a parametric evaluation of fogging …diem1.ing.unibo.it/dottorato/macchine/melino.pdfperformance...
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ALMA MATER STUDIORUM UNIVERSITA’ DEGLI STUDI DI BOLOGNA
Facoltà di Ingegneria
Dottorato di Ricerca in Ingegneria delle
Macchine e dei Sistemi Energetici
XVII Ciclo, 2002 – 2004
A PARAMETRIC EVALUATION
OF FOGGING TECHNOLOGY
FOR GAS TURBINE
PERFORMANCE ENHANCEMENT
PhD Candidate: Supervisors:
Dott. Francesco Melino Chiar.mo Prof. Ing. Antonio Peretto
Chiar.mo Prof. Ing. Pier Ruggero Spina
Dr. Rakesh Bhargava
PhD Course Coordinator:
Chiar.mo Prof. Ing. Antonio Peretto
«Certo, niente ho appreso, se non partendo, e niente ho insegnato all’altro
se non invitandolo a lasciare il nido. Partire esige uno sradicamento…
Chi non si sposta non apprende niente… Nessun apprendimento evita il viaggio…
Apprendere dà inizio all’erranza». M. SERRES, Il mantello di Arlecchino.
Il ricercatore è un viaggiatore che s’incammina sulle terre altrui, nomade per
vocazione, bracconiere per necessità, pronto ad impadronirsi delle ricchezze che vede
per goderne e per farsi guidare – finalmente – sui terreni meno battuti, forse del tutto
inesplorati, luoghi dove può lasciare un segno e, col tempo, perfino una traccia.
Questa fede ultima sorregge l’esistenza individuale impegnata nel conoscere.
Il mio cammino è cominciato sui territori dei miei Maestri, luoghi dove il prendere
non è mai stato così grande come il loro desiderio di dare e dove ho viaggiato sicuro
lungo i confini.
Questo debito – già così grande – posso restituirlo solo facendolo crescere di più.
Al prof. Negri devo l’esempio di un modello di vita donato alla ricerca con passione
e coinvolgimento senza limiti, arricchito da una profonda saggezza educativa per i
giovani.
Al prof. Peretto sono grato per la guida e l’orientamento, la misura critica del lavoro
scientifico e la coerenza nel perseguire gli obiettivi.
Al prof. Bianchi sono debitore del metodo instancabile del fare e rifare, provare e
riprovare, del persistere con tenacia lungo un percorso mai pago dei punti acquisiti.
Al prof. Spina sono riconoscente per l’esempio continuo del lavoro collaborativo,
formativo e produttivo e – soprattutto – mai privo della simpatia e della benevola
indulgenza per gli iniziati.
Ho cominciato il viaggio da studente, oggi, ancora per via, mi sento allievo.
Ai miei Maestri
grazie
Acknowledgements
The present Ph.D. thesis was inspired and realized with the financial
and technical support of BP Amoco. In particular many thanks to his
representative, Dr. Steve Ingistov, without which this work should not be
realized. The Author wish to thank him for the continuous and precious
suggestions on the realization of this work.
Many thanks also to PhD Rakesh Bhargava, Supervisor in this work and
fundamental help in the scientific and technical support
I
INDEX SUMMARY Page No 1 1 LITERATURE REVIEW “ 3 1.1 Influence Of Ambient Conditions On Gas Turbine And
Combined Cycle Power Plants Performances “ 4
1.2 Compressor Inlet Air Cooling Systems “ 14
1.3 Fogging Technology For The Gas Turbine Enhancement “ 23
1.3.1 Climatic and psychrometric aspects of inlet fogging “ 23
1.4 Behavior And Dynamics Of Fog Droplets “ 27
1.4.1 Nozzle type “ 27
1.4.2 Heat and mass transfer of droplets: analytical model “ 28
1.4.3 Factors affecting droplets size: experimental results “ 31
1.4.4 Effects of the operating pressure and position of measurement in the plume on the droplet size “ 31
1.4.5 Effect of distance of measurement and flow rate on the droplet size “ 32
1.5 Effects Of Gas Turbine Inlet Configuration And Nozzle Characteristics On Fogging “ 34
1.5.1 Nozzle location “ 34
1.5.2 Nozzle orientation and fog distribution “ 39
1.5.3 Number of nozzles and nozzle pattern “ 41
1.6 Wet Compression “ 42
1.7 Fogging Users Practical Considerations “ 48
1.7.1 Water quantity requirements “ 48
1.7.2 Water quality requirements “ 48
1.7.3 Foreign object damage (FOD) “ 49
1.7.4 Gas turbine inlet icing “ 49
1.7.5 Duct drainage “ 49
1.7.6 Compressor surge “ 51
1.7.7 Compressor intake temperature uniformity “ 51
1.7.8 Intake duct considerations “ 52
II
1.7.9 Axial compressor fouling “ 52
1.7.10 Compressor blade erosion “ 52
1.7.11 Corrosion in the inlet duct “ 53
1.7.12 Compressor blading and coating distress Page No 54
1.7.13 Off-frequency operation of gas turbines “ 55
1.7.14 Electrostatic build-up and bearing distress with overspray “ 55
1.7.15 Split-shaft gas turbines for mechanical drive or power generation service “ 56
1.7.16 Considerations relating to engine cooling air flow with overspray “ 56
1.7.17 Practical considerations during system design and implementation “ 58
1.8 Activities Of Gas Turbine Manufacturers With Inlet Fogging And Overspray “ 60
1.8.1 General Electric “ 60
1.8.2 Siemens-Westinghouse “ 60
1.8.3 Rolls-Royce “ 60
1.8.4 Alstom “ 61
1.9 Inlet Fogging Field Experience And Users Perspectives “ 62
1.9.1 Inlet evaporative fogging on 80 MW class heavy-duty industrial gas turbine in a cogeneration facility “ 62
1.9.2 Application of inlet fogging overspray to a Frame 5 cogenerative installation “ 63
1.9.3 Keyspan Corp power generation plants “ 64
1.9.4 Testing of four GT24/GT26 gas turbines with overspray by Alstom “ 64
1.9.5 GE Frame 6B gas turbine at Cardinal cogeneration facility “ 64
1.10 Unresolved Issues And Ongoing Research “ 66 2 ANALYSIS OF COMMERCIAL COMBINED CYCLE
POWER PLANTS PERFORMANCE EQUIPPED WITH INLET FOGGING AND OVERSPRAY “ 67
2.1 Combined Cycle Power Plants Selection “ 68
2.1.1 The influence of ambient conditions on CCPP performance “ 72
2.2 Inlet Fogging And Overspray Steady Models “ 76
2.3 CCPP Performance With Inlet Evaporative And Overspray Fogging “ 81
2.4 Inlet Fogging Application On An Existing Combined Cycle Power Plant “ 88
III
2.4.1 Plant lay out “ 88
2.4.2 Plant simulation model “ 89
2.4.3 Site ambient conditions “ 90
2.4.4 Plant performance with and without fogging “ 94
2.5 Evaluation Of Italian Combined Cycle Power Plant Set Of Units With Inlet Fogging “ 101
2.5.1 Evaluation of Italian set of combined cycle units with inlet fogging “ 103
2.5.2 Climatic assumptions “ 104
2.5.3 Results “ 105 3 ANALYSIS OF THE PERFORMANCE OF A GAS
TURBINE EQUIPPED WITH INTERSTAGE INJECTION BY THE USE OF A COMMERCIAL SIMULATION PROGRAM “ 108
3.1 Aero-Thermodynamic Modeling Of A Gas Turbine “ 109
3.1.1 Water evaporation and wet compression considerations “ 111
3.2 Compressor Stage Performance Maps “ 113
3.3 Gas Turbine Performance With Interstage Water Injection “ 122
3.4 Droplet Temperature Influence “ 140 4 DEVELOPMENT OF A CALCULATION CODE FOR
THE PREDICTION OF A GAS TURBINE PERFORMANCE WITH INTERSTAGE INJECTION “ 142
4.1 Evaporation Of A Water Droplet In An Air Stream “ 143
4.2 The Diffusion Coefficient “ 150
4.3 Physical–Mathematical Model Of Wet Compression “ 153
4.3.1 Wet compression model by Härtel & Pfeiffer “ 153
4.3.2 Wet compression model by White & Meacock “ 155
4.3.3 Wet compression model by Horlock “ 155
4.4 Implementation Of A Wet Compression Calculation Procedure “ 157
IV
4.5 IN.FO. G.T. E. Calculation Code “ 170
4.5.1 COMP routine for compressor performance evaluation “ 173
4.5.2 CC routine for combustion chamber evaluation “ 179
4.5.3 TURB routine for turbine evaluation “ 180
4.6 Wet Compression Model Tuning “ 182
4.7 IN.FO. G.T. E Application On GE Frame 7 EA “ 190
4.7.1 Frame 7EA Input Definition “ 190
4.7.2 Evaluation Of Frame 7EA With Water Injection By Using The IN.FO G.T. E. Calculation Code “ 199
4.7.3 Comparison Between IN.FO. G.T. E. And Commercial Software Package “ 205
5 CONCLUDING REMARKS Page No 212 6 REFERENCES “ 215 7 LIST OF FIGURES “ 222 8 LIST OF TABLES “ 234
1
Summary
In recent years, deregulation in the power generation market worldwide combined with significant variation in fuel prices and a need for flexibility in terms of power augmentation specially during periods of high electricity demand, has forced electric utilities, cogenerators and independent power producers to explore new power generation enhancement technologies.
Moreover, the strong influence of ambient temperature on gas turbine and combined cycle performance increases this critical problem. On this regard, it could be assumed that, for a gas turbine, the produced power output drops from approximately 0.50% to 0.90% for every 1ºC of ambient air temperature rise. For what concern the combined cycle power plants, instead, air ambient temperature effects the topping section (as previously explained) but also the bottoming section and in particular the efficiency of heat recovery steam generator (due to the changed conditions of exhaust gasses from the gas turbine) and the performance of the cooling section (depending on the type of adopted condenser system).
There are several strategies, known as gas turbine compressor inlet air cooling systems, to contrast this problem. The available inlet air cooling methods could be divided into two main categories: (i) “continuous cooling systems” in which the air cooling is realized by placing an heat exchanger into the compressor inlet duct (without direct contact between the incoming air and the cooling fluid) and (ii) evaporative cooling systems in which the air cooling is realized by water evaporation (by putting in contact the compressor inlet air and the cooling fluid).
It is not possible a choice among all the available inlet cooling techniques without taking into account many parameters such as historical climatic conditions profiles, air
2
flow to gas turbine output ratio, number of hours per day for needed power enhancement, cost of cooling system installation and electricity.
Nevertheless, during the last 10 ÷ 15 years, fogging approach has gained more and more attention and then a large-scale applications in particular due to its effectiveness and low initial and operative costs.
Although this technique is already adopted in a lot of installation worldwide (more than 1000 power plants) many aspects are not yet clear and need further investigations. In particular the “wet compression” strategy (in both overspray and interstage injection approach) has not yet fully studied and understood.
For the previously reasons the aim of the present work is to study the different aspects fogging (inlet fogging, overspray and interstage injection) under different points of view, both theoretical and practical, in its gas turbine and combined cycle application.
In the first chapter of the present work, (“Literature Review”), the influence of ambient conditions on the performance of both gas turbine and combined cycle power plants, together with the description of the most developed inlet air cooling systems is presented. Moreover a detailed and critical review of the current “fogging state of the art” is presented and discussed. This analysis highlights the fogging technology main aspects such as the behavior and the dynamic of fog droplets in the air stream, the influence of nozzle configuration and characteristics, practical considerations, the activities of gas turbine manufactures with inlet fogging and overspray and the field experience and users perspectives. This review has allowed to highlight that many topics related to fogging approach are not yet fully studied; among the others, the focus of this work is dedicated to the understanding of the change in performance of commercial gas turbine and combined cycle power plant units due to the fogging implementation and to the detailed study of the interstage injection by the use of wet compression analytical models.
More in details, in the second chapter (“Analysis Of Commercial Combined Cycle Power Plants Performance Equipped With Inlet Fogging And Overspray”) a parametric analysis of a large number of commercial plants with a power output ranging from 8 to 380 MW with and without inlet fogging and overspray (from 0.5 to 2%) is developed to understand the influence of this cooling technology on the plant performance (power output, electrical efficiency, air mass flow rate at compressor inlet, gas turbine exhaust gas mass flow rate and temperature, etc.).
The other issue investigated in this work is the analysis of performance for a gas turbine with interstage injection. This study is broken into two main parts. In the third chapter, (“Analysis Of The Performance Of A Gas Turbine Equipped With Interstage Injection By The Use Of A Commercial Simulation Program”) the investigation is carried out by the use of a commercial software for energy systems simulations while in the fourth chapter (“Development Of A Calculation Code For The Prediction Of A Gas Turbine Performance With Interstage Injection”) a calculation code (named IN. FO. G.T.E, INterstage FOgging Gas Turbine Evaluation) is developed and applied.
In fact, the use of commercial software allowed to understand the main topics related to the wet compression (such as the influence of the injection location or of the injected water mass flow rate on compressor characteristics and gas turbine performance), but its rigid structure showed some limitations in the understanding of water droplets evaporation during a compression process taking into account it as a dynamic process.
The IN.FO. G. T. E calculation code developed by DIEM – University of Bologna – overcomes the previously explained limitations due to its flexibility and to the use of evaporation models able to consider the dynamic aspects of this process.
3
1
Literature Review
The continuous growing request of power coupled with the need for flexibility has forced in the last years cogenerators and power producers to search for new techniques for power enhancement. Also the loss in performance of gas turbine and combined cycle power plants with the increase of ambient temperature becomes a key problem if is considered that nowadays the peaks of electricity demand often occur during the noon or during the summer months (this specially due to the heavy diffusion of home conditioning systems).
Obviously, every strategy able to recover the loss in power output and efficiency due to not favorable climatic conditions becomes a precious resource.
The previous sentences clarify the reason according to which during the last years the research focused its effort on the inlet air cooling methods.
These technologies allow to cool the air at gas turbine compressor inlet with the goal of an improvement in plant net output. There are several inlet air cooling systems, but in the last years fogging has gained more and more attention for its effectiveness and simplicity in its realization.
In this chapter the influence of ambient conditions, with particular attention to air temperature, on gas turbine and combined cycle power plants is analyzed under a theoretical point of view. Moreover, a review on the “state of the art” of inlet air cooling systems and a critical analysis about the actual level of fogging technology is carried out. This review, puts in evidence the main aspects of fogging approach highlighting both the practical and the theoretical aspects with also the yet unresolved issues of this system.
4
1.1 – INFLUENCE OF AMBIENT CONDITIONS ON GAS TURBINE AND COMBINED CYCLE POWER PLANTS PERFORMANCES
Ambient conditions (temperature, pressure and relative humidity) have a strong
influence on gas turbine performances in both single or combined cycle applications. Ambient temperature, pressure and relative humidity variations have a different
weigh on gas turbine and combined cycle efficiency and power output changes. The main influence on gas turbines or combined cycles performances is played by
ambient temperature; ambient pressure effects could be considered less important while relative humidity variations are almost negligible. In fact, considering a gas turbine cycle, while it is possible to estimate a power loss of about 0.50 ÷ 0.90 % for every 1°C of ambient temperature rise [1], for what concern ambient pressure, the loss in output due to its fluctuations, is about 0.10 % of ISO value for every 1 mbar of pressure decrease [2]. Similarly it could be estimated that for a variation of 1 percentage point in relative humidity, a change less than 0.01% in produced power output is expected. This last evidence confirm the insensibility of gas turbine cycle to relative humidity change. Although this, relative humidity will became a key parameter for gas turbine inlet air cooling in the case of evaporative methods (as largely explained on the follow).
In Figs. 1 and 2 the influence of ambient temperature on the produced power output and on the heat rate (as ratios to the ISO values) for an heavy duty gas turbine (GE 9351 FA) and for an aero-derivative gas turbine (GE LM-6000 PD) respectively, is presented. The trends in Figs. 1 and 2 are evaluated with a commercial program for energy system simulations [3] and clearly show higher sensitivity to ambient temperatures in case of aero-derivative gas turbines than heavy duty ones.
In the same way Figs. 3 and 4 put in evidence the influence of ambient pressure on the previous gas turbine models while Figs. 5 and 6 present the power output and heat rate changes due to relative humidity fluctuations. The range of the Y axis of Figs. from 1 to 6 is the same to better put in evidence the different weigh of ambient temperature, pressure and relative humidity on gas turbine performances.
0.70
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
10 15 20 25 30 35 40
PTG
(T)
PTG
(ISO)
Air Temperature [°C]
GE 9351 FA
LM 6000 PD
Figure 1 – Ambient temperature influence on gas turbine power output (@ 1.013 bar and 60% RH)
5
0.70
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
10 15 20 25 30 35 40Air Temperature [°C]
GE 9351 FA
LM 6000 PD
HRTG
(T)
HRTG
(ISO)
Figure 2 – Ambient temperature influence on gas turbine heat rate (@ 1.013 bar and 60% RH)
0.70
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
0.90 0.92 0.94 0.96 0.98 1.00 1.02Ambient Pressure [bar]
GE 9351 FA
LM 6000 PD
PTG
(T)
PTG
(ISO)
Figure 3 – Ambient pressure influence on gas turbine power output (@ 15 °C and 60% RH)
6
0.70
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
0.90 0.92 0.94 0.96 0.98 1.00 1.02Ambient Pressure [bar]
GE 9351 FA
LM 6000 PD
HRTG
(T)
HRTG
(ISO)
Figure 4 – Ambient pressure influence on gas turbine heat rate (@ 15 °C and 60% RH)
0.70
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
10 20 30 40 50 60 70 80 90 100Relative Humidity [%]
GE 9351 FA
LM 6000 PD
PTG
(T)
PTG
(ISO)
Figure 5 – Ambient relative humidity influence on gas turbine power output (@ 15 °C and 1.013 bar)
7
0.70
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
10 20 30 40 50 60 70 80 90 100Relative Humidity [%]
GE 9351 FA
LM 6000 PD
HRTG
(T)
HRTG
(ISO)
Figure 6 – Ambient relative humidity influence on gas turbine heat rate (@ 15 °C and 1.013 bar)
Moreover it should be observed that ambient pressure or relative humidity variations
are not linked to particular sites and that they do not much depend on the season (as it occurs for ambient temperature). For all these reasons their influence could be considered negligible on the respect of ambient temperature.
To better understand the influence of ambient temperature on gas turbine power output it could be considered its effect on the compression process. In fact a gas turbine compressor could absorbs more than the 60% of the total work produced by the gas turbine and so any strategy for reducing the work of compression will enhance the gas turbine power output [4].
The compressor specific work is given by:
⎥⎥⎥
⎦
⎤
⎢⎢⎢
⎣
⎡
−β=−= η⋅
−
1Tc)TT(cL polC
1k
1k
1P12PC (1)
For Eq. 1 is evident that increasing T1 (considering β as constant), compression
work would increase. If in the cycle no pressure losses and a constant specific heat are assumed, gas
turbine thermodynamic efficiency can be reduced to the equation below:
8
)TT(
1T11T
)TT(cLL
polC
polC
polE
1k
1k
13
1k
1k
1
k1k3
23P
CTTH
η⋅
−
η⋅
−
η−
β⋅−
⎥⎥⎥
⎦
⎤
⎢⎢⎢
⎣
⎡
−β−
⎥⎥⎥⎥
⎦
⎤
⎢⎢⎢⎢
⎣
⎡
β
−
=−
−=η (2)
Where:
β = compressor pressure ratio k = cp/cv = specific heat ratio ηpolE = turbine polytropic efficiency ηpolC = compressor polytropic efficiency T3 = turbine inlet temperature T2 = compressor discharge temperature T1 = compressor inlet temperature LT = turbine specific work LC = compressor specific work
As evident from Eq. 2 an increase in ambient temperature reduces the gas turbine
thermodynamic efficiency, by increasing compression specific work. The trend of gas turbine thermodynamic efficiency as function of ambient temperature evaluated by using the Eq. 2 in which was assumed T3=1400 °C, ηpolC = ηpolE= 0.9 and a constant isentropic coefficient k= 1.4, for different values of the pressure ratios ranging from 8 to 20, is presented in Fig. 7.
36
38
40
42
44
46
48
50
0 5 10 15 20 25 30 35 40
gas t
urbi
ne th
erm
odyn
amic
eff
icie
ncy
[%]
Air Temperature [°C]
β=201816
14
12
10
8
Figure 7 – Gas turbine thermodynamic efficiency as function of compressor pressure ratio and ambient temperature
9
Moreover a variation in ambient temperature determinates also a change in the operating point of both compressor (Fig. 8) and turbine (Fig. 9) performance maps, depending on the control system adopted for the gas turbine; for this evidence a variation in polytropic efficiency of compression and expansion process with also a reduction of gas turbine pressure ratio (β) are to be expected. The reduction in gas turbine thermodynamic efficiency due to high ambient temperature is lightly compensated by the increase of compressor discharge temperature as its polytropic efficiency drops.
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1. 4
1.11.0
0.90.8
0.70.6
0.50.40.3
pres
sure
ratio
/ pr
essu
re ra
tio d
esig
n
corrected flow / corrected flow design
corrected speed
corrected speed design = 1.2
ISO DAY
HOT DAY
Figure 8 – Operating point change passing from ISO day to an HOT day in a typical compressor performance map
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.85 0.90 0.95 1.00 1.05 1.10 1.15
pres
sure
ratio
/ pr
essu
re ra
tio d
esig
n
corrected flow / corrected flow design
corrected speed
corrected speed design = 1.1
1.0 0.9 0.8 0.7 0.6 0.5 0.4
ISO DAY
HOT DAY
Figure 9 – Operating point change passing from ISO day to an HOT day in a typical turbine performance map
10
On the basis of the previous consideration is possible to draw (Fig. 10) a qualitative T-s diagram to compare a gas turbine cycle operating at ISO ambient conditions with one at high ambient temperature. The drop in pressure ratio that occurs on the hot day is clear represented. The cycle peak temperature is limited and so the contemporary expansion ratio drop means that a less work is extracted from the turbine.
4.80
5.00
5.20
5.40
5.60
5.80
6.00
6.20
0.80 1.00 1.20 1.40 1.60 1.80 2.00 2.20
constant firingtemperature
ISODAY
HOTDAY
Tem
pera
ture
entropy
ambientpressure
Figure 10 – Entropy diagram on a hot day compared to one on ISO ambient conditions
Another fundamental aspect is the reduction in gas turbine air mass flow rate that
occurs in correspondence of high ambient temperatures. A gas turbine operating at a constant speed could be considered as a constant
volumetric machine, if no change in geometry are assumed (that means no Variable Inlet Guide Vane), the air mass flow rate at compressor inlet could be expressed as follow:
Vm airair &⋅ρ= (3)
where V& is the volumetric flow rate at compressor inlet. The density of the air at compressor inlet is given by:
1
1air RT
p=ρ (4)
where: p1 =compressor inlet pressure R = gas constant T1 = compressor inlet temperature
airρ = air density at compressor inlet
11
In Fig. 11 is presented the trend of the air density as function of ambient temperature for an ambient pressure of 1.013 bar and a relative humidity of 60%.
1.10
1.15
1.20
1.25
1.30
0 5 10 15 20 25 30 35 40
air d
ensi
ty [k
g/m
3 ]
Air Temperature [°C] Figure 11 – Air density as function of ambient temperature (@ 1.013 bar and 60% of relative humidity)
The gas turbine produced power is
( )CTairTG LLmP −= & (5) The combination of Eqs. 3 and 4 with 5 gives:
( )⎥⎥⎥⎥
⎦
⎤
⎢⎢⎢⎢
⎣
⎡
⎟⎟⎟
⎠
⎞
⎜⎜⎜
⎝
⎛
−β−
⎟⎟⎟⎟
⎠
⎞
⎜⎜⎜⎜
⎝
⎛
β
−⋅⋅=−⋅⋅= η⋅
−
η−
1T11TVRTpLLV
RTpP polC
polE
1k
1k
1
k1k3
1
1CT
1
1TG && (6)
The Eq. 6 shows that the reduction in gas turbine produced power is due to increase
of compression specific work, decrease of turbine specific work and also to the term 11 RT/p that is representative of the drop in compressor inlet mass flow rate.
If is supposed T3 and ηpolC and ηpolE as constant and is also supposed that no change in compressor pressure ratio occurs with ambient temperature variation, the ratio between the gas turbine power corresponding to a certain ambient temperature and the value at ISO conditions could be written as:
DTCT1
)ISO(P)T(P
1
1
TG
1TG⋅
⋅−= (7)
12
Where the terms C and D are constant for a datum compressor pressure ratio (β) and result as:
⎟⎟⎟⎟
⎠
⎞
⎜⎜⎜⎜
⎝
⎛
β
−
−β=
η−
η⋅
−
polE
polC
k1k3
1k
1k
11T
1C C
T1DISO
−=
The Eq. 7 puts in evidence the influence of ambient temperature on gas turbine
power output, supposing β as constant. In Fig. 12, the power ratio of Eq. 7 evaluated for different gas turbine pressure ratios
and for T3=1400 °C, ηpolC = ηpolE= 0.9 and a constant isentropic coefficient k= 1.4, is presented.
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
0 5 10 15 20 25 30 35 40Air Temperature [°C]
β=32
2416
8
PTG
(T)
PTG
(ISO)
Figure 12 – Gas turbine power ratio as function of compressor pressure ratio and ambient temperature
From the Fig. 12 it could be observed that, others factors being equal, in
correspondence of higher values of the compressor pressure ratio, the loss in gas turbine produced power output is greater. This last evidence partially explains why aeroderivative gas turbines (which are characterized by higher values of compressor pressure ratio) are more sensitive to the ambient temperature fluctuations on the respect of heavy duty machines, even if this distinction is actually not so marked due to the introduction of generation of heavy duty gas turbine with high turbine inlet temperature and pressure ratios.
The last considerations change in the case of gas turbine with two or more shafts. In fact, in aeroderivative machines for power generation, the produced power is controlled modifying fuel flow which consequently determinates a gas generator speed variation and this, as a result, affects the air flow rate. In this case the gas generator speed is free
13
to move to the level required to attain the desired power subject to the limits on maximum speed or maximum turbine inlet temperature. In such machines, there is a change in volumetric flow rate according to the mechanical speed of gas generator. The power turbine speed is fixed because it is coupled with generator. Instead in mechanical drive applications, both gas generator and power turbine speeds can change in accordance to the matching laws.
In the case of combined cycle power plants the effect of ambient temperature is evident also on the bottomer section. In particular, a reduction of thermal power discharged by gas turbine into the heat recovery steam generator occurs. This reduction is due to decrease in gas turbine exhaust gasses and is only lightly compensate by turbine outlet temperature increase. As consequence, a minor steam mass flow rate is produced that means a drop in steam turbine produced power. In addition, it should be noted that ambient temperature will effect also the cooling section, depending on the adopted condenser system.
A more detailed analysis about the influences of ambient temperature on combined cycle power plants is developed in the second chapter of this work.
In the following paragraphs the main gas turbine compressor inlet air cooling systems are presented and discussed.
14
1.2 – COMPRESSOR INLET AIR COOLING SYSTEMS There are several strategies to contrast the degradation of gas turbine performances
due to the high ambient temperatures [5 – 9]. In Fig. 13 the most common compressor inlet air cooling strategies are summarized. From the figure, it could be observed that the possible technologies are divided into two main categories: “continuous cooling systems” if no direct contact between incoming air (entering into the compressor) and refrigerant fluid occurs, and “evaporative cooling systems” in which the cooling effect is obtained by mixing the air with the refrigerant fluid.
In particular:
continuous cooling systems: the air cooling is realized by placing an heat exchanger upstream of the compressor in the inlet duct. Water or others refrigerant fluids could be the cold source. In this last case a refrigerant plant is required. The most common realizations of these systems are with compression or absorption refrigerant plant or with thermal energy storage.
evaporative cooling systems: the principle of these systems is the air cooling by
water evaporation. There are two main strategies for evaporative cooling systems on the basis of the method to put air and water in contact: traditional systems in which the air is forced through a wetted honeycomb (placed in the compressor inlet duct) and fogging that use a spray system. Ibrid systems, instead, combine the concept of evaporative cooling with other type of heat exchanger (usually rotary heat exchanger).
TRADITIONAL SYSTEMS
IBRID SYSTEMS
EVAPORATIVE COOLING SYSTEMS
CONTINUOUS COOLING SYSTEMS
FOGGING
HIGH PRESSURE FOGGING
OVERSPRAY INTERSTAGE FOGGING
CLOSED CYCLE
COMPRESSION PLANT
ABSORPTION PLANT
ICE STORAGE
WATER STORAGE
THERMAL STORAGE
REFRIGERANT PLANT
INLET AIR COOLING SYSTEMS
OPEN CYCLE
Figure 13 – Compressor inlet air cooling strategies
15
More in detail it could be observed that: Continuous cooling with open cycle: in this system water from an external source
(river, lake or sea) for the air cooling is adopted by placing an heat exchange in the compressor inlet duct. The main advantage of this strategy is the simplicity of the realization, but it should be noted that the cooling potential of this system depends on the temperature of the water from the external source and then on the climatic conditions of the sites in which it is adopted. Moreover the use of an heat exchange introduces an additional term of pressure losses into the compressor inlet duct.
Continuous cooling with compression refrigerant plant: this method uses an heat
exchanger, circulating a refrigerant fluid, placed in compressor inlet duct. In this plants a refrigerant compressor, an evaporative condenser and other auxiliaries need to be installed. Although other refrigerants can be used, ammonia is recommended to avoid formation of acids in the combustor. The optimum system is an ammonia liquid overfeed system utilizing screw compressor(s) and evaporative condenser(s). A liquid overfeed is utilized to increase the heat transfer coefficient by 25 ÷ 30%. Multi pressure ammonia can be used with multi stage coils to reduce the compressor power consumption. The compressor can be driven by electric power or gas. A gas fired unit results in less power consumption than an electrically driven compressor, though power is still required for refrigerant pumps, oil pumps and other auxiliaries. For a gas fired unit, there is an increase in net power enhancement but the equipment cost is higher, and it requires more floor space and maintenance. Gas driven compressors are recommended when the savings in auxiliary power consumptions absolutely needed. An alarm system can be used with a direct refrigeration system to detect refrigerant leakage. If refrigerant were to leak from the coils, it immediately flashes into vapor because of a reduction in the pressure. As soon as a leak is detected, the control system should shut the system off and pump the refrigerant out of the cooling coils with the compressors. The pumped refrigerant is stored in the low and high pressure receivers. The use of electrical chillers has been successfully applied to provide continuous cooling for several installations (Fig. 14).
M
Figure 14 – Continuous cooling system with compression refrigerant plant
16
Continuous cooling with absorption refrigerant plant: this system requires an heat source for the regenerator (usually gas turbine exhaust gas or steam when available). Absorption refrigeration, especially that using lithium bromide, has been successfully applied to provide cooling on many aeroderivative gas turbine installations for a number of years. The system can cool the inlet air to about 10 °C (50 ºF). The heat source for the regenerator can be gas, steam, hot water or turbine exhaust gases. Although the heat source spares the use of electrical power, it is still needed to drive the pumps, condenser fans and other auxiliaries.
The absorption units (Fig. 15) can be single or double effect. A single effect unit has a coefficient of performance (COP) of 0.7÷0.9 whereas a double effect unit has a COP of about 1.15. A single effect steam absorption system requires about 1 ÷ 1.4 bar steam and a double effect steam absorption unit requires a pressure of 7.9 ÷ 8.3 bar. Lower steam pressures can be used for double effect units, but they reduce the performance of the system. For example, a reduction in steam pressure to 4.8 bar reduces the chiller performance by 20%. A double effect steam absorption unit has about half the steam requirement of a single effect system.
For combustion turbines running in simple cycle mode, a significant amount of energy is wasted to the atmosphere with the exhaust gases. It would be beneficial if this energy could be utilized for some useful purpose. The turbine exhaust gases could be directly used as a heat source for the regenerator or they could be used to produce steam or hot water with the help of a heat exchanger at the turbine exhaust. Absorption system manufacturers normally recommend exhaust gas temperatures of 650ºC or more for its successful operation. However, care should be taken when steam or hot water usage requires the addition of a heat exchanger or heat transfer sections in the HRSG, because additional pressure drop through these sections is realized. This pressure drop can be significant in the case of combined cycle units when the sections are installed as retrofits, since the complete plant design is not optimized. For absorbers using steam or hot water, a means must be provided to use the steam/hot water when the absorbers are running on part load conditions. For example, if steam produced in an HRSG is used to run the generator for a lithium bromide absorption system, not all the available steam is utilized if the cooling load is less than the design conditions . In that case, either the steam may be bypassed to the condenser, vented to the atmosphere or used for some other applications.
QF
Qi
CK
E
P1
QSK’
QSK’’ a
b
a
c
A
TCu
TCi TFi
TFu
P
Figure 15 – Continuous cooling system with absorption refrigerant plant
17
Continuous cooling with thermal energy storage (TES): this cooling technique builds and stores a cold reserve (ice or cold water) during non peak hours and uses it to cool inlet air and enhance power production during peak hours. For this characteristic TES is economically useful only for few hours of operation a day and specially if the difference between the economic value of peak hours electric power and non peak ones is high. Although a relevant temperature reduction can be realized during certain period the average power output improvements are worse as compared to all the other systems. It also should be considered that the investment cost and plant complication is enormous is comparison with fogging.
Fig. 16 shows a typical thermal energy storage system using cold water, while Fig. 17 a plant for the ice production. Ice is often an optimum storage media because its energy holding capacity is several times that of other commonly used fluids for this application. Chilled water and other fluids, as an alternative to ice, can be considered. However, these systems may require seven times the storage capacity of ice and therefore limit their use for high cooling load applications.
COLD WATER STORAGE TANK
Refrigerant power plant
Figure 16 – Thermal energy storage with cold water
COLD WATER STORAGE TANK
receiver
Evaporativecondenser
Suction accumulator
HEAT
Make up
Blow down
Figure 17 – Thermal energy storage with ice
18
Traditional evaporative cooler systems: this system uses a wetted honeycomb media for water evaporation. The temperature drop that is possible to realize is function of the equipment design on one hand, and of ambient conditions on the other hand. With reference to Fig. 18, the effectiveness of the cooler, is given by:
wba
aca
TTTT
E−−
=
0
5
10
15
20
25
30
5 10 15 20 25 30 35 40 45 50Dry Bulb Temperature [°C]
a
wb
30
25
20
1510
Wet Bulb Temperature [°C]
[g H
2O/k
g dryA
IR]
ac
Figure 18 – Psychrometric chart for a traditional evaporative cooling system
A typical value for E is around 85 ÷ 90%. This means that it is not possible to reach
the wet bulb temperature. This value depends on the surface area of water exposed to the air stream and on the residence time.
Assuming E = 85% temperature drop ( TΔ ) could be expressed as:
( ) ( )WB1DB1WB1DB1DB2DB1 TT85.0TTETTT −⋅=−⋅=−=Δ considering, as example, ambient conditions equal to 40 °C of air temperature and
40% of relative humidity, it results a corresponding value of wet bulb temperature of about 28 °C which mean a maximum available cooling equal to 12 °C. With the previous hypothesis the total air cooling achievable with a traditional evaporative systems is equal to 0.85 × 12 = 10.2 °C.
19
Evaporative hybrid systems: these systems use the typical treatments of the air conditioning plant. In Fig. 19 a scheme of an hybrid evaporative air cooling system is presented while in Fig. 20 the corresponding humid air transformations on a psychrometric chart are presented. It could be observed from these figures that to increase the effect of the air cooling by adiabatic saturation first an adiabatic dehumidification (from point 1 to point 2 in Fig 20) is realized by using a rotative wheel (which is regenerated by the exhaust gasses from the turbine), an air cooling (from point 2 to point 3 in Fig 20) in which there is no change in the water amount in the air and then an adiabatic saturation of the air (from point 3 to point 4 in Fig 20).
1 2 3 4Ambient
air Evaporative cooler
Evaporative cooler
Water
To turbine inlet
From turbine exhaust
Rotary desiccant
bed
Rotary heat
exchanger
Figure 19 – Evaporative cooling hybrid system scheme
0
5
10
15
20
25
30
5 10 15 20 25 30 35 40 45 50Dry Bulb Temperature [°C]
1
3
30
25
20
1510
Wet Bulb Temperature [°C]
[g H
2O/k
g dryA
IR]
2
4
Figure 20 – Psychrometric chart for an evaporative cooling hybrid system
20
Fogging: is a method where demineralized water is converted into a fog by means of special atomizing nozzles operating at high pressure (from 70 to 200 bar). The cooling effect is provided by water evaporation. This means that an adiabatic saturation of inlet air mass flow rate occurs in the gas turbine inlet duct. In figure 21 is presented a psychrometric chart showing the cooling potential of fogging approach. The effect of the adiabatic saturation is to cool the air from the dry bulb temperature (point a in Fig. 21) to the wet bulb temperature (point wb in Fig. 21). This means a value of coling efficiency (E) close to 100%.
0
5
10
15
20
25
30
5 10 15 20 25 30 35 40 45 50Dry Bulb Temperature [°C]
a
wb
30
25
20
1510
Wet Bulb Temperature [°C]
[g H
2O/k
g dryA
IR]
Figure 21 – Psychrometric chart (@ p=1.01325 bar)
According with the injected water amount and injection location, three different
fogging strategies could be distinguished:
high pressure fogging (evaporative fogging): in this case the amount of injected water into the compressor inlet duct is strictly necessary for air saturation. With this strategy, water evaporation is completed before the air enters into the compressor;
overspray fogging (wet compression): the total injected water is more than the
amount requested for air saturation. As evident, a percentage of water in a liquid phase (usually less than 2% of saturated air flow) will enter into the compressor where the evaporation will end;
fog intercooling (interstage injection): with this method, the water injection is
realized (exclusively or in combination with one of the previous techniques) through the compressor stator blades. The effect of water injection is quite similar to the one of a traditional intercooling compression.
21
The choice among all the available inlet cooling techniques is not simple because many parameters such as air temperature, relative humidity, air flow to gas turbine output ratio or number of hours per day for needed power enhancement are to be considered [10, 11].
On the basis of the principal literature available on the matter [8 – 20] the Tab. 1 that summarized for all the cooling technologies the principal advantage and the corresponding disadvantages, was compiled.
Among the all inlet-cooling technologies available, inlet fogging (both high pressure fogging and overspray) has over the past few years, seen large-scale applications because of the advantage of low first-cost compared to the other techniques including media evaporative cooling and refrigeration technologies.
Coupled with a need for power augmentation is the requirement to attain incremental power with minimal capital investment. High pressure inlet fogging fits this niche of a low-cost and effective power augmentation technology.
In fact, the fogging approach resulted in a minimal impact on EPC (Engineering, Procurement and Construction) cost. Additionally, inlet fogging was the only option that provided a small improvement in heat rate, while all the other options worsened the heat rate value. The highest return on equity was obtained by combining inlet fogging and supplemental firing of the heat recovery steam generator (HRSG). In recent years, several combined cycle plants utilizing advanced technology gas turbines and peaking power units have adopted fogging as a power augmentation strategy. It is estimated that there are over 1000 gas turbines that use inlet fogging or overspray worldwide.
Jones and Jacobs [21] have recently presented a detailed informative study on the available options for enhancing combined cycle performance. Their study also included an economic assessment of performance enhancement alternatives by considering a combined cycle plant consisting of two GE PG7241(FA) gas turbines, two unfired three pressure-level HRSGs, and one GE D11 reheat condensing steam turbine with wet cooling tower. Among various alternatives examined, inlet fogging was found least sensitive to variations in the economic parameters considered due to its insignificant impact on non-peak period plant performance combined with low initial investment and modest gain in incremental peak-period power generation.
For the previous reasons, in the following paragraphs of this chapter, the main aspects of fogging are presented and discussed.
22
Table 1 – Inlet air cooling technologies main characteristics Technology Advantages Disadvantages
Open cycle continuous
cooling - relatively easy to realize
- water source required - additional pressure drop - cooling potential depends on climatic
conditions
Continuous cooling with compression
refrigerant plant
- provides an instantaneous cooling - air could be cooled lower than wet
bulb temperature - particularly effective when required
power enhanced time is long (more than 6÷8 hours/day)
- not much floor space is required using standard package chillers
- refrigerant fluid is required - high electrical power consumption
(from the almost the same to a third more than mechanical chilling)
- additional compressor inlet pressure drop
- refrigerant leakage possible danger
Continuous cooling with
absorption refrigerant
plant
- provides an instantaneous cooling - air could be cooled lower than wet
bulb temperature - particularly effective when required
power enhanced time is long (more than 6÷8 hours/day)
- heat source (gas or steam etc.) is required
- very hot gas (more than 650 °C) for a successfully application is needed (if heat source is gas)
- need a steam cycle (if heat source is steam)
- power consumption - additional compressor inlet pressure
drop
Continuous cooling with
Thermal energy Storage
- particularly effective when power enhanced is required for few hours a day
- relevant temperature reduction - particularly effective when the
difference in economic value between peak and non-peak hours is high
- complex installation - energy consumption for cold source
maintenance is required - additional compressor inlet pressure
drop
Conventional evaporative
cooling
- relatively inexpensive to install - easy to control
- is not possible to reach 100 % of relative humidity
- inadequate for application in high humidity regions
- demineralized water consumption - additional compressor inlet pressure
drop
Hybrid systems - additional cooling on the respect of the wet bulb temperature
- extremely complex - water or other refrigerant fluid is
required -quite expensive
High pressure fogging
- quite inexpensive - minimum plant modifications
required - minimum additional compressor inlet
pressure drop
- demineralized water consumption - impossible to cool air lower than wet
bulb temperature -not much effective for high humidity
regions application
Overspray fogging
- quite inexpensive - minimum plant modifications
required - minimum additional compressor inlet
pressure drop - additional power boost respect to inlet
fogging
- demineralized water consumption - impossible to cool air lower than wet
bulb temperature -not much effective for high humidity
regions application - possible compressor blades erosion
Interstage injection - compressor cleanness and efficiency
- compressor modification required - hard to realize as retrofit option - possible compressor blades erosion
danger
23
1.3 – FOGGING TECHNOLOGY FOR THE GAS TURBINE ENHANCEMENT Inlet fogging is a method of air cooling where demineralized water is converted into
fog droplets by means of specially designed atomizing nozzles operating at high pressure, 70 to 200 bars. This fog provides cooling when it evaporates in the inlet air duct of a gas turbine. This technique allows achieving close to 100% evaporative cooling effectiveness in terms of attaining the wet-bulb temperature at the compressor inlet.
A typical high pressure fogging system consists of series of reciprocating pumps providing demineralized water at high pressure to an array of fogging nozzles located typically downstream of the inlet air filter elements. High pressure water is required for the fogging nozzles because droplet size is proportional, to a limit, to the (applied pressure)a. Where, value of exponent “a” varies between –0.5 and –0.2 depending on the nozzle type, nozzle geometry, liquid characteristics and droplet diameter definition [22]. These nozzles create a large number of small droplets of varying size (less than 50 microns of diameter), which evaporate as they flow through the gas turbine intake system. A large gas turbine can have an array of several hundred nozzles. For high pressure fogging, demineralized water is necessary to minimize the potential for compressor blades fouling and hot gas path corrosion that can result from minerals naturally present in untreated waters.
In order to realize full potential of the fogging technology, understanding the physics of droplet sizing, the behavior of droplets in the inlet duct or inside the compressor (in the case of overspray or interstage injection) and the locations and types of nozzles.
Before a summary on the understanding of droplet dynamics and related topics is presented, it would be useful to discuss climatic and psychrometric aspects of inlet fogging.
1.3.1 – Climatic And Psychrometric Aspects of Inlet Fogging The main obstacle faced by gas turbine users who wish to evaluate the benefits of
fogging is the unavailability of easy-to-use climatic data. The obstacle may be broken into two factors:
1.operators can not easily locate the appropriate weather data for their site. Much
of the available data at a plant site may be based on average data points with no representation of the values of coincident dry and wet bulb temperatures. This type of data is invaluable when evaluating any evaporative cooling solution;
2.even when some appropriate data is available, additional data processing is essential before any meaningful estimate can be made of cooling potential.
Modeling Of Climatic Data. There are numerous problems when modeling climatic
data, several of which derive from the concept of “averaging” of the data. Averaging of the data can underestimate cooling potential because the highest relative humidity conditions do not occur coincident with high dry-bulb temperatures. Relative humidity has a marked systematic diurnal variation opposite to the temperature. The moisture holding capacity of air depends on its temperature. Warmer air can hold more moisture than a cooler air. Consequently, relative humidity is highest during the cool morning and evening hours and lowest in the hot afternoon hours. Actual data taken from a site, as shown in Fig. 22, clearly identifies the spread, between the dry bulb and wet bulb
24
temperatures, which determines evaporative cooling potential. Equivalent cooling degree hours (ECDH) is an important term used in understanding the climatic data. Detailed climatic analyses of several hundred sites within the US [23] and around the world [24] were recently performed by Chaker et al. using the concept of ECDH that showed considerable amount of evaporative cooling potentials throughout the world, even in very humid climates that would not be traditionally considered as possible sites for evaporative cooling. There are many fog systems currently in use in Southeast Asia and other hot and humid climates around the world.
Two sources of weather data are available; the first one is commercial and can be obtained for example from the National Climatic Data Center (NCDC), and the second is collected directly from sites. The ECDH gives information about the Wet Bulb Depression (WBD) defined as the difference between the Dry Bulb Temperature (DBT) and the Wet Bulb Temperature (WBT), and about the number of hours that this WBD occurs. WBT is calculated from measured DBT and RH or Dew Point Temperature (DPT). The conversion may be done using empirical equation or software. Useful conversion equations may be found in ASHRAE fundamental handbook [25].
The ECDH concept is important when using fogging systems to cool the air down to the Wet Bulb Temperature. However, when using fogging (or overspray) icing concern at the compressor bellmouth becomes an issue. Consequently, the ECDH calculations should take into account the number of hours during the year that the system can operate without creating an icing risk. The value of the static temperature depression, which occurs at the compressor bellmouth because of airflow acceleration, depend on the gas turbine, and consequently appropriate value of Minimum Wet Bulb Temperature (MWBT) needs to be applied.
The static temperature depression can be calculated using the following equation:
pcVT2
2
=Δ , where, V is the airflow velocity, and cp is the constant pressure specific heat
of air and can be taken equal to 1004 J/kg K For some gas turbines such as aeroderivatives and some advanced technology gas
turbines, the airflow velocity at the compressor bellmouth may reach in some locations 175 m/s, leading to a static temperature depression of 15 ºC. However, with the majority of gas turbines, the maximum air flow velocity at the compressor bellmouth is less than 100 m/s, in this case the static temperature depression can be as low as 5 ºC. Table 2 summarizes the number of hours that the MWBT is above 7.2 ºC for representative locations worldwide.
As can be seen in Tab. 2, some locations know as relatively humid such as Miami (Florida, USA), Belem (Brazil), Bangkok (Thailand), and Bombay (India), with MWBT above 7.2 ºC almost the whole year, overspray may be applied. In other locations considered as dry, such as Bakers Field, CA (USA), Shanghai (China), and Sevilla (Spain), with good cooling potential, overspray may be used during the 6 hot months.
25
0
10
20
30
40
50
0 2 4 6 8 10 12 14 16 18 20 22 24hour/day
0
20
40
60
80
100
T [°C] RH [%]Dry Bulb Temp.
Wet Bulb Temp.
RH
Figure 22 – Example of variations of site ambient climatic condition in a day [24]
Table 2 – ECDH and Number of Hours (MWBT is Above 7.2 ºC) MWBT: 7.2 C
Location Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec AnnECDH 3580 3103 3316 3347 2939 2770 2813 2768 2301 2425 2971 3617 35951Hours 744 672 744 720 740 720 744 727 704 744 720 744 8722ECDH 1969 2044 2577 2783 2892 2277 2462 2399 2135 2209 2028 1973 27749Hours 723 656 736 719 742 720 744 708 685 744 721 725 8623ECDH 1239 957 1057 1057 1385 1715 1871 1960 1906 1911 1859 1625 18543Hours 744 672 744 720 744 720 744 718 714 744 720 744 8727ECDH 4195 4125 4225 3305 3268 2257 1526 1506 1750 3108 4185 4398 37848Hours 744 673 744 720 742 719 743 743 720 744 720 744 8754ECDH 263 383 930 1695 2058 1624 1855 1926 1901 2019 1472 669 16794Hours 67 92 316 652 742 720 743 739 723 734 530 219 6277ECDH 2064 3386 4906 7657 10873 13138 14442 13565 11235 8860 4459 2227 96812Hours 357 486 629 675 734 720 744 738 709 724 544 323 7382ECDH 1272 1463 2528 2812 3684 4460 5879 5653 4589 3103 1893 1298 38633Hours 449 491 646 684 739 721 745 744 720 738 638 535 7850
Bombay, India
Spain, Sevilla
USA (FL), Miami
Brazil, Belem
Thailand, Bangkok
China, Shanghai
USA (CA), Bakers field
Psychrometrics Of Inlet Fogging. A psychrometric chart shown in Fig. 23 is used to
illustrate a method to estimate amount of cooling water required for a gas turbine with airflow capacity of 200 kg/s. The following ambient conditions are used: 40 ºC DBT, 40% of RH. First, find the ambient condition on the psychrometric chart (Start Point in Fig. 23). Note that the moisture content at this condition is 18.5 ·103 kg of H2O/kg of dry air. Assuming to cool the air to the ambient wet bulb condition (100% RH as the ending condition). Proceed left up the constant wet bulb temperature line until saturation is achieved (finish Point in Fig. 23). The moisture content corresponding to this condition is 24.0 ·103 kg of H2O/kg of dry air. Therefore, the amount of moisture to be added to the air stream to achieve wet bulb temperature (WBT) is 5.5 ·103 kg of H2O/kg of dry air. Thus the amount of water required is 1.1 kg/s. This is the amount of water required to cool 200 kg/s air by 12 ºC.
26
0
5
10
15
20
25
30
5 10 15 20 25 30 35 40 45 50Dry Bulb Temperature [°C]
1
2
30
25
20
1510
Wet Bulb Temperature [°C]
[g H
2O/k
g dryA
IR]
Figure 23 – Psychrometrics of Inlet Fogging
27
1.4 – BEHAVIOR AND DYNAMICS OF FOG DROPLETS
The reason for injecting water droplets is to increase the relative humidity of the airflow by evaporation of the droplets, which in turn leads to a decrease in the air temperature resulting in an increase in air mass flow rate due to increased air density and subsequently increased power output for the gas turbine. While this may intuitively seems to be a simple matter, there are several interacting factors that define the success and efficiency of this evaporation process.
A complete understanding of the atomization process is critical for the analysis of the droplet size produced by a fog nozzle: mean size, size distribution and patterns and penetration of the droplets in the plume for a given cone angle, are important parameters. Many variables (such as the properties of the water, the geometries of the nozzle and consequently the spray angle, pressure applied on the liquid and the air flow rate) can significantly influence the results of the fogging process.
The accuracy of the droplet size measurement is critical in the application of gas turbine inlet fogging. Different techniques exist to measure droplet size including the following: (1) light scattering technique and (2) imaging with high speed video camera. Light scattering technique has two approaches: First, the spatial technique (Malvern Spraytec), which allows sampling a large number of droplets in a given volume. This technique is more appropriate when sampling is done in a high-density spray and with a small distance between the measurement position and the sizing system detectors. Second, the temporal technique (known as Phase Doppler Particle Analyzer) samples and counts individual droplets passing through the sampling volume. This technique is more appropriate when sampling is done in a low-density spray and where there is a relatively large distance between the measurement position and the sizing system detectors. For further details on droplet size measurement techniques, reader can refer the work of Le Coz [26].
From the weather conditions one can calculate the quantity of water necessary to allow the airflow to reach saturation. Given that a droplet travels at the velocity of airflow, and that it typically spends around one to two seconds between the inlet air filter and the compressor inlet, the size of the droplets has to be small enough for evaporation to take place in the time allowed. If the droplet size is larger, then the droplets would not fully evaporate and this would result in an under-saturated condition. Therefore, to attain wet bulb temperature, more quantity (than required for saturation) of water would have to be injected. More quantity of injected water combined with bigger droplets may lead to erosion of the compressor blades. Thus, size of the droplets injected with airflow in the gas turbine inlet air duct is critical.
Small difference in the droplets size has significant implications. For example, consider droplets of 20 and 30 microns diameter. The larger size droplet will have the following characteristics as compared to the smaller droplet: (1) 3 to 4 times the mass and force of impact; (2) 33% less surface area/unit volume; (3) two times faster fall rate.
A discussion on importance of parameters to evaluate fogging nozzle’s atomizing performance is presented by Mahapatra and Gilstrap [27]. 1.4.1 – Nozzle Type
There are two types of nozzles commonly employed in the industry as shown in Fig. 24: with the swirl jet design nozzle, (Fig. 24 a and c) high pressure water is forced to enter tangentially into the swirl chamber before discharging through a cylindrical hole concentric to the swirl chamber. The discharging water is in the form of an
28
axisymmetric thin conical sheet that forms ligaments and small droplets; in the impaction pin nozzle, (Fig. 24 b and d) high-pressure water is forced through a smooth orifice and hits an impact pin, located above, with high velocity. This results in the formation of a thin sheet of water in a conical shape [28]. As the water leaves the orifice, the water sheet becomes unstable and disintegrates into small thread ligaments and subsequently into billions of small size droplets.
Based on the recent comprehensive experimental study to evaluate comparative performance of the two types of nozzles at varying operating pressures (20 to 207 bars) and air flow velocity of 15.2 m/s, some interesting results were obtained [29]: droplet sizes obtained for impaction pin nozzle were significantly lower even with higher flow rates than comparative swirl jet nozzles. Smaller droplet sizes are associated with higher values of the Weber number, defined as the ratio of aerodynamic and surface tension forces, for impaction pin nozzle that results in more secondary breakup of the larger size droplets. A more detailed discussion on the secondary breakup of liquid droplets can be found in the work of Pitch and Erdman [30]. A visual examination of the plumes for the two nozzle types shows a straight edge of the cone for swirl jet nozzles, which indicates high momentum and consequently larger droplets (Fig. 24 c).
Schick and Knasiak [31] conducted an experimental investigation to understand characteristics of the above mentioned two types of nozzles and examined differences in various parameters of interests including, spray angle, flow rate, volume flux, median volume diameter, Sauter mean diameter, average flow velocity and root-mean-square flow velocity. Their study indicated wider spray pattern and slow velocity field, which are indicative of better mixing, with swirl nozzle compared to the impaction pin nozzle. It is important to note that evaporation efficiency is influenced more by the droplet size than the mixing effect. This study, like many others, agrees that the impaction pin nozzle provides lower size droplets than the swirl design.
Close-up, high-speed photographs of the nozzle spray plumes (Fig. 25) were taken in order to better understand plume formation and atomization process. A properly designed impaction pin nozzle splits the water jet when it impinges on the sharp tip of the pin and a thin conically shaped sheet of water is formed. The water sheet thins as it expands, then breaks apart into small droplets. Breakup occurs when the aerodynamic forces, which result from turbulence caused by the extremely high velocity of the sheet, overcome the surface tension of the water.
1.4.2 – Heat And Mass Transfer Of Droplets: Analytical Model A detailed description of the physical-mathematical model of the droplet evaporation
in an air stream will be discussed in the following chapter. In this section only the basic concepts of this matter are presented.
In an effort to understand behaviour of the entire fog droplets coming out of a nozzle or set of nozzles, results of a simplified analytical model of heat and mass transfer for an individual droplet were recently presented showing transient behaviour of a droplet and its interaction to the surrounding air conditions as it reaches saturation within a limited volume [29]. Transient behaviour of 20 and 50 micron droplets under the same starting ambient air conditions (30 ºC and 20% RH) are shown in Figs. 26 (a) and 26 (b), respectively. A longer evaporation time for larger size droplets is quite evident. Also, droplet size reduces as the evaporation process progresses. A further examination on the effects of change in the values of RH and active radius of the droplet (active radius, RA, defines the volume around the droplet for which saturation conditions are being calculated) showed reduced evaporation rate with increased values of RH and RA. This simplified model concurs with the findings in the field.
29
(a) (b)
(c) (d)
Figure 24 – Swirl (a) and impaction pin nozzles (b) and plume characteristics at operating pressure of 138 bar of swirl jet nozzle (c) and impaction pin nozzle (d) [29]
Figure 25 – Close up picture of nozzle spray plumes operating at 138 bar [28]
30
The effect of water temperature on droplet evaporation, within the range of water temperature used in gas turbine inlet air fogging systems, is negligible. Due to the tiny size of droplets after atomization, no matter of their original temperature (within the range used), this temperature will converge quickly, in few milliseconds, to the wet bulb temperature.
It also points out a key gas turbine operational factor, namely, if there are a mix of droplet sizes present with some large droplets above 30 microns, then the smaller droplets will evaporate first increasing the local relative humidity so making it even more difficult for the larger droplets to evaporate, thus posing a bigger risk for blade erosion should they be ingested into the compressor.
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Figure 26 – (a) Interaction of droplet to surrounding air condition (30 ºC and 20% RH) – Starting droplet size 20 microns [29]; (b) Interaction of droplet to surrounding air condition (30 ºC and 20% RH) – Starting droplet size 50 microns [29]; (c) interaction of droplet to surrounding air condition (30 ºC and 60% RH) - Starting droplet size 20 microns [29]; (d) interaction of droplet to surrounding air condition (30 ºC and 60% RH) - Starting droplet size 50 microns [29]
31
1.4.3 – Factors Affecting Droplets Size: Experimental Results Factors that determine droplet size include: (a) Airflow velocity during the droplet
size measurement; (b) position of measurement (viz., at the center, or edge of the plume); (c) ambient conditions around the droplet; (d) water characteristics - the difference in liquid characteristics (density, viscosity, and surface tension) due to the presence of foreign matter and the ambient conditions may also account for differences in droplets size distribution. The variations of droplets size as a function of air flow velocity and position of measurement including different definitions of droplets size at an ambient condition of 30 ºC and 40% RH is shown in Fig. 27.
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Figure 27 – Variation of the droplets size as a function of airflow velocity (measurements taken at 7.6 cm from the nozzle orifice) [28]
1.4.4 – Effects Of The Operating Pressure And Position Of Measurement In The Plume On The Droplet Size
As can be seen in Figs. 28 and 29, droplet size decreases with the increase in operating pressure. This decrease is more important at low operating pressure up to 138 bar, and becomes insignificant when the operating pressure increase from 138 bar to 207 bar. It can also be seen that droplets size in the center of the plume is smaller (by 2 to 8 microns) than at the edge. This is due to the fact that as the droplets leave the tip of the nozzles at high velocity, the induced airflow creates a draught toward the center of the plume. This draught carries the smallest droplets to the center of the plume.
32
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Figure 28 – Variation of droplets size as function of the applied pressure in the center and at the edge of the plume
Figure 29 – Variation of droplets size as function of the applied pressure in the center and at the edge of the plume
1.4.5 – Effect Of Distance Of Measurement And Flow Rate On The Droplet Size Independent of type of nozzles, the increase in water flow rate by modifying the
nozzle configuration (increasing the hole size, for example) and keeping the other experimental parameters constant, will lead to the increase in droplet size. This may due to the modification in parameters governing the process of atomization such as increase in the thickness of water sheet, or after the atomization process due to coalescence between droplets. Figs. 30 and 31 shows the effect of coalescence on droplet size by taking the measurement at different distance from the nozzle and up to 30 cm, and by saturating the air in the measurement volume in order to neutralize the effect of evaporation during this time. This effect is shown for the two typical airflow velocities at which the nozzle manifolds is used to be installed and which are 2.5 m/s and 12.7 m/s. Dv90 and D32 are used to quantify the effect of coalescence.
In both figures, we can see that the coalescence effect is higher at low airflow velocity. The difference in droplet size is low close to the nozzle exit in the order of one micron for D32 and two microns for Dv90, and increase with the distance from the nozzle to reach up to 5 microns for D32 and 10 microns for Dv90.
As shown in Fig. 30 with flow rate of 11.5 kg/h at 138 bar, the major increase in droplet size due to coalescence occurs in the first 10 cm, while in Fig. 31 with flow rate of 16.5 kg/h at 138 bar, this increase does not seem to be attenuated especially for Dv90. Consequently, when characterizing a nozzle, it is important to measure the droplet size at a distance sufficiently far from the nozzle at which coalescence effect is attenuated.
Fig. 30 is done using an impaction pin nozzle, and Fig. 31 using a swirl jet nozzle. As shown by Schick [31], atomization from impaction pin leads in general smaller droplet size than atomization from swirl jet nozzle at the same experimental conditions. However, atomization from different impaction pin nozzles and swirl jet nozzles may lead to different droplet size, and the performance of each nozzle depends on the specific configuration of this nozzle.
33
As can be seen in Figs. 30 and 31, droplet sizes increases with the increase of distance between the nozzle tip and the measurement position. This may be due to two effects: the first is the coalescence effect between droplets, and the second is the relatively fast evaporation of smallest droplets in the spray. Close to the nozzle exist were the density of droplets is very high, the increase in size is due essentially to coalescence. With the increase in distance from the nozzle, plume expands and coalescence effect decreases, and effect of evaporation of smallest droplets on the increase in droplet size becomes more important. To quantify the each effect, measurement was done by Chaker et al [28] at ambient conditions and saturated air, and results show that, for the test nozzle, effect of coalescence is preponderant up to 20 cm from the nozzle.
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Figure 30 – Effects of distance between the nozzle tip and the measurement position and water flow rate on droplet size [29]
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Figure 31 – Effects of distance between the nozzle tip and the measurement position and water flow rate on droplet size [29]
34
1.5 – EFFECTS OF GAS TURBINE INLET CONFIGURATION AND NOZZLE CHARACTERISTICS ON FOGGING
Gas turbine inlet systems come in a wide variety of configurations and shapes. Some of the complexities associated with inlet systems that have to be addressed include:
multiple side entry configurations (2 or 3 side entry configurations) where the nozzle manifolds have to be arranged to ensure uniform coverage as shown in Fig. 32;
configurations with extremely steep curved roofs that require progressive angular changes in the nozzle angles to minimize roof wetting;
for short duct configurations, where residence time is minimal, special patterns of the fog nozzles may have to be used to optimize the flow. In most cases, horizontal spray nozzle manifolds are used, whereas, nozzle manifolds may have to be vertically oriented (often used in filter systems with three-sided entry) to allow nozzle angular changes to be made in a vertical plane;
complexities relating to unusual duct obstructions, or presence of blow in doors. Due to the wide range of the intake duct configurations, optimization of nozzle
locations, orientation and drain system design are often based on both modeling and experience, and in certain situations, with a systematic CFD analysis as discussed by Hoffmann [32].
Figure 32 – V-Shaped duct configuration requiring special V-shaped fog nozzle array
1.5.1 – Nozzle Locations Various alternatives for locating the nozzle array/manifold are possible, namely,
down stream of the inlet air filtration system and between the silencers and trash screens. The decision of where to locate the nozzle arrays depends on several factors, including installation related downtime, desire for overspray, existing inlet duct design, overall system’s cost and the fog droplets residence time. Among different available options, nozzle manifold located upstream of the silencers provides more residence time and is advantageous for both evaporative and overspray fogging [33]. To avoid any possibility of foreign object damage, the nozzle manifolds must always be installed upstream of the trash screens. In certain installations involving large heavy-duty gas
35
turbines, two nozzle manifolds may be required, one downstream of the inlet air filters for evaporative fogging and one upstream of the trash screen for fog inter-cooling.
Residence time is an important consideration that must be carefully evaluated. Typically, fog droplets attain the airflow velocity in a few milliseconds due to the large drag forces. As shown in Fig. 33, the response time for a fog droplet to attain air-stream velocity is a function of the droplet size [29].
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Figure 33 – Response time of droplets to attain air stream velocity as a function of droplet size [29]
More in details, it could be observed that: Fog Nozzles Upstream Filter And Silencers. The advantage of positioning the fog
nozzles upstream of the air filter and silencer is that installation can be accomplished without outage time. In this case is necessary to provide to an additional fog droplet filter (not represented in figure) downstream of fog nozzle. This additional filter is necessary to remove any unevaporated fog. In general about half of the water output by the fog is considered to be captured by droplet filter and then drained away. This system is now used only on the early installations, and is rarely applied to new gas turbine installations today. This system, compared to the other options, requires more fog nozzles and more water and is very expensive to operate. However in literature is presented an advantage for this system specially in the cases of excessive loading of inlet air filters. In this contest is realized a sort of "fog scrubber" that could increase air filter life. In Figs 34 and 35 an example of fog nozzles in operation upstream of air filters is represented.
36
Figure 34 – Fog nozzle disposition upstream filter and silencer
Figure 35 – Fog nozzles in operation upstream filter and silencer
Fog Nozzles Downstream Filter And Upstream Silencers. This is the most common
location for inlet high pressure fogging. This type of installation (Fig. 36) requires an outage of one to two days and are necessary only minor modifications to the inlet gas turbine structure. This type of installation becomes indispensably for fog overspray, but to realize it there are also other options. To install fog nozzles upstream of silencer means to increase the residence time for the fog droplets to evaporate in the air stream. The real problem of this system is that dirt silencer panels increase dramatically compressor fooling. It happens that fog washes dirt from silencer panels and deposit it on the axial compressor blades. This is confirmed by observing that in this cases silencer panels are cleaner than in the other installation options. In every case fog nozzles manifold must be always installed before trash screens to avoid any possibilities of foreign object damage (FOD). This is more of a problem considering the presence of a large number of nozzles in the air stream. The possibilities of danger comes from loosening of the nozzles or damage of the grid structure caused by vibrations induced by water flow. These problems can be eliminated by careful and conservative design. Specially to avoid loosening of nozzles a trash screen is installed as defense line (Fig. 37).
Figure 36 – Fog nozzles in operation upstream filter and downstream silencer
Figure 37 – Trash screen for FOD prevention
37
Fog Nozzles Downstream Filter And Silencers. Specially for fog overspray this location (Fig. 38), just upstream of trash screens, is the best option. This configuration is required to minimize the wetting of the duct and silencer. The outage time required and the inlet turbine structure modifications are essentially the same of the previous case. There are no problems for increasing of compressor fooling but the residence time for evaporate fog droplets decrease.
Fog Nozzles Downstream Filter And Downstream Silencers. Some installations,
specially in heavy duty gas turbines, present two fog nozzles manifold, the first before the silencer to realize high pressure fogging the second downstream for fog overspray, as shown in Fig. 39.
Figure 38 – Fog nozzles in operation downstream filter and silencer
Figure 39 – Fog nozzles in operation downstream filter and upstream silencer and downstream
In Fig. 40, an example of optimization of nozzle manifold position based on droplet
size as a function of airflow velocity, and evaporation efficiency as a function residence time, is presented. In this case three different positions for the nozzle mainfold are taken into account:
Zone 1: close to the inlet filter housing where the airflow velocity is around 2.5 m/s
with a residence time for the droplets of approximately 1 second. This position is commonly used for evaporative fogging applications.
Zone 2: after the silencer where the velocity is around 12.7 m/s. This position is also
used for evaporative fogging and for combined overspray fogging. The residence time is in the order of 0.4 seconds.
Zone 3: in the duct close to the axial compressor inlet, which is the typical
overspray installation, where the residence time will be of the order of 0.2 seconds. The velocity here would also be close to 12.7 m/s.
38
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Dv90=19 microns Airflow velocity=12.7 m/s Residence time= 0.2 second Evaporation Efficiency: 85.7%% of over-spray to approach saturation: 0.09%
Zone :
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Figure 40 – Optimization of nozzle manifold position based on droplet size as a function of airflow velocity, and evaporation efficiency as a function residence time
As previously explained, Fig. 40 confirms that the position of the nozzles manifold in the duct should be chosen with care. Tradeoff between the paradoxical influence of longer evaporation time of bigger droplet size and longer residence time needs to be evaluated. Installing the fog manifold close to the inlet filter housing (zone 1), as shown in Fig. 40, where the airflow velocity is low, leads to longer residence time, and, therefore, to a better evaporative cooling efficiency. However, the fog spray is poly-dispersed, and the penetration velocity of the bigger droplets emitted from the nozzle orifice is higher than the penetration velocity of the smaller ones; consequently, droplet collision and coalescence occurs.
By installing the nozzle manifold after the silencer (zone 2), the coalescence effect is reduced significantly due the change in shear conditions at the nozzle, leading to increased droplet break-up at higher flow rates and to the fast response time of the smallest droplets to the high airflow velocity. Large and small droplets are separated into different flow paths and collisions are greatly reduced.
A typical nozzle that produces a Dv90 of 25 microns at the lower velocity of the air-filter house will produce a Dv90 of around 19 microns in the higher air velocity that exists after the silencer. By installing the nozzles manifold after the silencer, the
39
residence time of the droplets in the duct is also reduced, typically from 1 second to just 0.4 seconds.
The position of the nozzle manifold in the duct should be chosen by taking into account the trade-off of droplet size and residence time in the duct. Since all the droplets are small enough to quickly take the velocity of the airflow, the effect of the velocity itself on the evaporation rate for droplets with sizes as small as the size atomized from this type of nozzles is negligible.
1.5.2 – Nozzle Orientation and Fog Distribution The fundamental reason for evaluating nozzle orientation and fog distribution is to
obtain uniform droplet distribution in the intake duct. The value of nozzle orientation angle chosen depends on factors such as airflow velocity, distance between nozzle manifolds, operating pressure, duct wall and roof shape constraints and duct geometry.
Several studies have been done by CFD modeling and then empirically verified using wind tunnel tests to visually see and optimize the nozzle layouts. Nozzle flow angle may vary between zero degree (co-flow, see Fig 41 a) and 90 degrees (perpendicular to the flow, see Fig 41 b). 90 degrees orientation can be used when the plume does not interact with the duct walls or other lines. At 90 degrees, the increased angle gives a marginally longer residence time and slightly better mixing as compared to the co-flow position. Furthermore, at 90 degrees, the relative velocity between the droplets and the airflow is slightly higher than the co-flow position. And when manifold is installed in high airflow velocity zone (after the silencer), Weber number may become sufficiently large to initiate a secondary breakup of the biggest droplets in the spray [34].
Experimental studies have been conducted to study the effect of nozzle angle with respect to the airflow. The plume diameter stays relatively constant in the axial length. This is an advantage when the distance between the nozzles is not too large as it provides a homogenous pattern across the duct. However, in a case of large spacing between nozzles, the 90 degree position may be advantageous because the plume diameter is much larger and consequently covers more duct volume. But care must be taken that the airflow does not blow the fog back onto the manifold tube where it can coalesce and form larger droplets. This can be a problem in filter houses where airflow is not always axial to the inlet duct.
Droplet to droplet coalescence will not occur in the region where the plumes intersect, as at this location, density of the droplets is much less than at a point close to the nozzle efflux. Droplet to droplet coalescence and the resulting formation of larger droplets can be significantly close to the orifice, where droplet populations are very dense and different sized droplets are moving at different velocities. Coalescence stops within 30 cm from the orifice for most inlet fogging nozzles.
It is important to understand the velocity and pressure profiles of the airflow in the intake duct and recognize that certain areas will have accelerated flow and decreasing pressure due to duct bends. Nozzles have to be specially oriented to accommodate these airflow patterns.
40
(a)
(b)
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Figure 41 – Plume shape with nozzle orientation of a fog nozzle in wind tunnel; airflow velocity is 4 m/s, Operating Pressure is 138 bar: (a) Co–flow; (b) Ninety degree; (c) Counter flow.
41
1.5.3 – Number of Nozzles and Nozzle Pattern The number of nozzles should be appropriate to provide uniform fogging in the gas
turbine inlet duct. There is a tradeoff between nozzle flow rate, number of nozzles and associated airflow pressure drop. Having a few number of higher flow rate nozzles can result in larger separation of plumes and less homogeneity in the fog distribution. Further, larger flow nozzles often produce larger droplet sizes. In addition, fewer nozzles will worsen the temperature distribution during the part-load operating condition. To give a feel for typical numbers of nozzles, a total of 1120 nozzles were used in an installation consisting of heavy-duty GE 7FA gas turbine [33]. The nozzle pattern itself should be such as to cover the maximum inlet duct area to avoid super-saturation in certain regions and under-saturation and coalescence in others.
42
1.6 – WET COMPRESSION Wet compression (in both overspray and interstage injection application) is derived
by the deliberate introduction of fog droplets into the axial flow compressor of the gas turbine. Compressor of a gas turbine cycle consumes considerable amount of the power output produced by the gas turbine as already explained in the paragraph 1.1 and confirmed in Fig. 42 which shows data for turbines ranging from old to new technologies. One of the main advantages of overspray fogging is that it enhances power output as a result of decrease in compression work associated with the continuous evaporation of water within the compressor. The other factors which contribute to enhance power out put are: (1) increased flow rate through the turbine; and (2) increase in heat capacity of the fluid mixture flowing through the gas turbine. Typically, the amount of overspray is in the range of 0.5 to 2% of the saturated air mass flow rate of the gas turbine.
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The concept of wet compression is as old as development of the gas turbine as is
evident from the work of Kleinschmidt in late forties [35]. He noted that wet compression has better effect than intercooling in improving gas turbine cycle performance as heat removed from air is returned to the working fluid in the form of steam. His analysis further showed that wet compression not only increases efficiency for a given pressure ratio but also increases the cycle pressure ratio at which maximum efficiency occurs for a given value of turbine inlet temperature. There is a lot of analytical and experimental work reported on overspray fogging in the literature [35 – 51].
The early work was related to thrust augmentation using water injection for aircraft engine. In early Fifties, a detailed procedure for estimating thrust augmentation of a turbojet engine with the help of psychrometric chart and Mollier diagram was presented by Wilcox and Trout [36], whereas, effects of water injection on the performance of a centrifugal compressor was examined by Beede et al [37]. The analytical study of Wilcox and Trout showed impact of factors such as altitude, flight Mach number and atmospheric temperature on the thrust boost associated with water injection. In early Sixties, Hill [38] presented a systematic aero-thermodynamic analysis procedure for
43
evaluating effects of water injection on performance characteristics of an axial compressor and found good agreement with the experimental results on medium and high pressure ratio compressors of turbo-shaft engines. For a given value of evaporation parameter ( θλw ), Hill noted that the amount of work reduction due to water injection was not affected by the compressor pressure ratio. Moreover, stage work distribution was affected by the evaporation inside the compressor such that initial stages were found loaded. He further noted that the process of mixing and evaporation depend on the geometry and speed of the compressor than on the initial drops size and their distribution. It is highly likely that this conclusion could be due to the full understanding of heat transfer effects, droplets measurement techniques and factors relating to the state of technology in 1963 might have contributed to the effects of droplet size in his study. However, as it will become evident later in the discussion, more recent studies clearly show impact of droplets size on the evaporation process both upstream and within the compressor.
A thermodynamic analysis procedure for wet compression, similar to the work of Hill, but including discussion on size, evaporation rate, evaporation time and breaking of water droplets was recently presented by Zheng et al [39]. Their analysis showed that polytropic index of actual wet compression is lower than dry air compression process which results in reduced compression work and compressor discharge temperature and increased compression efficiency in presence of wet compression. It was further shown that wet compression is more effective at high pressure ratio and for a given value of pressure ratio and evaporation factor, compression work can be reduced lower than that required for the isentropic dry air compression. In spite of lacking details, Zheng et al. noted that better evaporative effects can be achieved with smaller diameter droplets.
A detailed analytical and experimental investigation of overspray using 115 MW gas turbine (Hitachi Frame 9E) by Utamura et al. [40] showed 10% gain in power output and 3% increase in thermal efficiency with 1% overspray. Their simplified model of droplet evaporation showed incomplete evaporation inside the 17 stage compressor for a droplet size larger than 30 microns. However, complete evaporation could be attained within the compressor for droplet size less than 20 microns.
A linearized one-dimensional analysis of compressor off-design performance, with simplifying assumptions that the gas constant, polytropic efficiency, and blade speed remain constant through compressor in presence of overspray, was presented by Horlock [41]. The values of constant pressure specific heat and polytropic exponent were assumed to have first order change due to evaporation within the compressor. Also, the droplet temperature was assumed not to change through compressor while estimating evaporation rate which implies neglecting effect of convective heat transfer from droplets and over-estimating evaporation length for a given droplets size.
A detailed study using two analytical models for overspray fogging, considering effects of compression rate, droplet size, and polytropic efficiency on the overall compressor performance, was recently presented by Hartel and Pfeiffer [42]. In the first model (termed ideal model), air-liquid mixture was assumed to be in thermodynamic equilibrium through the compression process. This implies that droplets are very small and compression rate does not affect thermodynamic changes. In the second model (called droplet model), limitations of the ideal model were addressed. Their analysis showed that maximum spray flow rate, which corresponds to complete saturation at the compressor discharge for a given pressure ratio, depends on the pressure ratio and polytropic efficiency and its value is higher for a compressor with lower efficiency. The results further show that difference in specific work reduction due to polytropic efficiency change is insignificant for spray flow rate less than 2% (see Fig. 43). It is possible that negligible effect of polytropic efficiency for spray flow rate less than 2% is
44
due to the fact that stage mismatching and associated aerodynamic losses are not significant.
Effects of droplet size on compressor work ratio with compression rate of 870 bars/s or pressure ratio of 30 for polytropic efficiency values of 100% and 80%, based on droplet model of Hartel and Pfeiffer are shown in Figs. 44 (a) and 44 (b), respectively. It is evident that increased specific work reduction due to overspray fogging, for a given spray flow rate and droplet size, can be achieved for a compressor with lower efficiency. Because of the fact that the models proposed by Hartel and Pfeiffer did not consider effects of stage mismatching with high fogging, there exists a possibility of over-estimation of reduced specific work and associated reduced temperature drop.
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Figure 43 – Effects of spray flow rate and polytropic efficiency (denoted by “n” in the legend) on specific work ratio for pressure ratio of 30 [42]
A method combining droplet evaporation and mean-line compressor performance to
examine effects of wet compression, in particular on stage-by-stage performance was presented by White and Meacock [43]. The evaluation of overall compressor performance characteristics showed characteristics progressively moving to a higher flow and pressure ratio with the increase of injection rate. The increase in mass flow due to wet compression is associated with evaporative cooling and is more than the additional amount of injected water. Furthermore, analysis shows decrease in aerodynamic efficiency as the injection rate increases which is mainly attributed to off-design operation of all the stages as is evident from Fig. 45 which shows stage-by-stage normalized flow coefficient. This figure clearly shows unloading (increased flow coefficient compared to design value) of early stages and over-loading (decrease in flow coefficient) of later stages.
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Figure 44 – (a) Effect of droplet size on specific work ratio (compression rate 870 bar/s and polytropic efficiency 100%); (b) Effect of droplet size on specific work ratio (compression rate 870 bar/s and polytropic efficiency 80%)
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Figure 45 – Stage flow coefficient relative to their design value – effect of wet compression with 5 microns droplets
46
Another approach to wet compression, by using interstage water injection, has been also examined [52, 53]. Arsen’ev and Berkovich [52] compared overall performance considering cases of water injection in different stages including upstream of the compressor and for compressor modified for water injection effects. Their study showed that water injection effects are improved if water is injected within the compressor. The presented results further suggested that there is an optimum stage for water injection as in their study performance was higher with water injected at the third stage.
Recently, Bagnoli et al [54] investigated effects of interstage injection on the performance of a GE Frame 7EA gas turbine using aero-thermodynamic modeling. In addition to estimating the overall gas turbine performance changes achievable with the interstage injection approach, the study presented impact of interstage injection on the stage-by-stage compressor performance characteristics of the selected gas turbine. This study showed, similar to the work of White and Meacock [43], that various stages are at off-design conditions and first few stages are unloaded and later stages are over-loaded in presence of interstage injection. Increased amount of water injection causes 17th stage operating point to move closer to the surge line suggesting that one has to be cautious in selecting the water amount in case of the interstage injection approach.
Anyway it could be observed that the main issues of wet compression approach and in particular of interstage injection are not fully studied and understood. A more detailed analysis on the interstage injection will be presented in the following chapters 3 and 4 of the present work with also a more detailed analysis of the mathematical models of evaporation of water droplets in an air stream during a compression process.
Analytical studies on the effect of droplets on the blade surface reveal requirements for very small size (1 to 5 microns) droplets [55]. The use of swirl-flash technology, where pressurized hot water is sprayed through swirl nozzle, has demonstrated achieving droplets of 2.2 microns in the laboratory. Swirl Flash technology has been applied to a number of gas turbines including a GE Frame 6, Siemens V94.2, and an ABB 9D [56 – 59]. By using a swirl nozzle and combining with pressurized hot water, droplets roughly ten times smaller in diameter and thousand times smaller in volume and weight than the droplets from a normal swirl spray could be achieved as reported by Liere et al [58]. It was also shown that approximately 14% power boost can be obtained for ABB 9D gas turbine with 2% injection of fog water.
Overspray, compared to inlet evaporative fogging has advantage as it allows the air to reach saturation in ambient conditions with high relative humidity and low dry-bulb temperature due to increase in the quantity of small size droplets in the spray resulting in faster evaporation.
In case droplets are not evaporated in the allowed time, they may flow through the compressor and boost the power while evaporating in the compressor. However, in order to get full benefit of fogging systems, it is important to get as close as possible to saturation at the compressor bellmouth, because as shown in Fig. 46, the power boost due to fog injected to saturate the air is higher than the power boost due to over spray. For example, for the ABB GT8C, the power boost, up to the saturation point, is 2.5% for each 1% of water sprayed in the flow. This power boost is reduced to 1.25 % for each 1% of injected water as over spray.
Fig. 47 shows the progressive boost in power (red) with the increase of amount of water (blue) sprayed with the inlet airflow for a given ambient psychrometric conditions. Step power boost with the implementation of number of number of nozzle manifolds reflecting increased amount of injected water is evident from Fig. 47.
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Figure 47 – Experimental power boost due to water sprayed with the airflow in an advanced industrial gas turbine inlet air duct
48
1.7 – FOGGING USERS PRACTICAL CONSIDERATIONS
1.7.1 – Water Quantity Requirements The amount of water flow required for an evaporative fogging system is a strong
function of the climatic design conditions and the temperature depression that can be obtained. An illustration as to how to calculate the required amount of water needed under given ambient psychrometric conditions is provided in a previous paragraph. Fig. 48 may be used for a rough estimate of the water quantity required for evaporative fogging. The component of water in saturated air increases with the increase of the temperature of the air. Fig. 48 shows the required amount of water to cool the air to Minimum Wet Bulb Temperature (MWBT) of 15ºC (solid lines) and 30ºC (dashed lines) for a series of Wet Bulb Depression (WBD) between 5ºC and 15ºC. This figure shows that cooling the air, for example, from 45ºC to 30ºC require around 5% more water than to cool the air from 30ºC to 15ºC. If overspray is being considered, the water flow requirements significantly increase. Typically, with overspray, the water/airflow ratio rarely exceeds 2% in the field.
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Figure 48 – Rough estimate of water flow requirements for varying gas turbine airflow rates and temperature depressions ranging from 5 °C to 15 °C (Evaporative Fogging Only)
1.7.2 – Water Quality Requirements General water quality requirements for fogging applications are provided in Tab. 3. If
the water quality is not up to the requirements, there can be considerable distress due to compressor fouling or hot section corrosion. While the values given in this table provides a general guidelines, it is best that the gas turbine OEM’s on-line water wash specification be followed. Demineralized quality water is generally acceptable for inlet
49
fogging. It is imperative however, that only stainless steel piping and fittings be used to convey the water supply to the fogging skid.
Table 3 – General Water Quality Requirements WATER QUALITY INDICATOR VALUE Water pH 6.5÷7.5 Total Solids (dissolved and undisolved) 5 ppm
Total Alkali Metals and other metals that promote hot corrosion 0.5 ppm
Conductivity 0.5÷1 μOhm/cm
1.7.3 – Foreign Object Damage (FOD) As the nozzle manifolds are mounted in the air stream, care must be taken to avoid
any chance of foreign object damage resulting from the nozzles or nozzle array components themselves. Normally the nozzles are located in a much lower velocity area after the air filtration system. Extensive safety wiring of nozzles and analysis of the fog nozzle array for airflow induced vibration should be done to ensure that the structure is strong and cannot break. In the case of overspray applications, shorter time is need to approach saturation, and consequently, the fog nozzles are typically located much closer to the inlet bellmouth of the gas turbine, usually in the vertical duct section leading to the inlet cone of the gas turbine axial compressor. Experience has shown that with careful engineering, FOD issues can be easily avoided.
1.7.4 – Gas Turbine Inlet Icing The fog control system should automatically terminate fogging whenever there is any
chance of inlet icing due to the static temperature depression that occurs in the bellmouth due to the acceleration of the air to Mach numbers of about 0.5 for heavy duty gas turbines and 0.8 for aeroderivatives. Several systems default “out” at a temperature of 15 °C though about 13 °C can be easily set as a minimum temperature. Several OEMs publish a combination of relative humidity and temperature at which anti icing measures are turned on. With fogging applications where the ending relative humidity is close to 100%, temperatures as low as 10 °C can be utilized. However to be on the conservative side, temperatures of about 13 °C are typically used as a low minimum when situations warrant it. There are several considerations other than just calculating the intake static temperature depression caused by air acceleration to Mach number of 0.5 to 0.8. There is also some heating (although small in the order of 1 °C) due to the condensation and heat transfer from the number 1 bearing. 1.7.5 – Duct Drainage
This is an important area and there are a lot of practical issues involved. Drains should be strategically located both near the silencers and also in the intake bell mouth region. These should be carefully designed for continuous drains and the number of drains should be determined based on experience, configuration of the duct and obstructions that might result in water collection. Special shaped channel sections may
50
be located on the floor and duct sections to channel water collected to the drains. One such channel is shown in Fig. 49. Appropriate sealing systems must be used to prevent the flow of ambient air into the duct. In the case of p-trap seals, it is imperative that water be supplied to these to ensure that they do not run dry due to evaporation, which will allow the inlet of air.
Channel for water drain
Nozzle lines
Silencer
Figure 49 – Special channel system for water drain
Drain flow should be monitored and logged as a function of ambient conditions and
the number of stages. This can be done simply by measuring the flow for a duration of 5 minutes and then calculating the flow rate as a percent of the overall water flow.
Special considerations have also to be made for the bellmouth floor drains. At times, pooling of water in localized segments can be ingested into the compressor as it is pulled forward away from the continuous drains. Fig. 50 shows the proximity of the floor to the inlet of a large gas turbine. In such cases, a special arrangement may have to be made to prevent pooling.
Figure 50 – Proximity of the floor to the compressor inlet can at time cause vortex ingestion of pooled water
51
1.7.6 – Compressor Surge This is more of a concern on systems that are being fog intercooled (i.e., with
overspray). In most cases, the extent of the overspray is a little over the amount that is allowed for on-line water washing. The effect of overspray is to cause the operating points on the compressor map to move towards the surge line. Ingistov [14, 15] has reported on four Frame 7EA Units that are operating with overspray for over 4 years currently. No surge related problems are experienced especially when overfogging rate ranges between 0.5% to 1% (of the compressor saturated intake air). Normally, on-line water wash rates are themselves in the range of 0.4 ÷ 0.5 % on most heavy-duty gas turbines. Even with evaporative fogging, compressor inlet temperature distortion and rate of change has to be considered carefully in the design of the fogging system and the control system. With overspray, there is a movement of the engine operating line as can be seen in Fig. 51 Lecheler and Hoffmann [60].
Figure 51 – A schematic showing movement of engine operating line with respect to surge line [60]
1.7.7 – Compressor Intake Temperature Uniformity Axial compressors have stringent intake temperature and pressure uniformity criteria.
Issues relate to blading vibration that can be induced due to extensive distortions. Fogging systems are designed in multiple stages and each stage has manifolds distributed within the inlet duct to provide a relatively uniform intake temperature. This is an important consideration as the susceptibility of the compressor to stall or surge could be affected by severe temperature distortion an idea of the criteria for a typical high Mach number aeroderivative machine. Different machines would have different criteria. There are also criteria relating to rapid changes in the inlet temperature, but fogging cannot exceed these criteria, even in the event of an emergency shutdown of the fogging system due to the residual time and water flow in the system that never results in instantaneous shock temperature changes.
The location of gas turbine compressor inlet temperature (CIT) sensors, typically do not provide a means to evaluate temperature distortion. In several engines with complicated ducts, considerable temperature and pressure distortions exist even without fogging and it is not uncommon to find temperature difference of the CIT sensors of 2 °C. In some gas turbines, the CIT sensors may be located in the intake duct in a location
52
that is starved of flow. These sorts of natural distortions can be made more evident by inlet fogging. It is a good idea to map the compressor inlet sensor readings with different ambient conditions and operating parameters and the number of fog stages in operation. This can form a useful baseline for analysis. A common problem faced is that as the CIT sensors become wet, they tend to read lower temperatures than the bulk inlet temperature. To avoid the problem, special shielded sensors should be used.
1.7.8 – Intake Duct Considerations [61] There are several issues relating to the intake duct itself that needs to be considered.
On retrofit applications, it may not be possible to have ideal situations to obtain “ideal” conditions and hence careful compromises are needed. For example, it may be impossible to obtain a length of duct with no obstructions and there may be items such as bleed heating pipes (often present for DLN machines where air flow is bled for controlling the combustor conditions under certain operating regimes). Further, some engines utilize trash screens and these will act as agglomerators causing some larger droplets. In the intake bell region of cold end drives that have intake cones, cone wetting may occur and some water may progress towards the compressor. However due to the Weber number effect, the higher velocities often result in droplet shattering and consequently the extent of the problem is minimized. The fog system supplier should be knowledgeable of these practical considerations and design to minimize any problems. Much of the choice of selecting and designing a system is an art and involves CFD coupled with past experience, and visual observations.
1.7.9 – Axial Compressor Fouling [62, 63] It is important to distinguish between the problems of natural climatic fog and the
fog generated by the fogging system. High natural humidity and climatic fog often occurs during the nights and early mornings, can cause high filter differential pressure trips and sometimes the heavily fouled filters tend to unload (leaching effect through the filter), thus causing compressor fouling. However, if the air filtration system is working well, the increased humidity caused by the fogging system does not inherently increase fouling. Fouling is a situation that is so site specific that it is very difficult to predict the behavior. At times if the No.1 bearing is leaking oil then this may combine with the high humidity (caused by inlet fogging) to create some fouling. An important issue is to wash the silencers thoroughly to avoid dirt that has been accumulated here being washed into the compressor by the fogging system during startup. This is particularly important when fogging is being installed as a retrofit on older machines. On retrofit applications, it may be necessary to perform several crank washes before the problem resolves.
1.7.10 – Compressor Blade Erosion [64, 65] In pure evaporative cooling systems, system parameters can be adjusted to ensure
evaporation of fog prior to the compressor inlet. Yet, it is possible that some droplets do enter the compressor. For relatively small size (less than 15 ÷ 20 microns) droplet, CFD studies have shown that the flow will tend to follow the air stream. There is an issue of larger water particles forming on the trash screen and inlet cone of the gas turbine but with proper design and drainage approaches this can be minimized. It is important to note that the operational experience with overspray systems has not resulted in excessive erosion problems and several OEMs are offering this technology currently.
53
Successful operational results with respect to blade erosion are presented in Figs. 52 and 53.
(a) (b) (c)
Figure 52 – Level of erosion on compressor blades of GE Frame 6B gas turbine with wet compression: First stage blade tip (a), First stage blade mid-height (b), First stage blade hub (c) (Courtesy Caldwell Energy & Environmental, Inc.)
Figure 53 – Blade condition of first stage compressor blades of Alstom advanced heavy duty industrial gas turbine GT24 (Courtesy Caldwell Energy & Environmental, Inc.)
1.7.11 – Corrosion In The Inlet Duct The use of demineralized water can deteriorate inlet ducts that are already in a
deteriorated state the increased humidity is clearly a corrosion factor. With proper maintenance and painting this problem can be mitigated. The use of SS 316L as the duct material is gaining in popularity as life cycle studies have indicated that while the first cost is a little higher, the life cycle costs are significantly lower. It is important to note that ducts, around the world, often operate in a distressed condition such as seen in Fig. 54. It is important that such problems are addressed and corrected prior to the implementation of fogging. Careful consideration should be given to special paint systems or the use of stainless steel liners where appropriate.
54
Figure 54 – A view of a corroded floor of intake duct
1.7.12 – Compressor Blading And Coating Distress [66] Some gas turbines that have undergone overspray have experienced coating distress
in the first few stages of the axial flow compressor. In most cases, this can be minimized by careful location of fogging nozzles, avoidance of excessive water accumulation on ducts and inlet cones, and by several other proprietary approaches. Corrosive ambient conditions such as chlorides or even trace contents of HCl will cause acidity and hence coating damage. Airborne contaminants in even the parts/billion (ppb) range can often create very acidic environments for the gas turbine axial blading. This is an issue that must be resolved by proper inlet air filtration, especially in aggressive industrial environments. Coatings help protect in this area. In overspray situations, the situation should be evaluated on a case-by-case basis evaluating the blading material and coating technologies available. Haskell has described the criticality of ambient air quality.
Factors that can cause distress include:
excessive or improper use of overspray; improper orientation of fog nozzles; lack of drains or inappropriately located drains; corrosive ambient conditions that will cause acidity and hence coating damage.
This is an issue that must be resolved by proper inlet air filtration, especially in aggressive industrial environments.
In some cases, if some leading edge coating distress already exists prior to fogging
then this might progress with fogging due to the reasons mentioned above.
55
1.7.13 – Off-Frequency Operation Of Gas Turbines This is an important consideration for countries where the grid operates under off-
frequency conditions. When under-frequency operation occurs, the airflow rate drops considerably and there is a possibility for some gas turbines to operate much closer to rotating stall conditions in the early stages when the under frequency operation is coupled with high ambient temperatures. This sort of operation must be carefully considered during the design of the fogging system. The change in power output for a large heavy duty gas turbine operating under frequency is indicated in Fig. 55.
In the case of overspray, the aerodynamic loading through the compressor stages is modified and so the minimum corrected compressor speed needs to be determined. As it happens, some gas turbine models actually will modulate the inlet guide vanes at under frequency operation at high ambient temperatures.
Often, in power plants, guarantees have to be met even under under-frequency operation and overspray may be used to mitigate this effect.
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1.7.14 – Electrostatic Build-Up And Bearing Distress With Overspray [64, 67] Similar to the electrostatic charges that occur in the LP section of a steam turbine
where condensed particles induce an electrostatic charge, machines that are operated with wet compression need to have superior shaft grounding brush systems utilized. Several installations have used Sohre Turbomachinery grounding brush systems very successfully. An installation of such a brush is shown in Fig. 56 [64].
56
Figure 56 – Special Grounding brush utilized for overspray applications
1.7.15 – Split-Shaft Gas Turbines For Mechanical Drive Or Power Generation Service
In the case of split-shaft gas turbine such as commonly found in mechanical drive service, the changes in the gas generator speed result in changing airflow through the machine. It is important, therefore, that this be taken into account during the fogging system design to ensure that the injected water is appropriate with respect to the gas turbine airflow. If two-shaft gas turbines are used for power generation applications as is commonly done in the case of aeroderivatives, the same considerations should apply.
1.7.16 – Considerations Relating To Engine Cooling Air Flow With Overspray A detailed study on the overall engine and compressor performance using fixed and
variable design Secondary Air System (SAS) with overspray effects has been recently reported by Cataldi et al [68] and clearly shows impact of SAS system on the pressure distribution (see Fig. 57). Note, ex.1, ex. 2 and ex. 3, in Fig. 57 represent bleed air locations, whereas, “in” and “out” represent compressor’s inlet and outlet sections, respectively. Pressure ratio at various locations is normalized with its value for the dry case (no overspray). It was evident that for an engine designed for dry operation, amount of water injected is limited due to changes in the secondary air flow, compressor load redistribution, change in surge margin, etc. With overspray, the temperature and pressure distribution within the compressor stages changes and in some cases adjustments to the bleed air orifices may be required.
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The stage-by-stage change in pressure build-up with overspray on and without
overspray in a compressor is shown in Fig. 58 [60]. On y-axis is shown the relative change is pressure build-up in the compressor and effect of increased amount of overspray on pressure build-up including unloading in the initial stages and overloading in the rear stages can be clearly seen in this figure.
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Figure 58 – Change in pressure build with overspray up in a compressor [60]
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1.7.17 – Practical Considerations During System Design And Implementation Some important practical considerations to be made, in the form of a check-list, in
the implementation of any fogging system are provided below: check vendors design calculations with regards to flow requirements, off design
conditions, and evaluate the design under different climatic conditions; ensure that the fogging system vendor is provided with detailed sketches and
photos of the inlet system. Full details including air take off for cabin cooling should be provided as appropriate measures may have to be taken to avoid moisture entry into these areas;
evaluate amount of overspray required (i.e., intercooling effect) and ensure that the compressor can accommodate this. Surge margin should be evaluated;
if overspray is being considered, review generator capability and lube oil cooler capability,
insist on all stainless steel piping; in the event of a problem, does the design permit rapid isolation so operation
can proceed without the fogging system? optimize location of the fogging manifold with respect to air filters and inlet
system; check that sequential fogging capability (cooling stages) is adequate to meet
your demand profile and turndown profile; check design features to avoid potential FOD; check rigidity of the manifolds to avoid flow-induced vibration due to gas
turbine airflow. Vendor calculations should be reviewed; review vendors proposed manifold design for structural rigidity and strength;
review possible tie-in of skid (programmable logic controller (PLC) with plant data control system (DCS);
ensure that design does not impose a large pressure drop; evaluate proposed installation of nozzle manifold in the duct for maintainability
and accessibility; ensure that appropriate drain lines will be installed in the inlet duct system; galvanized material should not be used downstream of the fog manifolds; installation of the weather sensor can be at skid location if this is representative
of inlet conditions to the gas turbine. Ensure that there is no secondary effect causing wrong measurements (eg. close proximity to an air cooled heat exchanger or source of radiant heat);
demineralized water supply lines must be thoroughly flushed to ensure that no dirt has accumulated during installation. Supply lines must be of stainless steel construction;
thoroughly wash the inlet duct and silencers. Dirt in the inlet duct system can be washed by the fog into the compressor causing axial compressor fouling creating performance deterioration or compressor damage. This is of special importance on retrofit applications;
consider the installation of a viewing window at the compressor bellmouth. Lighting arrangement should also be made to enable a view as shown in Fig. 59.
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Figure 59 – View of intake from viewing window installed in a plenum
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1.8 – ACTIVITIES OF GAS TURBINE MANUFACTURERS WITH INLET FOGGING AND OVERSPRAY
The efforts pursued by major manufacturers in using inlet and overspray fogging is briefly summarized below and identifies importance of this technology in the gas turbine industry.
1.8.1 – General Electric A different approach to high pressure fogging, called Sprint (spray intercooled
turbine) has been implemented by GE for gas turbines with multi-spool compressor such as LM6000 (Fig. 60, [69] ). In this approach, power boost is achieved by fog spraying water, atomized with eighth-stage bleed air from the high-pressure compressor, at both low-pressure and high-pressure compressor inlet plenums. Power boosts of 12% at ISO and 30% or more at ambient temperature of 32 °C compared to the LM6000-PC gas turbine are reported. According to the latest available data. More than 260 LM6000 Sprint systems have been installed since June 1998 when the first two units went into commercial operation. With Sprint on, GE recommends 16,000 hours and 50,000 hours of service for HP and LP compressor blades, respectively. GE has also published some papers on fogging of heavy duty gas turbines using a GE system called SPRIT [70].
1.8.2 – Siemens-Westinghouse The wet compression technology cooperatively developed by Siemens-Westinghouse
and Dow Chemical Company, demonstrated in 1995 on W501A gas turbine, has been implemented on number of W501-series gas turbines (Smith, [71]). In one of the peaking applications using W501D5A gas turbine, wet compression combined with conventional evaporative cooler provided more than 20% power boost and heat rate reduction of 1.5 to 3% at ambient conditions of 38 °C and 14% relative humidity. Their experience further shows that overspray provides power boost in the ratio of 4:1 compared to the inlet evaporative fogging. In addition, W501D5A and W501D5 gas turbines equipped with standard combustor have shown a reduction in NOX emissions by 20 to 40% with wet compression.
Siemens Westinghouse has also done some research work relating to testing and evaluation of fogging system effectiveness (Willems and Ritland, [72]). This work includes the use of nine CIT probes that have special arrangements for shielding so that they accurately reflect the dry bulb temperature at the compressor inlet.
1.8.3 – Rolls-Royce A detailed analytical and experimental program including CFD analysis, termed
technology acquisition program, to evaluate wet compression technology (called IFB-inlet fog boosting) in combination with other options such as fog inter-cooling, inlet chilling and fog inter-cooling plus inlet chilling was initiated by Rolls-Royce in 1998 by Walsh et al [73]. In addition, characteristics of two main types of nozzles were examined using water and air for atomization process. Impaction pin type nozzles showed a smaller size droplets distribution compared to swirl nozzle.
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Figure 60 – GE LM 6000 Sprint
1.8.4 – Alstom Alstom has done extensive studies in fogging and overspray since 2000 on a range of
their gas turbines. Alstom’s term for overspray is “High Fogging”. High fogging has been implemented on test engines Alstom GT8C2 and a GT26 engine at their test facilities in Birr, Switzerland. According to Hoffman and Mckay [74] as of 2004, 10 fogging systems had been supplied by Alstom (1 × GT8, 3 × GT 13 and 6 × GT26) and 5 “High fogging” systems have been supplied (1 × GT8, 1 × GT24 and 3 × GT26). A map defined by ALSTOM [12] indicating temperature and Relative humidity where fogging and high fogging is permitted is shown in Fig. 61.
Figure 61 – Map defined by ALSTOM indicating temperature and Relative humidity where fogging and high fogging is permitted [68]
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1.9 – INLET FOGGING FIELD EXPERIENCE AND USERS PERSPECTIVES
In general, the field experience with inlet fogging has been quite positive as evidenced by a number of gas turbines (over 1000) fitted with fogging equipment in recent years. A list of gas turbine models operating in the field with inlet fogging systems installed by third-party vendors clearly shows inlet fogging implementation on gas turbines from all the major gas turbine manufacturers worldwide (see Tab. 4). It must be noted that both inlet evaporative and overspray fogging have been implemented on the gas turbines operating in wide ranging applications (base-load, peaking, simple cycle, and cogeneration and combined cycles, etc.). Furthermore, Tab. 4 does not include gas turbines where fogging systems have been provided by gas turbine OEMs. Further, on retrofit applications, careful duct and silencer washing must be done to avoid compressor fouling. If aggressive atmospheric pollution exists or if the gas turbine filter system is not up to par, then it is possible that the conditions at the inlet can be acidic causing damage. Effective drainage systems should be employed to avoid the compressor suctioning un-atomized water off duct walls and floors, which could lead to compressor blade distress.
A brief description on few field installations including results from a comprehensive experimental study by OEM is given below.
1.9.1 – Inlet Evaporative Fogging On 80 MW Class Heavy-Duty Industrial Gas Turbine In A Cogeneration Facility
A snap-shot of 8 hours of operation showing changes in key performance parameters, corrected to ISO conditions, for a 80 MW class industrial gas turbine with and without inlet evaporative fogging is shown in Fig. 62. With inlet evaporative fogging on, on the average power boost of approximately 2.5 ÷ 3 MW can be seen. Also decrease in compressor discharge temperature and associated decrease in compressor discharge pressure with inlet evaporative fogging effects can also be observed in Fig. 62. As the fogging system is brought offline (at around 5 PM), the gas turbine attains its dry rating power. The presented data was taken recently after a major overhaul and upgrade (change in firing temperature) of the machine.
Table 4 – List of gas turbines with third party furnished fog systems
Gas Turbine Manufacturer Gas Turbine Model
Number of Units With
Fogging Alstom GT26, 13D2, 13E, 13E2, GT-11N, GT-10B, Tornado, 9D, GT24 26 GE – Frame 5 5271, 5361, 5371, MS5001, 5001P, GT11 62 GE – Frame 6 6001, 6531B, 6541, 6541B, 6551B, 6561B, 6F, 6101FA 59
GE – Frame 7 7001, 7001E, 7001B, 7001C, 7121EA, 7061EC, 7191F, 7121FA, 7221FA, 7231FA, 7241FA, 7661, 7821 347
GE – Frame 9 9161E, 0171E, 9FA++ 18 GE – Aeroderivatives and others
LM2500, LM5000, LM6000, LM6000PD, LM2500PE, PGT10 28
Mitsubishi M701-F, M501G, MF111AB, MW-701D 6
Pratt & Whitney FT4, FT4 TwinPac, GG4/FT4, FT4-A5, FT8 TwinPac, GG4A9 Twin, TP4-2C1 51
Rolls-Royce 501KB7, AVON, SK30, RB211 7 Solar Turbines Mars, Mars100, Centaur 50, Taurus 60, Taurus 70 10 Siemens-Westinghouse V64.3, V64.3A, V84.2, V94.2, V94.3, W501F 31
Siemens-Westinghouse W191G, W251B11, W251E, W501AA, W501B4, WH E4-2, W501 D5A 37
63
Figure 62 – Effects of inlet evaporative fogging on a 80 MW class heavy-duty industrial gas turbine
1.9.2 – Application Of Inlet Fogging Overspray To A Frame 5 Cogenerative Installation
The first application of wet compression or overspray was an application on three Frame 5 gas turbines (air mass flow rate of about 73 kg/s) operating in cogeneration service has been reported by Nolan and Twombly [75]. This facility is located in Rifle, Colorado, USA which has long hot and dry summers. The facility examined traditional evaporative cooling approaches but chose to use a direct fogging system owing to a low initial cost. The system utilized 540 nozzles that allowed a total flow of about 1kg/s at 40 bar pressure supplied by three reciprocating pumps.
Wet bulb temperatures could be effectively attained (i.e., 100% evaporative effectiveness) throughout the temperature range of 21 °C to 37 °C. Figure 63 shows the average gas turbine output over the July to September time frame. In this figure, the lower curve (A) shows average output for all three units if no fogging would have been employed. The curve in the middle (B) shows the predicted output from the gas turbine manufacturer’s curves with average wet bulb temperatures attained, i.e., with 100% evaporative cooler effectiveness. The uppermost curve (C) was the actual curve of these engines. The difference between curve B and C was due to the fog intercooling (wet compression) effect. As reported by Nolan and Twombly [75], an economic analysis of this system indicated a payback of 3.2 months.
64
Figure 63 – First application of overspray on Frame 5 gas turbines (Nolan and Twombly [75])
1.9.3 – Keyspan Corp Power Generation Plants According to recently reported case studies, in the Second Quarter 2004 issue of
Combined Cycle Journal, KeySpan Corp of Brooklyn, New York is operating 50 gas turbines (FT4s, GE Frame 5s, and GE Frame 7s) with overspray (or wet compression) which provides aggregate power boost of 100 MW from all machines including NOx emissions reduction of 15% [76]. It was further reported that no significant damage to compressor blades, attributed directly to wet compression, has been experienced since operation of machines for more than three years.
1.9.4 – Testing Of Four GT24/GT26 Gas Turbines With Overspray By ALSTOM At ALSTOM, tests were conducted on four GT24/GT26 series heavy-duty industrial
gas turbines (reheat cycle) to evaluate overspray fogging. The test results, shown for three engines in Fig. 64, revealed linear power boost up to injection rate of 1.2% suggesting minor impact on compressor performance. This experimental study also showed that power boost of approximately 7.1% with 1% of overspray can be achieved for GT24/GT26 engines.
1.9.5 – GE Frame 6B Gas Turbine At Cardinal Cogeneration Facility Field tests on GE Frame 6B, installed at Cardinal cogeneration facility, using
overspray fogging were conducted in April 2002 [65]. The spray nozzles, installed downstream of silencers in the inlet duct, had a total flow rate of 1.42 kg/s (overspray of 1% of airflow). The gas turbine performance with and without overspray fogging is summarized in Tab. 5. The test data shows that approximately 9% power boost for 1% overspray was achieved. A drop in compressor discharge temperature of approximately 28 °C was noted, whereas, exhaust gas temperature reduced by approximately 6 °C.
65
0.00
1.00
2.00
3.00
4.00
5.00
6.00
7.00
8.00
0.0 0.2 0.4 0.6 0.8 1.0 1.2
Engine 1
Engine 2
Engine 3
(PG
T,W
ET -
P GT,
DR
Y) /
PG
T,D
RY [
%]
f = mH2O
/ mAIR
[%]
Figure 64 – Experimental test results on GT24/GT26 gas turbines by ALSTOM [68]
Table 5 – Field performance results on GE Frame 6B gas turbine with overspray [65]
Item Without Overspray With Overspray Ambient dry bulb 15 °C 15 °C Ambient wet bulb 11 °C 11 °C Compressor Inlet air temperature 15 °C 11 °C Compressor discharge temperature 351 °C 323 °C Compressor discharge pressure 10.55 bar 10.90 bar Fuel flow 2.77 kg/s 3.00 kg/s Turbine inlet temperature 1106 °C 1106 °C Turbine exhaust temperature 554 °C 548 °C Turbine power output 35333 kW 38579 kW
Some of the issues relating to overspray are totally different. If the overspray rates
are limited to 0.5 ÷ 1.0 %, then the potential for damage is very small, provided the system is designed and operated properly. For overspray rates higher than 1%, careful consideration should be given to the compressor aerodynamics and surge characteristics. There are also potential issues relating to the generation of electrostatic charges which might call for the use of special electric grounding brushes for heavy overspray cases.
The effects of overspray fogging on compressor casing temperature and dynamic pressure in the combustor through field measurements on GE Frame-6B gas turbine in cogeneration application were recently reported by Jolly [65]. In this application, two nozzle manifolds, one downstream of the inlet filters and another downstream of the silencers were installed. While for this installation no unusual distribution in the compressor casing temperature was noted, however, high values of dynamic pressure in the dry low NOx combustion system were reported in presence of 1% overspray fogging.
A hybrid inlet cooling system consisting of a traditional evaporative cooling system combined with inlet fogging having fog nozzles installed upstream of silencers experienced fouling in the compressor because the accumulation of dirt over years of operation of the gas turbine was washed into the compressor. The compressor fouling problem was alleviated by relocating the nozzle manifold downstream of the silencer.
66
1.10 – UNRESOLVED ISSUES AND ONGOING RESEARCH There are several issues of importance that need a more careful research today. These
include both theoretical and technological aspects; the main aspects are listed below:
Testing protocol for fog droplet sizing is required, which clearly defines the system for measurements, measurement locations, flow conditions, data analysis procedures and representative droplets diameters. This is important in specifying fogging nozzle and for their testing;
better understanding of practical issues relating to droplet wetting within inlet
ducts;
special design of duct water removal methods;
developments of new fog nozzles with lower value of Dv90 droplet size than currently achievable;
proper understanding of airflow within the inlet duct to optimize distribution of
fog nozzles in order to avoid under-fogged and over-fogged areas;
fog system performance testing protocols;
understanding compressor performance characteristics in presence of inlet evaporative and overspray fogging using CFD analysis and its verification with experimental data;
a systematic study of the performance of commercial gas turbine and combined
cycle power plants units with fogging implementations to give information to the gas turbine and combined cycle user about the potential of fogging as retrofit application;
improved droplet evaporation models that allow for accurate prediction of the
size of un-evaporated droplets, which may enter the compressor;
improved wet compression models that allow the evaluation of the performance of a gas turbine with the implementation of overspray and/of interstage injection
Among the several aspects highlighted, the aim of present work is the investigation
of the last three points of the presented list. This will be presented into the following chapter (2, 3 and 4).
67
2
Analysis Of Commercial Combined Cycle Power Plants Performance Equipped With Inlet Fogging And Overspray
Combined cycle power plants are very sensitive to ambient air temperature fluctuations. Among the all inlet air cooling methods the fogging approach has shown good prospective for power output and efficiency improvement. According to the Author’s acknowledgements there is not a systematic study with the aim of investigate the performance change of commercial combined cycle power plants with the implementation of fogging. In particular in this chapter, a parametric analysis on a large number of combined cycle power plants, representative of the “state of the art” of the commercial units, implemented with inlet fogging and overspray is carried out.
The interstage injection approach was not considered in this chapter because is more complicated to be realized (due to the heavy modifications required by the gas turbine compressor) and also because the performance evaluation of a machine with this technology requires a more detailed analysis which will be developed into the following chapters. Moreover the theoretical results highlighted by this study have been applied on an existing power plants and to the entire Italian set of units of combined cycle power plants.
68
2.1 – COMBINED CYCLE POWER PLANTS SELECTION The current state of the art of combined cycle power plants (CCPPs) is shown in Fig.
65 [76] which represents the LHV electric efficiency (ηel) versus the electric power size (Pel) in a range up to 400 MWel for different configurations of combined cycle systems.
CCPPs usually come in different configurations depending on the electric power size as is evident from Fig. 65. It could be seen from Fig. 65 that up to 25 MWel the most widespread CCPP configuration consists of a single pressure level HRSG. From about 25 to 250 MWel, CCPP with two pressure levels HRSG or, sometimes, CCPP with 3 pressure levels HRSG are usually adopted. For electrical power output greater than 250 MWel, the most developed configuration consists of a three pressure levels HRSG combined with reheat. It must be noted that number of pressure levels in a HRSG is also a function of gas turbine exhaust gas temperature as is evident from Fig. 66 [76].
0.25
0.30
0.35
0.40
0.45
0.50
0.55
0.60
0.65
0 50 100 150 200 250 300 350 400
eff 1L
eff 2L
eff 3L+RH
CCPP electric power size [MW]
LHV
ele
ctric
eff
icie
ncy
1 pressure level
2 pressure levels
3 pressure levels plus reheat
Figure 65 – LHV electric efficiency versus electric power size for combined cycle systems [76]
0.25
0.30
0.35
0.40
0.45
0.50
0.55
0.60
0.65
400 450 500 550 600 650
1 pressure level2 pressure levels3 pressure level S+ RH
TOT [°C]
LHV
ele
ctric
eff
icie
ncy
Figure 66 – LHV electric efficiency versus gas turbine exhaust gas temperature for combined cycle systems [76]
69
For this study, fifteen gas turbines with power output ranging from 5 MW to 260 MW (see Tab. 6 for main performance characteristics) employed in the CCPPs have been selected. In the present study, three most widely used combined cycle configurations discussed above, namely, one (Fig. 67) and two pressure levels HRSG (Fig. 68) and three pressure levels HRSG with reheat (Fig. 69) are simulated using a commercially available computer program [3].
The main design features of the adopted CCPP configurations, for this study, are described below and the key design parameters for bottoming cycle are reported in Tab. 7:
a surface condenser using wet cooling tower with mechanical draft (the cooling water for the condenser is cooled by mixing with air in a wet cooling tower);
integral deaerator; the steam sent from HRSG to the steam turbine is superheated at all pressure
levels; economizers are placed in parallel configuration (for two and three pressure
levels HRSG); parallel disposition of superheater and reheater (for three pressure levels HRSG). the overall performance parameters of CCPPs with selected gas turbines
corresponding to the ISO conditions are provided in Tab. 8.
Figure 67 – One pressure level combined cycle power plant lay-out
70
LPHP
Figure 68 – Two pressure levels combined cycle power plant lay-out
MP LP HP
Figure 69 – Three pressure levels with reheat combined cycle power plant lay-out
71
Table 6 – Overall performance of selected gas turbines1 @ ISO conditions
GT model HRGT [kJ/kWh] TIT [°C] pressure
ratio Pel,GT [MW] TOT [°C] Mair [kg/s]
Taurus 60 12121 1093 12.0 5.31 518 21.6 Titan 130 10942 1121 15.8 13.3 497 49.0 FT8 9677 1160 20.0 24.7 463 82.6 RB211/6562 10112 1227 20.8 27.9 495 92.1 GTX100 9836 1288 20.0 42.3 553 118.5 W251 B11/12 11111 1149 15.3 48.3 519 170.4 GT8C2 10651 1177 17.6 55.3 514 191.9 V64.3 10345 1310 15.8 65.9 588 186.2 MS 6001 FA 10682 1288 14.8 69.2 597 202.2 W501 D5A 10465 1177 14.2 117.9 536 374.4 MS 9001 EC 10345 1204 14.2 168.8 561 510.2 V94.2A 10169 1260 14.0 177.1 575 504.7 W 501 G 9326 1427 19.2 245.5 599 544.9 V94.3A 9449 1316 17.0 259.9 592 635.9 MS 9001 FA 9756 1327 15.8 254.8 606 641.6
1 values obtained considering the back pressure due to the HRSG
Table 7 – Bottoming cycle design parameters Pressure levels 1 2 3
Steam turbine inlet pressure 30 bar 15 bar 50 bar
3 bar 10 bar 120 bar
Reheat pressure - - 50 condenser pressure 0.05 bar Deaerating pressure 1.2 bar Minimum stack temperature allowed 110 °C Minimum ΔT of pinch point 10 °C Minimum ΔT of approach point 20 °C Minimum ΔT of sub cooling 5 °C Steam turbine isentropic efficiency 0.860 Mechanical efficiency 0.980 Electric generator efficiency 0.985
72
Table 8 – Overall performance of selected combined cycle power plants @ ISO conditions
GT model Pressure levels
HRCC [kJ/kWh] Pel,CC [MW] HRSG
efficiency Taurus 60 1 8139 7.92 0.7455 Titan 130 1 7770 18.72 0.7178 FT8 2 7214 33.15 0.7098 RB211/6562 2 7262 38.80 0.7417 GTX100 2 6920 60.21 0.7881 W251 B11/12 2 7579 70.93 0.7619 GT8C2 2 7348 80.30 0.7575 V64.3A 2 6992 97.57 0.8115 MS 6001 FA 2 7066 104.48 0.8196 W501 D5A 2 7187 171.44 0.7754 MS 9001 EC 3 6827 255.86 0.8223 V94.2A 3 6730 267.45 0.8304 W 501 G 3 6534 350.03 0.8373 V94.3A 3 6474 378.86 0.8336 MS 9001 FA 3 6545 380.32 0.8393
2.1.1 – The Influence Of Ambient Conditions On CCPP Performance
The variation of ambient conditions (T, p and RH) influences the GT and bottoming cycle performance and, consequently the overall performance of a CCPP. At a non-ISO condition, the GT operates at off-design point and consequently the thermal energy discharged to the HRSG varies (because of variation of exhaust gas mass flow rate and temperature). Also, the bottoming cycle performance is influenced because the condenser now operates under off-design conditions and its performance is linked to the condenser cooling system adopted. In this case, the air temperature increase determines an increase in the cooling water temperature and then in the condenser pressure. As a consequence, the steam turbine electric power output reduces.
The changes in performance parameters for different CCPP configurations, utilizing selected gas turbines as listed in Tab. 6 at the selected base ambient condition compared to the ISO condition are summarized in Tab. 9. The selected base ambient condition is characterized by an ambient temperature, pressure and relative humidity of 40 °C, 1.013 bar and 40% RH, respectively. This base ambient condition will be called the Hot Day Case (HDC). The values of ambient conditions for HDC are selected mainly because inlet fogging is effective at high ambient temperature and low humidity.
It should be mentioned that RH only slightly influences the CCPP performance in comparison to the effect of ambient temperature. Nevertheless, as it will be shown below, RH has an important role in fogging employment and strongly influences its benefits. Moreover, as far as the ambient pressure is concerned, its influence on CCPP performance is negligible in the fogging performance variation.
For these evaluations, GT was regulated with constant turbine inlet temperature and steam turbine in sliding pressure.
A review of the results presented in Table 9 highlights the following: Compared to the ISO condition, HDC condition reduces GT mass flow rate (10%
to 16%) and increases turbine outlet temperature (by about 20 to 29 °C); Loss in GT power output ranges from 13 to 25 % at the HDC condition compared
to the ISO condition; these results are consistent with the previous study made on
73
GT performances under hot ambient conditions and in simple cycle operation [77];
At HDC condition, compared to the ISO condition, steam turbine (ST) power output drops by about 8 to 13 %. The observed reduction in ST power output is due to the reduction of the GT thermal energy discharged and to the increase in condensing pressure. Nevertheless, small increase in the turbine outlet temperature (TOT) and HRSG efficiency (calculated with respect to ambient temperature) compensates the ST power output reduction;
The cooling system of the condensing steam turbine is strongly influenced by ambient temperature and relative humidity. In particular, the temperature rise of the cooling water (due to the hot ambient temperature) increases the condenser pressure from 0.05 bar to about 0.1 bar;
Net electric power output of CCPPs evaluated decreases by about 11 to 22%. The greater fraction of this decrease is associated with GT performance loss since the losses for bottoming cycle are comparatively smaller as is evident from Tab. 9.
Table 9 – CCPPs performances compared between ISO and hot day case (HDC) conditions GT model Solar Turbines
TAURUS 60 Solar Turbines
TITAN 130 Case ISO HDC Δ [%] ISO HDC Δ [%] Pressure Levels 1 1 CCPP Electrical Power Output [MW] 7.92 6.80 -16.43 18.72 15.72 -19.06 CCPP Heat Rate [kJ/kWh] 8140 8384 2.91 7754 8041 3.57 CCPP LHV Electrical Efficiency 0.4423 0.4294 -3.00 0.4633 0.4477 -3.48 GT Electrical Power Output [MW] 5.31 4.41 -20.59 13.27 10.76 -23.35 GT LHV Efficiency 0.2969 0.2774 -7.02 0.3291 0.3063 -7.45 GT Mass Flow Rate [kg/s] 21.92 19.55 -12.13 49.81 44.43 -12.09 GT Outlet Temperature [°C] 518 543 25 °C 497 519 22 °C ST Electrical Power Output [MW] 2.60 2.39 -8.78 5.45 4.97 -9.76 High Pressure Steam Mass Flow [kg/s] 2.76 2.63 -4.63 5.78 5.48 -5.48 Low Pressure Steam Mass Flow [kg/s] - - - - - - HRSG efficiency 0.7455 0.7833 4.82 0.7178 0.7516 4.49
Table 9 (continue) GT model Pratt & Withney
FT8 Rolls Royce RB 211/6562
Case ISO HDC Δ [%] ISO HDC Δ [%] Pressure Levels 2 2 CCPP Electrical Power Output [MW] 33.15 27.36 -21.18 38.80 31.92 -21.54 CCPP Heat Rate [kJ/kWh] 7215 7448 3.13 7262 7501 3.19 CCPP LHV Electrical Efficiency 0.4990 0.4833 -3.25 0.4957 0.4799 -3.29 GT Electrical Power Output [MW] 24.72 19.70 -25.48 27.86 22.23 -25.33 GT LHV Efficiency 0.3720 0.3480 -6.91 0.3559 0.3342 -6.51 GT Mass Flow Rate [kg/s] 83.97 72.19 -16.31 93.63 81.48 -14.92 GT Outlet Temperature [°C] 463 492 29 °C 495 516.0 21 °C ST Electrical Power Output [MW] 8.43 7.66 -10.10 10.94 9.70 -12.84 High Pressure Steam Mass Flow [kg/s] 7.68 7.52 -2.11 9.82 9.34 -5.16 Low Pressure Steam Mass Flow [kg/s] 1.45 1.03 -40.70 1.37 0.98 -40.68 HRSG efficiency 0.7098 0.7544 5.91 0.7417 0.7794 4.84
74
Table 9 (continue)
GT model Alstom Power GTX 100
SIEMENS WESTINGHOUSE
W 251 B11/12 Case ISO HDC Δ [%] ISO HDC Δ [%] Pressure Levels 2 2 CCPP Electrical Power Output [MW] 60.21 51.46 -17.00 70.93 60.67 -16.92 CCPP Heat Rate [kJ/kWh] 6920 7065 2.05 7578 7839 3.33 CCPP LHV Electrical Efficiency 0.5202 0.5096 -2.08 0.4750 0.4593 -3.42 GT Electrical Power Output [MW] 42.33 35.58 -18.98 48.30 40.01 -20.72 GT LHV Efficiency 0.3658 0.3523 -3.83 0.3235 0.3029 -6.80 GT Mass Flow Rate [kg/s] 120.81 105.42 -14.60 173.38 156.44 -10.83 GT Outlet Temperature [°C] 553 579 26 °C 519 540 21 °C ST Electrical Power Output [MW] 17.87 15.88 -12.55 22.63 20.66 -9.56 High Pressure Steam Mass Flow [kg/s] 15.52 14.88 -4.30 19.89 19.41 -2.47 Low Pressure Steam Mass Flow [kg/s] 1.22 0.78 -56.61 2.20 1.60 -37.75 HRSG efficiency 0.7881 0.8237 4.33 0.7619 0.7940 4.04 Table 9 (continue)
GT model Alstom Power GT8C2
Ansaldo Energia V64.3A
Case ISO HDC Δ [%] ISO HDC Δ [%] Pressure Levels 2 2 CCPP Electrical Power Output [MW] 80.30 69.41 -15.68 97.57 85.35 -14.32 CCPP Heat Rate [kJ/kWh] 7349 7558 2.77 6992 7189 2.74 CCPP LHV Electrical Efficiency 0.4899 0.4763 -2.86 0.5149 0.5008 -2.82 GT Electrical Power Output [MW] 55.33 46.63 -18.66 65.92 56.72 -16.23 GT LHV Efficiency 0.3376 0.3200 -5.49 0.3479 0.3327 -4.55 GT Mass Flow Rate [kg/s] 195.18 177.01 -10.26 189.99 173.11 -9.75 GT Outlet Temperature [°C] 514 534 20 °C 588 609 21 °C ST Electrical Power Output [MW] 24.97 22.78 -9.59 31.65 28.64 -10.52 High Pressure Steam Mass Flow [kg/s] 21.96 21.46 -2.33 27.59 27.10 -1.81 Low Pressure Steam Mass Flow [kg/s] 2.56 1.90 -35.00 1.29 0.80 -62.14 HRSG efficiency 0.7575 0.7895 4.05 0.8133 0.8436 3.59
Table 9 (continue)
GT model GE MS 6001 FA
SIEMENS WESTINGHOUSE
W 501 D5A Case ISO HDC Δ [%] ISO HDC Δ [%]Pressure Levels 2 2 CCPP Electrical Power Output [MW] 104.48 89.85 -16.29 171.44 147.60 -16.15CCPP Heat Rate [kJ/kWh] 7066 7318 3.44 7187 7385 2.68 CCPP LHV Electrical Efficiency 0.5095 0.4919 -3.58 0.5009 0.4875 -2.75 GT Electrical Power Output [MW] 69.18 57.93 -19.41 117.88 98.79 -19.33GT LHV Efficiency 0.3374 0.3172 -6.36 0.3444 0.3263 -5.56 GT Mass Flow Rate [kg/s] 206.30 186.15 -10.82 381.24 342.75 -11.23GT Outlet Temperature [°C] 597 622 25 °C 536 559 23 °C ST Electrical Power Output [MW] 35.30 31.92 -10.62 53.55 48.81 -9.72 High Pressure Steam Mass Flow [kg/s] 30.87 30.29 -1.91 46.38 45.02 -3.02 Low Pressure Steam Mass Flow [kg/s] 1.21 0.67 -81.32 4.34 3.04 -42.64HRSG efficiency 0.8196 0.8522 3.83 0.7754 0.8073 3.95
75
Table 9 (continue)
GT model NUOVO PIGNONE GE MS 9001 EC
SIEMENS WESTINGHOUSE
V94.2A Case ISO HDC Δ [%] ISO HDC Δ [%]Pressure Levels 3 + RH 3 + RH CCPP Electrical Power Output [MW] 255.9 217.9 -17.42 267.4 230.8 -15.87CCPP Heat Rate [kJ/kWh] 6827 7045 3.09 6730 6926 2.83 CCPP LHV Electrical Efficiency 0.5273 0.5110 -3.19 0.5349 0.5198 -2.90 GT Electrical Power Output [MW] 168.75 138.39 -21.94 177.14 147.88 -19.79GT LHV Efficiency 0.3478 0.3245 -7.16 0.3543 0.3330 -6.39 GT Mass Flow Rate [kg/s] 519.90 466.52 -11.44 514.69 464.47 -10.81GT Outlet Temperature [°C] 561 585 24 °C 575 601 26 °C ST Electrical Power Output [MW] 87.11 79.51 -9.56 90.31 82.95 -8.87 High Pressure Steam Mass Flow [kg/s] 57.41 57.2 -0.37 60.68 60.68 0.00 Medium Press Steam Mass Flow [kg/s] 14.4 11.99 -20.10 12.91 10.7 -20.65Low Pressure Steam Mass Flow [kg/s] 2.399 1.826 -31.38 2.274 1.731 -31.37HRSG efficiency 0.8223 0.8587 4.24 0.8304 0.8645 3.95
Table 9 (continue)
GT model SIEMENS
WESTINGHOUSE W 501 G
Ansaldo Energia V94.3A
Case ISO HDC Δ [%] ISO HDC Δ [%]Pressure Levels 3 + RH 3 + RH CCPP Electrical Power Output [MW] 350.025 313.802 -11.54 378.862 332.244 -14.03CCPP Heat Rate [kJ/kWh] 6533 6633 1.51 6474 6603 1.95 CCPP LHV Electrical Efficiency 0.5510 0.5428 -1.51 0.5561 0.5452 -2.00 GT Electrical Power Output [MW] 245.46 217.12 -13.05 259.85 223.63 -16.20GT LHV Efficiency 0.3864 0.3755 -2.90 0.3814 0.3670 -3.94 GT Mass Flow Rate [kg/s] 557.59 504.15 -10.60 649.51 586.58 -10.73GT Outlet Temperature [°C] 599 627 28 °C 592 615 23 °C ST Electrical Power Output [MW] 104.57 96.69 -8.15 119.01 108.62 -9.57 High Pressure Steam Mass Flow [kg/s] 72.5 72.8 0.41 81.61 80.8 -1.00 Medium Press Steam Mass Flow [kg/s] 11.82 9.612 -22.97 14.63 12.17 -20.21Low Pressure Steam Mass Flow [kg/s] 2.305 1.74 -32.47 2.746 2.097 -30.95HRSG efficiency 0.8373 0.8705 3.81 0.8336 0.8678 3.94
Table 9 (continue)
GT model NUOVO PIGNONE GE MS 9001 FA
Case ISO HDC Δ [%] Pressure Levels 3 + RH CCPP Electrical Power Output [MW] 380.318 321.009 -18.48 CCPP Heat Rate [kJ/kWh] 6546 6720 2.59 CCPP LHV Electrical Efficiency 0.5500 0.5357 -2.67 GT Electrical Power Output [MW] 254.81 209.157 -21.83 GT LHV Efficiency 0.3685 0.3490 -5.56 GT Mass Flow Rate [kg/s] 655.42 575.67 -13.85 GT Outlet Temperature [°C] 606 631 25 °C ST Electrical Power Output [MW] 125.50 111.85 -12.21 High Pressure Steam Mass Flow [kg/s] 87.22 84.21 -3.57 Medium Press Steam Mass Flow [kg/s] 13.01 10.34 -25.82 Low Pressure Steam Mass Flow [kg/s] 2.638 1.927 -36.90 HRSG efficiency 0.8393 0.8748 4.06
76
2.2 – INLET FOGGING AND OVERSPRAY STEADY MODELS
Considering an adiabatic process, the thermal energy loss from a non-saturated air stream ( airQ& ) with the injection of water may be expressed as the sum of energy, per unit time, consumed for water evaporation ( evQ& ) and energy stored in the water droplets ( stQ& ) not yet evaporated:
stevair QQQ &&& += (8)
Considering that m& , T and cp represent the air mass flow rate, temperature and
constant pressure specific heat, respectively and using the subscripts a and ac to distinguish the air before and after cooling, airQ& becomes:
acacpacaapaair TmcTmcQ &&& −= (9)
Eq. 9 is valid if the air temperature Tac is greater than the wet bulb temperature Twb
(Tac ≥ Twb). Moreover, the energy consumed for water evaporation evQ& may be expressed as:
( )shlatevev hhmQ Δ+Δ= && (10)
where evm& is the evaporated water mass flow rate, Δhlat the latent heat and Δhsh energy spent to heat the evaporated water from the evaporation temperature to Tac.
Also, Δhlat can be evaluated, expressing the droplet temperature Td in °C, with the following equation:
[kJ/kg] dlat T3704.27.2501h −=Δ (11)
Neglecting stQ& (that means to assume Td constant during the whole evaporation
process) and the difference between am& and acm& (due to the water evaporation) and assuming the average value of the air specific heat ( pc ), Eq. 8 reduces to:
( )shlatevacapa hhm)TT(cm Δ+Δ=− && (12)
Eq. 12 may be employed to evaluated air cooling due to water injection and
evaporation and may be utilized for the inlet evaporative and overspray fogging analysis. As was the case for Eq. 9, Eq. 12 is valid if Tac ≥ Twb.
In the case of inlet evaporative fogging and with reference to Fig. 70, a complete evaporation of the entire water mass flow rate injected ( injm& ) with the spray system before the compressor inlet occurs, implying that 0mm evinj =− && . Whereas, in the overspray fogging case, liquid water enters inside the compressor (i.e., 0mm evinj ≠− && ).
It should be noted that the total amount of water mass flow rate ( wm& ) consumed is greater than the water quantity ( injm& ) used for inlet evaporative or overspray fogging
77
because of water losses (leakage) due to deposit of the droplets against the silencer, the inlet air filter and the duct walls. To account for water losses, a spray system efficiency (ηs) is introduced and then the total water consumption can be expressed as:
s
injw
mm
η=&
& (13)
The amount of blowdown mass flow rate bdm& is reported in Fig. 70 as a function of
spray system efficiency and injected mass flow rate.
s
injw
mm
η=&
&
TaRHa
TacRHac
s
sinjbd
1mm
ηη−
= &&
Spra
y sy
stem
compressor( )evinjac mmm &&& −+am&
Td
Figure 70 – A schematic showing nomenclatures at the Compressor inlet with fogging
With the above mentioned hypothesis (such as, adiabatic process, droplet
temperature constant during the evaporation) and the air becoming saturated at the compressor inlet (RHac=100% and Tac = Twb) for the evaporative fogging case, air temperature decreases at the compressor inlet (from Ta to Twb, as shown in Fig. 71) and the evaporated water mass flow rate ( evm& ) can be calculated using Eq. 12 and a Psychrometric chart.
A decrease in the compressor inlet air temperature, due to air saturation, increases the corrected speed
TnCS = , the pressure ratio (β) and the corrected flow
( pTmCF = ) because the compressor operation point moves toward the design point
(ISO condition). In the case of overspray fogging, the amount of water injected in the air stream is
more than that required for the air to saturate at a given ambient condition. Thus, the saturated air enters inside the compressor together with liquid water.
As air temperature increases inside the compressor due to compression effect, it reduces the air relative humidity below the saturation point and, consequently, resulting in additional evaporation of water droplets.
In the overspray model presented here, it has been assumed that the whole overspray water evaporates simultaneously and immediately when air temperature inside the compressor is sufficient to evaporate all the water. Obviously, the pressure (pev = pi β ev) at which evaporation occurs depends on the values of ηp and the ratio (φ ) between the evaporated mass of overspray water ( os,evm& ) and saturated air mass flow rate at the compressor inlet ( acm& ):
78
ac
os,ev
mm&
&=φ (14)
In Fig. 71, for the three different values of compressor polytropic efficiency, the
value of φ and the increase of temperature during compression (starting from saturated air at 28 °C and pressure equal to 1.01325 bar), are presented versus the evaporation pressure ratio, β ev.
Figure 71 shows that for a typical compressor polytropic efficiency of 0.90 and limiting the value of φ at 0.02, all the water mass flow rate evaporates for β ev values around 2.
This model does not take into account the effects of number and size of water droplets which have significant influence on the evaporation time. Nevertheless, it is possible to take into account the droplet number and its dimension by introducing a “delay coefficient” μ that increases the evaporation pressure value (pev = pi μ β ev). This delay coefficient is a function of the droplet dimension and, in particular: - μ=1 means that no evaporation delay is considered (usually, this approximation is
correct when the size of droplets is less than 10 μm); - μ> 1 means that an evaporation delay is considered. Typically μ ranges from 1 to
1.4 depending on the droplet dimension [3]. It should be noted that, the compressor discharge temperature (To) depends on the pev
value (and, therefore, also on μ) and, with reference to Fig. 72, it may be expressed as:
pp
1k
1k
evpa
lat
1k
1k
wbo ch
TTη
−
η−
⎟⎟⎠
⎞⎜⎜⎝
⎛μβ
βΔφ−β= (15)
where, k is the average value of specific heat ratio and the value of ηp is assumed constant.
Eq. 15 shows that by increasing μ, the compressor discharge temperature and specific work increase.
-40
-20
0
20
40
60
80
100
120
140
0
0.005
0.01
0.015
0.02
0.025
0.03
0.035
0.04
0.045
1 1.2 1.4 1.6 1.8 2 2.2
Tem
pera
ture
[°C
]
φ
βev
ηp= 0.8
ηp= 0.9
ηp= 1.0
Figure 71 - Air temperature and evaporated water mass through compressor
79
Entropy
pi
Tem
pera
ture
wb
pi μ β
ev
o
po= p
i β
Standardcompression
Wetcompression
Figure 72 - Temperature-Entropy diagram for a compression process in presence of overspray fogging
Droplet diameter influences the evaporation time and, consequently, the evaporation pressure ratio and the compressor discharge temperature. In particular, as droplet diameter increases, evaporation pressure ratio and compressor outlet temperature increase.
In Fig. 73, two compression processes (both from pi to po) with different evaporation pressure pev, are presented. The compression process with dotted line takes into account effect of droplet size (or delay coefficient). Defining β the total pressure
ratio ⎟⎟⎠
⎞⎜⎜⎝
⎛=β
i
o
pp
and β ev the evaporation pressure ratio ⎟⎟⎠
⎞⎜⎜⎝
⎛=β
i
evev p
p, and assuming k
and ηp constant [41], it is possible to express the temperature T1 as:
p
1k
1k
evwb1 TT η−
β⋅= (16)
Entropy
pi
Tem
pera
ture
wb
pi μ β
ev
o
po= p
i β
o'
pi β
ev
2
2'
1
1'
Figure 73 – Temperature –Entropy diagram for overspray compression process considering different droplet diameters
80
Furthermore, indicating T1 - T2 = ΔT, for the thermal equilibrium (Eq. 12 in which Δhsh is neglected), values:
latos,evpaac hmTcm Δ⋅=Δ⋅ && (17)
pa
lat
pa
lat
ac
os,ev
ch
ch
mm
TΔ
⋅φ=Δ
⋅=Δ&
& (18)
Moreover, the compressor discharge temperature To can be expressed as:
( ) pp
1k
1k
ev1
1k
1k
ev2o TTTT
η−
η−
⎟⎟⎠
⎞⎜⎜⎝
⎛ββ
⋅Δ−=⎟⎟⎠
⎞⎜⎜⎝
⎛ββ
⋅= (19)
introducing Eq. 17 and 18, Eq. 16 reduces to:
pp
1k
1k
ev
1k
1k
wbo TTTη
−
η−
⎟⎟⎠
⎞⎜⎜⎝
⎛ββ
⋅Δ−β⋅= (20)
If now droplet dimension is taken into account, the water evaporation will be for a
pressure ratio (μ β ev) greater than β ev (μ≥ 1 is the “delay coefficient”). It should be noted that, if no variation occurs in the injected water mass flow rate,
TTTTT 21'2'1 Δ=−=− (21)
and Eq. 20 becomes
pp
1k
1k
evpa
lat
1k
1k
wbo ch
TTη
−
η−
⎟⎟⎠
⎞⎜⎜⎝
⎛μβ
βΔφ−β= (22)
Eq. 22 is same as Eq. 15 discussed previously. It should be highlighted that the compressor discharge temperature (To’) considering
the droplet dimension (μ> 1) is higher than that (To) obtained without evaporation delay (μ=1):
oo TT >′
81
2.3 – CCPP PERFORMANCE WITH INLET EVAPORATIVE AND OVERSPRAY FOGGING
To study the effects of high pressure inlet fogging on CCPPs performance
parameters, inlet evaporative and overspray fogging conditions were analyzed for various CCPPs using gas turbines listed in Tab. 6. In the case of overspray fogging, an additional water mass flow rate ranging from 0.5% to 2% of saturated air flow rate was considered. The various results are presented as change (percent change or just difference) in performance parameters with respect to the HDC condition. This approach allows comparison of various systems operating under same base ambient conditions. In addition, results for ISO condition are also included in each figure. This presentation allows to view changes in a performance parameter both with respect to ISO and HDC conditions for various fogging cases examined.
Net combined cycle plant output, comparing inlet fogging case to HDC, shows an increase in a range of 6% to 10% for various systems as is evident from Fig. 74. This observed change implies a gain of about 0.50% to 0.75% in net plant output for every 1°C of inlet air cooling. It may be noted that for the assumed base ambient condition (HDC), the maximum amount of cooling achievable is 12 °C. As anticipated and evident from Fig. 74, percent increase in net plant output becomes higher with the application of overspray fogging for each CCPP. In particular, with 1% overspray fogging, all the lost power due to high ambient temperature (HDC condition) is recovered for various machines examined in this study with few exceptions (such as MS9001FA, MS6001FA and FT8) as shown in Fig. 74. Whereas, with 2% of overspray a net gain in electrical power output compared to the ISO case could be achieved for various systems examined (Fig. 74). Based on the limited number of machines examined, results suggest that a higher amount (approximately, 4-9%) of boost in net power output can be achieved for a CCPP with an aero-derivative gas turbine (such as RB211 and FT8) than a CCPP with an advanced technology gas turbine (For example, MS9001FA, MS6001FA, W501G, etc) with 2% overspray fogging (Fig. 74). Furthermore, a CCPP using a traditional gas turbine (such as, W251 B11/12, TITAN 130, MS9001EC, etc.) shows higher (approximately, 2 ÷ 7%) gain in net power output with 2% overspray fogging compared to a CCPP using an advanced technology gas turbine.
The observed differences in the amount of power boost between an aero-derivative and advanced technology or traditional gas turbines can be attributed, among other factors, to the fact that the rate of loss in power output due to increased ambient temperature is comparatively higher for an aero-derivative gas turbine. The other factors could include differences in compressor performance characteristics and flexibility of changing rotational speeds for aero-derivative machines compared to the other two categories of gas turbines.
A further examination of the data shows that differences in power boost for three categories of gas turbines change (increase) with the increased amount of overspray. It is important to mention that, based on the field performance data, the commercial program used tend to over-estimate effects of fogging. Despite of differences in the absolute values of certain performance parameters, the study presented does show trends of performance gains due to fogging for CCPPs.
The observed increase in net plant output can mainly be attributed to GT performance enhancement as shown in Fig. 75. Whereas, the effect of fogging on steam turbine power output (Fig. 76) is small (maximum value, 6%), compared to the GT power output boost, and especially no significant change seems to exist between ST power boost and the type of applied fogging strategy. This observation could be explained
82
considering that drop in ST performance is principally caused by ambient temperature and relative humidity effects on the condenser cooling system and marginally by GT.
Comparing GT power output between inlet fogging case to HDC, an increase of about 8 ÷ 12% power boost (equivalent to approximately, 0.7 ÷ 0.8% power boost for every 1 °C of cooling of air at compressor inlet) can be observed (Fig. 75). Furthermore, with the use of overspray more power output gain is obtained. This amount of power boost grows with the increase in injected overspray water mass flow rate (Fig. 75).
Among the other parameters that influence CCPPs performance, GT mass flow rate and turbine outlet temperature are presented, in Figs. 77 and 78, respectively. Inlet evaporative and overspray fogging increase total mass flow rate through the compressor but obtained values are less than the ISO case (Fig. 77). This could be explained considering that air mass flow rate (for constant compressor speed) depends only on air density (and so by air temperature and relative humidity) at the compressor inlet. Air temperature at the compressor inlet is always 28 °C and relative humidity 100% for different fogging conditions; this implies that air density is lower compared to the ISO condition. The total air mass flow rate through a compressor increases as the fogging condition changes from inlet evaporative fogging to 2% overspray case and is exclusively due to the increased amount of injected water. A closer look at the data shows that increase in total air mass flow rate due to fogging (evaporative or overspray fogging) is higher, by approximately 2-3% for aero-derivative gas turbines (Fig. 77) compared to the other gas turbines with the exception of two gas turbines (GTX100 and MS9001FA).
It is not surprising to note that the value of TOT increases (by 20 to 30 °C) with the increase in ambient temperature (i.e., for the HDC condition) for various systems examined as shown in Fig. 77. Inlet and overspray fogging application reduce TOT value but does not reach its ISO value. It could be observed that the difference in TOT value compared to the ISO condition reduces with the increased amount of overspray fogging as shown in Fig. 78 mainly due to the inter-cooling effect.
The change, expressed in percentage point, in the plant LHV electric efficiency compared to the HDC condition is found small (less than one percentage point) due to different fogging conditions as shown in Fig. 79. A slightly higher increase in efficiency for aero-derivative gas turbine compared to other gas turbines is also evident in Fig. 79. Furthermore, efficiency loss for various CCPPs due to change in ambient condition to the HDC condition could not be recovered even with 2% overspray fogging.
It is quite evident from Fig. 80 that a CCPP with an advanced technology gas turbine requires considerably more amount of injected water per unit % power boost compared to the other gas turbine based systems. The main reason being that the injected water amount is a percentage fraction of the inlet air mass flow rate. Typically, an advanced technology gas turbine has a high air mass flow rate as is evident from Tab. 6. Thus, to compare injected water flow rates for different combined cycle systems, it would be more appropriate to normalize water flow rate by inlet air flow rate and percent power boost. A comparison of such a parameter does show that somewhat lower amount (less than 1 liter/hr per unit of air mass flow rate for 1% of power boost) of injected water is required for an aero-derivative based CCPP in comparison to the other types of gas turbines.
83
0
5
10
15
20
25
30TAURUS60
TITAN130
FT8
RB 211
GTX100
W 251B11/12
GT8C2
V64.3AMS6001FA
W501D5A
MS9001EC
V94.2A
W 501 G
V94.3A
MS9001FA
ISO Inlet Fogging0.5 % Overspray 1% Overspray2% Overspray
Figure 74 – % CCPP Power Output change @ Generator (Reference case - HDC)
48
1216202428323640
TAURUS60
TITAN130
FT8
RB 211
GTX100
W 251B11/12
GT8C2
V64.3AMS6001FA
W501D5A
MS9001EC
V94.2A
W 501 G
V94.3A
MS9001FA
ISO Inlet Fogging0.5 % Overspray 1% Overspray2% Overspray
Figure 75 – % GT Power Output change @ Generator (Reference case - HDC)
84
-4-202468
101214
TAURUS60
TITAN130
FT8
RB 211
GTX100
W 251B11/12
GT8C2
V64.3AMS6001FA
W501D5A
MS9001EC
V94.2A
W 501 G
V94.3A
MS9001FA
ISO Inlet Fogging0.5 % Overspray 1% Overspray2% Overspray
Figure 76 – % ST Power Output change @ Generator (Reference case - HDC)
02468
1012141618
TAURUS60
TITAN130
FT8
RB 211
GTX100
W 251B11/12
GT8C2
V64.3AMS6001FA
W501D5A
MS9001EC
V94.2A
W 501 G
V94.3A
MS9001FA
ISO Inlet Fogging0.5 % Overspray 1% Overspray2% Overspray
Figure 77 – % GT Mass Flow Rate change (Reference case - HDC)
85
-40
-35
-30
-25
-20
-15
-10
-5TAURUS60
TITAN130
FT8
RB 211
GTX100
W 251B11/12
GT8C2
V64.3AMS6001FA
W501D5A
MS9001EC
V94.2A
W 501 G
V94.3A
MS9001FA
ISO Inlet Fogging0.5 % Overspray 1% Overspray2% Overspray
Figure 78 – TOT [°C] change (Reference case - HDC)
-2.0-1.5-1.0-0.50.00.51.01.52.0
TAURUS60
TITAN130
FT8
RB 211
GTX100
W 251B11/12
GT8C2
V64.3AMS6001FA
W501D5A
MS9001EC
V94.2A
W 501 G
V94.3A
MS9001FA
ISO Inlet Fogging0.5 % Overspray 1% Overspray2% Overspray
Figure 79 – % points change of LHV CCPP Efficiency (Reference case - HDC)
86
0
500
1000
1500
2000
2500
3000TAURUS 60
TITAN 130
FT8
RB 211/6562
GTX 100
W 251 B11/12
GT8C2
V64.3AMS6001FA
W 501 D5A
MS 9001 EC
V94.2A
W 501 G
V94.3A
MS 9001 FA
Inlet Fogging 0.5 % Overspray
1% Overspray 2% Overspray
Figure 80 – Injected water consumption per unit percent Net CCPP output change [(liters/hr)/% power boost] (Reference case - HDC)
0.40
0.45
0.50
0.55
0.60TAURUS 60
TITAN 130
FT8
RB 211/6562
GTX 100
W 251 B11/12
GT8C2
V64.3AMS6001FA
W 501 D5A
MS 9001 EC
V94.2A
W 501 G
V94.3A
MS 9001 FA
Inlet Fogging 0.5 % Overspray
1% Overspray 2% Overspray
Figure 81 – Incremental efficiency for CCPP due to fogging (Reference case - HDC)
87
Incremental efficiency, defined as the ratio of incremental power divided by additional fuel consumption due to fogging [13], is indicative of fuel efficiency associated with fogging technology. The trend of incremental efficiency values shows that gas turbines having high efficiency without fogging also have high incremental fuel efficiency in presence of fogging when used in combined cycle applications as is evident in Fig. 81. It is interesting to note that for many advanced technology gas turbine based CCPPs, value of incremental fuel efficiency decreased with the increase of overspray amount. On the contrary, incremental fuel efficiency always increased with increased amount of overspray for aero-derivative and traditional gas turbine based CCPPs. Also, the gain in incremental efficiency, compared to the base case (HDC) is approximately 6% (max.).
88
2.4 – INLET FOGGING APPLICATION ON AN EXISTING COMBINED CYCLE POWER PLANT
The behavior of combined cycle power plants with inlet fogging analyzed in the previous paragraphs and the analysis methodology developed (from 2.1 to 2.4) could be applied to study also a specific case. In this paragraph an example about the study of inlet fogging implementation on an existing combined cycle power plant owned by Enipower and placed in Ravenna (Italy), is carried out. 2.4.1 – Plant Lay Out
In Fig. 82 the considered combined cycle power plant lay out is presented. The used gas turbine is a GE - MS9001E model fueled with natural gas (LHV = 49655 kJ/kg). The heat recovery steam generator, in which is present a duct burner, is characterized by two pressure levels and an integral deaerator. It should be noted that the lowest pressure level steam in not sent to the turbine but totally to the thermal utility. In addiction the steam turbine presents two steam extraction for the thermal utility.
MS9001EA
thermal utility #2
thermal utility #1
thermal utility #3
steam turbine
Integral deaerator
HP LP
duct burner
Open loop condenser
Make up water
inlet water pre-heater
Figure 82 – Two pressure levels combined cycle power plant (owned by Enipower and placed in Ravenna) lay out
89
2.4.2 – Plant Simulation Model The combined cycle performances were evaluated with a commercial program for
energy system simulations [3]. The calculation sheet which was realized for the thermodynamic study of the power plant is reported in Fig. 83.
COND: condenser CONTR: temperature controller HPECO: high pressure economizer HPEVA: high pressure evaporator HPPUMP: high pressure water pump
HPSH: high pressure superheater LPECO: low pressure economizer LPEVA: low pressure evaporator LPPUMP: low pressure water pump LPSH: low pressure superheater
PC: post combustion chamber PREECO: economizer PROC: thermal utility RCMIX: mixer ST: steam turbine
Figure 83 – Calculation sheet for power plant thermodynamic simulation [3]
The main assumptions of the realized model are the following:
Gas turbine
the gas turbine performances at ISO conditions are presented in Tab 10 [3]; the gas turbine is operated with constant turbine inlet temperature (1224 °C); gas turbine off design performances are evaluated by using non-dimensional
maps; an example of the used maps is presented in Fig. 84; the gas turbine is fueled with natural gas at 27 °C and 27 bar; used natural gas
low heating value it equal to 49655 kJ/kg.
90
Table 10 – GE MS9001 performances @ ISO conditions
Net electric power output 124.7 MW Net electric efficiency 34.0% Air mass flow rate 410 kg/s Pressure ratio 12.3 Turbine inlet temperature 1124 °C Turbine outlet temperature 541 °C
0.85
0.90
0.95
1.00
1.05
10 15 20 25 30 35Ambient air temperature [°C]
@ 60% relative humidity 1.01325 bar
compressor inlet air mass
flow ratefuel massflow rate
net electric power output
Figure 84 – GE MS9001 non-dimensional performances maps
Duct burner
the duct burner is used only if high pressure steam temperature is lower than 510 °C;
duct burner combustion efficiency is equal to 98%. Heat recovery steam generator
high pressure superheated steam temperature is always guaranteed higher than510 °C and lower than 525 °C;
the pressure of superheated steam sent to the thermal utility is constant at 9.7 bar;
deaerator pressure is constant at 4.9 bar. Cool source
the cool source is an open loop condenser with see water as cooling fluid; see water mass flow rate is constant at the value of 917 kg/s; see water temperature is considered a function of ambient air temperature.
91
Steam turbine steam inlet pressures are constant at the values of 119, 19 and 5.5 bar; the steam mass flow rate extracted from the turbine and sent to the thermal
utilities is evaluated with the logic presented in Fig. 85.
MST1 MST2
Mvap19bar Mvap5.5bar
MST3 ST1 ST2 ST3
MST1 = steam flow at ST1 inlet MST2 = steam flow at ST2 inlet MST3 = steam flow at ST3 inlet Mvap19bar= steam flow extracted between ST1 and ST2 Mvap5.5bar= steam flow extracted between ST2 and ST3
0
60
120
180
240
300
360
0 60 120 180 240 300 360
steam mass flow rate at ST1 inlet [t/h]
[t/h]
MST3
MST2
Mvap5.5bar
Mvap19bar
MST1
Figure 85 – Steam turbine control logic
92
2.4.3 – Site Ambient Conditions The ambient air temperature measured in Ravenna during the plant operation of the
year 2002 is presented in Fig. 86. For what concern the sea water temperature, the measured data in Ravenna combined
cycle site related to the ambient air temperature values are presented in Fig. 87.
Am
bien
t air
tem
pera
ture
[°C
]
-10
-5
0
5
10
15
20
25
30
35
40
1-Ja
n-02
1-Fe
b-02
1-M
ar-0
2
1-Ap
r-02
1-M
ay-0
2
1-Ju
n-02
1-Ju
l-02
1-Au
g-02
1-Se
p-02
1-O
ct-0
2
1-N
ov-0
2
1-D
ec-0
2
Dati relativi al periodo dal 01/01/2002 al 17/12/2002
Since 01/01/2002 to 12/17/2002
Figure 86 – Ambient air temperature in Ravenna since 01/01/2002 to 12/17/2002
0
5
10
15
20
25
30
35
0 5 10 15 20 25 30 35
valori misurati
Media Mobile su 50 per. (valori misurati)
Dati relativi al periodo dal 01/01/2002 al 17/12/2002
Sea
wat
er te
mpe
ratu
re [°
C]
Since 01/01/2002 to 12/17/2002
Ambient air temperature
Measured values
Average values on 50
Figure 87 – Sea water temperature as function of ambient air temperature in Ravenna since 01/01/2002 to 12/17/2002
From the hourly data reported in Fig. 87 is possible to determinate an average trend that shows the sea water temperature as a function of the ambient temperature. The
93
considered value of the sea water temperatures linked to the ambient air temperatures are presented in Tab. 11.
Table 11– Link between sea-water temperature and air ambient temperature air ambient temperature [°C] 10 15 20 25 30 35 40 sea-water temperature [°C] 12 16 22 24 25 26 27
In Fig. 88 the trend of the hourly values of air relative humidity versus air ambient temperature is plotted. This figure confirms that is not possible to find a relation between the air temperature and the relative humidity. It could be observed that Ravenna's site is characterized by values of relative humidity variable from 30% to 100% and they are independent from air temperature.
0
10
20
30
40
50
60
70
80
90
100
0 5 10 15 20 25 30 35
valori misurati
Dati relativi al periodo dal 01/01/2002 al 17/12/2002
Ambient air temperature [°C]
Rel
ativ
e hu
mid
ity [%
]
Since 01/01/2002 To 12/17/2002
Figure 88 – Ambient relative humidity as function of ambient air temperature in Ravenna since 01/01/2002 to 12/17/2002
94
2.4.4 – Plant Performance With And Without Fogging The main combined cycle parameters, as differences between the operation with
fogging and without it, as function of the ambient air temperature (from 10 °C to 40 °C) and for different values of relative humidity (from 40% to 80%), are presented in the Figs. from 89 to 95. The estimated values (with and without fogging application) are presented in Tabs. from 12 to 18.
0
2
4
6
8
10
12
10 15 20 25 30 35 40Ambient air temperature [°C]
CC
pow
er o
utpu
t cha
nge
[MW
]
RH=40%
50%
60%
70%
80%
Figure 89 – CC power output change as a function of ambient air temperature
0
2
4
6
8
10
12
10 15 20 25 30 35 40Ambient air temperature [°C]
GT
pow
er o
utpu
t cha
nge
[MW
]
RH=40%
50%
60%
70%
80%
Figure 90 – GT power output change as a function of ambient air temperature
95
0.0
0.2
0.4
0.6
0.8
1.0
10 15 20 25 30 35 40Ambient air temperature [°C]
ST p
ower
out
put c
hang
e [M
W]
RH=40%
50%
60%
70%80%
Figure 91 – ST power output change as a function of ambient air temperature
-0.10
0.00
0.10
0.20
0.30
0.40
0.50
10 15 20 25 30 35 40Ambient air temperature [°C]
CC
effi
cien
cy c
hang
e [p
erc.
poi
nts]
RH=40%
50%
60%
70% 80%
Figure 92 – CC efficiency change as a function of ambient air temperature
0
500
1000
1500
2000
2500
10 15 20 25 30 35 40Ambient air temperature [°C]
Fuel
flow
rate
cha
nge
[Sm
3 /h]4 RH=40%
50%
60%
70%
80%
Figure 93 – Fuel flow rate change as a function of ambient air temperature
96
0.0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
10 15 20 25 30 35 40Ambient air temperature [°C]
GT
effic
ienc
y ch
ange
[per
c. p
oint
s]
RH=40%
50%
60%
70%
80%
Figure 94 – GT efficiency change as a function of ambient air temperature
0
1
2
3
4
5
6
7
8
10 15 20 25 30 35 40Ambient air temperature [°C]
fogg
ing
wat
er c
onsu
mpt
ion
[ton/
h]
RH=40%
50%
60%
70%
80%
Figure 95 – Fogging water consumption as a function of ambient air temperature
97
Table 12 – CC power output [MW] Ambient air temperature [°C]
10 15 20 25 30 35 40
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
40 157.742 162.138 153.421 158.557 148.358 154.752 144.225 151.653 140.281 148.590 136.422 145.675 132.575 142.825 50 157.804 161.448 153.505 157.736 148.511 153.748 144.386 150.394 140.505 147.226 136.724 144.186 132.930 141.135 60 157.889 160.731 153.620 156.922 148.623 152.850 144.568 149.232 140.697 145.952 136.974 142.780 133.308 139.627 70 157.951 160.090 153.713 156.156 148.775 151.866 144.719 148.167 140.920 144.780 137.274 141.483 133.668 138.279 R
H [%
]
80 158.036 159.419 153.791 155.414 148.907 150.931 144.880 147.147 141.163 143.727 137.553 140.305 134.042 137.090
Table 13 – GT power output [MW] Ambient air temperature [°C]
10 15 20 25 30 35 40
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
40 121.261 125.444 117.699 122.616 113.833 119.908 110.099 117.191 106.375 114.322 102.726 111.553 99.066 108.830 50 121.306 124.746 117.759 121.784 113.914 118.937 110.206 115.939 106.514 112.927 102.905 110.025 99.295 107.092 60 121.350 124.072 117.819 120.987 113.994 118.019 110.312 114.773 106.652 111.651 103.084 108.594 99.523 105.543 70 121.394 123.419 117.879 120.226 114.074 117.042 110.418 113.686 106.791 110.473 103.262 107.267 99.750 104.151 R
H [%
]
80 121.439 122.789 117.939 119.497 114.155 116.115 110.524 112.675 106.929 109.374 103.440 106.055 99.976 102.892
98
Table 14 – ST power output [MW] Ambient air temperature [°C]
10 15 20 25 30 35 40
w
ithou
t fo
ggin
g
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g
40 36.481 36.694 35.722 35.941 34.525 34.844 34.126 34.462 33.906 34.268 33.696 34.122 33.509 33.995 50 36.498 36.702 35.746 35.952 34.597 34.811 34.180 34.455 33.991 34.299 33.819 34.161 33.635 34.043 60 36.539 36.659 35.801 35.935 34.629 34.831 34.256 34.459 34.045 34.301 33.890 34.186 33.785 34.084 70 36.557 36.671 35.834 35.930 34.701 34.824 34.301 34.481 34.129 34.307 34.012 34.216 33.918 34.128 R
H [%
]
80 36.597 36.630 35.852 35.917 34.752 34.816 34.356 34.472 34.234 34.353 34.113 34.250 34.066 34.198
Table 15 – CC efficiency [%] Ambient air temperature [°C]
10 15 20 25 30 35 40
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
with
out
fogg
ing
with
fo
ggin
g
40 42.82 42.77 42.66 42.60 42.31 42.33 42.17 42.22 42.06 42.13 41.93 42.04 41.77 41.9350 42.81 42.78 42.65 42.61 42.31 42.32 42.16 42.20 42.05 42.11 41.93 42.01 41.75 41.9060 42.82 42.78 42.66 42.62 42.30 42.33 42.16 42.19 42.03 42.09 41.90 41.98 41.75 41.8570 42.81 42.79 42.65 42.62 42.31 42.32 42.15 42.18 42.03 42.07 41.90 41.95 41.73 41.82R
H [%
]
80 42.81 42.79 42.64 42.62 42.30 42.31 42.14 42.16 42.02 42.05 41.89 41.92 41.72 41.78
99
Table 16 – Fuel flow rate [Sm3/h] Ambient air temperature [°C]
10 15 20 25 30 35 40
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
with
out
fogg
ing
with
fo
ggin
g
40 39276 40416 38343 39679 37385 38981 36466 38296 35563 37606 34688 36949 33838 3631350 39297 40234 38369 39465 37422 38731 36515 37994 35625 37274 34769 36591 33943 3591660 39316 40057 38396 39257 37457 38496 36563 37715 35688 36972 34850 36259 34047 3556870 39337 39888 38424 39062 37494 38259 36610 37454 35751 36696 34931 35956 34151 35257R
H [%
]
80 39356 39724 38450 38874 37529 38037 36659 37215 35813 36440 35012 35682 34254 34981
Table 17 – GT efficiency [%] Ambient air temperature [°C]
10 15 20 25 30 35 40
with
out
fogg
ing
with
fo
ggin
g
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
with
out
fogg
ing
with
fo
ggin
g w
ithou
t fo
ggin
g w
ith
fogg
ing
40 32.92 33.09 32.73 32.95 32.46 32.80 32.19 32.63 31.89 32.41 31.57 32.19 31.21 31.9550 32.91 33.06 32.72 32.90 32.46 32.74 32.18 32.53 31.88 32.30 31.56 32.06 31.19 31.7960 32.91 33.02 32.72 32.86 32.45 32.69 32.17 32.45 31.86 32.20 31.54 31.93 31.17 31.6470 32.90 32.99 32.71 32.82 32.44 32.62 32.16 32.36 31.85 32.10 31.52 31.81 31.14 31.49R
H [%
]
80 32.90 32.96 32.70 32.77 32.43 32.55 32.14 32.28 31.83 32.00 31.50 31.69 31.12 31.36
100
Table 18 – Fogging water consumption [ton/h] Ambient air temperature [°C] 10 15 20 25 30 35 40
40 3.373 3.989 4.626 5.242 5.825 6.390 6.919 50 2.772 3.265 3.758 4.234 4.673 5.098 5.483 60 2.189 2.563 2.941 3.287 3.618 3.920 4.198 70 1.631 1.897 2.167 2.405 2.639 2.844 3.031 R
H [%
] 80 1.084 1.260 1.429 1.580 1.728 1.854 1.966
More in detail, on the basis of Figs. from 89 to 95 and of the Tabs from 12 to 18, it could be observed that:
combined cycle power output increase is a function of both ambient temperature and relative humidity; in particular the maximum increase is for 40 °C and 40% of relative humidity (10.250 MW) and the minimum one for 10 °C and 80% (1.383 MW);
gas turbine power output increase shows the same trend of the combined cycle one; in particular is should be noted that the biggest fraction of combined cycle increase (between 95 and 98%) is due to gas turbine power boost while steam turbine power improvement is very little, from a maximum of 0.5 MW (@ 40°C and 40% of RH) to a minimum of 33 kW (@ 10°C and 80% of RH);
the combined cycle efficiency is quite constant (from +0.16 to -0.06 percentage points of variation due to fogging application) and no influence of ambient conditions is observed;
fogging water consumption increases with ambient temperature increase and relative humidity decrease.
It could be observed that the obtained results are coherent with the ones achieved and
discussed in the paragraph 2.3.
101
2.5 – EVALUATION OF ITALIAN COMBINED CYCLE POWER PLANT SET OF UNITS WITH INLET FOGGING
The analysis methodology developed in the previous paragraphs could be used to study the effects of inlet fogging implementations on the Italian combined cycle power plants set of units. This study finds its reasons by considering that in the last years the Italian energy scenario was characterized by a constant increase in electric power demand. In Fig. 96 [78] the peaks of electric power demand during summer and winter months since the years 2000 to 2004 are reported.
45
50
55
60
2000 2001 2002 2003 2004
summer peakwinter peak
47.4
48.6
51.0
53.153.5
49.7
52.052.6
53.4 53.6
elec
tric
pow
er [G
W]
year Figure 96 – Electric power demand peaks during summer and winter since 2000 to 2004 [78]
Observing the trend in Fig. 96 it is quite evident the progressive alignment of the summer peaks of electric demand with the winter ones. According to the prediction of the GRTN (the manager of national transmission net), for the year 2005 the overtaking of summer peak on winter one is expected. This phenomenon could be explained by considering the change of life habits of Italian citizens and in particular taking into account the heavy diffusion of air conditioning systems.
The problem of the increasing request of electric power during the summer period is also amplified by the following problems:
the loss in power output of gas turbine and combined cycles power plants due to the high air temperature and to the limitations regarding both the discharge water temperature of the cooling systems and the amount of pollutant emissions;
the reduction of the electric power imported from foreign countries (France and Swiss above all) due to the high summer temperatures (that reduce the transit capacity of the interconnection electric lines) and to the maintenance operations (which in Europe are usually scheduled during August);
102
the not availability of different thousand of megawatts due to repowering operations of old steam power plants or for maintenance operations; this interventions will realize a great economic advantage for the next years but now reduce the available net power of Italian generation set of units.
For all this reasons different critical periods in which the power demand could exceed the production capacity could occur both during the winter and, above all, during the summer. In Fig. 97, an example of a possible trend of the demand and availability of power, (as estimated by the GRTN), is presented [78].
30
35
40
45
50
55
7 8 9 10 11 12 13April May June July August Sept Oct
available electricpower
electric power
demandelec
tric
pow
er [G
W]
months Figure 97 – Trend of available electric power and of the electric power demand for Italy (only continent)
From the previous figure, it could be observed a power gap of power which has a maximum value greater than 7 GW.
It should be observed that the trend in Fig. 97 should be evaluated by assuming or predicting the climatic conditions of the considered period; this means that the shape of the curves in Fig. 97 could be also very different from the one which is sketched. By the way, this problem remains topical.
The possible strategies to limit the gap between production and demand of electric power could be different; one way could be to increase the limits both on the temperature of cooling systems discharge water and pollutant emissions amount; this intervention allows to gain more power from the generation set of units but obviously causes an increase in the environmental impact. Another possibility could be a carefully management of the scheduled maintenance in order to have more power plants as possible, able to produce when the electric demand is higher. Also the use of hydro storage gives the possibility of gaining additional power.
In this scenario, the use of fogging could be an additional help to reduce the possible problems due to the gap of power. In fact, its main characteristics, such as the high
103
effectiveness, the short downtime for installation and the low impact on the operating costs, present fogging approach as a fast method to obtain an additional contribution of electric power without heavy intervention on the electric generation set of units.
Moreover, the use of fogging has no environmental impact, differently from the increase of discharge temperatures of cooling water and from pollutant emissions amount. 2.5.1 – Evaluation Of Italian Set Of Combined Cycle Units With Inlet Fogging
To evaluate the Italian combined cycle power plant set of units potential with fogging implementation the set of plants presented in Fig. 98 was considered. The figure shows the CC electric efficiency (ηel) as function of produced power output (Pel) with reference to the ISO conditions. More in detail, it consists of 70 units placed in 47 different sites with a net produced output at ISO conditions of about 13 GW and with an average ISO electric efficiency of about 54%.
The most widely machines result to be the GE MS9001 (in both models FA and E) and the Siemens V94.3A or V94.2A. Moreover a considerable number of GE LM 6000 and some models of GE MS 6001 FA and E or Siemens V64.3A are also installed.
The performance evaluation of the set of CC units was conducted by using a self developed code able to calculate the electric power output, the LHV power introduced and the water consumption due to the climatic conditions and their corresponding variations due to the fogging use.
0.45
0.50
0.55
0.60
0 50 100 150 200 250 300 350 400
3 pressure levels plus reheat2 pressure levels
Pel [MW]
Reference @ ISO conditions
number of units
12
4
5 3
4
15 3
52 2
10
23
ηel
Figure 98 – Italian set of combined cycle units: electric efficiency versus produced power output (Reference: ISO conditions)
104
2.5.2 – Climatic Assumptions A period of 13 weeks since the June 1 to the August 31 of 2003 was considered as
reference for the ambient conditions. In correspondence of the daily peak hours, (during what fogging is assumed to operates) the values of ambient temperature (peak temperature, PT) and relative humidity (peak relative humidity, PRH) were assumed for the 47 considered sites [79] where the plants are located. The minimum and the maximum value of PT and PRH on the 47 sites, are respectively presented in Figs. 99 and 100. Moreover in these figures the weighted averages of PT and PRH are also reported. They are calculated as:
average peak temperature: ∑
∑
=
=
⋅= 70
1ii,el
70
1ii,eli
P
PPTAPT (23)
average peak relative humidity: ∑
∑
=
=
⋅= 70
1ii,el
70
1ii,eli
P
PPRHAPRH (24)
where i =1 ÷ 70 identifies the total number of units and the produced electric power is evaluated at ISO conditions.
These two figures give an overview of the climatic scenario showing the distribution of temperature and relative humidity with the respect to the distribution of the installed electric power on the territory. For example, in the 11th week APT value equal to about 30 °C and APRH equal to about 40 % indicate that in this week there is high capability of air cooling.
20
22
24
26
28
30
32
34
1st 2nd 3rd 4th 5th 6th 7th 8th 9 th 10th 11th 12th 13th
PT [°
C]
week1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th
APT
Figure 99 – Minimum, maximum and weighted average value of PT [79]
20
30
40
50
60
70
80
90
1st 2n d 3rd 4th 5th 6t h 7th 8th 9th 10 th 11 th 12 th 13 th
PRH
[%]
week1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th
APRH
Figure 100 – Minimum, maximum and weighted average value of PRH [79]
105
2.5.2 – Results On the basis of the previous assumptions (Figs. 98, 99 and 100) the main results of
fogging application are presented in Fig. 101 which shows the values of electric power output (Pel), of LHV power introduced with fuel (Pi), of average electric efficiency (ηel) and of the water consumption (FWC) with (fog) and without fogging (w/o fog).
In particular, the results show a power boost ranging between 0.29 GW and 1.05 GW (corresponding to about 9% of increase with reference to no fogging case) with an increase in the LHV power introduced which vary from 0.45 GW to 1.64 GW. The change in the average electric efficiency is always less than 0.80 percentage points while the maximum amount of water consumption is about 101 kg/s. It should be noted that the fogging water consumption is calculated considering an efficiency of the spray system equal to 100% and then the values of FWC, reported in Fig.101, may be considered as minimum values.
To evaluate the predisposition of a certain site (or of a generation set of units) to operates with fogging, the “Fogging Site Efficiency” (FSE) parameter could be introduced. This parameter could be expressed as the ratio between the achieved power boost and the achievable one, evaluated assuming 40 °C of ambient temperature and 40 % of relative humidity, (corresponding to about 12 °C of cooling by considering the difference between the dry and the wet bulb temperature):
( )( ) %40,C40@el
RH,T@el
PP
FSE°Δ
Δ=
In Fig. 102 the minimum and maximum value of the “Fogging Site Efficiency” for
the assumed period, is shown. In this figure it is also plotted the weighted average of FSE (AFSE) on the basis of the Eq. 23.
The average values of FSE exceed the 60% only in few cases; this shows that, on average, the installed power plants work in ambient conditions not completely favorable to fogging, for the most part of the time. Moreover, it should be observed that the maximum value of AFSE (greater than 80%) is found at the 11th week in which the maximum power boost is realized.
In Fig. 103, the minimum and maximum cooling degrees (evaluated as difference between dry and wet bulb temperature) and the corresponding weighted average (ACD) is presented. It could be observed from this figure that, on average, the 47 evaluated sites have from 2 to 8 °C of cooling in the most part of the cases. Only in one case, the ambient conditions allow a cooling greater than 10 °C.
Finally it could be observed that the average specific power boost of the generation set of units results equal to about 1.4 MW per °C of cooling degree corresponding to about 9.9 MW per kg/s of water consumption.
106
0
50
100
150
0.0
1.3
2.7
4.0
1th 2nd 3r d 4th 5th 6th 7th 8t h 9th 10th 11th 12th 13t h
FWC
[kg/
s] FWC
[kg/s]
150
100
50
0
52.5
53.0
53.5
54.0
0 .0
1 .3
2 .7
4 .0
1 t h 2 n d 3 rd 4 th 5th 6th 7th 8 th 9 th 10 th 1 1th 12 th 13 th
η el [%
]
w/o fog
fog
11.0
11.5
12.0
12.5
13.0
13.5
14.0
14.5
15.0
15.5
16.0
19.0
19.5
20.0
20.5
21.0
21.5
22.0
22.5
23.0
23.5
24.0
1s t 2n d 3 rd 4 th 5 th 6 th 7 th 8 th 9 th 10 t h 1 1 th 12 th 1 3 th
P el [G
W]
Pi [G
W]
fog
w/o fog
w/o fog
fog
week1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th
Figure 101 – Electric power output (Pel), LHV power introduced with fuel (Pi), average electric efficiency (ηel), water consumption (FWC) with fogging (fog) and without it (w/o fog)
107
0
20
40
60
80
100
1st 2n d 3rd 4th 5th 6th 7th 8th 9th 10 th 11 t h 12 t h 13 th
FSE
[%]
week1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th
AFSE
Figure 102 – Minimum, maximum and weighted average value of FSE [79]
0
2
4
6
8
10
12
14
1st 2n d 3rd 4th 5th 6th 7th 8th 9th 10 th 11 th 12 th 13 th
CD
[°C
]
week1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th
ACD
Figure 103 – Minimum, maximum and weighted average value of CD [79]
108
3
Analysis Of The Performance Of A Gas Turbine Equipped With Interstage Injection By The Use Of A Commercial
Simulation Program
The study of wet compression (as overspray approach) developed in the chapter 2 could be improved by taking into account also the interstage injection.
The analysis of the previous chapter was possible by using the “gas turbine” module into the “gas turbine library” of the adopted commercial software [3]. This makes possible the study of commercial models of gas turbines with inlet fogging and overspray.
As the interstage approach is not directly evaluated by the commercial software, the construction of an aero - thermodynamic model of a gas turbine (GE Frame 7 EA) able to take into account this method, was realized.
The advantage of this approach to the study is the possibility of achieving more detailed results on the behavior of a gas turbine in case of wet compression on the respect of the previously performed analysis.
In particular this approach allows to obtain information about the change of temperature and pressure or about the percentage of water which evaporates, stage by stage and also about the change of compressor surge margin due to the water injection.
In conclusion, the study presented in this chapter whish to extend the previous one with more attention to the water evaporation process during the compression process.
109
3.1 – AERO-THERMODYNAMIC MODELING OF A GAS TURBINE The selected gas turbine, a mid-size and heavy-duty industrial gas turbine developed
for 60 Hz applications, consists of a seventeen stage axial compressor and a three stage axial turbine.
To evaluate the gas turbine performance in presence of interstage injection, a stage-by-stage performance analysis has been performed by using a commercial software package [3].
A schematic of the selected gas turbine highlighting various injection points examined in the study and the cooling air streams (CAS) is shown in Fig. 104.
WIP1
WIP
2
WIP
3
WIP
4
WIP
5
CAS5
WIP
6W
IP7
WIP
8
CAS17
AIR
1 2 3 4 5 6 7 17161514131211109 8
Figure 104 – GE Fr7EA simulation model
Each compressor stage of the selected gas turbine has been simulated by using a
general “gas/air compressor” module of the program [3]. This compressor modeling allows simulation of the water injection into the compressor and the stage-by-stage analysis of the compressor’s behavior in presence of the interstage injection (i.e., determination of the working point of each compressor stage, of the stage in which all the injected water vaporize, etc.). The first injection point, WIP1, is considered to compare the interstage injection effects with the conventional evaporative fogging.
In order to make the Fr7EA gas turbine simulation model able to run, it is necessary to introduce a compressor stage performance map in each “gas/air compressor” component used in the model. These stage performance maps have been determined as explained in detail in the next section. A “gas/air turbine” component was also used to simulate the three stage Fr7EA turbine considering a default turbine performance maps
)FF,CS(F ***β=β . This map is presented in Fig. 105 [3]. The flow function (FF) is
defined as:
110
1k1k
0
1k2
kMWR
pTITmFF
−+
⎟⎠⎞
⎜⎝⎛
+
⋅⋅=&
where: R0 = Universal gas constant MW = Molecular weight k = cP/cV ratio between specific heat at constant pressure and volume
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.85 0.90 0.95 1.00 1.05 1.10
β*
FF*
CS*=1.1
1.0 0.9 0.8 0.7 0.6 0.5 0.4
Figure 105 – Default gas/air turbine performance map
)FF,CS(F ***β=β [3]
In Fig. 104 the two compressor air bleed ports (downstream from 5th and 17th stage)
for turbine blade cooling, are also highlighted; to take into account the effect of turbine blade cooling on gas turbine performance, the overall turbine cooling flow has been divided into two parts, considering that one mixes upstream and the other downstream the turbine [80, 81].
111
In particular it has been assumed that: the air extracted from 5th stage (CAS5) is mixed with the exhaust gases
downstream the turbine; the air extracted from the 17th stage (CAS17) is mixed with the gas upstream the
turbine. The main performance and thermodynamic parameters of Fr7EA gas turbine were
evaluated, by considering constant turbine inlet temperature, inlet guide vanes (IGVs) fully opened and neglecting inlet and exhaust pressure losses, in correspondence of two different climatic conditions: ISO case (ambient temperature 15 °C, pressure 1.013 bar and relative humidity 60 %) and HOT case (40 °C, 1.013 bar and 40 % RH). These data are presented in Tab. 19.
Table 19 – Comparison of the main performance and thermodynamic parameters of Fr7EA Thermoflex model at ISO and HOT conditions for constant TIT, IGV fully opened in absence of inlet and exhaust pressure losses
ISO case HOT case Diff. Power Output [kW] 85443 68203 -20% Heat Rate [kJ/kWh] 10991 11803 +7% Air mass flow rate [kg/s] 291.7 285.5 -2% Pressure Ratio 12.62 11.61 -9% Compressor outlet temperature [°C] 362 396 +34 °C Exhaust temperature [°C] 532 549 +17 °C
3.1.1 – Water Evaporation And Wet Compression Considerations
Due to the use of a commercial software package [3] for the simulation, the evaporation model used in this software was adopted. In this evaporation model, a thermal equilibrium between the two flows is assumed to occur instantaneously at the injection point. As a result, effects of droplet dynamics and evaporation rate are not considered.
The different injection points through the compressor have been modelled as represented in Fig. 104. The first injection (WIP1) occurs upstream of the compressor and it should be considered equivalent to a traditional evaporative fogging. The other injection points (from WIP2 to WIP8 in Fig. 104) are realized upstream from stage 2 to 8.
Some additional considerations about the calculation method of the used commercial software could be useful. In order to evaluate the behaviour of gas turbine model in the case of wet compression, the “mixer” and the “gas/air compressor” are the key components.
A “mixer” component with liquid water and air at inlet, simply mixes the two streams. The result of the mix is a stream that may contain two phases. In this case, (the liquid hasn't been completely evaporated), both phases must have the same temperature and the same pressure. Additionally, the gaseous phase must be at saturation (100% of relative humidity). The conditions that are satisfied by the “mixer” are the conservation of mass, energy (enthalpy balance) and the equilibrium of the water vapour in the air with the liquid water. The final temperature of the exit flow is determined as function of the pressure, of the relative amounts of air and water, and of enthalpy balance. In particular it should be noted that:
112
if both fluids enter at the same temperature and pressure, and if the air is already completely saturated at the “mixer” inlet (100% of relative humidity), no change in the temperature between inlet and outlet are obtained;
if both fluids enter at the same temperature and pressure, but the air is not completely saturated at the “mixer” inlet (less than 100% of relative humidity), the temperature of the exit stream reduces with respect to the inlet stream temperature due to the air saturation process;
the evaporation process in the “mixer” component certainly involves a convection thermal exchange evaluation but no dynamic process (such as water evaporation velocity, air and water contact surface, etc.) are considered. The last sentence means that no liquid water can enters into the compressor stage if the air saturation is not already reached.
For what concern the calculation of a “gas/air compressor” with liquid water at its
inlet, it could be assumed that in correspondence of each increment of temperature, a certain amount of water evaporates, cooling the air, in order to always have the 100% of relative humidity [3]. Obviously the total amount of liquid water that could be evaporated depends on the polytropic efficiency of the compressor.
113
3.2 – COMPRESSOR STAGE PERFORMANCE MAPS The determination of performance maps for each gas/air compressor component
used for Fr7EA gas turbine model has been carried out by a trial and error procedure in order to reproduce the performances of the real Fr7EA gas turbine in design and off-design conditions. This makes possible the use of simplified relationships (such as to consider constant values of the mean specific heats at constant pressure) without introducing unacceptable approximation in the performance calculations. In fact, independently of the initial guess assumptions and of the relationships used, the final result of the trial and error procedure are the performance maps of the compressor components which allows the obtainment of the Fr7EA gas turbine performances with the desired accuracy.
The first step of Fr7EA compressor modeling is the calculation of the design pressure ratio for each stage. For this aim, having no compressor design details available, it was assumed that, at the design condition, all the seventeen compressor stages are characterized by the same design value of the pressure coefficient:
( ) CONSTU
1Tc
Uh
DES
2
cR
SS,ip
DES
2is,S
DESp
S,p
=
⎥⎥⎥⎥⎥⎥
⎦
⎤
⎢⎢⎢⎢⎢⎢
⎣
⎡
⎟⎟
⎠
⎞
⎜⎜
⎝
⎛−β⋅⋅
=⎟⎟⎠
⎞⎜⎜⎝
⎛ Δ=ψ (25)
where: ΔhS,is: isentropic enthalpy variation across each stage βS: stage pressure ratio cp,S: specific heat at constant pressure (mean value between stage inlet and exit
temperatures) Ti,S: compressor stage inlet temperature
Obtaining βS from the Eq. 25, it results that:
( )
DES
Rc
S,ip
2p
DESS
S,p
TcU
1⎥⎥⎥
⎦
⎤
⎢⎢⎢
⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
⋅
⋅ψ+=β (26)
It can be noted that the temperature at the inlet of each stage (Ti,S) is virtually equal to the exit temperature of the preceding stage, that, in turn, may be evaluated starting from inlet temperature, pressure ratio and polytropic efficiency of the stage.
Therefore, if Ti,C is the air temperature at the compressor intake (inlet of the first compressor stage), the overall compressor pressure ratio (βTOT) can be written as:
114
( ) ( )
DES
n
1j
Rc
1j
1k
cR
Sk,ipk
2kp
C,ipj
2jp
n
1j
DES
Rc
Sj,ipj
2jp
n
1jDESSjDESTOT
S
pj
pkpk
S
pj
S
TcU
1Tc
U1
TcU
1 ∏∏
∏
∏
= −
=
η=
=
⎪⎪⎪⎪
⎭
⎪⎪⎪⎪
⎬
⎫
⎪⎪⎪⎪
⎩
⎪⎪⎪⎪
⎨
⎧
⎥⎥⎥⎥⎥⎥⎥
⎦
⎤
⎢⎢⎢⎢⎢⎢⎢
⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
⋅
⋅ψ+⋅
⋅ψ+=
⎥⎥⎥
⎦
⎤
⎢⎢⎢
⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
⋅
⋅ψ+=
=β=β
(27)
Equation 27 could be used to calculate the value of ( Ψ p)DES (which was assumed to
be the same for all the compressor stages) allowing the achievement of the desired design compressor pressure ratio. To solve Eq. 27, the small stage efficiency of each stage is also required. Assuming this last term as constant for each compressor stage and, therefore, equal to the polytropic efficiency of the overall multistage compressor, it can be calculated if the compressor discharge temperature is known. Applying Eq. 27 to the Fr7EA compressor, and assuming a compressor discharge temperature equal to 362 °C (that means a polytropic – or small stage – efficiency of about 90%), it results that: ( ) 208.0
DESp =ψ
The corresponding values of the stage pressure ratios in design conditions determined by using Eq. 26, are reported in Tab. 20.
Table 20 – Calculated design stage pressure ratio of axial compressor of GE Fr7EA gas turbine STAGE Stage pressure ratio STAGE Stage pressure ratio
1st 1.273 10th 1.141 2nd 1.245 11th 1.135 3rd 1.222 12th 1.129 4th 1.202 13th 1.124 5th 1.185 14th 1.119 6th 1.174 15th 1.114 7th 1.165 16th 1.110 8th 1.156 17th 1.106 9th 1.148
In order to determine the performance maps of each compressor stage (which link together stage pressure ratio, efficiency, corrected mass flow and corrected rotational speed), generalized relationships between pressure coefficient Ψ p
*= Ψ p/( Ψ p)DES, stage isentropic efficiency η*=η/(η)DES and flow coefficient φ *=φ /(φ )DES (all normalized to the design values) were used [82]. These relationships allow the complete evaluation of stage characteristics once the stage design point (( Ψ p)DES, (η)DES, (φ )DES) is known.
115
The first generalized relationship, [82]:
( )[ ] [ ( ) ] 1SF 11SF
1 2***2**
*max,p*
max,p*p max,pmax,p
max,pmax,p
φ−−φ⋅+φ⋅−−φ⋅+φ
−ψ−ψ=ψ ψψ
ψψ
(28)
where, “SF” is the “Shape Factor”, which allows the representation of different types of compressor stages.
In Fig. 106 three curves obtained using Eq. 28 with ]1,5.0[SF −∈ are shown.
0.2
0.4
0.6
0.8
1.0
1.2
0.4 0.6 0.8 1.0 1.2 1.4 1.6
SF=1.0
SF=0.0
SF=-0.5
ψ∗
P
φ∗
Figure 106 – Generalized stage characteristics
Values of 5.0SF −= for all the compressor stages have been determined at the end of
the trial and error procedure in order to reproduce the real Fr7EA gas turbine performances with the desired accuracy.
The second generalized relationship, was obtained from the generalized stage efficiency curve proposed by Howell and Bonham [83], by using the procedure presented in [82]:
( )
⎥⎥⎦
⎤
⎢⎢⎣
⎡⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ∈
φ
ψ⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ−
⎥⎥⎦
⎤
⎢⎢⎣
⎡⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ−
η−−=η
φψ1
***1
*1
11*
min
*p
*p
5.3*p
5.3
min
*p
/*
minp ,, (29)
( )
⎥⎥⎦
⎤
⎢⎢⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ∈
φ
ψ⎟⎟⎠
⎞⎜⎜⎝
⎛−
φ
ψ
⎥⎥⎦
⎤
⎢⎢⎣
⎡−⎟
⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ
η−−=η
φψ
max
*p
*p
2*p
2
max
*p
/
*1
*1
*1
*
11*
*maxp , , (30)
116
In Fig. 107 the generalized stage efficiency curve obtained by using Eqs. 29 and 30 is reported.
0.0
0.2
0.4
0.6
0.8
1.0
1.2
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6
η∗
ψp
∗/ φ∗
Figure 107 – Generalized stage efficiency curve
The stage performance maps in the form required by gas/air compressor components (βS
*=F(CFS*, CSS
*) and ηS*=F(CFS
*, CSS*), where βS
*, ηS*, CFS
* and CSS* are the
stage pressure ratio, isentropic efficiency, corrected mass flow rate and corrected rotational speed respectively, all normalized to the design value) have been calculated, starting from the generalized stage characteristics (Eqs. 28, 29 and 30), by using the following relationships:
( ) ( )( )DESS
Rc
S,ip
2*SDESis,S
*p*
S11
TcCSh
p
β⋅
⎥⎥
⎦
⎤
⎢⎢
⎣
⎡+
⋅Δ⋅ψ=β (31)
*S
**S CSCF ⋅φ= (32)
In the determination of the stage performance maps by using Eqs. 31 and 32 it was considered that:
the temperature and pressure at each stage inlet remain constant and equal to the design value;
the gas composition is the same for each stage and corresponds to the composition of air at ISO conditions.
The use of the stage performance maps obtained by using these equations introduces, however, an error when the operating gas composition changes with respect to the design one, such as in the case of inlet fogging and interstage water injection. In fact, while, as well known, the influence of gas composition on the generalized relationships expressed by the Eqs. from 28 to 30 can be neglected [84], the transformation of these
117
relationships into the ones required by the gas/air compressor component by the means of Eqs. 31 and 32 is influenced by the gas composition, due to the presence of the specific heat at constant pressure (cp) and gas constant (R) in Eq. 31. In particular, the variations of humid air composition due to variations of water content are more significant in the later compressor stages, where the temperature is higher and the water vapour amount which can be contained in the air is higher. An analysis was then performed to evaluate the stage performance maps modification due to variations of humid air composition. In particular, for the 17th stage, this analysis showed that, for a variation of the water content of 100% with respect of air water content in ISO conditions, the root mean square error on βS
* values evaluated by means of Eq. 31 is lower than 0.03 %. Therefore, it was deemed acceptable to neglect the influence of humid air composition on stage performance maps.
Once the compressor stage performance maps have been determined, the overall multistage compressor performance maps are obtained by matching all the stages by means of a “stage-stacking” procedure [82]. The graphs of all the seventeen compressor stage performance maps are reported in the Figs. from 108 to 124.
The values along the compressor of the temperature and of the ratio between the pressure and the pressure inlet value (1.013 bar) in both ISO and HOT cases are reported in Fig. 125 and Fig. 126 respectively.
118
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.2
1.11.0
0.90.8
0.70.60.50.40.3
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.1
1.00.9
0.80.70.6
0.50.40.3
Figure 108 – 1st stage performance map Figure 109 – 2nd stage performance map
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.1
1.00.9
0.80.7
0.60.50.40.3
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.1
1.00.90.8
0.70.60.50.40.3
Figure 110 – 3rd stage performance map Figure 111 – 4thstage performance map
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.0
0.90.80.70.60.50.40.3
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.0
0.90.80.70.60.50.40.3
Figure 112 – 5th stage performance map Figure 113 – 6th stage performance map
119
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.0
0.90.80.70.60.50.40.3
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.1
1.00.90.8
0.70.60.50.40.3
Figure 114 – 7th stage performance map Figure 115 – 8th stage performance map
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.00.9
0.80.70.60.50.40.3
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.0
0.90.80.70.60.50.40.3
Figure 116 – 9th stage performance map Figure 117 – 10th stage performance map
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.0
0.90.80.70.60.50.40.3
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.0
0.90.80.70.60.50.40.3
Figure 118 – 11th stage performance map Figure 119 – 12th stage performance map
120
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.00.90.80.70.60.50.40.3
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.00.90.80.70.60.50.40.3
Figure 120 – 13th stage performance map Figure 121 – 14th stage performance map
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.00.90.80.70.60.50.40.3
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.00.90.80.70.60.50.40.3
Figure 122 – 15th stage performance map Figure 123 – 16th stage performance map
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.2 0.4 0.6 0.8 1.0 1.2 1.4
β*
CF*
CS*=1.21.11.00.90.80.70.60.50.40.3
Figure 124 – 17th stage performance map
121
0
100
200
300
400
500
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
HOT case
ISO case
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
Figure 125 – Temperature profile along the compressor in ISO and HOT cases
123456789
10111213
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Pres
sure
ratio HOT case
ISO case
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
Figure 126 – Ratio between the pressure along the compressor and the inlet value (1.013 bar)
122
3.3 – GAS TURBINE PERFORMANCE WITH INTERSTAGE WATER INJECTION
The study of GE Frame 7EA gas turbine with interstage water injection was
developed referring to the previously defined two ambient conditions of ISO and HOT cases. The water injection was realized according to the schemes presented in Fig. 104.
The performance analyses were conducted considering the following assumptions: IGVs are maintained fully open (same as at the base-load and dry operation); gas turbine is regulated at constant TIT; inlet and exhaust pressure losses are neglected; the temperature and pressure of the injected water have been assumed always
equal to 15 °C and 30 bar, respectively; a thermodynamic equilibrium is assumed at the injection location (therefore,
effects of droplet dynamics have been neglected).
The simulation results, obtained with the above mentioned assumptions, are reported in Fig. 127 and Figs. from 131 to 133 and from 166 to 168 with dotted and continuous lines representing ISO and HOT conditions, respectively.
Figure 127 shows the amount of power boost as a function of injected water volume flow rate for different locations of water injection. The calculated value of power output with water injection (POUT) is normalized with the power output corresponding to the dry ISO case (POUT, ISO). This representation of power boost clearly identifies the amount of required injected water to recover the power loss due to high ambient temperature. For example, approximately 1.3% of water to air ratio injected at compressor stage #2 (injection point WIP2) is required to recover the lost power for the HOT case, as is evident from Fig. 127. It must be noted that the gas turbine lost approximately 20% power output due to high ambient temperature (HOT case) compared to the ISO case as shown in Fig. 127.
123
0.80
0.84
0.88
0.92
0.96
1.00
1.04
1.08
1.12
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8
WIP1
WIP2 WIP3
WIP4WIP5
water and air mass flow rate ratio [%]
ISOcase
HOTcase surge
limit
max powerlimit
WIP1
WIP2
WIP4
WIP5
WIP3
WIP6
WIP7
WIP8
WIP6
WIP7
WIP8
Pout
Pout,ISO
Figure 127 – Gas turbine power output change as function of injected water to air mass flow rate ratio and injection point
It should be noted that the total amount of water that is possible to inject into the compressor is mainly limited by two factors: (i) maximum acceptable gas turbine power output associated with the gas turbine mechanical integrity and the electric generator capacity, (ii) and the surge limit in the compressor. In the present study, the maximum power output condition conservatively limits POUT/POUT, ISO value to 1.08 which occurs in the ISO case. The second condition instead limits POUT/POUT, ISO to different values dependent on injected water mass flow rate and water injection location (see Fig. 127).
It is important to specify that the surge limit has not to be confused with the surge line. For a given corrected speed line, the surge limit is here calculated by augmenting the corrected flow at which the surge occurs by 5 %; this means that the surge limit is a point in which the compressor is able to operate.
It should be further observed from Fig. 127 that to obtain the same increase in gas turbine power output, water consumption increases as the injection point moves from upstream of the first stage to the eighth stage (both in ISO and HOT cases). This observation implies that, for a given amount of injected water mass flow rate, the power boost per unit of water consumption ( )( )INJISO,OUTOUT mPPPB &−= decreases as the injection point moves from the 1st to the 8th stage. Moreover, for a given injection point, the power boost per unit of water consumption decreases as the total amount of injected water increases.
Before discussing other performance parameters of interest, it would be important to examine effects of injection points and injected water amount on the compressor stage performance map. Performance characteristics of the two compressor stages, 1st and 17th stages, with varying amount of water injection at the three injection locations, WIP1, WIP2 and WIP5, for the ISO and HOT cases are shown in Fig. 128. For the first stage
124
with water injection upstream of the compressor, an increase in corrected speed occurs because of inlet cooling effect for the ISO case (Fig. 128 (a) ). As expected, a further increase in injected water flow accompanies decreased stage pressure ratio. In case of water injection points WIP2 and WIP5, operating points for various water injection flow rates lie on the design value of corrected speed since no inlet cooling occurs upstream of the first stage. Thus, for the ISO case, unloading (also indicated by increase in stage flow coefficient as shown in Fig. 129 (a) ) in the first stage occurs irrespective of location of the injection and the amount of injected water.
In the 17th stage for the ISO case, the corrected speed increases for each injection point because of a decrease in the air temperature as a result of continuous evaporative cooling in various compressor stages. Increase in speed compared to the design value results in increased pressure ratio as is evident from Fig. 128 (b). A decrease in stage flow coefficient indicative of increased stage loading is also evident from Fig. 129 (a). Redistribution of stage loading, unloading of first few stages and increased loading of later stages of the compressor, can be clearly seen in Figs. 129 (a) and 129 (b). Figures 129 (a) and 129 (b) respectively show the stage by stage values of flow coefficient and pressure ratio for the maximum allowed values of injected water flow rate at WIP1, WIP2 and WIP5.
The overall changes on 1st and 17th stage performance for the HOT case due to interstage injection are similar to that observed for the ISO case as is evident from Figs. 128 (c) and 128 (d). Some of the differences observed for the HOT case are briefly discussed here. Because of the high ambient temperature in the HOT case and without water injection, the operating point ‘A’ has reduced value of corrected speed and normalized pressure ratio compared to the design case as shown in Fig. 128 (c). Compared to the ISO case, a larger amount of injected water is required to achieve stage flow coefficients higher than the compressor design values in the 1st stage, as can be seen comparing Fig 129 (a) and 130 (a). As observed for the ISO case, stage pressure ratio increases compared to the dry case (operating point ‘A’) in the 17th stage for HOT case because of cooling effect inside the compressor due to interstage injection as is shown in Fig. 127 (d). For the HOT case, it is further noted that pressure ratio in the 17th stage increases as the injection point moves upstream of the compressor as is evident from Fig. 127 (d).
125
0.92
0.94
0.96
0.98
1.00
1.02
1.04
0.84 0.88 0.92 0.96 1.00 1.04
βS
*
CF*
CS*=1.0
0.9
surge line
surge limit
Reference: ISO case (a)
A
B
CB
CD
0.92
0.94
0.96
0.98
1.00
1.02
1.04
0.84 0.88 0.92 0.96 1.00 1.04
CF*
CS*=1.0
0.9
surge line
surge limit
A
B
C
D
E
F
B
C
D
E
B
C
D
E
F
Reference: HOT case (c)β
S*
0.98
0.99
1.00
1.01
1.02
1.03
1.04
0.94 0.96 0.98 1.00 1.02 1.04 1.06
WIP1
WIP2
WIP5
CF*
surge line
CS*=1.0
1.1
DCBA-0.400.330.00WIP1-0.450.340.00WIP2
0.780.680.340.00WIP5
water to air ratio [%] (ISO case)
Reference: ISO case (b)
A
B
C
B
CD
βS
*
surge limit
0.98
0.99
1.00
1.01
1.02
1.03
1.04
0.94 0.96 0.98 1.00 1.02 1.04 1.06
WIP1
WIP2
WIP5
CF*
surge line
CS*=1.0
A
B
C
D
E
F
1.1
FEDCBA1.771.401.100.700.400.00WIP11.681.401.100.700.400.00WIP2
1.461.100.800.400.00WIP5
water to air ratio [%] (HOT case)
Reference: HOT case (d)β
S*
surge limit
Figure 128 – (a) First stage performance maps and operating points with water injection (Reference: ISO case); (b) Seventeenth stage performance maps and operating points with water injection (Reference: ISO case); (c) First stage performance maps and operating points with water injection (Reference: HOT case); (d) Seventeenth stage performance maps and operating points with water injection (Reference: HOT case)
126
0.96
0.97
0.98
0.99
1.00
1.01
1.02
1.03
1.04
1.05
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
WIP2, C
WIP5, D
φ*
compressor stage @ inlet
Reference: ISO CASE
WIP1, C
(a)
A
0.97
0.98
0.99
1.00
1.01
1.02
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
WIP2, C
WIP5, D
βs*
compressor stage
Reference: ISO CASE
WIP1, C
(b)
A
Figure 129 – (a) Normalized flow coefficient trend at the inlet of each stage as function of injection point and water amount (ISO case); (b) Normalized stage pressure ratio as function of injection point and water amount (ISO case)
0.96
0.97
0.98
0.99
1.00
1.01
1.02
1.03
1.04
1.05
1.06
1.07
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
WIP2, B
A
WIP5, B
WIP5, E
WIP2, F
φ*
compressor stage @ inlet
Reference: HOT CASEWIP1, F
WIP1, B
(a)
0.94
0.95
0.96
0.97
0.98
0.99
1.00
1.01
1.02
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
WIP2, B
A
WIP5, B
WIP5, E
WIP2, F
βs*
compressor stage
Reference: HOT CASE
WIP1, F
WIP1, B
(b)
Figure 130 – (a) Normalized flow coefficient trend at the inlet of each stage as function of injection point and water amount (HOT case); (b) – Normalized stage pressure ratio as function of injection point and water amount (HOT case)
127
The variation of normalized heat rate, as a function of injected water volume flow rate and the locations of injection, shows that gas turbine cycle efficiency improves with various water injection locations for both ambient cases (ISO and HOT cases) as shown in Fig. 131. For the ISO case, it is further noted that the gas turbine heat rate is always lower than the dry case (without the water injection) in spite of the fact that its value increases as the injection point moves downstream in the compressor. For a given amount of injected water flow rate, the maximum improvement in gas turbine efficiency is seen with water injected upstream of the compressor (WIP1)compared to the other locations examined for both the ISO and HOT cases. In the HOT case, the design value of HR is reached only in the case of WIP1 with a water and air mass flow ratio of about 1.4%. This value is greater than that required to recover the loss in power output due to high ambient temperature (about 1.2 % of water and air flow rate ratio at WIP1, Fig. 127) and suggests that the interstage injection method is more effective to gain power than the gas turbine efficiency. The observed improvement in the values of HR with interstage injection approach can be attributed to the performance enhancement in the compressor section as will be seen later in Fig. 135.
0.960.970.980.991.001.011.021.031.041.051.061.071.08
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8water and air mass flow rate ratio [%]
WIP1 WIP5WIP4WIP3WIP2
ISO
surgelimit
HOT
max power limit
WIP1WIP2
WIP3
WIP5
WIP4
WIP6
WIP8 WIP7
WIP6WIP7WIP8HR
HRISO
Figure 131 – Gas turbine heat rate change as function of injected water to air mass flow rate ratio and injection point
Another effect of water injection is the increase of overall compressor pressure ratio (Fig. 132) and of the total air mass flow rate at the gas turbine inlet (Fig. 133) which are mainly caused by the cooling effects inside the compressor. Also for these two parameters the maximum increase is obtained at WIP1 and the highest water injection mass flow rate for both ambient conditions (ISO and HOT cases), as can be seen in Figs. 132 and 133.
A more detailed analysis of temperature and pressure profile along the compressor in correspondence of the maximum amount of injected water both in ISO and HOT cases are reported in Figs. from 134 to 149 and from 150 to 165 respectively.
128
11.6
11.8
12.0
12.2
12.4
12.6
12.8
13.0
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8
WIP1WIP2
WIP3WIP4 WIP5
water and air mass flow rate ratio [%]
β
ISO
HOT
max powerlimit
surgelimit
WIP1
WIP2
WIP3WIP4
WIP5
WIP6 WIP7 WIP8
WIP6WIP7
WIP8
Figure 132 – Compressor pressure ratio as function of injected water to air mass flow rate ratio and injection point
0.85
0.90
0.95
1.00
1.05
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8water and air mass flow rate ratio [%]
mAIR
mAIR,ISO
WIP1WIP2 WIP3
WIP4 WIP5
ISO
HOT
max power limit
surgelimit
WIP1
WIP2WIP3
WIP4WIP5
WIP6
WIP7 WIP8
WIP6 WIP7 WIP8
Figure 133 – Air mass flow rate at compressor inlet as function of injected water to air mass flow rate ratio and injection point
129
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP1
No injection
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
ISO case
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP2
No injection
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
ISO case
Figure 134 – Temperature profile along the compressor in ISO case and in ISO case with 1.20 kg/s of water injected @ WIP1
Figure 135 – Temperature profile along the compressor in ISO case and ISO case with 1.35 kg/s of water injected @ WIP2
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP3
No injection
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
ISO case
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]WIP4
No injection
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
ISO case
Figure 136 – Temperature profile along the compressor in ISO case and ISO case with 1.65 kg/s of water injected @ WIP3
Figure 137 – Temperature profile along the compressor in ISO case and ISO case with 1. 97 kg/s of water injected @ WIP4
130
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP5
No injection
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
ISO case
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP6
No injection
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
ISO case
Figure 138 – Temperature profile along the compressor in ISO case and ISO case with 2.30 kg/s of water injected @ WIP5
Figure 139 – Temperature profile along the compressor in ISO case and ISO case with 2.60 kg/s of water injected @ WIP6
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP7
No injection
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
ISO case
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP8
No injection
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
ISO case
Figure 140 – Temperature profile along the compressor in ISO case and ISO case with 2.80 kg/s of water injected @ WIP7
Figure 141 – Temperature profile along the compressor in ISO case and ISO case with 3.00 kg/s of water injected @ WIP8
131
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C]
WIP1
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
HOT case
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C]
WIP2
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
HOT case
Figure 142 – Temperature profile along the compressor in HOT case and HOT case with 5.00 kg/s of water injected @ WIP1
Figure 143 – Temperature profile along the compressor in HOT case and HOT case with 4.70 kg/s of water injected @ WIP2
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP3
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
HOT case
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP4
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
HOT case
Figure 144 – Temperature profile along the compressor in HOT case and HOT case with 4.50 kg/s of water injected @ WIP3
Figure 145 – Temperature profile along the compressor in HOT case and HOT case with 4.10 kg/s of water injected @ WIP4
132
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C]
WIP5
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
HOT case
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C]
WIP6
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
HOT case
Figure 146 – Temperature profile along the compressor in HOT case and HOT case with 3.90 kg/s of water injected @ WIP5
Figure 147 – Temperature profile along the compressor in HOT case and HOT case with 3.80 kg/s of water injected @ WIP6
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP7
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
HOT case
0
50
100
150
200
250
300
350
400
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
compressor stage
Tem
pera
ture
[°C
]
WIP8
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
HOT case
Figure 148 – Temperature profile along the compressor in HOT case and HOT case with 3.70 kg/s of water injected @ WIP7
Figure 149 – Temperature profile along the compressor in HOT case and HOT case with 3.60 kg/s of water injected @ WIP8
133
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP1
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP1ISO case
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP2
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP2ISO case
Figure 150 – Pressure ratio profile along the compressor in ISO case and ISO case with 1.20 kg/s of water injected @ WIP1
Figure 151 – Pressure ratio profile along the compressor in ISO case and ISO case with 1.35 kg/s of water injected @ WIP2
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP3
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP3ISO case
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP4
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP4ISO case
Figure 152 – Pressure ratio profile along the compressor in ISO case and ISO case with 1.65 kg/s of water injected @ WIP3
Figure 153 – Pressure ratio profile along the compressor in ISO case and ISO case with 1.97 kg/s of water injected @ WIP4
134
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP5
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP5ISO case
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP6
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP6ISO case
Figure 154 – Pressure ratio profile along the compressor in ISO case and ISO case with 2.30 kg/s of water injected @ WIP5
Figure 155 – Pressure ratio profile along the compressor in ISO case and ISO case with 2.60 kg/s of water injected @ WIP6
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP7
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP7ISO case
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP8
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP8ISO case
Figure 156 – Pressure ratio profile along the compressor in ISO case and ISO case with 2.80 kg/s of water injected @ WIP7
Figure 157 – Pressure ratio profile along the compressor in ISO case and ISO case with 3.00 kg/s of water injected @ WIP8
135
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP1
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP1HOT case
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 1 1 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP2
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP2HOT case
Figure 158 – Pressure ratio profile along the compressor in HOT case and HOT case with 5.00 kg/s of water injected @ WIP1
Figure 159 – Pressure ratio profile along the compressor in HOT case and HOT case with 4.70 kg/s of water injected @ WIP2
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP3
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP3HOT case
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP4
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP4HOT case
Figure 160 – Pressure ratio profile along the compressor in HOT case and HOT case with 4.50 kg/s of water injected @ WIP3
Figure 161 – Pressure ratio profile along the compressor in HOT case and HOT case with 4.10 kg/s of water injected @ WIP4
136
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP5
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP5HOT case
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP6
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP6HOT case
Figure 162 – Pressure ratio profile along the compressor in HOT case and HOT case with 3.90 kg/s of water injected @ WIP5
Figure 163 – Pressure ratio profile along the compressor in HOT case and HOT case with 3.80 kg/s of water injected @ WIP6
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP7
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP7HOT case
1.05
1.10
1.15
1.20
1.25
1.30
0
3
6
9
12
15
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Stag
e Pr
essu
re R
atio
Stage Discharge Pressure [bar]
WIP8
1st 2nd 3rd 4th 5th 6th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
No injection
No injection
WIP8HOT case
Figure 164 – Pressure ratio profile along the compressor in HOT case and HOT case with 3.70 kg/s of water injected @ WIP7
Figure 165 – Pressure ratio profile along the compressor in HOT case and HOT case with 3.60 kg/s of water injected @ WIP8
137
In Fig. 166 the trend of the turbine outlet temperature is presented: as evident from this figure, the maximum reduction of turbine outlet temperature ranges, in the HOT case, between 3 °C (about 1.4 % of water to air flow rate ratio at WIP8) and 15 °C (about 1.8 % of water to air flow rate ratio at WIP1), whereas, in the ISO case, from 4 °C (1.0 % of water to air flow rate ratio at WIP8) to 12 °C (0.4 % of water to air flow rate ratio at WIP1).
520
525
530
535
540
545
550
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8
WIP1
WIP2WIP3 WIP4
WIP5
water and air mass flow rate ratio [%]
TOT[°C]
HOT
ISO
max power limit
surgelimit
WIP1WIP2
WIP3WIP4
WIP5
WIP6
WIP7 WIP8 WIP6
WIP7WIP8
Figure 166 – Gas turbine outlet temperature as function of injected water to air mass flow rate ratio and injection point
As noted earlier, water injected upstream or within the compressor influences performance characteristics of the compressor. Compressor performance is improved as a result of water injection in the form of reduced compression work as is evident from Fig. 167. This reduction in compression work is mainly caused due to cooling effects in the compressor. For the present study, the maximum reduction in compression work is achieved for water injection taking place upstream of the compressor compared to the other injection locations (Fig. 167).
The variation of the turbine specific work as a function of injected water volume flow rates and injection locations shows that a small amount of water flow rate is required to recover the lost work due to high ambient condition in comparison to the amount of water required to affect the compression work, as is evident from Figs. 167 and 168.
Arsen’ev and Berkovich showed that there exists a interstage location for which power boost will be higher compared to the injection location upstream of the compressor [52]. Furthermore, if the injection location is moved downstream of the optimum location, power boost due to the interstage injection decreases compared to the injection point located upstream of the compressor. In the present study, however, power boost has been found largest with water injected upstream of the compressor compared to the other injection points. These differences in the two studies can be due to the simplified assumption of evaporation model in the present study including differences in the design of gas turbine’s compressor section of the two studies.
138
0.96
0.97
0.98
0.99
1.00
1.01
1.02
1.03
1.04
1.05
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8water and air mass flow rate ratio [%]
WC
WC,ISO
ISO
HOT
WIP6
WIP7WIP8
WIP5
WIP1WIP2WIP3WIP4
WIP1WIP2WIP3WIP4
WIP6WIP7
WIP8
Figure 167 – Compressor specific work change as function of injected water to air mass flow rate ratio and injection point
0.96
0.97
0.98
0.99
1.00
1.01
1.02
1.03
1.04
1.05
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8water and air mass flow rate ratio [%]
WT
WT,ISO
WIP1WIP2
WIP3WIP4WIP5
ISO
max power limit
surgelimit
WIP1WIP2 WIP3
WIP4
WIP5
WIP6, 7, 8
WIP6WIP7
WIP8
HOT
Figure 168 – Turbine specific work change as function of injected water to air mass flow rate ratio and injection point
139
The last result that is presented is the evaluation of the point of the compressor in which the complete evaporation of the water injected occurs. In Fig. 169 a graph is presented that shows all the evaporation points of injected water as function of the injection point (from WIP1 to WIP8), injected water volume flow rate and climate conditions (ISO and HOT cases).
As example, considering 1.5 % of water to air flow rate ratio, injected at WIP1 in the HOT case, it could be observed from the graph that about the 0.5 % of water to air flow ratio evaporates upstream of the 1st stage and the remaining 1.0 % enters into the compressor where the evaporation completes.
As already explained the interstage injection here presented is studied without taking into account the dynamic aspects of droplet evaporation. Nevertheless the presented results could be extended considering the delay of evaporation due to the droplet diameter and air velocity.
For example, if is kwon that a certain amount of water, injected at WIP2, all evaporates before WIP6 the power boost (such as all the other parameters reported in Figs. from 131 to 133 and from 166 to 167) could be considered ranging between the curves WIP2 and WIP6. The last sentence could be accepted considering that these curves could be considered two limit cases: on one hand, an instantaneous thermal exchange between air and water with the water evaporation only depending by the air saturation capacity (WIP2 in the example) and on the other hand an overestimated delay (WIP6).
0.0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.00
1
2
3
4
5
6
7
8
water and air mass flow rate ratio [%]
upstream stage 2 stage 2 stage 3WIP2
upstream st.2
upstream stage 3
upstream stage 3 st. 3 WIP3
HOT upstream st.1 stage 1 stage 2
ISO upst. st.1WIP1
stage 3 surgelimit
stage 1 max power
upstream stage 4WIP4
upstream stage 4
upstream stage 5WIP5
upstream stage 5
upstream stage 6WIP6
upstream stage 6
upstream stage 7
upstream stage 7WIP7
upstream stage 8
upstream stage 8WIP8
HOT
ISO
HOT
ISO
HOT
ISO
HOT
ISO
HOT
ISO
HOT
ISO
HOT
ISO
surgelimit
surgelimit
surgelimit
surgelimit
surgelimit
surgelimit
surgelimit
max power
max power
max power
max power
max power
max power
max power
Figure 169 – Evaporation points as function of injection point, water mass flow rate and climate conditions (ISO and HOT cases)
140
3.4 – DROPLET TEMPERATURE INFLUENCE
From a thermodynamic point of view, the influence of the droplet temperature on gas turbine power boost is relatively simple to be understood. Other factor being equal (injection pressure and point, water flow rate etc.), to inject hotter water means to obtain a lower cooling effect, and, so, a lower power boost. In a different way it could be observed that the increase of injected water temperature modifies the power boost per unit of water consumption. With reference to the ISO case, in Fig. 170 it could be observed that, in order to obtain a power output increase of about 7 % of the ISO value with a water injection realized at the first stage, the water consumption increases about the 16% passing from the temperature of 15 °C to 90 °C, that means a reduction of the power boost per unit of water consumption of about 0.8 MJ per kg of injected water. In Fig. 171 the changes of heat rate for the two considered injected water temperatures are presented. To increase the injected water temperature means to increase the water consumption, to obtain the same gas turbine performances.
The previous conclusions are consistent with the evaporation model which was used for this analysis. It should be observed that, if the residence and the evaporation time are taken into account, the last conclusion could change. Other factors being equal, the evaporation time decreases if the droplet temperature increases; this evidence could effect positively the cooling efficiency because the decrease of the cooling due to the high injected water temperature could be counterbalanced by the evaporation time reduction.
1.00
1.01
1.02
1.03
1.04
1.05
1.06
1.07
1.08
0.0 0.1 0.2 0.3 0.4 0.5water and air mass flow rate ratio [%]
Pout
Pout,ISO T
d=15°C
Td=90°C
Figure 170 – Power output change as function of injected water to air mass flow rate ratio and temperature for an injection upstream from the first stage
141
0.980
0.984
0.988
0.992
0.996
1.000
0.0 0.1 0.2 0.3 0.4 0.5water and air mass flow rate ratio [%]
HR
HRISO
Td=15°C
Td=90°C
Figure 171 – Heat rate change as function of injected water to air mass flow rate ratio and temperature for an injection upstream from the first stage
142
4
Development Of A Calculation Code For The Prediction Of A
Gas Turbine Performance With Interstage Injection
The use of a commercial software described in the previous chapter allowed the achievement of important results for what concern the understanding of the behavior of an axial compressor with water interstage injection.
Anyway, the adopted software showed some limitations in particular due to its rigid structure and to the impossibility of taking into account the dynamic aspects of the water evaporation.
To overcome these limitations a self developed software in Fortran 90 was written. The realization of this program is based on the implementation of a new physical-mathematical model of wet compression on the respect of the ones available in the literature.
First of the description and application of this calculation procedure, the physical aspects of the evaporation of a droplet in an air stream and of the existing wet compression models are exanimate with more detail on the respect of chapter 1.
143
4.1 – EVAPORATION OF A WATER DROPLET IN AN AIR STREAM The evaporation of a water droplet in an air stream is a complex phenomenon that
involves energy and mass transfer between the liquid water and the air. To better understand the physical phenomenon, it could be considered a simple
example [85]. In Fig. 172 a film of liquid water in contact with air on one side and with a solid
surface on the other side is presented. Both air and the surface are assumed to have a temperature greater than the one of the liquid water.
Figure 172 – Evaporation of a film of water in air [85]
In Fig. 173 the temperature distribution along the normal direction to the interface is
presented. It could be observed from this figure that the minimum value of the temperature distribution is reached in correspondence of the water to air interface. This last evidence means that it should be present a thermal power flux through the two side of the control surface. By considering the air as the reference phase, ''
Sq and ''Lq could
be assumed as the thermal power fluxes through the right side (S) and the left side (R) respectively. According to the assumed reference, it results:
0q ''
S ≥
0q ''L ≤
144
Figure 173 – Temperature and mass concentration distribution in case of evaporation of a film of water in air [85]
From Fig. 173 it could be also observed that the water vapor concentration in air
decrease as the distance from film of water increase, reaching its maximum value ( SOHm ,2 ) in correspondence of the water and air interface and the minimum one ( GOHm ,2 ) in the undisturbed air.
In fact as the thermal power flux tends to reduce the temperature difference, in the same way, the diffusion of water vapour tends to reduce the difference of vapour concentration. This last evidence explains the presence of a flux of vapour toward the right side ( ''
2OHm in Fig. 173). It could be observed that OHm 2 , differently from the temperature, shows a discontinuity in correspondence of the interface between air and liquid water; for the film of water it could be assumed 1,2 =LOHm while the gas mixture could absorbs only a little amount of water vapour depending on the total pressure, and on the temperature (that means depending on the saturation pressure).
As the liquid water contains only a small amount of air, this could be neglected and the only flux of mass could be assumed to be the diffusion of vapour.
This phenomenon could be described by the following equations:
)hmhm(qq L,j''
L,jS,jj''S,j
''L
''S −=− ∑ (33)
where Sjh , and Ljh , are the specific enthalpy differences on the right and on the left
side of the control surface respectively. In this specific case, it results:
''''L,O2H
''S,O2H mmm ==
145
and then the Eq. 33 becomes:
)hh(mqq L,O2HS,O2H''''
L''S −=− (34)
In Eq. 34, it results 0hh L,O2HS,O2H ≥− ; the difference between this two terms is
equal to the latent heat of evaporation. The heat and mass transfer between two phase is influenced by many factors which
could be divided into two main categories: thermodynamic characteristics of the air:
o relative humidity; o total pressure; o temperature;
physical and thermodynamic characteristics of the water droplet:
o droplet volume; o characteristics of the liquid phase; o droplet temperature.
With references to Fig. 174, is assumed that:
the water droplet is considered as a perfect sphere; the physical properties of air and liquid water are homogeneous in each
considered phase; an infinitesimal volume of saturated air surrounding the water droplet is
assumed; the air is not soluble into the liquid water.
LIQUID WATER
SATURATED AIR
LAYER
HEAT
WATER VAPOUR
Figure 174 – Scheme of a water droplet exchanging heat and mass with surrounding air
On the basis of the previously hypothesis, the variation of the vapour concentration
from the droplet surface to the undisturbed air is assumed to be an instantaneous process instead of a continuous variation. In the same way, the temperature variation could be assumed as an “average value” between the temperatures of the droplet and the undisturbed air, neglecting that there exists a continuous variation between the two phases.
146
Under the previously assumptions, the heat amount ( Φ ) stored in a droplet between the time t and t + Δt [29] could be expressed as:
t
)TT(cm t)tt( dd
pdd Δ
−=Φ Δ+
where: md: mass of the water droplet Td: droplet surface temperature cpd: specific heat at constant pressure then, at the time t + Δt the droplet temperature is equal to
dddd Cpm
tTTt)tt(
Φ×Δ+=
Δ+
Assuming the thermal equilibrium between droplet and air, it could be written:
latconv Φ+Φ=Φ The previous relation, affirms that the thermal power exchanged by air and water
droplet is the sum of a term representing the sensible heat ( convΦ ) and a term representing the latent heat of evaporation ( latΦ ).
Indicating with Sd the droplet surface and with Ta the air temperature, the convective heat transfer could be written as:
)TT(Sh dadcvconv −××=Φ
the coefficient of convective thermal exchange cvh (W/m2 K) could be evaluated
from the expression of the number of Nusselt:
a
dcvDhNuλ
= (35)
where aλ is the air conductivity (W/m K). With the hypothesis of natural convection, the number of Nusselt is function of the
number of Prandtl and of thermal number of Grashof:
33.025.0t PrGr6.02Nu ××+= (36)
where, it results:
2a
3dda
2a
tD)TT(gGr
μ⋅−⋅β⋅⋅ρ
= (37)
147
a
da CpPrλ×μ
= (38)
being:
aμ : dynamic viscosity of the air (kg/m s) g: the gravity acceleration
ap,C
a
a T1
T1
=⎟⎠⎞
⎜⎝⎛
∂ρ∂
ρ=β : the thermal dilatation coefficient (expressed for a perfect gas)
by matching the Eqs. From 35 to 38 it could be obtained:
⎥⎥⎦
⎤
⎢⎢⎣
⎡⎟⎟⎠
⎞⎜⎜⎝
⎛λ⋅μ
⋅⎟⎟⎠
⎞⎜⎜⎝
⎛⋅μ
⋅−⋅⋅ρ⋅+⋅
λ=
33.0
a
da
25.0
a2a
3dda
2a
d
acv
CpT
D)TT(g6.02D
h
For what concerns the latent heat of evaporation, it could be expressed as:
tLm vd
lat Δ⋅Δ
=Φ
where vL (J/kg) is the latent heat of evaporation and it could be evaluated by using
the following relationship:
[J/kg] )T413.22498(1000L dv ⋅−= The mass variation of the droplet ( dmΔ ) in a time interval tΔ could be expressed as:
evapdd Stm φ⋅⋅Δ−=Δ The water vapour mass flux could be expressed as:
massmassevap CDf Δ⋅=φ where massDf is the mass diffusion coefficient (m/s) and massCΔ (kg/m3) is the
difference in water vapour concentration between the vapour layer and the undisturbed air:
ad massmassmass CCC −=Δ assuming the air surrounding the droplet as a perfect gas, the term
imassC could be written as:
i
imass TR
pMCi ⋅
⋅=
148
and then massCΔ results:
⎟⎟⎠
⎞⎜⎜⎝
⎛−=Δ
a
vap
d
vmass T
pTp
RMC d
where M is the molecular weight (kg/mole), R is the universal gas constant (8.32
J/mole K), vapp and dvp are the vapour pressure of undisturbed air and of vapour layer
respectively. The mass diffusion coefficient could be written as:
d
amass D
DfShDf ⋅=
being aDf (m2/s) the diffusion coefficient that could be expressed as:
⎟⎠⎞
⎜⎝⎛
⎟⎟⎠
⎞⎜⎜⎝
⎛⋅⋅= −
15.273T
P1013251026.2D a
a
5v (39)
Under the hypothesis of natural convection, the number of Sherwood is a function of
the mass number of Schmidt. According to its empirical expression, it could be written as:
33.025.0
m ScGr6.02Sh ⋅+= the number of Schmidt is:
va
a
DSc
⋅ρμ
=
while the mass number of Grashof is equal to:
2a
3dmol
*2a
mDCgGr
μ⋅Δ⋅β⋅⋅ρ
=
being
⎟⎟⎠
⎞⎜⎜⎝
⎛−=Δ
d
v
a
vapmol T
pT
pR1C d
aP,T
a
a
* MC
1
aaρ
−=⎟⎠⎞
⎜⎝⎛
∂ρ∂
⋅ρ
−=β
In conclusion, by matching the previous equation, the complete expression of the
evaporative heat flux could be written as:
149
⎟⎟⎠
⎞⎜⎜⎝
⎛−
⋅⋅⋅
=φa
vapv
d
vevap T
pTdp
DRDShM d (40)
The Eq. 40, obtained by the use of theoretical considerations, is very useful to
understand the most important parameters which influence the evaporation process of a water droplet in an air stream.
In fact, it could be observed that the main factors effecting the evaporation process could be summarized as follows:
diffusion coefficient; droplet diameter; droplet temperature; vapour pressure at air/droplet interface and in the undisturbed air.
150
4.2 – THE DIFFUSION COEFFICIENT With the term “diffusion” is intended the mass transfer towards a single phase in
absence of mixing. Both experimental and theoretical consideration have shown that the diffusion could happens in presence of pressure gradient, temperature gradient and concentration gradient. This last case is of interest for the present work.
The coefficient of diffusion could be assumed as the proportionality constant between the mass flux and the concentration gradient that causes the process itself.
To better understand the last sentence, it could be considered a mixture of two components (indicated with A and B) in a container at constant pressure and temperature [86] in which is BA nnn += the total number of molecules per unit of volume. Assuming a not uniform distribution of the molecules An and Bn into the container the diffusion coefficient of the component A into a mixture of the components A and B is given by:
dydn
Dn AABA −=& (41)
where An& is the molar flux of the component A, that means the number of molecules
through the unit of surface in the unity of time. In the same way, for the component B, it could be written:
dydn
Dn BBAB −=&
being constnnn BA =+= , it follows that:
dydn
dydn BA −=
that means vBAAB DDD == Multiplying both members of the Eq. 41 with the molecular weigh MA, it could be
obtained:
dyd
Dm AvA
ρ−=&
where Am& represents the mass flux of the component A. This last equation, called
Fick law, could be also expressed in term of partial pressure gradient:
dydp
TRD
m A
A
vA −=&
After having clarify the physical meaning of the diffusion coefficient, the next step is
its evaluation. On this regards it should be observed that it does not exist an univocal
151
expression for the diffusion coefficient because is formulation is of semi-empirical nature. By the way, in all the different available definitions it could be observed the same following characteristics for what concerns the diffusion coefficient:
it is inversely proportional to the total pressure; it is directly proportional to the temperature; it lightly depends on the chemical characteristics of the involved
components; its unit of measure is equal to m2/s.
In Tab. 21, the most common relationships of the diffusion coefficient, are presented.
It should be observed that all these definitions are always applicable at low value of pressure. For higher value of pressure the relationship is not well known with the exception of the self-diffusion.
Table 21 – Diffusion coefficient formulas
Formula References
1 ⎟⎠⎞
⎜⎝⎛
⎟⎟⎠
⎞⎜⎜⎝
⎛⋅⋅= −
15.273T
P1013251026.2D a
a
5v Chaker
[29]
2 D
2AB
21AB
23
v MPT00266.0D
Ω⋅σ⋅⋅⋅
= Reid [86]
3 ( )( )[ ]
D2AB
21AB
23321AB
v MPT10M98.003.3D
Ω⋅σ⋅⋅−
=−
Wilke & Lee [87]
4 ( ) ( )[ ]231Bv
31Av
21AB
75.1
vMP
T00143.0DΣ+Σ⋅⋅
⋅= Fuller [87]
5 ( ) ⎟⎟⎠
⎞⎜⎜⎝
⎛+
+⋅=
BA231
B31
A
23
v M1
M1
VVPT0069.0D Gilliland
[88]
6 ( ) ( )
PM
1M
1TTPP
TTTaD
21
BA
125cBcA
31cBcAb
cBcAv
⎟⎟⎠
⎞⎜⎜⎝
⎛+
⋅⎟⎟⎠
⎞⎜⎜⎝
⎛=
Bird [89]
The formulation of the diffusion coefficient given in the first row of Tab. 21 is very
simple and it could be assumed to be valid only in case of diffusion of water into an air stream. In fact the constant 2.26·10-5 derives from the molecular weights of air and liquid water. Ta and Pa are the temperature and the pressure of the air respectively compared with their reference values chosen equal to 101325 Pa and 273.15 K.
The formulation of the diffusion coefficient proposed by Reid et ali (second row in Tab. 21) could be obtained by theoretical consideration on the diffusion of gas mixture at low pressure [86]. The parameter MAB is function of the molecular weighs of the components A and B involved in the diffusion process, according to the relationship:
BA
AB
M1
M1
2M+
=
152
The characteristic length ( ABσ ) and the collision integral ( DΩ ) used in the formula could be evaluated by assuming an intermolecular force law.
The relationship proposed by Wilke & Lee [87] (third row in Tab. 21) is very similar to the one from Reid et ali with the exception of the adopted constant of proportionality.
Fuller [87] (fourth row in Tab. 21) proposes to evaluate the diffusion coefficient as function, among the others parameters, of the atomic diffusion volumes ( vΣ ). The values of vΣ for some chemical elements and simple molecules are reported in Tab. 22.
Table 22 – Atomic Diffusion Volumes
Element Atomic diffusion Volume C 15.9 H 2.31 O 6.11 N 4.54
Aromatic ring -18.3 Heterocyclic ring -18.3
He 2.67 Ne 5.98 Ar 16.2 Kr 24.5 Xe 32.7 H2 6.12 D2 6.84 N2 18.5 O2 16.3 Air 19.7 F 14.7 Cl 21.0 Br 21.9 I 29.8 S 22.9
CO 18.0 CO2 26.9 N2O 35.9 NH3 20.7 H2O 13.1 SF6 71.3 Cl2 38.4 Br2 69.0 SO2 41.8
Finally, the proposal by Gilliland [88] (fifth row in Tab. 21) is based on the
relationship between the diffusion coefficient and the atomic volumes and the molecular weigh, while Bird [89] (sixth row in Tab. 21) shows the diffusion coefficient as function of the critical temperatures and pressures.
153
4.3 – PHYSICAL–MATHEMATICAL MODEL OF WET COMPRESSION After having analyzed the physical process concerning the evaporation of a water
droplet in an air stream, is necessary to take into account the modellization of a wet compression process. As evident, this last issue is very complex because two different phenomenon take place, linked each other. In fact, due to the compression process, the temperature rises depending on the pressure ratio and polytropic efficiency; this last event increase the saturation capacity of the air allowing a greater amount of water to evaporate; the evaporation of water produces a cooling effect which influences the pressure ratio and the efficiency of the whole compression.
To estimate a compression process in presence of liquid water, some hypothesis are made; in particular the most common assumptions are the following:
the water droplets are perfect sphere; the physical properties of air and liquid water are homogeneous in each
considered phase; an infinitesimal volume of saturated air surrounds the water droplet; the air is not soluble into the liquid water; there are no interactions among the droplets; the dynamic of the spray is not taken into account.
Under the previous hypothesis, the most widely wet compression models, will be
analyzed.
4.3.1 – Wet Compression Model By Härtel & Pfeiffer [42] An ideal evaporation model is based on the hypothesis of thermal equilibrium
between water droplets and surrounding air during the compression process. This last sentence means that every phase chance is enough slow to maintain unchanged the saturation conditions. It follows that under the last assumptions, the compression velocity dtdp has no effect on the evaporation of the water.
Unfortunately, the thermal equilibrium is never verifies, and the “differences” between the real process and the ideal process depend on the droplet diameters and on the compression velocity.
Assuming an infinite conductivity for the water droplets and applying the mass and energy balance, the droplet temperature (Tl) and diameter (D) could be written as:
( )
⎟⎟⎠
⎞⎜⎜⎝
⎛
−
−
⋅ρ⋅ρ⋅
−=
⋅⋅
+−⋅⋅ρλ⋅⋅
=
∞
∞
1Y1Y
lnD
D4dtdD
dtdD
cDh3
TTcD
Nu6dt
dT
s,v
,v
l
vref
pl
lvl
pl2
l
refl
(42)
where ∞ and s indicate the air and the droplet surface respectively while the index
ref is referred to a vapour layer in saturation condition which surrounds the water droplet. It could be observed that the change in droplet diameter and temperature depend on air conductivity ( refλ ), density of air ( refρ ) and water ( lρ ), on the diffusion coefficient ( vD ) and on the mass fractions of vapour in the humid air ( svY , and ∞,vY ), which could be written as:
154
( )
( ) va,v,v
vas,vs,v
MM1pp11Y
MM1pp11Y
⋅−+=
⋅−+=
∞∞
The Nusselt number could be evaluated with the following relationship:
⎟⎟⎠
⎞⎜⎜⎝
⎛−⎟⎟
⎠
⎞⎜⎜⎝
⎛⋅λ⋅π
⋅⋅λ⋅π
⋅
=
1D2
cmexp
Dcm
Nu
ref
ref,pvd
ref
ref,pvd
&
&
Integrating the Eqs. 42, it could be obtained:
⎟⎟⎠
⎞⎜⎜⎝
⎛
−
−⋅⋅⋅ρ⋅π=⋅⋅ρ⋅
π=
⋅ρ⋅π
=ρ⋅⋅π=
∞
s,v
,vvref
2ld
3ll
3d
Y1Y1
lnDD2dTdDD3
6m
D6
R34m
&
the previous equations could be numerically solved only if trend of the pressure as
function of the time is assumed as shown in Fig. 175.
Figure 175 – Pressure distribution as function of compressor axial length
155
4.3.2 – Wet Compression Model By White & Meacock [43] Considering a number n of droplets, the mass of liquid water for dry air could be
written as:
l3 nr
34f ρ⋅⋅⋅π=
with the non slip hypothesis the number of the droplets remains constant. Indicating
with ω the mass of vapour per unit of dry air, it could be written:
DtDrnr4
DtDf
DtD 2
lπρ−=−=ω
According to the evaporation model proposed by Spalding, the relationship which
describes the convection and the diffusion for a spherical droplet results as:
⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+
ρ⋅ρ
=ρ
−=ll
v
l 11ln
rDJ
DtDR
where J is the vapour flux and ρ the humid air density, vD the diffusion coefficient
and lω the specific humidity on the droplet surface. Assuming that the droplets reached rapidly the wet bulb temperature, the previous
equation could be integrated to evaluate the complete evaporation time (τ ), if the dimensionless parameter τp& is fixed, where p& is defined as:
DtDp
p1p =&
4.3.3 – Wet Compression Model By Horlock [41] This wet compression model is based on the assumption of a linear trend of the
temperature increase (from *IT to *
IIT ) trough a compressor of a length equal to xΔ . It could be written:
xTT
dxdT *
I*II
*
Δ−
=
assuming a constant axial velocity Vx the equation becomes:
( )*I
*IIx
*
x TTVxdT
Vdxdt
−⋅Δ⋅
== (43)
156
being l3dl r
34M ρ⋅⋅π= the droplet mass, its variation during the time could be written
as:
dtdrr4
dtdM d
l2d
l ⋅ρ⋅⋅π=
multiplying dtdM l with the latent heat of evaporation ( fgh ), the heat exchanged
between the air at the temperature T* and the water at temperature Td could be obtained:
∞
−
−⋅⋅π⋅χ=⋅⎟
⎠⎞
⎜⎝⎛ ⋅ρ⋅⋅π
r1
r1
TTr4hdtdrr4
d
*d2
dafgd
l2d
where aχ is the air conductivity. Putting 0r1 =∞ , the last equation becomes:
fgld
d*
ad
hrTT
dtdr
⋅ρ⋅−
χ−= (44)
by matching the Eqs 43 and 44, results:
( )*I
*IIxfgl
*d
*
add TTVhxdT)TT(drr
−⋅⋅⋅ρΔ⋅−
χ−=
this last equation could be integrated to evaluate the distance required through the
compressor to have to complete evaporation of the droplet. In particular assuming:
Idpwbd TTTT ≅≅= finally it results:
( ) ( )fgxl
*I
*II
a2
1d hVxTTr
⋅⋅ρΔ⋅−
χ=
157
4.4 – IMPLEMENTATION OF A WET COMPRESSION CALCULATION PROCEDURE
The performance calculation of a compressor stage in case of liquid water at its inlet
could be developed with a “trial and error” procedure able to estimate the thermodynamic conditions at the exit of the stage and the evaporated liquid water mass flow rate.
The key parameter of this “routine” is the liquid water mass flow rate able to evaporate during the compression process. A possible model able to estimate this variable could be obtained from the one presented by Horlock and could be summarized as:
⎥⎥⎥
⎦
⎤
⎢⎢⎢
⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
⋅⋅ρΔ
⋅−
⋅χ−−πρ⋅⋅=23
FGXWpa
ioa
21D
31DWEV,W hV
Xc
hhRRn
34M (45)
where: MW,EV: mass of liquid water evaporated during the stage compression process n: number of water droplets ρW: liquid water density RD1: droplet radius at the inlet of the stage
aχ : air conductivity
io hh − : enthalpy difference through the compressor stage
pgc : specific heat at constant pressure of air ΔX: compressor stage length VX: air axial velocity hFG: latent heat of evaporation The use of the evaporation model in Eq. 45 allows to evaluate the mass of liquid
water that evaporates as function of the droplet diameter (2·RD1), of the residence time
through the compressor stage ⎟⎠⎞⎜
⎝⎛Δ
XVX and of temperature difference between the exit
and the inlet of the stage ( ) ⎟⎠⎞
⎜⎝⎛ −
pa
ioc
hh .
The physical-mathematical model of wet compression adopted in the present procedure assumes the “wet” compression process as sketched in Fig. 176.
158
OUTC
T
s
IN
OUT
INC
ΔT1
ΔT2
pIN
pOUT
Figure 176 – Wet stage compression calculation: cooling effect
In general, during a wet compression the cooling effect due to the water evaporation
is a continuous process; in fact more the temperature increases due to the compression more the water evaporates cooling the air. Being MW,EV the total mass of liquid water that could evaporate during the whole compression process, the corresponding total cooling ΔT = f(MW,EV), is divided into two fractions:
ΔT1 = f[XW·MW,EV]
ΔT2 = f[(1- XW)·MW,EV]
whose effect were considered concentrated at the inlet of the stage (ΔT1) and at the
exit (ΔT2) by using a parameter [ ]1,0X W ∈ as presented in Fig. 176. With the previous assumptions, the physic state of the air at the exit of the
compressor stage is represented by the point OUTC in Fig. 176, but the specific work adsorbed during the compression is assumed to be the enthalpy difference between the point OUT and the point INC. This last consideration could be better understood by considering the h-s diagram in Fig. 177. In this figure is presented a “dry” compression (from point IN to point NC) and a “wet” compression (from point IN to OUTC). From the figure it could be observed that is assumed that:
INNCINOUT hhhh
C−=−
This evidence means that, enthalpy difference be constant (passing from the
compression IN-NC to INC-OUT) , the effect of water evaporation is to increase the stage pressure ratio (pOUT >pNC in Fig. 177) and to decrease the exit temperature ( NCOUT TT
C< in Fig. 177).
159
OUTC
h
s
IN
OUT
INC
ΝC
pIN
pOUT pNC<pOUT
Figure 177 – Wet and dry stage compression calculation comparison
With reference to Figs. 178 and 179, the iterative procedure of calculation which is
realized by the present “routine” could be explained as follow: 1) evaluation of the number of droplets by using the equation:
W3
1D
INJ
R34
Mn
ρ⋅⋅π=
where MINJ and RD1 are the mass of injected liquid water and the injected droplet radius respectively;
2) evaluation of the air conductivity (χ a, W/m °C) as function of the pressure (bar)
by using the following relationship
P103035,560756,0 5a ⋅⋅+=χ −
3) calculation of the stage assuming no liquid water at the inlet (Fig. 178) by using
the following equations
( )[ ] [ ( ) ] 2***2**
*max,p*
max,p*p 1SF
11SF
1max,pmax,p
max,pmax,p
φ−−φ⋅+φ⋅−−φ⋅+φ
−ψ−ψ=ψ ψψ
ψψ
160
( )
( )
⎪⎪⎪⎪⎪
⎩
⎪⎪⎪⎪⎪
⎨
⎧
⎥⎥⎦
⎤
⎢⎢⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ∈
φ
ψ⎟⎟⎠
⎞⎜⎜⎝
⎛−
φ
ψ
⎥⎥⎦
⎤
⎢⎢⎣
⎡−⎟
⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ
η−−=η
⎥⎥⎦
⎤
⎢⎢⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ∈
φ
ψ⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ−
⎥⎥⎦
⎤
⎢⎢⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ−
η−−=η
φψ
φψ
max
*p
*p
2*p
2
max
*p
/
min
*p
*p
5.3*p
5.3
min
*p
/
*1
*1
*1
*
11*
1***
1
*1
11*
*
*
maxp
minp
, ,
,,
at this step the temperature difference through the stage is known;
4) by using the Eq. 45, the mass of liquid water, MW,EV (i), that should evaporated
during the compression process with reference to the points IN and NC in Fig. 178 is calculated;
5) MW,EV (i) is compared to the mass of water (MW,S(i) ) that could saturate the air
with reference to the outlet state of the stage; if results MW,EV (i) ≥ MW,S(i) then is fixed MW,EV (i) = MW,S(i)
6) the value of MW,EV (i) is compared to the liquid water effectively available at the
inlet of the stage and then, according to the values of the parameter xW and is calculated the value of ΔT1(i);
7) calculation of the air temperature and composition at the inlet of the stage after
the water evaporation upstream of the compression process (point INC(i) in Fig. 179);
8) the point OUT(i) (Fig. 179) is found as previously explained;
9) from the value of ΔT2(i) the temperature and the composition of the point
OUTC(i) are calculated;
10) with reference to the points OUTC(i) and IN is calculated the MW,EV (i+1) by using the Eq. 45;
11) if the difference
( ) ( )1iMiM EV,WEV,W +−
is less than a tolerance value the procedure stops.
161
T
s
IN
NC
pIN
PNC
Figure 178 – Dry stage compression calculation
OUTC (i)
T
s
IN
OUT (i)
INC (i)
ΔT1 (i)
ΔT2 (i)
pIN
p'OUT pOUT(i)
Figure 179 – Wet stage compression calculation: iterative procedure
The previous iterative procedure allows the calculation of the compressor stage in
case of “wet” compression. Ones the convergence is reached, on the basis of the evaporated water is calculated the new mass of liquid water (if yet present) and the new radius of the droplets as:
EV,WIN,WOUT,W MMM −=
3
W
EV,W3IN,dOUT,d
34M
RRπρ
−=
where: MW,OUT: mass of liquid water at the exit of the stage MW,IN: mass of liquid water at the inlet of the stage Rd,OUT: droplets radius at the exit of the stage Rd,IN: droplets radius at the inlet of the stage
162
The described routine was tested in a large series of case to evaluate its stability and convergence. In particular, cases characterized by droplet diameters equal to 1, 5, 10 and 20 μm with values of the parameter Wx from 0 to 1 (0 – 0.25 – 0.50 – 0.75 – 1.00) were analyzed. In the most part of this cases the routine showed to not reach the convergence. In particular was observed the impossibility of reaching the convergence in calculation with small droplet diameters (less than 10 μm ), or great temperature differences across the compressor stage and great amount of injected water. This convergence problems are mainly due to the adopted evaporation model and not to the used iterative procedure.
In fact, the model realized by Horlock is based only on a thermal balance without taking into account the mass transfer and the theory of diffusion between two phase. This fact makes this model able to evaluate the wet compression only if the entire compressor is considered and not in case of a stage by stage calculation. Concluding, the model proposed by Horlock could be used only to calculate the wet compression in case of overspray for a first approach evaluation, but is not able to consider the interstage injection in which stage by stage evaluations are necessary.
For all these reasons this model was substituted by the one developed by Härtel & Pfeiffer, adapting the iterative calculation procedure to this case.
The differential equation representing the evaporation models proposed by Härtel & Pfeiffer [42] and White & Meacock [43] are below:
Meacock& White
Pfeiffer & Hartel
⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅ρ⋅ρ
=
⎟⎟⎠
⎞⎜⎜⎝
⎛
−
−
⋅ρ⋅ρ
−= ∞
ll
v
s,v
,v
l
vref
11ln
RD
dtdR
1Y1Y
lnRD
dtdR
first of all, it could be observed that results:
⎟⎟⎠
⎞⎜⎜⎝
⎛
−−
−=⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+ ∞
1Y1Y
ln11ln
s,v
,v
l
In fact, svY , and ∞,vY (kg of vapour / kg of humid air) are defined as
( )
( ) va,v,v
vas,vs,v
MM1pp11Y
MM1pp11Y
⋅−+=
⋅−+=
∞∞
considering that the specific humidity (kg of vapour / kg of dry air) for a vapour
pressure equal to saturation and for a pressure less than saturation could be written respectively as:
163
∞
∞
−=ω
−=ω
,v
,v
a
v
s,v
s,v
a
vl
ppp
MM
ppp
MM
the previous terms could be written as
( )
l
l
l
v
a
s,v
s,v
vas,vs,v
1
11
1
MM
ppp
1
1MM1pp1
1Y
ω+ω
=
=
ω+
=
=
⋅⎟⎟⎠
⎞⎜⎜⎝
⎛ −+
=
=⋅−+
=
in the same way, it results:
ω+ω
=∞ 1Y ,v
for the previous consideration it could be written:
ω+ω+
=
ω+ω−−ω
ω+ω−−ω
=−
ω+ω
−ω+
ω
=−−∞
11
11
11
11
11
1Y1Y l
l
ll
l
ls,v
,v
concluding, it results:
⎟⎟⎠
⎞⎜⎜⎝
⎛
−−
−=⎟⎠⎞
⎜⎝⎛
ω+ω+
−=⎟⎠⎞
⎜⎝⎛
ω+ω+
=⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+ ∞
−
1Y1Y
ln11ln
11ln
11ln
s,v
,vl1
l
l
by considering the last equivalence, the two evaporation model could be written as:
Meacock& White
Pfeiffer & Hartel
⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅⋅⎟⎟⎠
⎞⎜⎜⎝
⎛ρ⋅ρ
=
⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅⋅⎟⎟⎠
⎞⎜⎜⎝
⎛ρ⋅ρ
=
ll
v
ll
vref
11ln
R1D
dtdR
11ln
R1D
dtdR
As evident, the previous evaporation models are similar for the exception of the
phase chosen as reference. On this regard, appears better to adopt the model proposed
164
by Hartel & Pfeiffer which assumes a vapour layer surrounding the water droplet as reference phase.
By using the model proposed by Hartel & Pfeiffer instead of the one used by Horlock, it could be observed that the equation:
( )32D
31DWEV,W RRn
34M −⋅ρ⋅⋅π=
which estimates the mass of water evaporated across a stage remains true. The
relationship which calculates the droplet radius at the stage exit, instead, needs to be changed to take into account the new model. In fact the proposed differential equation must be integrated on the respect of the radius (from RD1 to RD2) and of the time (from 0 to xVxt Δ=Δ ). It could be written:
[ ]
xlv
l
ref21D
22D
lv
l
ref2
1D2
2D
t0
lv
l
refR
R
2
t
0 lv
l
refR
R
lv
l
ref
Vx
11lnD2RR
t11lnD
2RR
t11lnD
2R
dt11lnDdRR
dt11lnDdRR
2D
1D
2D
1D
Δ⋅⎟⎟
⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅⋅ρ
ρ⋅+=
Δ⋅⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅⋅ρ
ρ=
−
⋅⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅⋅ρ
ρ=⎥
⎦
⎤⎢⎣
⎡
⋅⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅⋅ρ
ρ=⋅
⋅⎟⎟⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅⋅ρ
ρ=⋅
Δ
Δ
∫∫
the last issue that needs to be clarify, regards the reference point on the respect of
which refρ , lω , and Dv are evaluated. For each stage the medium values of pressure (Pmed) and temperature ( medT ) are
estimated, as follows: medT : is the medium value between the inlet and the outlet state of the
compressor stage; with reference to Fig. 180, it is evaluated as:
⎟⎠
⎞⎜⎝
⎛ +=
2TT
T OUTINmed
Pmed: is evaluated as the isobar which passes for the point resulting by the
intersection between the isothermal Tmed and the politropic compression from IN to OUT.
165
T
s
IN
OUT
pIN
pOUT
TIN
TOUT
TMED
pMED
Figure 180 – Evaluation of the medium pressure and temperature through the compressor stage
Moreover, it could be taken into account the change in droplet surface temperature
( wlT ) during the compression process by using the following relationship:
( )wmedTLwwl TTXTT −⋅+= where:
wT : initial droplet surface temperature (at the injection point); TLX : tuning parameter ( 1X0 TL ≤≤ )
The use of the previous equation allows to take into account the variation of the
injected water droplets even if a convective heat transfer model is not included. By using the parameter TLX , in fact, the droplet surface temperature could be estimated between the value at the moment of water injection ( wT ) and the medium value of the air temperature into the compressor stage.
It could be observed that at a certain time, the droplet surface temperature is function of the droplet diameter. In fact, considering the relationship that express the thermal power exchanged between air and droplet through the droplet surface
t
)TT(CpV
t
)TT(Cpm t)tt(t)tt( dd
ddldd
dd Δ
−⋅⋅ρ=
Δ
−⋅=Φ Δ+Δ+
from which:
ddldd CpV
tTTt)tt( ⋅⋅ρ
Φ⋅Δ+=
Δ+ (46)
being:
166
( )
( )( ) vmassvddgdcv
vevapddgdcv
vddgdcv
latconv
LCDSTTSht
LStTTSh
tLm
TTSh
⋅Δ⋅⋅−−⋅⋅=
=Δ
⋅φ⋅⋅Δ−−⋅⋅=
=Δ
⋅Δ+−⋅⋅=
=Φ+Φ=Φ
(47)
by matching Eqs. 46 and 47 it could be obtained:
( )[ ]massvdgcvd
d
dldd CDTTh
VS
CptTT
tt)tt(Δ⋅−−⋅⋅
⋅ρΔ
+=Δ+
From this last equation, it could be observed that, others factor being equal, the
droplet surface temperature, after a time interval equal to tΔ , depends on the surface to volume ratio of the droplet; this means that the droplet surface temperature is inversely proportional to the droplet radius.
The choice of the coefficient TLX should be done by considering this evidence; than
TLX is inversely proportional to the droplet diameter. The temperature ( refT ) of the vapour layer which surround the water droplet could be
evaluated as:
( )wlmedTIwlref TTXTT −⋅+= where TIX is another toning parameter ( 1X0 TI ≤≤ ). Finally the values of the density for the saturated air ( refρ ), diffusion coefficient (Dv)
and maximum specific humidity ( lω ) are evaluated with reference to the total pressure Pmed and to the temperature Tref.
In particular, it results:
ref
medref TR
P⋅
=ρ
in which the gas constant R is evaluated with reference to the saturation pressure at
the temperature Tref, by using the following equation:
( )( )refsvmed
refsv
a
vl TpP
TpMM
,
,
−=ω
For what concerns the diffusion coefficient, among the available expression was
chosen the one proposed by Fuller [87]:
( ) ( )[ ]2313121
75.100143.0
BvAvAB
ABMP
TDΣ+Σ⋅⋅
⋅=
in which
167
216.22
016.181
970.281
2
M1
M1
2M
BA
AB =+
=+
= g/mole
( ) 7,19=Σ Av ( ) 1,13=Σ Bv The iterative calculation procedure (Fig. 181) to evaluate the performance of a
compressor stage with water injection results as follows: 1) evaluation of the number of droplets by using the equation:
W3
1D
INJ
R34
Mn
ρ⋅⋅π=
where MINJ and RD1 are the mass of injected liquid water and the injected droplet radius respectively;
2) calculation of the stage assuming no liquid water at the inlet (Fig.
178) by using the following equations
( )[ ] [ ( ) ] 2***2**
*max,p*
max,p*p 1SF
11SF
1max,pmax,p
max,pmax,p
φ−−φ⋅+φ⋅−−φ⋅+φ
−ψ−ψ=ψ ψψ
ψψ
( )
( )
⎪⎪⎪⎪⎪
⎩
⎪⎪⎪⎪⎪
⎨
⎧
⎥⎥⎦
⎤
⎢⎢⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ∈
φ
ψ⎟⎟⎠
⎞⎜⎜⎝
⎛−
φ
ψ
⎥⎥⎦
⎤
⎢⎢⎣
⎡−⎟
⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ
η−−=η
⎥⎥⎦
⎤
⎢⎢⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ∈
φ
ψ⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ−
⎥⎥⎦
⎤
⎢⎢⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛
φ
ψ−
η−−=η
φψ
φψ
max
*p
*p
2*p
2
max
*p
/
min
*p
*p
5.3*p
5.3
min
*p
/
*1
*1
*1
*
11*
1***
1
*1
11*
*
*
maxp
minp
, ,
,,
at this step the temperature difference trough the stage is known;
3) evaluation of Twl, TMED, PMED, Tref and Ps previous discussed; 4) if MEDS PP ≤ the procedure continue from the step 5; if MEDS PP ≥ the evaporable
water is fixed equal to all the available water and the procedure jump to the step 7;
5) calculation of lω , refρ , Dv;
168
6) evaluation of the water which could evaporate through the stage as:
⎥⎥⎥
⎦
⎤
⎢⎢⎢
⎣
⎡
⎟⎟⎠
⎞⎜⎜⎝
⎛ Δ⋅⎟⎟
⎠
⎞⎜⎜⎝
⎛ω+ω+
⋅⋅ρ
ρ⋅+−⋅ρ⋅⋅π=
23
xlv
l
ref21D
31DWEV,W V
x11lnD2RRn
34)i(M
7) comparison between the water which could evaporate with the available amount at the inlet of the stage and evaluation of the two fraction of water which evaporate upstream and downstream of the compression process by using the parameter [ ]1,0X W ∈ ;
8) calculation of the air temperature and composition at the inlet of the stage after
the water evaporation upstream of the compression process (point INC(i) in Fig. 179);
9) evaluation of the flux coefficient at the stage inlet after the evaporation of the
first fraction of water as: ( )desa
a
des
*
UVUV
=φφ
=φ
10) check to estimate if *φ is between the limits of the stage operation filed; if this
condition is realized the iteration goes on, on the contrary the calculation is stopped as will be better explained in a next paragraph;
11) the point OUT(i) (Fig. 179) is found as previously explained;
12) from the value of ΔT2(i) the temperature and the composition of the point
OUTC(i) are calculated; 13) evaluation of the new values of Twl, TMED, PMED, Tref and Ps;
14) calculation of the new values of lω , refρ , Dv;
15) evaluation of a new value of MW,EV (i+1)
16) if the difference
( ) ( )1iMiM EV,WEV,W +−
is less than a tolerance value the procedure stops.
169
Evaluation of number of droplets and of liquid water at stage inlet
Evaluation of pressure and temperature at the stage exit in the without liquid water
Evaluation of evaporable water (MWE1) according to Hartel e Pfeiffer model
Calculation of the thermodynamic conditions after theevaporation of the first fraction of water (point INC)
Evaluation of the compression process (point OUT)
Calculation of the thermodynamic conditions after the evaporation of the first fraction of water (point OUTC)
Evaluation of Dfrefl ,, ρω
Evaluation of evaporable water (MWE2) according to Hartel e Pfeiffer model
END PROCEDURE
Evaluation of ( )refSrefMEDMEDwl TPTPTT ,,,,
injection? Y
N
N
Y
ε<−RMWERMWE
211
Evaluation of the effective evaporable water (MWE1R) and of the fraction of water which should evaporate upstream of the compressor and downstream of it
Evaluation of the effective evaporable water (MWE2R) and ofthe fraction of water which should evaporate upstream of the
compressor and downstream of it
MWE2R
MWE1R
Evaluation of *φ in INC
*max
**min φφφ ≤≤
Y
EXIT N
medS PP ≥ Evaluation of frefl D,, ρω
Evaporable water (MWE1) is fixed equal to the available water
Evaluation of ( )refSrefMEDMED TPTPT ,,,
medS PP ≥
Evaporable water (MWE2) is fixed equal to the available water
N Y
N Y
Figure 181 – Wet compression calculation routine flow chart
170
4.5 – IN.FO. G.T. E. CALCULATION CODE The wet compression iterative procedure, described in the previous paragraph, was
written in Fortran 90 language to be included in an existing calculation code for the gas turbine performance evaluation.
This new program, named IN.FO. G.T. E (INterstage FOgging Gas Turbine Evaluation) developed by DIEM – University of Bologna, is able to evaluate the performance of a gas turbine in presence of fogging, overspray and in particular of interstage water injection.
The main structure of the calculation code is presented in Fig. 182. It could be observed that is divided into three calculation modules (COMP, CC and TURB) which contain, respectively, the physical-mathematical models used to evaluate their performance; these modules consist of a widely number of internal calculation routines, proposed in the following.
The management of the COMP, CC and TURB modules, of the code internal logic and of the all iterative processes of calculation, used to reach the convergence of the gas turbine simulation is realized by the “main program”. In the flow chart of the main program (Fig. 182) are also highlighted the flows of data (both input and output) between the various calculation modules.
In fact, the main program (MAIN) supplies the necessary inlet data (INPUTS and GAS TURBINE SETUP arrays in Fig. 182) to the three modules and reads the outputs of their internal calculation process. It should be observed that the data in the INPUTS array must be defined by the program User (by compiling two txt files) while the GAS TURBINE SETUP array contains all the parameters that characterize the machine.
In particular the data in the INPUTS array are:
ambient conditions (site temperature, pressure and relative humidity); injection point(s) location(s) and/or presence of high pressure fogging
system; temperature (TWATER), pressure (PWATER), mass flow rate(s) (MWATER) of
injected water and droplet radius (RD); gas turbine control logic (constant turbine inlet temperature (TIT) or
constant turbine outlet temperature (TOT) ); fuel composition; fuel injection pressure(PFUEL) and temperature (TFUEL); injected water and/or steam mass flow rate (MWSCC), pressure (PWSCC) and
temperature (TWSCC) in the combustion chamber;
171
INPUTS
MAIR (i)
COMP
CC
TURB
MAIR,CALC(i)
N
OUTPUT
Y
MAIR (i)
MA
IR(i+
1)=M
AIR
(i)+(
MA
IR,C
ALC
(i)-M
AIR
(i) )/
1000
GAS TURBINE SETUP
i=i+
1
TOLL)i(M
)i(M1
AIR
CALC,AIR <−
Figure 182 – Main program (MAIN) flow chart
172
the data in the GAS TURBINE SETUP array are: gas turbine performance and thermodynamic parameters @ ISO conditions; number of compressor stages; geometrical data of the compressor; compressor stage performance maps; compressor air bleed ports; turbine performance maps; inlet fogging (if presents) efficiency; pressure losses at compressor inlet, through the combustion chamber and at the
turbine outlet; mathematical models tuning parameters; compressor and turbine mechanical efficiency; generator efficiency.
The parameters calculated by the different modules of the code are collected by the
main program and are written in the OUTPUT array. In particular the data in the OUTPUT array are:
gas turbine power output and heat rate; pressure, temperature and air (and liquid water if presents) mass flow rate at the
exit of each compressor stage; compressor pressure ratio; pressure, temperature and mass flow rate of gasses at the exit of combustion
chamber; pressure, temperature and mass flow rate of gasses at the exit of turbine;
turbine expansion ratio; The performance evaluation of GE Frame 7 EA gas turbine is realized by using an
iterative calculation procedure based on the value of the air mass flow rate at compressor inlet as presented in Fig. 182. The management of this procedure is developed by the main program.
The steps of this procedure are: 1) the main program define a first value of the air mass flow rate MAIR(i); 2) the main program supplies to the COMP module the necessary inlet data (from
INPUTS and GAS TURBINE SETUP arrays) and the value of MAIR(i); 3) stage by stage, the module COMP calculates the performance of the compressor
(pressure and temperature at the stage exit, evaporated water mass flow rate, enthalpy difference between the inlet and the exit of the compressor stage);
4) according to both defined (INPUTS and GAS TURBINE SETUP arrays) and calculated (COMP OUTPUT array) data, the CC module evaluates the fuel consumption and pressure, temperature and the gasses mass flow rate at the turbine inlet;
5) on the basis of the outputs of the previous modules and of the defined inlet data, the gasses mass flow rate through the turbine are evaluated by using the turbine performance maps MGAS, MAP(i);
6) the main program calculates: ( ) ( ) LWSCCFUELWATERMAP,GASCALC,AIR MMMMiMiM +−−−=
where:
173
MWATER: injected water mass flow rate MFUEL: fuel mass flow rate MWSCC: steam and or water mass flow rate injected in combustion
chamber ML: leakage air mass flow rate
7) the main program evaluates:
( )AIR
CALC,AIR
MiM
1−
8) if this differences is less than a tolerance value (equal to 10-4), the iteration
stops; if not a new value of the air mass flow rate at compressor inlet is calculated as:
( ) ( ) ( ) ( )1000
iMiMiM1iM AIRCALC,AIR
AIRAIR
−+=+
in this last case, the calculation procedure starts again from the point 2) with the new value MAIR(i+1) of the air mass flow rate at the compressor inlet.
In the next paragraphs, the calculation logic and the mathematical models of the
COMP, CC and TURB modules are shown.
4.5.1 – COMP Routine For Compressor Performance Evaluation Before of starting the iterative procedure previously described, if the ambient
conditions are different from the ones at ISO conditions (15°C, 1.013 bar, 60% of RH) the design air mass flow rate (MAIR,DES, ridden in the GAS TURBINE SETUP array) is corrected to obtain an air mass flow rate of first trial (MAIR(i=1) as:
DES,AIRDES,AIR
AIRAIR M)1i(M
ρρ
==
With this value of air mass flow rate, the first run of the calculation procedure is
realized. First of all, the MAIN program, sets to zero all the terms of 12 arrays:
PRESS (pressures at the inlet and at the exit of each compressor stage); TEMP (temperatures at the inlet and at the exit of each compressor stage); WAR (specific humidity values at the inlet and at the exit of each compressor
stage); FAR (fuel to humid air ratios at the inlet and at the exit of each compressor stage); WATER (water mass flow rates at the inlet and at the exit of each compressor
stage); V_WM (air mass flow rates at the inlet and at the exit of each compressor stage); PHI (flux coefficient of each compressor stage, UVa=φ ); NG (number of injected droplets (if any) at each compressor stage); VX (air axial velocity of each compressor stage); P_ST (shaft power absorbed by each compressor stage); MWER (water mass flow rate evaporated in each compressor stage);
174
CONTR (control parameters to check the surge or the chocking of each compressor stage).
As evident by the definition of each of the previous arrays, it could be observed that
the first six arrays consist of a number of elements equal to two times the number of the compressor stages while the remaining six arrays consist ;of a number of elements equal to the number of the compressor stages.
At this point, the presence of water injection in the compressor inlet duct is checked. If inlet fogging is realized, a subroutine named INLET_FOG is activated which evaluates:
the number of water droplets from the injected water mass flow rate into the compressor inlet duct and the droplets radius and water density;
the wet bulb temperature and the corresponding evaporable mass of water (referred to the unity of dry air) by another subroutine named HUMIDITY;
the air temperature and the evaporated water at the first compressor stage inlet taking into account the efficiency of the inlet fogging system (this parameter if defined in the GAS TURBINE SETUP array);
the specific humidity, the air and the liquid water mass flow rates, the droplets radius downstream of the inlet fogging section (and then at the first stage inlet).
Back to COMP routine, a DO cycle is initialized. The number of these cycles is equal
to the number of the compressor stages. For each iteration of the MAIN program, the DO cycle should be realized.
First of all, the left (φ *lim,sx) and the right (φ *
lim,dx) limit of the stage performance map are evaluated. This limits correspond to the surge and to the chocking margin, respectively and are evaluated by using the following relationship:
( ) ( )[ ]
( )[ ]2**
2****max,p*
max,p*p
11SF
1SF1
max,pmax,p
max,pmax,p
−−φ⋅+φ
φ−−φ⋅+φ⋅−ψ−ψ=ψ
ψψ
ψψ
in which * indicates a parameter normalized to its design value. In particular as
φ *lim,sx is considered the value of φ * corresponding to the maximum value of the
pressure coefficient as:
( )1SF ***sxlim, max,pmax,p
−φ⋅+φ=φ ψψ while as φ *
lim,dx is considered the value of φ * corresponding to a pressure coefficient equal to zero:
CBB 2*
dxlim, ++=φ where:
( )1SFB **max,pmax,p
−φ⋅+φ= ψψ
2*
BA
C max,p −ψ
=
175
( )[ ]2**
*max,p
11SF
1A
max,pmax,p−−φ⋅+φ
−ψ=
ψψ
The value of φ * at the inlet of each stage is compared to the corresponding surge and
chocking limit [φ *lim,sx , φ *
lim,dx]. If results:
*sxlim,
* φ<φ or
*dxlim,
* φ>φ the corresponding element of the CONTR array is set equal to –1 or to 1
respectively. In case of surge or chocking of the stage, the trial value of the mass flow rate is corrected as follows:
( )
2WMWM
WM ok)old()new(
+=
where WM(old) is the current value of the air mass flow rate which causes the
reaching of the surge or of the chocking and WM(ok) is the last value of the air mass flow rate able to be into the limits of the performance maps of each stage. Obviously, this correction could be done only if the DO cycle of the COMP routine has run completely one time. This last sentence means that at least one iteration of the main program should be realized. On the contrary, if the surge or the chocking conditions are reached with the value of air mass flow rate evaluated at the first iteration of the MAIN program, it results that MAIR(i=1) coincides with WM(old) while WM(ok) is equal to zero. Under this conditions the correction of the air mass flow rate could not be done and then the value of MAIR(i=1) is increased or decreased of 0.1% according to the fact that the surge or the chocking conditions has been reached.
This procedure goes on until a value of air mass flow rate able to satisfy each compressor stage is found. It should be remembered that what described happens for each iteration of the MAIN program.
For each stage, after the check on the surge and chocking margin, the presence of liquid water at the stage inlet is observed. This liquid water could be injected into the current stage or it could be come from an injection in an upstream stage. The subroutines DRY_STAGE_CALC or WET_STAGE_CALC are activated according to the absence or to the presence of liquid water at the stage inlet, respectively.
DRY_STAGE_CALC receives the following inputs: pressure, temperature, air mass flow rate at the stage inlet;
design flux coefficient and all the geometrical parameters to completely define the stage, including its performance map.
First the value of φ * is calculated. With this value, the dimensionless pressure
coefficient ( Ψ *p) is estimated and then the pressure coefficient ( Ψ p) since its design
value. By using the relationship ( )**p
* F φψ=η η the values of η* and η are estimated.
176
By using these last two values, the isentropic (Δhis) and the real (Δhr) enthalpy difference through the stage could be evaluated as:
2
pis Uh ⋅ψ=Δ
2r Uh ⋅ψ=Δ
with the previous values, the temperature and the pressure at the exit of the stage and
then the stage pressure ratio could be evaluated. Moreover if the last compressor stage is evaluated, the compressor pressure ratio is calculated.
WET_STAGE_CALC routine, realizes the evaluation of the performance of a compressor stage in case of liquid water at its inlet, according to the calculation procedure explained in the previous paragraph.
The evaluation of the mass of liquid water which could evaporate (MW,EV(i) ) is developed according to the evaporation model already explained. This value is compared to the liquid water available at the inlet of the stage (MW,AV) to evaluate the real amount of water (MW,EVR) which could evaporate through the stage. More in details, if MW,EV(i) is less than MW,AV then results MW,EVR(i) = MW,EV(i); the mass of liquid water which evaporates upstream of the stage calculation is XW·MW,EVR(i) and downstream is (1 – XW)·MW,EVR(i).
On the contrary if results MW,EV(i) > MW,AV it results MW,EVR(i) = MW,AV; in this last case if also XW·MW,EVR(i) > MW,AV, then MW,AV evaporates upstream of the stage calculation and no water evaporation occurs downstream of it. Differently (XW·MW,EVR(i) < MW,AV), the water mass flow rate equal to XW·MW,EVR(i) evaporates upstream of the stage calculation and the remaining mass flow rate MW,AV – XW·MW,EVR(i) evaporates downstream of it.
After the evaluation of the mass of water which evaporates upstream of the stage calculation, the temperature and the specific humidity at the stage inlet need to be evaluated. With reference to Fig. 179, indicating with:
MAIR,DRY,IN = dry air mass flow rate at point IN; hAIR,DRY,IN = specific enthalpy of dry air mass at point IN; MAIR,VAP,IN = vapour mass flow rate in air at point IN; hAIR,VAP,IN = specific enthalpy of the vapour at point IN; XW·MW,EVR = evaporated water between point IN and INC that is to be assumed
liquid at point IN and vapour at point INC; hXW,IN = specific enthalpy of liquid water evaporated at point IN; hXW,INC = specific enthalpy of vapour at point INC; hAIR,DRY,INC = specific enthalpy of dry air at point INC; hAIR,VAP,INC = specific enthalpy of vapour at point INC; MAIR,IN = MAIR,DRY,IN + MAIR,VAP,IN = humid air mass flow rate at point IN; It will results: MAIR,INC = MAIR,IN + XW·MW,EVR = humid air mass flow rate at point INC; MAIR,DRY,INC = MAIR,DRY,IN = dry air mass flow rate at point INC; MAIR,VAP,INC = MAIR,VAP,IN + XW·MW,EVR = vapour mass flow rate in air at point INC;
177
Indicating with the subscript r the value assumed as reference for the evaluation of the enthalpy, which corresponds to the triple point of the water ( bar 00611,0P andK 16,273T rr == ), it could be written:
( ) ( )
( )( ) ( )( )rXW,INCXW,EVRW,W
rVAP,AIR,INCVAP,AIR,INCVAP,AIR,rDRY,AIR,INCDRY,AIR,INCDRY,AIR,
rXW,INXW,EVRW,W
rVAP,AIR,INVAP,AIR,INVAP,AIR,rDRY,AIR,INDRY,AIR,INDRY,AIR,
hhMXhhMhhM
hhMXhhMhhM
−⋅+
−⋅+−⋅=
=−⋅+
−⋅+−⋅
which could be also written as:
INCAIR,INC,AIR,INXW,EVRW,WINAIR,IN,AIR, hMhMXhM ⋅=⋅+⋅ from which hAIR,INC and then the temperature of the air in the point INC could be
evaluated. In the same way, the enthalpy of the point OUTC is calculated. After the evaluation of all the stage of the compressor (with DRY_STAGE_CALC or
WET_STAGE_CALC) the COMP routine ends and the calculation continues with the activation, by the MAIN program of the CC routine to evaluate the combustion chamber section.
Moreover the COMP routine evaluate the air extraction from the compressor by considering the flow function (FF) of the extracted air equal to its design value (FFDES). This means that the air mass flow rate (MEX) extracted from the stage is evaluated as:
1k1k
0
DESEX
1k2
kMWR
pT
FFM
−+
⎟⎠⎞
⎜⎝⎛
+
⋅
= (48)
where: R0 = Universal gas constant MW = Molecular weight k = cP/cV ratio between specific heat at constant pressure and volume The stages at which the air bleed ports are placed and the corresponding reference
values of the FFDES are specified in the GAS TURBINE SETUP array. The value of MEX is divided into two terms by using a parameter [ ]1,0x EX ∈ to take
into account the fraction of extracted air used for the turbine cooling (MCOOL) and the remain mass flow (ML) which is assumed to be lost (bearing cooling or other). It results:
EXEXCOOL MxM ⋅=
( ) EXEXL Mx1M ⋅−= The flow chart of the COMP module previously described is sketched in Fig. 183.
178
INPUTS
Inlet Fogging
Inlet foggingcalculation
Y N
STAGE (i) i = 1 to Ns
Interstage injectionand/or
liquid water @ stage inlet
Y
WET STAGE CALCULATION
DRY STAGE CALCULATION
Airextraction
Y N
Extractioncalculation
i = Ns
N
COMPRESSOROUTPUT
Y
STAGE (i) OUTPUT
N
GAS TURBINE SETUP
MAIR (i)
Figure 183 – Compressor module (COMP) flow chart
179
4.5.2 – CC Routine For Combustion Chamber Evaluation The combustion chamber calculation procedure is presented in the flow chart in Fig.
184. The module receives inlet data from the INPUTS, GAS TURBINE SETUP and from the COMPRESSOR OUTPUT arrays.
INPUTS
TIT (input)
Y
Liquid water@ compressor exit
COMBUSTOR OUTPUT
GAS TURBINE SETUP
M'AIR (i)
Nevaporationcalculation
Y
Nmix
calculation
Steam and/or water injection in CC
MFUEL calculation
COMPRESSOR OUTPUT
Figure 184 – Combustion chamber module (CC) flow chart
As performed in the figure, the module is able to calculate the fuel consumption
required for obtaining the desiderate value of turbine inlet temperature as function of the air mass flow rate at combustor inlet ( ) ( )( )∑−=′ EXAIRAIR MiMiM , of the water and/or the steam injected in the combustor and of the liquid water eventually present at the exit of the compressor by considering an energy and mass balance.
180
4.5.3 – TURB Routine For Turbine Evaluation The turbine performance calculation is developed following the flow chart in Fig.
185. From the figure it could be observed that two main calculation modules are highlighted: one dedicated to the calculation of the turbine coolant effects and the other to evaluate the turbine expansion.
INPUTS
Expansion ratio calculation
MGAS,MAP (i) TURBINE OUTPUT
GAS TURBINE SETUP
Isentropic efficiency calculation
COMBUSTOR OUTPUT
Flow Function calculation
Turbine coolant effects calculation
Turbine coolant effects calculation
COMPRESSOR OUTPUT
Figure 185 – Turbine module (TURB) flow chart
181
The effect of turbine cooling on the performance of the expansion are taken into account by considering the scheme in Fig. 186. The total turbine cooling air mass flow rate is divided into two fractions by a parameter [ ]1,0x C ∈ defined in GAS TURBINE SETUP array. The obtained fractions xC·MCOOL and (1-xC)·MCOOL are mixed respectively upstream and downstream of the turbine expansion [6, 7].
MCOOL
xC·MCOOL
(1-xC)·MCOOL
IN INC
OUT OUTC
Figure 186 – Turbine cooling scheme The evaluation of turbine performance is realized by using the performance map
presented in Fig 187 that shows the trend of the turbine expansion ratio (β) as function of the flow function (FF) and of the corrected speed (CS) normalized to the corresponding design values.
In particular the calculation procedure is developed into the following steps: 1. on the basis of the pressure ratio evaluated by the COMP module, the expansion
ratio could be evaluated; 2. from the turbine performance maps (Fig 187) the value of FF* and the
corresponding isentropic efficiency are evaluated;
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.85 0.90 0.95 1.00 1.05 1.10
β*
FF*
CS*=1.1
1.0 0.9 0.8 0.7 0.6 0.5 0.4
Figure 187 – Turbine performance maps
182
4.6 – WET COMPRESSION MODEL TUNING The tuning of the wet compression model used in the IN.FO. G.T. PE.P. program
was realized by comparison with the results obtained by Utamura et ali [13]. These results are summarized in Fig. 188 which shows the percentage of evaporated
water (on the respect of the injected water mass flow rate) as function of the compressor length for three value of the droplet diameters equal to 10, 20 and 30 micron, respectively.
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compressor stages
d = 10 μm
20 μm
30 μm
Figure 188 – Evaporated water to injected water ratio through the compressor for a Frame 9 E gas turbine [13]
The reported results are referred to a water injection which occurs upstream of the
first stage of a GE Frame 9 E gas turbine model operating at ISO conditions. The main performance of this machine are reported in Tab. 23, with reference to ISO ambient conditions.
Table 23 – GE Frame 9E main performance @ ISO conditions gross power output [MW] 115 RPM 3000 air at compressor inlet [kg/s] 411 pressure ratio 12.4 compressor isentropic efficiency [%] 89.9 number of stages 17 turbine inlet temperature [°C] 1155 turbine outlet temperature [°C] 560 fuel type Kerosene
183
The tuning of the wet compression calculation routine corresponds to the evaluation of the four tuning parameters (XIF, XTI, XTL and XW), already introduced in a previous paragraph. More in details:
XIF: evaluates the amount of water (Wevap) which could evaporate into the
compressor inlet duct and the corresponding temperature (T1) at the inlet of the first stage. The influence of this parameter could be simplify as follows:
o if XIF = 0 then Wevap = 0 and T1 = Tamb (no water evaporation occurs in the compressor inlet duct);
o if XIF = 1 then Wevap is the maximum value of water which could evaporate and T1 is equal to the wet bulb temperature.
XTI: which evaluates the temperature at water droplet to air interface. It could be
summarized as follows: o if XTI = 0 then the temperature of the interface is equal to the
temperature of the liquid water; o if XTI = 1 then the temperature of the interface is equal to the
medium value of the air across the considered stage.
XTL: which evaluates the temperature of the surface of the water droplet. It could be observed that:
o if XTL = 0 the surface water droplet temperature coincides to the water temperature at the injection point (which means that no thermal exchange between air and water droplet occurs and then the surface water droplet temperature remains constant);
o if XTL = 1 the surface water droplet temperature is equal to the medium value of the air across the considered stage.
XW: which divides the evaporation of water between the upstream and the
downstream of the stage calculation. This means: o if XW = 0 all the water evaporates downstream of the stage; o if XW = 1 all the water evaporates upstream of the stage;
First of all, the value of the XIF parameter was found by using the conditions
highlighted in Fig. 188. From this figure, it could be observed that about the 3.5% of the total amount of the injected water evaporates upstream of the inlet of the first stage, in case of droplet diameter equal to 10 μm. In case of droplet diameters equal to 20 and 30 μm, the percentage of evaporated water on the total injected amount is equal to 3.2% and 2.7%, respectively. To obtain results similar to the ones found by Utamura, a value of XIF equal to 0.16 could be adopted.
The next step of the tuning procedure was developed by considering a constant value of XIF equal to 0.16 and values of XTL, XW and XTI, varying from 1 to 0. In the Figs. from 189 to 197 the obtained results (dotted lines) compared with the ones of Utamura (continuous lines) are presented, for many combinations of the three tuning parameters.
184
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compressor stages
XIF
= 0.16
XTI
= 0.80
XTL
= 0.00
XW
= 0.50
d = 10 μmd = 20 μmd = 30 μm
Figure 189 – Wet compression model tuning (case XIF=0.16, XTI=0.80, XTL=0.00, XW=0.50)
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compressor stages
XIF
= 0.16
XTI
= 0.60
XTL
= 0.00
XW
= 0.50
d = 10 μmd = 20 μmd = 30 μm
Figure 190 – Wet compression model tuning (case XIF=0.16, XTI=0.60, XTL=0.00, XW=0.50)
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compressor stages
XIF
= 0.16
XTI
= 0.40
XTL
= 0.00
XW
= 0.50
d = 10 μmd = 20 μmd = 30 μm
Figure 191 – Wet compression model tuning (case XIF=0.16, XTI=0.40, XTL=0.00, XW=0.50)
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XIF
= 0.16
XTI
= 0.50
XTL
= 0.00
XW
= 0.50
d = 10 μmd = 20 μmd = 30 μm
Figure 192 – Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.00, XW=0.50)
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XIF
= 0.16
XTI
= 0.50
XTL
= 0.20
XW
= 0.50
d = 10 μmd = 20 μmd = 30 μm
Figure 193 – Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.20, XW=0.50)
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compressor stages
XIF
= 0.16
XTI
= 0.50
XTL
= 0.40
XW
= 0.50
d = 10 μmd = 20 μmd = 30 μm
Figure 194 – Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.40, XW=0.50)
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compressor stages
XIF
= 0.16
XTI
= 0.50
XTL
= 0.00
XW
= 0.00
d = 10 μmd = 20 μmd = 30 μm
Figure 195 – Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.00, XW=0.00)
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compressor stages
XIF
= 0.16
XTI
= 0.50
XTL
= 0.00
XW
= 0.50
d = 10 μmd = 20 μmd = 30 μm
Figure 196 – Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.00, XW=0.50)
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comp.inlet duct
compressor stages
XIF
= 0.16
XTI
= 0.50
XTL
= 0.00
XW
= 1.00
d = 10 μmd = 20 μmd = 30 μm
Figure 197 – Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.00, XW=1.00)
From the Figs. from 189 to 197, it could be observed that:
others factors been equal, as XTI decreases (which means that the temperature of the water to air interface decrease), the evaporation velocity through the compressor decreases (see Figs. from 189 to 191);
others factors been equal, as XTL increases (which means that the temperature of the droplet surface increases), the evaporation velocity through the compressor increases (see Figs. from 192 to 194);
others factors been equal, the influence of the parameter XW on the evaporation velocity through the compressor is negligible (see Figs. from 195 to 197);
According to the previous considerations, is quite evident that the tuning of the wet
compression model, could be done as follows: XIF = 0.16 as already explained;
XTI = 0.50 which means to consider the temperature of the water to air interface at a temperature equal to the medium value between the temperature of the surrounding air and the one of the liquid water;
XW = 0.50 which, as already seen, do not influences the evaporation velocity in sensible way;
XTL = f (d) which mean, others factors been equal, to consider the value of the droplet surface temperature as function of the droplet diameter. This last evidence is coherent with the real physic phenomenon; in fact a water droplet with a small diameter change is surface temperature more rapidly than a droplet with a large diameter. In the same way, it could be assumed higher values of the parameter XTL in correspondence of small droplet diameter.
189
The last considerations allows to drawn the trends in Fig. 198 in which was considered XIF = 0.16, XTI = 0.50, XW = 0.50 and for what concerns XTI was considered as follows:
⎪⎩
⎪⎨
⎧
μ=μ=μ=
=m30d00.0m20d10.0m10d45.0
XTL
for for for
0
20
40
60
80
100-2 -1 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
evap
orat
ed w
ater
/ in
ject
ed w
ater
[%]
Inje
ctio
n Po
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comp.inlet duct
compressor stages
XIF
= 0.16
XTI
= 0.50
XTL
= f (d)
XW
= 0.50
d = 10 μmd = 20 μmd = 30 μm
Figure 198– Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=f(d), XW=0.00)
From the Fig. 198 it could be observed that the results obtained with the wet
compression model implemented in the IN.FO. G.T. E. program reproduces the ones of Utamura with an acceptable approximation.
190
4.7 – IN.FO. G.T. E APPLICATION ON GE FRAME 7 EA The IN.FO. G.T. E. program was adapted to study the performance of an existing GE
Frame 7 EA gas turbine model in presence of interstage injection. This gas turbine (already introduced in the Chapter 4) is a medium size heavy duty
gas turbine which consists of a seventeen stages axial compressor and a three stages axial turbine with a produced power output of about 84 MW and an electrical efficiency of about 32.4 %.
First of all, by a “trial and error” procedure the program was tuned to reproduce the performance of the machine at different ambient conditions and at different values of turbine inlet temperature.
In particular the tuning was realized on 12 points of machine operation corresponding to four values of ambient temperature (5, 15, 30 and 40 °C) and three values of turbine inlet temperature (1113, 986 and 857 °C). The selected values of turbine inlet temperature correspond to the machine load equal to 100%, 80% and 60%, respectively.
The tuning procedure allows to determinate all the parameters required by the GAS TURBINE SET UP array.
On the following the main input used to reproduce the performance of the Frame 7 EA are presented and discussed.
4.7.1 – Frame 7EA Input Definition The performance of the compressor are evaluated by the use of the stage compressor
maps (already presented in the Chapter 4) according to the calculation procedure described in the paragraph 4.5. The values of pressure coefficient ( Ψ p), flux coefficient (φ ), shape factor (SF), small stage isentropic efficiency (η) used to define the compressor, are presented in Tab. 24
Table 24 – Compressor stages performance maps definition coefficients
STAGE ΨP,MAX φ(ΨP,MAX) SF φDES ΨP,DES ηDES η(ΨP/φ)MAX η(ΨP/φ)MIN (ΨP/φ)MAX (ΨP/φ)MIN
1 1.100 0.873 -0.5 0.559 0.271 0.892 0.910 0.200 1.456 0.040 2 1.100 0.873 -0.5 0.566 0.269 0.886 0.910 0.200 1.456 0.040 3 1.100 0.873 -0.5 0.575 0.271 0.886 0.910 0.200 1.456 0.040 4 1.100 0.873 -0.5 0.595 0.284 0.886 0.910 0.200 1.456 0.040 5 1.100 0.873 -0.5 0.602 0.283 0.887 0.910 0.200 1.456 0.040 6 1.100 0.873 -0.5 0.598 0.282 0.890 0.910 0.200 1.456 0.040 7 1.100 0.873 -0.5 0.598 0.282 0.891 0.910 0.200 1.456 0.040 8 1.100 0.873 -0.5 0.598 0.281 0.890 0.910 0.200 1.456 0.040 9 1.100 0.873 -0.5 0.599 0.280 0.890 0.910 0.200 1.456 0.040
10 1.100 0.873 -0.5 0.599 0.279 0.892 0.910 0.200 1.456 0.040 11 1.100 0.873 -0.5 0.599 0.279 0.892 0.910 0.200 1.456 0.040 12 1.100 0.873 -0.5 0.599 0.278 0.892 0.910 0.200 1.456 0.040 13 1.100 0.873 -0.5 0.599 0.278 0.892 0.910 0.200 1.456 0.040 14 1.100 0.873 -0.5 0.599 0.278 0.891 0.910 0.200 1.456 0.040 15 1.100 0.873 -0.5 0.599 0.277 0.892 0.910 0.200 1.456 0.040 16 1.100 0.873 -0.5 0.599 0.277 0.892 0.910 0.200 1.456 0.040 17 1.100 0.873 -0.5 0.598 0.280 0.891 0.910 0.200 1.456 0.040
191
Moreover the definition of some geometrical parameters of the compressor are required by the code. In Fig. 199 and in Tab. 25, the medium radiuses of the 17 compressor stages with also the tip and root radius implemented in the calculation code, are shown.
0.0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1.00.5 1.5 2.5 3.5 4.5 5.5 6.5 7.5 8.5 9.5 10.5 11.5 12.5 13.5 14.5 15.5 16.5
compressor stage
1st 2nd 3rd 4th4th 5th 6th5th 7th 8th 9th 10th 11th 12th 13th 14th 15th 16th 17th
[m]
RRO
OTR
MED
IUM
R TIP
Figure 199 – Mean, tip and root radiuses of the 17 compressor stages of the real Frame 7 EA
Table 25 – Mean, tip and root radiuses of the 17 compressor stages of the GE Frame 7 EA
MEAN RADIUS
BLADE HEIGHT
ROOT RADIUS
TIP RADIUS STAGE
RM [m] HB [m] RR [m] RT [m] 1 0.736 0.340 0.566 0.906 2 0.726 0.292 0.580 0.872 3 0.716 0.255 0.589 0.844 4 0.692 0.230 0.577 0.807 5 0.685 0.205 0.582 0.787 6 0.685 0.182 0.594 0.776 7 0.685 0.163 0.603 0.767 8 0.685 0.147 0.611 0.758 9 0.685 0.133 0.618 0.751
10 0.685 0.121 0.624 0.745 11 0.685 0.111 0.630 0.740 12 0.685 0.101 0.634 0.735 13 0.685 0.093 0.638 0.731 14 0.685 0.086 0.642 0.728 15 0.685 0.080 0.645 0.725 16 0.685 0.074 0.648 0.722 17 0.685 0.069 0.651 0.719
192
Finally the position of the air bleed ports, placed at the fifth and at the last stage, were defined.
The pressure losses through the combustion chamber are equal to the 4% of the pressure at its inlet while no pressure losses at inlet of the compressor and at the outlet of the turbine were considered. Moreover the gas turbine was evaluated without water or steam injection into the combustion chamber.
For what concerns the turbine, a map characterized by a constant value of the Flow Function was chosen as presented in Fig. 200.
0.75
0.80
0.85
0.90
0.95
1.00
1.05
1.10
1.15
1.20
0.94 0.96 0.98 1.00 1.02 1.04 1.06
β*
FF*
Figure 200 – Frame 7 EA Gas turbine performance map
This choice for the gas turbine performance map (together with the use of
compressor stage performance maps characterized by a shape factor equal to –0.5) allows to reproduce with high accuracy the performance of the Frame 7 EA as it will be shown on the follow. It should be observed that the characterization of the same machine developed in the previous chapter by the use of a commercial program [3] was possible only by using the turbine performance map presented in Fig. 105 (pag. 110). This was mainly due to the convergence problem that the commercial program highlighted with the implemented model. The aim of the developing a new calculation code was born also to overcome this limit. Obviously the first goal realized by the IN.FO. G.T. E. calculation code remains the possibility of study the wet compression by the use of an evaporative model able to take into account all the physic parameters involved in this process. The commercial codes for energy systems simulation do not allow this purpose.
According to the previous assumptions, in Figs. from 201 to 209 the comparison between the main performance parameters of the Frame 7 EA evaluated by the computational code (square markers) and the values of the real machine (continuous
193
lines) for three loads (equal to 100%, 80% and 60%) as function of ambient temperature are presented.
Moreover, in Tab. 26 the evaluated results and the percentage differences between the real machine and the computational code, are reported.
35000
45000
55000
65000
75000
85000
95000
5 10 15 20 25 30 35 40air ambient temperature [°C]
gas t
urbi
ne p
ower
out
put [
kW]
load = 60%load = 80%load = 100%
Figure 201 – Gas turbine power output of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
194
10000
10500
11000
11500
12000
12500
13000
13500
14000
5 10 15 20 25 30 35 40air ambient temperature [°C]
Hea
t Rat
e [k
J/kW
h]
load = 60%load = 80%load = 100%
Figure 202 – Heat Rate of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
250
260
270
280
290
300
310
320
5 10 15 20 25 30 35 40air ambient temperature [°C]
air m
ass f
low
rate
[kg/
s]
load = 60%load = 80%load = 100%
Figure 203 – Air mass flow rate of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
195
9.5
10.0
10.5
11.0
11.5
12.0
12.5
13.0
13.5
5 10 15 20 25 30 35 40air ambient temperature [°C]
pres
sure
ratio
load = 60%load = 80%load = 100%
Figure 204 – Pressure ratio of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
300
325
350
375
400
5 10 15 20 25 30 35 40air ambient temperature [°C]
com
pres
sor d
isch
arge
tem
pera
ture
[°C]
load = 60%load = 80%load = 100%
Figure 205 – Compressor discharge temperature of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
196
300
350
400
450
500
550
600
5 10 15 20 25 30 35 40air ambient temperature [°C]
turb
ine
outle
t tem
pera
ture
[°C]
load = 60%load = 80%load = 100%
Figure 206 – Turbine outlet temperature of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
80000
85000
90000
95000
100000
105000
110000
115000
120000
5 10 15 20 25 30 35 40air ambient temperature [°C]
com
pres
sor p
ower
[kW
]
load = 60%load = 80%load = 100%
Figure 207 – Compressor power of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
197
120000
140000
160000
180000
200000
220000
5 10 15 20 25 30 35 40air ambient temperature [°C]
turb
ine
pow
er [k
W]
load = 60%load = 80%load = 100%
Figure 208 – Turbine power of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
2.0
2.5
3.0
3.5
4.0
4.5
5.0
5.5
6.0
5 10 15 20 25 30 35 40air ambient temperature [°C]
fuel
mas
s flo
w ra
te [°
C]
load = 60%load = 80%load = 100%
Figure 209 – Fuel mass flow rate of real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
198
Table 26 – Performance parameters of real Frame 7 EA (Fr7EA) and evaluated by the code (IN.FO. G.T. E.) as function of ambient temperature and load
LOAD 100 % (TIT = 1113 °C) 80 % (TIT = 986°C) 60 % (TIT = 856 °C) Tamb [°C] IN.FO.
G.T. E. Fr7EA Diff. %
IN.FO. G.T. E. Fr7EA Diff.
% IN.FO. G.T. E. Fr7EA Diff.
% MAIR [kg/s] 311.35 309.80 0.5 311.40 309.80 0.5 311.40 309.80 0.5
pressure ratio 13.20 13.13 0.5 12.54 12.46 0.6 11.83 11.76 0.6 T2 [°C] 354 358 -1.1 344 344 0.0 334 327 2.0 T4 [°C] 534 528 1.1 454 454 0.0 377 381 -0.9
MFUEL [kg/s] 5.52 5.46 1.0 4.58 4.54 0.9 3.65 3.64 0.2 PTUR [kW] 204149 205572 -0.7 183771 182836 0.5 162447 159324 2.0 PCOM [kW] 110906 112011 -1.0 107722 107332 0.4 104288 101941 2.3 PGT [kW] 89347 89841 -0.5 72510 72180 0.5 54993 54376 1.1
5
HR [kJ/kWh] 11136 10957 1.6 11366 11331 0.3 11960 12069 -0.9
MAIR [kg/s] 296.90 297.20 -0.1 297.12 297.00 0.0 297.20 297.00 0.1 pressure ratio 12.60 12.61 0.0 11.97 12.00 -0.2 11.30 11.30 0.0
T2 [°C] 366 367 -0.3 356 353 0.8 345 337 2.4 T4 [°C] 537 534 0.6 458 460 -0.4 381 388 -1.8
MFUEL [kg/s] 5.21 5.20 0.2 4.31 4.31 0.0 3.42 3.45 -1.0 PTUR [kW] 194512 195581 -0.5 174603 173684 0.5 154472 151329 2.1 PCOM [kW] 106797 107580 -0.7 103753 103060 0.7 100483 97937 2.6 PGT [kW] 83998 84412 -0.5 67481 67385 0.1 50972 50444 1.0
15
HR [kJ/kWh] 11174 11096 0.7 11489 11520 -0.3 12089 12339 -2.0
MAIR [kg/s] 275.35 277.9 -0.9 275.6 277.9 -0.8 275.75 277.9 -0.8 pressure ratio 11.72 11.83 -0.9 11.13 11.22 -0.8 10.51 10.59 -0.8
T2 [°C] 382 380 0.5 372 366 1.6 362 349 3.7 T4 [°C] 542 546 -0.7 465 472 -1.5 391 399 -2.0
MFUEL [kg/s] 4.78 4.83 -1.1 3.94 4.00 -1.4 3.11 3.18 -2.2 PTUR [kW] 180308 180927 -0.3 161513 160417 0.7 141531 139161 1.7 PCOM [kW] 100398 100926 -0.5 97550 96700 0.9 94579 91811 3.0 PGT [kW] 76458 76586 -0.2 60840 60608 0.4 44177 44484 -0.7
30
HR [kJ/kWh] 11265 11368 -0.9 11653 11876 -1.9 12673 12872 -1.5
MAIR [kg/s] 260.10 264.50 -1.7 260.50 264.50 -1.5 260.73 264.50 -1.4 pressure ratio 11.11 11.29 -1.6 10.56 10.72 -1.5 9.98 10.12 -1.4
T2 [°C] 391 387 1.0 381 373 2.1 371 356 4.2 T4 [°C] 549 556 -1.3 476 481 -1.0 401 408 -1.7
MFUEL [kg/s] 4.52 4.61 -2.0 3.71 3.80 -2.4 2.92 3.02 -3.4 PTUR [kW] 170490 171268 -0.5 151310 151596 -0.2 132468 131461 0.8 PCOM [kW] 95669 96319 -0.7 93025 92271 0.8 90295 87641 3.0 PGT [kW] 71556 71637 -0.1 55352 56288 -1.7 39568 40996 -3.5
40
HR [kJ/kWh] 11374 11598 -1.9 12077 12174 -0.8 13290 13288 0.0 It could be observed from the figures and from table that the IN.FO. G.T. E. program
is able to reproduce the performance of the real machine with an acceptable approximation.
The greatest differences between the program and the real Frame 7 EA occur at lower gas turbine load at higher temperatures. In operation at full load (turbine inlet temperature equal to 1113 °C) the percentage difference between the program and the real Frame 7 EA is always less than the 2% while at a load of 60% (turbine inlet temperature equal to 856 °C) the maximum difference is of about the 3.5%.
199
4.7.2 – Evaluation Of Frame 7EA With Water Injection By Using The IN.FO G.T. E. Calculation Code
The performance change of GE Frame 7EA due to water injection has been evaluated by using the IN.FO. G.T. E calculation code. This evaluation has been developed also in the Chapter 3 by using a commercial software package [3]. Due to the use of this software for the simulation, the evaporation model used in this software was adopted. In this evaporation model, a thermal equilibrium between the two flows is assumed to occur instantaneously at the injection point. As a result, effects of droplet dynamics (influence of droplet diameter and temperature and the diffusion theory of two phase flow) and evaporation rate are not considered. The last sentence means that the results obtained in the previous chapter should be considered the maximum performance improvement due to water injection.
The use of the IN.FO. G.T. E calculation code, instead, allows to take into account the water evaporation dynamics and moreover to better evaluate the performance of the gas turbine, as already explained.
With the previous consideration the results regarding the performance change of Frame 7 EA in presence of interstage injection as function of the injected water to air ratio, are presented in the Figs. from 210 to 211. This figures are drawn considering a water injection from the upstream of the first stage (WIP1, that means into the compressor inlet duct) to the upstream of the eighth stage (WIP8) with reference to ambient conditions equal to 40 °C of air temperature, 40% of relative humidity and 1.013 bar of ambient pressure. The performance parameters in this figures, are presented normalized on the respect of the corresponding value at ISO conditions (15 °C, 1.013 bar and 60% of RH).
In Fig. 210, the change in power output is presented for the considered injection locations.
0.80
0.85
0.90
0.95
1.00
1.05
1.10
0.0 0.5 1.0 1.5 2.0 2.5
PGT
injected water to air ratio [%]
PGT,ISO
WIP1
WIP3
WIP2
WIP4
WIP8 WIP7WIP6
WIP5
Figure 210 – Power output change as function of injected water to air ratio and injection location
200
From Fig. 210, it could be observed that the gas turbine power output increase depends on the injection water location and injected water mass flow rate. In particular injected water mass flow rate being constant, the gas turbine power boost decrease passing from a water injection from the upstream of the first stage to the upstream of the eight stage. This evidence was also found in the evaluation developed with the commercial software package. From Fig. 210, it could be also observed that the maximum power boost is limited by the achieving of the compressor surge. In particular the compressor surge is reached for a water mass flow rate injected at the upstream of the first stage equal to about 7.0 kg/s while for the others stage the maximum mass of liquid water that could be injected is about equal to 6.1 kg/s. Finally, it could be seen that the power loss due to the ambient conditions is recuperated in case of water injection from the upstream of the first stage to the upstream of the sixth stage.
The change in heat rate due to water injection is presented in Fig. 211. As evident from the figure, it could be seen that the maximum reduction of hear rate is obtained with 7 kg/s of water injected at the upstream of first stage. In this case the gas turbine heat rate is less than the value at ISO conditions of about 1.5 %. Also by injecting from the second stage to the fifth stage the heat rate losses due to ambient conditions could be recovered. The evaluations performed with the IN.FO G.T. E. calculation code confirm that the fogging is a strategy more effective to gain power than efficiency, such as also the commercial software package shown.
0.98
0.99
1.00
1.01
1.02
0.0 0.5 1.0 1.5 2.0 2.5
HRGT
injected water to air ratio [%]
HRGT,ISO
WIP1
WIP3
WIP2
WIP4
WIP8 WIP7WIP6
WIP5
Figure 211 – Heat rate change as function of injected water to air ratio and injection location
The change of air mass flow rate at compressor inlet and of compressor pressure
ratio are presented in Fig. 212 and 213 respectively. From these figures it could be observed that the values corresponding to the ISO
conditions could not be achieved in any case.
201
0.88
0.90
0.92
0.94
0.96
0.0 0.5 1.0 1.5 2.0 2.5
MAIR
injected water to air ratio [%]
MAIR,ISO
WIP1WIP3
WIP2
WIP4
WIP8WIP7WIP6
WIP5
Figure 212 – Air mass flow rate at compressor inlet change as function of injected water to air ratio and injection location
0.88
0.90
0.92
0.94
0.96
0.98
0.0 0.5 1.0 1.5 2.0 2.5
βCOM
injected water to air ratio [%]
βCOM,ISO
WIP1WIP3
WIP2
WIP4
WIP8WIP7
WIP6
WIP5
Figure 213 – Compressor pressure ratio change as function of injected water to air ratio and injection location
202
In Fig. 214 the ratio between liquid water through the compressor and the injected water is presented as function of injection location for a total mass of injected water equal to 5 kg/s. It could be observed from the figure that as the injection location moves from the upstream of the first stage to the upstream of the eighth stage the evaporation velocity increases due to the higher air temperatures into the compressor.
0.0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1.0-2 -1 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
liqui
d w
ater
/ to
tal i
njec
ted
wat
er
Inje
ctio
n Po
int
1st stag
e IN
1st st
age
OU
T
2nd st
age
OU
T
3rd st
age
OU
T
4th st
age
OU
T
5th st
age
OU
T
6th st
age
OU
T
7th st
age
OU
T
8th st
age
OU
T
9th st
age
OU
T
10th
stag
e O
UT
11th
stag
e O
UT
12th
stag
e O
UT
13th
stag
e O
UT
14th
stag
e O
UT
15th
stag
e O
UT
16th
stag
e O
UT
17th
stag
e O
UT
comp.inlet duct
compressor stages
mass of injected water = 5 kg/s
Figure 214 – Liquid water through the compressor to injected water ratio as function of injection location for a total mass of injected water equal to 5 kg/s
The figures from 210 to 214 show the performance of Frame 7 EA with water
injection considering a water injected temperature and diameter equal to 15°C and 10 μm, respectively. In particular the influence of droplet diameter on performance improvement is a critical topic for fogging evaluation. To understand this influence three different droplet diameters were taken into account. The Figs. from 215 to 218 evaluates the performance improvement of GE frame7 EA in case of injection location in the compressor inlet duct for three different water droplet diameters (10, 20 and 30 μm) for an injected water mass flow rate passing from 0 to 4 kg/s always with reference to 40 °C of air temperature and 40% of RH.
From Fig. 215 it could be observed that the power boost due to water injection decrease of about 3.3 percentage points passing from a droplet diameter equal to 10 μm to 30 μm.
Also the increase in heat rate (Fig. 216), air mass flow rate at the compressor inlet (Fig. 217) and in compressor pressure ratio (Fig. 218) show a reduction passing from a droplet diameter of 10 μm to 30 μm.
The last evidence could be explained considering that as the droplet diameter increases, the evaporation velocity decreases and then the end of water evaporation occurs at a greater distance from the compressor inlet section. The delay in evaporation is responsible of the decrease of performance improvement. The delay of water evaporation is presented in Fig. 219 for the three considered droplet diameters with reference to 4 kg/s of water injected into the compressor inlet duct.
203
0.84
0.86
0.88
0.90
0.92
0.94
0.96
0.98
1.00
1.02
0.0 0.3 0.6 0.9 1.2 1.5
PGT
injected water to air ratio [%]
PGT,ISO
d = 10 μm
d = 20 μmd = 30 μm
Figure 215 – Power output change as function of injected water to air ratio and injected droplet diameters for an injection in the compressor inlet duct
0.98
0.99
1.00
1.01
1.02
0.0 0.3 0.6 0.9 1.2 1.5
HRGT
injected water to air ratio [%]
HRGT,ISO
d = 10 μm
d = 20 μmd = 30 μm
Figure 216 – Heat rate change as function of injected water to air ratio and injected droplet diameters for an injection in the compressor inlet duct
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MAIR
injected water to air ratio [%]
MAIR,ISO
d = 10 μm
d = 20 μmd = 30 μm
Figure 217 – Air mass flow rate at compressor inlet change as function of injected water to air ratio and injected droplet diameters for an injection in the compressor inlet duct
0.88
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0.95
0.0 0.3 0.6 0.9 1.2 1.5
βCOM
injected water to air ratio [%]
βCOM,ISO
d = 10 μmd = 20 μmd = 30 μm
Figure 218 – Compressor pressure ratio change as function of injected water to air ratio and injected droplet diameters for an injection in the compressor inlet duct
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Figure 219 – Liquid water through the compressor to injected water ratio as function of injected water droplet diameter for a total mass of injected water equal to 4 kg/s
4.7.3 – Comparison Between IN.FO. G.T. E. And Commercial Software
Package The results obtained by using the developed calculation code could be compared to
the ones which could be evaluated by the use a commercial program. To achieve this goal a model of Frame 7 EA was realized by the use of Thermoflow 14 calculation code [90]. This new version of the software package shows a stronger convergence on the respect of its previous version; in fact this new version allows the use of turbine performance map with constant flow function. This makes possible the comparison between the results obtained with the two codes. The Frame 7 EA model realized with Thermoflow 14 was tuned to reproduce the performance of the machine as performed by the IN.FO. G.T. E.
In Figs from 220 to 223 the comparison among the resulting gas turbine power output (Fig. 220), heat rate (Fig. 221), air mass flow rate at compressor inlet (Fig. 222) and compressor pressure ratio (Fig. 223) of the real machine (continuous lines) and evaluated by the IN.FO. G.T. E. (square shape markers) and by the commercial software package (circle shape markers) for three gas turbine loads (100%, 80% and 60%) and for an air ambient temperature changing from 5 to 40 °C.
From this figures it could be observed that there is a good correspondence between the results evaluated by the two codes with the exception of the heat rate. In particular the heat rate evaluated by the commercial software shows a small sensibility to the variation of ambient temperature. According to the Author’s knowledge this fact could be due to the convergence criteria used in this code which is not specifically realized for the aero-thermodynamic analysis.
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35000
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gas t
urbi
ne p
ower
out
put [
kW]
air ambient temperature [°C]
load = 60%load = 80%load = 100% reference
InFo GTEComm. Pr.
Figure 220 – Gas turbine power output of the real Frame 7 EA (continuous lines), evaluated by the IN.FO G.T E (square markers) and by a commercial software (circle markers) as function of ambient temperature and load
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load = 60%load = 80%load = 100% reference
InFo GTEComm. Pr.
Hea
t Rat
e [k
J/kW
h]
Figure 221 – Heat rate of the real Frame 7 EA (continuous lines), evaluated by the IN.FO G.T E (square markers) and by a commercial software (circle markers) as function of ambient temperature and load
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load = 60%load = 80%load = 100% reference
InFo GTEComm. Pr.
air m
ass f
low
rate
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Figure 222 – Air mass flow rate at compressor inlet of the real Frame 7 EA (continuous lines), evaluated by the IN.FO G.T E (square markers) and by a commercial software (circle markers) as function of ambient temperature and load
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Figure 223 – Pressure ratio of the real Frame 7 EA (continuous lines), evaluated by the IN.FO G.T E (square markers) and by a commercial software (circle markers) as function of ambient temperature and load
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The comparison of the two codes is realized to evaluate the difference in performance evaluation between the use of a dynamic evaporation model (performed by the IN.FO. G.T. E) and a stationary evaporation model (performed by the Thermoflow).
It should be remembered that the commercial software does not take into account the influence of droplet diameter and the delay in water evaporation model. In others word the water evaporation in the commercial software is evaluated only by considering the air saturation capacity.
In Fig. 224 the power output change (evaluated as ratio between the power output with and without water injection) for a water injection which occurs into the compressor inlet duct is presented with reference to 40°C of air temperature and 40% of relative humidity. The evaluation developed by the IN.FO. G.T. E was realized by considering a droplets diameter equal to 10 μm and a water temperature of 15 °C.
From the figure it could be observed that injected water being constant, the power boost evaluated by the commercial program is greater than the one evaluated by the IN.FO. G.T. E. This evidence is due to the different adopted evaporation models. Moreover it could be observed that the maximum amount of water which is possible to inject before the compressor surge achievement is less in the evaluation of the commercial program than the calculation of the IN.FO. G.T. E. Obviously the differences by the two codes increase considering greater droplet diameters in the use of IN.FO. G.T. E.
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PGT
injected water to air ratio [%]
PGT,HOT Comm.
Prog.
IN.FO. G.T. E.
Figure 224 – Power output change as function of injected water to air ratio evaluated by a commercial software and by IN.FO. G.T. E.
In Fig. 225, the comparison between the evaluation of heat rate is presented. It could
be seen that the reduction in heat rate is greater in the evaluation of the IN.FO. G.T. E. than in the ones of Thermoflow. This fact is not coherent with the assumed hypothesis. This evidence could be only explained by considering the small variation of heat rate to
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temperature variation showed by the gas turbine model realized by the use of the commercial program. Concluding the results presented in Fig. 225 could be not effective to compare the different evaporation model influence on the gas turbine performance evaluation.
The comparison between air mass flow rate and compressor pressure ratio are shown in Figs. 226 and 227 respectively.
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injected water to air ratio [%]
HRGT,HOT
Comm.Prog.
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Figure 225 – Heat rate change as function of injected water to air ratio evaluated by a commercial software and by IN.FO. G.T. E.
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MAIR
injected water to air ratio [%]
MAIR,HOT
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Figure 226 – Air mass flow rate change as function of injected water to air ratio evaluated by a commercial software and by IN.FO. G.T. E.
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injected water to air ratio [%]
βCOM,HOT
Comm.Prog.
IN.FO. G.T. E.
Figure 227 – Compressor pressure ratio change as function of injected water to air ratio evaluated by a commercial software and by IN.FO. G.T. E.
Figs. 226 and 227 confirms that the increase of the considered parameters evaluated
by the commercial code are greater than the one calculated by the IN.FO. G.T. E. Finally the evaporation rate through the compressor stages evaluated by the two
calculation codes in correspondence of an injected water mass flow rate equal to 5 kg/s is presented. As expected, the commercial program shows a greater evaporation velocity than the IN.FO. G.T. E.
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Figure 228 – Liquid water through the compressor to injected water ratio as function of injected water droplet diameter for a total mass of injected water equal to 5 kg/s evaluated by a commercial software and by IN.FO. G.T. E.
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5
Concluding Remarks
In the present work, the study of gas turbine and combined cycle power plants performance improvement due to fogging (considering inlet fogging, overspray and interstage injection strategies) has been developed.
The comprehensive analysis of a large number of commercial combined cycle power plants with varying net power output (from 8 MW to 380 MW) in presence of both evaporative and overspray fogging, showed the following concluding remarks:
high pressure fogging is effective for combined cycles in recovering the lost
power output due to high ambient temperatures. The amount of power boost increases with the increased amount of overspray fogging;
the gain in net power output due to fogging, both inlet evaporative and overspray
fogging, is mainly associated with the performance enhancement in the gas turbine section of a combined cycles;
the amount of net power boost is found higher for a combined cycle with
traditional or aero-derivative machines compared to a combined cycle using an advanced technology gas turbine;
the amount of injected water, normalized by saturated air flow rate at the
compressor inlet and percent net power boost, is smaller (by approximately, 1 liter/hr per unit flow rate and per unit % net power boost) for a combined cycle
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using an aero-derivative gas turbine than a combined cycle using other two categories of gas turbines. These results imply that an aero-derivative gas turbine based combined cycle plant will have lower operating cost with fogging compared to a combined cycle using an advanced technology or a traditional gas turbine;
In spite of the fact that a combined cycle using an advanced technology gas
turbine showed smaller power boost due to fogging, its incremental fuel efficiency is found higher compared to the systems with other two types of gas turbines.
Moreover the investigation on the effects of interstage injection on an existing gas
turbine model (GE Frame 7EA) performance has been conducted. In particular, stage-by-stage performance characteristics of the compressor and the overall gas turbine performance parameters have been evaluated considering different injection points, water mass flow rates and climatic conditions first by the use of a commercial software package for the energy systems simulation, then by a self developed calculation code (IN.FO. G.T.E). This code, developed by DIEM – University of Bologna, is able to evaluate the performance change of a gas turbine in presence of overspray or interstage injection taking into account all the physical parameters that influence the water evaporation in air (droplet diameter and temperature, diffusion coefficient, etc.). The developed calculation code evaporation model was compared to the results available in literature showing good precision and affidability.
Based on the results obtained, following concluding remarks can be summarized:
a careful examination of the gas turbine first and last stage performance characteristics shows redistribution of the stage loading (unloading of the first stage and increased loading of the last stage) for the injection locations and the ambient conditions examined. Increased amount of water injection causes last stage operating point to move closer to the surge line suggesting that one has to be cautious in selecting the water amount in case of the interstage injection approach;
the maximum power boost is obtained if the water injection is realized upstream
of the compressor independent of the climatic conditions and compared to the other injection locations;
the total water amount that is possible to inject into the compressor is limited by
ambient conditions, maximum allowable gas turbine power output (due to mechanical strength of the shaft and electric generator capacity) and the compressor surge limit;
the achievable power boost decrease as the injected water droplet diameter
increase.
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6
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[73] Walsh, P. P., Mathieson, D., Bicknell, G. and Matthews, K., Inlet Fog Boost Technology Acquisition Programme, Power-Gen Europe (2000).
[74] Hoffman, J and McKay T., “Customer Benefits of Air Inlet Cooling and ALFog Fogging and High Fogging,” Powergen Far East, October 2004
[75] Nolan, J.P. and Twombly, V.J., (1990), “Gas Turbine Performance Improvement Direct Mixing Evaporative Cooling System,” ASME Paper No: 90-GT-368, International Gas Turbine and Aeroengine Congress, Brussels, Belgium, June 11-14, 1990.
• [76] Gas Turbine World 2001-2002 Handbook”, 2002, Pequot Publishing Inc., 654 Hillside Rd., Fairfield - CT 06430, USA.
[77] Bhargava, R., and Meher-Homij, C. B., 2002, "Parametric Analysis Of Existing Gas Turbines With Inlet Evaporative And Overspray Fogging," ASME Paper No. GT-2002-30560.
• [78] “Audizione Del Gestore Della Rete Presso Le Commissioni Riunite Industria E Ambiente E Territorio Del Senato Della Repubblica” Roma, 23 settembre 2003
[79] Climate Diagnostic Center Archive of National Oceanic and Atmospheric Administration (NOAA) in collaboration with the Cooperative Institute For Research in Environmental Sciences (CIRES)
[80] Benvenuti, E., Bettocchi, R., Cantore, G., Negri di Montenegro, G. and Spina, P. R., 1993, “Gas Turbine Cycle Modeling Oriented to Component Performance Evaluation from Limited Design or Test Data”, Proc. 7th ASME COGEN - TURBO, Bournemouth, UK, IGTI Vol. 8, pp. 327-337.
[81] Bettocchi, R., Spina, P. R., 1997, “Influence of the Control Logic on Gas Turbine Operation”, ASME Paper 97-AA-127
[82] Spina, P. R., 2002, “Gas Turbine Performance Prediction by Using Generalized Performance Curves of Compressor and Turbine Stages”, ASME Paper GT-2002-30275.
220
[83] Howell, A. R., Bonham, R. P., 1950, "Overall and Stage Characteristics of Axial Flow Compressors", Proc. IMECHE, Vol. 163, pp. 235-248.
[84] Csanady, G. T., 1964, Theory of Turbomachines, McGraw-Hill, USA, pp. 47-52. [85] Spalding, D., B., (1963) “Convective mass transfer”, Edward Arnold Publishers,
London [86] Reid, R., C., Prausnitz, J., M., Poling, B., E., (1987) “The properties of gases and
liquids”, 4. ed., McGraw-Hill, USA [87] Eckert, E., R., G., Drake, Jr., R., M., (1959) “Heat and mass transfer”, 2. ed.,
McGraw-Hill, USA [88] Rohsenow, M., Choi, H., Y., (1961) “Heat, mass, and momentum transfer”,
Prentice-Hall, Inc., Englewood Cliffs, New Jersey [89] Bird, R., B., Stewart, W., E., Lightfoot, E., N., (1970) “Fenomeni di trasporto”,
CEA, Milano [90] Thermoflex 5.2, Thermoflow 14, Thermoflow Inc, Sudbury, MA, USA
221
222
7
List of figures Fig. 1 Ambient temperature influence on gas turbine power output
(@ 1.013 bar and 60% RH)
pag. 4
Fig. 2 Ambient temperature influence on gas turbine heat rate (@ 1.013 bar and 60% RH)
pag. 5
Fig. 3 Ambient pressure influence on gas turbine power output (@ 15 °C and 60% RH)
pag. 5
Fig. 4 Ambient pressure influence on gas turbine heat rate (@ 15 °C and 60% RH)
pag. 6
Fig. 5 Ambient relative humidity influence on gas turbine power output (@ 15 °C and 1.013 bar)
pag. 6
Fig. 6 Ambient relative humidity influence on gas turbine heat rate (@ 15 °C and 1.013 bar)
pag. 7
Fig. 7 Gas turbine thermodynamic efficiency as function of compressor pressure ratio and ambient temperature
pag. 8
Fig. 8 Operating point change passing from ISO day to an hot day in a typical compressor performance map
pag. 9
Fig. 9 Operating point change passing from ISO day to an hot day in a typical turbine performance map
pag. 9
Fig. 10 Entropy diagram on a hot day compared to one on ISO ambient conditions
pag. 10
223
Fig. 11 Air density as function of ambient temperature @ 1.013 bar and 60% of relative humidity
pag. 11
Fig. 12 Gas turbine power ratio as function of ambient temperature and compressor pressure ratio
pag. 12
Fig. 13 Compressor inlet air cooling strategies pag. 14
Fig. 14 Continuous cooling system with compression refrigerant plant
pag. 15
Fig. 15 Continuous cooling system with absorption refrigerant plant pag. 16
Fig. 16 Thermal energy storage with cold water pag. 17
Fig. 17 Thermal energy storage with ice pag. 17
Fig. 18 Psychrometric chart for a traditional evaporative cooling system
pag. 18
Fig. 19 Evaporative cooling hybrid system scheme pag. 19
Fig. 20 Psychrometric chart for an evaporative cooling hybrid system
pag. 19
Fig. 21 Psychrometric chart (@ p=1.01325 bar) pag. 20
Fig. 22 Variations of site ambient condition in a day pag. 25
Fig. 23 Psychrometrics of Inlet Fogging pag. 26
Fig. 24 Swirl (a) and impaction pin nozzles (b) and plume characteristics at operating pressure of 138 bar swirl jet nozzle (c) and impaction pin nozzle (d)
pag. 29
Fig. 25 Close up picture of nozzle spray plumes operating at 138 bar
pag. 29
Fig. 26 (a) Interaction of droplet to surrounding air condition (30 ºC and 20% RH) – Starting droplet size 20 microns (b) Interaction of droplet to surrounding air condition (30 ºC and 20% RH) – Starting droplet size 50 microns; (c) interaction of droplet to surrounding air condition (30 ºC and 60% RH) - Starting droplet size 20 microns ; (d) interaction of droplet to surrounding air condition (30 ºC and 60% RH) - Starting droplet size 50 microns
pag. 30
Fig. 27 Variation of the droplets size as a function of airflow velocity (measurements taken at 7.6 cm from the nozzle orifice
pag. 31
Fig. 28 Variation of droplets size as function of the applied pressure in the center and at the edge of the plume
pag. 32
Fig. 29 Variation of droplets size as function of the applied pressure in the center and at the edge of the plume
pag. 32
Fig. 30 Effects of distance between the nozzle tip and the measurement position and water flow rate on droplet size
pag. 33
Fig. 31 Effects of distance between the nozzle tip and the measurement position and water flow rate on droplet size
pag. 33
Fig. 32 V-Shaped duct configuration requiring special V-shaped fog nozzle array
pag. 34
224
Fig. 33 Response time of droplets to attain air stream velocity as a function of droplet size
pag. 35
Fig. 34 Fog nozzle disposition upstream filter and silencer pag. 36
Fig. 35 Fog nozzles in operation upstream filter and silencer pag. 36
Fig. 36 Fog nozzles in operation upstream filter and downstream silencer
pag. 36
Fig. 37 Trash screen for FOD prevention pag. 36
Fig. 38 Fog nozzles in operation downstream filter and silencer pag. 37
Fig. 39 Fog nozzles in operation downstream filter and upstream silencer and downstream
pag. 37
Fig. 40 Optimization of nozzle manifold position based on droplet size as a function of airflow velocity, and evaporation efficiency as a function residence time
pag. 38
Fig. 41 Plume shape with nozzle orientation of a fog nozzle in wind tunnel; airflow velocity is 4 m/s, Operating Pressure is 138 bar: (a) Co–flow; (b) Ninety degree; (c) Counter-flow.
pag. 40
Fig. 42 Compressor work input ratio versus turbine inlet temperature for existing machines
pag. 42
Fig. 43 Effects of spray flow rate and polytropic efficiency (denoted by “n” in the legend) on specific work ratio for pressure ratio of 30
pag. 44
Fig. 44 (a) Effect of droplet size on specific work ratio (compression rate 870 bar/s and polytropic efficiency 100%); (b) Effect of droplet size on specific work ratio (compression rate 870 bar/s and polytropic efficiency 80%)
pag. 45
Fig. 45 Stage flow coefficient relative to their design value – effect of wet compression with 5 microns droplets
pag. 45
Fig. 46 Power boost due to water sprayed in the airflow DBT= 38°C, WBT= 26°C
pag. 47
Fig. 47 Experimental power boost due to water sprayed with the airflow in an advanced industrial gas turbine inlet air duct
pag. 47
Fig. 48 Rough estimate of water flow requirements for varying gas turbine airflow rates and temperature depressions ranging from 5 °C to 15 °C (Evaporative Fogging Only)
pag. 48
Fig. 49 Special channel system for water drain pag. 50
Fig. 50 Proximity of the floor to the compressor inlet can at time cause vortex ingestion of pooled water
pag. 50
Fig. 51 A schematic showing movement of engine operating line with respect to surge line
pag. 51
Fig. 52 Level of erosion on compressor blades of GE Frame 6B gas turbine with wet compression: First stage blade tip (a), First stage blade mid-height (b), First stage blade hub (c) (Courtesy Caldwell Energy & Environmental, Inc.)
pag. 53
225
Fig. 53 Blade condition of first stage compressor blades of Alstom advanced heavy duty industrial gas turbine GT24 (Courtesy Caldwell Energy & Environmental, Inc.)
pag. 53
Fig. 54 A view of a corroded floor of intake duct pag. 54
Fig. 55 Under-frequency operation of gas turbine and its effect on output (Design speed is 50 Hz)
pag. 55
Fig. 56 Special Grounding brush utilized for overspray applications pag. 56
Fig. 57 Pressure variation along compressor showing pressure build-up in rear stages for different amount of overspray: Variable SAS system (a) and Fixed SAS system (b)
pag. 57
Fig. 58 Change in pressure build with overspray up in a compressor pag. 57
Fig. 59 View of intake from viewing window installed in a plenum pag. 59
Fig. 60 GE LM 6000 Sprint pag. 61
Fig. 61 Map defined by ALSTOM indicating temperature and Relative humidity where fogging and high fogging is permitted
pag. 61
Fig. 62 Effects of inlet evaporative fogging on a 80 MW class heavy-duty industrial gas turbine
pag. 63
Fig. 63 First application of overspray on Frame 5 gas turbines (Nolan and Twombly)
pag. 64
Fig. 64 Experimental test results on GT24/GT26 gas turbines by ALSTOM
pag. 65
Fig. 65 LHV electric efficiency versus electric power size for combined cycle systems
pag. 68
Fig. 66 LHV electric efficiency versus gas turbine exhaust gas temperature for combined cycle systems
pag. 68
Fig. 67 One pressure level combined cycle power plant lay-out pag. 69
Fig. 68 Two pressure levels combined cycle power plant lay-out pag. 70
Fig. 69 Three pressure levels with reheat combined cycle power plant lay-out
pag. 70
Fig. 70 A schematic showing nomenclatures at the Compressor inlet with fogging
pag. 77
Fig. 71 Air temperature and evaporated water mass through compressor
pag. 78
Fig. 72 Temperature-Entropy diagram for a compression process in presence of overspray fogging
pag. 79
Fig. 73 Temperature –Entropy diagram for overspray compression process considering different droplet diameters
pag. 79
Fig. 74 % CCPP Power Output change @ Generator (Reference case - HDC)
pag. 83
Fig. 75 % GT Power Output change @ Generator (Reference case - HDC)
pag. 83
Fig. 76 % ST Power Output change @ Generator (Reference case - HDC)
pag. 84
226
Fig. 77 % GT Mass Flow Rate change (Reference case - HDC) pag. 84
Fig. 78 TOT [°C] change (Reference case - HDC) pag. 85
Fig. 79 % points change of LHV CCPP Efficiency (Reference case - HDC)
pag. 85
Fig. 80 Injected water consumption per unit percent Net CCPP output change [(liters/hr)/% power boost] (Reference case - HDC)
pag. 86
Fig. 81 Incremental efficiency for CCPP due to fogging (Reference case - HDC)
pag. 86
Fig. 82 Two pressure levels combined cycle power plant (owned by Enipower and placed in Ravenna) lay out
pag. 88
Fig. 83 Calculation sheet for power plant thermodynamic calculation
pag. 89
Fig. 84 GE MS9001 non-dimensional performances maps pag. 90
Fig. 85 Steam turbine control logic pag. 91
Fig. 86 Ambient air temperature in Ravenna since 01/01/2002 to 12/17/2002
pag. 92
Fig. 87 Sea water temperature as function of ambient air temperature in Ravenna since 01/01/2002 to 12/17/2002
pag. 92
Fig. 88 Ambient relative humidity as function of ambient air temperature in Ravenna since 01/01/2002 to 12/17/2002
pag. 93
Fig. 89 CC power output change as a function of ambient air temperature
pag. 94
Fig. 90 GT power output change as a function of ambient air temperature
pag. 94
Fig. 91 ST power output change as a function of ambient air temperature
pag. 95
Fig. 92 CC efficiency change as a function of ambient air temperature
pag. 95
Fig. 93 Fuel flow rate change as a function of ambient air temperature
pag. 95
Fig. 94 GT efficiency change as a function of ambient air temperature
pag. 96
Fig. 95 Fogging water consumption as a function of ambient air temperature
pag. 96
Fig. 96 Electric power demand peaks during summer and winter since 2000 to 2004
pag. 101
Fig. 97 Trend of available electric power and of the electric power demand for Italy (only continent)
pag. 102
Fig. 98 Italian set of combined cycle units: electric efficiency versus produced power output (Reference: ISO conditions)
pag. 103
Fig. 99 Minimum, maximum and weighted average value of PT pag. 104
Fig. 100 Minimum, maximum and weighted average value of PRH pag. 104
227
Fig. 101 Electric power output (Pel), LHV power introduced with fuel (Pi), average electric efficiency (�el), water consumption (FWC) with fogging (fog) and without it (w/o fog)
pag. 106
Fig. 102 Minimum, maximum and weighted average value of FSE pag. 107
Fig. 103 Minimum, maximum and weighted average value of CD pag. 107
Fig. 104 GE Fr7EA simulation model pag. 109
Fig. 105 Default gas/air turbine performance map β* = Fβ(CS*, FF*) pag. 110
Fig. 106 Generalized stage characteristics ψp*=Fψ(φ*, SF) pag. 115
Fig. 107 Generalized stage efficiency curve ηS*=Fη(CFS*, CSS*) pag. 116
Fig. 108 1st stage performance map pag. 118
Fig. 109 2nd stage performance map pag. 118
Fig. 110 3rd stage performance map pag. 118
Fig. 111 4thstage performance map pag. 118
Fig. 112 5th stage performance map pag. 118
Fig. 113 6th stage performance map pag. 118
Fig. 114 7th stage performance map pag. 119
Fig. 115 8th stage performance map pag. 119
Fig. 116 9th stage performance map pag. 119
Fig. 117 10th stage performance map pag. 119
Fig. 118 11th stage performance map pag. 119
Fig. 119 12th stage performance map pag. 119
Fig. 120 13th stage performance map pag. 120
Fig. 121 14th stage performance map pag. 120
Fig. 122 15th stage performance map pag. 120
Fig. 123 16th stage performance map pag. 120
Fig. 124 17th stage performance map pag. 120
Fig. 125 Temperature profile along the compressor in ISO and HOT cases
pag. 121
Fig. 126 Ratio between the pressure along the compressor and the inlet value (1.013 bar)
pag. 121
Fig. 127 Gas turbine power output change as function of injected water to air mass flow rate ratio and injection point
pag. 123
Fig. 128 (a) First stage performance maps and operating points with water injection (Reference: ISO case); (b) Seventeenth stage performance maps and operating points with water injection (Reference: ISO case); (c) First stage performance maps and operating points with water injection (Reference: HOT case); (d) Seventeenth stage performance maps and operating points with water injection (Reference: HOT case)
pag. 125
228
Fig. 129 (a) Normalized flow coefficient trend at the inlet of each stage as function of injection point and water amount (ISO case); (b) Normalized stage pressure ratio as function of injection point and water amount (ISO case)
pag. 126
Fig. 130 (a) Normalized flow coefficient trend at the inlet of each stage as function of injection point and water amount (HOT case); (b) – Normalized stage pressure ratio as function of injection point and water amount (HOT case)
pag. 126
Fig. 131 Gas turbine heat rate change as function of injected water to air mass flow rate ratio and injection point
pag. 127
Fig. 132 Compressor pressure ratio as function of injected water to air mass flow rate ratio and injection point
pag. 128
Fig. 133 Air mass flow rate at compressor inlet as function of injected water to air mass flow rate ratio and injection point
pag. 128
Fig. 134 Temperature profile along the compressor in ISO case and in ISO case with 1.20 kg/s of water injected @ WIP1
pag. 129
Fig. 135 Temperature profile along the compressor in ISO case and ISO case with 1.35 kg/s of water injected @ WIP2
pag. 129
Fig. 136 Temperature profile along the compressor in ISO case and ISO case with 1.65 kg/s of water injected @ WIP3
pag. 129
Fig. 137 Temperature profile along the compressor in ISO case and ISO case with 1. 97 kg/s of water injected @ WIP4
pag. 129
Fig. 138 Temperature profile along the compressor in ISO case and ISO case with 2.30 kg/s of water injected @ WIP5
pag. 130
Fig. 139 Temperature profile along the compressor in ISO case and ISO case with 2.60 kg/s of water injected @ WIP6
pag. 130
Fig. 140 Temperature profile along the compressor in ISO case and ISO case with 2.80 kg/s of water injected @ WIP7
pag. 130
Fig. 141 Temperature profile along the compressor in ISO case and ISO case with 3.00 kg/s of water injected @ WIP8
pag. 130
Fig. 142 Temperature profile along the compressor in HOT case and HOT case with 5.00 kg/s of water injected @ WIP1
pag. 131
Fig. 143 Temperature profile along the compressor in HOT case and HOT case with 4.70 kg/s of water injected @ WIP2
pag. 131
Fig. 144 Temperature profile along the compressor in HOT case and HOT case with 4.50 kg/s of water injected @ WIP3
pag. 131
Fig. 145 Temperature profile along the compressor in HOT case and HOT case with 4.10 kg/s of water injected @ WIP4
pag. 131
Fig. 146 Temperature profile along the compressor in HOT case and HOT case with 3.90 kg/s of water injected @ WIP5
pag. 132
Fig. 147 Temperature profile along the compressor in HOT case and HOT case with 3.80 kg/s of water injected @ WIP6
pag. 132
229
Fig. 148 Temperature profile along the compressor in HOT case and HOT case with 3.70 kg/s of water injected @ WIP7
pag. 132
Fig. 149 Temperature profile along the compressor in HOT case and HOT case with 3.60 kg/s of water injected @ WIP8
pag. 132
Fig. 150 Pressure ratio profile along the compressor in ISO case and ISO case with 1.20 kg/s of water injected @ WIP1
pag. 133
Fig. 151 Pressure ratio profile along the compressor in ISO case and ISO case with 1.35 kg/s of water injected @ WIP2
pag. 133
Fig. 152 Pressure ratio profile along the compressor in ISO case and ISO case with 1.65 kg/s of water injected @ WIP3
pag. 133
Fig. 153 Pressure ratio profile along the compressor in ISO case and ISO case with 1.97 kg/s of water injected @ WIP4
pag. 133
Fig. 154 Pressure ratio profile along the compressor in ISO case and ISO case with 2.30 kg/s of water injected @ WIP5
pag. 134
Fig. 155 Pressure ratio profile along the compressor in ISO case and ISO case with 2.60 kg/s of water injected @ WIP6
pag. 134
Fig. 156 Pressure ratio profile along the compressor in ISO case and ISO case with 2.80 kg/s of water injected @ WIP7
pag. 134
Fig. 157 Pressure ratio profile along the compressor in ISO case and ISO case with 3.00 kg/s of water injected @ WIP8
pag. 134
Fig. 158 Pressure ratio profile along the compressor in HOT case and HOT case with 5.00 kg/s of water injected @ WIP1
pag. 135
Fig. 159 Pressure ratio profile along the compressor in HOT case and HOT case with 4.70 kg/s of water injected @ WIP2
pag. 135
Fig. 160 Pressure ratio profile along the compressor in HOT case and HOT case with 4.50 kg/s of water injected @ WIP3
pag. 135
Fig. 161 Pressure ratio profile along the compressor in HOT case and HOT case with 4.10 kg/s of water injected @ WIP4
pag. 135
Fig. 162 Pressure ratio profile along the compressor in HOT case and HOT case with 3.90 kg/s of water injected @ WIP5
pag. 136
Fig. 163 Pressure ratio profile along the compressor in HOT case and HOT case with 3.80 kg/s of water injected @ WIP6
pag. 136
Fig. 164 Pressure ratio profile along the compressor in HOT case and HOT case with 3.70 kg/s of water injected @ WIP7
pag. 136
Fig. 165 Pressure ratio profile along the compressor in HOT case and HOT case with 3.60 kg/s of water injected @ WIP8
pag. 136
Fig. 166 Gas turbine outlet temperature as function of injected water to air mass flow rate ratio and injection point
pag. 137
Fig. 167 Compressor specific work change as function of injected water to air mass flow rate ratio and injection point
pag. 138
Fig. 168 Turbine specific work change as function of injected water to air mass flow rate ratio and injection point
pag. 138
230
Fig. 169 Evaporation points as function of injection point, water to air mass flow rate ratio and climate conditions (ISO and HOT cases)
pag. 139
Fig. 170 Power output change as function of injected water mass flow rate and temperature for an injection upstream from the first stage
pag. 140
Fig. 171 Heat rate change as function of injected water mass flow rate and temperature for an injection upstream from the first stage
pag. 141
Fig. 172 Evaporation of a film of water in air pag. 143
Fig. 173 temperature and mass concentration distribution in case of evaporation of a film of water in air
pag. 144
Fig. 174 scheme of a water droplet exchanging heat and mass with surrounding air
pag. 145
Fig. 175 Pressure distribution as function of compressor axial length pag. 154
Fig. 176 Wet stage compression calculation: cooling effect pag. 158
Fig. 177 Wet and dry stage compression calculation comparison pag. 159
Fig. 178 Dry stage compression calculation pag. 161
Fig. 179 Wet stage compression calculation: iterative procedure pag. 161
Fig. 180 Evaluation of the medium pressure and temperature through the compressor stage
pag. 165
Fig. 181 Wet compression calculation routine flow chart pag. 169
Fig. 182 Main program (MAIN) flow chart pag. 171
Fig. 183 Compressor module (COMP) flow chart pag. 178
Fig. 184 Combustion chamber module (CC) flow chart pag. 179
Fig. 185 Turbine module (TURB) flow chart pag. 180
Fig. 186 Turbine cooling scheme pag. 181
Fig. 187 Turbine performance maps pag. 181
Fig. 188 Evaporated water to injected water ratio through the compressor for a Frame 9 E gas turbine
pag. 182
Fig. 189 Wet compression model tuning (case XIF=0.16, XTI=0.80, XTL=0.00, XW=0.50)
pag. 184
Fig. 190 Wet compression model tuning (case XIF=0.16, XTI=0.60, XTL=0.00, XW=0.50)
pag. 184
Fig. 191 Wet compression model tuning (case XIF=0.16, XTI=0.40, XTL=0.00, XW=0.50)
pag. 185
Fig. 192 Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.00, XW=0.50)
pag. 185
Fig. 193 Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.20, XW=0.50)
pag. 186
Fig. 194 Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.40, XW=0.50)
pag. 186
231
Fig. 195 Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.00, XW=0.00)
pag. 187
Fig. 196 Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.00, XW=0.50)
pag. 187
Fig. 197 Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=0.00, XW=1.00)
pag. 188
Fig. 198 Wet compression model tuning (case XIF=0.16, XTI=0.50, XTL=f(d), XW=0.00)
pag. 189
Fig. 199 Mean, tip and root radius of the 17 compressor stages of the GE Frame 7 EA
pag. 191
Fig. 200 Frame 7 EA Gas turbine performance map pag. 192
Fig. 201 Gas turbine power output of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 193
Fig. 202 Heat Rate of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 194
Fig. 203 Air mass flow rate of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 194
Fig. 204 Pressure ratio of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 195
Fig. 205 Compressor discharge temperature of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 195
Fig. 206 Turbine outlet temperature of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 196
Fig. 207 Compressor power of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 196
Fig. 208 Turbine power of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 197
Fig. 209 Fuel mass flow rate of the real Frame 7 EA (continuous lines) and evaluated by the code (square markers) as function of ambient temperature and load
pag. 197
Fig. 210 Power output change as function of injected water to air ratio and injection location
pag. 199
Fig. 211 Heat rate change as function of injected water to air ratio and injection location
pag. 200
Fig. 212 Air mass flow rate at compressor inlet change as function of injected water to air ratio and injection location
pag. 201
232
Fig. 213 Compressor pressure ratio change as function of injected water to air ratio and injection location
pag. 201
Fig. 214 Liquid water through the compressor to injected water ratio as function of injection location for a total mass of injected water equal to 5 kg/s
pag. 202
Fig. 215 Power output change as function of injected water to air ratio and injected droplet diameters for an injection in the compressor inlet duct
pag. 203
Fig. 216 Heat rate change as function of injected water to air ratio and injected droplet diameters for an injection in the compressor inlet duct
pag. 203
Fig. 217 Air mass flow rate at compressor inlet change as function of injected water to air ratio and injected droplet diameters for an injection in the compressor inlet duct
pag. 204
Fig. 218 Compressor pressure ratio change as function of injected water to air ratio and injected droplet diameters for an injection in the compressor inlet duct
pag. 204
Fig. 219 Liquid water through the compressor to injected water ratio as function of injected water droplet diameter for a total mass of injected water equal to 4 kg/s
pag. 205
Fig. 220 Gas turbine power output of the real Frame 7 EA (continuous lines), evaluated by the IN.FO G.T E (square markers) and by a commercial software (circle markers) as function of ambient temperature and load
pag. 206
Fig. 221 Heat rate of the real Frame 7 EA (continuous lines), evaluated by the IN.FO G.T E (square markers) and by a commercial software (circle markers) as function of ambient temperature and load
pag. 206
Fig. 222 Air mass flow rate at compressor inlet of the real Frame 7 EA (continuous lines), evaluated by the IN.FO G.T E (square markers) and by a commercial software (circle markers) as function of ambient temperature and load
pag. 207
Fig. 223 Pressure ratio of the real Frame 7 EA (continuous lines), evaluated by the IN.FO G.T E (square markers) and by a commercial software (circle markers) as function of ambient temperature and load
pag. 207
Fig. 224 Power output change as function of injected water to air ratio evaluated by a commercial software and by IN.FO. G.T. E.
pag. 208
Fig. 225 Heat rate change as function of injected water to air ratio evaluated by a commercial software and by IN.FO. G.T. E.
pag. 209
Fig. 226 Air mass flow rate change as function of injected water to air ratio evaluated by a commercial software and by IN.FO. G.T. E.
pag. 209
Fig. 227 Compressor pressure ratio change as function of injected water to air ratio evaluated by a commercial software and by IN.FO. G.T. E.
pag. 210
233
Fig. 228 Liquid water through the compressor to injected water ratio as function of injected water droplet diameter for a total mass of injected water equal to 5 kg/s evaluated by a commercial software and by IN.FO. G.T. E.
pag. 210
234
8
List of tables Tab. 1 Inlet air cooling technologies main characteristics pag. 22
Tab. 2 ECDH and Number of Hours (MWBT is Above 7.2 ºC) pag. 25
Tab. 3 General Water Quality Requirements pag. 49
Tab. 4 List of gas turbines with third-party furnished fog systems pag. 63
Tab. 5 Field performance results on GE Frame 6B gas turbine with overspray
pag. 65
Tab. 6 Performance for selected gas turbines pag. 71
Tab. 7 Bottoming cycle design parameters pag. 71
Tab. 8 Overall performance of selected combined cycle power plants @ ISO conditions
pag. 72
Tab. 9 CCPPs performances compared between ISO and hot day case (HDC) conditions
pag. 73
Tab. 10 GE MS9001 performances @ ISO conditions pag. 90
Tab. 11 Link between sea-water temperature and air ambient temperature
pag. 93
Tab. 12 CC power output [MW] pag. 97
Tab. 13 GT power output [MW] pag. 97
Tab. 14 ST power output [MW] pag. 98
235
Tab. 15 CC efficiency [%] pag. 98
Tab. 16 Fuel flow rate [Sm3/h] pag. 99
Tab. 17 GT efficiency [%] pag. 99
Tab. 18 Fogging water consumption [ton/h] pag. 100
Tab. 19 Comparison of the main performance and thermodynamic parameters of Fr7EA Thermoflex model at ISO and HOT conditions for constant TIT, IGV fully opened in absence of inlet and exhaust pressure losses
pag. 111
Tab. 20 Calculated design stage pressure ratio of axial compressor of GE Fr7EA gas turbine
pag. 114
Tab. 21 Diffusion coefficient formulas pag. 151
Tab. 22 Atomic Diffusion Volumes pag. 152
Tab. 23 GE Frame 9E main performance @ ISO conditions pag. 182
Tab. 24 Compressor stages performance maps definition coefficients
pag. 190
Tab. 25 Mean, tip and root radiuses of the 17 compressor stages of the GE Frame 7 EA
pag. 191
Tab. 26 Performance parameters of “reference” machine and evaluated by the code as function of ambient temperature and load
pag. 198
236