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DESIGN, PART LOAD AND TRANSIENT OPERATION OF
COMBINED CYCLE PLANTS WITH WATER FLASHING
PJ. Dechamps
Department of Nuclear Engineering and Power Plants
University of Liege
Rue E. Solvay 21 - Bat C3 - 4000 Liege
BELGIUM
ABSTRACT
NOMENCLATURE
The last decade has seen remarkable improvement in gas
turbine based power generation technologies, with the increasing
use of natural gas-fuelled combined cycle units in various regions
of the world. The struggle for efficiency has produced highly
complex combined cycle schemes based on heat recovery steam
generators with multiple pressure levels and possibly reheat. As
ever, the evolution of these schemes is the result of a technicoeconomic balance between the improvement in performance and
the increased costs resulting from a more complex system.
GT
HRSG
NTU
ST
TET
VIGV
IF
1P
2P
This paper looks from the thermodynamic point of view at
some simplified combined cycle schemes based on the concept of
water flashing. In such systems, high pressure saturated water is
taken off the high pressure drum and flashed into a tank. The
vapour phase is expanded as low pressure saturated steam or
returned to the heat recovery steam generator for superheating,
whilst the liquid phase is recirculated through the economizer.
With only one drum and three or four heat exchangers in the
boiler as in single pressure level systems, the plant might have a
performance similar to that of a more complex dual pressure level
system. Various configurations with flash tanks are studied based
on commercially available 150 MW-class E-technology gas
turbines and compared with classical multiple pressure level
combined cycles. Reheat units are covered, both with flash tanks
and as genuine combined cycles for comparison purposes. The
design implications for the heat recovery steam generator in terms
of heat transfer surfaces are emphasized.
Off-design considerations are also covered for the flash based
schemes, as well as transient performances of these schemes,
because the simplicity of the flash systems compared to normal
combined cycles significantly affects the dynamic behaviour of
the plant.
:
:
:
:
:
:
:
:
:
Gas Turbine
Heat Recovery Steam Generator
Number of Transfer Units
Steam Turbine
Turbine Entry Temperature
Variable Inlet Guide Vanes
Single pressure combined cycle with flash
Single pressure combined cycle
Dual pressure combined cycle
INTRODUCTION
In the race for fuel efficiency in power generation, the
combined cycle schemes have been gradually transformed from
single pressure level systems to dual pressure level system, and
possibly to triple pressure systems with reheat (Bolland, 1990)
(Mathieu, 1990) and (Dechamps et al., 1993). The main
drawback of these efficiency improvements is of course the
increased capital cost of the plant, mainly ascribed to the
increased heat transfer surface within the HRSG and to the more
complex multiple-section steam turbine. However, the complexity
of the system is in itself a drawback although perhaps less easily
quantifiable. This complexity turns into decreased availability,
increased operational complexity, etc. The idea behind this study
tends to reduce the complexity of these systems, while
maintaining most of their efficiency advantage.
If one compares the classical single pressure level combined
cycle with the state-of-the-art dual pressure system, the efficiency
advantage of this latter is obvious. Most of this advantage comes
from the improved HRSG effectiveness, which is defined as the
ratio of the heat extracted in this HRSG, over the maximum heat
one could extract from the gas bringing them to room
temperature. In other words, and in practical terms, the better
efficiency of the dual pressure system is mainly due to the lower
stack temperature that this system can achieve.
Presented at the International Gas Turbine and Aeroengine Congress and Exposition
Houston, Texas - June 5-8,1995
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The question is then to know why the simple single pressure
system cannot achieve comparable stack temperatures. The
answer lies into the consideration of the heat exchange : once the
pressure of the system is chosen, the steam-water mass flow is
given by the energy balance above the pinch. This mass flow,
and nothing more, is available to further cool the gas, below the
pinch. For a fixed condensate temperature, this immediately gives
the stack temperature using the energy balance below the pinch.
Extracting some of the water phase from the steam drum and
flashing it down to a lower pressure allows to further re-compress
the liquid phase to recirculate it back to the economizer. This
scheme is illustrated by Fig. 1. The steam phase resulting from
the flash can be expanded into the low pressure part of the steam
turbine. It could possibly be returned to the HRSG for
superheating in order to prevent the transportation of droplets into
the steam turbine. Energetically, however, this makes very little
difference.
t (°C)
600
400to
the
u erheater
300
and the HP
1-tx
turbine
1 +t(1-Y)
200
:f
to the LP
tt^^-x).1 .................... ....' turbine
1
100
0
0
1
2
3
4
5
6
7
8
9 10
s (kJ/kgK)
Fig. 2 - Flash in the (t, ․) diagram
The flash system is represented in the T,s diagram on Figure
2, where all the mass flows are for 1 kg/s at the condenser.
Design Characteristics
compressor pressure ratio : 10.71
compressor inlet flow : 498.9 kg/s
VIGV for off-design operation (flow reduced to 80% when
feed
water
economizer
All the mass flows are for 1 kg/s in the condenser
500 f=flash mass flow rate
fully closed)
TET : 1110°C
Iso Conditions
Pump
evaporator
gas path
blow
down
fuel : pure methane
net output : 153.6 MW
net efficiency : 33.5 %
exhaust flow : 506.5 kg/s
exhaust temperature : 545°C
flash
tank
Site Conditions
inlet pressure loss : 10 mbar
exhaust back pressure : 25 mbar
conditions :15°C, 1 atm., 60% rel. humidity
fuel : natural gas, 47.5 MJ/kg
net output : 145.8 MW
net efficiency : 32.48 %
exhaust flow : 506.8 kg/s
exhaust temperature : 552°C
superheater
LP
HP
superheated saturated
steam
steam
Table I - Gas Turbine Characteristics
The combined cycle operation takes a 10 mbar inlet pressure
drop, a 25 mbar back-pressure, a 47.5 MJ/kg natural gas fuel
into account and assumes that the control system adjusts the fuel
flow so that the TET is kept at its design value.
Fig. 1 - Schematic diagram of the flash system
For off-design operation, this gas turbine is supposed to be
VIGV-controlled in the first place, to be followed by a TET
reduction when the VIGV are fully closed. This is known to give
the smallest combined cycle efficiency penalty when reducing the
output (Dechamps et al., 1994).
THE GAS TURBINES
The genuine combined cycles used for comparison purposes,
as well as the flash schemes are based on commercially available
150 MW-class E-technology gas turbines. Table I gathers the
most important design parameters of this machine, together with
its iso-condition and site-condition performances.
2
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GENUINE COMBINED CYCLES
Normal combined cycles, with one pressure level and two
pressure levels were studied, based on the previously described
gas turbine, to provide the yardsticks for the comparison and the
appraisal of the flash-based schemes.
Their main design assumptions, together with the simulation
results, are grouped in table 11 and III respectively.
The cycle parameters (pressures and temperatures) were
optimized using the GTCALC and CCCALC codes described in
Dechamps, 1993.
It is obvious from the previous results (table III) that the 2P
system efficiency improvement is due to the progress in HRSG
effectiveness (i.e. lower stack temperature).
General Data
condenser pressure : 50 mbar
steam turbine isentropic efficiency : 85 %
idem, degradation with moisture content : 1 %/%
minimum stack temperature: 60°C
HRSG approach temperature differences : 1°C (economizer outlet undercooling)
HRSG pinch point temperature differences : 10°C
superheater min. temperature difference [gas-steam] : 35°C
feedwater pump efficiency : 75 %
minimum steam outlet quality from the steam turbine : 0.88
Single Pressure System
drum pressure : 45 bar
superheat temperature : 515°C
Dual Pressure System :
HP drum pressure : 60 bar
HP superheat temperature : 515°C
LP drum pressure : 6 bar
LP superheat temperature: 210°C
Table II - Genuine Combined Cycle Characteristics
GT power (MW) / ST power (MW)
net power output (MW)
net efficiency (%)
stack temperature (°C)
HRSG effectiveness (%)
Rankine cycle efficiency (%)
HRSG NTUs (-)
1P
2P
145.8 / 76.6
222.4
49.52
153
75
34
11.29
145.8 / 84.1
229.9
51.22
90
87
32
17.8
Table III - Genuine Combined Cycle Performances
OPTIMIZATION OF THE SINGLE PRESSURE SYSTEM
WITH FLASH
Optimization of the IF system has been conducted using the
GTCALC/CCCALC codes (Dechamps, 1993) which were
developed at the Department and allow the multi-variable
optimization of any combined cycle configuration. Results were
crossed-checked with the commercially available GATE/CYCLE
software (Erbes, 1988).
Compared with the classical single pressure system, the IF
system has two more parameters to be determined : the flash
pressure and the flash mass flow.
recirculation when the economizer temperature difference is
constant from inlet to outlet. We will therefore define the
recirculation ratio as the recirculated mass flow over this
maximum value.
A recirculation ratio of 0 gives the classical single pressure
system, whilst a recirculation ratio of 1 is the maximum that can
be done to increase the HRSG effectiveness.
We will first concentrate on choosing the right parameters for
the system with a recirculation ratio of 1.
This latter is linked to the mass flow which is recirculated
back to the economizer inlet. There obviously exists a maximum
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Fig. 3 presents the influence of the drum pressure, for both
the normal system and the flash system. Both curves exhibit the
same behaviour, with an optimum pressure around 60 bars for the
1F, higher for the flash system than for the conventional one at
45 bars. This optimum for the classical system is the result of a
compromise between the HRSG effectiveness and the Rankine
cycle efficiency which show opposite trends when increasing the
drum pressure. It is then normal that the flash system, with its
higher HRSG effectiveness has its optimum drum pressure shifted
towards higher values.
Fig. 4 shows the influence of the flash pressure on the cycle
efficiency and the stack temperature for a fixed value of the drum
pressure (60 bars, the optimum value). The ideal flash pressure
from the efficiency point of view is 6 bars, which happens to be
10% of the drum pressure. This is not too surprising as a 10%
value is classical between the LP drum pressure and the HP drum
pressure in dual pressure systems.
In the whole range, the stack temperature achieved by the
flash systems with a recirculation ratio of 1 is remarkably low,
comparable to that of multiple pressure systems.
Plant efficiency (%)
51.5
Fig. 5 shows the evolution of the HRSG NTUs with the same
parameter. Whatever the variations, the numbers are typically 3
times higher than those of the 1P system (see table III). It is
therefore unlikely that the flash system with a maximum
recirculation ratio will be an economically sound alternative. The
huge HRSG transfer surface is required by the economizer, which
has a "constant pinch" all along its tube banks within the boiler.
-With flash
(recirculation ratio = 1)
51
50.5
50
Without flash
49.5
49 1r
20
30
40
50
60
70
There is therefore an interest in varying the recirculation
ratio. Before this, table IV presents the main characteristics of the
system optimized with a recirculation ratio of 1. Most design
assumptions are kept similar to those of the genuine systems to
ease the comparison.
so
Drum pressure (bar)
Fig. 3 - Optimization of single pressure systems
Plant efficiency (%)
Total HRSG NTU5
Stack temperature (°C)
51.3
51.25
90
51.15
35
51.1
30
51.05
17
4
31
16
30
14
29
12
28 1
2
10
1s
13
75
3
4
5
6
7
8
9
10
70
11
Flash pressure (bar)
Fig.
18
95
51.2
51 1
2
LP steam mass flow (kg/s)
32
100
3
4
5
6
7
8
9
10
Flash pressure (bar)
- The influence of the flash pressure
Fig. 5 - Varying the flash pressure
General
condenser pressure : 50 mbar
steam turbine isentropic effic. : 85 %
idem, degradation with moisture : 1 %/ %
HRSG approach t° differences : 1°C
HRSG pinch point t° differences :10°C
drum pressure : 60 bar
superheat temperature : 515°C
flash pressure : 6 bar
flash mass flow : 54 kg/s
recirculation mass flow : 39.9 kg/s
LP saturated steam mass flow : 14.1 kg/s
Performances
GT power: 145.8 MW
ST power: 84.1 MW
net power output: 229.9 MW
net efficiency : 51.19 %
stack temperature : 85°C
HRSG effectiveness : 88 %
Rankine cycle efficiency : 32 %
HRSG NTUs : 29.8
steam outlet quality : 0.88
Table IV - IF System with Maximum Recirculation
4
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VARYING THE RECIRCULATION RATIO
Plant Efficiency (%)
Temperature ("C)
51.5
Fig. 6 shows the heat transfer curves for the single pressure
systems when the recirculation ratio is varied from 0 to 1. The
recirculation ratio of 1, which corresponds to the maximum
recirculation is when the economizer has a constant temperature
difference along its length.
F13
50.5
econo
cold endl
pinto
400
^11 D
^10 0
90
r ao
50
60
so
fcedwstcr
40
49.5
1-1' : without recirculation
2-7 : with recirculation ratio = 0.!
3-3' : with recirculation ratio = 1
(constant econo pinch)
D
L12 D
Temperature (C)
500
r 14 D
51
The progression of points 1-2-3 indicates how the stack
temperature and hence the HRSG effectiveness can be improved.
600
to 0
15 D
stack paaas
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
an
Recirculation Ratio
Fig. 8 - Varying the recirculation ratio
300
200
1
.
The main result of this section is that the linear evolution of
the efficiency, combined with the exponential-like evolution of
the transfer surface, might make the 0.7-0.8 range the optimum
for the recirculation ratio.
2
`1 3
100
0
i^
2• 3;
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
COMPARISON WITH GENUINE SYSTEMS
Heat transferred / maximum heat transferred
The consideration of the flash schemes only makes sense if
they are compared with the classical systems.
Fig. 6 - Heat exchange diagrams
Fig. 9 shows the situation of the 1P, 2P, and iF systems in
terms of efficiency and HRSG NTUs. The recirculation ratio is
plotted on the curves for the IF system.
Fig. 7 illustrates how the heat transfer surface depends on the
recirculation ratio. The levels marked 1P and 2P indicate what
are the heat transfer surfaces of these systems.
The 1P and 2P systems are the optimized cases described
above. It is obvious from this graph that the region of interest for
the IF system is with recirculation ratios in the vicinity of 0.60.8, when some of the performance of dual pressure systems is
achieved at the expenses of a reasonable transfer surface increase.
The maximum recirculation ratio of 1 appears non-attractive
because of its heat transfer surface requirement.
Total HRSG NTUs
30
25
20
NTU level
of the 2P system
1s
10
5
cycle efficiency (%)
NTU level
of the 1 P system
51.5
2P
^►
5
0 i 1
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
o
1F
o-e
0-7
Recirculation Ratio
50.5
0-
o-
0.
Fig. 7 - HRSG NTUs versus the recirculation ratio
4
Constant
pinch at the i
economizer
Recirculation
Ratio
0.3
50 o.
Most of the increase in transfer surface takes place above a
recirculation ratio of 0.6-0.8. For this reason it could be that this
sort of recirculation rate gives the best compromise between the
efficiency and the surface. The efficiency is on Fig. 8 and
hopefully exhibits a nearly linear behaviour with the recirculation
ratio. The temperatures of both the stack gas and the feedwater
are indicated on the left-hand scale, with their difference showing
the evolution of the temperature difference at the economizer cold
49.5
49
10
15
20
25
30
35
Total HRSG NTUs
Fig. 9 - Comparison of flash and conventional systems
end.
5
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OFF- DESIGN CONSIDERATIONS
Recirculation Ratio
Reducing the power output of combined cycle systems is
usually achieved first by closing the variable inlet guide vanes on
the gas turbine compressor. This enables to reduce the gas mass
flow, while keeping the exhaust temperature from the gas turbine
constant. On most recent machines, the exhaust temperature is
even increased when the VIGVs are closed, because the control
system maintains the TET constant.
Once the VIGVs are fully closed, the only option left is to
reduce the fuel flow, and hence the TET. This power reduction
strategy is accompanied by a more rapid drop in efficiency
(Dechamps et al., 1994)
Approach Delta T (°C)
1.25
t0
1.2
I8
1.15
16
1.1
14
1.05
12
10
0.95
0.9
0.85
0.8
0.75
J
o
40 60 80 100 120 140 160 180 200 220 240
Power Output (MW)
The off-design performance can be greatly improved by the
use of schemes with two gas turbines and two HRSGs for one
steam turbine, or by the juxtaposition of several single shaft
units. For reasons of simplicity, we will concentrate here on the
basic unit composed of one gas turbine, one HRSG and one steam
turbine. Fig. 10 shows the decrease in efficiency experienced
when reducing the output both for the classical single pressure
plant (1P) and for its counterpart flash-based system (1F), with
a recirculation ratio of 0.8.
Fig. 11 - Maintaining the approach off-design
design point approach of 1°C, the VIGV-controlled part of the
range sees a increased approach, due to the increase in gas
turbine exhaust temperature under these conditions. However,
further down the operating range, the TET control decreases the
approach until it reaches its design value of 1°C. At this point,
the recirculation ratio is deliberately increased to keep it constant.
cycle efficiency (%)
47.5
Regulation
45
42.540
37.5
TRANSIENT CONSIDERATIONS
TET
Reg
50
1 F System
Introduction
The transient behaviour of combined cycle plants is of
primary interest to the users. Most of them will have to operate
the plant on load following or on a daily load schedule.
1 P System
-
35
32.5
30 40 60 80 100 120 140 160 180 200 220 240
Power Output (MW)
Fig. 10 - Off-design performances
Clearly, the flash-based system retains its efficiency
advantage over the classical plant in the whole operating range.
In both cases, the higher part of the range, when the VIGVs are
used, exhibits a smaller efficiency drop with power than the
lower part of the range when the TET is used to control the
output. Quite interestingly, the flash-based system offers some
more flexibility when operating in off-design conditions. The
recirculation ratio can indeed be varied freely. Its design value
determines the economizer transfer surface and cold end
temperatures; but once this is done, the recirculation ratio can
further be varied in the actual plant, with a fixed economizer
surface. A recirculation is usually introduced in genuine
combined cycle plants at part load conditions to control the
economizer steaming. In the flash-based system, a slight increase
of the flash mass flow, which turns into an increase of the mass
flow recirculated to the economizer inlet provides the same
effect.
Fig. 11 shows the variation of the economizer approach (i.e.
the economizer outlet undercooling) and of the recirculation ratio
with the power output for the flash-based plant. Starting with the
A fast start-up procedure is essential to maximize the
integrated efficiency of the plant over a given load schedule. This
fast start-up depends on the boiler inertia if the steam turbine can
be loaded without restriction.
It is therefore interesting to compare the behaviour of a dual
pressure combined cycle plant with the flash system because they
have different boiler arrangements.
Dynamic Simulation
The CCDYN code (Dechamps, 1994) has been used to
simulate a cold start for the dual pressure combined cycle plant
and for the single pressure plant with a flash system and a
recirculation ratio of 0.8.
From fig. 7, it can be seen that these two systems have
approximately the same heat transfer surface requirement, but
with a basically different splitting between the economizers on
one side and the evaporators/superheaters on the other.
Fig. 12 presents the evolution with time of the reduced power
outputs for both plants during a cold start-up (cold meaning that
the boiler was at room temperature). The flash system exhibits a
more rapid loading capability in the 20-30 min. after the gas
turbine start. This is essentially because the dual pressure system
relies on the second pressure level start-up, depending on the
boiler inertia; whilst the flash and recirculation can be started as
soon as the pressure is established in the HP drum.
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HP drum level (m)
Output / design output
1.4
1
1.2
0.8
0.8
0.6
0.6
0.4
0.4
0.2 - dual pressure system (2P)
0
— dual pressure system (2P)
0.2-
......... flash system (1F)
flash system (19
0-
0
0
10
20
30
40
50
60
10
20
30
40
50
60
Time (min)
Time (min)
Fig. 14 - HP drum levels during start-up
Fig. 12 - Output J design output during cold start-up
Fig. 13 presents the evolution with time of the HP drum
pressures. They show similar trends to the previously discussed
power outputs.
ECONOMIC CONSIDERATIONS
The concept of incremental investment cost is helpful in
estimating the economic viability of power generation projects.
The incremental investment cost can be simply defined as the
amount of money to invest in order to increase the plant output
by one kW, without changing the fuel input.
HP pressure / design HP pressure
0.8
As far as the recirculation ratio is concerned, the main
parameter is the boiler heat transfer surface requirement from
which the boiler price can be estimated.
0.6
0.4
0.2
Fig. 15 shows this incremental investment cost as a function
of the recirculation ratio. On the same figure is the range of
incremental costs in use for the appraisal of power generation
projects. This range is from 1000 to 3000 5/kW. One has to
remember that this number is substantially higher than the overall
plant capital cost expressed in the same units, because it is a
incremental parameter, corresponding to a situation when the
plant output is varied without any modification in the fuel input.
— dual pressure system (2P)
flash system
(19
0
0 10 20 30 40 50 60
Time (min)
Fig. 13 - HP pressures during cold start-up
This commonly used range of incremental costs corresponds
to recirculations ratios in the 0.5 - 0.8 range, already mentioned
as the more interesting, from technical points of view.
System Stability
System stability is of primary importance for the transient
operation of power plants. The main concern is the behaviour of
the feedwater mass flow control loop which controls the
feedwater flow to maintain the drum level at its set-point.
Incremental capital cost
14
This system contributes to the drum level evolution after the
so-called swell effect resulting from the heating of the
evaporators during a cold start.
(S /kw) (thousands)
12
10
6
Fig. 14 shows the HP drum levels as a function of time. The
same control system settings have been used in both cases. The
similarity of the behaviour of the two systems demonstrates that
the important recirculation mass flow through the economizer
does not affect the feedwater mass flow control loop.
6
4
2
0
Some more severe tests have been performed involving the
sudden start or sudden stop of the flashed and recirculated flows
and have shown that the standard controls systems were able to
maintain the drum level within safe margins.
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 0
Recirculation Ratio
Fig. 15 - Incremental capital cost
7
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1P
recirculation ratio
2P
iF
-
0.6
0.8
1.0
-
45
60
60
60
60
-
6
6
6
6
gas turbine output (MW)
145.8
145.8
145.8
145.8
145.8
steam turbine output (MW)
76.6
81.3
82.7
84.0
84.1
plant output (MW)
222.4
227.1
228.5
229.8
229.9
plant efficiency (%)
49.52
50.59
50.90
51.19
51.22
HRSG total NTU
11.3
14.5
17.8
29.8
17.8
(-)
HP drum pressure (bar)
LP drum pressure (bar) or flash pressure
(-)
Table V - Final comparison of the various systems
FINAL COMPARISON
REFERENCES
Table V gathers information from all the previous tables and
shows clearly the trends between the efficiency and the heat
transfer surface requirements for all the considered systems. The
advantage of the 0.6-0.8 reciculation ratio appears; with an
efficiency close to the dual pressur system, at the expense of no
increase in transfer surface, and with a reduced complexity for
the plant.
Bolland 0., 1990, "A Comparative Evaluation of Advanced
Combined Cycle Alternatives", ASME paper 90-GT-335, Gas
Turbine and Aeroengine Congress and Exposition, Brussels.
Dechamps P.J., 1993, "GTCALC & CCCALC User's Guide
- Second Edition", University of Liege, Department of Nuclear
Engineering and Power Plants.
Dechamps P.J., Magain D. and Mathieu Ph., 1993,
"Advanced Combined Cycle Alternatives with Advanced Gas
Turbines", ASME Cogen Turbo Power '93, Bournemouth, UK IGTI Vol. 8.
Dechamps P.J., Pirard P. and Mathieu Ph., 1994, "Modelling
the Part Load Performance of Supplementary Fired Combined
Cycle Plants", ASME Cogen Turbo Power '94, Portland
(Oregon), IGTI Vol. 9.
Dechamps P.J., 1994, "Modelling the Transient Behaviour of
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CONCLUSIONS
This paper has considered simplified combined schemes
comprising flash systems with a recirculation of the water phase
back to the economizer.
Such systems were found to increase the HRSG effectiveness
in a similar way to the addition of a second pressure level,
without the structural complexity of the boiler that usually goes
with multiple pressure systems.
Consideration of the HRSG transfer surface as a function of
the recirculation ratio has lead to the conclusion that the
maximum recirculation rate is not the most sensible solution.
The consideration of off-design (especially part load) shows
that the flash and recirculation concept can be used to control the
economizer steaming at reduced loads.
The transient behaviour of flashed systems appeared attractive
compared to classical dual pressure systems, allowing a more
rapid output increase in cold starts when the steam turbine can be
loaded without restriction. The flash systems exhibited stable
operation characteristics in all the envisaged situations.
The economics of flash systems make sense in the 0.6-0.8
recirculation ratio range, when the incremental cost is in the
usual values for combined cycle power generation projects.
With these considerations, the single pressure system without
reheat, and a recirculation ratio of 0.6 to 0.8 appears as an
interesting alternative to classical dual pressure systems, or as an
upgrade to genuine single pressure solutions.
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