UTILIZATION OF SOLAR ENERGY TO SUPPLEMENT
THE COMBINED CYCLE POWER PLANT
Yuk L. Cheung
B.S., University of California, Berkeley, 2007
PROJECT
Submitted in partial satisfaction of
the requirements for the degree of
MASTER OF SCIENCE
in
MECHANICAL ENGINEERING
at
CALIFORNIA STATE UNIVERSITY, SACRAMENTO
SPRING
2010
UTILIZATION OF SOLAR ENERGY TO SUPPLEMENT
THE COMBINED CYCLE POWER PLANT
A Project
by
Yuk L. Cheung
Approved by:
__________________________________, Committee Chair
Tim Marbach, Ph.D.
____________________________
Date
ii
Student: Yuk L. Cheung
I certify that this student has met the requirements for format contained in the University format
manual, and that this project is suitable for shelving in the Library and credit is to be awarded for
the Project.
__________________________, Department Chair
Susan Holl, Ph.D.
Department of Mechanical Engineering
iii
________________
Date
Abstract
of
UTILIZATION OF SOLAR ENERGY TO SUPPLEMENT
THE COMBINED CYCLE POWER PLANT
by
Yuk L. Cheung
This paper presents an analytical review on the SMUD’s combined cycle power plant
modification project (in Rancho Seco, CA) with solar energy utilization to supplement the steam
turbine power generation process. During summer, the change in ambient temperature and air
density results in lower amount of steam produced and reduced the power capacity in the plant.
Thus, SMUD suggests using a concentrated solar power system to heat additional water and
injects the solar steam as a steam supplement to the system. SMUD has provided solar steam
parameters, heat balance data, and flow diagram of the plant, evaluation of different potent
designs is conduct. The result will serve as reference to SMUD for the actual project. The design
performance is measured in term of the net power gain, the cost of pumping throughout the
system, and the overall system modification. Among the discussed designs, injection of solar
steam between the IP steam turbine and the reheater, and condensate return by pumping from the
condenser directly without heating, would give the most optimal performance; providing a
theoretical power gain of 19.0 MW, or 9.8% capacity improvement, with flow rate of 168,600
lb/hr in the concentrated solar power system.
_______________________, Committee Chair
Tim Marbach, Ph.D.
_______________________
Date
iv
Acknowledgments
ACKNOWLEDGMENTS
I would like to thank my project advisor Dr. Tim Marbach for his help, support and guidance
throughout the work. In addition, I am in gratitude with SMUD for giving me this valuable
opportunity to use this Cosumnes Power Plant modification project as my master study. Without
them, this report and the analysis would be impossible to begin with.
v
TABLE OF CONTENTS
Page
Acknowledgments............................................................................................................................ v
List of Tables ................................................................................................................................ viii
List of Figures ................................................................................................................................. ix
Nomenclature ................................................................................................................................... x
Chapter
1. INTRODUCTION ....................................................................................................................... 1
2. BACKGROUND INFORMATION ............................................................................................ 3
2.1 Combined Cycle Power Plant ........................................................................................... 3
2.2 Combustion Gas Turbine .................................................................................................. 4
2.3 Heat Recovery Steam Generator ....................................................................................... 5
2.4 Steam Turbine ................................................................................................................... 6
2.5 Rankine Cycle ................................................................................................................... 7
2.6 Simplified Combined Cycle Flow Diagram ...................................................................... 8
2.7 The Complete Flow Diagram & Rated System Performance ........................................... 9
2.7.1 Gas Turbine Performance .................................................................................... 10
2.7.2 Steam Turbine Performance ................................................................................ 10
2.8 The Summer Time “Issue” .............................................................................................. 11
3. PROPOSED SOLUTION & DESIGN PROCEDURE ............................................................. 13
3.1 Concentrated Solar Power ............................................................................................... 13
3.2 The Design Parameters ................................................................................................... 15
3.3 Different Ways to Utilize Solar Steam ............................................................................ 16
4. DESIGN ANALYSES & DISCUSSION .................................................................................. 19
vi
4.1 Solar Steam Injection Location ....................................................................................... 19
4.2 Condensate Return Option .............................................................................................. 21
4.2.1 Condensate Return Option #1 ............................................................................. 22
4.2.2 Condensate Return Option #2 ............................................................................. 23
4.3 Different Scenarios of Solution ....................................................................................... 26
4.3.1 Scenario #1: Injection after Reheater with Return Option #1 ............................. 27
4.3.2 Scenario #2: Injection after Reheater with Return Option #2 & Solar Steam ..... 29
4.3.3 Scenario #3: Injection after Reheater with Return Option #2 & Hot Reheat ...... 31
4.3.4 Scenario #4: Injection before Reheater with Return Option #1 .......................... 33
4.3.5 Scenario #5: Injection before Reheater with Return Option #2 & Solar Steam .. 37
4.3.6 Scenario #6: Injection before Reheater with Return Option #2 & Hot Reheat ... 39
4.4 Summary of Performance Results ................................................................................... 41
4.5 The Baseline Scenario ..................................................................................................... 43
5. THE BEST DESIGN ................................................................................................................. 48
Appendix A1. CPP Flow Diagram (Courtesy to SMUD) ............................................................ 50
Appendix B1. Original HP Steam Turbine Performance Analysis .............................................. 51
Appendix B2. Original IP Steam Turbine Performance Analysis ............................................... 52
Appendix B3. Original LP Steam Turbine Performance Analysis .............................................. 53
Appendix C1. Properties of Solar Steam at Different Pressures .................................................. 54
Appendix D1. Flow Properties Analysis across IP & LP Turbines in Scenario #1 ..................... 55
Appendix D2. Flow Properties Analysis across IP & LP Turbines in Scenario #2 ..................... 56
Appendix D3. Flow Properties Analysis across IP & LP Turbines in Scenario #3 ..................... 57
Appendix D4. Flow Properties Analysis across IP & LP Turbines in Baseline .......................... 58
Bibliography .................................................................................................................................. 59
vii
LIST OF TABLES
Page
List of Tables
1.
Table 1: Summary of Original Inlet and Outlet Conditions across All Steam Turbines ....... 10
2.
Table 2: Flow Conditions at Different Stages of Process (Refer to Figure 11)..................... 22
3.
Table 3: Flow Conditions (with Solar Steam as Source) at Different Stages of Process ...... 24
4.
Table 4: Flow Conditions (with Hot Reheat as Source) at Different Stages of Process ....... 25
5.
Table 5: Flow Conditions from Stage A to D in Scenario #1 ............................................... 28
6.
Table 6: Flow Conditions from Stage A to D in Scenario #2 ............................................... 31
7.
Table 7: Flow Conditions from Stage A to D in Scenario #3 ............................................... 33
8.
Table 8: Flow Conditions from Stage A to E in Scenario #4 ................................................ 36
9.
Table 9: Flow Conditions from Stage A to E in Scenario #5 ................................................ 38
10. Table 10: Flow Conditions from Stage A to E in Scenario #6 .............................................. 40
11. Table 11: Summary of Performance Results for All Design Scenarios ................................ 42
12. Table 12: Flow Conditions from Stage A to D in Baseline Scenario .................................... 46
viii
LIST OF FIGURES
Page
List of Figures
1.
Figure 1: A Typical Combined-Cycle Power Plant................................................................. 3
2.
Figure 2: A Typical Combustion Gas Turbine ........................................................................ 4
3.
Figure 3: A Typical Counter-Flow Heat Exchanger, and the Temperature Distribution of the
Fluids in the Exchanger along Tube Axis ............................................................................... 5
4.
Figure 4: A Typical Multi-level Steam Turbine...................................................................... 7
5.
Figure 5: Schematic of a Basic Rankine Cycle ...................................................................... 8
6.
Figure 6: Schematic of a Simplified Combined Cycle System ............................................... 9
7.
Figure 7: Gas Turbine Performance as Function of Ambient Temperature .......................... 12
8.
Figure 8: Diagrams of the Four Popular Types of CSP Systems .......................................... 14
9.
Figure 9: Site Map of the CCPP and the CSP Field .............................................................. 15
10. Figure 10: Schematic of the Steam Flow Diagram ............................................................... 17
11. Figure 11: Schematic of the Condensate Return Option #1 .................................................. 22
12. Figure 12: Schematic of the Condensate Return Option #2 .................................................. 24
13. Figure 13: Schematic of the Design Scenario #1 .................................................................. 27
14. Figure 14: Schematic of the Design Scenario #2 .................................................................. 30
15. Figure 15: Schematic of the Design Scenario #3 .................................................................. 32
16. Figure 16: Schematic of the Design Scenario #4 .................................................................. 35
17. Figure 17: Schematic of the Design Scenario #5 .................................................................. 38
18. Figure 18: Schematic of the Design Scenario #6 .................................................................. 39
19. Figure 19: The Baseline Design Scenario ............................................................................. 45
20. Figure 20: Schematic of the Recommended Solution Design ............................................... 49
ix
NOMENCLATURE
Nomenclature
T
Fluid Temperature
P
Fluid Pressure

m
Fluid Mass Flow Rate

Thermal Efficiency (i.e. Turbine, Pump, or Overall)
h
Enthalpy Value of the Fluid
s
Entropy Value of the Fluid
E
Rate of Heat Transfer between Two Fluids
Q
Rate of Heat Transferred into Fluid
U
Overall Heat Transfer Coefficient inside a Heat Exchanger
A
Surface Area of the Pipe in a Heat Exchanger
∆T
Logarithmic Mean Temperature Difference
W turbine
Steam Turbine Power Output
W pump
Feed Pump Driving Power
HP
Acronym for High Pressure
IP
Acronym for Intermediate Pressure
LP
Acronym for Low Pressure
HPSH
Acronym for High Pressure Super Heater
IPSH
Acronym for Intermediate Pressure Super Heater
LPSH
Acronym for Low Pressure Super Heater
STG
Acronym for Steam Turbine Generation
CTG
Acronym for Combustion Turbine Generation
x
CND
Acronym for Condenser Hot Well
HRSG
Acronym for Heat Recovery Steam Generator
CSP
Acronym for Concentrated Solar Power
CCPP
Acronym for Combined Cycle Power Plant
xi
1
Chapter 1
INTRODUCTION
SMUD’s Cosumnes Power Plant, a combined cycle power plant (CCPP) that is located in
Rancho Seco, California, has an output under-capacity issue during summer time. In summer,
energy demand is the highest; however, the power plant is not capable of producing maximum
power output due to the ambient temperature and air density changes. As a result, the amount of
steam that is produced from the heat recovery steam generation (HRSG) system decreases. This
ultimately affects the steam turbine power generation (STG) process, which is one of the two
major power production processes in the CCPP. And the result of this summer “issue” is the loss
of power output capacity.
Some solutions to fix the issue are by burning more fuel in the combustion turbine to
compensate the effects of the ambient change, or by using an absorption chiller to cool off the
intake air temperature. However, they are either not sustainable idea or decrease the available
heat for steam production in the HRSG unit. Nevertheless, they do not follow the current the
trend of renewable energy development. Therefore, SMUD has come up with another “green”
solution; that is to utilize solar thermal collectors to make steam, and using this solar steam to
compensate for the lost steam in the STG process.
A solar thermal collector utilizes solar thermal energy, by concentrating the sun radiation
energy, to heat water into steam. Then the solar steam is injected and mixed with the regular flow
in the STG system to make up the lost amount of steam due to the temperature issue. The solar
steam brings in additional energy into the system and could be converted into electricity power
when passing thru the turbines. This additional gain will be used to compensate the lost output
capacity and serves as a means to solve the under capacity issue. In another sense, solar steam
2
could also be used to bring up the temperature of exiting flow from the high pressure (HP) steam
turbine, which could reduce the load on the reheater and the cost of operation in other seasons.
Either way, the utilization of solar thermal energy should give us a win-win situation to both the
company and the environment in the end.
The design parameters of the solar steam and return condensate from the Concentrated
Solar Power (CSP) system, the heat balance data and the detail flow diagram of the CCPP are
provided by SMUD. In return, evaluation of the idea using solar energy as a supplement in the
power generation cycle is conducted in this project.
The main goal of this study is to determine the feasibility of the injection of solar steam
into the existing power generation cycle. Thermodynamics knowledge and heat transfer analysis
are applied to investigate the effects of mixing solar steam with the regular flow in the system.
The potential condensate return options are also included in the discussion. Different scenarios of
injection are designed and performance is evaluated based on the net power gain and capacity
improvement, the cost of pumping, the additional heating/fueling requirement, and the degree of
overall system modification.
3
Chapter 2
BACKGROUND INFORMATION
2.1 Combined Cycle Power Plant
The Rancho Seco combined cycle power plant consists of two gas turbine generators,
equipped with heat recovery steam generator unit (HRSG), and three steam powered turbines.
The plant is large and rated in the hundred of mega-watts. Figure 1 shows a typical CCPP. The
combined cycle plant combines the Brayton (gas turbine) and the Rankine (steam turbine)
thermodynamic cycles by using heat recovery steam generator (HRSG) to capture the energy in
the gas turbine exhaust for steam production. Pressurized steam are generated from the HRSG
and used to drive the steam turbines for power output. The combustion turbines energy
conversion typically ranges between 25% to 40% efficiency as a simple cycle. And combined
cycles have a higher thermal efficiency (up to 60%) than the steam or gas turbine cycles operating
alone [1] [2].
Figure 1: A Typical Combined-Cycle Power Plant
4
2.2 Combustion Gas Turbine
The combustion gas turbine being used in this natural gas fuelled power plant consists of
inlet section, compressor, combustion system, turbine, exhaust system, and exhaust diffuser as in
figure 2.
Figure 2: A Typical Combustion Gas Turbine
The compressor draws air into the engine, pressurizes it and feeds it to combustion at
high speed. The combustion system is made of fuel injectors that inject steady streams of fuel into
the combustion chambers where it mixes with air. The mixture is then burned at a high
temperature (>1093 deg. C). Turbine has a stationary and rotating aerofoil-section, as hot
combustion gas expands through the turbine and spins the rotating blades. The rotating blades
would drive the compressor to draw more pressurized air into the combustion system, and spin a
generator to produce electricity. The exhaust gases exiting the engine could be recovered and
used to heat up water into steam and generate more electricity from the steam turbine [2] [3].
5
2.3 Heat Recovery Steam Generator
The HRSG unit is installed after the gas turbine exhaust to recover the exhausting heat in
gases mixture by generating superheat steam to operate the steam turbines. The HRSG system
consists of reheater, super heater, evaporator, economizer, preheater, drum and dearator. See
Appendix A for details.
The HRSG system employs the counter-flow heat exchanger principle to extract the heat
from the flue gases in order to generate steam for the steam turbine. The equations involve with
counter-flow heat exchanger are shown below [4]. Figure 3 illustrates a typical counter-flow
exchanger and the temperature distribution inside the exchanger.
Figure 3: A Typical Counter-Flow Heat Exchanger, and the Temperature Distribution of the
Fluids in the Exchanger along Tube Axis
In a counter-flow heat exchanger, fluids flow in the opposite direction. If the specific heat
capacity of fluids is constant, it can be shown that:
E  U  A  T
(Eq. 1),
6
where E is the rate of heat transfer between two fluids, U is the overall heat transfer coefficient,
A is the surface area of the tube, and ∆T is the logarithmic mean temperature difference.
If the inlet and outlet conditions of the fluids in the heat exchanger are available, by
conservation of energy, the energy transferred between the two fluids could be calculated
according to (Eq. 2) below,
 a  (ha _in  ha _out )  m
 b  (hb _out  hb _in )
E = m
(Eq. 2),
 is the fluid mass flow rate in the pipe,
where E is the rate of heat transfer between two fluids, m
and h is the enthalpy value of the fluid at the defined temperature and pressure.
2.4 Steam Turbine
A steam turbine is a mechanical device that converts thermal energy in pressurized steam
into useful mechanical work. Steam turbine is ideal for the very large power configurations used
in power stations. The steam turbine derives much of its thermodynamic efficiency using multiple
stages steam expansion. This results in a closer approach to the ideal reversible process.
Steam turbines are made in a variety of sizes ranging from small 0.75 kW to 1,500 MW
units. They are widely used for marine vessel propulsion systems. In recent times gas turbines, as
developed for aerospace applications, are being used more and more in the field of power
generation once dominated by steam turbines [5][6].
Steam energy is converted to mechanical work by expansion through the turbine. The
expansion takes place through a series of fixed blades and moving blades. The moving blades
rotate on the central turbine rotor and the fixed blades are concentrically arranged within the
circular turbine casing, which is substantially designed to withstand the steam pressure.
On large output power plant, if the duty is too large for one turbine, numbers of turbine
casing/rotor units are combined to achieve the duty. These are generally arranged on a common
7
centerline (tandem mounted). For this project, the steam turbine system used is one of those types
mentioned previously and is shown in figure 4 below.
Figure 4: A Typical Multi-level Steam Turbine
2.5 Rankine Cycle
The Rankine Cycle is a steam cycle for a steam power plant operating under the
theoretical conditions for high efficiency. It is an ideal imaginary cycle against which all other
real steam working cycles can be compared. This ideal cycle is shown below in figure 5, along
with steam flow reference. The cycle assumes no radiation loss, steam leakage, or frictional loss
in the mechanical components. The condenser will condense the superheated steam to saturated
water vapor. The feed pump and turbine are isentropic, which means they are reversible &
adiabatic [5].
Inside the turbine, the power output is equal to fluid flow rate times the difference of
fluid enthalpies across the turbine, namely (h1 – h2). Heat transfer from the gas turbine flue gases
thru the HRSG system supplies most of the energy in the steam. The stored energy in the flow
will be the difference in the enthalpy of the steam leaving the boiler and the water entering the
feed pump, or simply (h1 – h3). The thermal efficiency of the Rankine cycle and is defined as,
8
(Eq. 3),
where h is the enthalpy value of fluid at different state of process. In addition, the turbine power
output can be calculated as,
  (h1  h2 )
Turbine Output = W turbine = m
(Eq. 4),
 is the fluid mass flow rate, and h is the enthalpy of
where W turbine is the turbine power output, m
the fluid across the turbine. Also the pump power input can be defined in a similar way,
  (h4  h3 )
Pump Input = W pump = m
(Eq. 5),
 is the fluid mass flow rate, and h is the enthalpy of the
where W pump is the pump power input, m
fluid across the pump.
Figure 5: Schematic of a Basic Rankine Cycle
2.6 Simplified Combined Cycle Flow Diagram
Combined cycle power generation involves many different thermodynamics concepts.
Figure 6 below shows the various energy streams flowing in a simplified combined cycle system.
It is clear that the working fluid is in a closed circuit apart from the free surface of the hot
9
well. Every time the working fluid flows at a uniform rate around the circuit, it experiences a
series of processes making up a thermodynamic cycle.
When the turbine system is operating under steady state conditions, the law of
conservation of energy dictates, so that the flow of energy entering any system boundary must be
equal to the rate of energy leaving the system boundary. This allows energy equations to be
developed across each component unit and the whole plant.
Figure 6: Schematic of a Simplified Combined Cycle System
2.7 The Complete Flow Diagram & Rated System Performance
In order to design and evaluate the new system performance with the solar steam
injection, SMUD has provided a basis set of heat balance data in a spreadsheet. The heat balance
10
 ), temperature (T), pressure (P),
data includes the steam flow conditions, i.e. mass flow rate ( m
and enthalpy (h), across all devices (gas turbine, steam turbine, pump, etc.) at ambient
temperature of 104 deg. F [7]. The complete flow diagram for the plant is also included in the
spreadsheet and is shown in Appendix A1.
2.7.1 Gas Turbine Performance
The operating parameters of gas turbine unit (two identical units) are as follows:
Gas turbine (GTD 1 or GTD 2) - consumes 7,358 lb/hr of natural gas, and 3,338,800 lb/hr
of air at ambient temperature of 104 deg. F and pressure of 14.7 psia, to produce 159.2
MW of electricity and emits 3,434,000 lb/hr of exhaust gas at a temperature of 1137.0
deg. F and pressure of 15.0 psia. [7] Data are extracted from the attached heat balance
data provided by SMUD. Since there are two combustion gas turbines, the total power
output by combustion is equal to 318.4 MW.
2.7.2 Steam Turbine Performance
The operating parameters of steam turbine units (HP, IP, & LP) are extracted
from the heat balance data [7] and are summarized in the following table.
Table 1: Summary of Original Inlet and Outlet Conditions across All Steam Turbines
Properties
Q (lb/hr)
HP Turbine
Inlet (S63)
Outlet (S68)
827,881
827,881
IP Turbine
Inlet (S67)
Outlet (S86)
948,370
948,370
LP Turbine
Inlet (S83)
Outlet (S78)
1,072,500
1,072,500
T (F)
1051.1
688.2
1050.8
540.9
544.5
101.2
P (psia)
1787.2
439.3
406.8
56.5
56.5
0.99
h (btu/lb)
1512.7
1353.5
1550.5
1303.1
1304.9
1034.4
11
With the given data, the work output from each steam turbine is calculated according to (Eq. 4).
In addition, the ideal turbine output can be calculated with assumption of isentropic properties;
solving the isentropic outlet enthalpy of the steam (h2S is equal to h(s = s1 & P = P2)) with the
extrapolation method on the steam properties table in reference [5]. The isentropic efficiency (η)
of the turbines can be calculated as,

WTurbine, Actual
W
(Eq. 6),
Turbine, Ideal
where η is the isentropic turbine efficiency, and W turbine is the turbine power output.
The performance of the steam turbines (HP, IP, and LP) in this project are evaluated, and
results are included in Appendix B1 to B3. See Appendix B1 to B3 for the calculation details.
The total power output by steam expansion is equal to the sum of 38.6 MW (from HP turbine),
68.8 MW (from IP turbine) and 85.0 MW (from LP turbine), or roughly ~192.4 MW.
2.8 The Summer Time “Issue”
During summer period with high-energy demand, the CCPP experiences an undercapacity issue, which is not capable of generating the rated output power. The major cause
behind it is due to the effects of the ambient air temperature rise, the air density decreases. As
a result, fewer amount of flue gas is available from the combustion turbine thus reduces the
amount of steam produced in the HRSG system. As shown in figure 7 below, there is an
obvious drop in the power output as the ambient air temperature increases and the air density
decreases. This is particularly relevant in the summer time where the temperature is easily
over 100 deg. F [8].
Since the amount and the quality of the steam production in the HRSG are directly
related to the heat in the gas turbine exhaust. A smaller flow rate of flue gas would decrease the
12
temperature of the steam flow, or if the flow conditions are kept the same, less steam is available,
since the HRSG is not as efficient as before. In consequence, less enthalpy (or energy content) is
in the steam flow and less power could be generated from the steam turbine power (STG)
process.
Figure 7: Gas Turbine Performance as Function of Ambient Temperature
According to SMUD, data suggests that the power plant has “lost” capacity during
summer time. At ambient temperature of around 104 deg. F, the STG system only operates at
88% of its rated capacity; with maximum power output of ~193 MW, only ~171 MW of power is
being produced from the steam turbines [9]. This matches the curve and expectation shown in
figure 7 earlier. Ultimately, the goal, or design solution, is to bring back the STG lost capacity
(~22 MW) of power as much as possible.
13
Chapter 3
PROPOSED SOLUTION & DESIGN PROCEDURE
There are many ways to improve the capacity of the STG output. One is to control the
intake air to the ISO conditions (15 deg. C. and 100% relative humidity) before entering the gas
turbine. However, cooling the intake air would consume additional power (if air conditioning is
applied) or reduce the total heat content available in the system (if absorption chiller is used and
heat input is extracted from the system). Comparison between different solutions will be included
in the final chapter of this report, in order to show the feasibility of the design. Besides, those
methods are counter-intuitive to the trend of developing sustainable and renewable energy.
To compensate for the lost capacity due to the reduction in the steam flow, SMUD has
suggested the use of concentrated solar power (CSP) field as additional boiler to make additional
steam and inject the solar steam to make up the lost steam mass in the STG system during
summer time when the ambient temperature rises. The fundamental concept behind the idea is
similar to adding additional heat into the existing steam flow, in order to make up the lost steam
quality in the HRSG process.
3.1 Concentrated Solar Power
Concentrated solar power (CSP) is an indirect way to utilize solar power (contrary to
photovoltaic), in which the sun’s energy is focused to boil water into steam and then used to drive
a turbine to produce electricity in a steam power generation system. These systems use lenses or
mirrors and tracking systems to focus a large area of sunlight into a small beam. The concentrated
heat beam is then used to heat up the working fluid in the thermal collector [10].
14
A wide range of concentrating technologies exists; the most developed are the parabolic
trough, the concentrating linear fresnel reflector, the Stirling dish and the solar power tower [10]
[11]. Below in Figure 8 illustrates the four types of CSP systems previously mentioned.
Figure 8: Diagrams of the Four Popular Types of CSP Systems
CSP field usually requires large area of sunlight coverage for pure solar power plant. But
in this project, the design only requires solar power to supplement the steam power generation
process; that is to create some new steam to supplement the lost steam mass, instead of driving
the whole plant alone. Therefore, the size of the CSP system is relatively smaller than a pure solar
power plant or solar farm. In figure 9 below, a Google map view is shown for the project site
location. In order for this modification project to become economically feasible, SMUD has
expected an average improvement of ~15 MW (or ~7.8% capacity) requirement from the
installation of CSP system and the design [9].
15
Decommissioned Rancho
Seco Power Plant
Proposed Re-power of
Existing PV Facility Using
Thermal Solar Technology
- CSP Field
Solar Steam & Condensate
Return Pipe to CPP Fence
Steam Meter Set
inside CPP Fence
Existing CPP Site
14295 Clay East Road
Figure 9: Site Map of the CCPP and the CSP Field
3.2 The Design Parameters
Design parameters, i.e. target conditions of the solar steam and condensate return, are
required for system designing. These target conditions from the CSP field have been estimated by
SMUD design engineers and put on a contractual term by the provider. The actual values and any
variations should not fluctuate much from the estimated values and will be determined by the
thermal energy purchase agreement with the independent power producer of the CSP field.
The reference values, such as the pressure, temperature and enthalpy for steam supply
and condensate return at the fence line meter, are showed following,
i) Steam (superheated) entering the CPP plant: T = 675 deg. F, P = 499.7 psia, and
h = 1343 btu/lb; and
16
ii) Condensate (compressed) leaving the CPP plant: T = 325 deg. F, P = 814.7 psia, and
h = 297 btu/lb.
Given these conditions of the solar steam and condensate, an critical flow rate of the solar
steam required in order to make up the lost capacity (~22MW) in STG process could be
calculated as shown below, based on the energy flow concept of the incoming hot solar steam and
outgoing cold condensate and ideal/isentropic heat transfer process.
 critical  (hsolar _ steam  hcondensate )
Wlost _capacity  m
22 MW  7.51  10 7
btu
btu
 critical  (1342  297)
m
hr
lb
 critical  71,866
m
lb
hr
Hence, the critical flow rate of the solar steam required to make up the lost capacity is around
71,866 lb/hr, assuming there is no energy loss in the process. However, that is never possible
because heat transfer process is not isentropic in reality; the energy loss has to be (and should be)
taken into consideration. Knowing the steam turbine’s limitation of less than 10% increment in
flow capacity, the CSP system should only be generating up to 200,000 lb/hr of solar steam
(without exceeding the steam turbine maximum design capacity) for STG steam supplementation
[9]. This limitation (& requirement) will become a design criterion for different designs/scenarios
of solar steam injection and condensate return option, along with the design parameters of the
CSP system.
3.3 Different Ways to Utilize Solar Steam
In order to create a sound design that can utilize the solar steam efficiently to the STG
supplementation, the understanding of the steam flow process is required. This is important
17
because the injection point of the solar steam has to be carefully selected in order for the steam
power generation to achieve its best efficiency. Therefore, a brief discussion of the steam flow in
the plant is included here. A schematic of the steam flow diagram showing the flow process is
listed below in figure 10.
To HRSG
To Ambient
Reheater
LPSH
HRSG #2
Gas Turbine
(GTD2)
HP
Steam
IP
Steam
LP
Steam
Condenser Hot Well
(CND1)
Gas Turbine
(GTD1)
Reheater
LPSH
HRSG #1
To HRSG
To Ambient
Figure 10: Schematic of the Steam Flow Diagram
Basically the steam comes from the heat recovery steam generators. Since the HRSG #1
is identical to HRSG #2, the discussion would only focus on either one side of the process (with
the other side being synchronous). The high pressure (HP) steam first goes thru the HP steam
turbine, and the steam would drive the turbine to make the generator produce electricity. The exit
18
steam (or the cold reheat) is then sent to the reheater and get heated back to the intermediate
pressure (IP) steam (or the hot reheat). The hot reheated steam goes thru the IP steam turbine and
continues on to the LP steam turbine. Superheated steam is injected around the process, i.e. into
the exit steam from the IP turbine, with complete utilization of the heat in the flue gases. As the
energy from the steam flow is converted into power, the cold condensate would
condense/accumulate in the condenser hot well (CND1). This condensate would be sent to the
HRSG and get heated into steam again, and the whole STG cycle repeats.
Looking into the system, there are many potential locations for injection or mixing of the
solar steam, i.e. injection before the steam turbine, or mixing the solar steam with the cold
condensate to reduce the heating load in the HRSG system, etc. Also depending on the
condensate return option, i.e. the availability of addition boiler for heating the cold condensate,
many different designs/scenarios are possible. With the consideration of solar steam condition,
the condensate return option, and the system modification limitation, different scenarios of
injection and condensate return are built and evaluated in term of turbine power output, pumping
cost, system’s change, net power gain, overall capacity improvement, and additional fuel
requirement (if applicable).
19
Chapter 4
DESIGN ANALYSES & DISCUSSION
4.1 Solar Steam Injection Location
The solar steam could be injected before the HP steam turbine, however, the inlet
condition of the HP steam turbine requires steam flow at P = 1787.2 psia, T = 1051.1 deg. F and h
= 1512.7 btu/lb (from Table 1). While the solar steam entering the plant at the fence line meter is
only at P = 864.7 psia, 675 deg. F and h = 1343 btu/lb. This means if injection is chosen at before
the HP turbine, additional pumping is required to increase the solar steam pressure in order to mix
with the steam flow at the HP turbine’s inlet. Nevertheless, the temperature of the steam mixture
would be lower than before because the solar steam is originally cooler than the regular steam,
thus affecting the whole STG process afterward. If additional heating is applied on the solar
steam to bring its temperature up to the inlet condition, much energy is required for the heating.
The purpose of this project is to increase the power output with minimal system changes and
additional heating. To summarize this, injection before HP turbine is a bad idea because of the
fact that additional cost for pumping and heating could outweigh the final gain in the steam power
output. Not forget to mention the difficulty of mixing high-pressure high-temperature steam flow
and the hardware compatibility problem are always some limiting factors to the design.
What about injection of solar steam in between the IP and LP turbines? If that is the case,
it would require throttling the pressure of solar steam from 499.7 psia to 56.5 psia. This not a
smart idea, since throttling process would reduce the energy content in the solar steam, and this is
a waste of energy. The solar steam also loses the driving pressure that could be used to transport
within the system. Moreover, letting the additional steam pass thru only one steam turbine would
20
give little improvement to the STG process. Thus, injection in between the IP and LP turbine is
not a good way to fully utilize the solar steam energy.
Similar reasoning could be applied to injection of solar steam into the cold condensate
from the condenser hot well (CND1). It would, theoretically, help reduce the heating load in the
HRSG system, since heating up water at warm temperature is always easier than doing so from
cool temperature. However, injection into the flow in HRSG would affect many parts of the STG
process. This designing requires many other details information (i.e. maximum flow capacity,
highest temperature limitation, etc.) for equipments involved in the process in order to complete
the analysis. And this is out of the range of the study, as certain information are confidential and
not provided by SMUD. One other noteworthy point is that the condensate returning to the CSP
field is at higher temperature than the cold condensate (from CND1), which means cold
condensate has to be heated before sending it back to the CSP field. This requires the system to
return some of the “heated” flow back to the CSP field and becomes somewhat contrary to the
project’s goal. So injection into the cold condensate is a complicated idea, and impossible with
the current available information.
Therefore, the remaining injection location for the solar steam is in between the HP and
IP turbines. Looking at the solar steam conditions and the steam conditions before the IP turbine,
it is clear that SMUD has intentional chosen the IP region for injection. As the IP turbine’s inlet
conditions are P = 406.8 psia, T = 1050.8 deg. F , and h = 1550.5 btu/lb, or the steam flow before
the reheater with P = 433.9 psia, T = 687.6 deg. F , and h = 1353.5 btu/lb. The difference between
the solar steam and the regular steam is insubstantial. Also the steam mixture would only go thru
the IP and LP turbines to generate power without affecting much of the remaining processes in
the system; little or no change is required for the system to utilize the solar steam energy. As
21
result, injection before the IP turbine becomes the most reasonable location. In fact, it is the best
location for injection.
4.2 Condensate Return Option
Injection of solar steam is important in the design process, so is the condensate return to
the CSP field. Different condensate return options would directly influence the ultimate design to
be employed, since the conditions for the condensate leaving the CCPP at the fence line meter is
specified and must be maintained for the whole project becomes retainable. The return
condensate is being sent back to the CSP field at a higher energy content (h = 297 btu/lb) and
pressure (P = 864.7 psia) than the saturated water directly pumped from the condenser hot well
(at S80: h = 68.3 btu/lb & P = 130.3 psia). This means either, 1) the return flow (equal to the
incoming solar steam flow) has to be heated up by an additional boiler, or 2) be mixed with some
high temperature steam (which is extracted from the system), and be pressurized to the target
condition before being pumped back to the CSP field.
If the second option is used, the location of steam extraction has to be carefully decided
without influencing other parts of the system, while the design still gives the a good power
improvement. With this in consideration, hot steam could be extracted either from the solar steam
directly (analogous to the principle of saving up a fraction of the solar energy to heat up the cold
water), or from the hot reheated steam. These two sources represent the energy outside and inside
the system. Extraction from HP steam is fundamentally same as extraction from IP steam; same
amount of energy is split from the STG system to heat the cold condensate. But HP steam is not
preferred since it affects many parts of the remaining processes. Also LP steam flow is best not be
chosen since most of the STG power output is relied on the flow thru the LP turbine.
22
Analyses on the required energy input into the cold saturated water (for the first option)
and the required extraction amount from the hot steam (for the second option) are calculated
below to aid the design process in the latter part of the report.
4.2.1 Condensate Return Option #1
In option #1, the return condensate is pumped from condenser hot well and heated up by
an additional boiler. The concept is illustrated below in figure 11, along with the fluid conditions
at different stages of process in table 2 below.
Wpump 1
A
○
Saturated
Water from
Condenser
Qin
B
○
Pump
#1
Wpump 2
C
○
Additional
Boiler
D
○
Pump
#2
Return
Condensate to
CSP Field
Figure 11: Schematic of the Condensate Return Option #1
Given the solar steam flow rate of 200,000 lb/hr, the return flow rate has to equal to the incoming
flow for a fully controlled system. The required heating cost and the pumping power consumption
could be calculated using the data in table 2.
Table 2: Flow Conditions at Different Stages of Process (Refer to Figure 11)
Properties
Location A
Location B
Location C
Location D
T (F)
100.0
100.1
324.5
325
P (psia)
0.99
130.3
130.3
864.7
h (btu/lb)
67.8
68.3
295.6
297
23
The required energy input for heating is calculated as,
lb
btu
btu
  (hc  hb )  200,000  (295.6  68.3)
Q in  m
 4.55  10 7
.
hr
lb
hr
The heating cost is around 4.55x107 btu/hr, or equal to ~13.3 MW.
Similarly, the required pumping power consumption to return the fluid is calculated as,
lb
btu
btu
  (hb  ha )  200,000  (68.3  67.8)
W pump1  m
 10
.  105
, and
hr
lb
hr
lb
btu
btu
  (hd  hc )  200,000  (297  295.6)
W pump 2  m
 2.8  105
.
hr
lb
hr
The total pumping cost is around 3.8x105 btu/hr, or roughly equal to 111 kW.
4.2.2 Condensate Return Option #2
In option 2, the return condensate is made of the mixture with cold water and hot steam.
The hot steam can be extracted from either the solar steam or the hot reheated steam. Depending
on the source, the fluid conditions (i.e. temperature and enthalpy value) would be different, and
the amount of extraction required to satisfy the return target condition would be different. The
concept of this return option is illustrated in figure 12 below, along with the fluid conditions at
different stages of process in table 3 and table 4 (each table with a different source of hot steam
listed).
Given the hot steam conditions, the amount of extraction from each steam source could
be calculated according to the following equations, with conservations of mass and energy.
a  m
 b (Conservation of Mass) ,
m
b  m
c  m
d  m
 e  200,000
m
lb
(Conservation of Mass), and
hr
24
 b  hb  m
 c  hc  m
 d  hd (Conservation of Energy).
m
The subscript under each variable refers to the specified stage of process shown in figure 12. In
addition, for this return option mixing of the cold water and hot steam is involved. The detail of
the mixing chamber calculations is not included in this report and is common seen in any
thermodynamics textbook.
Wpump 1
A
○
Saturated
Water from
Condenser
Qin
Pump
#1
Wpump 2
B
○
D
○
Mixing Chamber
Solar Steam
(S.S.)
or
Hot Reheat
(H.R.)
E
○
Pump
#2
Return
Condensate to
CSP Field
C
○
Figure 12: Schematic of the Condensate Return Option #2
Table 3: Flow Conditions (with Solar Steam as Source) at Different Stages of Process
Properties
Location A
Location B
Location D
Location E
100.3
Location C
(Solar Steam)
675
T (F)
100.0
324.7
325
P (psia)
0.99
499.7
499.7
499.7
864.7
h (btu/lb)
67.8
68.7
1343
296.1
297
Using the data in table 3 for the return method using solar steam as the heat source, the required
amount of solar steam extraction is calculated as
25
c 
Solar Steam = m
 d  (hd  hb )
m

(hc  hb )
lb
btu
 (2961
.  68.7)
hr
lb  35700 lb .
btu
hr
(1343  68.7)
lb
200,000
And the amount of saturated water required for mixing is determined as,
 b  200,000
Saturated Water = m
lb
lb
 c  164,300 .
m
hr
hr
The required pumping power cost in the process is calculated as,
lb
btu
btu
 b  (hb  ha )  164,300  (68.7  67.8)
W pump1  m
 15
.  105
, and
hr
lb
hr
lb
btu
btu
 e  (he  hd )  200,000  (297  2961
W pump 2  m
.)
 18
.  105
.
hr
lb
hr
The total power required for pumping in this return option is about 3.3x105 btu/hr, or ~96.7 kW.
Table 4: Flow Conditions (with Hot Reheat as Source) at Different Stages of Process
Properties
Location A
Location B
Location D
Location E
100.2
Location C
(Hot Reheat)
1050.8
T (F)
100.0
324.7
325
P (psia)
0.99
406.8
406.8
406.8
864.7
H (btu/lb)
67.8
68.5
1550.5
296.1
297
Similarly, when the hot reheated steam is used as the heat source, the required amount of hot
steam extraction and saturated water can be calculated using the values in table 4.
c 
Hot Reheat Steam = m
 d  (hd  hb )
m

(hc  hb )
lb
btu
 (2961
.  68.5)
hr
lb  30700 lb , and
btu
hr
(1550.5  68.5)
lb
200,000
 b  200,000
Saturated Water = m
lb
lb
 c  169,300 .
m
hr
hr
26
The required pumping cost is again calculated as follow,
lb
btu
btu
 b  (hb  ha )  169,300  (68.5  67.8)
W pump1  m
 12
.  105
, and
hr
lb
hr
lb
btu
btu
 e  (he  hd )  200,000  (297  2961
W pump 2  m
.)
 18
.  105
.
hr
lb
hr
The total pumping power consumption in this case is about 3.0x105 btu/hr, or ~87.9 kW.
4.3 Different Scenarios of Solution
From the previous discussion, the most optimal injection location is determined in
between the HP and IP turbines; the solar steam can either be injected after the reheater (in point
S67 on Appendix A1) or at before the reheater (in S44). For the condensate return option, two
potential methods are available; either by heating the cold saturated water in the condenser with
an extra boiler, or by mixing the cold water with hot steam extraction from the STG process. And
if the hot steam extraction is selected, the source could be from the solar steam or the hot reheated
steam.
Based on these injection and return options, six different scenarios are built and
evaluated. Beware that all these design scenarios are based on the previously defined design
parameters for the solar steam and return condensate. Other possible scenarios are available (but
not satisfying the criteria), and a baseline case is included as comparison at the end of this
chapter. The six design scenarios are:
Scenario #1: Injection after reheater with return option #1,
Scenario #2: Injection after reheater with return option #2 (and solar steam),
Scenario #3: Injection after reheater with return option #2 (and hot reheat),
Scenario #4: Injection before reheater with return option #1,
27
Scenario #5: Injection before reheater with return option #2 (and solar steam), and
Scenario #6: Injection before reheater with return option #2 (and hot reheat).
4.3.1 Scenario #1: Injection after Reheater with Return Option #1
This design advocates injection of solar steam into the flow after the reheater. The solar
steam is throttled to the intermediate pressure (around 407 psia) and then mix with the hot
reheated steam in a mixing chamber. The steam mixture would pass thru the IP turbine, mix with
some superheated steam from the LPSH, and goes thru the LP turbine to generate power. The exit
flow from the LP turbine condenses in the condenser hot well. And the return condensate is sent
to an additional boiler and get heated to the required target condition before being pumped back
to the CSP field. The main flow would go to the HRSG system and the whole STG process
repeats again. The schematic of design is illustrated in figure 13 below.
A
○
Solar
Steam
B
○
Mixing
Chamber
Hot
Reheat
C
○
E
○
D
○
IP
Turbin
e
To Return
Option #1
F
○
LP
Turbin
e
Seal Steam +
Superheated Steam
Condenser
(CND)
To HRSG
Figure 13: Schematic of the Design Scenario #1
The design performance is based on the improved power output from each affected
turbine after the solar steam injection. Conditions of the steam flow from stage A to D are
available, and the steam turbine efficiency is already calculated in Appendix B1 to B3. Assuming
the isentropic efficiency of each turbine is constant, which is a fair assumption with this level of
28
flow increment; addition of 200,000lb/hr new steam across the IP and LP turbines. Also the
throttling process is assumed to be ideal; pressure is dropped without lowering the flow
temperature. This is not true in reality, but is kept consistent with the rest of the analyses, so this
assumption is fine for the theoretical power improvement analysis being done in follow.
Using these available steam conditions, such as mass flow rate, temperature, pressure,
and enthalpy values, the new power output from each steam turbine can be calculated with (Eq.
4), as shown below.
 d  (hd  he ) & W LP,Turbine  m
 e  (he  h f ) ,
W IP,Turbine  m
 is the steam mass flow rate thru the turbine, and h is the actual enthalpy value (or the
where m
energy stored in the steam flow) across the steam turbine. The steam conditions from stage A to
D are listed in table 5 below. The detail of calculating steam condition at location D is just a
mixing chamber problem, thus is not replicated here. With the flow condition in location D
known, the enthalpies across the IP steam turbine can be calculated with the turbine’s isentropic
efficiency (in Appendix A2) and the steam table in reference [5]. The enthalpies across the LP
turbine can be estimated once the IP turbine outlet condition is found. The detail of finding
enthalpy values across each affected steam turbine is included in the Appendix D1.
Table 5: Flow Conditions from Stage A to D in Scenario #1
Properties
Location A
Location B
Location C
Location D
m (lb/hr)
200,000
200,000
948,370
1,148,370
T (F)
675
664.8
1050.8
982.6
P (psia)
499.7
406.8
406.8
406.8
H (btu/lb)
1343
1343
1550.5
1514.4
29
With the enthalpies across the IP and LP turbines known, the turbine output can be solved easily.
From Appendix D1, the steam mixture flow rate going thru the IP turbine is 1,148,370 lb/hr
(original flow rate is 948,370 lb/hr and increased by 200,000 lb/hr of solar steam addition), with
inlet enthalpy of 1514.4 btu/lb and outlet enthalpy of 1280.6 btu/lb. The rate of work output is,
lb
btu
btu
W IP ,Turbine  m d  (hd  he )  1,148,370  (1514.4  1280.6)
 2.68  10 8
.
hr
lb
hr
The power output from IP turbine is around 2.68x108 btu/hr, or ~78.5 MW.
Similarly, the steam mixture flow rate going thru the LP turbine is 1,272,500 lb/hr
(original flow rate is 1,072,500 lb/hr and increased by 200,000 lb/hr of solar steam addition), with
inlet enthalpy of 1284.3 btu/lb and outlet enthalpy of 1021.9 btu/lb. The LP power output is,
lb
btu
btu
W LP ,Turbine  m e  (he  h f )  1,272,500  (1284.3  1021.9)
 3.34  10 8
.
hr
lb
hr
So the power output from LP turbine is around 3.34x108 btu/hr, or ~97.9 MW.
Compared with the original performance, in which the IP turbine only gives 68.8 MW
and the LP turbine gives around 85.0 MW. The power gain from the solar steam injection is about
22.6 MW. For the additional heating and pumping of the return condensate, the energy costs are
13.3 MW and 111 kW. Therefore, this design scenario would give a net output gain of 9.2 MW.
The performances of this design, along with other scenarios (which are discussed in following
sections), are summarized in the end of this chapter. Thus, only calculations on the improved
power output of each affected turbine is going to be shown in the rest of the design scenarios.
4.3.2 Scenario #2: Injection after Reheater with Return Option #2 & Solar Steam
The fundamental in this design scenario #2 is same the first design, with only difference
on condensate return option. Instead of using an extra boiler to heat the cold condensate, a
fraction of the solar steam is extracted to mix with the cold condensate for the heating
30
requirement on return condensate. The advantage of this design is that no additional boiler is
needed. However, less amount of solar energy is available for STG process since less solar steam
is injected into the STG process (only 164,300 lb/hr instead of the original 200,000 lb/hr). In
figure 14 illustrates the design scenario.
A
○
Solar
Steam
B
○
D
○
Mixing
Chamber
Hot
Reheat
C
○
E
○
IP
Turbin
e
To Return
Option #2
F
○
LP
Turbin
e
Seal Steam +
Superheated Steam
Condenser
(CND)
To HRSG
Figure 14: Schematic of the Design Scenario #2
Much of the information required for power output analysis is already determined in the previous
design. The only difference is the steam mixture condition at location D, since different amount
of solar steam is injected into the STG cycle. However, the method used to find the equilibrium
mixing condition is the same. So the steam flow can be fully defined up to location D. The steam
conditions from stage A to D are listed in table 6 below. In addition, the detail of finding enthalpy
values across each affected steam turbine is included in the Appendix D2.
31
Table 6: Flow Conditions from Stage A to D in Scenario #2
Properties
Location A
Location B
Location C
Location D
m (lb/hr)
164,300
164,300
948,370
1,112,670
T (F)
675
664.8
1050.8
992.9
P (psia)
499.7
406.8
406.8
406.8
H (btu/lb)
1343
1343
1550.5
1519.9
With the enthalpies across the IP and LP turbines known, the turbine output can be solved
like in the previous design. From Appendix D2, the steam mixture flow rate going thru the IP
turbine is 1,112,670 lb/hr (original flow rate is 948,370 lb/hr and increased by 164,300 lb/hr of
solar steam addition), with inlet enthalpy of 1519.9 btu/lb and outlet enthalpy of 1284.4 btu/lb.
The rate of work output is,
lb
btu
btu
 d  (hd  he )  1112
W IP ,Turbine  m
, ,670  (1519.9  1284.4)
 2.62  108
.
hr
lb
hr
The power output from IP turbine is around 2.62x108 btu/hr, or ~76.8 MW.
Similarly, the steam mixture flow rate going thru the LP turbine is 1,236,800 lb/hr
(original flow rate is 1,072,500 lb/hr and increased by 164,300 lb/hr of solar steam addition), with
inlet enthalpy of 1287.8 btu/lb and outlet enthalpy of 1024.0 btu/lb. The turbine power output is,
lb
btu
btu
 e  (he  h f )  1,236,800  (1287.8  1024.0)
W LP ,Turbine  m
 3.26  108
.
hr
lb
hr
So the power output from LP turbine is around 3.26x108 btu/hr, or ~95.5 MW.
4.3.3 Scenario #3: Injection after Reheater with Return Option #2 & Hot Reheat
The scenario #3 is almost the same with scenario #2, except the source of hot steam used
to mix with the cold condensate is different. Instead of the solar steam, the hot reheated steam is
32
used to mix with the cold condensate. This scenario is built and evaluated. The pumping costs of
the two designs are already found to be different as shown in section 4.2.2. The solar energy is
deemed as energy outside the system, while the hot reheated steam is energy inside the plant,
extraction of energy at different parts of the system would affect the usage effectiveness of solar
energy. Thus, the turbine power output is expected to be different than in others and is worthy of
investigation. In figure 15 below the schematic of scenario #3 is shown for reference.
A
○
Solar
Steam
To HRSG
B
○
D
○
Mixing
Chamber
Hot
Reheat
C
○
E
○
IP
Turbin
e
F
○
LP
Turbin
e
Condenser
(CND)
Seal Steam +
Superheated Steam
To Return
Option #2
Figure 15: Schematic of the Design Scenario #3
Similar information of flow conditions from stage A to D could be found from the previous
scenarios. The difference for the steam condition in location D is considered. The steam
conditions from stage A to D are listed in table 7 below. And the detail of enthalpy values
calculation across each the IP and LP steam turbines is included in the Appendix D3.
33
Table 7: Flow Conditions from Stage A to D in Scenario #3
Properties
Location A
Location B
Location C
Location D
m (lb/hr)
200,000
200,000
917,670
1,117,670
T (F)
675
664.8
1050.8
980.8
P (psia)
499.7
406.8
406.8
406.8
H (btu/lb)
1343
1343
1550.5
1513.4
From Appendix D2, the steam mixture flow rate going thru the IP turbine is 1,117,670
lb/hr (hot reheat flow is dropped to 917,670 lb/hr due to extraction and then increased by 200,000
lb/hr of solar steam addition), with inlet enthalpy of 1513.4 btu/lb and outlet enthalpy of 1279.9
btu/lb. The steam turbine power output is,
lb
btu
btu
 d  (hd  he )  1117
W IP ,Turbine  m
, ,670  (1513.4  1279.9)
 2.61  108
.
hr
lb
hr
The power output from IP turbine is around 2.61x108 btu/hr, or ~76.5 MW.
Similarly, the steam mixture flow rate going thru the LP turbine is 1,241,800 lb/hr (the IP
exit flow rate is 1,117670 lb/hr and increased by 124,130 lb/hr of seal steam & superheated steam
addition), with inlet enthalpy of 1283.8 btu/lb and outlet enthalpy of 1021.6 btu/lb. The LP
turbine power output is,
lb
btu
btu
 e  (he  h f )  1,241,800  (12838
W LP ,Turbine  m
.  10216
. )
 3.26  108
.
hr
lb
hr
Therefore, the power output from LP turbine is around 3.26x108 btu/hr, or ~95.5 MW.
4.3.4 Scenario #4: Injection before Reheater with Return Option #1
For this one and the remaining designs, solar steam is injected into the flow before the
reheater. The reheater would then heat the steam mixture back to the regular flow conditions (as
34
in the original state). Since the solar steam would lower the temperature of the regular steam flow
after injection and mixing, therefore, additional fuel is required to burn in the CTG system to
make more flue gases for the steam mixture heating in order to maintain the original flow
conditions across the steam turbines. The amount of required fuel is just enough to heat the steam
mixture back to the original state. And the remaining devices within the HRSG unit would not be
affected by the injection.
This design requires the gas turbine being capable of burning more fuel and intake more
air for combustion. And steam mixture flow conditions across the system should be maintained as
in the original system (except the mass flow rate is now larger due to the solar steam addition).
Otherwise, if no additional fuel is burned and no extra heat is created to compensate the lost
quality in the steam mixture, this design would eventually become the design in scenario #1,
since equal amount of “solar energy” is added into the system in both cases, and the STG would
extract the same amount of added energy into power generation.
Back to the design, after the steam mixture is heated back to the design conditions, it
would pass thru the IP turbine, mix with the superheated steam from the low pressure superheated
(LPSH), and then pass thru the remaining LP turbine before going into the condenser hot well.
The cold condensate would then be pumped to an additional boiler, as discussed in return option
#1, before sent back to the CSP field. The overall process of this design is illustrated in figure 16
below.
35
A
○
B
○
Solar
Steam
C
○
Seal
Steam
E
○
IP
Turbin
e
D
○
Cold
Reheat
Mixing
Chamber
F
○
Reheater
To Return
Option #1
G
○
LP
Turbin
e
Condenser
(CND)
Seal Steam +
Superheated Steam
To HRSG
Figure 16: Schematic of the Design Scenario #4
The improved power output after the solar steam injection could be calculated
accordingly in the previous designs, since the steam flow conditions are already defined in the
system. From stage A to D in figure 16, the solar steam and the cold reheat conditions are known.
The condition at location D is solved by the conservation of energy principle in a mixing
chamber. And then the flow conditions from stage E to G are just the same as in the original
system, with the mass flow rate increased by the amount of injected solar steam. The conditions
data across the IP and LP turbines could be found in Appendix B2 & B3 separately. Thus, the
power output of each steam turbine can be calculated with the flow rate and the enthalpies across
the turbine, as defined below,
 e  (he  h f ) & W LP,Turbine  m
 f  (h f  hg ) ,
W IP,Turbine  m
 is the steam mass flow rate thru the turbine, and h is the actual enthalpy value at the
where m
specified stage. The flow conditions from stage A to E for this scenario are given in table 8
below.
36
Table 8: Flow Conditions from Stage A to E in Scenario #4
Properties
Location A
Location B
Location C
Location D
Location E
m (lb/hr)
200,000
200,000
935,854
1,135,854
1,148,370
T (F)
675
667.9
674
674.1
1050.8
P (psia)
499.7
433.9
433.9
433.9
406.8
H (btu/lb)
1343
1343
1343.4
1343.3
1550.5
With the flow rate and enthalpy data, the turbine power outputs could be calculated. From
Appendix B2, the original steam mass flow rate going thru the IP turbine is 948,370 lb/hr, with
inlet enthalpy of 1550.5 btu/lb and outlet enthalpy of 1303.1 btu/lb. With solar steam addition of
200,000 lb/hr, the new flow rate going thru the IP turbine becomes 1,148,370 lb/hr. So the IP
power output is,
lb
btu
btu
W IP ,Turbine  m e  (he  h f )  1,148,370  (1550.5  1303.1)
 2.84  10 8
.
hr
lb
hr
The power output from IP turbine is around 2.84x108 btu/hr, or ~83.2 MW.
Similarly, the original steam mass flow rate going thru the LP turbine is 1,072,500 lb/hr,
with inlet enthalpy of 1304.9 btu/lb and outlet enthalpy of 1034.4 btu/lb. With the injected solar
steam, the mass flow rate becomes 1,272,500 lb/hr. The rate of work output becomes,
lb
btu
btu
W LP ,Turbine  m f  (h f  hg )  1,272,500  (1304.9  1034.4)
 3.44  10 8
.
hr
lb
hr
The power output from LP turbine is around 3.44x108 btu/hr, or roughly equal to 100.8 MW.
Since the steam mixture is reheated back to the original flow conditions, the energy used
for reheating could be estimated by multiplying the injected solar steam flow rate with the
37
enthalpy difference across the reheater (he – hd). Thus, the reheating energy used in the process is
calculated as below,
lb
btu
btu
 solar _ steam x (he  hd )  200,000  (1550.5  1343.3)
E Re heat  m
 4.14  10 7
.
hr
lb
hr
The total energy required for reheating in this design is around 4.14x107 btu/hr, or ~12.1 MW.
4.3.5 Scenario #5: Injection before Reheater with Return Option #2 & Solar Steam
This scenario is analogous to scenario #4 but with return option #2 and solar steam as hot steam
source. The solar steam is still injected before the reheater, mix with the cold reheat, and reheat
back to the original design condition. The remaining processes are the same as in design #4. This
design has only 164,300 lb/hr of solar steam for injection since some fraction of the solar steam is
extracted for heating the return condensate. See section 4.3.2 for reference of condensate return
description. Figure 17 below illustrates the schematic of this design.
The flow conditions in the system are already defined, as explained in the previous
scenario. The only difference is the amount of solar steam injected into the system. The steam
mixture condition after the mixing process (in location D) should not deviate too much from the
one calculated in scenario #4 because the extracted amount of solar steam is not substantial
enough to create a significant change in the steam’s enthalpy. Therefore, solving the turbine
power outputs is simple with flow conditions from stage A to E shown in table 9 below.
38
A
○
B
○
Solar
Steam
C
○
Seal
Steam
E
○
IP
Turbin
e
D
○
Cold
Reheat
Mixing
Chamber
F
○
Reheater
To Return
Option #2
G
○
LP
Turbin
e
Condenser
(CND)
Seal Steam +
Superheated Steam
To HRSG
Figure 17: Schematic of the Design Scenario #5
Table 9: Flow Conditions from Stage A to E in Scenario #5
Properties
Location A
Location B
Location C
Location D
Location E
m (lb/hr)
164,300
164,300
935,854
1,100,154
1,112,670
T (F)
675
667.9
674
674.1
1050.8
P (psia)
499.7
433.9
433.9
433.9
406.8
H (btu/lb)
1343
1343
1343.4
1343.3
1550.5
Similar to the calculations in previous scenario, the amount of solar steam injected into
the system is 164,300 lb/hr instead of 200,000 lb/hr. Thus, the mass flow rate going thru the IP
turbine becomes 1,112,670 lb/hr, with inlet enthalpy of 1550.5 btu/lb and outlet enthalpy of
1303.1 btu/lb. The IP turbine output is,
lb
btu
btu
 e  (he  h f )  1112
W IP ,Turbine  m
, ,670  (1550.5  13031
.)
 2.75  108
.
hr
lb
hr
The IP turbine output is around 2.75x108 btu/hr, or ~80.6 MW.
39
And to calculate the LP turbine output, the original steam mass flow rate going thru the
LP turbine is 1,072,500 lb/hr and is increased to 1,236,800 lb/hr after the solar steam injection.
Therefore, the LP power output is calculated as,
lb
btu
btu
 f  (h f  hg )  1,236,800  (1304.9  1034.4)
W LP ,Turbine  m
 3.35  108
.
hr
lb
hr
So the LP turbine output from LP turbine is ~3.35x108 btu/hr, or equal to 98.2 MW.
The reheating cost in the process is calculated below,
lb
btu
btu
 solar _ steam x (he  hd )  164,300  (1550.5  1343.3)
E Re heat  m
 3.40  10 7
.
hr
lb
hr
The total energy required for reheating in this design is around 3.40x107 btu/hr, or ~10.0 MW.
4.3.6 Scenario #6: Injection before Reheater with Return Option #2 & Hot Reheat
The scenario #6 is kind of like scenario 3 in nature, but with solar steam injection before
the reheater and flow conditions maintained as in original flow by the reheater. Return option #2
and hot reheat are employed in this case. The flow process in this design is shown in figure 18
below.
A
○
B
○
Solar
Steam
C
○
To HRSG
Seal
Steam
E
○
IP
Turbin
e
D
○
Cold
Reheat
Mixing
Chamber
F
○
Reheater
G
○
LP
Turbin
e
Seal Steam +
Superheated Steam
Figure 18: Schematic of the Design Scenario #6
Condenser
(CND)
To Return
Option #2
40
Flow conditions data is given in table 10 below. It is almost the same with scenario #4,
but the amount of hot reheat is fewer due to the extraction for the condensate return heating.
Table 10: Flow Conditions from Stage A to E in Scenario #6
Properties
Location A
Location B
Location C
Location D
Location E
m (lb/hr)
200,000
200,000
935,854
1,135,854
1,117,670
T (F)
675
667.9
674
674.1
1050.8
P (psia)
499.7
433.9
433.9
433.9
406.8
H (btu/lb)
1343
1343
1343.4
1343.3
1550.5
From Appendix B2, the original hot reheated steam flow rate going to the IP turbine is
948,370 lb/hr. Due to the 200,000 lb/hr solar steam addition, the new steam flow rate is raised to
1,148,370 lb/hr. However, some hot reheat (~30,700 lb/hr) is extracted to mix with the cold
condensate for return purpose. The remaining amount of steam flow rate is adjusted to 1,117,670
lb/hr (as shown in location E in table 10). The inlet and outlet enthalpy values are same as the
previous two scenarios. So the turbine power output can be calculated as,
lb
btu
btu
 e  (he  h f )  1117
W IP ,Turbine  m
, ,670  (1550.5  13031
.)
 2.77  108
.
hr
lb
hr
The power output from IP turbine is about 2.84x108 btu/hr, or ~81.2 MW.
Similarly, the steam flow rate going thru the LP turbine, with the seal steam and super
heated steam from the LPSH added, is found to be 1,241,800 lb/hr (as the difference of 1,272,500
lb/hr – 30,700 lb/hr). The rate of work output is
lb
btu
btu
 f  (h f  hg )  1,241,800  (1304.9  1034.4)
W LP ,Turbine  m
 3.36  108
.
hr
lb
hr
The power output from LP turbine is around 3.36x108 btu/hr, or ~99.1 MW.
41
And for the cost of reheating, it is the same as scenario #4 since the same amount of solar
steam is injected and reheated in the system. So the reheating energy used in the process is
calculated below,
lb
btu
btu
 solar _ steam x (he  hd )  200,000  (1550.5  1343.3)
E Re heat  m
 4.14  10 7
.
hr
lb
hr
The total energy required for reheating in this design is about 4.14x107 btu/hr, or ~12.1 MW.
4.4 Summary of Performance Results
Six different scenarios with solar energy utilization are discussed in section 4.3
previously. The power output improvement are calculated and results are compared with each
other in order to determine the most optimal design. These designs are based on the design
criteria with the specified target conditions of the solar steam and return condensate, provided
from the CSP provider’s contract, with the assumed maximum injection flow rate of 200,000
lb/hr.
From Appendix B2 and B3, the original power outputs of the IP and LP steam turbines
are calculated. See Appendix B for the calculation details. So knowing the original amount of
power generated by each turbine; IP turbine and LP turbine could generate 68.8 MW and 85.0
MW individually; the improved power gain from each design can be found. Also know the
amount of lost capacity in power output during the summer time; roughly 22 MW (as the
difference of the name plated STG output 193 MW and the actual value of 171 MW); the
improved capacity can be estimated as well. In table 11 below, the summary of the performance
results, along with the net gain in power and capacity, for all designs discussed earlier is included.
42
Table 11: Summary of Performance Results for All Design Scenarios
Performance
Results

WIP (MW)
Design #1
Design #2
Design #3
Design #4
Design #5
Design #6
78.5
76.8
76.5
83.2
80.6
81.2
W IP (MW)
9.7
8.0
7.7
14.4
11.8
12.4
W LP (MW)
97.9
95.5
95.5
100.8
98.2
99.1
W LP (MW)
12.9
10.5
10.5
15.8
13.2
14.1
WTotal
22.6
18.5
18.2
30.2
25.0
26.5
W Pump (kW)
(111)
(96.7)
(87.9)
(111)
(96.7)
(87.9)
Extra Boiler?
YES
NO
NO
YES
NO
NO
Q in (MW)
(13.3)
N/A
N/A
(13.3)
N/A
N/A
Extra Fuel?
NO
NO
NO
YES
YES
YES
E Re heat
N/A
N/A-
N/A
(12.1)
(10.0)
(12.1)
9.2
18.4
18.1
4.7
14.9
14.3
4.8%
9.5%
9.4%
2.4%
7.7%
7.4%
(MW)
(MW)
Net Gain
(MW)
Capacity
Gain
Based on the performance results, some observations are made about these design
scenarios. It is obvious that designs with injection after reheater produce better improvement than
those with injection before reheater. This makes sense because reheater has its own thermal
efficiency as well. It serves like a heater in nature and is not isentropic. And reheating is not as
efficient as direct injection and mixing, since direction injection can transfer almost all the energy
into the regular flow. While reheater exchanges heat between the flue gas and the steam flow, and
energy is lost to the surrounding in the process. Also, it is better off to use steam from the system
for heating in the condensate return than with an additional boiler. The total turbine gain of
43
Design #1 or Design #4 is certainly higher than those within the same injection category are.
However, the heating cost from the additional boiler outweighs the gain, and as a result, turns the
designs with return option #1 down.
Nevertheless, all these values are theoretical, with assumption of ideal throttling, heating,
pumping and reheating processes. For instance, as in the analysis shown, when calculating the
pumping power consumption, only the required energy for raising the enthalpy of the fluid from
low pressure to high pressure is included. However, for any real pump, additional energy is
needed to compensate for the friction and vibration losses. Similar reasoning applies to other
devices in the STG process. In reality, the actual power gain from the solar steam injection would
be less, due to the energy loss in piping transmission, turbine’s efficiency fluctuation by variable
flow rate, friction, and throttling. And the cost of power consumption with equipments, such as
pump, boiler and reheater, would become larger since their efficiencies are never ideal.
Therefore, the actual net improvement on power output (and capcity) would be less than the
values shown in table 11.
Based on the defined design criteria, Scenario #2 would give the best improvement
(theoretically) on power output with a net gain of 18.4 MW, about 9.5% of total capacity. The
real output gain will be less than 18.4 MW, but it is still likely to meet the SMUD’s minimally
required improvement of 15 MW, even including all types of energy losses previously discussed
(i.e. throttling, piping transmission, friction, pumping, etc.).
4.5 The Baseline Scenario
All the previous scenarios are based on design parameters with the specified target
conditions for the solar steam and return condensate at the injection flow rate of 200,000 lb/hr,
and these terms are specified on the energy agreement with the CSP provider. There is a problem
44
with all these design. That is the required heating on the return condensate (due to the design
condition), is always deemed as a penalty. Since the condensate leaving the CCPP at the fence
line meter is saturated water at 325 deg. F, and it is at a higher temperature than the only source
of saturated water available for return option in the system. Thus, heating is necessary in order for
the return condensate to meet the specified condition by the agreement. This is already explained
in the previous section.
However, what if the CSP field can take in cold condensate directly from the condenser
hot well (CND) and pump it to the CSP system without heating it to the required condition? As
long as the solar steam condition is maintained at the specified condition (T = 675 de. F, P =
499.7 psia, and h = 1343 btu/lb), this could happen with less than 200,000 lb/hr of saturated water
being pumped from the condenser hot well. At certain selected flow rate, the CSP can heat up the
water to the specified solar steam condition and the steam can be used for injection right away for
power generation in STG. This design would be the simplest scenario of solar energy utilization,
or namely the baseline case, since it requires the least amount system changes; no extra boiler nor
additional fuel is required, and simple injection design. In fact, only pumping power for
transporting fluid in and out of the CSP field are required in this case.
Assuming the CSP system can originally heat up 200,000 lb/hr of water from 325 deg. F
to 675 deg. F (from 297 btu/lb to 1343 btu/lb), the amount of solar energy utilized for heating can
be calculated as,
lb
btu
btu
 flow x (hout  hin )  200,000  (1343  267)
E CSP  m
 2.15  108
.
hr
lb
hr
The amount of solar energy utilized for heating is around 2.15x108 btu/hr. And using this same
amount of energy to heat up water (that is pumped directly from CND) to the target condition, the
adjusted mass flow rate can be found as follow.
45
 adjusted  (ht arg et  hwater ,CND )  2.15  108
m
btu
hr
btu
hr  168,600 lb

btu
hr
(1343  67.8)
lb
2.15  108
 adjusted
m
So if cold water is sent directly to the CSP system and is heated to the target condition for solar
steam, only roughly 168,600 lb/hr of water is needed from the CND.
Using the adjusted rate of solar steam for the STG supplementation process, the steam
can be injected into system either before or after the reheater. Looking into table 11 on previous
section, the designs with injection after the reheater will always produce more net power gain
than those with injection before the reheater of the same condensate return option. Therefore, the
baseline scenario will inject the solar steam after the reheater and pump condensate from the
condenser hot well directly without any heating. The following figure 19 illustrates the baseline
scenario design.
A
○
CSP Field /
Solar Steam
Hot
Reheat
C
○
Mixing
Chamber
B
○
D
○
E
○
IP
Turbin
e
Pump
F
○
LP
Turbin
e
Condenser
(CND)
Seal Steam +
Superheated Steam
Figure 19: The Baseline Design Scenario
To HRSG
46
Most of the information required for power output analysis are already calculated in the
previous sections. The flow conditions from stage A to D are given in table 12 below. And the
actual enthalpy values across each turbine for this case are calculated in Appendix D4.
Table 12: Flow Conditions from Stage A to D in Baseline Scenario
Properties
Location A
Location B
Location C
Location D
m (lb/hr)
168,600
168,600
948,370
1,116,970
T (F)
675
664.8
1050.8
991.6
P (psia)
499.7
406.8
406.8
406.8
H (btu/lb)
1343
1343
1550.5
1519.2
Using the results from Appendix D4, with the mass flow rate going thru the IP turbine
and the enthalpy values across the turbines, the rate of work outputs are calculated as follow,
lb
btu
btu
 d  (hd  he )  1116
W IP ,Turbine  m
, ,970  (1519.2  1283.9)
 2.63  108
, and
hr
lb
hr
lb
btu
btu
 e  (he  h f )  1,241100
W LP ,Turbine  m
,
 (1287.4  1023.7)
 3.27  108
.
hr
lb
hr
The power output from IP turbine is around 2.63x108 btu/hr, or ~77.1 MW. And the power output
from LP turbine is around 3.27x108 btu/hr, or ~95.8 MW.
For the cost of pumping, it is simply the power used for pressure raise from 130.3 psia (at
S80) to 864.7 psia (at CSP inlet). So the pumping energy cost is,
lb
ft 3
btu
 flow  P  v  2168,600  (864.7  130.3) psia  0.01613
W pump  m
 3.71  105
hr
lb
hr
Or W pump =108.7 kW.
47
The net gain of power output for the baseline design is around 19.0 MW, which is
equivalent to 9.8% improvement on the output capacity. Obviously, as compared with the
performance results in table 11, the baseline design provides better power improvement and
utilizes solar thermal energy more efficient.
48
Chapter 5
THE BEST DESIGN
Compared the performance results of all the design scenarios, the design with the best
power improvement seems to be the baseline scenario, with 19.0 MW (or ~9.8% of total capacity)
improvement on power output. Without the energy agreement specified by SMUD and CSP
provider, which requires heating on the return condensate, there is no doubt about this design
being the best. However, based on the specified design criteria, Scenario #2, which has injection
after reheater and with solar steam extracted for condensate heating, would become the most
optimal design and give a theoretical power gain of 18.4 MW (or ~ 9.5% of total capacity).
Although the actual value might be less that 18.4 MW, but the design itself should give the
highest usage efficiency of the solar steam, out of those six usage scenarios.
Once again, the best design is determined by the terms on the final energy agreement
specified by SMUD, and it would be depending on the design conditions of the solar steam, the
return condensate, the injection flow rate. But for this study, based on the solar energy utilization
efficiency, the baseline design will be the recommended solution. For reference, the schematic of
the recommended solution design, which utilizes the baseline scenario, is shown in figure 20
below.
To conclude this study, the performance of other available methods used to improve
power plant output is included. For instance, using absorption chiller that makes use of the
exhaust heat at the gas turbine outlet to cool off the intake air from ambient condition to 15 deg.
C, would provide a raise of CTG capacity by 8 to 13% and STG capacity by 7 to 10%. [12]
Obviously, utilization of solar thermal energy to supplement the steam turbine power generation
49
process in a combined cycle power plant is indeed a viable and sustainable solution comparable
to other available methods.
To HRSG
To Ambient
Reheater
LPSH
HRSG #2
Gas Turbine
(GTD2)
Solar
Steam
CSP
System
Pump
Condensate
Return
Mixing Chamber
HP
Steam
Condenser Hot Well
(CND1)
IP
Steam
LP
Steam
Gas Turbine
(GTD1)
Reheater
LPSH
HRSG #1
To HRSG
Figure 20: Schematic of the Recommended Solution Design
To Ambient
50
APPENDIX A1
Appendix A1. CPP Flow Diagram (Courtesy to SMUD)
CPP Flow Diagram (Courtesy to SMUD)
S116
S115
EVC1
EVC2
S118
S117
S109
S92
S93
M11
S56
GTD2
1799.7 P T 1051.7
V9
S100
420198W H 1512.7
HX1
S55
GTD1
S105
410.09 P T 1051.9
S98
S89
815780W H 1353.5
439.27 P T 688.19
S97
467927W H 1551.1
M9
S96
SP7
V7
V8
S69
TMX3
S2
S58
SEAL2
S64
406.75 P T 1051.8
S63
SEAL1
S66
935854 W H 1551.1
1787.2 P T 1051.1
S68
HPTURB
840396W H 1512.7
SP10
S104
S110
V13
S44
S45
S87
S46
S42
S65
HPEVAP
M2
S70
S6
S59
S88
M3
IPSH2
S32 S43
V3
436.87 P T 687.93
S86
IPTURB
S40
S39
S8
S53
M4
HPEC2
57716W H 1315.5
61.39 P T 566.98
S114
LPSH2
S41
V4
S60
V11
407890W H 1353.5
S67
V15
S7
V5
S108
58121W H 1308.1
434.13 P T 607.76
433.87 P T 687.60
TMX2
S47
HPSH1
S5
407890W H 1353.5
S48
REHR1
S4
1811.1 P T 1052.2
S50
S3
420198W H 1512.7
V10
V2
S49
HPSH2
1959.7 P T 303.66
3084.8W H 276.79
S103
S102
M8
S9
S84
S54
S91
57.28 P T 566.34
V14
57716W H 1315.6
S95
56.54 P T 541.90
S71
S72
957078W H 1303.6
SEAL3
S85
SSR1
S90
S82
56.57 P T 566.23
S83
HPECN1
S73
S16
S61
M1
S30
HRSG #2
S75
S78
S76
S17
PUMP3
Dry Bulb
495.41 MW
104.00 F
6134.6 Btu/kWh LHV
Net Output
Net Heat Rate
CTG Unit Output 159.20 MW
189346 kW
12340 kW
STG Output
BOP Aux Load
0.99 P T 101.20
S18
S107
S77
SP6
S24
S74
CND1
1.07E6W H 1034.4
S25 V16
HRSG #1
S26
LPDAEV
S28 SP1
S27
PUMP2
S29
30.00 P T 80.00
S112
GLD1
CT1
PUMP5
S111
S81
576529 W H 69.82
128.34 P T 101.64
6.90E7W H 47.97
S113
30.00 P T 95.05
S80
6.90E7W H 62.99
PUMP4
192.29 P T 280.90
S21
TMX1
S19
155216W H 250.18
PUMP1
S23S22
PRHTR
S79
128.34 P T 101.64
MU1
M7
576529 W H 69.82
15.00 P T 60.01
S35
4723.3W H 27.94
56.54 P T 544.52
LPTURB
V6
72439W H 48.04
455.33 P T 78.91
S62
S34
SP4
SP3
SP8
1.07E6W H 1304.9
115432W H 1315.5
M6
M5
SP5
S12
S14
IPEC S31
94631W H 271.34
524.70 P T 300.93
S15
1959.7 P T 303.66
S38
S11
SP2
455.69 P T 436.70
S37
S10
LPSH1
419199W H 276.79
S36
IPSH1
IPEVAP
36220 W H 415.17
S33
S13
SMUD - Cosumnes Power Plant, Performance, 1/15/03
M10
S106
S99
V12
S101
V1
S51
S52
REHR2
467927W H 1551.1
413.04 P T 1052.1
SP9
S57
S1
SMUD Provided CT Heat Rate
524.70 P T 300.93
1916.2W H 271.34
Choked LP Flow - to lower Pipe Velocity in LP piping.
1 x 1 - 100% CT load cases, HPST floor pressure limit set
S94
S20
51
APPENDIX B1
Appendix B1. Original HP Steam Turbine Performance Analysis
Original HP Steam Turbine Performance Analysis
52
APPENDIX B2
Appendix B2. Original IP Steam Turbine Performance Analysis
Original IP Steam Turbine Performance Analysis
53
APPENDIX B3
Appendix B3. Original LP Steam Turbine Performance Analysis
Original LP Steam Turbine Performance Analysis
54
APPENDIX C1
Appendix C1. Properties of Solar Steam at Different Pressures
Properties of Solar Steam at Different Pressures
55
APPENDIX D1
Appendix D1. Flow Properties Analysis across IP & LP Turbines in Scenario #1
Flow Properties Analysis across IP & LP Turbines in Scenario #1
56
APPENDIX D2
Appendix D2. Flow Properties Analysis across IP & LP Turbines in Scenario #2
Flow Properties Analysis across IP & LP Turbines in Scenario #2
57
APPENDIX D3
Appendix D3. Flow Properties Analysis across IP & LP Turbines in Scenario #3
Flow Properties Analysis across IP & LP Turbines in Scenario #3
58
APPENDIX D4
Appendix D4. Flow Properties Analysis across IP & LP Turbines in Baseline
Flow Properties Analysis across IP & LP Turbines in Baseline
59
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