i

THE EFFECT OF INTEGRATING SOLAR ENERGY ON THE

RANKINE CYCLE IN THE COMBINED CYCLE POWER

PLANT TO THE STEAM TURBINE OUTPUT USING A

PARABOLIC TROUGH SOLAR COLLECTOR

A final project report

presented to

the Faculty of Engineering

By

Gusti Armando Ginting

003201600011

in partial fulfillment

of the requirements of the degree

Bachelor of Engineering in Mechanical Engineering

President University

January 2020

i

DECLARATION OF ORIGINALITY

I declare that this final project report, entitled “THE EFFECT OF INTEGRATING SOLAR

ENERGY ON THE RAKINE CYCLE IN THE COMBINED CYCLE POWER PLANT TO

THE STEAM TURBINE OUTPUT USING A PARABOLIC TROUGH SOLAR

COLLECTOR” is my own piece of work and, to the best of my knowledge and belief, has not

been submitted, either in the whole or in the part, to another university to obtain a degree. All

sources that are quoted or referred to are truly declared.

Cikarang, Indonesia, 17th January 2020

Gusti Armando Ginting

ii

THE EFFECT OF INTEGRATING SOLAR ENERGY ON THE

RANKINE CYCLE IN THE COMBINED CYCLE POWER

PLANT TO THE STEAM TURBINE OUTPUT USING A

PARABOLIC TROUGH SOLAR COLLECTOR

By

Gusti Armando Ginting

003201600011

Approved by

Prof. Tohru Suwa, Ph.D. Lydia Anggraini, ST., M.Eng., Ph.D.

Final Project Supervisor Head of Study Program

Mechanical Engineering

iii

APPROVAL FOR SCIENTIFIC PUBLICATION

I hereby, for the purpose of development of science and technology, certify and approve to give

President University a non-exclusive royalty-free right upon my final project report with the title:

THE EFFECT OF INTEGRATING SOLAR ENERGY ON THE RANKINE CYCLE IN

THE COMBINED CYCLE POWER PLANT TO THE STEAM TURBINE OUTPUT

USING A PARABOLIC TROUGH SOLAR COLLECTOR

along with the related software or hardware prototype (if needed). With this non-exclusive royalty-

free right, President University is entitled to conserve, to convert, to manage in a database, to

maintain, and to publish my final project report. These are to be done with the obligation from

President University to mention my name as the copyright owner of my final project report.

Cikarang, 17th January 2020

Gusti Armando Ginting

003201600011

iv

ABSTRACT

Solar energy is a source of energy that is clean, pollution-free and will never run out. The energy

can be used to substitute the petroleum produced energy for different purposes. In this project,

solar energy is integrated into a combined cycle power plant or commonly referred to as the

Integrated Solar Combined Cycle (ISCC). In solar fields, parabolic troughs are used as solar

collectors which stretch from north to south and move in the direction of the sun from east to west.

In the process of integration, there are two technologies commonly used in the configuration of

this power plant, Direct Steam Generator (DSG) and also Heat Transfer Fluid (HTF). What

distinguishes these two configurations is the DSG system, the steam coming from the feed water

pump is directly flowed into the solar field so there is no need for heat exchangers like those in

HTF technology. By calculating the mass flow rate of water, it can be determined how much

additional capacity of a steam turbine, and the efficiency of the plant. In this study, the results

obtained where by using DSG technology, the capacity of a steam turbine has more than three

times with same efficiency. Whereas in HTF technology, the capacity of a steam turbine is slightly

lower than a DSG.

Keywords: ISCC, Rankine cycle, Parabolic Trough Collector, HTF, DSG

v

ACKNOWLEDGEMENT

First of all, I praise and thank you for the presence of God Almighty, the Holy Trinity, Father, Son,

and Holy Spirit for the blessings that have been given, so that this thesis can be completed with

the guidance and inclusion from Him. I am also grateful for the people around me who are very

supportive and always provide encouragement in the completion of this thesis:

1. My deep gratitude for my mother, Ernawaty br Sembiring. And my beloved brothers,

Fernando Ginting, Evrando Ginting, and Risky Rinaldo Ginting. Thank you for always

being there when I am at my lowest point and need more energy. Thank you for always

trust me, love me, supporting me, and all the prayers that you give to me.

2. Prof. Tohru Suwa, Ph.D., as my thesis advisor. Thank you for your time, suggestions,

advices, patience, and support for me for completing this thesis.

3. Mrs. Lydia Anggraini, ST., M.eng., Ph.D. as the head of Mechanical Engineering study

program. Thank you for all kindness, help, patience, and all the knowledge that you have

shared to me.

4. All lectures of Mechanical Engineering Study Program, staff of Mechanical Engineering.

Thank you for every knowledge, moments, and assists that you gave to me so I can finish

this thesis

5. My comrade-in-arms and my permanent classmates of Mechanical Engineering 2016,

Adelya, Aprilia, Annisa, Haryo, Liwiryon, Lutfi, Quinn, Mahendra, Revi, and Azhar.

Thank you for all the moments, see you on top guys!

6. All my PUCatSo family. Thank you for coloring my university life with so many moments.

Especially for the PUCatSo batch 2016, thank you for being my little family in my

university life.

Finally thank you for everyone else that supports me during this research that I have

not mentioned above. I hope that this research could inspire the readers and useful for

betterment of education in President University. I also realize that there are some weakness

and mistakes of this thesis. Therefore, I accept all suggestions to improve my

understanding.

vi

TABLE OF CONTENTS

DECLARATION OF ORIGINALITY ............................................................................................ i

APPROVAL FOR SCIENTIFIC PUBLICATION ....................................................................... iii

ABSTRACT ................................................................................................................................... iv

ACKNOWLEDGEMENT .............................................................................................................. v

TABLE OF CONTENTS ............................................................................................................... vi

LIST OF FIGURES ..................................................................................................................... viii

LIST OF TABLES ......................................................................................................................... ix

NOMENCLATURE ....................................................................................................................... x

CHAPTER I .................................................................................................................................... 1

INTRODUCTION .......................................................................................................................... 1

1.1. Background ...................................................................................................................... 1

1.2. Identification of the Problem............................................................................................ 2

1.3. Objective of the Study ...................................................................................................... 2

1.4. Benefits............................................................................................................................. 3

1.5. Thesis Writing Systematics .............................................................................................. 3

CHAPTER 2 ................................................................................................................................... 5

LITERATURE REVIEW ............................................................................................................... 5

2.1. Combined Cycle Power Plant .......................................................................................... 5

2.2. Integrated Solar Combined Cycle (ISCC) ........................................................................ 6

2.3. Parabolic Trough Solar Collector ..................................................................................... 7

2.4. Heat Transfer Fluid (HTF) Technology ........................................................................... 8

2.5. Direct Steam Generation (DSG) .................................................................................... 11

2.6. Design of The Solar Integration in HRSG ..................................................................... 12

CHAPTER 3 ................................................................................................................................. 14

RESEARCH SCHEME AND ANALYSIS .................................................................................. 14

3.1. Initial Observation .......................................................................................................... 15

3.2. Data Collection ............................................................................................................... 15

3.3. Data Analysis ................................................................................................................. 17

3.3.1. Gas Turbine Engine ................................................................................................ 17

vii

3.3.2. Heat Balance in HRSG ........................................................................................... 18

3.3.3. Rankine Cycle ......................................................................................................... 20

3.3.4. HTF pumping power ............................................................................................... 21

3.3.5. DSG pumping power .............................................................................................. 22

3.3.6. Parabolic trough efficiency and solar collector area ............................................... 23

3.3.7. Overall thermal efficiency of the CC and ISCC ..................................................... 24

CHAPTER 4 ................................................................................................................................. 26

RESULT AND DISCUSSIONS ................................................................................................... 26

4.1. Current Situation ............................................................................................................ 26

4.1.1. Gas turbine engine analysis .................................................................................... 26

4.1.2. HRSG heat balance ................................................................................................. 27

4.1.3. Ideal Rankine cycle heat balance ............................................................................ 28

4.1.4. Combined cycle thermal efficiency ........................................................................ 30

4.1.5. Current condition heat capacity .............................................................................. 31

4.2. Analysis of Implementing Solar Field ........................................................................... 33

4.2.1. Direct Steam Generator (DSG) effect ..................................................................... 33

4.2.2. Heat Transfer Fluid (HTF) effect ............................................................................ 35

4.3. Comparison of Simple Combined Cycle and ISCC ....................................................... 38

4.3.1. Combined cycle and DSG configuration ................................................................ 38

4.3.2. Combined cycle and HTF configuration ................................................................. 39

4.3.3. DSG and HTF configuration ................................................................................... 40

CHAPTER 5 ................................................................................................................................. 42

CONCLUSIONS AND RECOMMENDATION ......................................................................... 42

5.1. Conclusions .................................................................................................................... 42

5.2. Recommendation ............................................................................................................ 43

REFERENCES ............................................................................................................................. 44

viii

LIST OF FIGURES

Figure 2.1: Combined Cycle Power Plant flow diagram ................................................................ 5

Figure 2.2: Integrated Solar Combined Cycle Diagram ................................................................. 6

Figure 2.3: Parabolic Trough Solar Collector ................................................................................. 7

Figure 2.4: HTF-configuration scheme........................................................................................... 8

Figure 2.5: Direct Steam Generator Scheme ................................................................................ 11

Figure 2.6: HTF-ISCC Design ...................................................................................................... 12

Figure 2.7: DGS-ISCC Design ..................................................................................................... 12

Figure 3.1: T-s diagram of Brayton cycle ..................................................................................... 15

Figure 3.2: Rankine Cycle Flow Chart ......................................................................................... 16

Figure 3.3: Open cycle gas-turbine engine ................................................................................... 17

Figure 3.4: HRSG layout .............................................................................................................. 18

Figure 3.5: T-s diagram of Rankine Cycle.................................................................................... 20

Figure 3.6: DSG configuration ..................................................................................................... 22

Figure 3.7: Thermal efficiency of parabolic trough solar collector .............................................. 23

Figure 3.8: Combined Cycle scheme ............................................................................................ 24

Figure 3.9: ISCC scheme .............................................................................................................. 25

Figure 4.1: Gas turbine engine T-s diagram ................................................................................. 26

Figure 4.1: Current condition T-h diagram ................................................................................... 32

Figure 4.2: DSG-Configuration T-h diagram ............................................................................... 33

Figure 4.3: HTF-Configuration T-h diagram ................................................................................ 35

Figure 4.4: Steam Turbine working principal ............................................................................... 38

ix

LIST OF TABLES

Table 2.1 Advantages and disadvantages of working fluid compared to thermal oil ................... 10

Table 3.1 Gas Turbine Operation Condition................................................................................. 15

Table 3.2 Rankine Cycle Operation Condition ............................................................................. 16

Table 3.3 Condenser Cooling Water Operation Condition ........................................................... 17

Table 4.1 Enthalpy for the steam line ........................................................................................... 31

Table 4.2 Enthalpy for the gas line ............................................................................................... 31

Table 4.3 Comparison between DSG and HTF technologies ....................................................... 40

x

NOMENCLATURE

Symbol Description Unit

mw Water mass flow rate kg/s

ma Air mass flow rate kg/s

mo Oil mass flow rate kg/s

qin Heat inlet kJ/kg

qout Heat outlet kJ/kg

Qsf Solar field heat capacity kW

h Enthalpy kJ/kg

s Entropy kJ/kg.K

𝜂th Thermal efficiency

T Temperature °C

Wturb,DSG Turbine work with DSG configuration kW

Wturb,HTF Turbine work with HTF configuration kW

DNI Direct Normal Irradiation W/m2

Asf Solar field area m2

1

CHAPTER I

INTRODUCTION

1.1. Background

The need for natural gas as the main fuel to produce electricity in power plants is

currently one of the major concerns in the energy sector because natural gas is one of the

non-renewable energy sources. Solar energy is a source of renewable, pollution-free energy

that will never run out. The energy can be used for various purposes to replace the

petroleum-generated energy. Indonesia is a very strategic country to do various things with

its agricultural natural resources and is located between 6 °NL-11 ° SL latitudes and

crossed by the equator so that Indonesia's earth gets solar energy throughout the year and

has the potential to be used in terms of power generation. [1]

In the power generation sector, the use of solar power as additional energy using

solar collectors in solar fields has been widely introduced. Solar power that is integrated

together with combined cycle power plants is one solution in this sector. The combination

of a concentrated solar plant with a combined cycle power plant is known as the Integrated

Solar Combined Cycle. This technology has the potential to increase the capacity of power

plants or to reduce operational costs for power plants. Solar power is used to produce

additional steam as a driving force for steam turbines in the combination cycle. With the

integration with power, operational flexibility will be greater than that of a combined cycle

standalone power plant [2].

Parabolic trough solar thermal power plants focus solar radiation and increase their

temperature on linear receivers located in the parabolic focal lines and those flowing

through heat transfer fluid (HTF). The heat transfer liquid is closely associated with the

solar field operating temperature, ranging from simple demineralized water to synthetic

oil. To paralyze the collection of solar radiation, the parabolic trough collector moves all

day after the position of the movement of the sun, usually around a parallel axis in each

collector's focal line. This is very important because only direct normal irradiation (DNI)

can be obtained by the solar concentrator. DNI is the amount of solar radiation to a unit of

surface area that is perpendicular to sunlight coming in a straight line from the sun's

2

position. By keeping the position of the solar collector maintained, the amount of radiation

obtained can be maximized. Parabolic trough technology shows demonstrated commercial

success, with initial experience gained in the 1980s, and the design and implementation of

such technology is particularly advanced than other solar thermal systems.

At present, there are several ISCCs that have operated throughout the world. In the

United States, the largest ISCC was inaugurated in December 2010. Three power plant

already operated North Africa (Egypt, Algeria and Morocco). In Iran, ISCC Yazd with a

capacity of 467 MW was installed in August 2010, and in Italy the ISCC Archimede in

Priolo Gargallo has been operating since July 2010 [3].

In integrating with solar power, there are two common configurations, namely

Direct Steam Generation (DSG) and Heat Transfer Fluid (HTF) technology. In the DSG

configuration, water from the condenser is directed directly to the solar collector to be

raised in temperature and converted to steam. This steam is then flowed into the turbine

steam. Whereas in the HTF configuration, there is a working fluid which acts to collect

heat from solar irradiance. This working fluid is then flowed to the heat exchanger to heat

the feed water from the condenser and then flowed to the HRSG for further heating. The

working fluid used is usually thermal oil. But at this time other working fluids have been

developed including molten salt, water/steam, and also gas (CO2).

1.2. Identification of the Problem

In this study, analysis for the steam turbine capacity increases will be calculated

with the integration of solar power into a combined cycle power plant. Here, the solar field

will be installed in the Rankine of combined cycle power plant cycle. As well as a

comparison of the effectiveness between DSG and HTF with thermal oil as the working

fluid.

1.3. Objective of the Study

In this study, the objectives are:

1. To know the effects of implementing solar field on the Rankine Cycle

2. Knowing the additional power that occurs through integration with solar energy in

the steam turbine

3

3. To know the area needed for the solar field

4. Comparing the Direct Steam Generator and Heat Transfer Fluid system

1.4. Benefits

Based on the above objectives, it is expected that this research can provide benefits

to other parties, such as: providing information and education related to hybrid power plant

and provide recommendations of the renewable energy implementation in power plant.

1.5. Thesis Writing Systematics

This thesis divided into three parts, namely: the preliminary matter of the thesis,

the content of the thesis, and the final part of the thesis.

The thesis preliminary matter consists of title page, declaration of originality,

approval, abstract, table of contents, list of figures, list of tables, and preface.

The content of the thesis consists of five chapters arranged with the following

systematics:

Chapter 1 Introduction

This chapter consists of problem background, problem statements as the

things to be solved, objectives to be achieved in this research and the

benefits, scopes of the limitation of the research, and research outline of the

study.

Chapter 2 Literature Review

This chapter provides data that supports the analysis of the research such as

Natural Gas Combined Cycle (NGCC) power plant, the cycle at the NGCC

power plant, Integrated Solar Combined Cycle, Direct Steam Generator

(DSG), and Heat Transfer Fluid (HTF)

4

Chapter 3 Scheme and Analysis

This chapter describes the flow of this research and explanation of each step

to conduct this research start from initial observation until analyze the

collected data which comes up with an improvement and recommendation

Chapter 4 Results and Discussions

In this chapter the result from the project can be found. The integration of

the solar into combine cycle effect are calculated and the capacity of the

steam turbine are defined using each configuration both HTF and DSG

configuration.

Chapter 5 Conclusion and Recommendations

Here, the conclusion and the recommendation of this project are mentioned.

Also, the comparison results between the HTF and DSG are shown.

5

CHAPTER 2

LITERATURE REVIEW

In this study, solar power integration to the combined cycle power plant uses two types of

configurations, namely Direct Steam Generation and Heat Transfer Fluid. Comparison of these

two configurations is the main point in the discussion of this study.

2.1. Combined Cycle Power Plant

In the power industry, Combined Cycle Power Plants (CCPP) is a common plant in

the electricity generation. Basically, a combined cycle power plant is a power plant that

combines a gas turbine system with a steam turbine that is connected through a Heat

Recovery Steam Generator (HRSG) to produce greater thermal efficiency compared to a

standalone power plant. The results of the exhaust gas from the turbine gas with a high

enough temperature flowed into HRSG which is then used to heat water from the condenser

to the superheated point and then flow into the steam turbine to produce electricity. [4]

Figure 2.1: Combined Cycle Power Plant flow diagram

Figure 2.1 is a simple combined cycle system representation. Air is compressed by

the compressor and then flowed into the combustion chamber which is then used to turn

the gas turbine (the Brayton cycle). The exhaust gas from the turbine gas is then flowed to

the HRSG which is used to raise the temperature of the feed water. Superheated steam that

has passed through HRSG is then flowed into a steam turbine to produce electricity. Steam

6

that comes out of the steam turbine flows into the condenser and goes back to HRSG

(Rankine cycle).

Although the combined cycle achieved higher thermal efficiency compared to the

gas-turbine engine or Rankine cycle, carbon dioxide emission is unavoidable.

2.2. Integrated Solar Combined Cycle (ISCC)

Integrated Solar Combined Cycle is a technology that integrates a solar thermal

collector into a combined cycle plant. Solar energy is used as a part of heat source in ISCC

plants to support the steam cycle, resulting in increased generation capacity or decreased

fossil fuel usage. [5]

The diurnal nature of solar resources makes it necessary to have a storage system

in a Concentrated Solar Power (CSP) system. This problem can be solved by integrating

CSP with the combined cycle power plant and also provides a reduction in operational and

capital costs, as well as the possibility of operational flexibility compared to running alone

combined cycle power plant [2].

Figure 2.2: Integrated Solar Combined Cycle Diagram

Steam produced through the ISCC system can be used for two alternatives, namely

to increase the power output of the power plant by maintaining the same fuel input during

7

the day (power boost) or also to save fuel by using energy storage that can still be used at

night day so that it produces the same power output. And to increase the steam produced,

the ISCC with Rankine cycle will be applied as shown in the figure 2.6 above.

To produce electricity, there are several components that must be included in the

CSP system, including concentrators, receivers, storage systems, and also heat exchanger

device. By utilizing solar energy, the carbon dioxide emission is reduced for the same

electricity output comparing to a simple combined cycle.

2.3. Parabolic Trough Solar Collector

Basically, a parabolic trough is a long, curved mirror to focus the heat from solar

energy. The solar energy is concentrated on absorber tube surface. Using a motorized

device, the mirror tracks the sun position. Typically, a parabolic reflector is made of silver

mirror with a thickness of about 4-5 mm. On the other hand, polished metal, plastic film,

and also thin glass are used as reflectors. Figure 2.7 shows a picture of the collector. [6]

Figure 2.3: Parabolic Trough Solar Collector

In the mirror’s focus line, there is a steel receiver tube. The steel tube outside surface

is painted with special coatings to improve the absorption of energy and reduce heat losses.

A heat transfer working fluid absorb heat from focused sunlight that flowing through the

tube. The metal tube wrapped in a glass tube. To decrease heat losses, the gap among the

glass tube and the absorber is retained under vacuum. A metal-based support structure kept

the collector in a precise position to absorb maximum solar irradiation.

8

Many collectors are connected as a loop. To generate the heat needed to bring HTF

(heat transfer fluid) to the maximum temperature expected, many loops are required. With

good solar radiation, an area of approximately 4-5 hectares is needed to produce a capacity

of 1 MW, followed by a location with good solar radiation.

In the case of heat transfer fluid technology, HTF heat is transmitted to a heat

exchanger which usually called as Heat Solar Steam Generator (HSSG) where at the first

part, HTF heat is transmitted to water to be turned to steam and then send to steam in the

next part a producing superheated steam. Henceforth, power blocks converting steam into

electricity include traditional components: steam turbines, heat sinks, feed heaters,

condenser and boilers [6]. Power block in the solar thermal power generation is a power

generation unit that supplying energy to the power plants [7].

2.4. Heat Transfer Fluid (HTF) Technology

Conceptually, HTF technology collects solar energy through solar fields and

collects DNI (Direct Normal Irradiance) into absorbing pipes where in this pipe solar

energy is converted into thermal energy. This hot fluid is then transferred to a heat

exchanger (HSSG) that connecting Rankine cycle of the combined cycle and the solar

field[3]. Figure below showing the scheme of the HTF technology.

Figure 2.4: HTF-configuration scheme

9

Usually, thermal oil used as the heat transfer fluid. But thermal oil has a

disadvantage which maximum temperature obtained using HTF is 400 °C. This results in

limited efficiency of the steam turbine because temperature of the superheated steam

supplied to the turbine is not greater than 390 °C [8]. Currently some working fluid has

been used to replace thermal oil as a working fluid, including molten salt (mixture of KNO3

and NaNO3), water/steam, and gas (CO2) [9]. Table below showing the advantages and

disadvantages of each working fluid

10

Table 2.1 Advantages and disadvantages of working fluid compared to

thermal oil

Fluid Advantages over thermal oil Disadvantages over thermal oil

Molten Salt • More efficient heat

storage

• Higher working

temperature

• No pollution or fire

hazards

• Higher thermal losses at

night

• Complex solar field design

• High electricity

consumption

Water/steam • Simple plant design

• Higher working

temperature

• No pollution or fire

hazard

• Minimum of suitable

storage system

• Complex solar field

control

• Higher pressure on the

solar field

Gas • Higher steam

temperature

• Thermal storage

enhancement

• No pollution or fire

hazard

• Poor heat transfer in the

receiver tubes

• Complex solar field

control

• Higher pressure on the

solar field

11

2.5. Direct Steam Generation (DSG)

In the Rankine cycle in the DSG configuration, water is used as working fluid that

flows through a parabolic trough and is directed as superheated steam to the steam turbine

as shown in the Figure 2.5.

Figure 2.5: Direct Steam Generator Scheme

One of the ISCC drawbacks at the moment is the lack of knowledge about Thermal

Energy Storage. The storage media currently still use liquid salt where the storage

temperature must be kept constant above 240 °C. This is because molten salt will be

solidified at room temperature. This means that at night with low temperatures, additional

energy is needed to maintain the temperature in the storage room. Technological

developments in DSG can be a breakthrough because there is no need for heat exchangers

at the plant, this can increase the efficiency of the plant itself. [10].

Although both DSG and HTF technologies are applicable to integrated solar

combined cycles, the difference in the thermodynamics performances of these two

technologies is not well known. At the same time, the methodology to identify how to

determine the total HRSG output heat of the combined solar integrated cycle has never

been documented. In this thesis, the thermodynamic performance of the combined cycle

utilizes the two working fluid technologies are compared. While analyzing the

thermodynamics performance, the methodology to identify the output power of the

integrated solar combined cycle is discussed.

12

2.6. Design of The Solar Integration in HRSG

Here are two types of designs that will be used in integrating solar in the Rankine

cycle. The first is using a Heat Transfer Fluid that comes with the Heat Solar Steam

Generator (HSSG) and the next is the concept of the Direct Steam Generator where there

is no heat exchanger needed, so the steam flows directly to the solar field.

Figure 2.6: HTF-ISCC Design

From Figure 2.6 it can be seen that the water pumped through the condenser enters

first at the end of the HRSG and then flows into the HSSG which will be raised in

temperature by a heat exchanger from the HTF. After going through HSSG the steam will

flow back to HRSG where there is a result of exhaust from the turbine gas and then flow

into the steam turbine to conduct electricity.

Figure 2.7: DGS-ISCC Design

13

In the DSG-ISCC there is no heat exchanger like that of the HTF-ISCC system. In

this flow, the water coming from the condenser is channeled to the low temperature HRSG

and then directly flowed to the solar field to be raised in temperature and then flowed back

to the high temperature HRSG.

14

CHAPTER 3

RESEARCH SCHEME AND ANALYSIS

Start

Initial Observation

Problem Identification

Study Literature

Problem Scope

Data Collection

Data Analysis

Result

Finish

Yes

No

15

3.1. Initial Observation

This observation was carried out at one of the power plants in Jababeka, Cikarang.

This power plant uses a combined cycle by integrating 2 gas turbines and 1 steam turbine.

Natural gas is used as a fuel supplier to supply gas turbines. Bekasi Power Plant was built

to guarantee the supply of Uninterruptible Power Supply (UPS) to seven industrial areas

including Jababeka. By providing reliable electricity supply at competitive prices, the plant

has a positive effect on the Jababeka business district and surrounding areas, also,

increasing Indonesia's electricity capacity and industrial growth in the coming years.

3.2. Data Collection

The table below is an operational data in the Bekasi Power (CCPP). This data was

taken when the air temperature at 34 °C and air pressure at 100.8 kPa.

Table 3.1 Gas Turbine Operation Condition

Location T Unit P Unit Reference

1 35 °C 100.8 kPa Compressor inlet

2 366 °C 1050.8 kPa Compressor exit

3 1040 °C 1050.8 kPa Turbine inlet

4 560 °C 103.7 kPa Turbine exit

5 150 °C 101.8 kPa HRSG exit

Figure 3.1: T-s diagram of Brayton cycle

16

Item Value Unit

Air flow rate 111.1 kg/s

Fuel flow rate 2.3 kg/s

Electronic power 32.9 MW

Table 3.2 Rankine Cycle Operation Condition

Location T Unit P Unit Reference

1 45 °C 10.8 kPa Condenser exit

2 45 °C 12100.8 kPa Pump exit

3 45 °C 12100.8 kPa HRSG inlet

4 538 °C 8900.8 kPa HRSG exit

5 525 °C 8300.8 kPa Turbine inlet

6 45 °C 10.8 kPa Turbine exit

Figure 3.2: Rankine Cycle Flow Chart

17

Item Value Unit

Steam flow rate 22.51 kg/s

Electronic power 46.2 MW

Table 3.3 Condenser Cooling Water Operation Condition

Location T Unit P Unit Reference

Inlet 30 °C 2 bar Condenser cooling water inlet

Outlet 36 °C 1.5 bar Condenser cooling water outlet

Cooling water flow rate

3.3. Data Analysis

3.3.1. Gas Turbine Engine

Figure 3.3: Open cycle gas-turbine engine

At the gas turbine engine, the heat inlet (Qin) can be determined from the enthalpy

difference at the boiler as:

Qin = h3 − h2 (equation 1)

and the work output of the gas turbine (Wturb, out) and compressor work input can be

obtained from following equation:

Wturb,out = h4 − h3 (equation 2)

Wcomp,in = h2 − h1 (equation 3)

Cooling water 12000 m3/h 3320 kg/s

18

3.3.2. Heat Balance in HRSG

HRSG is a unitary component between gas turbine and steam turbine in a

combined cycle power plant system. HRSG utilizes gas turbine exhaust gas to heat

water in the pipes inside HRSG until it becomes superheated steam that is capable

of turning the steam turbine.

Figure 3.4: HRSG layout

As shown in figure 3.4, water drives into HRSG and heated by exhaust gases

from the combustion of Brayton cycle and flowed into the steam turbine to produce

electricity. To obtain a heat balance in HRSG there are several things that must be

considered, including:

Water side:

(𝑞𝑖𝑛 − 𝑞𝑜𝑢𝑡) − (𝑊𝑜𝑢𝑡 − 𝑊𝑖𝑛) = 𝑚��(ℎ2 − ℎ1) + ∆𝑘𝑒 + ∆𝑃𝑒

As at the water side there are no heat out, work out, work in, kinetic energy and

potential energy, so we conclude that:

𝑞𝑖𝑛 = 𝑚��(ℎ2 − ℎ1) (equation 4)

where,

qin = heat gained by water (kW)

19

��w = mass flow rate of water (kg/s)

h1 = inlet enthalpy of water (kJ/kg)

h2 = outlet enthalpy of steam (kJ/kg)

From the exhaust gas:

(𝑞𝑖𝑛 − 𝑞𝑜𝑢𝑡) − (𝑊𝑜𝑢𝑡 − 𝑊𝑖𝑛) = 𝑚��(ℎ4 − ℎ3) + ∆𝑘𝑒 + ∆𝑃𝑒

At the exhaust gas side there are no heat inlet, work inlet, and outlet work. Also, in

this case the kinetic and the potential energy are neglected, so it can be found that:

𝑞𝑜𝑢𝑡 = 𝑚𝑎 (ℎ4 − ℎ3) (equation 5)

where:

qout = heat loses by gas-turbine exhaust (kW)

��a = mass flow rate of air (kg/s)

h3 = inlet enthalpy of gas (kJ/kg)

h4 = outlet enthalpy of gas (kJ/kg)

Then to find the heat lost from HRSG to the environment can be calculated as:

𝑞𝑙𝑜𝑠𝑡 = 𝑞𝑜𝑢𝑡 − 𝑞𝑖𝑛 (equation 6)

and the efficiency of the HRSG can be obtained as:

𝜂𝐻𝑅𝑆𝐺 =𝑞𝑤𝑎𝑡𝑒𝑟 𝑔𝑎𝑖𝑛𝑒𝑑

𝑞𝑔𝑎𝑠 𝑙𝑜𝑠𝑡 (equation 7)

20

3.3.3. Rankine Cycle

Pump, boiler, turbine, and condenser are four components in the Rankine

cycle. Figure below showing the temperature-entropy diagram of the Rankine cycle

as steady-flow devices, so that the four cycles that make up the Rankine cycle can

be viewed as a fixed flow system. [11]

Figure 3.5: T-s diagram of Rankine Cycle

Thermal efficiency of the ideal Rankine cycle can be derived as the work

net divided the heat inlet, or:

ηth =Wnet

qin=

Wturb,out−Wpump,in

qin (equation 8)

where:

Wturb,out = h3 − h4 (equation 9)

Qin = h3 − h2 (equation 10)

as,

ηth = Thermal Efficiency

Wturb,out = Steam turbine work (kJ/kg)

Wpump,in = Feed water pump work (kJ/kg)

qin = heat comes to steam turbine (kJ/kg)

h = enthalpy (kJ/kg)

21

3.3.4. HTF pumping power

In the working fluid flowing through the pipe, a pressure drop occurs.

Therefore, an additional pump is needed in the solar field to overcome this problem.

Pumping power can be calculated by finding the average velocity of working fluid

and the pressure drop in the solar field with:

vavg =v

Ac=

mHTF/ρ

Ac (equation 11)

where:

vavg = Average velocity of working fluid (m/s)

v = volume flow rate of working fluid (m3/s)

Ac = cross sectional area of pipe (m2)

mHTF = HTF mass flow rate (kg/s)

ρ = HTF density (kg/m3)

and the pressure loss can be calculated as:

∆P = fL

D

ρ vavg2

2 (equation 12)

where,

∆P = pressure loss (kPa)

f = friction factor (from moody chart)

L = Tube length (m)

D = Tube inner diameter (m)

so that, the pumping power can be determined as:

Wpump = v ∆P =mHTF

ρ ∆P (equation 13)

Wpump = pumping power (kW)

22

3.3.5. DSG pumping power

Figure 3.6: DSG configuration

Similar with HTF configuration, the pumping power of the DSG

configuration can be determined through equation:

Wpump = v ∆P =mHTF

ρ ∆P

but in DSG case, the pressure loss at the solar field can be determined as:

∆P = ∆PHRSG + ∆PSolar,field (equation 14)

23

3.3.6. Parabolic trough efficiency and solar collector area

A heat collector's efficiency can be defined as the energy that the working

fluid obtains against direct normal radiation that leads to the collector. Efficiency

is usually calculated between the difference in heat with ambient temperature and

temperature in operation. The figure below is the performance curve for the

working temperature of the solar collector [12].

Figure 3.7: Thermal efficiency of parabolic trough solar collector

as the heat capacity of solar field is:

Qsf = Asf × DNI × ηsc

so that, the area of solar collector can be determined as:

Asf =Qsf

DNI×ηsc (equation 15)

where,

Asf = solar field area (m2) Qsf = solar field heat capacity (kW)

DNI = Direct Normal Irradiation (W/m2)

ηsc = solar collector thermal efficiency

24

3.3.7. Overall thermal efficiency of the CC and ISCC

Figure 3.8: Combined Cycle scheme

As shown in figure above, there are two gas turbine engine and one Rankine

cycle connected in this combined cycle power plant. So that the efficiency can be

determined as:

ηth,CC =Wnet,GT1+ Wnet,GT2+Wnet,Rankine

Qin1+Qin2 (equation 16)

where,

WnetGT = 2 × ma × Wnet,GT (equation 17)

Wnet,Rankine = mw × Wnet,Rankine (equation 18)

Qin = 2 × ma × Qin,GT (equation 19)

25

Figure 3.9: ISCC scheme

And from figure 3.9 there are Qsolar that added in ISCC, thus the thermal

efficiency can be determined with

ηth =Wnet,GT1+ Wnet,GT2+Wnet,Rankine

Qin1+Qin2+Qsolar (equation 20)

26

CHAPTER 4

RESULT AND DISCUSSIONS

4.1. Current Situation

4.1.1. Gas turbine engine analysis

From the data obtained at the table, the T-s diagram of the gas turbine engine

can be seen from the figure below

Figure 4.1: Gas turbine engine T-s diagram

at the state 1, the temperature inlet (T1) is 35 °C, the enthalpy can be

obtained from Ideal-gas properties of air.

State 1:

T1 = 308 K h1 = 308.23 kJ/kg

State 2:

T2 = 639 K h2 = 648.16 kJ/kg

State 3:

T3 = 1313 K h3 = 1411.43 kJ/kg

State 4:

T4 = 833 K h4 = 858.35 kJ/kg

As the enthalpy from each state is found, so the heat input of the gas turbine

engine can be determined with equation (1):

27

Qin = 763.27 kJ/kg

also, the gas turbine work output and compressor work input can be determined

from equation (2) and (3):

Wturb, out = 553.08 kJ/kg

Wcomp, in = 339.93 kJ/kg

Thus, the net-work of gas turbine engine is

Wnet = 213.15 kJ/kg

4.1.2. HRSG heat balance

From the data in table 3.1, an analysis can be performed to calculate the heat

balance in HRSG. As shown in the figure 3.4 the water flows come up through state

1 and state 2, so the heat occurs at water side (qin) is:

State 1:

T1 = 45 °C; h1 = 188.44 kJ/kg (compressed liquid)

State 2:

P2 = 12.1 MPa; h2 = 3427.8 kJ/kg (superheated vapor)

T2 = 530 °C

As the steam flow rate is 22.51 kg/s, so the heat received by HRSG can be

determined from equation (4):

qin = 72918 kW

and the state 3 and 4 showing the flow of the exhaust gas, so:

State 3:

T3 = 560 °C (833 K); h3 = 858.35 kJ/kg

State 4:

T4 = 150 °C (423 K); h4 = 424.31 kJ/kg

28

Where the exhaust gas mass flow rate is 111.1 kg/s and there are 2 gas turbines so

from equation 4, the qout of the HRSG is:

qout = 96443.69 kW

and heat lost by the HRSG can be determined with equation (6):

qlost = qout − qin = 96443.69 − 72917 = 23526.69 kW

So that, the efficiency of the HRSG can be obtained from equation (7) as:

𝜂𝐻𝑅𝑆𝐺 =𝑞𝑖𝑛

𝑞𝑜𝑢𝑡=

𝑞𝑤𝑎𝑡𝑒𝑟 𝑔𝑎𝑖𝑛𝑒𝑑

𝑞𝑔𝑎𝑠 𝑙𝑜𝑠𝑡=

72917

96443.69= 0.756 = 75.6%

4.1.3. Ideal Rankine cycle heat balance

To analyze the heat balance in the Rankine cycle, data from table 3.2 is

needed. Assuming steady operation condition and the kinetic and potential energy

changes are negligible so:

State 1:

T1 = 45 °C; h1 = 188.44 kJ/kg

P1 = 10.8 kPa v1 = 0.001029 m3/kg

State 2:

P2 = 12.1008 MPa

s2 = s1

Wpump, in = v1 (P2 – P1) = 0.001029 m3/kg (12100.8 – 10.8)

= 12.440 kJ/kg

h2 = h1 + wpump, in = 188.44 + 12.44

= 200.88 kJ/kg

29

State 3:

T3 = 530 °C h3 = 3463.2 kJ/kg

P3 = 8900.8 kPa s3 = 6.7804 m3/kg

State 4:

P4 = 10.8 kPa hf = 197.27 kJ/kg

s4 = s3 hfg = 2388.9 kJ/kg

sf = 0.6661 m3/kg

sfg = 7.4600 m3/kg

𝑥4 =𝑠4−𝑠𝑓

𝑠𝑓𝑔=

6.7804−0.6661

7.4600= 0.82

h4 = hf + x4hfg = 197.27 + (0.82) (2388.9)

= 2156.2 kJ/kg

Thus,

the steam turbine work output can be determined by equation (9), so:

Wturb, out = (3462.2 – 2156.2) kJ/kg

= 1306 kJ/kg

so that, the work net of the Rankine cycle can be determined as:

Wnet = Wturb, out – Wpump, in = (1306 – 12.44) kJ/kg

= 1293.56 kJ/kg

as the Qin is the enthalpy difference at state 2 and 3 as shown in equation (10), so:

Qin = 3262.32 kJ/kg

The Rankine cycle efficiency can be obtained from equation (8) as:

ηth =1292.56

3262.32= 39.62%

30

4.1.4. Combined cycle thermal efficiency

As the work net and heat input on each cycle has been determined, so using

equation (16), (17), (18), and (19), thermal efficiency of the combined cycle can be

obtained.

Wnet,GT = 47361 kW

Wnet, Rankine = 29118 kW

Qin = 169598 kW

So, the combined cycle thermal efficiency is:

ηth,CC = 45 %

31

4.1.5. Current condition heat capacity

In the water side, the enthalpy (kW) for each state are shown in the table

4.1. The enthalpy could be derived as the enthalpy different multiplied by the steam

flow rate (22.51 kg/s).

Table 4.1 Enthalpy for the steam line

For the exhaust gas line, the exit temperature of the gas turbine is 560 °C

and leaving the HRSG at 150 °C. The enthalpy of the exhaust gas can be found with

the multiplication of the enthalpy difference, mass flow rate of air (111,1 kg/s),

thermal efficiency of the HRSG and also the number of the gas turbine. It can be

derived as:

𝐻 = 2 × ��𝑎 × 𝜂𝐻𝑅𝑆𝐺 × ∆ℎ

Table 4.2 Enthalpy for the gas line

From the diagram below, we can see the T-h diagram for current

configuration without solar integration on it. The blue line stand for the steam line

and the red line is stand for the exhaust gas line. For the steam line, water comes

T [°C] h [kJ/kg] H (kW)

45 188,44 0

302 1356,12 26284,48

302 2745,52 57559,87

530 3462,16 73691,44

T [°C] h [kJ/kg] H (kW)

150 424,31 0

197 472,24 8051,435

277 555,74 22078,03

357 638,63 36002,16

437 724,04 50349,6

507 800,03 63114,65

560 858,35 72911,43

32

from 45 °C, then at 302 °C it changes from saturated liquid to the saturated vapor

and goes to the steam turbine inlet at 530 °C. And for the exhaust gas line, the gas

flowing from state 1 at 560 °C and ends up at 150 °C as the HRSG exit temperature.

Figure 4.1: Current condition T-h diagram

0

100

200

300

400

500

600

0 20000 40000 60000 80000

Tem

per

atu

re

Enthalpy (KW)

Steam Line

Exhaust Gas

33

4.2. Analysis of Implementing Solar Field

4.2.1. Direct Steam Generator (DSG) effect

In the DSG system, there are no heat exchanger, so the water that comes out

from the low-temperature HRSG is directly channeled into the solar fields to be

heated then flowed back to the high-temperature HRSG. in this case, the pinch point

is at 350 °C with a temperature difference of 10 °C. So, the critical point gas turbine

line is at a temperature of 360 °C as shown in figure 4.2.

Figure 4.2: DSG-Configuration T-h diagram

with,

h530 °C = 3462.16 kJ/kg h560 °C = 858.35 kJ/kg

h350 °C = 2957.3 kJ/kg h360 °C = 642.87 kJ/kg

ma = 168 kg/s

As the heat gained by water (Qin) is equal with the heat lost by the air (Qout), the

mass flow rate of the steam can be determined from equation (3) and (4), so:

mw = 71.72 kg/s

45

164,57

303 303

350

530

150

360 360

560

0

100

200

300

400

500

600

0 10000 20000 30000 40000 50000 60000 70000 80000

Tem

pe

ratu

re (

°C)

Enthalpy (kW)Steam Line Exhaust Gas Line

Tx

34

After finding the mass flow rate of the steam, the enthalpy at DSG inlet (Tx)

can be determined to get the solar field heat capacity. As the heat capacity at lower

temperature of exhaust gas (Q@360°C-150°C) is equivalent with the heat capacity at

lower temperature of the steam line (Q@Tx-45°C),

with,

ma = 168 kg/s mw = 71.72 kg/s

H360°C = 642.8 kJ/kg h45°C = 188.44 kJ/kg

h150°C = 424.31 kJ/kg

the enthalpy at Tx is

hTx = 695.41 kJ/kg

from the saturated water table properties, the temperature at enthalpy 695.41 kJ/kg

is 164.57 °C. By determining the temperature at DSG inlet (Tx), the heat capacity

in the solar field can be found by multiplying the mass flow rate with the enthalpy

difference, so:

Qsf = 162223.7 kW

As the solar field heat capacity found, the area for the solar field can be determined

from equation (15), so that:

Asf = 259973.9 m2

The pumping power of DSG configuration can be analyzed with:

ρDSG = 958 kg/m3 mDSG = 71.72 kg/s

vavg = 15.67 m/s f = 0.013(from moody chart)

number of tubes = 372 tube

from equation (13), the pressure drop is:

ΔPDSG = 1971.6 kPa

So that, from equation (12) the pumping power is: Wpump,DSG = 147.6 kW

35

4.2.2. Heat Transfer Fluid (HTF) effect

At the HTF configuration, there are heat exchanger between the steam side

and the solar side. This heat exchanger is useful for increasing the temperature of

the steam coming from the low temperature of HRSG and then flowing it back to

the high temperature of HRSG using the working fluid media. The temperature out

of the fluid (thermal oil) from the heat exchanger is 350 °C and by assuming the

temperature difference at the pinch point is 10 °C, so the exit temperature of the

steam is 340 °C, as shown in the figure below.

Figure 4.3: HTF-Configuration T-h diagram

with:

h530 °C = 3462.16 kJ/kg h560 °C = 858.35 kJ/kg

h340 °C = 2917.22 kJ/kg h360 °C = 642.87 kJ/kg

ma = 168 kg/s

The mass flow rate of water at this configuration can be calculated through equation

(3) and equation (4), so:

mw = 66.78 kg/s

45

191,85

303 303340

530

150

350 350

560

0

100

200

300

400

500

600

0 10000 20000 30000 40000 50000 60000 70000 80000

Tem

pe

ratu

re (

°C)

Enthalpy (kW)Steam Line Exhaust Gas Line

Tx

36

To analyze the heat capacity of the HTF configuration in the solar field, the

enthalpy of the steam at inlet temperature of heat exchanger (Tx) is identified with,

ma = 168 kg/s mw = 66.78 kg/s

H350°C = 673.7 kJ/kg h45°C = 188.44 kJ/kg

h150°C = 424.31 kJ/kg

So that, the enthalpy at Tx is 815.70 kJ/kg

from the saturated water table properties, the temperature at enthalpy 815.70 kJ/kg

is 191.85 °C. By determining the temperature at Tx, the heat capacity in the solar

field can be found by multiplying the mass flow rate with the enthalpy difference,

so the result is

Qsf = 140349 kW

As the solar field heat capacity found, the area for the solar field can be determined

through equation (14) where the solar field area needed is 227839.3 m2

In HTF configuration the mass flow rate of the thermal oil is also need to

be considered. With h350 °C = 673.7 kJ/kg and h313 °C = 584.88 kJkg the oil mass

flow rate is:

mo = 86.69 kg/s

The pumping power in the HTF configuration can be analyzed with the data below:

Ac = 0.00477 m2 mHTF = 86.69 kg/s

ρHTF = 1068 kg/m3

So, the average velocity of the working fluid can be obtained from equation (10):

vavg = 17 m/s

as,

f = 0.014(from moody chart) L = 100 m

D = 0.078 m number of tubes = 326 tube

37

the pressure drop of this configuration can be determined from equation (11), so:

ΔPHTF = 2768.55 kPa

So that, from equation (12) the pumping power of HTF configuration is:

Wpump, HTF = 225.53 kW

38

4.3. Comparison of Simple Combined Cycle and ISCC

4.3.1. Combined cycle and DSG configuration

As the mass flow rate of the steam in integration with DSG is calculated,

the capacity of the steam turbine work output can be determined by considering the

inlet temperature of the steam turbine and the exit temperature leaving the steam

turbine.

Figure 4.4: Steam Turbine working principal

The steam turbine work output (Wturb, DSG) with the DSG configuration can

be determined from equation (8). Where mw = 71.72 kg/s, h1 = 3462.16 kJ/kg,

h2 = 2156.2 kJ/kg the turbine work output is:

Wturb,DSG = 93664 kW

Compared to the current configuration without solar integration, the steam

turbine capacity is 29420.57 kW. It means with the configuration with the solar

integration, the steam turbine capacity is more than three times bigger. From this

comparison, by using solar integration with DSG configuration, two more identical

steam turbine can be added.

39

The Rankine cycle thermal efficiency of this configuration can be

determined from equation (8) where the pumping power needed is 147.6 kW, so

the Rankine cycle thermal efficiency is:

ηth,DSG =(93664−147.6)

(234792.6)= 39.83%

From equation (20) the overall thermal efficiency of DSG-ISCC

configuration can be obtained where:

Wnet, GT = 47361 kW Qin, GT = 169598 kW

Wnet, Rankine = 93516.42 kW Qin, solar = 162223.7 kW

ηth,ISCC−DSG = 42.4%

4.3.2. Combined cycle and HTF configuration

Similar with section 4.3.1, the steam turbine capacity can be determined by

the inlet and the exit temperature of the steam turbine. With the same inlet and exit

temperature and the mw = 66.78 kg/s, the steam turbine work output at HTF

configuration is:

Wturb,HTF = 87217.92 kW

Almost similar with the DSG configuration, the steam turbine capacity with

HTF configuration is nearly two times bigger than the current combined cycle

steam turbine capacity. Its mean that the identical steam turbine can be added with

this configuration.

And the Thermal efficiency of this cycle also can be determined through

equation (7) with pumping power is 225.53 kW, thus the thermal efficiency is:

ηth =(87217.92−225.53)kW

218633.8 kW= 39.78%

40

From equation (20) the overall thermal efficiency of DSG-ISCC

configuration can be obtained where:

Wnet, GT = 47361 kW Qin, GT = 169598 kW

Wnet, Rankine = 86992.39 kW Qin, solar = 140349 kW

ηth,ISCC−DSG = 43.34 %

4.3.3. DSG and HTF configuration

From the calculation above, we can see that the heat capacity produced by

the Direct Steam Generation configuration is bigger than the Heat Transfer Fluid

configuration. It means that the DSG configuration can produce more turbine

output work (WST, out). And by this condition by applying the DSG configuration,

the installation of additional steam turbines can be done using one more identical

steam turbine. Table below show the comparison between the DSG and HTF

configuration.

Table 4.3 Comparison between DSG and HTF technologies

DSG Technology HTF Technology

Heat exchanger No Yes

Operating temperatures Promising Limited

Efficiency Higher, promising Medium, limited

Fluid toxicity No Yes

Configuration Complex Simple

Phase Flow Two phases One phase

Thermal development

storage

Expensive, demonstrative stage Less expensive, commercial

plants exist

Temperature gradients Higher High

Scaling up With additional cost Easier

Performance enhancement Promising Limited

Environmental risks Low High

Operation and maintenance

costs

Lower Higher

41

As shown in table 4.3, configuration using DSG technology has more

advantages compared to HTF technology. This is because in the DSG system there

is no need for a heat exchanger that requires more power for the circulation pump.

The installation of DSG technology in the combined cycle also provides more

benefits where the steam temperature increase in the Rankine cycle is more

significant. However, due to the two-phase flow in the DSG cycle, process stability

and also constant pressure pose more challenges in this process. And for the impact

on the environment, HTF is worse because it includes liquid oil and salt or certain

gases that can affect the environment. But in the solar field side, the HTF

configuration need less area compared to the DSG configuration. It means that the

pressure drops at DSG configuration need more attention.

Process stability Less stable Stable

Leaks Higher Low

Solar field size Larger Smaller

Advancements Very promise Limited

Power output Higher High

42

CHAPTER 5

CONCLUSIONS AND RECOMMENDATION

5.1. Conclusions

Based on the results of data and experimental analysis with reference to the formulation of

the problem, this study can be summarized as follows:

1. Integration of solar power in combined cycle power plants is very possible. By

applying solar power through a parabolic trough system, the capacity of the steam

turbine can be increased.

2. By using the DSG configuration, the turbine steam capacity has increased more

than three times (93664 kW) compared to the current configuration (29420.57 kW).

Thus, the addition of identical steam turbines can be done to drive the generator.

With the addition of an identical steam turbine, the steam turbine in the current

configuration does not need to be replaced. Solar collector area needed for this

configuration is 259973.9 m2. The Rankine and ISCC efficiency are 39.83% and

42.4%. The weakness of this configuration is due to the length of the steam flow

pipe that flows directly to HRSG, it is very possible for leakage to the steam pipe.

This causes the care and maintenance of the steam pipe must be paid more attention

to prevent leakage that will disrupt the entire Rankine cycle.

3. Through the HTF configuration, the increase in capacity in the turbine also occurs

about less than three times (87217.92) the capacity of the turbine in the current

conditions. Solar collector area needed is 227839.3 m2. Rankine and ISCC

efficiency are 39.78% and 43.34%. Another disadvantage of this configuration is

that an additional pump is needed to supply thermal oil to the solar fields. Thus, the

extra power needed in the solar field.

43

5.2. Recommendation

In this project, constant solar radiation calculations are used. As solar energy is very low

at night, ISCC operation can only be carried out during the day. At night, normal

combination cycles are used without involving the solar field. And if a constant output is

desired from a steam turbine, more fuel is needed to obtain a constant output. For more

accurate output power analysis, measured DNI data must be used.

44

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