1998, Paper No. 98172
[5] - Boccaletti, C., Cerri, G., Khatri, D. S., Seyedan, B., 1999, "An Application of Neural
Network in Combustion Processes Evaluations," Proceedings of International Conference on
Enhancement & Promotion of Computational Methods in Engineering & Science (EPMESC
VII), Macao, China, August 2-5, 1999
[6] - Cerri, G., Borghetti, S., Salvini, C., 2006, "Neural Management for Heat and Power
Cogeneration Plants," Engineering Applications of Artificial Intelligence, Vol. 19, pp. 721-
730.
[7] - Cerri, G.; Monacchia, S.; Salvini, C., 1994, “Development of Gas - Steam Combined
Cycles Equipped with Coal PFBC by Using an ECRQP Simultaneous Solution Method ,”
Workshop on Cycle Development, University of Essen, 15 dic.
[8] - Cerri G., Boccaletti C., Salvini C. (2000): “Algoritmi deterministici ed evolutivi naturali
nell’ottimizzazione della gestione di impianti cogenerativi”,55° Congresso ATI, Matera, 15-
20 Settembre, 2000
[9] - Cerri G., (1996): “A Simultaneous Solution Method Based on a Modular Approach for
Power Plant Analyses and optimized Designs and Operations”, ASME paper 96-GT-302,
International Gas Turbine and Aeroengine Congress and Exhibition, Birmingham, UK, June,
10-13, 1996.
[10] - Biggs M. C., 1972, “Constrained Minimization Using Recursive Equality Quadratic
Programming,” Numerical Methods for Non Linear Optimization, 1F. A. Lootsma
ed.,Academic Press, London.
Pag. 70 of 202
Chapter III
IGCC Component Models
3.0 Introduction
A model is a numerical structure able to replicate one or more components of a complex
system such as IGCC power plants. From a qualitative point of view a physical model is
based on observation of phenomena and on the sensitiveness of the model designer.
Meanwhile, from the quantitative point of view, a mathematical model must rely on a
mathematical formulation able to describe as well as possible the overall phenomena. By
adopting ‘physical-mathematical’ models description, in the following paragraphs details on
the formulation of the models developed and adapted to represent the H2-IGCC power plant
behaviour is given. Fluid properties (gas and water), gas turbine component models
(compressor, expander, etc.) and steam cycle component models are treated.
3.1 Fluid Properties
Processes occurring both in Gas Turbine and in Steam Cycle components have been evaluated
by the working fluid models [4]. Existing subroutines used to evaluate gas properties,
enthalpies and other quantities had been developed taking Oxygen, Nitrogen, Water (Steam)
and Carbon Dioxide into consideration. In this work, to account the H2-IGCC Project Syngas,
such subroutines have been updated to take also other species (carbon monoxide, Hydrogen)
related to the Syngas production into account. Accordingly, in the next paragraphs, more
details concerning the gas and steam properties are given.
3.1.1 Gas Properties
The updated subroutines have been reported:
ENGA5: this routine calculate the specific enthalpy, steam partial pressure and
saturation temperature of gas phase mixture for a given pressure (pi), temperature (Ti),
and fractions of compositions [xx]k (consisting of O2, N2, CO2, H2O, CO, H2, SO2). In
this routine the heat of vaporization of water vapor is being considered. Subroutine
input and output are given in figure 3.1.
COGAS5: this routine calculate the specific enthalpy (hg), the constant pressure and
volume specific heat (cp,cv), the gas constant (R), the heat ratio (k), the isentropic
exponent (ɛ), the sound velocity (cs) of gas phase mixture for a given pressure (pi),
temperature (Ti), and fractions of compositions [xx]k (consisting of O2, N2, CO2, H2O,
CO, H2, SO2). Subroutine input and output are given in figure 3.2.
Pag. 71 of 202
Fig. 3.1: Block Scheme of the ENGA 5 Subroutine
Fig. 3.2: Block Scheme of the COGAS 5 Subroutine
Thg: temperature[K]
pi: gas total pressure [MPa ]
Fo2: oxygen mass fraction
Fn2: nytrogen mass fraction
Fco2: carbon dioxide mass fraction
Fh2o: vapore mass fraction
Fh2: hydrogen mass fraction
Fco: carbon monoxide mass
fraction
Fso2: sulfhur mass fraction
Subroutine ENGA5
Hga5: specific enthalpy [kJ/kg]
pv:steam partial pressure [kPa]
Tsat: saturation temperature
[K]
Thg: temperature[K]
pi: pressure [MPa]
Fo2: oxygen mass fraction
Fn2: nytrogen mass fraction
Fco2: carbon dioxide mass fraction
Fh2o: vapore mass fraction
Fh2: hydrogen mass fraction
Fco: carbon monoxide mass
fraction
Fso2: sulfhur mass fraction
SubroutineCogas5
cpg: constant pressure specific heat [kJ/kg*K]
rg: gas constant[kJ/kg*K]
cv: constant volume specific heat [kJ/kg*K]
xk: cpg/cv [#]
eps: rg/cpg [#]
cs: sound velocity of gas m/s
rhog: density [kg/m3]
Pag. 72 of 202
Possibility of establish Low Heating Value, Molecular Wight, mass composition and other
relevant quantities of a gas mixture, taking the molar fraction of the syngas produced by the
gasification island into consideration, leads to develop a subroutine. Adoption of empirical
correlation taken from the State of the Art [1,2,3] allow to establish the LHV taking molar
fraction or mass fraction of the whole gas mixture into consideration.
SYGPROP: this routine calculate the LHV, mass fractions of compositions [xx]k
(consisting of O2, N2, CO2, H2O, CO, H2, SO2), constant pressure specific heat (cp),
molecular weight of the mixture (W) of a gas mixture characterized by a molar
composition (consisting in O2, N2, CO2, H2O, CO, H2, SO2) a temperature Ti and
pressure pi. Subroutine Input and output are given in figure 3.3
Fig. 3.3: Block Scheme of the SYGPROP Subroutine
3.1.2 Steam properties
Existing water steam thermodynamic functions [23] have been improved to have a shorter
computing time. Modifications to traditional subroutines have been introduced, to have an
exact matching of enthalpy values (and of other quantities) on both saturated water and
saturated steam lines when such quantities are calculated with functions valid in two
adjoining regions. In order to calculate the thermodynamic quantities below the triple point,
properties of solid phase and solid-vapor mixture phase have been added to the routines.
Ti: temperature[K]
pi: pressure [MPa]
Fo2: oxygen molar fraction
Fn2: nytrogen molar fraction
Fco2: carbon dioxide molar fraction
Fh2o: vapore molar fraction
Fh2: hydrogen molar fraction
Fco: carbon monoxide molar
fraction
Fso2: sulfhur molar fraction
Subroutine SYGPROP
Fo2: oxygen mass fraction
Fn2: nytrogen mass fraction
Fco2: carbon dioxide mass fraction
Fh2o: vapore mass fraction
Fh2: hydrogen mass fraction
Fco: carbon monoxide mass fraction
Fso2: sulfhur mass fraction
cpg: constant pressure specific heat [kJ/kg*K]
LHV: Low Heating Value [MJ/kg]
W: Molecular Weight
[kg/mol]
Pag. 73 of 202
3.1.3 Working fluid properties
Air composition and combustion product composition of a CH4 combustion with an Air Fuel
Ratio (AFR) of some 45 are given in table 3.0a and 3.0b, respectively.
Table 3.0a: Wet Air – RH60% mass fraction composition
Wet Air - RH 60%
O2 0.2312
N2 0.7620
CO2 0.0005
H2O 0.0063
Table 3.0b: Gas Mass Fraction Composition of CH4 combustion with an 45 AFR
Combustion
Product Gas
O2 0.1372
N2 0.7462
CO2 0.0605
H2O 0.0561
The Nitrogen of these compositions being the so called Atmospheric Nitrogen taking the
Argon and other minor species into account. Properties of the ISO air and of the table 3.0b
combustion products gas are given in the following tables 3.0c and 3.0d, respectively.
The first column gives the Celsius Temperatures, the corresponding values are gas constant
R= [kJ/(kgK)]; specific constant pressure heat cp [kJ/(kgK)]; heat ratio k=cp/(cp-R)
[#], the isentropic exponent ɛ =R/cp = (k-1)/k [#] and the sound velocity a [m/s]. The average
values of cp , k and are calculated as mean values in the temperature range from 15°C to
the actual.
Pag. 74 of 202
Table 3.0c: ISO Air Properties
Actual Values Average Values
T R (T) cp (T) k (T) ɛ (T) a (T) cp (15,T) k (15,T)
(15,T)
[°C] [kJ/(kg/K)] [kJ/(kg/K)] [#] [#] [m/s] [kJ/(kg/K)] [#] [#]
15.0 0.289 1.016 1.398 0.285 341.4 1.016 1.398 0.285
50.0 0.289 1.018 1.397 0.284 361.4 1.017 1.398 0.285
100.0 0.289 1.022 1.395 0.283 388.0 1.019 1.397 0.284
150.0 0.289 1.029 1.391 0.281 412.7 1.022 1.395 0.283
200.0 0.289 1.037 1.387 0.279 435.7 1.026 1.393 0.282
250.0 0.289 1.047 1.382 0.276 457.3 1.031 1.390 0.281
300.0 0.289 1.058 1.376 0.273 477.7 1.037 1.387 0.279
350.0 0.289 1.070 1.371 0.270 497.1 1.043 1.385 0.278
400.0 0.289 1.082 1.365 0.267 515.6 1.049 1.382 0.276
450.0 0.289 1.094 1.359 0.264 533.3 1.055 1.379 0.275
500.0 0.289 1.107 1.354 0.261 550.3 1.061 1.376 0.273
550.0 0.289 1.119 1.349 0.259 566.7 1.067 1.374 0.272
600.0 0.289 1.130 1.344 0.256 582.7 1.073 1.371 0.270
650.0 0.289 1.141 1.340 0.254 598.1 1.078 1.369 0.269
700.0 0.289 1.152 1.335 0.251 613.2 1.084 1.367 0.268
750.0 0.289 1.162 1.332 0.249 627.8 1.089 1.365 0.267
800.0 0.289 1.171 1.328 0.247 642.1 1.093 1.363 0.266
850.0 0.289 1.179 1.325 0.245 656.1 1.098 1.362 0.265
900.0 0.289 1.187 1.322 0.244 669.9 1.102 1.360 0.264
950.0 0.289 1.195 1.319 0.242 683.3 1.105 1.359 0.263
1000.0 0.289 1.202 1.317 0.241 696.5 1.109 1.358 0.263
1050.0 0.289 1.209 1.315 0.239 709.4 1.112 1.356 0.262
1100.0 0.289 1.215 1.312 0.238 722.0 1.115 1.355 0.261
1150.0 0.289 1.221 1.310 0.237 734.5 1.118 1.354 0.261
1200.0 0.289 1.227 1.309 0.236 746.8 1.121 1.353 0.260
1250.0 0.289 1.232 1.307 0.235 758.8 1.124 1.353 0.260
1300.0 0.289 1.237 1.305 0.234 770.7 1.126 1.352 0.259
1350.0 0.289 1.242 1.304 0.233 782.4 1.129 1.351 0.259
1400.0 0.289 1.246 1.302 0.232 794.0 1.131 1.350 0.259
1450.0 0.289 1.250 1.301 0.231 805.3 1.133 1.350 0.258
1500.0 0.289 1.255 1.300 0.231 816.5 1.135 1.349 0.258
1550.0 0.289 1.259 1.298 0.230 827.5 1.137 1.348 0.257
Pag. 75 of 202
Table 3.0d:Gas Properties of CH4 combustion with an 45 AFR
Actual Values Average Values
T R (T) cp (T) k (T) ɛ (T) a (T) cp (15,T) k (15,T)
(15,T)
[°C] [kJ/(kg/K)] [kJ/(kg/K)] [#] [#] [m/s] [kJ/(kg/K)] [#] [#]
15.0 0.295 0.951 1.448 0.310 350.6 0.951 1.448 0.310
50.0 0.294 1.062 1.384 0.277 362.8 1.007 1.416 0.293
100.0 0.294 1.068 1.381 0.276 389.5 1.010 1.415 0.293
150.0 0.295 1.076 1.377 0.274 414.2 1.014 1.412 0.292
200.0 0.295 1.087 1.372 0.271 437.2 1.019 1.410 0.290
250.0 0.295 1.098 1.366 0.268 458.8 1.025 1.407 0.289
300.0 0.295 1.111 1.361 0.265 479.2 1.031 1.405 0.287
350.0 0.295 1.124 1.355 0.262 498.6 1.038 1.402 0.286
400.0 0.295 1.138 1.349 0.259 517.1 1.045 1.399 0.284
450.0 0.295 1.152 1.343 0.256 534.9 1.052 1.396 0.283
500.0 0.295 1.166 1.338 0.253 551.9 1.059 1.393 0.281
550.0 0.295 1.180 1.333 0.250 568.3 1.066 1.390 0.280
600.0 0.295 1.194 1.328 0.247 584.3 1.072 1.388 0.278
650.0 0.295 1.206 1.323 0.244 599.7 1.079 1.386 0.277
700.0 0.295 1.219 1.319 0.242 614.8 1.085 1.384 0.276
750.0 0.295 1.230 1.315 0.239 629.4 1.091 1.382 0.275
800.0 0.295 1.241 1.311 0.237 643.7 1.096 1.380 0.273
850.0 0.295 1.251 1.308 0.235 657.7 1.101 1.378 0.272
900.0 0.295 1.261 1.305 0.234 671.4 1.106 1.377 0.272
950.0 0.295 1.270 1.302 0.232 684.8 1.111 1.375 0.271
1000.0 0.295 1.279 1.299 0.230 697.9 1.115 1.374 0.270
1050.0 0.295 1.287 1.297 0.229 710.9 1.119 1.373 0.269
1100.0 0.295 1.295 1.295 0.228 723.5 1.123 1.371 0.269
1150.0 0.295 1.302 1.292 0.226 736.0 1.127 1.370 0.268
1200.0 0.295 1.309 1.290 0.225 748.2 1.130 1.369 0.267
1250.0 0.295 1.315 1.289 0.224 760.2 1.133 1.368 0.267
1300.0 0.295 1.321 1.287 0.223 772.1 1.136 1.368 0.266
1350.0 0.295 1.327 1.285 0.222 783.8 1.139 1.367 0.266
1400.0 0.295 1.333 1.284 0.221 795.3 1.142 1.366 0.265
1450.0 0.295 1.338 1.282 0.220 806.6 1.145 1.365 0.265
1500.0 0.295 1.344 1.281 0.219 817.8 1.148 1.365 0.264
1550.0 0.295 1.349 1.279 0.218 828.7 1.150 1.364 0.264
Steam and water enthalpies are given in table 3.0e for the three different pressure,
representing the reference HRSG steam production. Accordingly, for each pressure liquid
region is marked by blue cells and steam one by red cells. Moreover, saturation temperature
for each pressure is given.
Pag. 76 of 202
Table 3.0e: Steam Properties for different pressure
Blue – Water ; Red - Steam
Tsat [°C] 336.6 Tsat [°C] 254.7 Tsat [°C] 143.6
T[°C] h[kJ/kg] T[°C] h[kJ/kg] T[°C] h[kJ/kg]
30.0 138.3 30.0 129.5 30.0 126.0
40.0 179.7 40.0 171.1 40.0 167.7
50.0 221.1 50.0 212.8 50.0 209.4
60.0 262.6 60.0 254.5 60.0 251.2
70.0 304.2 70.0 296.3 70.0 293.1
80.0 345.8 80.0 338.1 80.0 335.0
90.0 387.5 90.0 380.0 90.0 377.0
100.0 429.3 100.0 422.0 100.0 419.1
110.0 471.3 110.0 464.2 110.0 461.3
120.0 513.3 120.0 506.4 120.0 503.7
130.0 555.5 130.0 548.9 130.0 546.2
140.0 597.8 140.0 591.5 140.0 588.9
150.0 640.4 150.0 634.3 150.0 2753.4
160.0 683.1 160.0 677.3 160.0 2777.0
170.0 726.2 170.0 720.7 170.0 2799.1
180.0 769.5 180.0 764.4 180.0 2820.4
190.0 813.1 190.0 808.5 190.0 2841.3
200.0 857.2 200.0 853.0 200.0 2862.1
210.0 901.6 210.0 898.0 210.0 2882.7
220.0 946.6 220.0 943.6 220.0 2903.2
230.0 992.2 230.0 989.9 230.0 2923.7
240.0 1038.4 240.0 1037.0 240.0 2944.2
250.0 1085.4 250.0 1085.1 250.0 2964.6
260.0 1133.4 260.0 2818.4 260.0 2985.1
270.0 1182.4 270.0 2855.2 270.0 3005.6
280.0 1232.7 280.0 2888.7 280.0 3026.1
290.0 1284.6 290.0 2920.0 290.0 3046.6
300.0 1338.4 300.0 2949.5 300.0 3067.1
310.0 1394.9 310.0 2977.8 310.0 3087.7
320.0 1455.1 320.0 3005.1 320.0 3108.3
330.0 1520.9 330.0 3031.7 330.0 3128.9
340.0 2656.6 340.0 3057.6 340.0 3149.6
350.0 2740.2 350.0 3083.1 350.0 3170.3
360.0 2801.0 360.0 3108.1 360.0 3191.0
370.0 2855.2 370.0 3132.8 370.0 3211.8
380.0 2904.3 380.0 3157.2 380.0 3232.6
390.0 2949.2 390.0 3181.4 390.0 3253.5
400.0 2990.8 400.0 3205.4 400.0 3274.4
410.0 3029.7 410.0 3229.3 410.0 3295.4
420.0 3066.5 420.0 3252.9 420.0 3316.5
430.0 3101.4 430.0 3276.5 430.0 3337.5
440.0 3134.9 440.0 3300.0 440.0 3358.7
450.0 3167.2 450.0 3323.4 450.0 3379.9
460.0 3198.5 460.0 3346.7 460.0 3401.1
470.0 3228.9 470.0 3370.0 470.0 3422.5
480.0 3258.6 480.0 3393.2 480.0 3443.8
490.0 3287.7 490.0 3416.4 490.0 3465.3
500.0 3316.2 500.0 3439.6 500.0 3486.8
510.0 3344.4 510.0 3462.7 510.0 3508.3
520.0 3372.1 520.0 3485.8 520.0 3529.9
530.0 3399.5 530.0 3508.9 530.0 3551.6
p [bar] = 140 p [bar] = 43 p [bar] = 4
Pag. 77 of 202
3.2 Gas Turbine Component Models
Description of GT component models and of the global model developed to perform the
Equivalent Brayton Cycle calculation and overall coolant flows calculation is given in the
following paragraphs.
3.2.1 300MW F Class GT Brayton Cycle Evaluation Model
On the basis of the State of the Art and on the Reference Data Base (1.3.1, tables 1.4,1.5) a
preliminary evaluation of an Equivalent Brayton Cycle (EBC) has been performed to establish
global parameters for the development of the Generic 300MW F Class GT Simulator.
Nomenclature of main quantities has been assumed according to figure 3.4, in which main
sections of GT are given and according to figure 3.5, in which Brayton Cycle is represented in
a T-S chart:
Fig. 3.4-a: Scheme of a Generic 300MW F Class GT
Fig.3.4-b: Scheme of a GT Brayton Cycle
Pag. 78 of 202
For the evaluation of the cycle calculation some equations can be written taking fluid
properties and processes into consideration. With reference to the schematic representation of
the figs. 3.4a and 3.4b the equation can be sorted for Sections (Compressor, Combustion
Chamber, Expander, Gas Turbine and Equality Constraints). For every section equations
representative of the constitutive equation of the fluids and are sorted under the ‘voice’
Stations, while the equation describing the processes (compression, expansion, pressure loss,
etc.) are listed in the respectively Section. Generic Station j-th, in which fluid properties are
evaluated., include RO3 Fluid Properties calculation methodology that has been discussed in
the [7].
3.2.1.1 Compressor Section
For a fixed shaft speed the following equations for the compressor can be written. The inlet
mass flow is characterized by the ISO conditions. Air properties are expressed by the (3.1)
and (3.2)
Station 1 – Constitutive Equations :
1 1 1 1 1( , , ,[ ] ) 0f T p xx (3.1)
2 1 1 1 1( , , ,[ ] ) 0f h T p xx (3.2)
The air at the compressor exit is characterized by a pressure p2 and a temperature T2. Air
properties at the compressor exit for the an isentropic compression and for the polytropic one
are expressed by (3.3), (3.4), (3.5), (3.6):
Station 2 - Constitutive Equations:
3 2 2 2 1( , , ,[ ] ) 0s sf T p xx (3.3)
4 2 2 2 1( , , ,[ ] ) 0s sf h T p xx (3.4)
5 2 2 2 1( , , ,[ ] ) 0f T p xx (3.5)
6 2 2 2 1( , , ,[ ] ) 0f h T p xx (3.6)
Equations of the Process:
7 1 2( , , ) 0f p p (3.7)
8 2 1( , , ) 0Cs sf L h h (3.8)
9 2 1( , , ) 0Cf L h h (3.9)
10( , , ) 0C Cs icf L L (3.10)
Pag. 79 of 202
3.2.1.2 Combustion Chamber Section
Pressure loss across the combustor, conservation of energy and chemical equations are taken
in the combustion chamber process into consideration, as written in (3.11), (3.12) and (3.13)
Equations of the Process:
11 3 2( , , ) 0ccf p p p (3.11)
12 3 1([ ] ,[ ] ,[ ] , ) 0ff xx xx xx (3.12)
13( , , ) 0ff Q LHV (3.13)
3.2.1.3 Expander Section
Equations for isentropic and polytropic expansion can be written as well as for the
combustion product gas properties both for the inlet Station 3 (3.14), (3.15), both for the
outlet Station 4 (3.16), (3.17), (3.18) and (3.19).
Station 3 - Constitutive Equations:
14 3 3 3 3( , , ,[ ] ) 0f T p xx (3.14)
15 3 3 3 3( , , ,[ ] ) 0f h T p xx (3.15)
Station 4 - Constitutive Equations:
16 4 4 4 3( , , ,[ ] ) 0s sf T p xx (3.16)
17 4 4 4 3( , , ,[ ] ) 0s sf h T p xx (3.17)
18 4 4 4 3( , , ,[ ] ) 0f T p xx (3.18)
19 4 4 4 3( , , ,[ ] ) 0f h T p xx (3.19)
Equations of the Process:
20 4 1( , , ) 0ef p p p (3.20)
21 3 4( , , ) 0ef p p (3.21)
22 3 4( , , , ) 0Es sf L h h (3.22)
23 3 4( , , , ) 0Ef L h h (3.23)
24( , , ) 0E Es ief L L (3.24)
Pag. 80 of 202
3.2.1.4 Gas Turbine Equations
Global relations between Gross Power, Gross efficiency, Mechanical Efficiency, Electric
Generator Efficiency, exhaust mass flows, fuel and others related to the GT specifications are
reported below:
25( , , ) 0ex f cif m m m (3.25)
26( , , , ) 0f GTf m LHV P (3.26)
27 ( , , ) 0GTi E Cf L L L (3.27)
28( , , , ) 0GT GTi m gef L L (3.28)
29( , , ) 0GT f GTf L Q (3.29)
EQUALITY Constraints
These equality constraints (1-5) represent the assumption made, according with the Data
available from the State of the Art:
*
4(1) 0exge T T (3.30)
*(2) 0GT GTge (3.31)
*(3) 0ex exge m m (3.32)
*(4) 0ge P P (3.33)
*(5) 0ge (3.34)
The following equality constraints are auxiliary for the non-linear equation solution
methodology
(6) ( / ) 0GT cige L P m (3.35)
(7) ( / ) 0ci fge m m (3.36)
2 1 1 2(8) ( , , , ) 0s sge T f T (3.37)
2 1 1 2(9) ( , , , , ) 0pcge T f T (3.38)
4 3 3 4(10) ( , , , , ) 0e pege T f T (3.39)
4 3 3 4(11) ( , , , ) 0s s ege T f T (3.40)
3 2(12) , , , 0ccge h f h LHV (3.41)
Pag. 81 of 202
Variables involved in the 41 equations are the following:
17 - 1 2 1 1 1 1 2 2 2 2 2 2, , , , ,[ ] , , , , , , , , , , ,s s s pc Cs C icp p T xx h T h T h L L
8 - 3 3, ,[ ] , ,[ ] , , ,cc f cc fp p xx xx LHV Q
16 - 4 3 3 3 4 4 4 4 4 4, , , , , , , , , , , , , , ,e e s s s pe Es E iep p T h T h T h L L
9 - , , , , , , , ,ex f ci GTi m ge GT GTm m m P L L
5 - * * * * *, , , ,ex GT exP m T
variables: v =55
equations: eq =41 unknown variables = 41
independent variables + data: 𝝃 =(55-41) =14
Boundary Conditions Data b =3
1. inlet pressure 1p
2. inlet temperature 1T
3. inlet Air Composition ([xx]j are considered as a single variable) 1[ ]xx
Data and Assumption related to the Gas Turbine (fuel, pressure drop, etc.) are:
- From the Available Manufacturers Data
1. Gas Turbine Gross Power *P
2. Exhaust Mass Flow *
exm
3. net efficiency *
GT
4. Exhaust Temperature *
exT
5. Pressure Ratio *
- Assumption made according to the State of the Art
6. fuel composition [ ] fxx
7. Low Heat Value LHV
8. pressure drop across Comb. Chamber 2/ccp p
9. pressure drop across the exhaust duct (back pressure depending of the downstream
system HRSG and Chimney) 1/ep p
10. mechanical efficiency m
11. electric generation efficiency ge
Data: d =11
Under this Assumption and this Boundary Condition Data, the set of equation is satisfied.
Pag. 82 of 202
3.2.1.5 GT Global Model for the evaluation of the overall cooling mass flow
Taking results of the uncooled equivalent Brayton cycle calculation as well as the
technological level (Class) of the Gas Turbine into consideration, the overall coolant mass
flow required to perform the cooling purposes can be evaluated by the adoption of global
models. Such models relate the overall coolant mass flow to some relevant temperatures
(compressor outlet temperature, metal temperature, firing temperature that is strictly related to
the TIT), to the compressor inlet or expander inlet mass flow, to the main flow and coolant
flow properties and to some parameters that well represent the Class of the Gas Turbine
(introduction of some coefficients). The overall coolant flow can be express as a function of
such parameters:
2
1
k
g pg f b
pc b cex
m c T Tmc k
c T T
(3.42)
Similar Correlation have been found in the technical background [12, 13, 14].
TIT is a relevant temperature because it relates the overall coolant mass flow to the inlet hot
gas mass flow entering the gas expander and the coolant temperatures to the firing
temperature. Scheme of TIT calculation is given in figure 3.5. TIT is defined in [24].
Fig. 3.5: Turbine Inlet Temperature Nomenclature
Accordingly, TIT can be approximated by the rule (3.43) according to figure .5:
g pg g cj pj cj
j
mix pmix
m c T m c T
TITm c
(3.43)
mixm being the sum of the various coolant flows cjm and of the gas mass flow
gm and pmixc
being the pressure constant specific heat of the mixture depending on many parameters.
0 0[ ( , ), ( , )]pmix pg pcjc f c TIT T c TIT T (3.44)
Pag. 83 of 202
3.2.2 Compressor
Compressor component models is based on conservation equations of mass, energy,
momentum and entropy. Moreover, auxiliary and constitutive equations have been adopted to
establish the source terms of conservation equations. Accordingly, work exchange has been
evaluated taking losses in a global manner into consideration. Such losses are correlated to
incidence (i) and deviation (δ) angles:
i = f ( m, ρ, u, Ω) (3.45)
δ = f (i, θ, Ma, Re, l/s) (3.46)
Profile losses on the blade surfaces, skin friction losses on the annulus walls and secondary
losses are taken into account using various empirical correlations available in literature.
Different relationships have been used for different blading whose features are stored in DBs
embedded in the model.
Compressor has been modelled by following a modular approach that takes each blade raw
FV’s into consideration. Lumped approach described in 2.2 paragraph has been followed. In
figure 3.6 the sketch of the stations and central nodes FV representation of the axial
compressor component model is depicted.
Fig. 3.6: Sketch of compressor through Flow Section
Pag. 84 of 202
According to fig.3.7, compressor has been schematically divided in four bodies. The first one
represents stages from 1 to 5, the second one takes stages from 6 to 9 into account and so on
for the other bodies. Bleed extraction of compressed air has been taken at the end of each
body into account.
Fig. 3.7: Compressor sub-components to account the bleed extraction
3.2.3 Combustion Chamber
Combustion Chamber CC has been treated as a zero dimensional model in which performance
are lumped. The model takes the combustion products, pressure losses and the pollutants into
account. In figure 3.8 scheme of the combustion chamber model is given.
Fig. 3.8: Sketch of combustion chamber component model
Pag. 85 of 202
Inlet
mai [kg/s] Air inlet mass flow
pai [kPa] Air inlet pressure
Tai [°C] Air inlet temperature
[xx]ai [#] Air inlet mass fraction composition
mf [kg/s] Fuel inlet mass flow
pf [kPa] Fuel inlet pressure
Tf [°C] Fuel Inlet Temperature
[xx]f [#] Fuel mass fraction composition
LHV [MJ/kg] Fuel Low Heating Value
Alias quantities (Oi) are introduced into the model to take the injection of some ballast mass
flow (nitrogen) or steam characterized by a composition, temperature and pressure into
consideration. Equations describing the CC model allow to described any kind of injected
flows into combustion chamber.
Outlet
mgo [kg/s] Outlet combustion products mass flow
pgo [kJ/kg] Outlet combustion products pressure
Tgo [kg/s] Outlet combustion products temperature
[xx]go [#] Outlet combustion products mass fraction composition
Combustion chamber component model is based on mass, energy, momentum and entropy
conservation and equations describing its behavior. Implicit formulation of such equations is
given:
1( , , , ) 0ai f go Oif m m m m (3.47)
2( , , , , , , , , ,[ ] , , ) 0Gi Go F Oi Gi Go Oi F F F bf m m m m h h h p T XM LHV (3.48)
3( , , , ) 0ai f go Oif p p p p (3.49)
being , ,Gi Go Oih h h inlet, outlet gas and alias specific enthalpies, respectively:
4( , , ,[ ] ) 0Gi Gi Gi Gif h p T XM (3.50)
5( , , ,[ ] ) 0Go Go Go Gof h p T XM (3.51)
6( , , ,[ ] ) 0Oi Oi Oi Oif h p T XM (3.52)
Pag. 86 of 202
Chemical reaction have been assumed at equilibrium and stoichiometric equations are given,
according to the fuel composition.
2 2C O CO (3.53)
2 2S O SO (3. 54)
2 2 22 2H O H O (3. 55)
The gas outlet composition [ ]GoXM depends both on the product of fuel oxidation and on air
fuel ratio r :
7 ([ ] ,[ ] , ,[ ] ) 0Gi F r Gof XM XM XM (3.56)
For the off-design calculation, thermal losses are established as a function of pressure and
temperature difference between two streams in combustion chamber using generalized
relationships with respect to the reference data. Figure 3.9 shows the curves for off-design
calculation of combustion efficiency versus the temperature difference between two streams
varying inlet pressures:
f8(pgi, Tcc, kcc, b )=0 (3.57)
(pi) being the combustion chamber inlet pressure, (Tcc) the difference between the two
stream temperatures in across combustion chamber, and (kcc) is the combustion chamber
correction factor and is obtained from the design calculation, respectively.
Fig. 3.9: Combustion Chamber Off-Design Curves
0.6
0.7
0.8
0.9
1.0
0.1 0.3 0.5 0.6 0.8 1.0
η/η
*
ΔT/ΔT* 0.3bar 0.7bar 1bar 20.7bar
Pag. 87 of 202
3.2.4 Expander Model
According with the modelling approach described in the paragraph 2.2, the Expander model
has been developed on a Finite Volume (FV) approach, in which cascade lumped features
(work, losses, etc.) are reduced to the FV central node, while spatial distribution of
temperature, pressure, velocity, etc. are averaged on the boundary surfaces on an approximate
stream line. Such a representation is given in figure 3.10
Fig. 3.10: Sketch of the Expander through Flow Section
For each volume (blade row) the following equations are established:
conservation of mass, momentum, energy, entropy.
constitutive of the system.
auxiliary.
Auxiliary equations express the real behaviour of the system describing processes and
phenomena. They substantially describe the source terms in the conservation equations.
During the development of the expander component model the complex phenomena of the
turbine cooling has been treated, taking additional losses related to the various aspects
(momentum conservation, heat transfer process, mixing, etc) into account. Uncooled and
cooled expansion, as also said in [5,6] are characterized by different isentropic efficiencies
because of the various losses.
The Cooled Expander Model (CEM) accounts for the uncooled losses [11] and for other
additional losses related to the various phenomena, previously stated. These losses are strictly
connected to the entropy sources owing to the cooling process. In figure 3.11, a schematic
representation of a expander cooled row is given:
Pag. 88 of 202
Fig. 3.11: Schematic representation of a expander cooled row
With reference to Fig.3.11, the various aspects of the cooling process are described.
AB
Uncooled expansion of the main flow mgi (hot gas entering the j-th row) is taken into
consideration as well as its losses.
o Airfoil Profile Loss
o Windage Loss
o Nozzle End Loss
o et.c
All this losses (some 6%) can be connected with the loss of kinetic energy related to the
difference between the isentropic flow and the real flow:
2
is
v
v
(3.58)
v and vis being respectively the velocity and the isentropic velocity. For a Nozzle Vane, v is
the absolute velocity c, while for a Rotor Blade, v is the relative velocity w.
During the expansion the blade has been seen as a heat transfer device and a heat rate of the
main flow is removed and sent to the coolant flow. On the gas side, the reduction of its
temperature implies ad entropy reduction.
Pag. 89 of 202
CD
On the other hand, taking the coolant stream mc into consideration, the heat removed from the
main stream increase the coolant temperature. The coolant stream temperature increase
represents an entropy production.
DE
Before mixing with the main gas, in the simplified schematization of Fig. 3.11, the coolant
flow reduces its pressure from pD to pE. Coolant stream entropy production is connected to
the pressure loss.
B+EF
Main flow, the hot gas at the end of the expansion, and the coolant flow (heated by the heat
transfer process schematically represented in Fig. 3.11 mix together in the section M of the
scheme of Fig. 3.12. Mixing of two streams implies some entropy production and
consequently some dissipative work.
a - b
Fig. 3.12 Schematic Representation of the Mixing:
a) Momentum Conservation – b) Thermal Equilibrium
The main stream has an higher velocity then the coolant one. This implies for the momentum
conservation that the velocity of the mixture is lower that the initial velocity of the main flow
as depicted in figure3.12–a. Related to this, a kinetic loss as well as a pressure loss have been
taken into consideration. According with [6], in the model for the rotor blades, the rotor
cooling air acceleration is taken into consideration. Due to this aspect a specific pumping
power is required.
Moreover, the two streams, characterized by a certain number of moles nj, have different mass
compositions and for each flow the species have their partial pressure (Dalton’s Low). When
Pag. 90 of 202
the mixing occurs, an entropy source is connected with the expansion of each stream to its
partial pressure (figure 3.12-b).
The achievement of the Thermal Equilibrium is another entropy source that has to be taken
into consideration. All this entropy sources are taken into the model into account as a
Dissipative Work. Losses Connected with the mixing of two streams at different pressures,
temperatures and velocities are some 6-7%, according with [4].
A h-s chart summarizing the mixing of the two streams is represented in figure 3.13.
Fig. 3.13: Cooling and main flow expansion on h-s chart
Pag. 91 of 202
3.2.5 GT Cooling Model
According with the paragraph 2.2 modelling approach is based on a FV lumped feature and
performance discretisation of components. This approach can be easily adopted for a heat
transfer device. The Gas Turbine cooling system can be seen as a complex arrangement of
series and parallel heat transfer devices. Heat transferred from a fluid to the other (the
performance) is related to the lumped flow features and to the geometric features of the
various components by adapting classical heat transfer model. The connection between
component features and heat transfer model is established according to the amount of data
available by detailed simulations.
Accordingly, the GT Simulator takes a GT cooling lumped model into account which implies
transfer of heat from the main flow (hot gas) to the coolant flows, through various
components (blade row, disk, etc.). Moreover, some heat flows from the hottest GT
components (i.e. combustion chamber) to the colder ones (i.e. shaft, casing, etc.).
In the following paragraph a description of the heat transfer scheme and cooling paths in the
gas turbine as well as the cooling model is given.
3.2.5.1 Heat transfer scheme and cooling scheme
Various heat transfer phenomena have to be taken during the design of the cooling system
into consideration to properly design the cooling system [13, 14, 16, 17, 18] . Under the effect
of convection, radiation and conduction the heat of the main stream (the hottest one) flows
through the various GT components, each of them characterized by a thermal gradient.
Accordingly, lot of the Gas Turbine components (disk, shroud, sidewall, blade, cavity, etc.)
need to be cooled by some ‘cold’ air to maintain their temperature under a defined threshold
value. Some coolant flows are required to achieved this purpose. Moreover, coolant flows are
used for services (sealing, balance, etc.). Thus, coolant flows are not solely used for blade
surface cooling, but for all the aspects concerning the GT cooling. Figure 3.14 represents
schematically the cooling flows along a Generic 300MW F Class Gas Turbine.
Moving from the 1st vane of the compressor to the last rotor row of the gas expander, the main
flow path is split in various stations for various purposes, as schematically represented in fig.
3.14. Some fractions of the compressor inlet mass flow are extracted at different compressor
stages and move to the expander stages mixing with the hot gas main flow. Main flow at the
compressor exit is split in various fraction. One is directed to the 1st Nozzle Row, a second
one is addressed to the 1st Rotor Row while the major of them is used for the combustion
process.
Pag. 92 of 202
Fig. 3.14: Cross Section of the Cooling Paths (SIEMENS)
Fig. 3.15: Schematic View of the main stream and coolant streams along the combustor
and of the heat fluxes moving through the GT to the casing and to the inner components (shaft, disk, etc.)
Pag. 93 of 202
All the fluxes are also adopted to cool the combustion chamber externally and internally,
respectively. Indeed, Combustor is also taken in the complex cooling path into consideration
because of the high temperature of the combustion process. Liners of the Annular Combustor
are cooled inside where the flame or combustion occurs. The inner of the liner is cooled by
film and also the Liner Metal Temperature (LMT) is reduced by the interposition of the
Thermal Barrier. The outer of the combustor is protected by the coolant
flows directed to the 1st expander stage. The extracted mass flows, both from the compressor
stages both at the combustor inlet, are not used for the 100% to the surface blade cooling but
also for other features.
As an example of the high complexity of the heat transfer phenomena occurring in the Gas
Turbine, in figure 3.15 a sketch of the heat fluxes moving from the combustion chamber to
the casing and to the shaft is given.
Convection, radiation (especially for the combustion chamber) and conduction phenomena
have to be taken for the GT cooling into account. Indeed, high temperatures are reached
during the combustion process so systems to maintain the component temperature under a
threshold upper limit are usually adopted. Both the coolant flow addressed to the expander
and the main flow sent to the burner lap the outer surface of the liner, while the inner of the
liner is cooled by film and also the Liner Metal Temperature (LMT) is reduced by the
interposition of the Thermal Barrier. Even if the complex system of the combustion chamber
is cooled, some heat fluxes flow through the metal to the casing and to the shaft, respectively.
Taking the outer (casing) and the inner (disks, shaft, etc.) components of the machine into
account, main flows and coolant flow are subjected to convection and radiation heat transfer
phenomena. According to figure 3.15, these streams increase their temperatures moving along
the combustor.
3.2.5.1.1 Flow in the expander stages
Description of the purposes that the coolant flows has to perform allows to better understand
which peculiarities of the GT cooling are taken by the RO3 Lumped Model into account.
According to figure 3.14, the ‘cooling channels’ of the compressor rotor rows extractions are
highlighted by the red circle. This channels lead the coolant flows to the respectively
expander stages in order to cool all the components thermally stressed.
List of the main row components that required to be cooled to maintain their temperatures
under the threshold value is given and by the help of some exemplificative pictures the
expander flow paths are described.
Pag. 94 of 202
Disk
Disk Cavity
Shroud
Platform
Shank
Sidewall
Airfoil Surface
Tip Cap
others
Moreover coolant flows are used for the services. Such a services are as an example the piston
balance, the sealing and other as shown in figure 3.16 below:
Fig. 3.16: Schematic view of the cooling paths along the disks – As example
Coolant mass flows extracted from the compressor stages have different paths and are
addressed both for stator row and for the rotor row. By the simplified adoption of Fig. 3.17, is
possible to better understand which are the various coolant flow paths along the gas turbine.
The path from the extraction (bleeding) sections to the respectively expander rotor row is the
cooling passage represented by L in Fig. 3.17.
The coolant flows pass through the shaft before entering the disk cavity and the disk.
Bleed extractions addressed to the stator (nozzle) rows pass externally (around) the machine
lapping the case before re-entering in the respectively row.
The coolant flow addressed to a Stator Row assuming the schematization of Fig. 8 is used for
various purposes:
Cooling of the Airfoil Surface (inner and outer) - D in the Fig. 3.17
Cooling of the Platform and Sidewall (inner and outer) - E in the Fig. 3.17
Mixing with the main stream, downstream the Stator Vane - G in the Fig. 3.17
Pag. 95 of 202
Extracted mass flow for the Stator Row cooling is used for various cooling surfaces. For this
reason the overall extracted mass flow is split in various fractions, adopted for the various
cooling purposes, respectively. A schematic view of the Stator Row cooled components is
given in Fig. 3.18:
Fig. 3.17: Example of a Generic Gas Turbine Cooling Path along Stator and Rotor Row
Fig. 3.18: Schematic View of the Cooled components of the Stator Row – As Example
L
D
E
E
G
B
A
C
A
F
H H
Pag. 96 of 202
As for the Stator Row also for the Rotor Row, coolant flows are used for various row
components cooling and the overall mass flows (extracted from the compressor) are divided
into minor flows for different purposes:
Cooling of the disk outer – H in the Fig. 3.17
Cooling of the Airfoil Surface (inner and outer) - B in the Fig. 3.17
Cooling of the Shank - C of Fig. 3.17
Sealing - F in the Fig. 3.17
Mixing with the main stream, downstream the Rotor Blade – A in the Fig. 3.17
From the Rotor Blades, heat fluxes move to the shaft passing through the various components.
A typical temperature distribution along the disk is given in Fig. 3.19:
Fig. 3.19: Typical Temperature Distribution along a 1st Stage Aeronautic Rotor Disk – As Example
All these aspects (components cooling, services, sealing, etc.) have to be considered to
evaluate the coolant mass flows and the various temperatures of the phenomena. Indeed, heat
removed from all the hot components (disk, shank, etc.) flows towards the fractions of the
overall coolant flow designed to perform the defined purpose (cooling, service, sealing, etc.).
To ensure that all the temperature of the various components are sufficiently lower than the
threshold value, the bled mass flow is split in various fluxes. A first assumption for all the
Stator Rows, except for the first one, is that a 60% of the overall coolant flow is addressed to
the blade surface cooling and the other 40% is used for the sidewall, platform and for all the
other components previously described. In Fig. 3.20 a detailed figure of the Stator (Nozzle)
Row cooling components is given:
Pag. 97 of 202
Fig. 3.20: Schematic View of a 1st Nozzle Vane Cooling Components – As Example
Coolant mass flows distribution for the Rotor Rows is pretty similar to the Stator Row. Some
65% of the overall extracted coolant flow is used for the blade surface cooling and the rest
some 35% is addressed to the other row components (dovetail serration, shank, platform,
etc.). In Fig. 3.21 a detailed representation of a Rotor Blade cooling components is given:
Fig. 3.21: Schematic View of a 1st Rotor Blade Cooling Components – As Example
Of course even if the stator row and the rotor row blades of the last expander stage are
uncooled (not cooled by internal coolant flows and not film cooled). Some coolant flows are
addressed anyway to that stage because the disks have always to be cooled. Thus a heat flux
from the hot parts to the cold one exists. In figure 3.22, a schematically comparison of the
Pag. 98 of 202
various coolant flows between the cooled blade and uncooled blade is sketched. Moreover,
the RO3 GT cooling model is schematically represented in figure 14.
Fig. 3.22: Comparison between cooled blade and uncooled blade coolant flow
Each Heat transfer process is characterized by a ‘heat transfer effectiveness’ if a
Effectiveness - Number of Transfer Unit (ɛ-NTU) approach is adopted to model the GT
cooling system. Taking the various heat transfer phenomena characterized by a certain
effectiveness into account, the coolant flow fraction distributions has been evaluated
according to the above:
Table 3.1: Fractions of the overall mass flow for each row (in percentage %)
STAGE
ROW 1S 1R 2S 2R 3S 3R 4S 4R
Airfoil Surface 50 65 60 65 60 65 0 0
Other Purposes
(endwall, shroud,
sealing, etc)
30 35 40 35 40 35 100 100
Jet Cooling 20 0 0 0 0 0 0 0
Coolant Mass Flow Percentage % for the various purposes
1st Stage 2nd Stage 3rd Stage 4th Stage
Pag. 99 of 202
Fig. 3.23: Schematic View of the cooling path
from the compressor bleeding station to the expander row injection station
Pag. 100 of 202
Taking the description of the cooling paths along the machine into consideration, in such a
lumped model the coolant flows consider both the airfoil blade cooling and the cooling of the
other parts (disk cavities, shrouds, endwalls (sidewall) and the action of coolant as sealant
flow re-entering into the main flow. Temperatures (coolant, blade, etc.) have the meaning of
lumped reference temperature of the complex cooling process. In figure 3.23, the schematic
view of the complex parallel and series equivalent heat transfer devices taken into account by
the model is given.
3.2.5.2 Blade Cooling Model
Gas Turbine Blade Cooling can be seen as a series of layers characterized by different heat
transfer phenomena. In figure 3.24 sketch of that schematization is depicted:
Fig. 3.24: Sketch of a Rotor Blade temperature distribution along the layers
Moving from the inner side (coolant) to the outer one (main stream) the following heat
transfer layers can be described:
o Internal Cooling Flow Bulk Material: convection heat transfer
Coolant mass flow entering the blade is used to remove the heat flowing from the
metal. Flow velocity, gas composition, architecture and geometry of the blade are
some parameters that influence the internal convection heat transfer phenomena.
o Bulk Material and Thermal Barrier Coating: conduction heat transfer
Both for the Bulk Material (BM) and for the Thermal Barrier Coating (TBC) the
heat flux coming from the outer surface passes through the various conductive
Pag. 101 of 202
layers characterized by a thickness js and by a thermal conductivity j , that is a
function of the heat transfer temperatures.
o Thermal Barrier Coating Hot gas: prevalent convection heat transfer
The hot gas exiting the combustion chamber and entering the expander is at high
temperature. The model takes both the radiation effects and the convection into
account by considering the heat transfer as a prevalent convection phenomena.
By the adoption of the most suitable expression, the hot gas prevalent convection
heat transfer coefficient Ug can be evaluated:
Re Pr
q
gm n
W
TNu A
T
(3.59)
Nu being the non-dimensional group of Nusselt, , Re being Reynolds number, Pr
being Prandtl number, Tg being the gas temperature, TW being the wall temperature
and A, m, b, q coefficients depending on the phenomena. By the adoption of
different value of these coefficients also internal convection heat transfer
coefficient Uc0 can be evaluated.
The various heat layers can be seen as a thermal equivalent circuit and schematically the heat
transfer phenomena previously described can be represented as a series of thermal resistance
as shown in the figure 3.25:
Fig. 3.25: Simplified view of the thermal resistance for a generic blade
Pag. 102 of 202
To improve the performance of the blade in terms of life consumption rate, is desirable to
increase the outer thermal resistance (reduce the external heat transfer coefficient), to reduce
the inner thermal resistance (increase the internal heat transfer coefficient) and to adopt a
thermal barrier coating layer characterized by a high thermal resistance (low conductivity).
Various techniques are employed both on cold side (jet impingement, turbulence promoters,
etc.) and on the hot side (film cooling) to remove the heat from the blade.
On the coolant flow side, adoption of some architectural devices as the turbulence promoter,
the rib arrangement, the pin fins and of jet impingement technique leads to increase the
internal heat transfer coefficient. Each enhancing system can be seen as a corrective
coefficient fjk greater than 1 of the equivalent smooth heat transfer coefficient Uc0.
Accordingly, in figure 3.26 temperature profile modification on the coolant side owing to the
enhancing system of the heat transfer coefficient is given:
jet impingement fji >1
turbulence promoter ftp >1
rib arrangement ftp >1
pin fins fpf >1
Fig 3.26: Schematic view of the enhance system of the internal heat transfer coefficient
Turbulence promoter are widely employed in Heavy Duty Gas Turbine inner channels in
order to enhance the internal heat transfer coefficient. Taking ribs configuration according to
Data Base (figure 3.27-a) into consideration, in figure 3.27-b the increase of heat transfer
coefficient is shown.
Pag. 103 of 202
Fig. 3.27 a-b: a) rib distribution – b) Influence of Turbulent promoter on the NU number
Adoption of impingement concepts leads to enhance the internal heat transfer coefficient. The
overall increase of the Nusselt number depends on many architectural and geometrical
parameters taken from Data Base and from the HDGT State of the Art. Nusselt non-
dimensional group versus some architectural ratios is shown in figure 3.28:
Fig 3.28 : Influence of jet impingement architecture on internal heat transfer coefficient
Pag. 104 of 202
Heat flux coming from the blade layers is mitigated by the coolant mass flow taken from
compressor. Internal heat transfer coefficient 0cU is evaluated taking internal diameter,
velocity, mass flow, viscosity, etc. into account. Internal devices, suitably arranged (pins, rib,
etc.) as well as the jet impingement are designed to enhance internal heat transfer coefficient.
Coolant mass flow passes through multi-pass channel, increasing the effective surface of the
heat transfer, before exiting from the blade and mixing with the main hot gas stream. The
contribution of turbulence promoters, ribs arrangement, pin fins and jet impingement are
taken into account by expressing the coolant heat transfer coefficient:
0c c Tp ji ra pfU U f f f f (3.61)
On the other side, the hot one, introduction of techniques to reduce the external heat transfer
coefficient are taken into consideration. The adoption of film cooling allows to depress the
hot gas heat transfer coefficients (Ug0) by the correction of a ffilm coefficient, lower than 1,
because of the cold insulating layer between the hot gas stream and the wall of the blade.
Accordingly, film cooling can be also seen as an additional thermal resistance layer
characterized by an equivalent thickness and thermal conductivity. In figure 3.29, temperature
profile with and without film cooling is depicted:
Fig 3.29: Schematic view of the depression of the external heat transfer coefficient
owing to the film cooling
Pag. 105 of 202
Main stream prevalent convection heat transfer coefficient 0gU is related to some parameters
such as velocity, efflux area, conductivity, viscosity, etc. External heat transfer coefficient
assumes different values for different points among the blade profile as shown in fig.3.30. By
the adoption of lumped model hot gas heat transfer coefficients have been evaluated for the
various blade rows. When the film cooling occurred, external heat transfer coefficient is
depressed by the coolant mass flow exiting from the blade row holes realizing a thin cold film
that protects the blade.
Fig. 3.30: Typical heat transfer distribution among the blade row surface
Heat transfer coefficient distribution on pressure and suction side and film cooling influence
on the phenomena are shown in figure 3.31:
Fig. 3.31: External heat transfer coefficient depressed by the film cooling
Pag. 106 of 202
The hot gas heat transfer coefficient can so be expressed:
0g g filmU U f (3.60)
1filmf being the film cooling coefficient.
Also BM and TBC layer influences the heat transfer process. Bulk Material and Thermal
Barrier Coating thermal resistances are evaluated taking the thickness sj and the thermal
conductivity of the layer into account. Changing the thickness of the TBC layer and the
TBC material composition, the coolant mass flows required to maintain the same ratio of life
consumption change significantly. As an example, in figure 3.32 modification of the coolant
flows versus the TBC thickness is presented:
Fig. 3.32: Influence of the Thickness TBC layer on the coolant flows
3.2.5.2.1 Cooling Effectiveness
Combining the various heat transfer processes (phenomena) together the overall GT blade
temperature profile is sketched in figure 3.33.
From the technology point of view a global relationship exists between the characteristic
temperatures of cooling phenomena and cooling effectiveness. For each blade row the cooling
effectiveness can be expressed:
g W
c
g c
T T
T T
(3.61)
Pag. 107 of 202
Fig. 3.33: Temperature profile along the various blade layers
Such a cooling effectiveness is an empirical result and is an empirically established
relationship among architecture, geometry of the coolant system (platform, blade, shroud,
etc.) thermic and thermal barrier, bulk material as well as main stream and coolant parameter
relevant for the heat transfer process. It is a results of coupling, of a coolant stream and blade
seen as an heat transfer device and of the outer stream.
3.2.5.2.2 Effectiveness – Number of heat Transfer Unit
To establish a global relation to express cooling effectiveness c in terms of characteristic
quantities of the overall phenomena, such as coolant and hot gas mass flows, architectural and
geometric parameters as well as heat transfer coefficients, the problem can be addressed by
adopting the Effectiveness VS Number of heat Transfer Unit NTU approach.
Effectiveness represents the effective heat Q that can be exchanged versus the heat Q that
could be hypothetically exchanged by a heat transfer device of infinite surface (3.62):
Q
Q
(3.62)
The Number of heat Transfer Unit is expressed by the relation (3.63):
p
U SNTU
c m
(3.63)
Pag. 108 of 202
U being the heat transfer coefficient, S the characteristic Surface of phenomena, pc the
specific heat of the fluid and m the mass flow.
Depending on geometry, holes arrangement, streams directions (equicurrent, countercurrent)
and so on, the most adequate formulation that relates to NTU can be adopted, taking Data
Base and the State of the Art into consideration [10, 15]:
( , , ,....., , , )g c gf m m U NTU geometry architecture (3.64)
According to nomenclature of figure 3.33 cooling effectiveness c can be evaluated as a
combination of effectiveness related to elementary heat transfer processes taking film cooling,
impingement, conduction and all aspects into consideration. Moreover, evaluation procedure
of cooling effectiveness c has been performed taking relationship of counter current heat
exchange from Data Base into account. Accordingly, the following heat transfer process have
been described:
Hot gas – Thermal Barrier effectiveness
1
1 1g NTU
g TB
Te
T T
(3.65)
1
1
g
g pg
U SNTU
m c
(3.66)
0g g FilmU U f (3.67)
gU being the heat transfer coefficient of the hot stream corrected by film cooling
coefficient (if film cooling is adopted) depressing the hot gas heat transfer coefficient
0gU . Filmf is lower than 1.0.
In the various gas expander row, hot gas stream reduces its temperature both because
of the expansion (uncooled) and because of the injection of the coolant flows into the
main stream. The latter aspect lead to a temperature differencegT strictly connected
to the heat transfer process. A schematic equivalent representation, not to scale, of the
cooling effect on the gas side is given in figure 3.34:
Pag. 109 of 202
Fig. 3.34: schematically main stream temperature decrease – Not to scale
Hot Gas – Bulk Material
2
2 1g NTU
g W
Te
T T
(3.68)
2
2
g
g pg
U SNTU
m c
(3.69)
2
1
1 TB
g TB
Us
U
(3.70)
2U being the heat transfer coefficient taking convection of the main stream and conduction of
the TB layer into consideration.
Hot Gas – Coolant
In this case two different fluids take part at the heat transfer phenomena. According to Data
Base, expression of effectiveness is different from the (3.65) and (3.69) because Thermal
Capacity Ratio TCR must be considered (3.71):
c pc
g pg
m c
m c
(3.71)
Coolant stream is the lower heat thermal capacity fluid that must be put at the dominator of
NTU expression:
Pag. 110 of 202
3
3
(1 )
3 (1 )
1
1
NTU
Co Ci
NTU
g Ci
T T e
T T e
(3.72)
33
i
c pc
U SNTU
m c
(3.73)
3
1
1 1 1c cTB
g g TB g BM c
US Ss
U S S U U
(3.74)
0c C TP IU U f f being the internal coolant heat transfer coefficient corrected by enhancing
coefficient related to turbulence promoter and impingement effect, respectively.
BMU being the heat transfer coefficient of the bulk material, TB
TB
s
being the heat transfer
coefficient trough the thermal barrier.
In this simple application, expressing cooling effectiveness as (3.75):
g W
c
g c
T T
T T
(3.75)
and combining effectiveness of sub-process (3.71) and (3.75):
3
2 ( )
g W
g c c co
T T Tg
T T T T
(3.76)
Substituting (3.75) into (3.76), cooling effectiveness is expressed in terms of mass flows,
architectural and geometrical parameters (3.78) taking conservation of energy into account
(3.77):
( )c pc c co g pg gm c T T m c T (3.77)
3
2
c pc
c
g pg
m c
m c
(3.78)
Finally combining (3.78) with (3.68), (3.71) and (3.72) analytic expression, for a really simple
case, of cooling effectiveness c is obtained and given as rule (3.79):
3
3
2
(1 )
(1 )
1
1
1
NTU
NTU
c NTU
e
e
e
(3.79)
Pag. 111 of 202
Such an effectiveness depends on many parameters and empirically known aspects:
( , , , , , , , , , , ....)c g pg c pc i o j jf m c m c U U s arch geo etc (3./0)
Accordingly, GT Cooling model based on lumped performance features includes all the
aspects previously described (airfoil, platform, sidewall cooling and others). The best fit
relation to establish the cooling effectiveness (taking architecture, technology, flow feature,
etc. into account) can be described by the following equation:
2
1
k
cj k e (3.81)
k1 and k2 being coefficients with a suitable value for cooling modern technologies and χ being
the thermal capacity ratio between the coolant stream and the hot gas. In figure 3.35 the
cooling curves given the TCR vs cooling effectiveness are presented. These curves refer to the
cooling of the foil (not taking the discs, shroud, etc. cooling into account).
Degradation phenomena (fouling, corrosion, erosion, etc.) influence pressure losses and heat
transfer coefficients. Furthermore, at part-load, the mass flows change. Such aspects influence
continuously the expander cooling and they are taken into account off-design effectiveness-
TCR. For each Stator Vane and Rotor Blade a peculiar relation can describe the off-design
behavior of the cooling system.
4
3( )k
c k e (3.82)
k3 and k4 being coefficients taking the variation of the TCR and of the impingement and film
cooling into account.
Fig. 3.35: RO3 Cooling Design Curve – Stator Row and Rotor Row
Pag. 112 of 202
3.3 Steam Cycle Component Models
According to H2-IGCC power island layout, given in figure 3.36, steam cycle sizing and off-
design component models representing the real plant set-up have been developed and adapted.
In the following paragraphs description of quantities involved in the sizing and off-design
calculations is given.
Fig. 3.36: Sketch of H2-IGCC Steam Cycle
Pag. 113 of 202
3.3.1 Heat Transfer Devices
Surface heat transfer devices have been modeled considering two aspects. The heat transfer
from the hot stream to the cold one, through the tube surfaces, and the pressure losses across
the device both for the hot stream both for the cold stream. Suitable equipment’s to take
advantage of this kind of heat transfer phenomena are heterogeneous systems constituted by
three spatially distinct subsystems: cold fluid, tube walls and hot fluid, interacting through
boundary surface. A representative scheme of the above is given in figure 3.37:
Fig. 3.37: Heat Transfer Device scheme
In the plot of figure 3.37, the orifices is representative of the pressure losses while the arrow
on the boundary surface is representative of the heat transfer phenomena. By adopting this
scheme, heat transfer devices of the Heat Recovery Steam Generator such as Economizer
(ECO), Evaporator (EVA) and Super-heater (SH) have been modeled.
Sizing and off-design model have been developed for each device. Input data for the sizing
problem have been assumed according to cycle calculation results and to the present state of
the art of the 3 pressure levels HRSG. Results of sizing process are stored and data are used as
input for the off-design analysis.
For a generic heat transfer devices the following quantities are inlet and outlet variables for
the hot (gas) and the cold (water/steam) side, respectively:
Pag. 114 of 202
Gas Side
mhi [kg/s] Inlet gas mass flow
Thi [°C] Inlet gas temperature
phi [kPa] Inlet Pressure
Tho [°C] Outlet gas temperature
pho [kPa] Outlet gas pressure
Water/Steam Side
mci [kg/s] Inlet water/steam mass flow
Tci [°C] Inlet water/steam temperature
pci [kPa] Inlet water/steam pressure
hci [kJ/kg] Inlet water/steam enthalpy
Tco [°C] Outlet water/steam temperature
pco [kPa] Outlet water/steam pressure
hco [kPa] Outlet water/steam enthalpy
Both for sizing both for off-design process some equations describe the heat transfer devices
behavior taking heat transfer phenomena as well as pressure losses into account. Introducing
moreover auxiliary and constitutive equations each aspects can be described.
Mass conservation
1( , ) 0hi hof m m (3.83)
2( , ) 0ci cof m m (3.84)
Energy conservation
3( , , , ) 0h hi hi hof Q m h h (3.85)
4( , , , ) 0c ci ci cof Q m h h (3.86)
Auxiliary & Constitutive Equations
Heat Transfer
5( , , ) 0c pc cif C c m (3.87)
6( , , ) 0h ph hif C c m (3.88)
7 min( , , ) 0c hf C C C (3.89)
8 max( , , ) 0c hf C C C (3.90)
Pag. 115 of 202
9 min max( , , ) 0f C C (3.91)
10 min( , , , , ) 0gi cif Q C T T (3.92)
11( , , , , ) 0f NTU architecture geometry (3.93)
12 min( , , , ) 0f NTU U S C (3.94)
13( , , ) 0h cf U U U (3.95)
Terms of equation 3.95 have been established by the adoption of the most adequate
correlation concerning the 3 pressure levels HRSG finned tube banks [10, 15]. Many
parameters such as the maximum available speed, the arrangement of the tubes (in-line or
staggered) and the number of the tubes rows influence the heat transfer coefficient on the gas
side. Moreover, the tube banks type (i.e. super-heater, economizer and boiler) influenced
practically the internal heat transfer coefficients because of many aspects such as the status of
the fluid (steam and water) and the different boiling conditions (nucleate boiling, film boiling,
etc.) steps into the evaporator tubes. Looking at the external heat transfer coefficients, in
Forced Convection for Tube bundle in cross flow, figure 3.38 is exhaustive [15].
Fig. 3.38: Sketch illustrating nomenclature for in-line tube arrangements [15]
In figure 3.39 the trend of Nusselt number versus the flow condition expressed by the
Reynold number is reported for in-line tube arrangement. Similar trend can be observed also
for staggered tube arrangements [15].
Pag. 116 of 202
Fig. 3.39: NU vs Re max for in-line tube arrangement [15]
Influence of the transverse rows on the heat transfer coefficients has been taken from [15] into
account. In figure 3.40, the correction coefficient taking the number of the rows into
consideration is given.
Fig.3.40: Correction Factor to account the number of the Row [15]
Accordingly, various correlation are presented in the SoA to evaluate the heat transfer
coefficients on the external side (hot gas side) of finned tube banks of HRSG. Such
coefficients Uh has been evaluated by means of the rule 3.96.
( ,Re,Pr, , , , , ) 0h h al Row Jf U D n k (3.96)
λh, ɛal, kJ, nRow, D being the thermal conductivity of the hot gas, the ratio of total surface area
with fins to the bare tube surface area without fins, the coefficients j-th of the empirical
correlation, the number of the transversal rows and the characteristic dimension, respectively.
Pag. 117 of 202
On the other side, the cold one (steam or water), the heat transfer coefficients have been
established by considering both for super-heaters and for economizers forced convection
inside tubes. Accordingly, Uc has been evaluated by the rule 3.97.
( ,Re,Pr, , , ) 0c c Jcf U D k (3.97)
λc, and D being the thermal conductivity of the cold side fluid and the characteristic
dimension, respectively.
In case of evaporator tube bundles, empirical results taken from the Available Technologies
of the State of the Art have been adopted to perform the heat transfer coefficient evaluation
[15]. Heat flux has been taken into account as well as the steam quality and the temperature
difference between the tubes wall and the fluid. In figure 3.41, trend of heat flux coefficient
versus such temperature difference is given.
Fig.3.41: Heat Flux VS Temperature Difference
Accordingly, in boiling process two contribution can be defined by following [19].
Convection and boiling heat transfer mechanism have been taken into account.
Pag. 118 of 202
Pressure Losses
14( , , , , ) 0hi h ho h pgf p p p f (3.98)
15( , , , , ) 0ci c co c pcf p p p f (3.99)
16( , , , ) 0h hi hi hif m T p (3.100)
17 ( , , , ) 0c ci ci cif m T p (3.101)
Also pressure losses have been evaluated taking the most adequate correlation from the State
of the Art into consideration, both on hot side and on cold side [15]. Equation 3.99 assumes a
different formulation when water is considered instead of steam.
18( , , ) 0hi hi hif h T p (3.102)
19( , , ) 0ho ho hof h T p (3.103)
20( , , ) 0ci ci cif h T p (3.104)
21( , , ) 0co co cof h T p (3.105)
22( , , , , ) 0ph hi hi ho hof c T p T p (3.105)
23( , , , , ) 0pc ci ci co cof c T p T p (3.106)
The above equations describe phenomena interesting the heat transfer devices. During the off-
design analysis, the hot and cold heat transfer coefficients are evaluated by a relation between
reference (*) and actual conditions. Exponent (a,b,c,d) of the relations (3.107) and (3.108)
have been chosen according to heat transfer devices Data Base [20] and assumes different
values if they refer to the hot side or to the cold side (table 3.1)
* * * * *
24( , , , , , , , , , , , , , ) 0h h h h ph ph h h h hf U U m m c c a b c d (3.107)
* * * * *
25( , , , , , , , , , , , , , ) 0c c c c pc pc c c c cf U U m m c c a b c d (3.108)
, being the viscosity and thermal conductivity, respectively. Coefficients a,b,c,d are given
in table 3.1:
Pag. 119 of 202
Table 3.1: coefficient exponents of the heat transfer coefficient calculation
gas Water/steam
a 0.6 0.8
b -0.27 -0.47
c 0.67 0.67
d 0.33 0.33
In case of an evaporator heat transfer devices relation (3.108) assumes a different formulation:
* * *
26( , , , , , ) 0c cf U U T T (3.109)
φ being the thermal heat flux. In the model the critical flux is taken into consideration.
3.3.2 Condenser
Heat transfer device model described in paragraph 3.3.1 has been adapted to the specification
of a condenser, adopting the best relations to account the heat transfer coefficient and pressure
loss calculations on the two sides [15,21].
A multi-zone modeling approach has been adopted and two heat transfer device have been
included in the condenser model. Such a modelling formulation allows to takes the wide
variability of the plant operating conditions into account and the relative modification of the
steam quality at the inlet of the condenser. Accordingly, condenser can be fed by superheated
or saturated steam depending on plant operations. Such a scheme is given in figure 3.42.
Fig. 3.42: Multi-Zone Condenser
Pag. 120 of 202
The model is structured to handle de-superheating and condensing zones of variable surfaces
where the above phenomena occur. The extension of the zones changes according to
operations. Of course the sum of surfaces of the various zones must be equal to the overall
heat transfer area established during the sizing procedure.
3.3.3 Steam Turbine
Two options are available to model Steam Turbine (ST). The first one consists in curves
connecting steam consumption, power production and steam extractions (Willan’s Curves).
According to the second option, the ST model is built up by using sub-models describing
groups of non-controlled or controlled stages.
The model refers to a group of stages of a steam turbine. At inlet and outlet stations of the
stage, admission and extraction of fluid has been taken into consideration as the following
figure 3.43 points out:
Fig. 3.43: Scheme of a generic steam expander
Taken the inlet and outlet section into account, ST variables have been defined:
Inlet
mi [kg/s] Inlet steam mass flow
Tvi [°C] Inlet steam temperature
pvi [kPa] Inlet steam pressure
hvi [kJ/kg] Inlet steam enthalpy
n [rpm] Shaft Speed
Pag. 121 of 202
Outlet
mo [kg/s] Outlet steam mass flow
Tvo [°C] Outlet steam temperature
pvo [kPa] Outlet steam pressure
hvo [kJ/kg] Outlet steam enthalpy
P [MW] Power
The model considers the mass flow constant through the stage. Therefore the conservation
low of mass is satisfied as the following relation indicates:
1( , ) 0i of m m (3.110)
The conservation of energy is expressed as follows:
2( , , , ) 0i o lossf h h W W (3.111)
Wloss being the mechanical losses.
Inlet steam enthalpy hvi and outlet steam enthalpy hvo can be evaluated by constitutive
equations:
3( , , ) 0wi wi wif h T p (3.112)
4( , , ) 0vi vi vif h T p (3.113)
Such groups operate at constant steam mass flow (i.e. inlet mass flow is equal to that at the
exit). Thus the model of a ST with intermediate steam admissions or extractions can be
arranged by using the above constant mass flow sub-models followed by nodes where steam
is added or extracted. Sketch of steam turbine is given in figure 3.44.
Fig. 3.44: Stodola Ellipse Sketch and steam turbine body with governing valve
Pag. 122 of 202
For each group of stages an equivalent sizing aimed at evaluating some reference quantities
required for off-design calculations is performed according to [8]. Input quantities are
reference inlet pressure, temperature, mass flow, steam quality and discharge pressure.
Non controlled stages off-design behaviour is described by adopting a modified Stodola
ellipse law [9]. The efficiency of the group of stages is evaluated as a function of the actual
mass flow and expansion pressure ratio by using correlation curves stored in the model DB.
The above curves refer to different kinds of real machines. The most suitable ones can be
chosen according to the features of the turbine to be modelled and scaled on the basis of the
reference quantities. Suitable RFs allow the model to fit the real component behaviour at
N&C conditions. Actuality Functions AFs affecting flow functions and efficiency are
introduced in the off-design model.
In case of controlled stage groups, such groups are modelled by adding a partial admission
controlling stage in front of a non-controlled group of stages. The model gives the controlling
stage efficiency and mass flow as a function of the degree of partialization, inlet pressure and
temperature and expansion pressure ratio. Such correlation have been represented by maps
expressing relation between pressure ratio , steam mass flow m , partialization ratio and
adiabatic efficiency have been adopted:
1( , , ) 0F m (3.114)
1( , , ) 0F m (3.115)
The maps are provided in non dimensional form and are scaled in relation to the referenced
quantities * * *, ,m .
As for the non-controlled stage groups, a set of correlations concerning different kinds and
sizes of partial admission stages is stored in the model DB. The selection of the most suitable
set can be made on the basis of reference stage data.
Pag. 123 of 202
3.3.4 Deaerator
In the H2-IGCC plant a deaerator has been adopted to remove gases from the water stream. A
steam mass fraction extracted from overall steam mass flow addressed to the Low Pressure
Steam Turbine. Component model of this component has been properly adapted to the plant
specifications. Deaerator inlet and outlet quantities are described taking the scheme given in
figure 3.45 into consideration:
Fig. 3.45: Deaerator scheme
Inlet
mwi [kg/s] Inlet water mass flow
Twi [°C] Inlet water temperature
pwi [kPa] Inlet water pressure
hwi [kJ/kg] Inlet water enthalpy
mvi [kg/s] Inlet steam mass flow
Tvi [°C] Inlet steam temperature
pvi [kPa] Inlet steam pressure
hvi [kJ/kg] Inlet steam enthalpy
Outlet
mwo [kg/s] Outlet water mass flow
Two [°C] Outlet water temperature
pwo [kPa] Outlet water pressure
hwo [kJ/kg] Outlet water enthalpy
Pag. 124 of 202
Deaerator component model is based as a water/steam mixer with the water continuously
maintained at the thermodynamic equilibrium. Accordingly, mass and energy conservation
equations describing its behavior are given:
1( , , ) 0wo wi vif m m m (3.116)
2( , , , , , ) 0wo wi vi wo wi vif m m m h h h (3.117)
3( , ) 0vi wif p p (3.118)
4( , ) 0vi wof p p (3.119)
The system leads to equilibrium, thus adopting constitutive equation of the fluids, the outlet
enthalpy hwo can be established taking saturation enthalpy in relation to the deaerator
pressure into account:
5( , ( )) 0wo sat vif h h p (3.120)
6( , ( )) 0wo sat vif T T p (3.121)
Inlet water enthalpy hwi and steam enthalpy hvi can be evaluated with constitutive equations:
7 ( , , ) 0wi wi wif h T p (3.122)
8( , , ) 0vi vi vif h T p (3.123)
Deaerator component model is characterized by some constraints and by some inequalities to
ensure that the physical aspects of the phenomena will be taken into consideration.
3.3.5 Pump
Pump component model has been developed on the basis of maps, representing the kinematic
similitude. Sketch of main variables involved in the pump component model is given in figure
3.46.
Fig. 3.46: scheme of a generic pump
Pag. 125 of 202
inlet
mwi [kg/s] Inlet water mass flow
Twi [°C] Inlet water temperature
pwi [kPa] Inlet water pressure
hwi [kJ/kg] Inlet water enthalpy
n [rpm] Rotational Speed
P [MW] Power
outlet
mwo [kg/s] Outlet water mass flow
Two [°C] Outlet water temperature
pwo [kPa] Outlet water pressure
hwo [kJ/kg] Outlet water enthalpy
The model is able to reproduce a fixed rotational speed pumps and variable speed ones.
Power consumption of the machine is expressed by the following relation:
wmi
m
m pP
(3.124)
p being the pressure increase between the inlet and the outlet section, m being the average
density of the work fluid and being the efficiency.
The enthalpy difference across the machine is evaluated as follows:
mio i
w
Ph h
m (3.125)
In the cycle calculation the efficiency value is set by the user.
For the off design operating conditions, two different cases have to be taken into account. One
considering n cost and the other considering n cost . In both cases the model takes the
characteristic curves that give the head and the efficiency versus the mass flow and velocity
into account. In figure 3.47 such maps are schematically represented.
Pag. 126 of 202
Fig. 3.47: Pumps Characteristic Non-Dimensional Curves
Referring to a constant rotational speed of the shaft, these curves are expressed as the product
of the nominal value for the normalized value related to it. Expressing the normalized mass
flow0
m
m , the head and the efficiency , they are expressed by the following relations:
0
1( )f (3.126)
0
2( )f (3.127)
0( ) being representative of the nominal conditions, 1 2( ), ( )f f being representative of the
normalized curves previously described.
In the case of variable rotational speed, the curves depend on the normalized velocity defined
as follows:
0
n
n (3.128)
and are expressed as a function of the both variables 1 2( , ), ( , )f f . These functions are
tabulated for different normalized speed and during the calculations the most adapted curve is
selected. Thus, operating in kinematic similarity, efficiency and head1 curves are determined
vs the actual mass flow and actual velocity.
Pag. 127 of 202
3.3.6 Pressure Loss Devices
The model is used to describe the devices that establish a pressure difference between two
domains. Models for devices operating with gas and water flows are taken into account.
The fluid evolves according to an isenthalpic transformation. Inlet pressure pi and outlet
pressure pu are correlated as follows:
(1 )o i pp p k (3.129)
pk being loss coefficient.
In the cycle calculation, the previous coefficient can be either assigned or not. Two different
cases can be distinguished: device with fixed or variable opening.
In the first case, the flow rate is correlated to the pressures at the extremities of the device as
follows for the gas:
2 i o
i
p pk
p
(3.130)
being the corrected inlet flow rate and k being a constant obtained with respect to reference
values:0, pi
0 e pu
0. In case of water, instead:
2
i oQ k p p (3.131)
Q being the volumetric flow rate k being a constant obtained with respect to reference values
In the second case, device opening is automatically adapted in order to maintain the
controlled variables (pressure or flow rate) at the assigned values. Different expression taken
from the SoA can be adopted to established the opening of the device under the design
conditions.
f (kv, pi, po, Ti, To, mi, kind of fluid)=0 (3.131-a)
3.3.7 Junctions
Economizer, Evaporator, Super – Heater, Pumps, Deaerator, Desuper – heater and other plant
components have been connected together by the adoption of dummy Junctions. Junctions can
be divided into mixer, in which two or more streams mix in a only one stream exiting the
device, and into splitter where one stream enters the device and two or more streams exit.
Pag. 128 of 202
3.3.7.1 Water/Steam Mixer
Water/Steam Mixer component model is addressed to calculate mixing of some water and
steam flows. The model considering two inlet quantities and one outlet quantities is described.
Mixer inlet and outlet quantities are described taking the scheme given in figure 3.48 into
consideration:
Fig 3.48: Mixer scheme
Inlet
m1 [kg/s] Inlet 1st stream mass flow
h1 [kJ/kg] Inlet 1st stream enthalpy
m2 [kg/s] Inlet 2nd
stream mass flow
h2 [kJ/kg] Inlet 2nd
stream enthalpy
Outlet
m3 [kg/s] Outlet mass flow
h3 [kJ/kg] Outlet enthalpy
Mixer component model is based on mass and energy. Equations describing its behavior are
given:
1 1 2 3( , , ) 0f m m m (3.132)
2 1 2 3 1 2 3( , , , , , ) 0f m m m h h h (3.133)
MIXW component model is characterized by some constraints to ensure that the physical
aspects of the phenomena will be taken into consideration.
3.3.7.1 Gas Mixer
Gas Mixer component model is addressed to calculate mixing of two gas flows of the same
mass composition. For the further development the possibility of mixing different gas species
will be taken into consideration. Mixer inlet and outlet quantities are described taking the
scheme given in figure 3.49 into account:
Pag. 129 of 202
Fig 3.49: Gas Mixer scheme
Inlet
m1 [kg/s] Inlet 1st stream mass flow
p1 [kPa] Inlet 1st stream pressure
T1 [°C] Inlet 1st stream temperature
m2 [kg/s] Inlet 2nd
stream mass flow
p2 [kPa] Inlet 2nd
stream pressure
T2 [°C] Inlet 2nd
stream temperature
Outlet
m3 [kg/s] Outlet mass flow
p3 [kJ/kg] Outlet pressure
T3 [kg/s] Outlet temperature
The model allows, by the adoption of orifices, to adapt the highest pressure of the two gas
streams entering the component model, to the lowest one. Mixing takes place at this lowest
level of pressure. Outlet pressure is thus assumed as the mixing pressure value.
Gas Mixer component model is based on mass, energy, momentum and entropy conservation
and equations describing its behavior are given:
1 1 2 3( , , ) 0f m m m (3.134)
2 1 2 3 1 2 3( , , , , , ) 0f m m m h h h (3.135)
3 1 2 3( , , ) 0f p p p (3.136)
Pag. 130 of 202
Inlet and outlet enthalpy hj of the respective streams are evaluated by adopting constitutive
equations:
4 1 1 1( , , ,[ ]) 0f h T p xx (3.137)
5 2 2 2( , , ,[ ]) 0f h T p xx (3.138)
6 6 6 6( , , ,[ ]) 0f h T p xx (3.139)
MIXER GAS component model is characterized by some constraints to ensure that the
physical aspects of the phenomena will be taken into consideration
3.3.8 Splitter
The component model has been developed to take ramification of gas, water as well as water
flow into consideration. Scheme of the splitter is given in figure 3.50
Fig. 3.50: Splitter Scheme
Splitter inlet and outlet quantities are given referring to the scheme of 3.50
Inlet
m1 [kg/s] Inlet mass flow
Outlet
m2 [kg/s] Outlet 1st stream mass flow
m3 [kg/s] Outlet 2nd
stream mass flow
Model is based on conservation low of mass:
1 1 2 3( , , ) 0f m m m (3.140)
Moreover is possible to assign a desired fraction φ of the two outlet mas flows. In this case:
2 1 2 3( , , , ) 0f m m m (3.141)
Pag. 131 of 202
3.4 Gasification Island Simulator
Gasification Island block scheme is given in figure 3.51. Various interaction between sections
are schematically represented. The gasification model is focused on the description of
phenomena and processes giving a significant contribution to the electric power production.
Accordingly, less relevant aspects are neglected or treated in a simplified way.
Fig. .3.51: Gasification Island Block Scheme
According with figure 3.51, the gasification model is sub-divided in macro blocks
representing the Air Separation Unit (ASU) Section, the Gasification (GASIF.), the Water Gas
Shift (WGS)and the Carbon Capture and Storage (CCS), respectively. Concerning the ASU
section, it has been considered to take the compressor power requirements into consideration.
In the following paragraphs a description of the macro blocks is given.
3.4.1 Gasification Block
In the Gasification Block two sub-components have been considered. The first one is the
gasifier reactor, the second one is the syngas cooler. The gasification zone is modeled taking a
Chemical Model (CM) and a Heat Transfer Model (HTM) into consideration. Scheme of the
gasification reactor is given in figure 3.52.
Pag. 132 of 202
Fig. 3.52: Gasifier Reactor Model Scheme
Syngas Composition (Xsgi) and mass flow (msgi) are evaluated by the CM section and thermal
power transferred (QSG) to the liquid slag layer is established by the HTM section. Such
quantities are given as input for the gasification zone model. Mass, Energy, Momentum and
Entropy conservation equations have been adopted to model the various components.
The composition of syngas is calculated by imposing the mass balance to the chemical
species constituting the syngas itself, which is assumed formed by CO, CO2, H2, N2 and
H20. Other component such as sulfur compounds (H2S and COS), methane, HCN etc. are
neglected. Chemical reactions are assumed at equilibrium.
On the basis of the compositions and mass flows of streams entering the CM, the syngas
composition is evaluated. Syngas composition exiting the node itself is determined by the
element balances (C, N, H, O) and by imposing the chemical equilibrium of water shift
reaction, evaluated at the temperature of the gas inside the volume.
CO2 + H2 CO +H2O (3.141)
The produced syngas mass flow is calculated by applying the conservation of mass
(3.142)
mash being known from coal input composition.
GZ
MSG, TSG, pSG, XSG
mc, Tc, Xc
mH20 TH2O
mO2 TO2 XO2
mO2 XO2 mH20 mc
CM
TSG mSG
XSG
mASH
mSG0, TSG, PSG
TSG, mSLAG
mSLAG
QMW
HTM
QSG
Pag. 133 of 202
The gasification process is exothermal, thus to achieve the desired syngas temperature some
heat has to be removed. The gasification volume is surrounded by membrane tube walls
where IP saturated steam is produced. Such a membrane tube walls is protected by a
refractory liner which during operations is covered by a slag layer. Heat is transferred from
the syngas to the slag layer by radiation and convection and then to the water wall for
conduction trough slag layers, the refractory and the tube walls. The HTM, given the syngas
temperature Tsg evaluates the thermal power transferred to the water wall Qww .Thermal power
transferred from syngas to the slag Qsg is evaluated as follows:
(
) ( ) (3.143)
being Tsg and Tsl the temperature of the syngas and the temperature of the liquid slag layer
contacting the syngas respectively, A the heat transfer surface, the Boltzmann constant and
the syngas emissivity. Qsg is transferred to the gasification zone model as a source term in
the energy conservation equation. The raw syngas at the exiting the gasifier at some 1600 °C
is cooled to 900°C by adding a stream of recycled cold syngas before entering the syngas
cooler where High Pressure (HP) and Intermediate Pressure (IP) steam is produced. Finally
the cooled raw syngas passes through a wet scrubber where the water soluble species are
removed together with the particulate matters. The model is constituted by a mixing node
(syngas quenching), a heat transfer section (syngas cooling) and a saturator node (scrubber) as
shown in Fig. 3.53.
Fig. 3.53: Syngas Cooler Model Scheme
msg1, hsg1
msg2, hsg2
msg4, hsg4
msg3, hsg3
msg5, hsg5
msg6, hsg6
mwin, hwin
mwex, hwex
Q
Pag. 134 of 202
The syngas quenching is simulated by a mixing node with two entering stream and an exiting
one. The node solves energy and mass conservation equations. The heat transfer apparatuses
constituting the syngas cooler (tube bundles ) are modelled by adopting the e-NTU approach.
Finally, the saturator node evaluates the mass flow, temperature, enthalpy and composition of
the exiting syngas given the features of the syngas and water inlet flows
3.4.2 Water Gas Shift Block
The section is modelled as a mixing node accounting for the steam addiction to the syngas,
two plug flow adiabatic reactors and heat transfer devices required to cool the syngas before
and after the second shift reactor (Fig. 3.54). The reference CO conversion rate is really high
(99%), therefore the chemical equilibrium is considered for the CO shift reactions in both the
reactors. The syngas hold up is taken into consideration in mass ans energy conservation in
both the reactors.. Heat transfer devices are modelled as described in the previous section.
Fig. 3.54: WGS Block Scheme
3.4.3 Carbon Capture and Sequestriation Block
The reference plant is designed for CCS, thus a sour syngas shift followed by a single AGR
plant section has been adopted. A Selexol based process has been selected for CO2 and H2S
removal. H2S and CO2 removal, although separate steps, are integrated in such a way that
solvent absorption and regeneration are combined to use one column for each operation.
According to the aim of the present analysis, focused on plant power production, the main
aspect taken into consideration is the H2 rich syngas availability along the time. The section is
simply modelled by a separation node where the CO2 is removed from the syngas flow.
The CO2 separation rate is assumed constant and equal to the design value. Pressure losses
inside the column are modelled by introducing a fixed opening orifice at the exit of the
storage volume.
.
ms, hs
msg1, hsg1, Xsg1 R1
Q
R2 HT
msg3, hsg3, Xsg3 msg4, hsg4, Xsg4 msg5, hsg5, Xsg5
msg2, hsg2, Xsg2
Pag. 135 of 202
3.5 Reference
[1] - Buckley T.J., Domalski E.S., 1988 ‘Evaluation of data on higher heating values and
elemental analysis for refuse-derived fuels’, Chemical Thermodynamics Division National
Bureau of Standards Gaithersburg, Maryland
[2] - Mott, R. A., and Spooner, C. E. "The Calorific Value of Carbon in Coal: The Dulong
Relationship." Fuel 19 (1940): 226-231, 242-251.
[3] - Annaratone D., 2008: ‘Generatori di Vapore’, Maggioli Editore, ISBN: 8838741492
[4] - Cerri G (2011): ‘Deliverable 4.2.2 – Description of the Model adapted or developed ad
hoc for the IGCC&CCS plants’.
[5] - Jonsson M., Bolland O., Bucker D., Rost M. (Siemens), 2005, ‘Gas Turbine Cooling
Model for Evaluation of Novel Cycles’. Proceedings of ECOS 2005, Trondheim, Norway,
June 20-22, 2005.
[6] - Final Report of the RTO Applied Vehicle Technology, 2007: ‘Performance Prediction
and Simulation of Gas Turbine Engine Operation for Aircraft, Marine, Vehicular, and Power
Generation’
[7] - SIEMENS AG, Siemens Gas Turbine SGT6-5000F, Answer for Energy, 2008
[8] - Baily, F. G., Cotton, K. C., Spencer, R. C., (1967): “Predicting the Performance of
Large Steam Turbine- Generators with Saturated and Low Superheat Steam Conditions”, 28
Annual Meeting of American Power Conference, Ger-2454-A.
[9] - Cooke D. H., (1985): “On Prediction of Off-Design Multistage Turbine Pressures by
Stodola’s Ellipse”, Transaction of the ASME, 596/ Vol. 107, July 1985
[10] - Rohsenow W. M, Hartnett J. P, Cho Y. I.,(1998): “Handbook of Heat Transfer – Third
Edition”, McCraw-Hill Handbooks
[11] - Torbidoni L., Horlock J.H., 2005, “Calculation of the expansion through a cooled gas
turbine stage” Asme Turbo Expo 2005, Reno-Tahoe, Nevada (USA), June 6-9, 2005;
[12] - Jonsson M., Bolland O., Bucker D., Rost M. (Siemens), 2005, ‘Gas Turbine Cooling
Model for Evaluation of Novel Cycles’. Proceedings of ECOS 2005, Trondheim, Norway,
June 20-22, 2005
[13] Han J.C., Dutta S., Ekkad S.V., Gas Turbine Heat Transfer and Cooling Technologiy,
Taylor and Franis, 2000
[14] - Boyce M.P., Gas Turbine Engineering Handbook 2nd
edition, Gulf Publishing
Company, 2002
[15] – Kreith F., Manglik R.M., Bohn M. S., 2011: ‘Principles of Heat Transfer – 7th
Edition’
Pag. 136 of 202
[16] Cohen H., Rogers G.F.C., Saravanamuttoo H.I.H., Gas Turbine Theory 3rd
edition,
Longman Scientific & Technical, 1987.
[17] Logan E., Roy R., Handbook of Turbomachinery 2nd
edition Revised and Expanded,
Marcel Dekker, 2003.
[18] Sanjay K., Singh O., Thermodynamic Evaluation of different gas turbine blade cooling
techniques, Thermal Issues in Emerging Technologies, ThETA 2, Cairo, Egypet, Dec 17-20th
2008.
[19] – Chen J. C. 1966.: “Correlation for Boiling Heat Transfer to Saturated Liquids in
Convective Flow,” Ind. Eng. Chem. Proc. Des. Dev., vol. 5, p. 332, 1966.
[20] – G. Cerri, G. Castiglione, A. Sorrenti, ”Model lo per l'analisi del potenziamento di
impianti a vapore con turbomotori a gas”. III Convegno Nazionale Gruppi Combinati
Prospettive Tecniche ed Economiche, Bologna, 23 maggio, 1989.
[21] - M. M. Chen, “An Analytical Study of Laminar Film Condensation,” part 1. “Flat
Plates,” and part 2, “Single and Multiple Horizontal Tubes,” Trans. ASME, Ser. C, vol. 83,
pp. 48–60, 1961.
[22] - Ebrahimi P., KArrabi H., Ghedami S., Barzegar H., Raoulipour S., Kebriyaie M., 2012:
“Thermodynamic modeling and Optimization of Cogeneration Heat and Power System Using
Evolutionary Algorithm (Genetic Algorithm)”, ASME TURBOEXPO 2010, July 14-18,
2010, Glasgow, USA.
[23] – Grigull U. et al., 1984: ‘Steam Tables in SI-Units / Wasserdampftafeln’, Springer-
Verlag Berlin Heidelberg 1984
[24] – Gas Turbine – Acceptance Tests, 2009, International Standard ISO 2314, pp. 54.
Pag. 137 of 202
Chapter IV
Gas Turbine and Steam Cycle Simulators
4.0 Introduction
Adoption of modelling approach and the development of component models for the various
H2-IGCC plant components that have been described in the chapters II and III, respectively,
leads to establish the simulator of the gas turbine and of the steam cycle. Taking the
methodological approach described in paragraph 2.3 into account, the various steps
concerning the cycle or process calculation, sizing, off-design and matching have been
performed for the various modelled components (i.e. compressor, expander, combustion
chamber, heat transfer devices, steam turbine, etc.). In the following paragraphs, results of the
various calculation steps are reported. Processes that take place in GT and SC have been
established by models that include also the working fluid properties. Accordingly, quantities
both for a gas mixture and for water/steam are presented.
4.1 Gas Turbine Component Simulators
Gas Turbine Equivalent Bryton Cycle, Cooling Overall mass flow, compressor, expander,
combustion chamber and cooling model calculation are reported. Moreover, matching of the
off-design maps of the component models leads to the development of the Generic 300MW F
Class Gas Turbine Simulator. Nominal Running point of such a machine as well as part load
analysis of the simulator when operating conditions change and different fuel feeding have
been reported. Accordingly, gas turbine behaviour when 33 MJ/kg H2-Rich Syngas is used as
fuel has been analysed.
4.1.1 Reference GT Brayton Cycle Evaluation and overall coolant flows
According to the paragraph 1.3.1 Ansaldo AE94.3A and Siemens SGT5 – 4000F GT’s have
been selected as reference ones. Taking Data available in the technical background of the
manufactures into consideration, the Gas Turbine cycles have been evaluated under the
Boundary Conditions and Data given in table 4.0a.
Brayton Cycle model has been used to perform the calculation of both Ansaldo and Siemens
Gas Turbine Cycle, evaluating polytropic efficiency for the compressor and for the expander
as well as the combustion efficiency. Mass flows of inlet compressor air, of fuel and of
exhaust are given in table 4.0b. Moreover relevant quantities estimated during the calculations
(temperature, pressure, work, etc.) are given in table 4.0b for the main station of the GT
Cycle.
Pag. 138 of 202
Table 4.0a: Input Data for Cycle Calculation
AE 94.3A SGT5-4000F
BOUNDARY CONDITIONS (b)
p1 [kPa] 101.3 101.3
T1 [°C] 15.0 15.0
[xx]1 [#]m dry air + RH60%
DATA (d)
[xx]f [#]m pure methane
LHV [kJ/kg] 50060 50060
P* [MW] 294 292
ηGT* [#] 0.397 0.397
mex* [kg/s] 702 692
Tex* [°C] 580 577
* [#] 18.2 18.2
∆pcc/p2 [#] 0.05 0.05
∆pe/p1 [#] 0.03 0.03
ηm [#] 0.998 0.998
ηge [#] 0.968 0.968
Table 4.0b: Cycle Mass Flows and Outlet Quantites
AE 94.3A SGT5 - 4000F
Mass Flow
mCi [kg/s] 687.2 677.3
mf [kg/s] 14.8 14.7
mex [kg/s] 702.0 1844.1
COMPRESSOR
T1 [°C] 15 15
T2 [°C] 409 409
p2 [kPa] 1844.1 1844.1
etapc [#] 0.929 0.928
WC [kJ/kg] 411 411
COMBUSTION CHAMBER
etacc [#] 0.99 0.99
AFR [#] 46 46
EXPANDER
T3 [°C] 1246 1246
p3 [kPa] 1751.9 1751.9
etape [#] 0.866 0.871
WE [kJ/kg] 854 858
GAS TURBINE
WTg [kJ/kg] 428 431
Pag. 139 of 202
Taking 3.2.1.5 paragraph into account, adoption of global models leads to establish the
overall coolant mass flows requirements. Thus, for the preliminary evaluation of the overall
coolant mass flow, the coolant temperature Tc has been assumed in the range of some 400-
500 °C and the blade temperature Tb in the range of 830-895 °C. Such temperatures are
defined arbitrarily as reference lumped temperatures of the GT global model. In table 4.0c,
evaluation of the overall coolant mass flow for the extreme values of the coolant temperature
and blade temperature is reported.
Table 4.0c: Evaluation of the overall coolant mass flow
for various coolant and blade temperature, respectively
By the assumption of some coefficients taken for the SoA, the overall coolant mass flow has
been calculated and by averaging these results, an overall value is of some 26% of the
compressor inlet mass flow.
26%cj
j
m of inlet compressor mass flow
Evaluation of the overall coolant mass flow given by the model leads to a value of that mass
flow that agrees with the coolant ratio (mcool/mcompr) found in the technical background [1,2]
4.1.2 Compressor
Sizing of the compressor made has led to obtain a certain value of temperature, pressure,
outlet mass flow (including the coolant flows both for the 1st Nozzle Row and for the 1
st Rotor
Row ) and power consumption assuming inlet air ISO condition. According with [3] in table
mc 165 kg/s
cpc 1.1 kJ/(kgK)
mg 523 kg/s
cpg 1.3 kJ/(kgK) mci 685 kg/s
Tcexit 400 °C mc/mci 24.1 %
Tf 1440 °C
Tb 830 °C
k1 0.1884 #
k2 1 #
mc 215 kg/s
cpc 1.1 kJ/(kgK)
mg 523 kg/s
cpg 1.3 kJ/(kgK) mci 685 kg/s
Tcexit 500 °C mc/mci 31.4 %
Tf 1440 °C
Tb 830 °C
k1 0.1884 #
k2 1 #
mc 129 kg/s
cpc 1.1 kJ/(kgK)
mg 526 kg/s
cpg 1.3 kJ/(kgK) mci 685 kg/s
Tcexit 400 °C mc/mci 18.8 %
Tf 1440 °C
Tb 895 °C
k1 0.1884 #
k2 1 #
mc 162 kg/s
cpc 1.1 kJ/(kgK)
mg 526 kg/s
cpg 1.3 kJ/(kgK) mci 685 kg/s
Tcexit 500 °C mc/mci 23.6 %
Tf 1440 °C
Tb 895 °C
k1 0.1884 #
k2 1 #
Pag. 140 of 202
4.10 inlet and outlet quantities of the compressor sizing are reported. Through flow shape of
the compressor has been obtained by the sizing procedure. Such a scheme is depicted in figure
4.1.Moreover, as a result of sizing procedure the blade to blade scheme with lumped blade
profile has been established and reported in figure 4.2. Blue and red blades represent rotor
blade and nozzle vane, respectively. On the right, number of the blades for each cascade is
given.
Table 4.1: Compressor Sizing Quantities
Compressor Quantities
INLET
1 Inlet Mass Flow mCi 685 kg/s
2 Inlet Air Composition (ISO) [xx]a [#] [#]m
3 Inlet pressure pci 101.3 kPa
4 Inlet Temperature tCi 15 °C
OUTLET
5 Exit Mass Flow mCo 582 kg/s
6 Exit pressure pCo 1843.6 kPa
7 Exit Temperature tCo 397 °C
8 Compressor Power Pc 264 MW
Such a through flow shape has been obtained by the sizing process of the compressor.
Abscissa and ordinate are non-dimensional. Geometric quantities refers to VIGV Hub
Leading Edge Radius. The value of the reference radius is 0.6925m.
4.1: H2-IGCC Compressor Through Flow Shape
Pag. 141 of 202
Fig. 4.2: Compressor blade to blade overview
Pag. 142 of 202
Once the compressor has been sized, off-design behvaiour has been investigated. Compressor
maps have been evalutated varying both the inlet compressor temperature and the Variable
inlet guide vane opening. Outcomes of the maps, in respect of the corrected mass flow, are the
pressure ratio and the isentropic efficiency. Accordingly, in figure 4.3 and 4.4 results of
compressor behaviour when ambient temperature changes, for a fixed VIGV opening, have
been presented.
Fig. 4.3: Pressure ratio versus corrected mass flow curves at different
compressor inlet temperatures
Fig. 4.4: Compressor isentropic efficiency versus pressure
ratio curves at different compressor inlet temperatures
12
14
16
18
20
22
24
100 105 110 115 120 125 130
Corrected mass flow
Pre
ss
ure
ra
tio
T=15°C -20°C -10°C 0°C 10°C
20°C 30°C 40°C 50°C
0,90
0,91
0,92
0,93
0,94
0,95
12 14 16 18 20 22 24
Pressure ratio
Ise
ntr
op
ic e
ffic
ien
cy
T=15°C -20°C -10°C 0°C 10°C
20°C 30°C 40°C 50°C
surge
choke
Pag. 143 of 202
On the other hand, by fixing the shaft speed (3000rpm) and the inlet temperature (15°C),
VIGV opening has been varied in the range 50% to 100%. Results of these calculations are
given in figure 4.5 and 4.6.
Fig. 4.5: Pressure ratio versus corrected mass flow curves at different VIGV openings
Fig. 4.6: Compressor isentropic efficiency versus pressure
ratio curves at different IGV openings
8
10
12
14
16
18
20
22
85 90 95 100 105 110 115 120
Corrected mass flow
Pre
ss
ure
ra
tio
IGV 100% IGV 90% IGV 80% IGV 70%
IGV 60% IGV 50% IGV 47%
0,9
0,91
0,92
0,93
0,94
0,95
8 10 12 14 16 18 20 22
Pressure ratio
Ise
ntr
op
ic e
ffic
ien
cy
IGV 100% IGV 90% IGV 80% IGV 70%
IGV 60% IGV 50% IGV 47%
surge
choke
Pag. 144 of 202
4.1.3 Combustion Chamber
According to the combustion chamber component model, described in paragraph 3.2.2, sizing
quantities concerning the CH4 burning process have been established, taking cycle calculation
and compressor and expander sizing outcomes into consideration. Input data are given in table
4.2. Combustion Chamber efficiency has been established according with the thermal power
related to the combustion process and to the typical CC heat losses.
Table 4.2: Combustion Chamber Input Data
INPUT QUANTITIES
1 Inlet Mass flow mai 514 kg/s
2 Inlet Pressure pai 1844 kPa
3 Inlet Temperature Tai 397 °C
4 Inlet Fuel Mass Flow Rate mf 15 kg/s
5 Low Heating Value LHV 50.0 MJ/kg
6 Firing Temperature Tf 1440 °C
7 Pressure Loss Δpcc 5.6 %
8 Combustion Chamber Efficiency ηcc 95.0 %
9 Air Composition
[N2] 23.1 %m
[O2] 76.3 %m
[H20] 0.0 %m
[CO2] 0.6 %m
10 Fuel Composition
[C] 75 %m
[H2] 25 %m
[H20] 0.0 %m
[N2] 0.0 %m
[O2] 0.0 %m
[S] 0.0 %m
Sizing procedure has led to evaluate the loss factor cck related to the pressure loss across the
combustion chamber that occurs mainly due to the frictional losses.
2
ai aicccc
ai ai
m Tpk
p p
(4.1)
In the sizing process, all quantities of equation 4.1 are known and the only unknown quantity
is cck . Output quantities of combustion chamber sizing are given in table 4.3.
Pag. 145 of 202
Table 4.3: Combustion Chamber output quantities
OUTPUT QUANTITIES
1 Outlet hot gas mass flow mgo 528.3 kg/s
2 Outlet pressure pgo 1788 kPa
3 Hot Gas Composition
[N2] 74.2 %m
[O2] 11.5 %m
[H20] 6.8 %m
[CO2] 7.5 %m
Off-Design of the combustion chamber has been investigated taking Data Base DB of Heavy
Duty Gas Turbine HDGT into consideration. In order to establish the off-design efficiency,
different quantities have been taken into consideration: inlet pressure gip , temperature
difference across the combustion chamber ccT , defined as rule 4.2
cc go aiT T T (4.2)
and the loss factor cck evaluated in sizing process. Related to this coefficient, pressure losses
across combustion chamber are evaluated by the rule 4.3
2
cc cc ip k (4.3)
being the correct mass flow evaluated at the combustion chamber inlet.
Fig. 4.7: Combustion Chamber Off-Design Curves data
Pag. 146 of 202
According to generalized relationship of HDGT combustion chamber off-design behaviour,
the model for calculating the combustion chamber efficiency in conditions different from the
nominal one depends on ccT across the combustion chamber and on the inlet pressure gip . In
figure 4.7 data related to combustion chamber are given for different inlet pressures
4.1.4 Expander
By the adoption of the expander component model, described in the paragraph 3.2.4, sizing
and off-design behaviour of the four stages 300MW F Class Gas Turbine expander have been
performed.
The preliminary cycle calculation has led to establish thermodynamic quantities at several
stations, mass flows and overall performance of the machine. According with Data Base DB,
firing temperature Tf, overall coolant mass flow mc, coolant temperatures Tcj as well as
exhaust pressure pex have been assumed as input for the expander sizing procedure.
Exhaust pressure has been set according to the typical pressure loss across the bottom heat
recovery steam generator (HRSG). Pressure loss HRSGp is assumed of some hundred water
mm. Number of stages Z has been established according with expander power P, mass flow
entering the gas expander, loading factor, degree of reaction and other parameters, taken from
DB. Inlet Gas mass flow mgi has been assumed as the sum of outlet compressor mass flow mco
and of fuel mass flow mf.
Coolant mass flows are given in column ‘mcj‘ of table 4.6 while ratios between coolant flows
and inlet compressor mass flow are given in column ‘mcj / mCi‘ of table 4.4, mCi being
compressor inlet mass flow (685.4 kg/s). Coolant mass flows have been assumed according
with a first estimation of cooling.
Table 4.4: Blade Cooling input
Stage Row mcj mcj / mCi
1 1S ms1 42.9 [kg/s] 6.3 [%]
1R mr1 39.6 [kg/s] 5.8 [%]
2 2S ms2 29.9 [kg/s] 4.4 [%]
2R mr2 23.0 [kg/s] 3.4 [%]
3 3S ms3 12.4 [kg/s] 1.8 [%]
3R mr3 15.5 [kg/s] 2.3 [%]
4 4S ms4 - - - -
4R mr4 - - - -
Input data of sizing process are given in table 4.5.
Pag. 147 of 202
Table 4.5: Expander Sizing Input Data
INLET QUANTITIES
1 Inlet Gas Mass Flow mgi 528.4 kg/s
2 Inlet Gas Composition [xx]gi [#] [#]m
3 Inlet Pressure (Total) pgi 1788 kPa
4 Inlet Temperature tf 1440 °C
5 1st Nozzle Inlet Angle α0 90 deg
6 Outlet Pressure (Static) pgo 104.3 kPa
7 Outlet Temperature tgo 577 °C
GLOBAL QUANTITIES
1 Shaft Speed n 3000 rpm
2 Stage Number nst 4 [#]
3 Mechanical Efficiency ηmec 99.5 %
STATOR QUANTITIES*
1 tip radius rts m
2 hub radius rhs m
ROTOR QUANTITIES*
1 tip radius rtr m
2 hub radius rhr m
* these quantities are given for any rotor/stator row
Output quantities of expander sizing have been calculated with an iterative procedure and
results of calculation, at the last step, are given in table 4.6.
Table 4.6: Expander output quantities
1 Exhaust Mass flow mex 687 kg/s
2 Exhaust Temperature (static) Tex 568 °C
3 Exhaust Temperature (total) Tex0 578 °C
5 Exhaust Pressure (total) pex0 108.6 kPa
6 Mechanical Power P 570 MW
Results of the sizing procedure are moreover given in the tables 4.7 in which geometric
quantities have been evaluated referring to the reference length:
0 0.6925r m (2.0)
0r being VIGV Hub Leading Edge Radius.
Table 4.7: Row by Row geometric quantities
Pag. 148 of 202
r mid Z t chax ch
Stage Row [m] [#] [m] [m] [m]
1 1 1.1228 69 0.1022 0.0945 0.1705
2 1.1297 79 0.0898 0.0999 0.1351
2 1 1.1523 66 0.1097 0.1152 0.1947
2 1.1793 76 0.0975 0.1016 0.1437
3 1 1.2127 51 0.1494 0.1507 0.2482
2 1.2496 61 0.1287 0.1511 0.2028
4 1 1.2943 35 0.2324 0.2000 0.2884
2 1.3507 41 0.2070 0.2196 0.2474
In figure 4.8 the through flow shape of the expander is given.
Fig. 4.8: Expander through Flow Section including the rear frame
Taking quantities of tables into account, velocity diagrams of rotor blades are plotted in
figures 4.9 – 4.12.
Velocities have been plotted with different colours as a briefly nomenclature shows
u green
w blu
c red
0
1
2
3
10 11 12 13 14 15 16
Pag. 149 of 202
Fig. 4.9: 1st Rotor Velocity Diagrams
Pag. 150 of 202
Fig. 4.10: 2nd
Rotor Velocity Diagrams
Pag. 151 of 202
Fig. 4.11: 3rd
Rotor Velocity Diagrams
Pag. 152 of 202
Fig. 4.12: 4th
Rotor Velocity Diagrams
Moreover, as a result of sizing procedure the blade to blade scheme with lumped blade profile
has been established and reported in figure 4.13. Blue and red blades represent rotor blade and
nozzle vane, respectively.
Pag. 153 of 202
Fig. 4.13: Expander blade to blade overview
As it has been done for the compressor component models, also for the expander off-design
behavior have been evaluated. Pressure ratio and total to static efficiency, versus the corrected
mass flow, have been established changing the turbine inlet temperature (firing temperature)
in the range 1240°C – 1640°C, 1440°C being the reference value. Moreover, exhaust pressure
has been kept constant at the nominal one as well as the shaft speed. Expander off-design
maps are summarized in figure 4.14 and 4.15,respectively.
Pag. 154 of 202
Fig. 4.14: Pressure Ratio vs Corrected mass flow
for different firing temperature
Fig. 4.15: Total to Static Efficiency vs Pressure Ratio
for different firing temperature
Pag. 155 of 202
4.2 Cooling System
The properly cooling off-design curve has been derived taking the design (nominal) point
described by an effectiveness and TCR of each row into account. In figure 4.15b, off-design
curves for each row of the 4 stage Generic 300 MW F Class GT are depicted. According to
the GT cooling lumped model, off-design curves have been presented also for the uncooled
stages in which anyway some heat is conducted by the blade to the disk and also drained to
the other components, as previously said. Such curves can be expressed by the following
exponential correlations
22.39
1 (1.01 3.01 )S e (4.1)
20.01
1 (1.01 2.78 )R e (4.2)
28.28
2 (1.01 3.85 )S e (4.3)
30.85
2 (1.01 4.93 )R e (4.4)
48.34
3 (1.01 8.48 )S e (4.5)
25.44
3 (1.01 5.50 )R e (4.6)
28.25
4 (1.01 8.27 )S e (4.7)
26.95
4 (1.01 7.28 )R e (4.8)
Fig. 4.15b: Off-Design cooling effectiveness VS TCR
Pag. 156 of 202
4.3 Gas Turbine Simulator
Gas Turbine simulator has been developed by assembly together the off-design maps of
compressor, combustion chamber, cooling and gas expander. Matching process leads to size
connections between compressor, combustion chamber and expander, by the introduction of
some components such as orifices connecting the compressor bleed exits and the expander
cooling ducts inputs. Connections between compressor and expander have been sized
according to the nominal coolant requirement of gas expander as well as the compressor exit
bleeding pressures. Once the connections have been sized, an off-design program of the
matching has been carried out and part load analyses have been performed. According with
the simultaneous solution described in the second chapter, a block diagram of gas turbine
matching is given in figure 4.16.
Fig. 4.16: ECRQP block scheme of the gas turbine matching
Pag. 157 of 202
A sketch of the generic 300MW F Class gas turbine representing the coolant flows, the shaft
and the other components is depicted In figure 4.17.
Fig. 4.17: Sketch of the Generic 300MW F Class GT Simulator
4.3.1 CH4 Gas Turbine
By means of such a gas turbine simulator the CH4 300MW F Class Gas Turbine looking like
the Siemens and the Ansaldo, described in chapter first, has been replicated and results at the
nominal running point are given table 4.9 and the through flow shape of the whole GT is
depicted in figure 4.18.
Fig. 4.18: Gas Turbine Through flow shape
0
1
2
3
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
Combustion
Chamber
Pag. 158 of 202
4.3.1.1 Nominal Running Point
Results of nominal CH4 GT running point are given in figure 4.9. The calculation has been
performed by assuming ISO ambient conditions (Tin=15°C, p=101.325kPa, RH=60%) and by
fixing some constraints on the firing temperature (Tf=1440°C) and on the exhaust pressure
(p=104.3 kPa). Results concerning the cooling component model are reported in table 4.10.
Table 4.9: RO3 Simulator - Nominal Running Point CH4 Fed
CH4 - 300MW F Class GT
COMPRESSOR
VIGV OPENING % 100
Inlet pressure kPa 101.3
Inlet temperature °C 15
Relative Humidity % 60
Inlet Mass Flow kg/s 685.4
1st bleed mass flow kg/s 13.7
2nd bleed mass flow kg/s 26.9
3rd bleed mass flow kg/s 52.7
4th bleed mass flow kg/s 83.7
Exit mass flow rate kg/s 592
Exit pressure kPa 1844.4
Exit temperature (total) °C 399
Compressor Power MW 262.9
COMBUSTOR
Compressed Air Mass Flow kg/s 508.4
LHV kJ/kg 50060
Fuel Mass Flow kg/s 15.0
Firing Temperature (total) °C 1440
EXPANDER
Inlet Mass Flow kg/s 523.4
Iinlet Pressure (total) kPa 1756.1
ISO TIT °C 1227
Exhaust mass flow kg/s 686.8
Exhaust temperature (static) °C 568
Exhaust temperature (total) °C 578
Exhaust pressure (static) kPa 104.2
Exhaust pressure (total) kPa 108.6
Expander Power MW 569.3
GAS TURBINE
Net Power MW 299.7
Efficiency % 39.8
Heat Rate kJ/kWh 9045.2
Pag. 159 of 202
Table 4.10: Results of the Lumped Model for cooling requirement - CH4
Stage Row mc [kg/s] Tb [°C]
1 s1 ms1 44 Tbs1 895
r1 mr1 40 Tbr1 880
2 s2 ms2 30 Tbs2 820
r2 mr2 23 Tbr2 810
3 s3 ms3 12 Tbs3 790
r3 mr3 15 Tbr3 760
4 s4 ms4 6 Tbs4 724
r4 mr4 7 Tbr4 613
The temperatures have the significant of the overall ‘cooling row phenomena’, thus the
coolant temperature TC is not the injection temperature but the lumped reference temperature.
Moreover, the coolant mass flows have the significant of the overall flow required to cool
airfoil surface, disk, sealing and all the other aspects previously described.
4.3.1.2 Part Load Analysis
Off-design GT matched simulator allows to perform a GT part load analysis to investigate the
response of the simulator when conditions change. Keeping the exhaust temperature constant
at the nominal value, relevant quantities have been plotted when load decreases from 100% to
70%. Accordingly, VIGV opening is reduced as well as pressure ratio, fuel mass flow and GT
efficiency. Such a behaviour is typical of such kind of mono-shaft heavy duty gas turbine.
Results of such an analysis are given in figure 4.19.
Fig. 4.19: CH4 fed GT part load behaviour
Pag. 160 of 202
4.3.1.3 Simulator Validation
Once the matching phase has been realized, the GT simulator has been validated by
replicating the real machine behaviour [3] for various ambient conditions. Taking the ambient
temperature change into consideration, the CH4 GT base load map and the Real Machine map
are given in fig. 4.20 and fig. 4.21, respectively. In the charts the Power Output and
Efficiency at Generator Terminals versus the ambient temperature are depicted.
Fig. 4.20: RO3 Power Output and Efficiency at Generator Terminals
Fig. 4.21: Siemens SGT5-4000F Power Output and Efficiency at Generator Terminals [3]
Pag. 161 of 202
4.3.2 Hydrogen Rich Syngas Gas Turbine
Due to the plant operating policies changing, also fuel composition changes influencing plant
operating conditions. When gas turbine is fuelled by a 33MJ/kg H2-Rich Syngas some
modifications have been carried out on the existing machine to ensure a stable and safe GT
running point. The optimum modification solution has been the 1st Nozzle Vane Expander re-
staggering [4,5,6]. Changing on the stagger angle leads the machine to work at the nominal
pressure ratio of the CH4 GT and re-design of coolant system allows the 33H2R Gas Turbine
to be operated without exceeding the threshold temperature of the hot components maintain
the life consumption rates at the desired values.
Taking the CH4 GT simulator into consideration, investigation on different fuel feeding leads
to explore the modification options that allows the GT to be fed by Hydrogen Rich Syngas. In
figure 4.22 nominal running point of CH4 GT Simulator fed by CH4 (red cross) and fed by
33H2R Syngas are reported. 33H2R fuel feeding off-design behaviour when conditions
change has also been depicted. In the figure 4..22 the Lowest Allowable Stall Margin
(LASM) line is given together with the expected compressor safe operation limits (surge and
choke lines). LASM is the limit that should not be overtaken to ensure safe and stable gas
turbine operating conditions.
Fig. 4.22: CH4 Gas Turbine Simulator Running Point for different fuel feeding
Pag. 162 of 202
The expected nominal running point using the H2-Rich syngas is almost on the LASM line
and such a phenomenon becomes more significant closing the VIGV. To avoid such
phenomena, re-staggering the 1st nozzle vane of the GT expander has been taken into
consideration. Re-staggering consists in opening the 1st Nozzle Guide Vane throat to allow the
GT to be operated CH4 nominal pressure ratio. Such an option ensures a stable and safe gas
turbine behaviour when it is fuelled by the H2-Rich Syngas. In figure 4.23, re-staggering
procedure has been sketched.
Figure 4.33: Stagger angle modifications
Such a modification implies a variation on metal blades temperatures that overcome the
threshold values, if no modification on the cooling system are carried out. Hence, by iterative
procedure, cooling system has been re-designed to allow the GT to be operated under safe and
stable condition.
During the iterations, both for the compressor and for the combustion chamber no
modification have been taken into consideration during the various re-staggering steps, while
the expander and the cooling systems have been modified [5]. Briefly description of the steps
of the iterative process are described and results of each step are summarized in table 4.11.
Case 0, represents the CH4 GT fed by methane syngas performance (power, pressure
ratio, coolant flows and temperature, etc.) at the nominal condition. Such a nominal
point is the benchmark for the comparison with the results of the new gas turbine
simulator, fed by the Hydrogen Rich Syngas (33H2RGT).
CASE A: The CH4 GT is fed with the H2-Rich Syngas characterized by 33MJ/kg
LHV. The outlet compressor pressure is higher than the value that assures safe and
stable GT operating conditions. Cooling flows are a little bit different from that of
Case 0 and blade temperatures are some 1-3°C higher than the threshold value;
Pag. 163 of 202
CASE B: 33H2R GT -1 Expander First Nozzle Vane (1st NV) has been opened,
modifying the stagger angle of some -1.68 degs to reduce the compressor pressure
ratio. No changes have been made on the lumped cooling simulator and the blade
temperature (red values) exceeds the maximum allowable values.
CASE C: 33H2R GT -2 Both expander and cooling component model have been
updated to allow the GT to be fed with 33MJ/kg Syngas. As a result the opening of the
1st NV is of some -0.68 degs and higher coolant flows are required to have the same
Blade Temperatures of the CH4 GT.
Table 4.11: GT Simulator and Cooling System Performance Results for the various Re-Staggering Steps
1 2 3 4
0 A B C
CH4
GT
CH4
GT
33H2R
GT-1
33H2R
GT-2
N N N Y
CH4 33H2R 33H2R 33H2R
56.4 56.4 54.72 55.77
/ / -1.68 -0.63
90 90 91.68 90.63
18.38 18.38 20.06 19.01
685.4 685.3 685.4 685.4
1844.4 1894.6 1844.3 1844.5
18.2 18.7 18.2 18.2
1440 1440 1440 1440
39.8 41 40.8 40.7
299.6 329.6 335.6 324.5
Stator 1 43.7 45.4 41.1 45.5
Rotor 1 40 41.6 39 42.8
Stator 2 29.8 30.5 29.3 31.5
Rotor 2 22.9 23.4 22.6 24.5
Stator 3 12 12 11.8 12.8
Rotor 3 14.9 15 14.8 17.2
Stator 1 895.3 897.8 917.9 895.4
Rotor 1 879.4 882.2 902.5 878.2
Stator 2 820.3 825 840.4 821
Rotor 2 807.1 811.1 825.2 806.4
Stator 3 786.7 789.8 800.7 786.5
Rotor 3 756.8 761 774.2 756.3
GT Efficiency [%]
GT Power [MW]
Cooling Mass Flows
[kg/s]
Blade Temperatures
[°C]
CASE
1st
vane blade inlet angle [°]
1st
vane blade exit angle [°]
Compressor inlet mass flow [kg/s]
Compressor outlet pressure [kPa]
Pressure Ratio [#]
Firing Temperature [°C]
MACHINE
Cooling update
Fuel
1st
vane Stagger angle [°]
1st
vane Stagger variation [°]
Pag. 164 of 202
4.3.2.1 33H2R Base Load Map
33H2R GT base load map has been evaluated following the same approach adopted to
perform the CH4 GT base load map calculation. Keeping exhaust temperature as constant at
the reference value (some 575°C) and changing the variable inlet guide vane (VIGV) owing
to the ambient temperature (Tamb) changes (figure 4.24), power and efficiency have been
obtained. Assumption on conditioning (pre-heating) inlet compressor flow when Tamb is
lower than 5°C, lead to justify the constant trend of power between 0°C and 5°C. Results of
such an investigation are given in the chart of figure 4.23.
Fig. 4.23: 33H2R GT –Load and efficiency non dimensional value versus ambient temperature
Fig. 4.24: 33H2R GT –Tex and VIGV non dimensional data versus ambient temperature
Pag. 165 of 202
4.3.3 Gas Turbine Control Rules
Definition of Gas Turbine Control Rules have been performed by the development of the GT
Neural Model (GTNM) that allows to evaluate, owing to the changing of the optimization
parameters (i.e. boundary conditions, prices, etc.) in real time, the best solution in terms of
VIGV and Tex. Feasible domain for the best GT control policy of the 33H2R GT, when
boundary condition changes, has been explored. Part load of the machine for each temperature
has been evaluated by adoption of control rules characterized by a VIGV and Tex value in
respect of the nominal running point. Such values are the results of the optimization of an fob
taking various aspects into account. In figure 4.24-a, non-dimensional values of Tex and
VIGV have been evaluated owing to the load reduction, when GT is operated under ISO
conditions.
Fig. 4.: 33H2R Gas Turbine Behaviour versus Ambient Temperature
Similar trends have been encountered when conditions different from the ISO occur. VIGV
and Tex trends for temperature in the range 0°C – 50°C have been explored, but not reported.
Proper control rules have been established for different boundary conditions (i.e. ambient
temperature) by means of GT simulator. Analyses have been carried out varying the GT load
from the peak load to a minimum load (i.e. 40%) for each ambient temperature.
Pag. 166 of 202
Changing load and ambient temperature, 33H2R- GT maps have been obtained. Each
calculated point is characterized by a VIGV and Tex value, established according to the
minimum of the above objective functions.
Life consumption rates of the hot components (i.e. blade metal temperature), limits on the
outlet compressor pressure and on the power driven by the shaft have been taken during the
gas turbine control rule definition into account.
Results, in a non-dimensional form, of such investigations are given in figure 4.24-b.
Optimized GT map is reported for various Tamb and for different loads, moving from the
peak value to the minimum one. Taking 33H2R GT nominal running point is possible to
calculate the power values for each temperature.
Fig.4.: 33H2R Gas Turbine Behaviour versus Ambient Temperature
Pag. 167 of 202
4.4 Steam Cycle Component Simulator
Hydrogen rich gas turbine nominal running point data as well as boundary conditions and H2-
IGCC H2R nominal operating conditions have been assumed to develop the steam cycle
component simulators. Sizing and off-design maps of the most important section of the steam
cycle have been performed. Accordingly, Heat Recovery Steam Generator (HRSG), Steam
Turbines and Condenser data have been reported in the following paragraphs. Pumps,
junction, valves have been modelled and simulated, too.
Taking the methodological approach described in the paragraph 2.3, off-design simulators of
the steam cycle section have been assembly together and nominal running point of the whole
steam section has been performed. Scheme of the steam cycle is given in figure 4.25.
4.4.1 HRSG
Cycle calculation and HRSG sizing has been performed taking exhaust 33H2R GT quantities
as well as characteristic temperature differences (approach, sub-cooling, pinch point) into
account. Both on gas and steam side, pressure drop across Heat Transfer Devices (HTD) have
been evaluated as well as heat transfer coefficients. Such quantities have been established by
the adoption of the most adequate correlation taken from the SoA of the three pressure level
HRSG of combined power plants, as described in the chapter third [7]. In table 4.12, gas
turbine exhaust quantities (mex=gmf and Tex=Tgi), stack temperature (Tgo) and HRSG outlet
pressure (pgo) are reported.
Table 4.12: Gas Side quantities for HRSG calculation
Gas Side
gmf Inlet Mass Flow [kg/s] 708
pgo Outlet Pressure [kPa] 101.3
Tgi Inlet Temperature [°C] 574.0
Tgo Outlet Temperature [°C] 110.0
The characteristic temperature differences refer to Super-Heaters (∆TAp), Boilers (∆TPP) and
Economizers (∆TSC), respectively.
∆TAP Approach Temperature Difference
∆TPP Pinch Point Temperature Difference
∆TSC Sub – Cooling Temperature Difference
Pag. 168 of 202
Fig. 4.25: Gas Steam Combined Cycle plant layout
Pag. 169 of 202
According to the present Data Base and to existing data in the scientific literature, approach
and sub-cooling temperature differences have been assumed while pinch point temperature
difference has been checked to be in the range of typical values for such kind of power plants.
Temperature difference for the three pressure lines are given in table 4.13
Table 4.13: Temperature Differences of HRSG
∆TAP [°C] ∆TPP [°C] ∆TSC [°C]
HP line High Pressure 44 35 - target 12
IP line Interm. Pressure 44 10 - target 12
LP line Low Pressure 39 10 – target 12
According with the plant layout and with the Data [8] concerning the deaerator pressure, the
steam mass flows entering and exiting the HRSG, pressure of the three lines and other, results
of sizing calculation are given in table 4.15 in which temperature profile (gas side and steam
side), pressure distribution (gas side and steam side) and heat transfer output (internal surface
of finned tube bundles and thermal power) are reported for each HTD. Nomenclature of HTD
has been chosen according with figure 4.25. Moreover, temperature profiles along the HRSG
both for the gas stream and for the three steam streams are presented in figure 4.26. Slope of
the lines refer to the HRSG sections and not to the heat or to the heat transfer device surface.
Tube banks in parallel arrangement (i.e. HP-SH and IP-SH, etc.) have been represented in the
same section (i.e. 0-1)
Fig. 4.26: Gas Side and Steam Side Temperature Profile along the HRSG stations.
Pag. 170 of 202
Table 4.15: HRSG Sizing Results
HT
D [
#]
GM
Fi
[kg/
s]V
mf
[kg
/s]
Tgi
[°
C]
Tgo
[°
C]
Tvi
[°
C]
Tvo
[°
C]
pgi
[kP
a]pg
o [
kPa]
pVi
[kP
a]pV
o [k
Pa]
S I
nt [
m2]
Qsc
[k
W]
151
5.4
143.
757
4.0
392.
733
9.1
530.
010
4.0
103.
714
250.
014
200.
036
50.7
1103
95.6
HP
_S
H
219
3.5
99.2
574.
039
2.7
353.
753
0.0
104.
010
3.7
4350
.043
00.0
1601
.741
434.
9IP
_S
H
370
8.9
13.1
392.
737
4.7
326.
033
8.0
103.
710
3.4
1425
0.0
1425
0.0
498.
814
719.
6H
P_
BO
470
8.9
50.5
374.
734
7.7
244.
332
6.0
103.
410
3.1
1430
4.8
1425
1.7
473.
621
907.
6H
P_
EC
3
570
8.9
17.8
347.
734
0.6
147.
930
0.0
103.
110
2.8
450.
040
0.0
90.6
5773
.8L
P_
SH
670
8.9
34.2
340.
626
6.7
242.
725
4.7
102.
810
2.5
4300
.042
99.9
2375
.659
300.
5IP
_B
O
743
3.3
50.5
266.
721
3.5
130.
724
4.3
102.
510
2.2
1435
8.3
1430
4.8
871.
625
699.
6H
P_
EC
2
827
5.6
34.2
266.
721
3.5
135.
924
2.7
102.
510
2.2
4350
.143
00.0
560.
516
344.
1IP
_E
CO
970
8.9
17.9
213.
516
4.4
135.
914
7.9
102.
210
1.9
450.
045
0.0
1644
.438
551.
8L
P_
BO
1033
0.4
50.5
164.
414
7.8
105.
013
0.7
101.
910
1.6
1441
0.5
1435
8.3
248.
860
28.8
HP
_E
C1
1137
8.5
52.1
164.
414
7.8
105.
013
5.9
101.
910
1.6
500.
045
0.0
308.
669
06.5
LP
_E
CO
1270
8.9
100.
114
7.8
110.
020
.090
.010
1.6
101.
318
0.0
130.
063
2.6
2931
5.6
PR
E
HR
SG
SIZ
ING
Pag. 171 of 202
4.4.2 Steam Turbine
According with the Steam Turbine component model described in the paragraph 3.3.3, the
steam turbine has been modelled taking the Equivalent Stodola Ellipse low into consideration.
The H2-IGCC steam turbine layout with admission and extraction between the various steam
turbine bodies (HP-ST, IP-ST and LP-ST) has been taken into account to perform the sizing
and the off-design maps of the components. Steam Turbine HP and IP bodies are governed by
the adoption of two throttling valves. In figure 4.27 H2-IGCC ST scheme is reported.
Fig. 4.27: Sketch of the three turbine bodies and the HRSG interactions
Equivalent sizing procedure leads to establish the unknown Stodola Ellipse parameters
required to perform steam turbine off-design calculation. Sizing has been performed assuming
the HRSG sizing outcome as reference quantities for the various steam turbine bodies and the
boundary conditions set according to the plant specifications (i.e. Gassifier pressure, required
steam mass flow, etc.). In table 4.16 sizing quantities, input and output, are summarized.
Reported inlet quantities refer before the governing valves. Temperatures (Tvi,Tvo), pressures
(pvi,pvo) as well as steam mass flows (smf) are the input data for the calculation and the
corrected mass flows (corr.smf) and the Stodola paramters (kSiz) are the outcomes.
Pag. 172 of 202
Table 4.16: Steam Turbine Sizing Quantities
Sizing outcomes have been used to perform the steam turbine off-design analyses. In figure
4.28 ST bodies off-design behaviours are given, varying inlet mass flows and for fixed
exhaust HP ad LP pressures. Results of these off-design analyses lead to developed the ST
simulator
Fig. 4.28: Steam turbine bodies (HP, IP, LP)
Off-Design behaviour for various steam mass flowand fixed condensing pressure.
Steam Turb smf [kg/S] Tvi [°C] Tvo [°C] pvi [bar] pvo [bar] corr.smf kSiz
Hp_ST 170 530 355 14000 4300 0.22 17.92
IP_ST 130 530 230 4300 400 0.58 2.99
LP_ST 116 240 21 400 2.5 0.00 4.42
INPUT OUTPUT
STEAM TURBINE SIZING
Pag. 173 of 202
4.4.3 Condenser
H2-IGCC plant condenser is a surface cooling water system fed by sea water. Condensing
pressure has been assumed to be at the nominal conditions of 2.5mbar, according with some
manufacturer declarations [9,10]. Taking the FV’s Multi-zone approach described in the
chapter second, sizing and off-design maps have been established for the condenser
component model.
Sizing has been performed by the assumption of some characteristic parameters (approach
temperature) and by the evaluation of the pressure loss and of the heat transfer coefficients
according with the state of the art correlations [7,11]. In table 4.17 some design parameters
are given. Moreover, steam inlet conditions (h steam), steam mass flow (smf), condensing
pressure and cooling water inlet temperature (Twi) have been given as input to the condenser
sizing simulator. Outcomes of the calculations are the required coolant mass flow (WMF), the
thermal power transferred from the hot stream to the cold stream (Qth) and the surface (S).
Table 4.17: Condenser relevant sizing quantities
Fig. 4.29: Condenser Off-Design behaviour for different steam flows
Condensing pressure and cooling water temperature VS steam mass flow
Twi ∆Tapp pcond h Steam Smf WMF Qth S
[°C] [°C] [mbar] [kJ/kg] [kg/s] [T/h] [MW] [m^2]
10 2 25 2230 118 1.8 248 9600
INPUT
CONDENSER SIZING
OUTPUT
Pag. 174 of 202
Once sizing phase has been performed, condenser off-design component model has been
carried out. Outcome of sizing have been used to perform the condenser off-design maps
Pressure and cooling temperature trends versus the condensing mass flow reduction are
similar to that established by [12], adopting a similar condenser component model for a
different power plant layout [12,13]. Such an analysis has been carried out by keeping
constant (at nominal value) the cooling water mass flow. Results of such evaluation are
summarized and reported in figure 4.29.
4.5 Steam Cycle Simulator
Matching of the various components (HRSG, Steam Turbine, Condenser, pumps, junctions,
degasser, etc.) constituting the Steam Section leads to develop the Steam Cycle Simulator.
In figure 4.30, a skecth of the steam cycle matching algorithm is depicted.
Fig. 4.30: ECRQP block scheme of the steam cycle matching
Nominal running point of the steam section has been obtaind taking reference the 33H2R GT
exhaust quantites as well as boundary and operating conditions into account. Off-Design
Pag. 175 of 202
steam cycle component models have been assembly together and connection between them
such as the opening of the valves have been sized. In table 4.18 results of the calculation are
reported. Pressures, temperatures, powers and efficiencies have been reported. Accordingly,
once the connections have been sized, matching simulator has been adopted to perform part
load analyses.
Table 4.18: Steam Cycle Simulator – Nominal Running Point
Pag. 176 of 202
4.6 Power Island Simulator
Once gas turbine and steam cycle component models and simulators have been developed,
they have been matched together to achieve the power island simulator. Block scheme of the
macro components constituting the power island and of inlet and outlet quantities crossing the
power island boundary is given in figure 4.31.
According with the H2-IGCC power plant layout, the various interactions between power
island and gasification one have been taken into consideration. Such interactions concern
water and steam mass flows taken from the HRSG and sent to the gasification island (i.e.
Syngas Cooler) et vice versa. Extracted and admitted mass flows values have been assumed at
the nominal point taking the available data concerning the IGCC plant behaviour when the
gas turbine is fed by Hydrogen rich syngas into account. Under these conditions, water and
steam flows are required by the gas treatment and carbon capture and storage sections to
obtain such a 33MJ/kg Hydrogen Rich GT fuel.
Fig. 4.31: Schematic view of the H2-IGCC Power Island
According with the methodological approach (paragraph 2.3), matching process between the
GT and the SC leads to size the connections of the various components (i.e. valves opening).
Opening of valves and all other sized quantities during the sizing matching phase have been
adopted to perform the power island part load analysis. In table 4.19, nominal running point
of the power island, under ISO conditions and for 33H2R fuel, is reported.
Pag. 177 of 202
Table 4.19: Power Island Nominal Running Point – ISO Conditions
Gas Turbine
VIGV % 100.0
m1 kg/s 683.5
mf kg/s 23.8
Tf °C 1437.8
mex kg/s 693.7
Tex °C 574.0
pex kPa 104.0
P MW 324.2
ETA % 40.7
BETA # 18.2
TIT °C 1221.0
fS1 # 0.9
fR1 # 0.9
fS2 # 1.0
fR2 # 0.9
fS3 # 0.9
fR3 # 0.9
TwS1 °C 894.5
TwR1 °C 877.4
TwS2 °C 820.2
TwR2 °C 805.7
TwS3 °C 785.6
TwR3 °C 755.7
Steam Cycle
GMFoGT kg/s 707.3
TGoGT °C 574.0
pGoGT kPa 104.0
TStack °C 110.5
pStack kPa 101.3
VMFoHP kg/s 14.1
VMFoIP kg/s 34.6
VMFoLP kg/s 18.6
TVoHP_SH °C 530.5
TVoIP_SH °C 531.8
TVoLP_SH °C 298.9
pVoLP_ST kPa 2.5
PST_HP MW 39.6
PST_IP MW 55.7
PST_LP MW 80.7
PST MW 176.0
Pag. 178 of 202
Power island part load behaviour has been investigated taking the control rules described in
the previews paragraphs (4.3.3) into consideration. Under ISO conditions, power island
outcomes have been evaluated, when GT load change from the peak value to a 60% of the
nominal. In this section, results of such a part load investigation have been reported in a non-
dimensional form. Dimensional values can be evaluated taking results of table 4.19 into
consideration. In figure 4.32, relevant quantities connected with the gas turbine part load
behaviour are reported. Variable inlet guide vane (VIGV) and exhaust temperature (Tex) have
been evaluated according to the optimum control polices described in paragraph 4.33, hence
their variation is related to the ambient temperature and to the GT load. Efficiency (ETA) and
pressure ratio (BETA) are outcome of the calculation. It can be highlighted that the adoption
of suitable control rules allow the machine to be operated under the Lowest Allowable Stall
Margin (LASM), in a safe and stable domain. Fuel mass flow (mf) decreases as a
consequence of the load reduction and of the adopted control rule.
Fig. 4.32: Non dimensional values of GT relevant quantities
for ISO conditions and changing GT load
Under the same operating conditions, gas turbine wall temperatures have been monitored to
ensure that life consumption rates of the machine components do not exceed the desired
values. As a result of the optimization of the fob, the gas turbine control rules allow to
maintain the expander blades temperatures lower than the reference ones, for the all domain
of the load variation. It means that the ratio between gas turbine virtual operating life and
Pag. 179 of 202
operating hours is almost all the time lower than 1.0, reducing the terms of the objective
function related to the gas turbine cost.
Steam Cycle (SC) outcomes have been evaluated by means of the power plant simulator,
depending both on the gas turbine control rules, because of the exhaust quantities (exhaust
temperature and mass flow) and on the integration of the power island with the gasification
section, because of the entering and exiting steam and water streams. Accordingly, in figure
4.33 the trend of the main steam section quantities versus load are given.
Reducing GT load, intermediate (IMF_BO) steam production decreases with a similar
reduction ratio of the exhaust gas mass flow entering the HRSG, till to 70% load. More
relevant reduction takes place when Tex is reduced. A limited decrease in high pressure steam
production (HMF_BO) is observed, but this fact does not give a significant contribution to
power production owing to that the HP steam generated in HRSG is lower than that produced
in other IGCC plant sections (i. e. the syngas cooler). The super-heated steam temperatures
(THP_SH, TIP_SH, TLP_SH) remain practically unchanged, therefore there is no need to
control them by introducing attemperators.
The relevant reduction of the high pressure steam turbine output P_HPST is related to the
huge governing valve throttling required to accommodate the decrease in steam mass flow.
Such a reduction is partially compensated by the behaviour of the intermediate and low
pressure turbines. Pressure at condenser is also influenced by load variation. When steam
mass flow decreases, condensing pressure decreases, too.
Pag. 180 of 202
Fig. 4.33: Non dimensional values of the steam side relevant quantities for various ISO conditions loads
Pag. 181 of 202
4.7 Reference
[1] - Jonsson M., Bolland O., Bucker D., Rost M. (Siemens), 2005, ‘Gas Turbine Cooling
Model for Evaluation of Novel Cycles’. Proceedings of ECOS 2005, Trondheim, Norway,
June 20-22, 2005.
[2] - Ashok Rao., 2010, ‘1.3.2 Advanced Bryton Cycles’.
[3] - SIEMENS AG, Siemens Gas Turbine SGT5 – 4000F. Answer for energy, 2008.
[4] – Cerri G., et Al. 2011: ‘Selection of the best IGCC Cycles – Cycle options analysis’,
Milestone 4.1, H2-IGCC EU Project.
[5] – Cerri G., Chennaoui L., Giovannelli A., Mazzoni S., 2014: ‘Expander Models for a
generic 300MW F Class Gas Turbine for IGCC’, GT2014 – 26493, Proceedings of ASME
Turbo Expo 2014, Dϋsseldorf, Germany, June 16-20, 2014.
[6] - Jones R., Golmeer J., Monetti B., GE Energy, Addressing Gas Turbine Fuel Flexibility,
GER4601 (05/11) revB.
[7] – Kreith F., Manglik R.M., Bohn M. S., 2011: ‘Principles of Heat Transfer – 7th
Edition’
[8] - http://www.h2-igcc.eu/default.aspx
[9] – Siemesn AG 2010: ‘Siemens Steam Turbine SST-3000 Series for combined cycle
application’.
[10] – Emberg H., Alf M., SCC5-4000F Single Shaft (SST5-5000): ‘A single shaft concept
for cold cooling water conditions’.
[11] - Rohsenow W. M, Hartnett J. P, Cho Y. I.,(1998): “Handbook of Heat Transfer – Third
Edition”, McCraw-Hill Handbooks.
[12] - Cerri G., Chennaoui L., 2013, “General Method for the development of Gas Turbine
based plant simulators: an IGCC application”, Asme Turbo Expo 2013, San Antonio, Texas
(USA), June, 3-7, 2013.
[13] – Cerri G., Mazzoni S., Salvini C, 2013: ‘Steam Cycle Simulator For CHP Plants’, Asme
Turbo Expo 2013, San Antonio, Texas (USA), June, 3-7, 2013.
[14] - Mohagheghi M., Shayegan J., 2009: “Thermodynamic optimization of design variables
and heat exchangers layout in HRSGs for CCGT, using genetic algorithm”, Applied Thermal
Engineering, 29 (2009) pp. 290-299.
[15] - Franco. A, Russo A., 2001: “Combined cycle plant efficiency increase based on the
optimization of the heat recovery steam generator operating parameters”, International Journal
of Thermal Sciences, 41 (2002) 843 – 859.
[16] - Bonataki E. T., Giannakoglou K. C., 2005: “Preliminary Design of Optimal Combined
Cycle Power Plants Through Evolutionary Algorithms”, EUROGEN 2005.
Pag. 182 of 202
Chapter V
IGCC Plant Simulator
5.0 Introduction
Power island component models description as well as gas turbine and steam cycle simulator
development have been widely discussed in the chapter three and four. In this chapter
Gasification Island, Power Island and H2-IGCC Plant are described.
The matching between the gasification and the power island leads to obtain the whole H2-
IGCC power plant simulator. As shown in Fig.5.1a, the plant is characterized by a high level
of integration between sections constituting the plant itself. There are mass, heat and work
exchanges among components that lead to a strong interaction in the behaviour of the
different sections. IGCC simulator takes all these aspects into consideration.
Fig. 5.1a: Sketch of IGCC Plant
Power, Heat and mass flow interactions between the various plant sections
By adopting such a simulator tool, plant operating policies, boundary conditions, prices (i.e.
coal, CO2, electricity, etc.) and other aspects that influence the plant performance have been
taken into account during the whole system mapping. Feasibility domain of the solution and
safe and stable components behaviours (i.e. not exceeding the threshold thermal and
mechanical stresses, etc.) have been taken into consideration to find the best solution.
A Description of the IGCC simulator is given in the following paragraphs. Plant control
philosophy and control policies are presented. H2-IGCC plant performance when boundary
conditions and gas turbine load change have been evaluated and discussed and time
dependent plant ramp have been reported. Discussion of results is given.
POWER
ISLAND
NET POWER
Pag. 183 of 202
5.1 Gasification Island Simulator
Taking figure 5.1 into account, interactions between electric grid, power section and
gasification island have to be taken into consideration by the whole plant simulator. Load,
primary coal consumption, H2 Rich Syngas mass flow GT demand, power consumptions,
CO2 production and others have been correlated by the adoption of some correlations.
Various interaction between sections are schematically represent also in figure 5.1b.
The gasification model is focused on the description of phenomena and processes giving a
significant contribution to the electric power production. Accordingly, less relevant aspects
are neglected or treated in a simplified way.
Fig. 5.1b: IGCC Layout Block Scheme
Pag. 184 of 202
According with figure 5.1b and with the paragraph 3.3.9, the interaction between Gasification
Island Simulator and the whole system have been taken into account during the H2-IGCC
plant simulator development. Therefore, Gasification island and power island interaction in
terms of power, heat, pressure and mass flows have been taken into consideration by the
following functional correlations.
Coal mass flow is related with many plant quantities such as the gas turbine fuel demands,
the plant load and the coal composition. In equation 5.4 a correlation between the whole
system main relevant quantities is given, kcoal being a proportional parameter.
f1(load, mf, mcoal, kcoal, [xx]coal)=0 (5.4)
When GT load and fuel mass flow change, power consumption of the gasification section
changes too. Introduction in the model of typical polynomials that correlate the centrifugal
compressor power to the inlet mass flow has been taken into account. For some components
such as the coal mill, the power consumption has been kept constant at the nominal value also
when coal mass flow changes owing to the operating conditions. Relation between
gasification island power consumption (Pj), load and syngas mass flow (mf) and its
composition is presented, kjp being a coefficient or a function (5.5).
f2(load, mf, Pj, kjp)=0 (5.5)
According with the figure 5.1b and the description of the many interactions between the
whole system sections (i.e. HRSG, WGS, Syngas Cooler, etc.), steam and water mass flows
(msj) interconnections between GI and PI are related to load and fuel mass flow by means of
proportional coefficients kjs (5.6)
f3(load, mf, msj, kjs)=0 (5.6)
CO2 mass flow has been evaluated by the adoption of the CCS model and it mainly depends
on the coal and syngas compositions, on the plant load and on the fuel mass flow. CO2 is
related to the other quantities by the coefficient kCO2 (5.7).
f4(load, mf, CO2, Coal, kCO2, [xx])=0 (5.7)
Pag. 185 of 202
5.2 Control policies for optimum, safe and stable operating conditions
Engines and thermo-mechanical devices are generally designed to work under safe and stable
conditions to ensure a certain operating life. When operating conditions change threshold
values of temperatures, pressures, powers and other quantities may be overcome and
operating costs increase. In order to avoid such a drawback, suitable control policies have
been adopted. Such aspects lead to define a feasible domain in which pressures, temperatures,
power and other quantities should be limited.
To establish the most convenient control policy of the plant components, optimization of an
Objective Function (fob) has to be performed.
5.2.2 Plant Control Philosophy
Highly integrated power plants performance is affected by many parameters. Some of them
are strictly connected to the plant components (i.e. hot components temperatures, life
consumption rates, pressure limits, etc.) and others are related to techno-economic aspects
(i.e. electricity price, fuel cost, etc.) and to environmental aspects (i.e. CO2 emission, taxes,
etc.). All the above aspects are taken by the objective function into consideration. To establish
the control philosophy of the plant, minimum (or maximum) of the fob has to be searched,
taking the variables feasible domain into account.
The earning representing the fob is the difference between sold and purchased assets.
In detail power selling, fuel cost, operating costs, taxes and complementary products (costs
and sales) have be taken in the formulation of the fob into account.
The Earning, that should be maximized can be expressed as follow (5.1).
2 2
2 2 2 2
1 1 1 1
t t t t
p p s s
el el el el CO CO coal coal
t t t t
E p P dt p P dt p m dt p m dt
2 2
1 1
t t
lcr j j k K
j kt t
p f dt p q dt (5.1)
Pel, mCO2, mcoal, fJ, qk being the electric power, the CO2 mass flow, the primary coal mass flow,
the life consumption rate of the j-th component and the complementary products k-th,
respectively. pi are the prices of the above mentioned quantities. The super-script p and s
represent the purchased and sold electricity, respectively.
Pag. 186 of 202
In the simulator correlations between plant operating parameters, such as the Variable Inlet
Guide Vane (VIGV) opening and the Turbine Exhaust Temperature (Tex), Syngas Cooler
Temperature and Gasification Pressure, and variables are embedded into the models. With
reference to the Gas Turbine control policy, two relationships between GT control quantities
and variables involved in the optimization procedure are described by the rules (5.2) and (5.3)
1( , , , 2, , ,...) 0f Jf VIGV m f CO prices power (5.2)
2( , , , 2, , ,...) 0f Jf Tex m f CO prices power (5.3)
All the functions should be represented in a well-defined domain to take the feasibility aspect
of the solution into account.
The plant control philosophy to safe operate the whole system and to ensure a revenue takes
various aspects into consideration. Gasification Island as well as gas turbine have to be
operated under well define conditions to do not exceed the life consumption rates of
components owing to thermal and mechanical stresses (i.e. pressure, components metal
temperatures, shaft, electric generator, etc.).
Gas Turbine has to be operated maintaining the pressure ratio and the blade metal
temperatures under the reference values and the gasification section has to be operated
keeping some temperatures (i.e. Syngas Cooler Temperature 900°C) and some pressures (i.e.
gasifier pressure 43bar) at the nominal value during the plant life. Accordingly, monitoring,
controlling and regulation systems, whose input quantities are power, temperatures and
pressures and others, allows to restore the set point values to assure the desired plant
behaviour. An example of the GT syngas admission valve system is depicted in figure 5.2b.
When GT and plant load decreases nominal running point of the gas turbine decreases also. A
reduction of the required fuel mass flow is encountered. Accordingly, control valve opening is
reduced to increase the pressure losses along the fuel paths. Such operation allow the
gasification island pressure to be maintained as constant.
In figure 5.2a the pressure loss introduced by the GT fuel control valve is sketched as well as
the pressure trends of the GT burner (red line) of the inlet injectors pressure (green line) and
of the p1 pressure (blue line). Opening of the valve is also plotted.
Pag. 187 of 202
Fig. 5.2a: pressures trends and valve opening versus plant load
Fig. 5.2b: Sketch of the control system of the GT fuel admission valve
Pag. 188 of 202
5.3 H2-IGCC Plant Simulator
Matching between Gas Turbine, Steam Turbine and Gasification Island simulator has been
performed to establish the connections (valves opening, proportional coefficients, etc.)
between the various simulators, according with the methodologies described in chapter
second. Interactions between power island and gasification island have been taken into
consideration during the calculations. Opening of valves, proportional coefficients and all
sized quantities during the matching phase have been adopted to perform the part load
analysis, according with the methodological approach described in the paragraph 2.3.
5.4 H2-IGCC Plant Mapping
Due to the changing of ambient conditions, coal composition, prices (i.e. CO2, Coal,
electricity, etc.) and of all the other aspects that influence the plant behaviour, the best plant
managing policy changes continuously. Adoption of the H2-IGCC plant simulator gives the
possibility to forecast in short time the plant outcomes (internal quantities, power, efficiency)
that allow to better operate the whole system (i.e. maximize an earning).
The short computational occupancy and short calculation times allow to evaluate many case
changing parameters such as ambient temperature, pressure, coal composition, power demand
and others. Accordingly, full mapping of the whole system can be performed.
In this work, investigations on plant performance trend when boundary conditions change
have been carried. Ambient temperature has been changed in the range 5°C to 45°C.
Moreover, the simulator has been used to map the plant for ISO conditions varying the gas
turbine load between some 60% of the nominal value and to the peak value. Test case that
have been evaluated are reported in table 5.1.
Table 5.1: Whole System Map – Test Case
Results and discussion of these investigations are reported in the following paragraphs.
5 15 25 35 45
x x x x x
1.02 x
1.00 x
0.90 x
0.80 x
0.70 x
0.60 x
Test CaseAmbient Temperature [°C]
GT LOAD
[#]
GT BASE LOAD
Pag. 189 of 202
5.4.1 GT Load Changes
GT Exhaust temperature and mass flows trend versus the GT load variation have been
presented in figure 5.3. As a results of the control rules defined in the last chapter (4.3.3), it
can be remarked that exhaust temperature is reduced only when the VIGV cannot close
anymore and that is kept as constant for the bottomed HRSG. Accordingly, the exhaust mass
flow shows a trend similar of that of the VIGV.
Fig. 5.3: ISO Conditions – GT Exhaust Mass Flow and Temperature VS GT Load
Another result of the adoption of the previously described GT control rules is shown in figure
5.4. In such a figure the trend of the life consumption rates (ffj) related to the cooled expander
blades is given. The ratio ffj between gas turbine virtual operating life and operating hours is
almost all the time lower than 1.0 when the GT load is lower than the nominal one and
becomes higher Moving to the peak value, 1.015 of the nominal Power.
Reference CH4 fed gas Turbine has been designed to produce 300MW at the nominal
conditions. Under the same compressor nominal running point (i.e. pressure ratio=18.2 and
inlet mass flow=685kg/s), the power of the re-staggered gas turbine is some 324MW with an
increase of the GT efficiency. For sure, according with electro-mechanical limitations the
maximum power that can be given also by the 33H2R GT is 330MW, because no re-design of
shaft and electric generator has been performed. For such a reason the load peak value is
1.015 (330:324=1.015). According to that, the plot of figure 5.4 shows that gas turbine
feasible domain is limited for electro-mechanical reasons and not only for thermal ones.
Compressor
Nominal Point
Pag. 190 of 202
Fig. 5.4: ISO Conditions – GT Nozzle Vane and Rotor Blade life consumption rates VS GT Load
On the steam cycle side, according with the results given in the paragraphs 4.6, super heating
temperature and steam mass flows trends of the three pressure lines (140, 43, 4 bar) versus the
gas turbine load are given in figure 5.5 and 5.6, respectively.
Fig. 5.5: ISO Conditions –Superheating temperature (HP, IP, LP) VS GT Load
Pag. 191 of 202
Taking figure 5.5 into account, HP and IP super-heater outlet temperatures trends are quite
similar to that of the gas turbine exhaust temperature. Load reduction from peak value to the
70% leads a very small increase of the temperatures (only some °C) while moving from the
70% to the 60% a significant temperature reduction can be observed both on HP and IP line.
LP super-heater outlet temperatures shows a trend similar to the HP and IP temperature, even
if it decreases less than the other two. It depends on the fact that only 20kg/s of steam are
produced in such a line, while some 140 and 100 kg/s are produced in the HP and IP SH’s.
Such a small variation on the temperature makes not necessary the adoption of attemperator
systems, used to ensure safe behaviour of the HRSG tube banks.
In figure 5.6 evaporator outlet steam mass flows trends versus the GT load are given. The
trends are not very different for GT load changes between the 100% and 70% while show
differences between 70% and 60%. Such differences are connected to the high integration
level of the steam section with the whole system. Admissions and extractions of steam and
water mass flows affect also the temperature profile distribution of the HRSG streams.
Accordingly, different thermal power are exchanged between the hot gas and the ‘cold’
stream in the different tube bundles. Thus, a small change on the steam produced in the
boilers has been encountered.
Fig. 5.6: ISO Conditions – Boiler Outlet Steam Mass Flow (HP, IP, LP) VS GT Load
Pag. 192 of 202
By the adoption adoption of the gasification island simulator, the whole system power
consumption and production have been evaluted. In figure 5.8, the combined cycle power PCC
(5.8), the whole system power PIGCC (5.9), the overall power consumption P- and the steam
turbine power trends versus gas turbine load changes have been presented.
CC GT STP P P (5.8)
IGCC CCP P P (5.9)
P- being the sum of power consumption of the pumping system, of the CO2 and ASU
compressors and of the other power requirements of the gasification components. It is the
Aux. Power represented in figure 5.1b.
Fig. 5.7: ISO Conditions – Whole System power VS GT Load
In figure 5.8 the ratios between IGCC power, steam turbine power and consumed power
versus the gas turbine power are given. From the analysis of such a figure, it can be remarked
that ratio between steam turbine and gas turbine power remains practically unchanged, when
gas turbine load decrease. Moreover, taking figure 5.5 into account, a similarity between
temperature trends and such power ratio can be observed. Ratio between P-
and GT Power
shows that reducing the load the power consumption is reducing less than the other power,
giving a contribution to the whole plant efficiency reduction. Accordingly, to the above
describe trends, IGCC power vs GT power trend is practically constant between 100% and
70% and decrease between 70% and 60% GT load range.
Pag. 193 of 202
Fig. 5.8: ISO Conditions – Power Ratio VS GT Load
In figure 5.9, primary coal consumption and 33H2-Rich fuel mass flow GT demands have
been presented. Coal mass flow has been evaluated by means of relation (5.4) Gas turbine
load reduction leads to a reduction of the required fuel mass flow and consequently of the
primary coal demand.
Fig. 5.9: ISO Conditions –33H2R Syngas and primary coal mass flow VS GT Load
Pag. 194 of 202
Efficiency of the whole plant (ETAIGCC) is defined as the ratio between the net electric power
and the chemical power introduced by coal, characterized by its Coal Heating Value (CHV).
Such an efficiency is expressed as the rule (5.9)
IGCCIGCC
Coal
P
m CHV
(5.9)
To perform such calculations, the mass composition of the adopted coal is given in table 5.2
Table 5.2: Coal mass fraction composition [6]
In figure 5.10, trends of whole plant efficiency and power is given. Gas Turbine load
reduction leads to a reduction of efficiency and of power, in according with the consideration
given in the previews plots and paragraphs.
Fig. 5.10: ISO Conditions –IGCC Power and Efficiency VS GT Load
Pag. 195 of 202
5.4.2 Ambient condition changes
Charts similar to that shown in the paragraph 5.4.1 ‘Load Changes’ have been presented in
this section. An analysis on the H2-IGCC plant behaviour when the ambient conditions
(Ambient Temperature) change has been carried out. Temperature variability has been
considered in the range 5°C – 45°C.
In figure 5.11, GT exhaust temperature and mass flow versus ambient temperature have been
reported. According with the 33H2R Base Load Map (4.3.2.1), exhaust temperature is
constant and the exhaust mass flow is higher for low temperature and become lower when hot
conditions occur. Such behaviour is typical of such king of heavy duty GT engines.
Fig. 5.11: GT Exhaust Mass Flow and Temperature VS Ambient Temperature
On the steam side, trends different from that presented in the previews paragraph have been
obtained and shown in figure 5.12 and 5.13. Super heating temperatures trends versus the
ambient temperature are similar for the three pressure lines. Also in such conditions can be
remarked that no attemperators are needed.
Steam mass flows trends versus ambient temperature are presented in figure 5.13. Interactions
between Gasification Island and Power Island lead to have different behaviours on the three
pressure line. IP and LP Boiler outlet mass flow decrease when ambient temperature
increases. An inverse trend is shown by the HP line.
Pag. 196 of 202
Fig. 5.12: ISO Conditions –Superheating temperature (HP, IP, LP) VS Ambient Temperature
Fig. 5.13: ISO Conditions – Boiler Outlet Steam Mass Flow (HP, IP, LP) VS GT Load
Pag. 197 of 202
In figures 5.14 and 5.15 Power and Ratio between IGCC, Steam Turbine and overall power
consumption versus gas turbine power are presented.
When ambient temperature increases, the various power decrease. Each power is
characterized by its slope, but the trends are similar for all the quantities reported in figure
5.14. Steam power and P-
decrease less than the IGCC and combined cycle power. Such a
difference is related to the fact that gas turbine power is twice the amount of the steam turbine
power and decrease much more, as shown in the paragraph 4.3.2.1, concerning the GT map.
On the other hand, in figure 5.15 ratios increase when ambient temperature increase too. In
this case the most significant change is observed on the steam turbine trend. Similar trend
affects the IGCC power ratio while for the power required by pumps, compressors and others
a really flat curve is given.
Accoding to the GT base load map, also fuel mass flow and consequently the primary coal
mass flow decrease when hotter climate conditions take place. Such behaviours are
summarized in figure 5.16. Also whole plant power and efficiency show a similar trend owing
to the temperature increases. According with the consideratio of the above trends, whole plant
output are presented in figure 5.17.
Fig. 5.14: ISO Conditions – Whole System power VS Ambient Temperature
Pag. 198 of 202
Fig. 5.15: ISO Conditions – Boiler Outlet Steam Mass Flow (HP, IP, LP) VS GT Load
Fig. 5.16: ISO Conditions –33H2R Syngas and primary coal mass flow VS GT Load
Pag. 199 of 202
Fig. 5.17: ISO Conditions –IGCC Power and Efficiency VS GT Load
Pag. 200 of 202
5.4.3 Discussion and Concluding Remarks
H2-IGCC power plant behaviour has been analysed by means of the IGCC power plant
simulator. Changing on ambient conditions and on power demands have been taken into
consideration and whole system performance as well as internal quantities have been
calculated and monitored. Accordingly, the simulator capability of being the replica of the
reference plant has been shown during the mapping of the IGCC plant behaviour.
Results of the investigations allow to make some consideration about the possibility of
burning Hydrogen Rich Syngas (HRS) in the gas turbine. By adopting a minor structural
modification on the gas turbine, the opening of the 1st nozzle of the expander, a CH4 designed
gas turbine can be easily adapted to be fuelled with HRS under safe and stable operating
conditions. No increase on the life consumption rates of the gas turbine hot components have
been observed and GT efficiency similar or higher than the CH4 one have been disclosed.
Hence, more than 300MW are produced by the gas turbine with some 40% of efficiency.
Concerning the combined cycle section, the adoption of a three pressure level HRSG has been
strictly connected with the whole plant steam and water requirements. The steam turbine
layout with the three bodies (HP, IP, LP) and the various steam admissions and extractions is
typical of such a kind of high integrated power plants. Such a configuration allow to produce
more the 170MW by the bottomed cycle, under the H2R nominal running point exhaust
conditions (710kg/s mass flow , 575°C exhaust temperature). Accordingly, combined GT and
Steam Turbine ST cycles are the highest efficiency energy converters, because of the
peculiarities related to the high peak cycle temperature of the GT and the low heat rejection
temperature of the steam condenser in which the major heat fraction (some 80%) is
discharged at the constant temperature of the condenser and of the stack in which the minor
fraction (some 20%) is released at the exhaust temperature. Therefore, combined cycle
efficiency, without taking the gasification isle into consideration, is of some 60% at the
nominal conditions.
Integrating the power island with the gasification island with the carbon capture and storage,
the whole plant efficiency has been established. Taking the LHV of some 25MJ/kg, the coal
mass flow of some 44kg/s and the net power of the plant (produced power minus auxiliary
power) into consideration, more than 36% IGCC plant efficiency has been evaluated. For
sure, the various processes to capture and storage the CO2 require some power and for such a
reason the IGCC plant efficiency decreases from the some 44% in the case of turned off CCS
section. Value of such an efficiency is a good results because it is higher than the typical
Pag. 201 of 202
IGCC efficiency [9] and allows economic revenues ensuring the answer for the clean coal
energy.
The ability of the simulator to reply the H2-IGCC plant performance in a wide operating
domain and the short computational time allows to employ such a IGCC plant simulator tool
for optimum plant managing and planning purposes. Accordingly, taking formulation of the
objective function (earning that has to be maximized) (5.1) into account, performance of the
plant can be evaluated along the time when power demands and operating conditions are far
from the nominal one. Terms of the various integrals can be established by such plant
simulator. Accordingly, primary coal mass and GT fuel flow demands and other plant
outcomes such as life consumption rates of components (i.e. expander blades), power, CCS
CO2 mass flow can be established and it is possible to evaluate, for any instant or for a
period, costs and revenues related to the plant behaviour on the basis of the values that prices
(electricity, CO2, coal, etc.), taxes and all the other costs assume along the time.
Pag. 202 of 202
5.5 Reference
[1] - G. Cerri et Al., H2-IGCC Milestone M4.1, “Selection of the best IGCC Cycle(s)
finished: Cycle options analysis”, H2-IGCC project, 29 July 2011.
[2] - G. Cerri et Al., H2-IGCC Milestone M4.1 Part 2, “Selection of the best IGCC Cycle(s)
finished: Cycle options analysis”, H2-IGCC project, 15 May 2013.
[3] Internal flow sheet, H2-IGCC project. Ensimm Calculation PDF including control loops:
SECTION 1000: ASU, SECTION 2000: Coal Milling and Drying; SECTION 3000:
Gasification; SECTION 4000: SOUR CO-Shift + COS-Hydrolysis; SECTION 5000:
H2S/CO2 Removal; SECTION 6000: Power Production.
[4] - G. Cerri et Al., H2-IGCC Milestone M5.8, “Selected Thermodynamic Optimized IGCC
Cycles”, H2-IGCC project, 31 May 2013.
[5] - “Cost and Performance Baseline for Fossil Energy Plants, Volume 1: Bituminous Coal
and Natural Gas to Electricity, Revision 2, November 2010”, DOE/NETL-2010/1397, 2010.
[6] - Nikolett Sipöcz, Mohammad Mansouri, Peter Breuhaus & Mohsen Assadi, “Plant
specification and detailed thermodynamic performance analysis of selected IGCC cycle”, H2-
IGCC Report, October 2010.
[7] - Mansouri, M., Breuhaus, P., Assadi, M., “Results from the thermodynamic simulations
and preliminary layout of best cycle option(s)”, H2-IGCC Report, October 2011.
[8] - G. Cerri et Al., H2-IGCC Deliverable D4.2.4, “Optimum Plant Operating Maps and
Control Policies – Part 1”, H2-IGCC project, 31 March 2014.
[9] – Domenichini R., Mancuso L., Ferrari N., Davison J., 2012: ‘Operating Flexibility of
Power Plants with Carbon Capture and Storage (CCS)’, Elsevier, Energy Procedia, 2012.