Vivek Asthana, Kanpur, IND - Eurotherm 2008

5th European Thermal-Sciences Conference, The Netherlands, 2008
Abstract
PERFORMANCE OF POWER PLANTS WITH HIGH
TEMPERATURE CONDITIONS AT SUB-CRITICAL
PRESSURES
Vivek Asthana 1 and P. K. Panigrahi 2
1. Panki Thermal Power Station, Kanpur, Uttar Pradesh, India
2. Indian Institute of Technology, Kanpur, Uttar Pradesh, India
Technology for coal fired power generation is switching over to supercritical (SC) and ultra
supercritical (USC) plants which operate with steam on higher temperature and above critical
pressure to produce power output at higher thermal efficiency. Due to involvement of high heat
resistant material, manufacturing costs of the components for these plants are higher. High pressure
designing of these components further increases the cost. An analysis has been made in this study
to explore the possibilities of operating power plants with steam at same temperatures as used in SC
and USC plants but at sub-critical pressures to produce the output comparable with supercritical
plants on comparable thermal efficiencies as that of supercritical plants. Analysis shows that with
proper selection of reheat pressure and extraction steam parameters with in technically feasible
limits of design of turbine and steam piping, power plants operating on high temperature conditions
at sub-critical pressures can be designed to produce output comparable with supercritical plants.
Thermal efficiency of these plants will be comparable with supercritical plants. How ever these
plants shall be more economical and involve low capital as compared with supercritical and ultra
supercritical plants due to relatively low pressure designing of plant components. Moreover the
components of these plants can be manufactured with presently developed high heat resistant
materials along with better factor of safety, which will reduce maintenance cost of the plant due to
lesser numbers of failures and damages of components.
1. Introduction
Power generation is heavily dependent on coal fired plants through out the world. These plants
contribute for about 50 percent of global power generation. This dependence of power generation
on coal is likely to be continued in near future also, as the coal is cheap, easily available and evenly
distributed through out the world. Conventional coal fired power generating plants operate on subcritical
technology with low thermal efficiency, which results in huge environmental pollution and
rapid depletion of limited reserves of coal. Due to these reasons super-critical and ultra supercritical
generating plants have attracted the attention of power utilities. These plants have high
generating capacity on higher thermal efficiency. Presently a large number on supercritical plants
are in commercial operation in many of the countries. However due to inevitable requirement of an
efficient power generation, the coal fired power generation is switching over to super-critical and
ultra super-critical technology from its existing phase of conventional sub-critical technology
through out the world.
Yamamoto et al. (2003) reported the application of heat resistant material of high creep rupture
strength and high oxidation resistance up to 650 °C, which have already been developed for Boiler
and turbines of ultra super critical power plants. Viswanathan (2001) discussed the materials for
ultra supercritical (USC) plants to withstand up to operating steam conditions at 760 °C
temperature and 35 MPa pressure, which are under development. A large number of super-critical
plants with steam parameters in range of 240 bar/600 °C /600 °C are confirmed to be in commercial
operation in many countries by Shigeru (2001). Ultra super critical plants with parameters in the

5th European Thermal-Sciences Conference, The Netherlands, 2008
range of 350 bar/700 °C /720 °C or higher are under development. As reported by Alto et al.
(2001), these plants are proposed to operate on net thermal efficiency of 48-52%. However due to
higher present cost of heat resistant materials, as discussed by Booras (2003) and designing of
components to operate on higher supercritical pressures, a large capital is involved in these plants.
Steam at a given temperature on sub-critical pressures consists of more specific enthalpy as
compared with steam at super-critical/ultra super-critical pressures. Possibilities of designing coal
fired power plants operating at higher temperatures of the order of SC & USC plants but on subcritical
pressures to provide same output as that of SC & USC plants on thermal efficiency
comparable with SC & USC plants have been explored here. These plants will be more cost
effective due to involvement of components designed on relatively low pressures.
2. Design features
Conventional sub-critical coal fired power generating units operate on Rankine cycle. These units
are invariably provided with one reheat of steam up to main steam temperature at some suitable
intermediate pressure and regenerative heating of condensate through five to seven numbers of
regenerative feed water heaters. Three to five numbers of these heaters (called as low pressure
heaters) are provided to heat the condensate before its discharge to the deaerator. Subsequent
heating of condensate takes place in deaerator which is an open type heater and facilitating removal
of dissolved gases from condensate. Two more feed heaters (called as high pressure heaters) are
usually provided to heat the feed water in its way from discharge of boiler feed pump to the inlet of
economizer. The same design configuration has universally been adopted for super-critical and
ultra super-critical units also. However in some of the units double reheat arrangement of steam is
provided. Schematic of a coal fired power generating unit is shown in fig. 1, which has been
analyzed here under different sub-critical and supercritical operating conditions.
Fuel
Input
25
Boiler
24
23
6
14&15
22
13 12
K J
1
H P
Turbine
14
13
2
I
4
3 5
12
I P
Turbine
G
11
F
10
9
E
L P Turbine
8 7 15 9
19
D
C
B
A
20
21 26
Generator
6
H
Condenser
16 17 18
Figure 1: Schematic of system analyzed under different operating conditions.
Condensate is extracted from the condenser by means of condensate extraction pump (A) and
passes through ejector (B) for heating by means of steam used in ejector for vacuum rising in
condenser. Subsequently this condensate is heated in low pressure heaters (C to G) and finally
stored in a feed water tank provided below the deaerator (I). The feed water is pumped to the boiler
through boiler feed pump (H), which is heated on its way to boiler in high pressure heaters (J & K).

5th European Thermal-Sciences Conference, The Netherlands, 2008
Drips collected in low pressure heaters and ejector (16, 17 & 18) is added in main condensate line
(at 19). Water and steam losses from the cycle have been neglected. Thus no make-up water is
assumed to be added in cycle for analysis purpose.
3. Analysis
To evaluate the performance of unit under different operating conditions, thermal efficiency of
turbine has been worked out in accordance with following universally adopted standard relation.
Energy output from Turbine (E21
)
ηTurbine
=
------------------------(1)
(E + E ) − E + E + E + E + E + E + E − E )
1
3
(
2 7 8 9 10 11 12 20
A portion of energy associated with turbine exhaust (6) is recovered back along with condensate
(20). Hence it is also deleted along with energy associated with other extractions from total energy
input (E1 + E3
) to find net energy input to turbine.
Thermal efficiency of boiler based on gross calorific value of coal, in accordance with following
relation (equation 2), is assumed as 86% for analysis purpose. For most of the sub-critical units,
which are in operation through out the world, boilers operate on approximate this value (86%) of
thermal efficiency.
( E24 − E23)
+ ( E6
− E22
)
η
Boier
=
----------------(2)
Fuel Energy ( associated with GCV of coal)
input to Boiler (E )
To evaluate thermal efficiency of unit in accordance with following relation (equation 3), generator
efficiency is assumed as 98% (for most of the sub-critical units, which are in operation through out
the world, generator efficiency is the same as 98%). Heat and friction losses from turbine is
assumed as 2.5%. In coal fired power plants, boiler & turbine are connected through steam pipes,
condensate water and feed water system. There is a net transfer of energy from boiler to turbine
through above pipes & systems. Based on the data collected from a running 110 MW unit in India,
as discussed in sub-section 3.3 of the paper, it is found that about 2% of the net energy being
transferred from boiler to turbine is lost in above pipes & systems due to heat dissipation to
surroundings. For the plants running at high temperatures, there is no change in designing of these
pipes & systems except the change of material for two steam pipes only. Thus an energy loss of
2% is assumed here in these pipes & systems to evaluate thermal efficiency of the unit.
Power output at generator terminals ( E26
)
ηUnit
=
------------------(3)
Fuel Energy ( associated with GCV of coal)
input to Boiler (E )
Considering above assumptions, thermal efficiency of unit is evaluated directly with the help of
turbine and boiler efficiencies.
3.1 Cycle analysis
Expansion of steam with respect to different parameters on T-S plane is shown in fig.2. Process 1-2
shows the expansion of steam in high pressure cylinder of turbine (H P turbine) in conventional
sub-critical units. Switching over to super-critical parameters, expansion 3-4 provides more output
on higher thermal efficiency. Instead of switching over to super-critical pressures, expansion 5-6 on
same sub-critical pressure will provide enhanced output (comparing with 3-4) on relatively higher
thermal efficiency. However due to higher specific volume at turbine inlet (point 5), size of turbine
shall be larger comparing with sub-critical turbine (expansion 1-2). This problem is compensated
25
25

5th European Thermal-Sciences Conference, The Netherlands, 2008
by adopting turbine inlet conditions at slightly higher pressures (p b instead of p a ) and selecting
expansion 7-8 instead of 5-6. Specific volume of steam at point 7 is the same as that of at point 1,
which provides same size of steam piping and feasible design of turbine, however with more stages
and larger dimensions at end stages comparing with conventional sub-critical turbine (expansion 1-
2). This expansion 7-8 provides more output comparing with output of 3-4 in super-critical units on
almost same thermal efficiency. Work output obtained from H P turbine and I P & L P turbine
solely depends upon the intermediate pressure selected for reheating of steam. This pressure is
governed by various other design constraints also including specific volume of steam and quality of
steam at the end of expansion in L P turbine at condenser pressure.
p d (300 bar)
p c (246 bar)
T
p b (142 bar)
T b
600ºC
3
7
5
T a
535ºC
1
4
2
8
6
S
Figure 2: Expansion of steam in turbine under sub-critical and super-critical conditions.
3.2 Choice of intermediate pressure for reheating of steam
For conventional sub-critical units an intermediate pressure in the range of 25-45 bars has
universally been adopted for reheating of steam. For most of the super-critical units, which are in
operation or under development, an intermediate pressure in the range of 60-80 bars has been
adopted for reheating of steam (AD 700 project). Power output and thermal efficiency of the unit
depends on the choice of intermediate pressure, which is decided by various other design
constraints also apart from the work output based on enthalpy drop.
Output vs. IP
Efficiency vs. IP
Output in MW
680.00
660.00
640.00
620.00
600.00
580.00
560.00
540.00
20 30 40 50 60 70 80 90 100
Intermediate Pressure (IP) in bar
Efficiency in %
47.2
47.1
47
46.9
46.8
46.7
46.6
20 30 40 50 60 70 80 90 100
Intermediate Pressure (IP) in bar
Figure 3: Variation of output and efficiency w.r.t. intermediate pressure
under supercritical conditions.

5th European Thermal-Sciences Conference, The Netherlands, 2008
An analysis has been made to demonstrate the variations in output from turbine and the variations
in thermal efficiency of turbine with respect to different intermediate pressures under supercritical
conditions (246 bar/600°C/600°C) and high temperature sub-critical pressure conditions (142
bar/600°C/600°C) of steam.
Output from turbine is evaluated neglecting the extractions from I P & L P turbines and assuming
isentropic expansion of steam. Thermal efficiency of turbine is evaluated in accordance with
equation 1. Variation of output and thermal efficiency with respect to different intermediate
pressures under supercritical conditions is shown in fig. 3. Variation of output and thermal
efficiency with respect to different intermediate pressures under high temperature condition of
steam on sub-critical pressures as proposed here to design the units is shown in fig. 4.
Output vs. IP
Efficiency vs. IP
700
46
Output in MW
650
600
550
Efficiency in %
45.5
45
44.5
500
10 20 30 40 50 60 70 80
44
10 20 30 40 50 60 70 80
Intermediate Pressure (IP) in bar
Intermediate Pressure (IP) in bar
Figure 4: Variation of output and efficiency w.r.t. intermediate pressure under high
temperature condition of steam on sub-critical pressures.
3.3 Analysis under different operating conditions
Analysis has been made on the cycle as shown in fig. 1 with rated steam parameters for most
common sub-critical unit (130 bar/535°C/535°C) of 110 MW (case A) assuming cylinder
efficiencies of H P turbine, I P turbine and L P turbine as 0.90, 0.92 and 0.90 respectively as
available with modern design of turbines used in USC plants. Intermediate pressure for reheating of
steam is taken as 38 bar as per common design. Rated main steam flow rate has been modified to
maintain the rated output, in view of assumption of higher cylinder efficiencies as compared with
actual values of cylinder efficiencies for conventional sub-critical units. Output from the unit and
thermal efficiency of unit has been evaluated in accordance with equations and assumptions as
given in the beginning of this section.
Further analysis has been made on the cycle with common conditions of commercial super critical
plants as 246 bar/600°C/600°C, with same flow rates under same conditions (case-B). The cycle
was subsequently analyzed considering high temperature conditions on sub-critical pressures (142
bar/600°C/600°C) with same flow rates under same conditions with intermediate pressure of 38 bar
(case-C). To show the benefits of selecting other technically feasible intermediate pressures for
proposed steam conditions of 142 bar/600°C/600°C, subsequent analysis at intermediate pressure of
15 bar has been done (Case-D). Results of analysis have been tabulated in table 1 and parameters
for above analysis are given in table 2 to 5.
3.4 Analysis based on exergy
Exergy analysis has emerged as an effective tool in recent days to study and analyze the
performance of energy conversion systems including coal power plants through pinpointing the

5th European Thermal-Sciences Conference, The Netherlands, 2008
magnitude and quantum of irreversibilities in the system and to improve the performance of the
system based on minimization of exergy losses and consumption.
Quijano (2000), Kotas (1995) and Bejan et al. (1996) have beautifully defined the term “Exergy”
and concept of “Exergy Analysis”. Rosen (2001) compared the working of a coal fired and a
nuclear power plant based on energy and exergy. Subsequent exergy studies of conventional steam
power plants by Sengupta et al. (2006) and Bandpy and Ebrahimian (2007) revealed similar
results. Another important study in this field by Rosen and Dincer (2004) reported that exergy
analysis is not significantly sensitive to reasonable variations in reference (dead state) properties.
To perform the exergy analysis under different operating conditions (case A, B, C & D) as given in
previous sub-section, the “environment” as modeled by Bejan et al. (1996) has been taken. The
value of uniform temperature T o and pressure p o have been taken as that of average environmental
conditions i.e. 25 o C and standard atmospheric pressure of 1.01325 bar. Physical exergies of
flowing streams at different locations are evaluated with the relation Exergy =m (h - T o s) and
shown in table 2 to 5. Fuel input to the boiler is evaluated in accordance with operating parameters
and thermal efficiency. To evaluate exergetic efficiency of boiler, exergy (chemical exergy)
associated with fuel (coal) input is calculated as per empirical relation given in text by Kotas
referred above.
Exergetic efficiency of unit is evaluated with the relations (1), (2) and (3) modified for Exergy
(A ph ) in place of Energy (E) along with suitable assumptions as mentioned for calculation of
thermal efficiency. Exergy losses & consumption in steam pipes, condensate water and feed water
system is assumed as 5 % instead of 2% of heat losses as considered for calculation of thermal
efficiency. The exergetic efficiencies obtained under different operating conditions are tabulated in
table 1.
Table 1
Sl.
No.
4. Conclusions
Different conditions of analysis
Output at
generator
terminals
(MW)
Thermal
Efficiency
of Unit (%)
Exergetic
Efficiency
of Unit (%)
1 Sub-critical conditions (case-A) 110.54 MW 37.94 33.89
2 Super-critical conditions (case-B) 131.22 MW 40.85 36.63
3 high temperature conditions at sub-critical
pressure with intermediate pressure of 38
bar (case-C)
122.11MW 39.41 34.93
4 high temperature conditions at sub-critical
pressure with intermediate pressure of 15
bar (case-D)
133.49 MW 40.09 34.92
Analysis shows that higher output can be obtained with proposed design of units operating with
high temperature steam on sub-critical pressures comparing with the output of supercritical units
operating with same steam flow rates. Thermal efficiency of these proposed units is comparable
with super critical units, which may be improved up to the efficiency of supercritical units with
proper selection of design conditions. However exergy analysis reveals that these proposed units
are relatively less efficient to utilize the available exergy of coal comparing with supercritical units.
The units shall be more economical and cost effective comparing with super critical units due to
relatively low pressure designing of plant components.

5th European Thermal-Sciences Conference, The Netherlands, 2008
References
AD 700 Project, https://projectweb.elsam-eng.com/AD700/default.aspx.
Bandpy, M. J. and Ebrahimian, V., 2007, Exergy analysis of a steam power plant: a case study in
Iran, Int. J. Exergy, Vol. 4, pp. 54-73.
Bejan, A.; Tsatsaronis, G. and Moran, M., 1996, Thermal design and optimization, John Wiley &
Sons Inc., 1996.
Booras, G., Combustion technology university alliance workshop, Columbus, OH, August 4, 2003.
Kotas, T. J., 1995, The exergy method of thermal plant analysis, Reprinted, Butterworths, London,
1995.
Quijano, J., 2000, Exergy analysis for the Ahuachapan and Berlin geothermal fields, Proc. World
Geothermal Congress, Kyushu-Tohoku, Japan, May 28-June10, 2000.
Rosen, M. A., 2001, Energy and exergy based comparison of coal fired and nuclear steam power
plants, Exergy an Int. Journal 1 (3), pp 125-127.
Rosen, M. A. & Dincer, I., 2004, Effect of varying dead-state properties on energy and exergy
analysis of thermal systems, Int. J. of Thermal Sciences, 43, pp 121-133.
Sengupta, S.; Gupta, S. D. and Dutta, A., 2006, Exergy analysis of a coal-based 210 MW thermal
power plant, Int. J. Energy Research, Vol. 31, pp. 14-28.
Shigeru Azuhata, 2001, IEEE Power Engineering Review, March 2001.
Viswanathan, R., 2001, Boiler materials for ultra supercritical coal power plants, USC Materials
quarterly report, EPRI Inc., Oct-Dec 2001.
Viswanathan, R.; Alto, P.; Henry, J. F.; Chattanooga, A.; Tanzosh, J.; Stanko, G.; Shingledecker, J.
and Weighardt, K., 2001, The advanced supercritical 700°C pulverized coal fired power plant,
Powwergen Europe, Brussels, May 2001.
Yamamoto, K.; Kajigaya, I. and Umaki, H., 2003, Operational Experience of USC Steam Condition
Plant and PFBC Combined Cycle System with Material Performance, Materials at High
Temperatures, Vol. 20, pp. 15-18.
Table 2
Parameters for analysis under case A
Points Temp Press Flow Enthalpy Entropy Energy Exergy
°C Bar Ton/Hr kJ/kg kJ/kgK kW kW
1 535.0 130.00 343.00 3430 6.557 326802.78 140537.62
2 351.2 38.00 343.00 3100 6.616 295361.11 107419.94
3 535.0 35.00 311.50 3530 7.258 305443.06 118199.31
4 2.50 2861 7.385
5 2.50 2861 7.385
6 41.5 0.08 211.88 2366 7.558 139252.24 6626.09
7 81.8 0.51 8.96 2604 7.467 6481.07 940.09
8 86.4 0.61 11.82 2629 7.456 8631.88 1333.01
9 141.9 1.40 17.77 2757 7.413 13608.86 2699.13
10 214.9 3.00 17.75 2896 7.375 14278.89 3437.31
11 324.5 8.00 16.08 3108 7.321 13882.40 4132.76
12 427.0 17.30 27.24 3311 7.285 25053.23 8618.26
13 353.0 38.00 31.50 3104 6.623 27160.00 9881.83

5th European Thermal-Sciences Conference, The Netherlands, 2008
Table 3
Parameters for analysis under case B
Points Temp Press Flow Enthalpy Entropy Energy Exergy
°C Bar Ton/Hr kJ/kg kJ/kgK kW kW
1 600.0 246.00 343.00 3495 6.371 332995.83 152014.39
2 315.0 38.00 343.00 3008 6.464 286595.56 102972.26
3 600.0 35.00 311.50 3678 7.434 318249.17 126464.92
4 239.5 2.50 2948 7.562
5 239.5 2.50 2948 7.562
6 41.5 0.08 211.88 2425 7.746 142724.72 6799.59
7 93.2 0.51 8.96 2669 7.649 6642.84 966.81
8 107.0 0.61 11.82 2695 7.637 8848.58 1372.53
9 180.3 1.40 17.77 2834 7.591 13988.94 2817.25
10 258.6 3.00 17.75 2985 7.550 14717.71 3618.87
11 376.3 8.00 16.08 3217 7.496 14369.27 4386.57
12 486.1 17.30 27.24 3440 7.462 26029.33 9195.05
13 315.0 38.00 31.50 3008 6.464 26320.00 9456.64
Table 4
Parameters for analysis under case C
Points Temp Press Flow Enthalpy Entropy Energy Exergy
°C Bar Ton/Hr kJ/kg kJ/kgK kW kW
1 600.0 142.00 343.00 3589 6.709 341951.94 151368.92
2 278.4 38.00 343.00 3202 6.774 305079.44 112649.96
3 600.0 35.00 311.50 3678 7.434 318249.17 126464.92
4 239.5 2.50 2948 7.562
5 239.5 2.50 2948 7.562
6 41.5 0.08 211.88 2425 7.746 142724.72 6799.59
7 93.2 0.51 8.96 2669 7.649 6642.84 966.81
8 107.0 0.61 11.82 2695 7.637 8848.58 1372.53
9 180.3 1.40 17.77 2834 7.591 13988.94 2817.25
10 258.6 3.00 17.75 2985 7.550 14717.71 3618.87
11 376.3 8.00 16.08 3217 7.496 14369.27 4386.57
12 486.1 17.30 27.24 3440 7.462 26029.33 9195.05
13 315.0 38.00 31.50 3008 6.464 26320.00 9456.64
Table 5
Parameters for analysis under case D
Points Temp Press Flow Enthalpy Entropy Energy Exergy
°C Bar Ton/Hr kJ/kg kJ/kgK kW kW
1 600.0 142.00 343.00 3589 6.709 341951.94 151368.92
2 278.4 15.00 343.00 2989 6.831 284785.28 90736.59
3 600.0 14.00 311.50 3695 7.871 319720.14 116662.06
4 318.8 2.00 3110 7.958
5 318.8 2.00 3110 7.958
6 41.5 0.08 211.88 2554 8.156 150317.09 7197.36
7 73.4 0.15 8.96 2636 8.118 6560.71 536.65
8 111.1 0.25 11.82 2707 8.078 8887.98 980.22
9 266.8 1.28 17.77 3007 7.981 14842.89 3097.24
10 346.3 2.50 17.75 3165 7.946 15605.21 3924.23
11 401.1 3.80 16.08 3276 7.926 14632.80 4077.46
12 499.5 7.50 27.24 3480 7.896 26332.00 8518.61
13 278.4 15.00 31.50 2989 6.831 26153.75 8332.95