Comprehensive Energy Systems, vol.4a - Energy Conversion [4a, 1 ed.] 978-0-12-814925-6

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4.1 The Role of Energy Conversion Shahid Islam, University of Ontario Institute of Technology, Oshawa, ON, Canada and King Fahd University of Petroleum and Minerals, Dhahran, Saudi Arabia Ibrahim Dincer, University of Ontario Institute of Technology, Oshawa, ON, Canada r 2018 Elsevier Inc. All rights reserved.

4.1.1 Introduction 4.1.2 Background 4.1.3 Systems and Applications 4.1.3.1 Conversion of Chemical Energy of Fuel to Heat 4.1.3.1.1 Heating value of a fuel 4.1.3.2 Thermal Efficiency of Energy Conversions 4.1.4 Energy Conversions Analysis of Systems in Steady State 4.1.4.1 Turbines and Compressors 4.1.4.2 Heat Engines 4.1.4.3 Thermal Efficiency 4.1.4.4 Refrigerators 4.1.4.4.1 Coefficient of performance 4.1.4.5 Heat Pump 4.1.4.6 Absorption Chillers 4.1.4.7 The Carnot Heat Engine 4.1.5 Renewable Energy Conversions 4.1.5.1 Biomass Energy Conversion 4.1.5.2 Wind Energy Conversion 4.1.5.3 Ocean Current Energy 4.1.5.4 Solar Thermal Energy Conversions 4.1.5.5 Geothermal Energy Conversion 4.1.6 Case Studies 4.1.6.1 Case Study 1 4.1.6.1.1 Thermodynamic assessment 4.1.6.1.2 Results and discussion 4.1.6.2 Case Study 2 4.1.6.2.1 Thermodynamic assessment 4.1.6.2.2 Results and discussion 4.1.7 Future Directions 4.1.8 Concluding Remarks References Further Reading Relevant Websites

Nomenclature A C _ Ex ex g GHV h H HHV K L LHV

Area (m) Compressor Exergy rate (kW) Specific exergy (kJ/kg) Gravity (m/s2) Gross heating value (kJ/kg) Specific enthalpy (kJ/kg) Total enthalpy of the flow High heating value (kJ/kg) Thermal conductivity (W/m2 K) Length of semiconductor (m) Lower heating value (kJ/kg)

Comprehensive Energy Systems, Volume 4

_ m n P _ Q r R S_ gen s T v V _ W Z

doi:10.1016/B978-0-12-809597-3.00401-6

2 4 7 7 7 8 9 10 14 15 16 17 18 19 20 21 21 22 24 25 28 29 30 30 33 34 34 35 36 37 37 38 39

Mass flow rate (kg/s) Number of moles Pressure (kPa) Heat rate (kW) Radius (m) Refrigerant Entropy generation (kW/K) Specific entropy (kJ/kg K) Temperature (K) Specific volume (m3/kg) Velocity (m/s) Work rate (kW) Figure of merit

1

The Role of Energy Conversion

2

Greek Symbols D Change r Density (kg/m3) Z Energy efficiency

y c o

Total energy of flowing fluid Exergy efficiency Humidity ratio (kgwater/kgair)

Subscripts a abs act avail avg b c c chrg cond cool d dest dis E en eva gen h

Charging water inlet state Absorber Actual Available Average Charging water exit state Cold Discharging water inlet state Charging Condenser Cooling effect of absorption chiller Discharging water exit state Destruction Discharging exit state Energy Evaporator Generator Hot

H HE hp i L n p P prod rev s st t th TEC TEG u vap 0 1, 2,….67

High Heat exchanger Heat pump Inlet state Low n type semiconductor p type semiconductor Pump Products Reversible Source Storing Turbine Thermal Thermoelectric cooler Thermoelectric generator Utilization Vaporize Reference environment State number

Acronyms AFUE CHP COP EBE EES EnBE ExBE HAWT HEX

Annual fuel utilization efficiency Combined heat and power Coefficient of performance Energy balance equation Engineering equation solver Entropy balance equation Exergy balance equation Horizontal axis wind turbine Heat exchanger

HTF LiBr–H2O multigen MBE ORC PTSC TEC TEG TES VAWT

Heat transfer fluid Lithium bromide–water Multigeneration Mass balance equation Organic Rankine cycle Parabolic trough solar collector Thermoelectric cooler Thermoelectric generator Thermal energy storage Vertical axis wind turbine

4.1.1

Introduction

A huge number of energy conversion processes occur naturally. Humans have invented a large number of additional energy conversion methods throughout history. Energy conversion devices can be classified according to chemical and physical principles, and the different forms of energy at the inlet and exit states of the device. In this chapter, different types of energy conversion fundamentals including renewable energy will be discussed and explained. The simplified process of changing energy from one form to another is referred to as energy conversion as represented in Fig. 1. The term “heat-power engineering” was commonly used for “energy conversion engineering” before World War II. The energy of available sources like the sun, fossil fuels, and nuclear fuels can be transformed into useful energy such as electricity, rotation, propulsion, cooling, and heating through energy conversion engineering [1]. Energy conversion engineers face the biggest challenges, like selection of the appropriate method, minimizing losses, reducing pollution, and reducing overall cost of the developed systems, respectively.

Energy source Fig. 1 Conversion of energy sources to useful outputs.

Energy conversion technology

Useful energy

The Role of Energy Conversion

3

Energy consumption (1000 TJ)

45 40 35 30 25 20 15 10 5 0 1970

1980

1990

2000

Petroleum products Natural gas Liquid biofuels

2010 Year

2020

2030

2040

2050

Hydro Coal Other renewable energy

Electricity generation capacity (GW)

Fig. 2 Global energy consumption and projection. Data from Energy Information Administration US. Annual energy outlook report. Available: https://www.eia.gov/outlooks/aeo/pdf/0383(2017).pdf; 2017 [accessed 20.03.17].

40 30

History

Projections

Additions

Solar Wind Oil and gas Nuclear Coal

20 10 0 10 20 30

Retirements 2005

2010

2015

2020

2025

2030

2035

2040

Year

Fig. 3 Additions and retirements of annual electricity generating capacity in gigawatts. Reproduced from Energy Information Administration US. Annual energy outlook report. Available: https://www.eia.gov/outlooks/aeo/pdf/0383(2017).pdf; 2017 [accessed 20.03.17].

The demand for energy is estimated to increase at a much faster rate in the near future. Fig. 2 provides an overview of the energy consumption through previous years and an estimate of the future energy consumption, based on different energy sources [2]. Renewable energy sources (except hydropower) continue to offer more potential than actual energy production. The demand of energy is increasing worldwide due to growing population and higher living standards. The combustion of fossil fuels, mainly coal, natural gas, and petroleum, supply most of the energy demand of the world. The utilization of fossil fuels through combustion to meet the increasing energy demand results in fast depletion of fossil fuel reserves and environmental degradation like acid rain, smog formation, global warming, ozone depletion, and health hazards. The energy conservation and search for alternative sources of energy are crucial to encounter with energy crisis and pollution. In the past, various investigations have been done on the conservation of energy in fossil fueled power generation systems. It is important to exploit alternative sources of energy to mitigate environmental concerns and global warming. Conservation and energy efficiency have exhibited compelling results over the past three decades. In addition to this, energy efficient processes offer potential to relieve some of the dependence on the import of petroleum products and compensate increased electric power demand to some extent [3]. Fig. 3 represents the additions and retirements of the annual electricity generating capacity in gigawatts from different energy sources according to the US Energy Information Administration annual energy outlook report [2]. The projections show that the use of coal to generate electricity is going to end by 2040, whereas the use of solar, oil, and gas energy sources will increase considerably. The environmental concerns and the limitations of the resources available on Earth are the deriving factors to increase the conversion of solar energy into electricity. The use of new concepts and alternatives to convert renewable energy are gaining acceptance due to strict environmental legislations. Currently the focus on the renewables is short term as compared to the long term objectives due to limited research and budget constraints. Nevertheless, open attitudes are established as new and previously discarded ideas are now being implemented. Some

4

The Role of Energy Conversion

Table 1

Illustrations of energy conversion processes corresponding to the initial form of the available energy and converted form of energy

Initial energy (form)

Converted energy (form) Heat

Mechanical

Electrical

Radiant

Chemical

Heat

Convector, radiator, heat pipe

Thermodynamic engines

–

–

Mechanical

Friction churning

Turbines

–

–

Electrical

Resistance, heat pump

Electric motor

Thermionic and thermoelectric generators Electric generator, MHDa –

Lamp, laser

Radiant Chemical

Absorber Burner, boiler

– –

– –

Electrolysis, battery charging Photolysis –

Nuclear

Reactor

–

–

–

Photovoltaic cell Fuel cell, battery discharge –

a

Magneto-hydrodynamic generators. Source: Modified from Sorensen B. Renewable energy conversion, transmission and storage. Burlington, MA: Elsevier/Academic Press.

System Energy in

Process Energy out

Surroundings Loss Fig. 4 Layout of energy conversion process with loss to surroundings.

of the examples are combined gas and steam turbine cycles, solar and wind energy-based farms, rotary combustion engines, multigeneration, solar concentrators, photovoltaic (PV) solar power, waste driven devices, turbocharged engines, fluidized-bed combustors, and coal- and biomass-based integrated gasification combined cycle power plants. The modification in the old technologies and the continuous development of new ones are necessary to meet the rapidly growing energy demand of the world. Hence, energy conversion engineering is an exceedingly appealing, complicated, and feasible field at present due to increased environmental concerns [1]. The conversion of one form of energy to another form depends on the process, requirements, and the quality of the available initial form of the energy. The process and the connections between the different forms of available energies are tabulated in Table 1. The most usable form of energy in a substance is usually in the form of extremely structured chemical bonds such as gasoline and sugar. It is possible to convert all other forms of energy to heat completely but the complete conversion of the heat to other energy forms is not possible because of factors like heat loss, friction, and other constraints. Therefore, the thermal efficiencies (Zth) of the energy conversion are considerably less than 100%. Typical steam power converts about 35%–40% of the heat to electricity whereas typical automobile engines based on gasoline operate at about 25% efficiency. Some amount of thermal energy is lost when it is converted to other forms of energy as represented in Fig. 4. In some cases it is very difficult to conserve the waste energy, for instance, 90% of the electrical power/energy is wasted in the form of heat in a light bulb while only 10% is converted to electricity. This clearly indicates that energy conversion efficiency has some distinct limits. Hence, a low temperature thermal reservoir is required as a sink to convert energy to useful work [4–7].

4.1.2

Background

The study of the history of science and engineering presented in Table 2 shows some of the key ideas and inventions along with the names of inventors. Some of these inventions are the landmarks for the energy conversion engineers. The significant inventions are tabulated in Table 2 as it is not possible to present the complete history associated with energy conversion. Most of these milestones were the achievements of teams of recognized individuals whose talents were essential part toward success. Table 2 also portrays how these ideas, events, and scientific and technological advancements are linked with each other and their dependence on their predecessor [8].

The Role of Energy Conversion

Table 2

5

Significant developments of techniques in energy conversion

Name of scientist/company

Year

Development/invention

Giovanni Branca James Watt James Watt John Barber Benjamin Thompson

1629 1765 1775 1791 1798

Robert Fulton Robert Stirling N. L. Sadi Carnot Michael Faraday Robert Mayer James Joule James Joule Rudolph Clausius William Thompson (Lord Kelvin) Etienne Lenoir A. Beau de Rochas James C. Maxwell Niklaus Otto Charles Parsons C.G.P. de Laval Rudolph Diesel – Albert Einstein Ernst Frank Whittle Frank Whittle Otto Hahn Hans von J. Ackeret, C. Keller Enrico Fermi – NASA Electricité de France Junghans, Germany Linus Torvalds iRobot Corporation Andre Geim and Konstantin Novoselov Nicholas Negroponte Apple Apple Elon Musk Jean-Pierre Sauvage, J. Fraser Stoddart, and Bernard Feringa

1807 1816 1824 1831 1842 1847 1849 1850 1851 1860 1862 1865 1876 1884 1889 1892 1895 1905 1926 1930 1937 1938 1939 1939 1942 1957 1969 1986 1990 1991 2002 2004 2005 2007 2010 2013 2016

Proposal of impulse steam turbine The idea of separate steam condenser First Boulton and Watt condensing steam engine Ideas and patent of gas turbine Mechanical energy conversion (Count Rumford) to heat observed in boring process of cannon First commercial steamboat Stirling engine Fundamentals of an ideal heat engine (foundation of thermodynamics) First generator (electric current) Heat and work equivalence 1st Law of Thermodynamics (basic ideas developed) Mechanical equivalent of heat measured 2nd Law of Thermodynamics Alternate of 2nd Law of Thermodynamics Internal combustion (IC) engine with no mechanical compression 4-stroke cycle (IC engine ) Mathematical principles linked with electromagnetics Four-stroke cycle internal combustion engine Steam turbine with multistage and axial-flow reaction Convergent-divergent nozzle used in impulse steam turbine Diesel engine with mechanical compression 1st Hydroelectric power (Niagara Falls) Equivalence mass/energy Schrodinger mechanics of quantum wave Patent application for turbojet engine Static test of 1st jet engine Nuclear fission discovered 1st turbojet engine flight (Ohain) Gas turbine (GT) electric power generation in a closed-cycle University of Chicago demonstrated nuclear fission Electricity produced through nuclear fission at Shippingport, Pennsylvania Man landed on moon with vehicle powered by rocket Fast breeder reactor with a capacity of 1200 MW at Superphénix, 1st grid power Radio-controlled clocks quartz clocks and watches Collaboratively written first version of computer operating system (Linux) First version of vacuum cleaning robot (Roomba) was introduced Graphene discovered Low-cost laptop called OLPC launched by MIT Touchscreen cellphone called the iPhone was introduced Touchscreen tablet computer, the iPad, came to market A giant, pneumatic tube transport system (Hyperloop) was developed Miniature machines out of molecules were built

Source: Modified from Sorensen B. Renewable energy conversion, transmission and storage. Burlington, MA: Elsevier/Academic Press and Woodford C, Technology timeline. Available: http://www.explainthatstuff.com/timeline.html; 2017 [accessed 25.10.17].

The capacity to perform work is called energy. The different forms of available energy are heat, gravitational, electrical, chemical, light, and nuclear. The summation of all forms of energy such as internal, kinetic energy, and potential energy in a system is the total energy possessed by the system. The sum of chemical, nuclear, sensible, and latent energies is the internal energy associated with a system. The rotation and vibration effects of atoms and molecules are the main cause of sensible internal energy. The combination of latent and sensible forms of internal energy is the thermal energy of a system. The natural sciences have placed boundaries on the classification of different forms of energy, such as interactions between neutrons and protons in the nucleus are atomic energy and energy of chemical bonds is chemical energy. The energy due to the molecular structure of a substance is independent of reference conditions [9]. The energy resources available in nature are connected to the energy services through energy systems. An extensive array of energy harnessing technologies is used to transform primary energy resources into useful commodities. The advanced technologies have also been developed to move energy to the location where energy service is required. For example, human travel is service, the

6

The Role of Energy Conversion

car is service technology, and gasoline is energy commodity produced in a refinery through crude oil. Therefore, crude oil can be categorized as a primary energy resource [10]. The shares of several primary sources available in a country are referred to as an energy mix. The main factors that define the structure of an energy mix are availability of the resources, geopolitical and environmental context, and the development of specific technologies to extract a particular energy resource in the context of political interest [10]. The stored potential energy in hydrogen can be released through fusion as in the sun. Sunlight is among the one of the converted energies as result of fusion in sun. Furthermore, this sunlight is converted into gravitational potential as it helps water to evaporate, then this water rains on the elevated areas of the earth. This rain water is stored in dams, which produce electricity with the help of turbines due to the potential energy of the rain water. Solar energy is also responsible for weather conditions like hurricanes, rain, wind, snow etc. Plants capture and convert water and carbon dioxide into proteins and carbohydrates through a chemical process called photosynthesis. This potential of chemical energy is the cause of the development and growth of biological cells [9]. The energy extracted from the environment is called primary energy and can be categorized in three groups: 1. renewable energy (e.g., solar energy, wind, biomass, ocean, hydropower, and geothermal energy) 2. nonrenewable energy (e.g., crude oil, coal, natural gas, and nuclear fuel) 3. waste Primary energy can be transformed to a useful form of secondary energy through the transformation process depicted in Fig. 5, [11]. Primary energy from nonrenewable energy sources like oil, natural gas, and coal provide 85% of the total demand of the world [7,12]. The projections of the world energy show that fossil fuels will still be dominating by 2035. According to the principle of supply and demand, the increase in the price of fossil fuels due to depletion will in turn make it economical to exploit renewable energy sources such as wind, solar, and biomass [7,12]. In most of the cases electrical energy is the final useful form of energy. The type of final energy form required sets a limit between the energy consumption and production [13,14]. The transformation of other forms of energy to electrical is based on the seven fundamental approaches [9]: 1. Static electricity is generated from the transport of charge and physical separation. The mechanical separation and transportation of the charge increase the difference between positive ( þ ve) and negative ( ve) charges, which results in static electricity. The lightning phenomenon is the best natural illustration of static discharge. 2. Electromagnetic induction is a process in which kinetic energy is transformed into electrical energy through three main devices: an electrical generator, an alternator, and a dynamo. This is the most widely used commercial process of electricity generation that uses the mechanical energy to drive a generator. The mechanical energy can be produced through different modes like hydro, tidal, wind, and heat engines. 3. Electrochemistry is a method of transforming chemical energy into electricity, such as, fuel cells and batteries. 4. Thermoelectric effect refers to the phenomenon of conversion of temperature differences at low and high temperature junctions of the device to electricity. Thermocouples, thermionics, and thermopiles are the best examples of the thermoelectric effect. 5. Photoelectric effect is a process in which light is transformed into electricity, like solar cells. 6. Piezoelectric effect is the generation of electricity as a result of mechanical strain induced in electrically anisotropic crystals. 7. Nuclear transformation is the generation and acceleration of the charged particles. It is possible to directly convert nuclear energy to electrical energy through beta decay in small scale projects. Some most commonly used devices or the processes to convert one form of energy to another form are tabulated in Table 3. The output energy in the form of electrical and mechanical are in high demand from energy sources as these forms can be converted to any other useful form easily.

Secondary energy

Waste

Solar, biomass, wind, hydro, geothermal

Petroleum products, derived solid fuels and gases

Electricity and heat

To consumption

Crude oil, coal, natural gas, nuclear

Transformation

Primary energy

Bio fuels

Fig. 5 Transformation of primary form of energy to secondary form of energy. Modified from Schobert HH. Energy and society. New York, NY: Taylor & Francis; 2002.

The Role of Energy Conversion

Table 3

7

Conversion of one form of energy to other through some devices/processes

Process

Input/available form of Energy

Useful form of energy

Steam engine Hydroelectric dams Photosynthesis Diesel/petrol engine Windmills Electric motor Fuel cells Generator/electric Bulb/electric Battery Ocean thermal Resistance heater Nerve impulse Bioluminescence Muscular activity Wave power Geothermal power Thermoelectric Friction Piezoelectric

Heat Potential/gravitational Solar Chemical Mechanical Electricity Chemical Mechanical Electricity Chemical Heat Electricity Chemical Chemical Chemical Mechanical Heat Heat Kinetic Strain

Mechanical Electrical Chemical Mechanical Electrical Mechanical Electrical Electrical Light and heat Electrical Electrical Heat Electrical Light Mechanical Electrical Electrical Electrical Heat Electrical

Source: Modified from Wu C. Thermodynamic cycles: computer-aided design and optimization. New York, NY: Dekker; 2004 and Granet I, Bluestein M. Thermodynamics and heat power. 6th ed Upper Saddle River, NJ: Prentice Hall; 2000.

Energy conversion technology blossomed after the establishment of thermodynamic fundamentals and electromagnetism, prior to which the technological advancement was significantly slow. The development and advancement in the nuclear technology is a result of theoretical and experimental research performed in the first half of the 20th century. The geothermal production of heat and power depends on the depth of reservoirs. Hydrothermal systems with temperatures more than 453K are found near the boundaries of plate tectonics. Intermediate temperatures with range 373–453K and low temperatures below 373K geothermal systems are also present in continental settings with/without hydrothermal resources. On the basis of the temperature, geothermal energy sources can be categorized as low temperatures below 363K, moderate temperatures between 363 and 423K, and high temperatures above 423K [15].

4.1.3

Systems and Applications

Some important energy conversion systems with their applications are discussed in this section.

4.1.3.1

Conversion of Chemical Energy of Fuel to Heat

The chemical energy contents of a fuel can be converted to heat through direct combustion. The heat is released as a result of combustion as this reaction is exothermic. The heat of reaction for a combustion is the same as the heat released as a result of combustion. The standard heat of combustion of a chemical can be calculated using standard heat of formation of the species in the product of chemical reactions. For example, combustion reaction of one mole of propane (C3H8) can be represented as C3 H8 ðgÞ þ 5O2 ðgÞ-3CO2 ðgÞ þ 4H2 OðgÞ þ Heat

ð1Þ

Considering the ideal gas state at standard pressure of 1 bar and temperature of 25ºC, the standard heat of chemical reaction can be written as X o o v DHfio ¼ vCO2 DHfCO þ vfH2 O D þ vC3 H3 DHfC ð2Þ DHro ¼ 2 3 H3 i i where, vi represents the stoichiometric coefficient of species i, which is negative for reactants and positive for products. In the above example these coefficients are 3, 4, and 1 for carbon dioxide, water, and propane, respectively.

4.1.3.1.1

Heating value of a fuel

The total amount of heat rejected during the process of combustion is called heating value of the fuel [16]. The heating value is usually measured in units of energy per unit mass of the substance, for example, kJ/kg, kcal/kg, or Btu/m3. The three ways to express the heating values or calorific values of fuels are the lower heating value (LHV), higher heating value (HHV), and gross heating value (GHV). HHV is obtained by reversing all the products taking part in combustion to the original precombustion temperature as well as condensing the water vapor as a result of combustion. The hydrocarbon-based fuels are combusted in the

8

The Role of Energy Conversion

presence of air, which can be expressed as ½C þ HŠ ðFuelÞ þ ½O2 þ N2 Š ðAirÞ-CO2 þ H2 O ðLiquidÞ þ N2 þ Heat ðHHVÞ

ð3Þ

½C þ HŠ ðFuelÞ þ ½O2 þ N2 ŠðAirÞ-CO2 þ H2 O ðVaporÞ þ N2 þ Heat ðLHVÞ

ð4Þ

where, C, H, O, and N represent carbon, hydrogen, oxygen, and nitrogen, respectively. The GHV in combustion reaction includes the heat required to vaporize water and the amount of liquid water present in fuel before burning. This GHV value is extremely significant for fuels like wood and coal as these substances contain some amount of water before burning. Net calorific value or LHV is obtained from the difference between HHV and the heat of vaporization (water vapor). The HHV is commonly correlated to the LHV as    nH2 O;out MW H2 O;out ð5Þ LHV ¼ HHV ðDHvap Þ nfuel;in MW fuel;in or LHV ¼ HHV

  mH2 O;out ðDHvap Þ mfuel;in

ð6Þ

where, DHvap, nH2 O;out , nfuel,in, MW H2 O , and MWfuel represent heat required to vaporize the water in kJ/kg, number of moles of vaporized water, moles of fuel combusted, molecular weight of water, and molecular weight of fuel, respectively. Illustrative Example 1: The HHV of methane (CH4) at room temperature (21ºC) is measured as 55,533 kJ/kg. Convert this HHV to LHV for a given heat of vaporization of water; DHvap,water ¼2454 kJ/kg at 21ºC. Solution: Assumption: Combustion of methane is complete, no unburnt methane in products. The equation of chemical reaction of the combustion process can be defined as CH4 þ 3=2O2 -CO2 þ H2 O HHV of methane ¼ 55;533 kJ=kg From periodic table MW H2 O ¼ H2 þ1=2O2 ¼ 2 þ 16 ¼ 18 kg=kmol

MW CH4 ¼ C þ 4H ¼ 12 þ 4 ¼ 16 kg=kmol nCH4 ¼ 1 kmol and nH2 O ¼ 2 kmol n

Therefore, nHCH2 O ¼ 2=1 4 Heat of vaporization: DHvap,water ¼2454 kJ/kg at 21ºC Hence, using Eq. (5) LHV ¼ HHV

ðDHvap Þ

LHV ¼ 55;533 kJ=kg

LHV ¼ 50;010 kJ=kg

4.1.3.2



  nH2 O;out MWH2 O;out nfuel;in MWfuel;in ð2454 kJ=kgÞ ð2=1Þ ð18=16Þ

Thermal Efficiency of Energy Conversions

The percentage of the thermal energy converted to other forms of useful energy is called thermal efficiency and it is denoted by “Zth.” Thermal efficiency for devices like furnaces or boilers is the ratio of measure of useful energy to the input energy. In general, it is not possible to achieve 100% efficiency. Thermal efficiency for steam power plants and other such systems is low. For instance, the efficiency of most light emitting bulbs ranges between 5% and 10% because of the losses to the environment in the form of heat. The thermal efficiency of a heater based on electric resistance is close to 100%, whereas, the efficiency of the natural gas-based furnace is about 80%. The selection of better heating unit is based on efficiency as well as other factors like economic analysis for the cost-effective selection [11,17–19]. Thermal efficiency associated with the chemical fuels is generally an evaluation of the chemical energy and the useful form of energy recovered as kinetic energy. The efficiency depends upon the type of heating value of fuel used as input energy because efficiency is obtained by the ratio of the useful energy extracted and released input energy. The production of saturated water in boilers involves the sum of latent heat of vaporization and increase in its sensible heat. The rate of heat gained by the water can be written as _ p;av DT þ DHvap Þ q_ ¼ mðC

ð7Þ

Vaporization: H2 O ðliquidÞ þ DHvap -H2 O ðvapor Þ

ð8Þ

_ m; _ Cp ; DT, and DHvap represents rate of heat gain, mass flow rate, temperature difference at inlet and outlet states, and heat where, q; of vaporization, respectively. Heat is gained in the vaporization process whereas, heat is released in the condensation process as

The Role of Energy Conversion

9

Condensation: H2 O ðvapor Þ-H2 OðliquidÞ þ DHcond

ð9Þ

The heat of vaporization at the constant temperature and pressure is the same as heat of condensation. At constant temperature and pressure: DHvap ¼ DHcond

ð10Þ

The LHV is actual magnitude of the heat that a boiler can produce because some of the combustion heat is required to evaporate water that is lost with the flue gases. The precise amount of the air supply is important for the efficiency of the boilers. The excess amount of air will result in loss of heat through the furnace. Conversely, less air will result in incomplete combustion and unburnt fuel. The water vapors are released to the stack during the combustion process so the net calorific value of the fuel does not include this portion of energy. The manufacturers of the furnaces often state the thermal efficiency in steady state but annual fuel utilization efficiency (AFUE) should also be considered before selection, which provides better estimate of energy effectiveness for a year.

4.1.4

Energy Conversions Analysis of Systems in Steady State

Many engineering devices like engines, compressors, and turbines are classified as steady-flow devices once they start operating under steady state for long hours after completing their transient start-up duration. Under steady state operation the fluid flow across the control volume is steady so such devices are called steady-flow processes. This means that the properties of fluid may change while flowing through a control volume but these remain constant during the entire process. The amount of mass energy and volume at inlet state of a steady-flow system must be equal to the exit state as shown in Fig. 6. The mass balance at inlet and exit states can be written as X X _ ¼ _ m m ðkg=sÞ ð11Þ in in oui out The fluid properties may change over a cross section between inlet and exit but the rate of change remains constant at inlet and/ or exit. The interactions of heat and work between surroundings and steady-flow system remain unchanged with time. Hence, in a steady-flow process heat transfer rate and power obtained to or from the system remain constant. Fig. 7 represents the balance of mass flow rate and enthalpy when a stream is split in two streams at the exit. The enthalpy is considered in systems with ideal fluid flow instead of internal energy and the work (PV) performed by fluid flow is included within enthalpy. The mass balance of the above system can be written as _2þm _3 _1 ¼m m

ð12Þ

_ 1 h1 ¼ m _ 2 h2 þ m _ 3 h3 m

ð13Þ

The energy balance equation (EBE) can be written as

The general equation of all forms of energy (mass, heat, and work) balance across a control volume can be expressed as E_ in ¼ E_ out ðkW Þ

ð14Þ

Mass in Control volume mCV = Constant ECV = Constant

Mass out

Fig. 6 Mass and energy balance in a control volume.

. m1

. m2

h1 Control volume

h2 . m3 h3

Fig. 7 Mass flow rate and enthalpy in a control volume.

The Role of Energy Conversion

10

. Qout

Heat loss Hot water out . . m2 = m1 Control volume

Hot water tank

Electric heating element

. Win

. m1 Cold water in

Fig. 8 Water heater operating in steady state.

Fig. 9 Industrial steam turbine with capacity from 2 to 250 MW. Modified from Siemens Steam Turbines. Available from: https://www.energy. siemens.com/hq/en/fossil-power-generation/steam-turbines/steam-turbine-products.htm; 2017 [accessed 12.04.17].

_ in , where _ 1 enters the tank with heat Q A water heater in a steady state operation is represented in Fig. 8. Mass of cold water m _ out . The energy balance in _ in and heats up the water. The heat lost to the environment is Q electric heating element performs work W the form of heat, work, and mass can be written as X X _ in þ W _ out þ W _ in þ _ out þ _ ¼Q _ Q my my ð15Þ in out The unit mass of the fluid flowing possesses the summation of enthalpy, kinetic energy, and potential energy. y¼hþ

V2 þ gz 2

By substituting the value of “y” in above equation    X X  V2 V2 _ in þ W _ out þ W _ out þ _ in þ _ hþ _ hþ Q m m þ gz ¼ Q þ gz out in 2 2

ð16Þ

The energy balance relation is easy to use in the case when the amount and the direction of transfer of heat and work are well known. Conversely, when both of these parameters are unknown then direction for work or heat interactions are assumed. The general form of energy balance or the first law associated with a steady-flow system can be represented as   X   X V2 V2 _ W _ ¼ _ hþ _ hþ þ gz þ gz ð17Þ Q m m out in 2 2

_ or W _ shows that the assumption of heat or work transfer was made in the wrong direction and needs to The negative value of Q be reversed. The above expression can be simplified for a single stream as   V 2 V12 _ W _ ¼m _ h2 h1 þ 2 Q þ g ðz2 z1 Þ ð18Þ 2

4.1.4.1

Turbines and Compressors

Turbines drive electric generators in steam, hydro, or gas-based power plants. The turbine blades are fixed to a shaft, which rotates due to the force of the working fluid on the blades. The rotating shaft is coupled with the generator, which converts mechanical energy into electricity. Fig. 9 represents the blades of an industrial steam turbine.

The Role of Energy Conversion

11

Fig. 10 An oil-free SST-600 steam turbine including active magnetic bearing. Modified from Siemens Steam Turbines. Available from: https:// www.energy.siemens.com/hq/en/fossil-power-generation/steam-turbines/steam-turbine-products.htm; 2017 [accessed 12.04.17].

Fig. 10 shows an oil-free steam turbine (SST-600), which includes an active magnetic bearing for efficient operation. Pumps, fans, and compressors work opposite to the turbine as they increase the pressure of the working fluid. These devices are driven through mechanical energy, which rotates the shaft so the work input is required to drive them. The function of these three devices is similar but there is a difference in the tasks. Compressors: Intake low pressure gas and compress it to a very high pressure. Fans: Mobilize gas while there is slight increase in the pressure as well. Pumps: Perform work similar to the compressors except that they handle liquids unlike gas in the compressors.   _ The turbines are very well insulated so heat transfer is approximately equal to zero QE0 . The heat transfer for compressors is also assumed zero as except for the cases where inlet cooling is provided. It is also important to note that fans, pumps, and compressors require work input to drive whereas, turbines produce work output. The change in the potential energies associated with these devices is negligible therefore it is assumed zero (Dpe E 0). The change in the velocities in these devices is also too low to cause any major change so the kinetic energies of these devices are also taken as zero (Dke E 0). In turbines, the fluctuation in the fluid velocities are often too high, which results in substantial change in the kinetic energy but this change is minor as compared to change in the enthalpy, therefore, kinetic energy is evaded from the analysis [6,20]. Recent developments have successfully produced highly efficient and sustainable compressors for numerous applications within process industries as presented in Fig. 11. A wide range of turbocompressors are available to suit specific needs.

• • •

Illustrative Example 2: The intake of an air compressor receives air at 101.325 kPa and 290K and compresses it to 607.95 kPa and 450K. The air flows at a rate of 0.025 kg/sec. The heat loss during the compression process is 20 kJ/kg. Calculate the power required to drive the compressor assuming that net change in kinetic and potential energies is zero (Figs. 12 and 13). Solution: Compression process is steady and power input is required. Assumptions: 1. Process is steady-flow so no change with time at any point, so Dmcv ¼ 0 and DEcv ¼0. 2. Ideal properties of air are taken from Engineering Equation Solver (EES). 3. Negligible changes in kinetic and potential energies (Dke ¼ 0 and Dpe ¼ 0). Analysis: Compressor is assumed as a system with control volume. Work is supplied and heat is lost. Mass crosses the system _1¼m _2¼m _ boundary so m E_ in

dEsystem ¼ 0 ðSteady stateÞ E_ out ¼ dt

_ out þ mh _ in þ mh _ 1 ¼Q _ 2 W

E_ in ¼ E_ out ðDke ¼ 0 and Dpe ¼ 0Þ _ in ¼ mq _ out þ mðh _ 2 h1 Þ W

12

The Role of Energy Conversion

Fig. 11 Single-stage turbocompressor. Adapted from A unique portfolio of turbocompressors for all industries. Available from: https://www. energy.siemens.com/hq/en/compression-expansion/product-lines/; 2017 [accessed 12.04.17].

qout = 20 kJ/kg P2 = 607.95 kPa T2 = 450K Air m = 0.025 kg/s . Win =? P1 = 101.325 kPa T1 = 290K Fig. 12 Layout of compressor with inlet and exit parameters.

P1 = 3 MPa T1 = 673K V1 = 60 m/s z1 = 12m Steam turbine Wout = 6 MW

P2 = 15 kPa x2 = 90% V2 = 190 m/s z2 = 5m Fig. 13 Layout of turbine with inlet and exit parameters.

Using EES to find the values of ideal air at specified temperatures h1 ¼ h@290K ¼ 290:4 kJ=kg h2 ¼ h@450K ¼ 452:1 kJ=kg Substituting the values in above equation _ in ¼ ð0:25 kg=sÞ ð20 kJ=sÞ þ ð0:25 kg=sÞ ð452:1 W _ in ¼ 4:54 kW W

290:4ÞkJ=kg

The Role of Energy Conversion

13

Illustrative Example 3: Steam enters an adiabatic steam turbine at 3 MPa and 773K, and leaves at 15 kPa with 90% quality of steam. Turbine produces 6 MW power. The velocities and the heights at inlet and exit are 190 m/s and 12 m, and 60 m/s and 5 m, respectively. 1. Calculate change in enthalpy, kinetic energy, and potential energy. 2. Calculate work done per kg of the steam. 3. Determine rate of mass flow of steam. Solution: The inlet and exit temperature, pressure, velocity, and quality of the steam are given as shown in Fig. 8. Calculations of changes in kinetic and potential energies, work done per unit mass, and mass flow rate have to be performed. Assumptions: 1. Process is steady-flow so no change with time at any point, so Dmcv ¼ 0 and DEcv ¼0. 2. Real fluid properties of steam are taken from EES. 3. Turbine is adiabatic. Analysis: _2¼m _ _1 ¼m The turbine is considered as a system with control volume. Mass crosses the system boundary so m 1. Using EES to find the value of enthalpy h1 as P1 ¼ 3 MPa T1 ¼ 773K h1 ¼ 3456 kJ=kg The fluid is a mixture of liquid and vapor at the outlet of turbine so h2 ¼ hf þ x2 hfg h2 ¼ 2361 kJ=kg h1 ¼ ð2361

Dh ¼ h2

Dh ¼ Dke ¼

V22

V12 2

¼

3456Þ kJ=kg

1095 kJ=kg

ð190 m=sÞ2

2

ð60 m=sÞ2



1 kJ=kg 1000 m2 =s2



Dke ¼ 16:25 kJ=kg Dpe ¼ g ðz2

2. E_ in

E_ out ¼

dEsystem dt

  z1 Þ ¼ ½ 9:8 m=s2 ð12 Dpe ¼

¼ 0 ðSteady stateÞ wout ¼



ðh2

wout ¼

h1 Þ þ

V22

5ÞmŠ



1 kJ=kg 1000 m2 =s2

0:068 kJ=kg V12

2

½ 1095 þ 16:25

þ g ðz2

 z1 Þ ¼

ðDh þ Dke þ DpeÞ

0:068Š kJ=kg ¼ 1078 kJ=kg

3. Mass flow rate needed to produce 6 MW of power _ ¼ m



_ out 6000 kJ=s W ¼ ¼ 5:57 kg=s wout 1078 kJ=kg

The Role of Energy Conversion

14

Discussion: It is important to note that the magnitude of potential energy is too low while the magnitude of kinetic energy is negligible as compared to the enthalpy. Therefore, for the calculation of most of the engineering problems both of these energies are usually omitted as inclusion will not make any significant difference.

4.1.4.2

Heat Engines

It is easy to convert work to energy but the conversion of energy to work is not easy. For example, if a shaft rotates in a container of water the work will be converted to heat up water but if we heat up water with a shaft in it then the shaft will not rotate. This means that conversion of heat to work is possible with some mechanical device called a heat engine. Heat engines: 1. 2. 3. 4.

Use high temperature source like furnace, solar energy, nuclear reactor, etc. to provide heat. Convert partial amount of heat to work. Reject waste heat to sink (low temperature reservoirs like rivers, etc.). Operate based on a complete cycle.

A simple heat engine that absorbs heat from a high temperature source and rejects it to the sink is shown in Fig. 14. The devices that operate in a closed loop or cycle including heat engines use a fluid that absorbs heat from the source and rejects it to a sink; this fluid is called a working fluid. The wider sense of heat engine includes internal combustion engines, which burn hydrocarbon line gasoline, and this fluid does not complete the thermodynamic cycle. The definition of heat engine is best suited to the external combustion engine such as the steam turbine or steam power plant. Steam is generated by heating the water through an external source, which runs the turbine, and then working fluid is cooled close to the atmospheric conditions in a condenser and then is pumped back to the boiler with the help of a pump. The simplified schematic of a power plant is shown in Fig. 15. Heat source high temperature

Qin Heat engine

Wnet, out Qout

Heat sink low temperature Fig. 14 Schematic of a heat engine operating between heat source and sink.

Energy source Qin Boiler

Turbine

Pump Win

Condenser Energy sink

Fig. 15 Layout of a power plant with steam as a working fluid.

Qout

Wout

The Role of Energy Conversion

15

The different quantities represented in Fig. 15 may be explained as Qin ¼Heat addition in boiler from high temperature source to the steam Win ¼Work output delivered by the turbine Qout ¼ Heat rejection in condenser to atmosphere or any other low temperature sink Wout ¼ Work input required to pump the water back to the boiler The net amount of the work obtained through this power plant is the difference between the total work output and total work input as ð19Þ Wnet;out ¼ Win Wout ðkJÞ The net change in the internal energy of a closed system operating in a cycle is zero. There are four components connected with each other through pipes in the above steam power plant and the same working fluid is flowing through them; hence this can be treated as closed loop. Therefore, the net amount of work output can also be calculated from the difference between the heat supplied and heat rejected as Wnet;out ¼ Qin

4.1.4.3

Qout ðkJÞ

ð20Þ

Thermal Efficiency

The amount of heat rejected from the system cannot be zero, therefore, the net amount of the work obtained from a heat engine is always less than the heat supplied to it. The fraction of the heat input converted to the useful work output is called the thermal efficiency Zth. The thermal efficiency of a heat engine can be written as Thermal efficiency ¼

Net amount of work output Total amount of heat input

Zth ¼ where, Wnet,out ¼Qin

Wnet;out Qin

ð21Þ

Qout Qin

ð22Þ

Qout therefore, Zth ¼ 1

Illustrative Example 4: Two heat engines are operating at the same heat source of 250 kJ. The heat sink of engine 1 and engine 2 are 187.5 and 175 kJ, respectively. Find net amount of work done and thermal efficiencies of both engines. Solution: 1. Heat engine 1 with sink 187.5 kJ Qin ¼ 250 kJ Qin ¼ 187:5 kJ Eq. (20) can be used to calculate the net amount of work output. Wnet;out ¼ Qin Wnet;out ¼ 250

Qout

187:5 ¼ 62:5 kJ

Thermal efficiency can be obtained with the help of Eq. (22) as Zth ¼ 1

Zth ¼ 1

Qout Qin

187:5 ¼ 25% 250

2. Heat engine 2 with sink 175 kJ Qin ¼ 250 kJ Qin ¼ 175 kJ

The Role of Energy Conversion

16

W net,out = 60 kW

Fuel in m fuel Qin

Qout Exhaust out Fig. 16 Schematic of an internal combustion engine.

Eq. (20) can be used to calculate net amount of work output. Wnet;out ¼ Qin Wnet;out ¼ 250

Qout

175 ¼ 75 kJ

Thermal efficiency can be obtained with the help of Eq. (22) as Zth ¼ 1

Zth ¼ 1

Qout Qin

175 ¼ 30% 250

Illustrative Example 5: An engine of a car delivers 60 kW of power while working with 27% of thermal efficiency. The car is burning liquid propane whose heating value is 46,340 kJ/kg. Calculate rate of fuel consumption. Solution: The rate of fuel consumption or the mass flow rate of fuel is to be calculated while power output and thermal efficiency are given. Assumptions: The engine of the car provides constant power (Fig. 16). Analysis: The car engines convert chemical energy of the fuel to mechanical. In this case, 27% of the total energy of fuel is being converted to power. According to the thermal efficiency amount of energy required to produce 60 kW power can be found as _ _ in ¼ W net;out ¼ 60 kW ¼ 222:2 kW Q Zth 0:27 which means 222.2 kW thermal energy should be supplied to the engine to get 60 kW output for a 27% efficient engine. Hence mass flow rate can be determined as _ ¼ m

222:2 kW ¼ 0:0048 kg=s ¼ 17:28 kg=h 46;340 kJ=kg

Discussion: The fuel consumption depends on efficiency of the engine, which means engines with high efficiency consume less fuel as compared to the engines with less efficiency.

4.1.4.4

Refrigerators

The heat transfer from high temperature source to low temperature sink occurs naturally in the universe and there is no need for a special device for this purpose. However, the reverse process is not possible without some special device called a refrigerator. The refrigerators transfer the heat from low temperature space to high temperature region. The operation of refrigerators is also cyclic just like heat engines. The refrigerant is a working fluid, which circulates in a closed loop in refrigerators. The vapor-compression refrigeration cycle is most widely used, and is based on four main components:

• • • •

Compressor: Compresses refrigerant from vapor phase to liquid phase and condenser pressure. Condenser: The high temperature compressed refrigerant condenses after rejecting its heat through the condenser to the surrounding medium. Expansion Valve: The temperature and pressure of the refrigerant is dropped remarkably because of the throttling effect as it passes through the capillary tube. Evaporator: The low temperature refrigerant evaporates in the evaporator as it absorbs the heat of the refrigerated area/space. Then working fluid is directed back to the compressor and the loop is completed.

A schematic of the refrigeration cycle is shown in Fig. 17. The symbols Wnet,in, QL, and QH represent the net amount of work input, heat gained from the refrigerated space at TL, and heat rejected to the high temperature region at TH, respectively.

The Role of Energy Conversion

17

Surroundings

QH

Compressor

Condenser

Expansion valve

W net,in

Evaporator Refrigerated space

QL

Fig. 17 Schematic of a refrigeration cycle.

Warm heated space TH >TL

Compressor

QH

Expansion valve

Cold outside air

Wnet,in

QL

Fig. 18 Schematic of a heat pump device.

4.1.4.4.1

Coefficient of performance

The efficiency of the refrigerators is represented as coefficient of performance (COP) and denoted by COPR. The COP of the refrigerator is the ratio of desired output and work input as (Fig. 18) COPR ¼

Desired output QL ¼ Wnet;in Required work input

ð23Þ

The ratio of the above equation can also be written in the form of rate for both output and input. The amount of work input in this case for the closed loop can be expressed as Wnet;in ¼ QH

ð24Þ

QL

Then, Eq. (23) can be rewritten as COPR ¼

QL ¼ QH QL

1 QH QL

1

ð25Þ

It is important to note that the COPR can be more than one, which means that the amount of heat removed from the refrigerated space can be more than the work input whereas, thermal efficiencies are always less than one. This is why the efficiencies of the refrigerators are expressed as COP to avoid any confusion.

18

The Role of Energy Conversion

Illustrative Example 6: The heat is rejected from a refrigerator at the rate of 480 kJ/min. The power required to drive the compressor is 3.2 kW. Find out 1. Coefficient of performance. 2. The rate at which heat is rejected to the surroundings. Solution: The power required to drive the compressor is provided whereas, COP and rate at which heat is rejected are to be calculated. Assumption: Refrigerator operates in steady state. Analysis: (1) The COP can be determined as COPR ¼

  _L Q 480 kJ=min 1 kW ¼ 2:5 ¼ _ net;in 3:2 kW 60 kJ=min W

This means that each kJ of the work input removes 2.5 kJ of heat. (2) The rate of heat rejection can be calculated as   _H¼Q _L þW _ net;in ¼ 480 þ 3:2 kW 60 kJ=min ¼ 672 kJ=min Q 1 kW Discussion: The amount of energy removed from the refrigerator is transferred to the room, which shows that energy can be converted to another form but cannot be destroyed.

4.1.4.5

Heat Pump

A device that transfers the heat from low temperature region to high temperature medium is called a heat pump. The objective of the heat pump is to maintain heat in a high temperature medium, which is done by absorbing heat from the low temperature medium. The household heat pumps usually absorb heat from ambient air and supply it to the inside of the house and thus maintain high temperature. The performance of the heat pump is also expressed as COP of heat pump. COPHP ¼ whereas, Wnet,in ¼QH

Desired output QH ¼ Wnet;in Required work input

ð26Þ

QL so the above expression can be written as COPHP ¼

QH ¼ QH QL 1

1 QL =QH

ð27Þ

Heat pumps driven by heat source instead of electricity are called absorption heat pumps. The heat source can be propane, water, natural gas, or heat transfer fluid (HTF) heated up by solar/geothermal source. Natural gas operated heat pumps are known as gas-fired heat pumps. Gas-fired coolers are also available but these cannot be reversed to be used as heat source [21]. A heat pump driven by R410a is shown in Fig. 19, which can provide a green way of heating/cooling to the buildings with a COP of 1.5 [22]. Illustrative Example 7: A house has to be maintained at 23ºC by a heat pump device. The amount of heat lost from the house is 90,000 kJ/h when ambient temperature is 3ºC. If the COP of the heat pump is 3, then calculate 1. Power input required to operate heat pump. 2. The rate at which heat is absorbed from ambient air. Solution: The COP of the heat pump is known whereas, the power required to drive heat pump and rate of heat absorbed from ambient are to be calculated.

6

3

2

8

Evaporator

Space heating

Condenser

7

5

4 Compressor

Fig. 19 Schematic of an R410a driven heat pump.

Heat source

1

The Role of Energy Conversion

19

Assumption: Heat pump operates in steady state. Analysis: 1. Power consumed by the heat pump can be calculated using the definition of COP as _ net;in ¼ W

_H Q 90; 000 kJ kJ 1h ¼ 30;000 or 30;000  ¼ 8:33 kW ¼ COPHP 3 h h 3600 s

2. The rate at which the house is rejecting heat is 90,000 kJ/h to maintain 23ºC so the heat must be supplied at the same rate by the heat pump. _L ¼Q _H W _ net;in ¼ 90;000 30;000 ¼ 60;000 kJ=h Q Discussion: It is important to note that 2/3 of the heat requirement of the house is extracted from the ambient air. Conversely, if an electric heater is used the total amount of heat 90,000 kJ/h will be delivered by electric power input, which is why heat pump devices are preferable even though initial investment is high.

4.1.4.6

Absorption Chillers

Absorption chillers are generally classified on the basis of refrigerant and the combination of absorber. Most of the absorption chillers either use the combination of lithium bromide/water or ammonia/water as refrigerant [23]. For space cooling applications, lithium bromide/water chillers are preferred. The absorption cycle represented in Fig. 20 consists of six major components, namely generator, condenser, expansion vale, evaporator, absorber, and solution pump [24]. The cost of absorption refrigeration system is high as compared to the vapor-compression refrigeration systems. Moreover, these are more complex, low in efficiency, require more space, and large cooling towers. The absorption refrigeration systems should only be adopted when thermal energy is available at an inexpensive rate. These systems are more suitable for industrial and commercial applications due to the cost and space needed for the installation. The COP of the absorption system can be defined as COPabsorption ¼

Desired output QL QL ¼ D Qgen þ Wpump;in Qgen Required input

ð28Þ

The amount of work input required to drive the pump of the system is neglected due to its lower magnitude. When the whole absorption refrigeration cycle is considered as reversible then the amount of COP is maximum. The COP of a reversible absorption refrigeration system can be written as    QL T0 TL COPrev;absorption ¼ ¼ Zth;rev  COP ¼ 1 ð29Þ Qgen Ts T0 TL where, T0, TL, and TS are the temperatures of environment, refrigerated space, and heat source, respectively. Illustrative Example 8: The evaporator of an absorption chiller is producing 2200 kJ of cooling effect. The amount of heat generated in the generator is 1470 kJ. Calculate the COP of the absorption chiller. Solution: The amount of heat at the generator and the cooling produced in the evaporator are known whereas, COP is to be calculated. Assumption: Absorption refrigeration system is reversible. Analysis: The COP of the reversible absorption refrigeration can be calculated using Eq. (28) as QL ¼ 2200 kJ Qgen ¼ 1470 kJ

Heat source

1 3

8

Pump 2

HE 5

Generator

7 Absorber

5 4 Absorption refrigeration system

6 12

13

9 Condenser

Evaporator

16

Fig. 20 Schematic of absorption refrigeration chiller.

Cooling water

10 15

11

Cooling 14

The Role of Energy Conversion

20

COPrev;absorption ¼

QL 2200 ¼ 1:5 ¼ Qgen 1470

Discussion: It can be observed that 1.5 kJ of cooling can be produced for each kJ of heat supplied at the generator of the absorption chiller.

4.1.4.7

The Carnot Heat Engine

A theoretical heat engine based on reversible Carnot cycle is known as a “Carnot heat engine.” Thermal efficiency associated with the heat engine whether reversible or irreversible can be written as QL ð30Þ QH where, QH and QL are the ratio of the heat supplied and heat rejected at the high temperature source TH and low temperature sink TL, respectively. Moreover, this ratio of heat transfer is same as the ratio of the absolute temperatures of heat source and sink for all reversible heat engines [6]. Therefore, the Carnot efficiency of a reversible heat engine can also be described as Zth ¼ 1

Zth ¼ 1

TL TH

ð31Þ

Now the comparison of Eq. (30) and Eq. (31) yields 

QL QH



rev

¼

TL TH

ð32Þ

The above relation yields the highest possible efficiency for reversible engines operating between temperatures TH and TL and it is known as “Carnot efficiency.” It is impossible for a power cycle to yield 100% efficiency as there are irreversibilities associated with the actual cycle [25]. It is obvious from Eq. (31) that the efficiency can be increased either by increasing the temperature of the source reservoir or reducing the temperature of the sink. Both of these temperatures have constraints such as the high temperature supplied to a power cycle is limited by the material of the device and low temperature is restricted by the temperature of the sink like ambient, lakes, or rivers. The thermal efficiencies of most of the actual heat engines are under or around 40%, which seems to be very low as compared with 100%. As a matter of fact, the efficiencies of the actual heat engines should be correlated or compared to the efficiency of the reversible one under identical operating conditions as this is the factual theoretical limit for comparison. Fig. 21 represents the efficiencies of the reversible, irreversible, and impossible heat engines while operating between identical heat source and sink at 1000 and 300K, respectively. The thermal efficiency of the reversible heat engine is determined 70% using Eq. (31). Therefore, it is obvious that it is impossible to achieve the magnitude of thermal efficiency more than the efficiency of a reversible heat engine. There are two corollaries associated with the maximum possible theoretical efficiency of a Carnot cycle operating between high and low temperature reservoirs. These Carnot corollaries are: Corollary 1: When two power cycles, an irreversible and a reversible, operate between identical temperature reservoirs then the maximum thermal efficiency of the irreversible power cycle is always less than the reversible one. Corollary 2: The thermal efficiencies of reversible power cycles are the same when these operate at identical high and low temperature reservoirs.

High-temperature source TH = 1000K

Rev. HE th = 70%

Irrev. HE th = 45%

Impossible HE th = 80%

Low-temperature sink at TL = 300K Fig. 21 Schematic of the efficiencies of the reversible, irreversible, and impossible heat engines (HE). Reproduced from Cengel YA, Boles MA. Thermodynamics an engineering approach. 7th ed. New York, NY: McGraw-Hill; 2011.

The Role of Energy Conversion

21

The comparison of the thermal efficiencies associated with reversible and the actual heat engines while operating between identical temperature reservoirs is as follows: 8 oZth;rev irreversible heat engine > < reversible heat engine ð33Þ Zth ¼ Zth;rev > : 4Z impossible heat engine th;rev

Illustrative Example 9: A Carnot heat engine operates between high and low temperature reservoirs at 705 and 29ºC, respectively. The engine extracts 600 kJ of heat from the source in a cycle. Calculate

1. Thermal efficiency of the engine. 2. Heat lost to sink during each cycle. Solution: The temperatures of heat source and sink, and heat extracted are known Carnot efficiency and heat lost to sink are to be calculated. Assumption: Heat engine is reversible. Analysis: (1) The thermal efficiency of a reversible Carnot heat engine can be calculated using Eq. (31) as Zth;C ¼ 1

TL ¼1 TH

ð273 þ 29Þ ¼ 0:69 ¼ 69% ð273 þ 705Þ

(2) Now by using Eq. (33), the heat rejected to the sink can be calculated as QL;rev ¼

TL ð273 þ 29ÞK ð600 kJÞ ¼ 185:3 kJ  QH ¼ TH ð273 þ 705ÞK

Discussion: It can be observed that 31% loss in the efficiency is because this percentage of the heat supplied is lost to the sink.

4.1.5

Renewable Energy Conversions

The natural sources of energy like sunlight, wind, geothermal, biomass, rain, and tides are treated as renewable energy resources, which are replenished naturally after use. The energy from the sun is considered as a sustainable and a renewable source. In addition to this, solar thermal systems can provide power indirectly. Some of the important renewable energy sources are discussed below.

4.1.5.1

Biomass Energy Conversion

Biomass is a biological material that is largely extracted from living or dead matter available on the earth [26]. The biomass resources can be converted biofuel, heat, and power [27]. There are four technologies available to convert biomass into useful energy depending upon the types of biomass and specific energy product [28]. 1. Thermal conversion: Biomass feedstock is heated with or without oxygen and converted to other forms of energy. Direct combustion, torrefaction, and pyrolysis are the main processes involved in thermal conversion. 2. Thermochemical conversion: The combination of heat and chemical processes is applied to biomass to convert it into other forms of energy products. Gasification is one of the most important thermochemical conversion processes. 3. Biochemical conversion: The biomass is broken down into liquid-based fuels through bacteria, enzymes, or microorganisms which involves fermentation and anaerobic digestion. 4. Chemical conversion: Biomass is converted into liquid fuels by using chemical agents. The generation of electricity from biomass is considered as an inefficient use of biomass as the efficiency of such plants is only 25%. Efficiency levels of up to 80%–90% can be achieved through cogeneration technologies, whereas biopower efficiencies are 25%. The waste heat at the exit of the steam turbine is used to produce heat in combined heat and power (CHP) plants. The steam exiting the turbine loses a portion of the heat contents in it. This heat at the exit of the turbine is usually wasted to the atmosphere. The fuel gases also contain significant amount of thermal energy, which is vented to the outside air. The efficient recovery and utilization of the waste heat is the main focus of CHP systems. The production of heat along with the electricity generation yields higher overall energy and exergy efficiencies, reduced cost, and less CO2 emissions. The ideal application of these kind of systems is in pulp and paper mills as the demand of both heat and electricity is high in such industries. The high cost associated with the waste heat recovery unit can only be justified when the demand of heat is high. The district heating plants in Europe are successfully operating based on CHP technology with proven efficiency. Many researchers have studied biomass energy-based cogeneration systems for numerous industries like palm oil, rice, wood, sugar, and paper [29]. The technoeconomical overview of the variety of biomass samples has yielded competitive results for almond shell and olive stone [30]. The variety of biomass gasification processes are feasible energetically and exergetically for the

22

The Role of Energy Conversion

production of hydrogen and power [31]. The study of ignition and combustion characteristics of small particles of biomass and biomass energy-based multigeneration have yielded imperative results [31–34]. A biomass feedstock-based CHP plant is displayed in Fig. 22.

4.1.5.2

Wind Energy Conversion

Wind energy conversion machines have been developed by the mankind over the past 2000 years. In the early stages of the development this process was mostly based on hit and miss trial method. There are two major classifications of the windmills 1. Horizontal-axis wind turbine (HAWT): Uses a rotor that rotates about a horizontally placed shaft. 2. Vertical-axis wind turbines (VAWT): Uses a rotor that rotates about a vertically placed shaft. Both types of wind turbines are displayed in Fig. 23 along with their main features. The whole drivetrain mechanism is located on the tower in HAWT. The major two disadvantages of this type of windmill are gravity causes cyclic stresses on the blades so these should be oriented in the direction of the wind, and servicing is difficult as the driving mechanism is mounted on the tower. The advantage of this type of wind mill is that it can access stronger winds available at high altitudes by placing the rotor at a tall tower. Conversely, the drivetrain of VAWT is located on the ground. Hence, there are no gravitational stresses on the rotor as well as no problem with the orientation of the blades. However, in VAWT, it is not possible to install the blades at high altitudes to

Combustion gases Flue gas cleaning

Chimney

Heat exchanger Combustion chamber Biomass feedstock

Generator HP steam

Feed water

Boiler

Electricity

Residual heat Heat

Fig. 22 Biomass-based combined heat and power power plant. Modified from Biomass Innovation Centre: fueling rowth through clean technology. Available from: http://www.biomassinnovation.ca/CombinedHeatAndPower.html; 2017 [accessed 28.10.17].

Rotor diameter

Rotor blade Generator Rotor diameter Wind direction for an upward rotor

Hub height

Gear box Rotor tower

Tower

Wind direction for a donward rotor Fixed pitched rotor blade Rotor base

Equator height Generator

Gear box Fig. 23 A schematic of horizontal-axis wind turbine and vertical-axis wind turbine. Modified from Goswami DY, Krieth F. Energy conversion. Boca Raton, FL: Taylor & Francis Group; 2007.

The Role of Energy Conversion

23

extract more useful work available due to strong winds. In addition to this, the blades of VAWTs experience severe fluctuating aerodynamic load arising as a result of rotation. The propeller-type HAWT are generally classified as 1. Rotor orientation: Upwind of the tower or downwind of the tower. 2. Blade articulation: Rigid or teetering. 3. Number of blades: Usually composed of two or three blades. The most common types of modern VAWT have curved blades that are fixed in pitch. A massive 722 ft (220 m) wind turbine located at Maade, Denmark is considerably taller than the London Eye. It is the biggest and most powerful in the world with a production capacity of 260,000 kWh. The power produced by this wind turbine in 24 h can meet the electricity requirement of hundreds of homes for a month [35]. The harmonization of new wind turbine rotor blade development is possible using the PC computer with the CATIA designing system and the Gerber Garment cutter system. The blade fabricated from composite laminated materials yields better results. The contour of the airfoil can be formed by a continuous structural pocket and a fiberglass skin [36]. The kinetic energy of the wind is converting into rotational kinetic energy in the turbine. This rotational energy is converted to electrical energy with the help of a generator coupled with wind turbine and then this electrical energy can be supplied to domestic or industrial users through the national grid. The speed of the wind and the swept area of the turbine are the two main factors that define the conversion of available energy. It is important to calculate the economic viability by making initial estimates of output power and energy output of each turbine before planning or establishing a new wind form. In 1919, German physicist Albert Betz concluded that it is impossible to convert more than 59.3% of the kinetic energy of the wind into mechanical energy turning a rotor. This statement is known as the Betz limit or Betz’s law. The theoretical maximum power efficiency of any design of wind turbine is 0.59, which means that maximum energy extracted by any windmill is no more than 59% of the wind energy. This is known as the “power coefficient” and is defined as CP;max ¼ 0:59 Also, wind turbines cannot operate at this maximum limit. The Cp value is unique to each turbine type and is a function of wind speed that the turbine is operating in. Once we incorporate various engineering requirements of a wind turbine – strength and durability in particular – the real world limit is well below the Betz limit with values of 0.35–0.45 common even in the best designed wind turbines. In general, when all the losses associated with gear box, bearings, and generator are considered then about 10%–30% of the wind energy is transformed to electricity. Therefore, the power that can be extracted from the wind is represented as Pavail ¼

1 rAV 3 CP 2

ð34Þ

where, r, A, and v3 represent density of air, swept area, and wind speed respectively. The swept area depends upon the length of the turbine blades. Fig. 24 represents direct drive type 2 MW wind turbine with adaptronic blades. Illustrative Example 10: A windmill has length of the blades 48 m and power coefficient is 0.4. The wind is flowing at the speed of 15 m/s. Calculate the power produced by the wind turbine. Take the density of the air as 1.23 kg/m3.

Fig. 24 Schematic of direct drive type wind turbine of 2 MW with adaptronic blade. Reproduced from Rasuo B, Dinulovic M, Veg A, Grbovic A, Bengin A. Harmonization of new wind turbine rotor blades development process: a review. Renew Sustain Energy Rev 2014;39:874–82.

The Role of Energy Conversion

24

Solution: Length of blade, wind speed air density, and power coefficient are given while power produced by the wind turbine needs to be calculated. Assumption: Wind is flowing at constant speed. The swept area of the turbine can be calculated as A ¼ pr 2 A ¼ 3:14  482 A ¼ 7234:6 m2 The power produced by the turbine can be calculated using Eq. (34) 1 rAV 3 CP 2 1 Pavail ¼  1:23  7234:6  153  0:4 2 Pavail ¼ 6 MW

Pavail ¼

Discussion: The wind turbine can produce 6 MW of power. It is important to check the behavior of the turbine at different wind speeds because smooth operation of the wind turbine is more important. In actuality, a design with high power output at high wind speeds requires more maintenance.

4.1.5.3

Ocean Current Energy

The continuous directed movement of seawater generated by the forces like solar heating, wind, salinity differences, breaking waves, wind, salinity and temperature differences, and the gravitational pull of the sun and moon causes ocean current energy. The direction and strength of the current depend heavily on the depth of contours, configurations of shoreline, and the interactions with other currents. Primarily, the ocean currents are horizontal water movements and can flow for miles and create a global chain. The ocean currents’ influence on the temperature of the regions also plays an important role in forecasting the climate of many regions on the earth. The kinetic energy of the movement of the tides can be harnessed by underwater turbines, which are similar to small wind turbines. The ocean current turbines are usually installed on the seabed experiencing high velocities of ocean currents. The electricity is generated through the rotation of the tidal turbine blades due to the movement of the water. The conversion of wind energy is an aerodynamic process whereas, the conversion of water movement is hydrodynamic process. Ocean current turbines can convert the water velocities of as low as 1 m/s to generate electrical power [37]. An ocean current turbine is represented in Fig. 25, where the ocean current is seen to rotate the blade of the turbine, converting the kinetic energy of the current to electrical power. An ocean current turbine can produce the same power as a wind turbine with a larger rotor size due to the large density difference between seawater and air (water is typically 800 times denser than air). It has been determined that for the same size turbine, a water speed of around one-tenth of the speed of the wind can generate the same electrical power [37]. The power density in flowing water such as an ocean current can be expressed as follows: Pa ¼

1 3 rV 2

ð35Þ

For a turbine having Zelec as the conversion efficiency for ocean current energy to electricity, the electrical power density generated is found as Pe ¼ Zelec Pa

Ocean level

Ocean current turbine

Ocean current

Ocean floor Fig. 25 Schematic illustration of the working of ocean current turbine.

ð36Þ

The Role of Energy Conversion

25

Oil prices greatly influence efforts to harness ocean energy, with activity increasing when oil prices are high and vice versa. Recent concerns about increasing global CO2 emissions and other environmental issues have increased efforts to derive energy from oceans. Illustrative Example 11: A turbine of radius 6 m is generating electricity driven through an ocean current moving with a speed of 3.7 m/s. Calculate the electrical power density and the total electrical power generation. Assume that 37% of the ocean current can be converted to the electrical power. The density of ocean water is 1028 kg/m3. Solution: Assumptions: Conversion of ocean current to electricity is assumed to be 37%. The power density in the ocean current can be calculated by using Eq. (35), as Pa ¼

  1 3 rV ¼ 0:5  1028  3:73 ¼ 26; 036 W=m2 2

Now Eq. (36) can be used to get the electrical power density generated using ocean current Pe ¼ 0:35 Pa ¼ 0:35  26; 036 ¼ 7713 W=m2 ¼ 9:113 kW=m2 The total electrical power generated can be evaluated as Ptotal;e ¼ Pe  A ¼ 7:713  p  62 ¼ 1030 kW Discussion: The total electrical power generated by the ocean current plant is 1.03 MW. The power produced is directly proportional to the radius and cube of velocity of the ocean current.

4.1.5.4

Solar Thermal Energy Conversions

Solar thermal energy has been used for space and water heating in the past. Additional solar thermal applications have been developed, which include refrigeration, air conditioning, crop drying, process heat for industries, and electric power generation. Solar collectors are divided in many categories depending on the geometry [38]. Solar thermal collector consists of an absorber, insulation, a trap, and a heat transfer medium. 1. 2. 3. 4.

Absorber: An absorber is made out of a thermally conducting dark surface. Insulation: Insulation reduces heat loss where it is placed. Trap: Trap allows to pass radiation with short wavelength while it blocks radiation with long wavelength. Heat transfer medium: Mediums like air, water, or oil transfer the solar heat to useful work.

The solar radiation is concentrated on the absorber with the help of reflectors. The advanced solar collectors are able to provide temperatures in the range of 10001C or even higher. The desired temperature and economics of the solar thermal application depend upon the design and the selection of working. Some types of solar thermal collectors are tabulated in Table 4 on the basis of their temperature range. Illustrative Example 12: The mass flow rate of isobutane in the closed PTSC cycle is 25 kg/s as shown in Fig. 26. Therminol VP-1 exits parabolic solar trough collectors at 202, then enters the solar heat exchanger at 3201C and leaves at 701C after heating up isobutane to 239.41C. The water at ambient pressure enters the condenser at 251C and leaves at 42.41C. Isobutane exits ORC turbine at 351C. The pressures at inlet and exit of the pump are 75 and 2500 kPa, respectively. Determine (1) net rate of work done by the ORC turbine; (2) the net rate of work done required to drive the pump; and (3) the energy and exergy efficiencies of the ORC turbine. Take the ambient temperature to be 251C, and the isentropic efficiencies of the turbine and pump to be 85% assuming no pressure drop across the solar heat exchanger. Solution: For the schematic of the closed PTSC system (Fig. 26), the rate balance equations can be written. For the solar heat exchanger the balance equations can be expressed as follows: _1 ¼m _ 2 ðfor Therminol VP Mass balance equation ðMBEÞ: m Table 4

1Þ

Types of solar collectors with concentration ratio and typical range of temperature

Type of collector

Concentration ratio

Typical range of temperature (1C)

Flat plate solar collector Flat plate collector with high efficiency Fixed concentrator Parabolic trough solar collector (PTSC) Parabolic dish collector Solar tower with central receiver

1 1 2–5 10–50 200–2000 200–2000

Z70 60–120 100–150 150–350 250–700 400–1000

Source: Reproduced from Goswami DY, Krieth F. Energy conversion. Boca Raton, FL: Taylor & Francis Group; 2007.

26

The Role of Energy Conversion

ORC turbine 3

1

Therminol VP-1

Solar heat exchanger

Solar collectors

Isobutane

2

Electricity generator 5

Pump 4

Condenser

6

7

8

Fig. 26 Schematic of a parabolic trough solar collector power generation cycle.

_4 _3¼m m

ðfor isobutaneÞ

_ 4 h4 ¼ m _ 2 h2 þ m _ 3 h3 _ 1 h1 þ m EBE: m _ 1 s1 þ m _ 4 s4 þ S_ gen;evap ¼ m _ 2 s2 þ m _ 3 s3 Entropy Balance Equation ðEnBEÞ: m _ des;HEX _ 1 ex1 þ m _ 4 ex 4 ¼ m _ 2 ex2 þ m _ 3 ex 3 þ Ex Exergy Balance Equation ðExBEÞ: m For the condenser, the balance equations can be written as _5¼m _6 MBE: m _7 _8 ¼m m

ðfor isobutaneÞ

ðwater circulation for coolingÞ

_ 7 h7 ¼ m _ 6 h6 þ m _ 8 h8 _ 5 h5 þ m EBE: m _ 5 s5 þ m _ 7 s7 þ S_ gen;cond ¼ m _ 6 s6 þ m _ 8 s8 EnBE: m _ des;cond _ 5 ex 5 þ m _ 7 ex 7 ¼ m _ 6 ex 6 þ m _ 8 ex 8 þ Ex ExBE: m For turbine, the balance equations can be written as _3¼m _5 MBE: m _ act;turb _ 3 h3 ¼ m _ 5 h5 þ W EBE: m _ 3 s3 þ S_ gen;turb ¼ m _ 5 s5 EnBE: m _ act;turb þ Ex _ des;turb _ 3 ex 3 ¼ m _ 5 ex5 þ W ExBE: m For pump, the balance equations can be written as _4¼m _6 MBE: m _ act;p ¼ m _ 4 h4 þ W _ 4 h4 EBE: m _ 4 s4 þ S_ gen;p ¼ m _ 4 s4 EnBE: m _ act;p ¼ m _ des;p _ 4 ex4 þ W _ 4 ex4 þ Ex ExBE: m For Therminol VP-1 one can obtain the following properties from EES. For state 1, ) h1 ¼ 601:4 kJ=kg T 1 ¼ 251C P1 ¼ 202 kPa

s1 ¼ 1:393 kJ=kg K

The Role of Energy Conversion

Table 5

27

Input and calculated data for the PTSC system in illustrative example 9

State no.

Fluid type

P (kPa)

_ (kg/s) m

T (1C)

h (kJ/kg)

s (kJ/kg K)

0 0 0 1 2 3 4 5 6 7 8

Water Isobutane Therminol VP-1 Therminol VP-1 Therminol VP-1 Isobutane Isobutane Isobutane Isobutane Water Water

101.3 101.3 101.3 202 202 2500 2500 75 75 101.3 101.3

– – – 35 35 25 25 25 25 25 25

25 25 25 320 70 239.4 35 75 35 25 42.4

104.8 598.9 20.14 601.4 93.17 1034 284.9 689.5 616.7 104.8 177

0.3669 2.513 0.06853 1.392 0.296 3.164 1.278 2.837 2.614 0.3669 0.602

Similarly for state 2, T 2 ¼ 701C P2 ¼ 202 kPa

)

h2 ¼ 93:17 kJ=kg s2 ¼ 0:296 kJ=kg K

For reference state enthalpy and entropy for Therminol VP-1 ) h0 ¼ 20:14 kJ=kg T 0 ¼ 251C P0 ¼ 101:321 kPa s0 ¼ 0:06853 kJ=kg K Also, for the reference state enthalpy and entropy for isobutane, ) h0 ¼ 598:7 kJ=kg T 0 ¼ 251C P0 ¼ 101:321 kPa s0 ¼ 2:513 kJ=kg K For reference state enthalpy and entropy for water )

h0 ¼ 104:8 kJ=kg s0 ¼ 0:3669 kJ=kg K

T 0 ðs1

s0 Þ ¼ 186:7 kJ=kg

T 0 ¼ 251C

P0 ¼ 101:321 kPa The specific exergy at state 1 can be calculated as ex 1 ¼ h1

h0

Similarly for state 2 ex2 ¼ 5:18 kJ=kg Table 5 provides the input and calculated process data required for the system in illustrative Example 9. 1. Using the energy rate balance for the turbine, the rate of work done by the turbine can be determined as _ act;turb ¼ m _ 3 ðh3 W

h5 Þ ¼ ð25 kg=sÞð1034

689:5ÞkJ=kg ¼ 8612:5 kW

2. Similarly, the volume of isobutane at state 6 using EES v4 ¼ 0:5775 m3 =kg _ act;pump ¼ v4 ðP4 W

P6 Þ ¼ 0:5775ð2500

75Þ ¼ 1400 kW

So, the net output work rate of the system is _ act;turb _ net ¼ W W

_ act;p ¼ 8612:5 W

1400 ¼ 7212:5 kW

3. The energy efficiency of the turbine can be found as Zen;orc ¼

_ net 7212:5 W ¼  0:85 ¼ 34:5% _ 1 ðh1 h2 Þ m 35ð601:4 93:17Þ

4. The exergy efficiency of the ORC turbine can be found as cen;orc ¼

_ net 7212:5 W ¼  0:85 ¼ 54:7% _ 1 ðex1 ex 2 Þ m 35ð325:4 5:18Þ

ex (kJ/kg) 0 0 0 325.4 5.18 241.5 54.14 5.61 12.25 0.0038 2.1

28 4.1.5.5

The Role of Energy Conversion Geothermal Energy Conversion

The word “geothermal” means heat of the earth, as it is adapted from the Greek word “geo” (earth) and “therme” (heat) [39]. The amount of heat and power that can be produced through geothermal source depends largely on the reservoir depth. More than 453K temperatures are possible to achieve near plate tectonic boundaries. Geothermal energy sources can be classified on the basis of the temperature as below 363K (low temperatures), between 363 and 423K (moderate temperatures), and above 423K (high temperatures) [15,40]. Intermediate and low temperatures are present with or without hydrothermal resources in continental settings [41]. Geothermal fields can be used for both direct utilization of heat or power production depending upon the reservoir temperature and the requirements. In binary cycles, hot geothermal fluid is not in direct contact with the turbine; instead it is used to heat up a secondary fluid, which vaporizes and runs the turbine. There are many technical variations associated with binary plants and Kalina cycles [42]. The geothermal hot fluid and the ORC working fluid circulate in separate closed loops. The organic Rankine cycle (ORC) is feasible even for the low temperature geothermal resource. The binary geothermal cycle is environmentally friendly as it does not produce any emissions. The basic layout of a geothermal power plant is depicted in Fig. 27. The utilization efficiency of the geothermal power plants can be found as Zu ¼

_ b ½hA m

_ net W h0 T0 ðsA

ð37Þ

S0 ފ

_ net represents the net power obtained through geothermal cycle; m _ b , hA, h0, T0, sA, and s0 represent mass flow rate of brine, where, W enthalpy of brine at inlet, enthalpy at dead state, dead state temperature, entropy at inlet, and entropy at dead state, respectively. Geothermal plants with huge capacity have been installed around the world. Geysers Geothermal Complex, installed in California with a capacity of 1517 MW, is one of the largest geothermal plants in the world, as shown in Fig. 28 [43]. The plant is actively producing 900 MW power. Some geothermal power plants in the world on the basis of the installed capacity for power production are tabulated in Table 6. Illustrative Example 13: Brine enters the heat exchanger at a temperature of 440K. The mass flow rate of brine is 95 kg/s and net power of the geothermal cycle is 1200 kW. Find the utilization efficiency at dead state temperature 25ºC. Solution: The mass flow rate of geothermal fluid and net power of the cycle are given while utilization efficiency need to be calculated. Assumption: Pressure losses in heat exchangers are negligible. Analysis: (1) The utilization efficiency of a geothermal cycle can be calculated using Eq. (37) as Zu ¼

ORC turbine

Electricity generator

8

10

4

12 Geotherrmal reservior

Preheater

Mixing chamber 11

14

15 5

6

7 Pump 2

13

Condenser

Pump 1

3

Flash seperator 9

S0 ފ

Superheater

2 1

_ b ½hA m

_ net W h0 T0 ðsA

Pump 3

Reinjection

Fig. 27 Schematic of a basic geothermal power plant. Reproduced from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

The Role of Energy Conversion

29

Fig. 28 Geysers Geothermal Complex, California, United States. Reproduced from The top 10 biggest geothermal power plants in the world. Available from: http://www.power-technology.com/features/feature-top-10-biggest-geothermal-power-plants-in-theworld/; 2017 [accessed 25.05.17].

Table 6

Top 10 geothermal power plants around the world

No.

Name

Location

Installed capacity (MW)

1 2 3 4 5 6 7 8 9 10

Geysers Geothermal Complex Larderello Geothermal Complex Cerro Prieto Geothermal Power Station Makban Geothermal Complex CalEnergy Generation’s Salton Sea Geothermal Plants Hellisheidi Geothermal Power Plant Tiwi Geothermal Complex Darajat Power Station Malitbog Geothermal Power Station Wayang Windu Geothermal Power Plant

California, United States Italy Mexico Philippines United States Iceland Philippines Indonesia Philippines Indonesia

1517 769 720 458 340 303 289 259 232.5 227

Source: Reproduced from The top 10 biggest geothermal power plants in the world. Available from: http://www.power-technology.com/features/feature-top-10-biggest-geothermalpower-plants-in-theworld/; 2017 [accessed 25.05.17].

The enthalpy and entropy values at inlet temperature 440K and dead state 298K are determined using EES as hA @440K ¼ 705 kJ=kg sA @440K ¼ 2:009 kJ=kg K h0 @298K ¼ 103:93 kJ=kg s0 @298K ¼ 0:36384 kJ=kg K Zu ¼

1250 ¼ 11:92% 95  110:39

Discussion: Typical geothermal plants have thermal efficiency of about 15% but the utilization efficiency is quite low. This is because preheat is required to heat up the working fluid and the preheater accounts for large irreversibilities due to high finite temperature difference.

4.1.6

Case Studies

In this section, there are multiple case studies presented to include the analyses of selected renewable energy-based integrated systems for multigeneration. Case study 1 presents the advantages of integrating thermoelectric devices in multigeneration systems;

The Role of Energy Conversion

30

Solar field

Parabolic trough solar collectors

Heat

Solar heat exchanger

Organic rankine cylce

Heat

Electricity

Heat Heat Solar field

Photovoltaic panels Electricity

Heat Cooling Heat

Heat pump

Absorption refrigeration chiller

Electricity Thermoelectric generator

Electrolyzer

Heating

Thermoelectric cooler

Heat

Thermal energy storage system

Cooling

Heat

Heat Condenser Heat

Fig. 29 Flow diagram of multigeneration system including thermoelectric generator.

whereas, solar and geothermal-based renewable energy are combined in Case study 2 for improved performance of the multigeneration system. These studies include the assessment of subsystems as well as overall systems.

4.1.6.1

Case Study 1

The primary objective of this presented case study is to improve the performance of a solar energy-based integrated system through a unique integration of thermoelectric devices (see Fig. 29). The novel integrated multigeneration system is assessed thermodynamically and an exergy analysis is conducted. The major subsystems of the proposed system are an absorption chiller an ORC, an electrolyzer, and a heat pump. The presented system generates electricity for off grid areas of countries with abundant amount of solar radiations like Saudi Arabia. The heat of the PV panels is extracted to run thermoelectric devices, as well as the efficiency of the PV panels is increased due to decrease in the operating temperature. The absorption chiller delivers the cooling to the cold store, heat is supplied to industries, and thermal energy storage (TES) system supplies uninterrupted hot water to industrial and domestic users even in the absence of solar energy (at night), and hydrogen for chemical and petrochemical industries. Islam et al. [44] made following assumptions in order to investigate the performance of the proposed multigeneration system.

• • • • • • • • • • •

System operating conditions are steady. The dead state temperature and pressure are 298K and 101.325 kPa, respectively. Negligible or no changes in kinetic and potential energies. R410a and water are treated as actual fluids. All pumps and turbines are adiabatic. Isobutane is the working fluid used in ORC turbine. All pumps and turbines operate with 85% isentropic efficiencies. TES system losses 10% of heat while discharging period. All pressure losses are neglected. The typical evaporator used by Al-Ali and Dincer [45] is used. Parabolic solar collectors operate with 80% efficiency.

4.1.6.1.1

Thermodynamic assessment

Here, the energy and exergy efficiencies of all subunits of the multigeneration are presented. The model parameters used for the thermodynamic analysis of the PV panels are tabulated in Table 7. The figure of Merit has a great influence on the efficiency of the thermoelectric device and can be written as pffiffiffiffiffiffiffiffiffiffiffiffiffiffi DTð 1 þ ZT 1Þ  ZTEG ¼ pffiffiffiffiffiffiffiffiffiffiffiffiffiffi ð38Þ Th 1 þ ZT þ Tc =Th

where, ZT represents figure of Merit, Th denotes high temperature junction, and Tc symbolizes low temperature junction of the thermoelectric. Magnitude of figure of merit is selected as “1” for the calculations. The COP associated with thermoelectric cooler (TEC) can be calculated as [46] COPTEC ¼

Qc PTEG

where, Qc and PTEG represent the amount of heat absorbed and power input to TEC.

ð39Þ

The Role of Energy Conversion

Table 7 hydrogen

31

Parameters of commercial photovoltaic (PV) cell for the production of 0.75 m2 36 (commercial) 530 2.5% 850 W/m2 and 251C

PV module area (Ac) No. of PV cells No. of PV modules Wiring losses (Ohm) Standard operating conditions of PV

Source: Compiled from Paola A, Fabrizio Z, Fabio O. Techno-economic optimisation of hydrogen production by PV – electrolysis: ‘RenHydrogen’ simulation program. Int J Hydrogen Energy 2011;36(2):1371–81.

The energetic efficiency of ORC turbine can be found as Zt ¼ The overall energy efficiency of TES can be calculated as ZTES ¼

Hd Ha

_t W _ i hi þ m _ e he m

ð40Þ _ Q

Hc ¼1 Hb

Ha

ð41Þ

Hb

The work rate required to drive the compressor of the heat pump can be expressed as _ c ¼m _ e he W

_ i hi m

ð42Þ

_ e he m _ i hi m _ _ W c þ W pump

ð43Þ

The COP of heat pump can be defined as COPen;hp ¼ The COP of the absorption chiller can be found as COPen;chiller ¼

_ eva;in heva;in _ eva;out heva;out m m _ gen;out hgen;out _ gen;in hgen;in m m

ð44Þ

The net amount of work rate obtained by the system can be written as _ tþW _ TEG _ net ¼ W W

_ TEC W

_ P;overall W

_c W

ð45Þ

The energy efficiency of multigeneration system can be expressed as Zoverall ¼

_ hp þ Q _ cool þ Q _ HE þ Q _ cond þ Q _ TES _ net þ Q W _ solar Þ ðEGH þ Q

ð46Þ

where, EGH denotes the electricity generated through PV panels. The efficient use of energy is possible through developing exergy analysis strategies. Exergy analysis identifies the components that are responsible for huge losses, hence it becomes possible to extract a larger amount of useful work. The correlation between energy, exergy, environment, energy policy making, and ecological development is established by Dincer and Rosen [47]. Many researchers has widely used exergy analysis for simulation, design and performance evaluation of energy systems [23,48–51]. The exergy efficiencies of the subsystems and the overall system can be found by assuming no changes in kinetic and potential energies. ct ¼

_t W _ e ex e _ i ex i þ m m

ð47Þ

The exergetic COP of the heat pump driven by R134a can be found as COPex;hp ¼

_ e ex e m _ i exi m _ pump 4 _ c þW W

ð48Þ

The exergetic COP of the absorption refrigeration chiller can be found as COPex;chiller ¼

_ eva;out heva;out ðm _ gen;in hgen;in ðm

ex 47 cTES ¼ ex 46

_ eva;in heva;in Þ TTeva0 m _ gen;out hgen;out Þ 1 m

ex48 ¼1 ex44

_ 1 Q

ðex46

T0 Ts



ex 44 Þ

1

T0 Ts

ð49Þ

ð50Þ

The Role of Energy Conversion

32

Table 8 Work rate, efficiency, and coefficient of performances of subsystems and multigeneration system Component

Values

W_ DORC;excluding TEG W_ DORC;including TEG W_ Dhp W_ DTEG W_ DTEC;in W_ D

4331 kW 4310 kW 1537 kW 20.8 kW 20.8 kW 19.1 kW 5.6% 5.9% 10.1% 10.7% 16.73% 33.2% 16.7% 33.1% 1.5 0.71 3.3 0.46 50.6% 39.8% 51.3% 40.3% 671.8 kW 1630.66 kW

TEC;out

ZPV,excluding TEG cPV,excluding TEG ZPV,including TEG cPV,including TEG ZORC,excluding TEG cORC,excluding TEG ZORC,including TEG cORC,including TEG COPen,chiller COPex,chiller COPen,hp COPex,hp Zmultigen,excluding TEG cmultigen,excluding TEG Zmultigen,including TEG cmultigen,including TEG W_ Dcomp W_ Dpumps

Source: Based on data provided by Islam S, Dincer I, Yilbas BS. Energetic and exergetic performance analyses of a solar energy-based integrated system for multigeneration including thermoelectric generators. Energy 2015;93:1246–58.

Table 9 Significant exergy destruction occurring in subsystems of the proposed system Subsystem

Rate of exergy destruction ðE x_ D Þ ðkWÞ

Parabolic trough solar collectors Absorption refrigeration unit Heat pump Generator of absorption refrigeration unit Compressor of heat pump unit Condenser of ORC Thermal energy storage (TES) unit

13,032 937 275 247 195 111 26.97

Source: Modified from Islam S, Dincer I, Yilbas BS. Energetic and exergetic performance analyses of a solar energy-based integrated system for multigeneration including thermoelectric generators. Energy 2015;93:1246–58.

The exergy efficiency of the overall system can be calculated as

coverall ¼

_ hp 1 _ net þ Q W

T0 Ts;hp



_ cool þQ



T0 Ts;eva



_ cond 1 1 þQ _ solar 1 EGH þ Q

T0 Ts;cond



T0 Ts;solar



_ HE 4 1 þQ

T0 Ts;HE



_ TES 1 þQ

T0 Ts;TES



ð51Þ

where  _ solar ¼ m _ PTSC;in hPTSC;in Q

_ PTSC;out hPTSC;out m



ð52Þ

The Role of Energy Conversion

Geothermal fluid

Geothermal cycle

Heat

Heat Super-heater

Organic rankine cycle 1

33

Electricity

Heat Solar field

Heat

Parabolic trough solar collectors

Solar heat exchanger

Heat

Organic rankine cycle 2

Electricity

Heat Fig. 30 Schematic of combined solar and geothermal energy-based integrated system for single generation. Modified from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

Geothermal fluid

Heat

Geothermal cycle

Super-heater

Heat

Organic rankine cycle 1 Electricity

Heat

Organic rankine Electricity cycle 2

Heat Solar field

Parabolic trough solar collectors

Heat

Solar heat exchanger Heat

Heat Absorption refrigeration chiller

Cooling

Fig. 31 Flow diagram of combined solar and geothermal energy-based integrated system for cogeneration. Modified from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

Geothermal fluid

Heat

Geothermal cycle

Heat

Organic rankine cycle 1 Electricity

Heat

Organic rankine Electricity cycle 2

Super-heater Heat

Solar field

Parabolic trough solar collectors

Heat

Solar heat exchanger Heat

Heating

Heat pump

Heat Absorption Refrigeration Chiller

Cooling

Fig. 32 Flow diagram of combined solar and geothermal energy-based integrated system for trigeneration. Modified from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

4.1.6.1.2

Results and discussion

The overall energy and exergy efficiencies of the system presented in Fig. 20 with thermoelectric devices are 51.3% and 40.3%, respectively. Both efficiencies of the integrated multigeneration system are less excluding thermoelectric devices because of the reduction in the efficiencies of PV panels as a result of increase in temperature throughout the day. The exclusive integration of TEC driven PV cooling system enhances the energy and exergy efficiencies of PV panels, which in turn enhances the performance of the overall system.

34

The Role of Energy Conversion

The organic Rankine turbine of the system produces 4331 kW excluding thermoelectric generator (TEG) and 4310 kW including TEG. Despite the fact that the organic Rankine turbine including TEG provides slightly less amount of work done, the overall system is still more efficient. This is because of the presence of the PV cooling system, which increases the energy and exergy efficiencies of PV panels in the order of 4.5% and 4.7%, respectively. The COP of the absorption refrigeration chiller and heat pump are found to be 1.5 and 3.3, respectively. The high magnitude of heat pump reflects how efficiently the waste heat is conserved in the proposed system and this is another contributing factor for the increased performance of the overall system. Table 8 tabulates the main findings of this case study. The significant rate of exergy destruction occurring in major subunits/components of the overall system is tabulated in Table 9. It is obvious that the major irreversibilities are associated with the solar cycle of the overall system followed by the LiBr driven absorption refrigeration chiller. Moreover, the significant magnitude of irreversibilities in the heat pump is due to the exergy destruction occurring in the compressor and the TES system accounts for the least rate of exergy destruction due to less temperature difference between charging and discharging.

4.1.6.2

Case Study 2

In this case study, an integrated multigeneration system based on solar and geothermal energy is assessed thermodynamically (see Figs. 21–24). The multigeneration system combines two types of renewable energies and provides heating for domestic users or industries, space cooling for cold stores, dryer for drying of crop, two TES systems, and electrical energy. Fig. 30 represents a single generation system that produces electricity through combined geothermal and solar energy. The proposed system produces electricity and cooling after incorporating the absorption refrigeration chiller with the ORC 2 (see Fig. 31). The waste heat is effectively conserved and it produces cooling for cold storage. The proposed system produces electricity, cooling, and space heating after incorporating the absorption refrigeration chiller with the ORC 2 and a heat pump with solar heat exchanger (see Fig. 32). The waste heat rejected by the solar heat exchanger is utilized to produce space heating through R134a driven heat pump. The proposed system produces electricity, cooling, space heating, drying system, and TES system after incorporating the additional subunits dryer and TES system to the trigeneration system to conserve waste heat as depicted in Fig. 33.

4.1.6.2.1

Thermodynamic assessment

The energy efficiencies of the single generation system to multigeneration system can written as Zsingen ¼  Zcogen ¼

_ net;singen W  _ solar _ i hi þ Q m

ð53Þ

_ cool _ net;cogen þ Q W _ _ i hi þ Qsolar Þ ðm

ð54Þ

Thermal energy storage system 1 Heat Geothermal fluid

Geothermal cycle

Heat

Heat

Organic rankine cycle 1 Electricity

Heat

Organic rankine Electricity cycle 2

Super-heater Heat

Solar field

Parabolic trough solar collectors

Heat

Solar heat exchanger Heat

Heating Heat pump

Heat

Crop dryer

Heat Absorption Refrigeration chiller

Cooling

Heat Thermal energy storage system 2 Fig. 33 Flow diagram of combined solar and geothermal energy-based integrated system for multigeneration. Modified from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

The Role of Energy Conversion

Ztrigen ¼

Zmultigen ¼

_ hp þ Q _ cool _ net;trigen þ Q W _ _ ðmi hi þ Qsolar Þ

35

ð55Þ

_ hp þ Q _ cool þ Q _ Dryer þ Q _ TES _ net;multigen þ Q W _ _ i hi þ Qsolar Þ ðm

ð56Þ

The exergy efficiencies of all proposed systems from single generation to multigeneration system can calculated as csingen ¼

_ net;singen W _ solar 1 _ i exi þ Q m

T0 Ts;sun

ð57Þ





_ cool T0 _ net;cogen þ Q 1 W Ts;evp

ccogen ¼ T0 _ solar 1 _ i exi þ Q m Ts;sun ctrigen ¼

_ hp 1 _ net;trigen þ Q W

ð58Þ



_ cool T0 þQ Ts

T0 _ _ i ex i þ Qsolar 1 Ts;sun m T0 Ts

1

The exergy efficiency for the multigeneration or overall system can be found as



_ hp 1 T0 þ Q _ Dryer 1 _ cool T0 1 þ Q _ net;multigen þ Q W Ts Ts

cmultigen ¼ T0 _ solar 1 _ i ex i þ Q m

T0 Ts

ð59Þ



_ TES 1 þQ

Ts;sun

_ solar ¼ ðm _ i hi where Q

4.1.6.2.2

T0 Ts



ð60Þ

_ e he Þ. m

Results and discussion

Islam and Dincer [22] assessed the presented system thermodynamically through energy and exergy methodologies. The energy and exergy efficiencies of the single generation, cogeneration, trigeneration and multigeneration system are found to be 22% and 54%, 34% and 60%, 44% and 60.4%, and 51% and 62%, respectively as represented in Fig. 34. It is important to note that the energy efficiency of trigeneration system is higher than the cogeneration system but the exergy efficiencies of the cogeneration and trigeneration systems are almost same because of the exergy losses in the compressor of the heat pump. The rate of exergy destruction in each subunit of the proposed system is displayed in Fig. 35. It can be observed that the magnitude of the rate of exergy destructions occurring in solar cycle is the highest followed by the geothermal cycle. The heat pump accounts for the third highest rate of exergy destruction mainly because of the exergy losses in the compressor. The generator of the absorption chiller is found to have the fourth highest exergy destruction because of the increased temperature difference across its inlet and exit states. The rate of exergy destruction in the generator is possible to reduce either by incorporating an efficient absorber or by operating it with a low temperature source. The fifth highest rate of exergy destruction is found to be in the Energy efficiencies Exergy efficiencies

Energy and exergy efficiencies

0.7 0.6 0.5 0.4 0.3 0.2 0.1 0 Multigeneration system

Trigeneration system

Cogeneration system

Single generation system

Fig. 34 Graphical representation of efficiencies of single generation, cogeneration, trigeneration, and multigeneration system. Data from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

36

The Role of Energy Conversion

Exergy destruction rate (kW)

10,000

1000

100

10

eo G

So

la

th e

rc ol

le c rm tors al cy Ab He cl e so at rp pu tio m p n C chil om la pr r e H ea sso tp r um p TE S 2 TE S 1 H E 2 Su Dr y pe e rh r ea te r H E 3

1

Fig. 35 The rate of exergy destruction occurring in subsystems of the multigeneration system. Data from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

Exergy efficiency

0.64

Exergy efficiency of multigeneration system

0.62 0.6

Exergy efficiency of trigeneration system

0.58 0.56

Exergy efficiency of cogeneration system

0.54 0.52 0.5 297

Exergy efficiency of single generation system

302 307 Ambient temperature (K)

Fig. 36 Effect of ambient temperature on the exergy efficiencies of proposed system. Data from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

compressor of heat pump as it operates at a larger temperature difference. The heat losses in other subunits like TES, heat exchangers, and dryer are less, therefore, the rate of exergy destruction associated with these is insignificant. The variation in the reference environment conditions affects the performance of the proposed multigeneration system. Fig. 36 displays the fluctuation in the exergy efficiencies of the single generation, cogeneration, trigeneration, and multigeneration systems against the variation in the ambient temperature from 298 to 310K. The exergy efficiencies of single generation, cogeneration, trigeneration, and multigeneration system are enhanced from 54% to 57.6%, 60% to 62.8%, 60.4% to 63%, and 62.2% to 63.5%, respectively. The exergy efficiencies are enhanced because of less temperature difference as a result of increased ambient temperature. The rate of variation in the exergy destructions in subunits of the multigeneration system against the fluctuation in the ambient temperature from 298 to 310K are graphically represented in Fig. 37. The magnitude of exergy destruction rate decreases with the increase in the ambient temperature. This is due to the fact that the temperature difference between operating conditions and ambient conditions is reduced with the increase in ambient temperature.

4.1.7

Future Directions

The exclusive integration of more than one energy conversion method is of great importance to meet the increasing demand of electricity in upcoming years. Further advancements in the best possible integration of more than one energy resource and waste energy harness methods will enhance the energy and exergy efficiencies of the conversion methods. The integration of TEG and their best possible configuration necessitates further research. Moreover, development of improved thermoelectric materials is vital to achieve increased output from thermoelectric devices.

The Role of Energy Conversion

Rate of exergy destruction in;

10,000

Exergy destruction rate (kW)

37

Solar cycle

1000

Geothermal cycle Heat pump

100

Absorption chiller TES 2

10

TES 1 Dryer

1 297

302 307 Ambient temperature (K)

Fig. 37 Effect of ambient temperature on exergy destruction rate in subunits of the multigeneration system. TES: thermal energy storage. Data from Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43.

In addition to this, gasification of the oil and biomass waste integrated with multigeneration system is an attractive option to convert the waste into useful energy with less emissions. Aspen Plus simulation of the animal waste-based dual gasifier reveals that CO2 emissions can be reduced by further using the produced CO2 as a gasification agent [52]. Farrukh et al. [53] combined solar and biomass energy sources and optimized multigeneration system. They found the newly developed hybrid system more efficient and economical as compared to the operation of solar and biomass systems individually. Recently, Islam et al. [54] developed a novel renewable energy-based multigeneration system and reported overall increase in the energy and exergy efficiencies. Moreover, the energy and exergy efficiencies associated with the ORC were also increased due to multigeneration. Shim et al. [55] proposed an enhanced energy conversion device using ultrafast magnetic cooling effect phenomenon. The magnetic cooling effect originates from a large change in entropy by the forced magnetization alignment, which has long been considered to be utilized as an alternative environment-friendly cooling technology compared to conventional refrigeration. Wu et al. [56] developed a novel energy conversion method based on hydrogel material for self-powered sensor system applications that can harvest energy from environment vibrations and supply power to sensors without any external power source. Moreover, the integration of various types of renewable energy systems with thermoelectric devices including the assessment of their best configuration to reduce pollution is another direction for research. Ali and Yilbas developed an innovative design for an extended leg TEG with tapering and segmented pin configuration for improved thermal performance [57].

4.1.8

Concluding Remarks

Several energy conversion methods are presented and discussed from energy, exergy, and environmental perspectives. There are advantages as well as disadvantages associated with each method of energy conversion. The novel integration of two or more renewable energy conversion methods can help to moderate or swap the usage of fossil fuels for electricity generation. The high initial investment is one of the major challenges in the expansion and commercialization of the renewable energy-based integrated multigeneration systems. The profitability of such systems can be achieved through the advances in the technologies to extract increased amount of useful energy from a source through multigeneration and manufacturing of tough materials that can withstand increased temperatures. The transportation and effective utilization of the byproducts such as hydrogen and other chemicals is another challenge associated with such systems. Case studies showed that integrating two renewable energy resources is advantageous to generate multiple outputs with high efficiency.

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The Role of Energy Conversion

Smith JM, Van Ness HC, Abbott MM. Introduction to chemical engineering thermodynamics. 7th ed. New York, NY: McGraw Hill; 2005. Demirel Y. Nonequilibrium thermodynamics. transport and rate processes in physical, chemical and biological systems. 2nd ed. Amsterdam: Elsevier; 2007. Hargreaves CM. The philips stirling engine. Amsterdam: Elsevier; 1991. Coskun C, Oktay Z, Dincer I. Investigation of some renewable energy and exergy parameters for two geothermal district heating systems. Int J Exergy 2011;8(1):1–15. Sheng C, Azevedo JL. Estimating the higher heating value of biomass fuels from basic analysis data. Biomass and Bioenergy 2005;28(5):499–507. Department of Energy. Where does the energy go? Advanced technologies and energy efficiency. US Department of Energy; 2009. Lynd LR, Larson E, Greene N, et al. The role of biomass in America’s energy future: framing the analysis. Biofuels Bioprod Biorefining 2009;3(2):113–23. Smil V. Energy at the cross roads global perspective and uncertainties. Cambridge: MIT Press; 2003. Moran MJ, Shapiro HB, Boettner DD, Bailey M. Fundamentals of engineering thermodynamics. 7th ed. New York, NY: John Wiley and Sons Ltd; 2011. Absorption Heat Pumps. Available from: https://energy.gov/energysaver/absorption-heat-pumps; 2017 [accessed 15.04.17]. Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43. Suleman F, Dincer I, Agelin-Chaab M. Development of an integrated renewable energy system for multigeneration. Energy 2014;78:196–204. Islam S, Dincer I, Yilbas BS. System development for solar energy-based hydrogen production and on-site combustion in HCCI engine for power generation. Sol Energy 2016;136:65–77. Moran MJ, Shapiro HN. Fundamentals of Engineering Thermodynamics. 6th ed. New York, NY: Wiley; 2008. Cohce MK, Dincer I, Rosen MA. Energy and exergy analyses of a biomass-based hydrogen production system.Bioresour Technol 2011;102(18):8466–74. Capareda S. Introduction to biomass energy conversions. Boca Raton, FL: CRC Press, Taylor & Francis Group; 2013. Bioenergy conversion technologies. Available from: http://www.wgbn.wisc.edu/conversion/bioenergy-conversion-technologies; 2017 [accessed 29.05.17]. Mujeebu MA, Jayaraj S, Ashok S, Abdullah MZ, Khalil M. Feasibility study of cogeneration in a plywood industry with power export to grid. Appl Energy 2009;86(5):657–62. García R, Pizarro C, Lavín AG, Bueno JL. Biomass sources for thermal conversion. Techno-economical overview. Fuel 2017;195:182–9. Kalinci Y, Hepbasli A, Dincer I. Exergoeconomic analysis and performance assessment of hydrogen and power production using different gasification systems. Fuel 2012;102:187–98. Riaza J, Gibbins J, Chalmers H. Ignition and combustion of single particles of coal and biomass. Fuel 2017;202:650–5. Soltani R, Dincer I, Rosen MA. Thermodynamic analysis of a novel multigeneration energy system based on heat recovery from a biomass CHP cycle. Appl Therm Eng 2015;89:90–100. Taheri MH, Mosaffa AH, Garousi Farshi L. Energy, exergy and economic assessments of a novel integrated biomass based multigeneration energy system with hydrogen production and LNG regasification cycle. Energy 2017;125:162–77. The world’s biggest wind turbine. Available from: http://www.dailymail.co.uk/sciencetech/article-4342966/Wind-turbine-world-s-biggest-722-feet.html; 2017 [accessed 02.06.17]. Rasuo B, Dinulovic M, Veg A, Grbovic A, Bengin A. Harmonization of new wind turbine rotor blades development process: a review. Renew Sustain Energy Rev 2014;39:874–82. Meisen P, Loiseau A. Ocean energy technologies for renewable energy generation. Global Energy Network Institute. Available from: http://citeseerx.ist.psu.edu/viewdoc/ download?doi=10.1.1.454.792&rep=rep1&type=pdf; 2009 [accessed 29.10.17]. Goswami DY, Krieth F. Energy conversion. Boca Raton, FL: Taylor & Francis Group; 2007. Ahmadi P, Dincer I, Rosen MA. Energy and exergy analyses of hydrogen production via solar-boosted ocean thermal energy conversion and PEM electrolysis. Int J Hydrogen Energy 2013;38(4):1795–805. Ghasemi H, Sheu E, Tizzanini A, Paci M, Mitsos A. Hybrid solar–geothermal power generation: optimal retrofitting. Appl Energy 2014;131:158–70. Geothermal electricity. Available from: http://www.geoelec.eu/about-geothermal-electricity/; 2014 [accessed 25.11.14]. Kalina AI, Leibowitz HM. Application of the Kalina cycle technology to geothermal power generation. Geotherm Resour Counc Trans 1989;13:605–11. The top 10 biggest geothermal power plants in the world. Available from: http://www.power-technology.com/features/feature-top-10-biggest-geothermal-power-plants-in-theworld/; 2017 [accessed 25.05.17]. Islam S, Dincer I, Yilbas BS. Energetic and exergetic performance analyses of a solar energy-based integrated system for multigeneration including thermoelectric generators. Energy 2015;93:1246–58. Al-Ali M, Dincer I. Energetic and exergetic studies of a multigenerational solar-geothermal system. Appl Therm Eng 2014;71:16–23. Yu J, Wang B. Enhancing the maximum coefficient of performance of thermoelectric cooling modules using internally cascaded thermoelectric couples. Int J Refrig 2009;32(1):32–9. Dincer I, Rosen MA. Energy, environment and sustainable development. Oxford: Elsevier; 2013. Khaliq A, Kumar R, Dincer I. Performance analysis of an industrial waste heat based trigeneration system. Int J Energy Res 2009;33:737–44. Ahmadi P, Dincer I, Rosen MA. Performance assessment and optimization of a novel integrated multi-generation system for residential buildings. Energy Build 2013;67:568–78. Islam S, Dincer I. Comparative performance study of an integrated air-cycle refrigeration and power generation system. Appl Therm Eng 2016; Khalid F, Dincer I, Rosen MA. Energy and exergy analyses of a solar-biomass integrated cycle for multigeneration. Sol Energy 2015;112:290–9. Fernandez-Lopez M, Pedroche J, Valverde JL, Sanchez-Silva L. Simulation of the gasification of animal wastes in a dual gasifier using Aspen Pluss. Energy Convers Manag 2017;140:211–7. Khalid F, Dincer I, Rosen MA. Thermoeconomic analysis of a solar-biomass integrated multigeneration system for a community. Appl Therm Eng 2017;120:645–53. Islam S, Dincer I, Yilbas BS. Analysis and assessment of a biomass energy-based multigeneration system with thermoelectric generators. Energy Fuels 2017;31 (10):10901–15. Shim J-H, Syed AA, Kim C-H, et al. Ultrafast giant magnetic cooling effect in ferromagnetic Co/Pt multilayers. Nat Commun 2017;8(1):796. Wu X, Li G, Lee D-W. A novel energy conversion method based on hydrogel material for self-powered sensor system applications. Appl Energy 2016;173:103–10. Ali H, Yilbas BS. Innovative design of a thermoelectric generator of extended legs with tapering and segmented pin configuration: Thermal performance analysis. Appl Therm Eng 2017;123:74–91.

Further Reading Akrami E, Chitsaz A, Nami H, Mahmoudi SMS. Energetic and exergoeconomic assessment of a multi-generation energy system based on indirect use of geothermal energy. Energy 2017;124:625–39. Bejan A, Tsatsaronis G, Moran MJ. Thermal design and optimization. New York, NY: John Wiley & Sons; 1996. Capareda S. Introduction to biomass energy conversions. Boca Raton, FL: CRC Press, Taylor & Francis Group; 2013. Cengel YA, Boles MA. Thermodynamics an engineering approach. 7th ed. New York, NY: McGraw-Hill; 2011. Demirel Y. Nonequilibrium thermodynamics. Transport and Rate Processes in Physical, Chemical and Biological Systems, 2007; 2nd ed. Amsterdam: Elsevier; 2007.

The Role of Energy Conversion

Demirel Y. Energy: production, conversion, storage, conservation and coupling. London, Heidelberg, New York, Dordrecht: Springer; 2012. Dincer I. Refrigeration systems and applications. 3rd ed. London: John Wiley and Sons Ltd; 2017. Dincer I, Rosen MA. Energy, Environment And Sustainable Development. Oxford: Elsevier; 2013. Dincer I, Rosen MA, Ahmadi P. Optimization of Energy Systems. New York, NY: John Wiley and Sons Ltd; 2017. Goswami Y, Dharendra , Krieth F. Energy conversion. Boca Raton, FL: Taylor & Francis Group; 2007. Moran MJ, Shapiro HB, Boettner DD, Bailey M. Fundamentals of engineering thermodynamics. 7th ed. New York, NY: John Wiley and Sons Ltd; 2011.

Relevant Websites http://www.fao.org/docrep/u2246e/u2246e02.htm Basic Energy Concepts. http://energyeducation.ca/encyclopedia/Energy_conversion_technology Energy Education. https://ocw.mit.edu/courses/mechanical-engineering/2-60-fundamentals-of-advanced-energy-conversion-spring-2004/ Massachusetts Institute of Technology. http://study.com/academy/lesson/types-of-energy-conversions.html Study.com: Types of Energy Conversions. https://www.teachengineering.org/lessons/view/cla_lesson4_forms_states_conversions Teach Engineering.

39

4.2 Heat Exchangers Almıla G Yazıcıog˘lu, Middle East Technical University, Ankara, Turkey Selin Aradag˘, Ece Aylı, Gizem Gülben, and Sadık Kakaç, TOBB University of Economics and Technology, Ankara, Turkey r 2018 Elsevier Inc. All rights reserved.

4.2.1 Introduction 4.2.1.1 Classification of Heat Exchangers 4.2.1.1.1 Type of interaction 4.2.1.1.2 Recuperation/regeneration 4.2.1.1.3 Flow configuration 4.2.1.1.4 Heat transfer mechanism 4.2.1.1.5 Geometry 4.2.1.2 Heat Exchanger Selection 4.2.1.3 Recent Developments 4.2.1.3.1 Microscale heat exchangers 4.2.1.3.2 Nanofluids 4.2.2 Background/Fundamentals 4.2.2.1 Design Methods for Heat Exchangers 4.2.2.1.1 Log mean temperature difference method 4.2.2.1.2 Effectiveness-number of transfer units method 4.2.2.2 Background on Gasketed-Plate Heat Exchangers 4.2.3 Specifics of Gasketed-Plate Heat Exchanger Design 4.2.4 Experimental Work and Software Development 4.2.4.1 Experiments 4.2.4.2 Computer Software Development 4.2.5 Case Study 4.2.6 Validation and Verification 4.2.6.1 Validation of the Software Using Experimental Data 4.2.6.2 Verification of the Computer Software With Correlations in Literature 4.2.7 Future Directions 4.2.8 Closing Remarks Acknowledgments References Further Reading Relevant Websites

Nomenclature A A1 A1,p b cp C C* Dp De Dh f F Gc h k Lh Lp Lv Lw

40

2

Area, m Actual effective area, m2 Projected plate area, m2 Mean channel spacing, m Specific heat, J kg 1 K 1 Heat capacity rate, W K 1 Heat capacity rate ratio Port diameter, mm Equivalent diameter, m Hydraulic diameter, m Fanning friction factor LMTD correction factor Channel mass velocity, kg m 2 s 1 Heat transfer coefficient, W m 2 K 1 Thermal conductivity, W m 1 K 1 Horizontal port distance, m Length of the plate between the ports, m Vertical port distance, m Plate width, m

_ m Ncp Ne Np Nt NTU Nu OS p Pr DP Q Re R t T DTlm U _p W

41 41 41 41 42 42 42 45 45 46 48 49 49 50 50 51 55 56 56 56 58 59 59 60 65 67 67 67 69 69

Mass flow rate, kg s 1 Number of channels per pass Number of plates for heat transfer Number of passes Total number of plates Number of transfer units Nusselt number Over surface Plate pitch, m Prandtl number Pressure drop, kPa Heat transfer rate, W Reynolds number Thermal resistance, m2 K1 W 1 Thickness, m Temperature, K Logarithmic mean temperature difference (LMTD), K Overall heat transfer coefficient, W m 2 K Pump power, kW

Comprehensive Energy Systems, Volume 4

1

doi:10.1016/B978-0-12-809597-3.00402-8

Heat Exchangers

Subscripts 1 2 b c eff

Inlet Outlet Bulk Cold Effective

Greek Symbols b Chevron angle,1  HEX effectiveness

4.2.1

f h i o w

Fouled Hot Inside Outside Wall

m r F

Dynamic viscosity, Pa s 1 Density, kg m 3 Surface enlargement factor

41

Introduction

Heat exchangers (HEX) are essential components in various industries, including aerospace, chemical, food, electronics, health, petroleum, power, and transportation, among others. They allow heat transfer between two or more fluids for the purpose of heating or cooling in processes, such as refrigeration, residential heating, electronics cooling, power production, wastewater treatment and heat recovery, and combustion. This chapter aims to provide a brief background on the classification, selection, and design of HEX, and focuses on gasketed-plate HEX (GPHEX) with a summary of recent developments in literature, including experimental and numerical work, and the specifics of GPHEX design methodology. The first objective of this work is to provide the classification and types of HEX (Section 4.2.1.1) and to provide general information on the HEX selection process, design methods, and methodology (Sections 4.2.1.2 and 4.2.2.1). After providing general information on the classification and design of HEX, the main focus of the chapter is on GPHEX; specifically, the work aims to shed light on the design, experiments, computations, and selection of plates for GPHEX. General information on GPHEX including advantages, utilization, plate geometries, and design is provided. A detailed literature survey on the state of the art is in Section 4.2.2.2. Section 4.2.3 explains the design procedure of GPHEX in detail. In Section 4.2.4, the details of a GPHEX selection software coded based on the correlations developed using experimental findings on several types of plates are explained. Section 4.2.5 is devoted to a case study of this software. Results are discussed in Section 4.2.6, whereas future directions and closing remarks are provided in Sections 4.2.7 and 4.2.8.

4.2.1.1

Classification of Heat Exchangers

HEX are classified according to several criteria: type of interaction (direct or indirect contact), recuperation/regeneration, flow configuration (parallel, counter, cross, or mixed), heat transfer mechanism (single or two-phase convection), and geometry (tubular, plate, or extended surfaces). These classification criteria are usually based on the application they are intended for. As a result, the terms boiler, condenser, steam generator, evaporator, heater, cooler, and radiator are often used in place of the word HEX. Many HEX are unique to a particular application, due to specialized requirements of the application; thus the designs are proprietary. Regardless, classification of HEX allows for the development of general design methodologies.

4.2.1.1.1

Type of interaction

In a direct contact HEX, the fluids are allowed to get in contact with each other as heat is transferred from the hot to the cold fluid at the interface between the two fluids. Mass transfer usually accompanies heat transfer in such a case. The cooling tower of a steam power plant is a good example for this type of HEX, where a stream of cool air flows over a spray of warm water. In an indirect contact (or transmural) HEX, the two fluids are separated by a solid wall, generally a tube or a plate, across which heat is transferred. The fluids are not in contact and mixing does not occur.

4.2.1.1.2

Recuperation/regeneration

Another classification is based on whether recuperation (direct transfer) or regeneration (storage) occurs in the HEX. Direct-transfer should not be confused with direct contact, explained above. Most conventional HEX, such as double-pipe, shell-and-tube, or gasketed-plate are recuperators, where the heat is transferred directly (either through contact or through a separating wall) from the hot fluid to the cold fluid and both flows are continuous. In contrast, in a regenerator, a matrix is utilized to store the thermal energy as it is first transferred from the hot fluid to the matrix, then from the matrix to the cold fluid, as a result of which the flow is generally periodic. Regenerators may be fixed matrix, such as air preheaters for furnaces, or rotary – either disk-type or drum-type. In a rotary regenerator, the matrix alternately flows in and out of the two gas streams as it stores and releases thermal energy; these are often utilized in steam power plants and gas turbines.

42

Heat Exchangers

Hot fluid

Heat transfer

Hot fluid

Cold fluid

Cold fluid

(A)

(B)

Fluid 1

(C)

Heat transfer

Fluid 1

Fluid 2

(D)

Fluid 2

Fig. 1 Possible flow configurations for two fluids in a heat exchanger (HEX). (A) Parallel-flow; (B) counter-flow; (C) cross-flow, Fluid 1 mixed, Fluid 2 unmixed; and (D) cross-flow, both fluids unmixed.

4.2.1.1.3

Flow configuration

Flow configuration (or arrangement) refers to the geometric relationship between the fluid streams, relating to the fluid flow paths. In parallel-flow (Fig. 1(A)) the two streams enter the HEX on the same side, flow parallel to each other, and leave on the other side of the HEX, while in counter-flow (Fig. 1(B)) the fluids flow in opposite directions. In terms of temperature efficiency, i.e., making good use of the available temperature difference between the two fluids, counter-flow HEX are superior [1,2], however, parallelflow HEX generally have more uniform wall temperatures. In cross-flow (Fig. 1(C) and (D)), the two fluids flow perpendicular to each other and the flows may each be mixed or unmixed. A mixed flow refers to the fluid being allowed to mix in itself in the transverse direction (see Fig. 1(C), Fluid 1), whereas unmixed flows refers to the fluid being guided through several individual channels (plates or tubes) without mixing between the adjacent channels (see Fig. 1(C), Fluid 2 and Fig. 1(D), both fluids). Depending on the combination, the HEX may be referred to as mixed–unmixed cross-flow, unmixed–unmixed flow cross-flow, etc. Often the flow configuration is not simple or ideal, as described above, but a combination of these flows occurs due to the use of multipass arrangements, U-tube junctions, and baffles.

4.2.1.1.4

Heat transfer mechanism

When the heat transfer mechanism is single-phase convection, the fluid stream(s) often encounter a significant amount of temperature change. In a counter-flow HEX, it is even possible for one of the fluids to leave the HEX at a temperature close to the inlet temperature of the other fluid, based on their respective heat capacity rates. However, in a parallel-flow HEX, the most that can occur is the two outlet temperatures being close to each other, since temperature cross-over is not thermodynamically possible. When there is two-phase convection on one or both sides of the HEX, often the purpose is to boil or condense the fluid(s) with phase change. Two-phase flow also refers to a case where a mixture of gas and solid particles are utilized for heat transfer, such as a fluidized bed. When there is phase change, the temperature change in the related stream is small, usually as a result of the small changes in pressure. Many times, subcooling and superheating sections occur before and after the saturation region in the HEX; in such a case, design is sometimes performed as if there are three HEX in series.

4.2.1.1.5

Geometry

The final classification is based on the geometry of the HEX design and the major features of the construction, thus the type of HEX. As new and specialized designs are added to the market each day, the geometry becomes more complex and unique. Nevertheless, three major categories may be described: tubular, plate, and extended-surface. In a tubular HEX, one fluid flows inside the tube, while the other flows on the outside, with the number of tubes, flow configuration, and tube diameter, length, and arrangement being key variables in the design. The most common tubular HEX are double-pipe (hairpin) and shell-and-tube HEX; while spiral-tube HEX are also utilized. A simple double-pipe HEX has one pipe placed concentrically inside a larger-diameter pipe, with one fluid flowing inside the inner tube and the other flowing through the annular space between the two tubes. U-shaped return bends enclosed in a housing are used to connect inner tubes and the whole structure resembles a hairpin, hence the name. Modular use of these HEX is possible; individual hairpins may be combined in series of parallel configurations to meet the pressure drop and temperature requirements in the process. Heat transfer surface areas are generally small, unless multiple inner tubes and/or fins on the surfaces of the inner tubes are utilized, thus double-pipe HEX are usually more suitable for low heat duty applications. They can withstand high pressures, especially on the tube side, due to small tube diameters, and cleaning and maintenance are relatively easy, compared to other types of HEX. Shell-and-tube HEX are perhaps the most widely used type of tubular HEX due to their versatility in construction and application. They are formed by placing a large tube bundle inside a shell, with one fluid flowing through the tubes, while the

Heat Exchangers

43

Fig. 2 A shell-and-tube heat exchangers (HEX) with single shell-and-tube side passes. Courtesy of Konuk Isı.

Fig. 3 The U-tube bundle, tube sheet, and baffles of a shell-and-tube heat exchanger (HEX). Courtesy of Konuk Isı.

other is on the shell-side, in parallel, counter, cross-flow or a combination of these, depending on the specifics of construction. An example is shown in Fig. 2. Baffles are often used to guide the shell-side flow across the tubes, enhance the heat transfer coefficient on the shell-side by promoting turbulence, and support the tube bundle against sagging and vibration. Fig. 3 is an example of the tube bundle inserted into the shell, with supporting baffles, and the tube sheet at one end of the bundle. It is possible to employ different tube and shell types or number of passes, tube layouts, baffle types and geometries, based on requirements of design; such as heat duty, pressure drop, fouling, and cleaning, and maintenance. In general, shell-and-tube HEX designs are more complex and costly compared to double-pipe HEX, but they have much larger heat transfer surface areas, thus they can accommodate larger heat duties. Spiral-tube HEX have coaxial-flow and contain spirally wound coils inside a shell, making benefit of the increased heat transfer coefficients in the curved tubes compared to straight tubes. These type of HEX are preferred in applications with clean fluids, since cleaning is nearly impossible. In plate HEX, instead of tubes, thin plates form the flow channels. Fluid streams flow between the plates, which may have corrugations or wavy surfaces to promote heat transfer depending on the type of fluid. GPHEX, which are the main focus of this chapter, are the most widely used type of plate HEX. They were initially used in the food industry due to their ease of cleaning, but have now come to be preferred as alternatives to shell-and-tube HEX in relatively lower pressure and heat duty liquid-to-liquid applications. They are composed of a pack of plates as the heat transfer surface with gaskets and the components of the frame, which includes a fixed plate, a compression (pressure) plate, an upper carrier bar, a guidance bar, a support column, and tightening bolts and nuts. A sample GPHEX assembly is provided in Fig. 4. The flow pattern for each fluid is through the passages formed by

44

Heat Exchangers

Fig. 4 A gasketed-plate HEX (GPHEX) assembly showing the plate pack, end plates, bolts, and nuts.

Fig. 5 The plates and gaskets of a gasketed-plate HEX (GPHEX) demonstrating the alternating flow patterns of each fluid. Courtesy of Alfa Laval Corporate AB.

alternate plate pairs, as shown in Fig. 5. Plate corrugation design and the specialized gaskets direct the fluids across the plates and prevent intermixing of the fluids and leakage to the outside. A major advantage is that the heat transfer area (plates) is easily accessible, which allows for changing the configuration to suit different process requirements by changing the number of plates and cleaning. In addition, heat transfer coefficients are very high due to small channel sizes and increased turbulence. GPHEX will be investigated in detail in the following sections. Other types of plate HEX are also available, such as spiral plate and lamella HEX.

Heat Exchangers

45

Fluid 2

Fluid 1

Fluid 2

Fluid 1 (A)

(B)

Fig. 6 Common core types of extended surface heat exchangers (HEX). (A) Plate-fin HEX and (B) tube-fin HEX.

Extended-surface HEX make use of fins on the main heat transfer surface area – tube or plate – to enhance heat transfer by both promoting turbulence and increasing the heat transfer surface area. They are also known as compact HEX since surface area-tovolume ratios are high (over 700 m2 m 3) compared to other types of HEX. As a result, they are widely preferred in applications in the transportation and aircraft industries, where limited space is available. Plate-fin HEX, generally used for gas-to-gas applications, and tube-fin HEX, used with the liquid on the tube side and the gas on the finned side, are the major categories, but finned tubes, such as those used in double-pipe HEX, are also common. In the plate-fin HEX, whose simple core structure is shown in Fig. 6(A), corrugated fins are sandwiched between flat plates, forming the fluid channels. The fins may be plain, perforated, serrated, or wavy, but specialized designs are also possible. Good flow distribution in the channels tends to be a problem, therefore manifold design requires special attention in the construction of the HEX. In addition, pressure drop can be high due to small hydraulic diameters. Tube-fin HEX are utilized when low heat transfer coefficients are encountered on one side due to gas flow and liquid flow may still be used inside tubes. The fins are fixed on the outside of an array of tubes, which may be round, flattened, or elliptical. Due to relative ease of construction, the fins are generally continuous plates, as sketched in Fig. 6(B), attached to the tubes by brazing, welding, or mechanical fit. Extended surfaces may also be utilized on the inside of tubes, especially for condensing and boiling applications, through the use of microfins (rough surfaces) and helical wire or twisted tape inserts.

4.2.1.2

Heat Exchanger Selection

Basic design methodology for HEX will be provided in the next section. However, in order to initiate the design, a suitable HEX type must be selected, considering the classification criteria summarized previously and other aspects that affect the selection process. Firstly, the HEX must satisfy all the process requirements related to flow rates, stream temperatures, operating pressures, pressure drop limitations, and size, and it must do so until a predetermined maintenance period. It must tolerate the environmental conditions of the plant where it will function, for instance by resisting corrosion. Another crucial issue is the resistance to fouling, which is the accumulation or growth of unwanted material on the heat transfer surface, leading to reduced performance, or even shutdown. Fouling affects initial and operating costs as well as heat transfer performance and pumping power requirements. Provisions for fouling must be made in selecting the suitable materials, fans/pumps, and factors that affect the size (heat transfer surface area), for example, tube length, diameter, and number. Fouling has been cited as the major unresolved problem in heat transfer [3]. Another issue in the selection process is maintenance. Fluid stream distribution and the HEX configuration must allow for cleaning and/or replacement of required parts, such as tubes or gaskets, that are affected by fouling, corrosion, erosion, or even vibration. Suitable space must be allocated around the HEX for maintenance or transportation, should a replacement unit be necessary. There may be further restrictions affecting selection, such as material/replacement parts availability, or any other plant requirement. Finally, the HEX must be cost effective in terms of capital, operational, and maintenance costs.

4.2.1.3

Recent Developments

In recent years, traditional materials, both for the working fluids and HEX structure, have started to be replaced by more advanced materials, such as the use of nanofluids and polymer HEX, in order to provide a more efficient performance or a more durable structure. Although metals have been the material of choice in the HEX industry since the earliest designs, in some applications, they give rise to limitations on the operation. This has resulted in the need to develop alternative designs with different materials, such as polymers. Polymers are generally able to resist fouling and corrosion and they may offer considerable reduction in weight. Although their thermal conductivities are low, resulting in high thermal resistance as the HEX construction material, using thin walls and/or polymer composites helps reduce this problem and utilize the benefits of easier machinability and lower maintenance costs. For example, in one of the earliest designs, a plastic two-phase HEX was utilized for desalination application [4]. In addition to alteration of the working fluid or the HEX material, the need for heat removal in smaller-scale applications has promoted the development of microscale HEX.

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Heat Exchangers

Fig. 7 A microchannel heat exchanger (HEX). Flow channels have a width of 350 mm. Reproduced from Çetin M. Design and experimental investigation of a microchannel heat exchanger [M.Sc. Thesis]. Ankara: Middle East Technical University; 2010.

Heat transfer efficiency can be improved by several means. The use of extended surfaces is one of the most widely utilized methods for this purpose, as was described in Section 4.2.1.1.5. Reducing the characteristic dimensions of the HEX via the use of microchannels is another solution. In one of the initial works in this area, it was shown that significantly higher heat fluxes are possible in smaller-scale channels [5]. The authors investigated heat sinks with liquid water flow to cool high-speed planar integrated circuits. This heat sink was able to dissipate 790 W cm 2 and the researchers claimed that 1000 W cm 2 heat dissipation is possible. This pioneering study made an impact on the field of heat transfer and motivated other researchers to investigate smallscale heat transfer, hence leading to the development of microscale HEX. Thermal conductivities of commonly used heat transfer fluids like water, ethylene glycol, or engine oil are relatively low; thus the fluid in the HEX generally possesses a higher heat transfer resistance compared to the solid components of the device. A direct intervention to the heat transfer fluids to improve the performance of the systems is an attractive idea; high-performance cooling can be obtained by increasing the thermal conductivity of fluids. This notion led to the development of nanofluids, briefly defined as the suspension of nanoparticles in a base fluid. The idea was initially investigated by Ahuja [6,7], who was able to achieve a heat transfer enhancement with mini-sized polystyrene particles in his system. However, clogging of the channels became a serious problem because of the deposition of the particles. The research required smaller (nano) sized particles called nanoparticles. They were dispersed in a base fluid, mixed, and homogenized with special techniques. One of the leading scientists who used this idea in a heat transfer system was Choi [8]. Since then, the research in this area has exponentially increased, especially to alleviate the drawbacks, such as sedimentation or aggregation of particles, clogging of channels, and erosion in channel walls. In the following sections, more detail on microscale HEX and nanofluids will be provided.

4.2.1.3.1

Microscale heat exchangers

Devices with dimensions between 1 mm and 1 mm are generally classified as microdevices, which include micropumps, microfuel cells, microprocessors, microbiosensors, and micro-HEX, to name a few. In many microdevices, especially those used in the electronics industry, overheating has become a serious issue. In fact it is one of the primary causes of device failure [9], due to heat fluxes exceeding 100 W cm 2. However, the new challenge for the coming decade is achieving 600–1000 W cm 2 values observed in, for example, supercomputers or military radar applications. In addition, available temperature differences are becoming smaller, and in some cases as low as only a few degrees. These high levels of heat dissipation require a dramatic reduction in the channel dimensions, matched with suitable coolant loop systems to facilitate the fluid movement away from the heat source [10]. Thus the development of compact and high-performance heat dissipation devices has become inevitable. Fig. 7 presents a sample micro-HEX with a channel size of 350 mm [11]. The theory governing flow and heat transfer in small-scale channels, under certain conditions, may differ from that in macroscale systems, especially for gases. At lower pressures, the flow becomes rarefied, and the degree of rarefaction is quantified by the Knudsen number, the ratio of the mean free path of the molecules to the characteristic length of the flow conduit, Kn¼l/L. The degree of rarefaction relates to whether the flow can be solved by the continuum model, i.e., traditional Navier–Stokes and energy equations with no-slip and no-temperature jump boundary conditions, or if further effort is necessary, like velocity-slip and temperature-jump boundary conditions, the use of direct simulation Monte Carlo (DSMC) methods, or the Boltzmann transport

Heat Exchangers Table 1

47

Knudsen number-based classification of flow regimes

Kn range

Flow regime

Solution method

Knr0.001 0.001rKnr0.1

Continuum Slip-flow

0.1rKnr10 Kn410

Transition Free-molecular

Navier–Stokes and energy equations with no-slip and no-temperature jump boundary conditions Navier–Stokes and energy equations with slip and temperature-jump boundary conditions or direct simulation Monte Carlo (DSMC) Boltzmann transport equation (BTE), DSMC BTE, DSMC

Source: Adapted from Gad-el-Hak M. The fluid mechanics of microdevices – the freshman scholar lecture. J Fluids Eng 1999;121:5–33.

equation (BTE). Flow in micro-HEX often falls into the slip-flow regime, as a result, traditional Navier–Stokes and energy equations are solved with velocity-slip and temperature-jump boundary conditions. Velocity-slip increases the predicted heat transfer while temperature-jump reduces it, due to the reduced temperature gradient at the wall. Table 1 summarizes flow regimes based on the value of Knudsen number. Many efforts, both experimental and theoretical, have been made since the early 1980s to predict the friction and heat transfer coefficients in microchannels, to aid in the design process of micro-HEX. Although these proposed correlations are out of the scope of the current chapter, exceptional differences from macroscale equations and flow/heat transfer phenomena will be pointed out. In micro-HEX, the heat transfer surface-to-volume ratio is high, which results in surface effects to become significant in gaseous flow, in turn resulting in mass flow rates and pressure drop higher than predicted by the continuum theory [2]. Higher pressure drop prevents high velocities, thus reducing the Reynolds number. In fact, transition from laminar to turbulent flow has been observed to be at much lower Reynolds numbers, as low as a few hundred [12], depending on the test condition and the geometry. This early transition improves heat transfer. In addition, in the laminar flow regime, Nusselt number is no longer constant, but is a function of Reynolds number, unlike continuum flow. Similarly, Poiseuille number, the product of friction factor and Reynolds number, also varies in both regimes, in contrast to being constant in laminar continuum flow. These variations from continuum behavior stress the importance of accurate prediction of heat transfer and friction coefficients during design. Since liquids have much higher thermal conductivities than gases, they are more frequently utilized in HEX applications, even though liquid-flow micro-HEX are more challenging to design especially for electronics applications. For liquid flow at microscale, scaling effects may need to be considered in design. Major scaling effects include entrance effects, axial heat conduction and viscous heating of the fluid, conjugate heat transfer, surface roughness effects, and temperature-dependent properties of the fluid. The measurement uncertainties also tend to be more critical at the microscale. In the design of most conventional HEX, entrance effects are generally disregarded and the flow is assumed to be both hydrodynamically and thermally developed. However, when length scales are small, the flow may not reach fully developed conditions through the HEX, which results in heat transfer coefficients being higher. In conventional HEX, since channel sizes are large, Reynolds numbers tend to be large as well. As a result, with the exception of very small Prandtl number fluids (e.g., liquid metals), the effect of fluid advection is dominant compared to the effect of conduction in the fluid. The ratio of these two effects may be termed as the nondimensional Peclet number, essentially the product of Reynolds and Prandtl numbers [13]. In other words, Peclet numbers are generally high in conventional HEX. In contrast, at the microscale, axial conduction may be effective, especially for thermally developing flow, and needs to be considered in design. A similar phenomenon affecting heat transfer at the microscale is the viscous heating of the fluid in the channel, which varies with the cube of hydraulic diameter [14]. The amount of viscous heating is quantified with the nondimensional Brinkman number. In one study, it has been argued that neglecting viscous heating effects at the microscale would greatly affect the predicted friction factor [15]. Conjugate heat transfer occurs when the axial conduction in the fluid and in the channel wall material are considerable. In macrochannels, the wetted perimeter, fluid velocity, hydraulic diameter, and Reynolds number are large, resulting in the effects of conjugate heat transfer being small. However, in microchannels, the opposite may be true. For example, Maranzana et al. analytically and numerically investigated the effect of conjugate heat transfer in a parallel plate counter-flow micro/mini HEX. The authors have developed two analytical models and have shown that neglecting conjugate heat transfer reduces the HEX efficiency [16]. They proposed the nondimensional Maranzana number, which compares the wall axial conduction to the wall-to-fluid convection, which may be used to predict the effect of conjugate heat transfer in the system. Surface roughness depends on the production method of the channel. Especially for similar manufacturing methods, relative roughness (the ratio of average height of roughness elements to the channel diameter) tends to be much higher for microchannels. Although it has a negative effect on the pressure drop, and thus pumping power requirement of the HEX, it improves the heat transfer, due to enhanced turbulence and surface area [17]. In the design of micro-HEX, a careful selection of suitable production methods could result in a considerable change in the pumping power or heat transfer rate. Temperature-dependent properties need to be considered in all HEX designs for accurate prediction of heat transfer coefficients. However, in micro-HEX, the change in inlet-to-exit temperatures of the fluids may be considerable, varying the density, thermal conductivity, and the viscosity of the fluids significantly. Especially at higher heat flux values, taking the thermophysical properties of the fluids constant could reduce the accuracy in design. In one work, it was shown that the change in Nusselt number for variable fluid properties, compared to constant fluid properties, was 10%, even at a relatively low heat flux of 30 W cm 2 [18].

48

Heat Exchangers

The general design methodology (see Section 4.2.2) of micro-HEX are similar to conventional HEX, with the exception of the abovementioned effects being included in the prediction of friction factors and heat transfer coefficients. In addition, the manufacturing methods for micro-HEX may be specialized, for example, micromechanical machining, X-ray micromachining, photolithography, or chemical etching; see Ref. [19] for a comprehensive review.

4.2.1.3.2

Nanofluids

With the progresses in nanotechnology, producing nanoparticles (sizes on the order of nanometers) are possible. As a result, suspending these nanoparticles in a base liquid for improving thermal conductivity has been proposed, hence the new class of working fluids called nanofluids. The enhanced thermal conductivity of the nanofluid in turn enhances heat transfer. Due to their small size, nanoparticles fluidize easily inside the base fluid, and as a consequence, clogging of channels and erosion in channel walls are no longer a problem. In addition, sedimentation problems (due to the use of larger particles) can be prevented with the addition of suitable dispersants, thus improving the stability of the fluid. It is also possible to use nanofluids in microchannels [20]. When preparing nanofluids, traditional working fluids commonly used in HEX applications, for example, water, ethylene glycol, and engine oil, are mixed with metallic or nonmetallic nanoparticles, with thermal conductivities much higher than those of base fluids, such as Al2O3, CuO, TiO2, SiC, TiC, Ag, Au, Cu, Fe, and even carbon nanotubes. In order to improve the stability of the nanoparticles inside the base fluid, some additives are added to the mixture in small amounts. The particles have sizes less than 100 nm usually, with the exception of carbon nanotubes, whose lengths can exceed this value, but their diameters are generally much smaller [20]. Nanoparticles may form clusters or aggregates with sizes of the order of micrometers, which may adversely affect the flow due to clogging and sedimentation, but may improve heat transfer due to enhanced conduction within the cluster. During the preparation of nanofluids, two possible methods are utilized. In the one-step method, the production of the nanoparticles and the dispersion of nanoparticles in the base fluid are combined into a single step. However, mass production is not possible with this method [21]. In the two-step method, as the name suggests, first the nanoparticles are produced, then they are mixed with the base fluid, making the method more suitable for mass production. In literature, two-step methods are more common in part due to their flexibility. As mentioned above, nanofluids are produced and used with the intention of increasing the effective thermal conductivity of the working fluid. However, predicting this thermal conductivity value accurately is still a challenge for many researchers. Because of the definition of Nusselt number, researchers expect to see an equal amount of enhancement in the convection heat transfer coefficient as the thermal conductivity. Surprisingly, this is not the case, as observed in many experiments – the enhancement in the heat transfer coefficient surpasses the thermal conductivity enhancement of nanofluids [22–25]. Several theoretical and empirical models were proposed to predict the effective thermal conductivity of the nanofluid and also to clarify this additional enhancement. Most models offer the effective thermal conductivity as a function of the thermal conductivities of the base fluid and nanoparticles, particle volume fraction (the volumetric concentration of the nanoparticles in the nanofluid), and often particle size. Other models also include factors, such as temperature and particle shape, and may even be proposed for a particular base fluidnanoparticle combination. In addition to the functional relations, it has also been observed that the surfactant used for stabilization, the acidity of the base fluid, the amount of clustering/aggregation of the nanoparticles, and ultrasonication time for nanofluid preparation (with the two-step method) alter the effective thermal conductivity. An excellent review of the enhanced thermal conductivity of nanofluids reviews and compares experimental findings and theoretical predictions in literature [26]. The chapter summarizes that there exist significant discrepancies among the available experimental data and between the experimental findings and the theoretical model predictions. The discrepancies are as surprising as a reverse trend in the functional relations. For instance, the increase in the nanoparticle size has been observed to decrease [27] or increase [28] the effective thermal conductivity. One of the major reasons for this discrepancy has been offered to be the thermal conductivity measurement technique [20]. In one of the commonly used methods, the transient hot-wire method, free convection effects at higher experimental temperatures result in the thermal conductivity data being higher [29]. Measured thermal conductivities of nanofluids in literature are as high as 160% of that of the base fluid [30]. Researchers suggest that several mechanisms could be responsible for this anomalous enhancement. These are Brownian motion of particles (random motion of particles suspended in the fluid, which transports energy directly by nanoparticles); clustering of nanoparticles (results in fast transport of heat along relatively large distances over the solid aggregate); and formation of thin liquid nanolayers with higher thermal conductivity than the base fluid over nanoparticles. For feasible application of nanofluids in HEX systems, all thermophysical properties should be accurately predicted and used in the estimation of convection heat transfer coefficient, and problems, such as clogging and sedimentation should be resolved. Aside from thermal conductivity, during the design process, density, specific heat, and viscosity also need to be calculated. Density and specific heat can be accurately calculated using mixing theory and the thermal equilibrium method [31]. However, the viscosity of nanofluids is not straightforwardly predicted and the literature contains similar debates as those regarding thermal conductivity. Viscosity is critical, since it directly affects the required pumping power in the HEX system. Similar to thermal conductivity, most models for the effective viscosity of the nanofluid include particle volume fraction and size, in addition to the viscosities of the base fluid and nanoparticles, with fewer models also depending on temperature and cluster size. The increase in viscosity may be as high as 250% [32], for 5% particle volume fraction. While classical models, such as the Einstein model, underpredict the effective nanofluid viscosity, other newer methods include a wider selection of system parameters and better match experimental data [33].

Heat Exchangers

49

In some works, the concept of nanofluids is investigated from an overall performance and system-application point of view. This is especially important for the use of nanofluids in HEX. Even though the overall heat transfer coefficient of the system increases by changing its working fluid to a nanofluid, the nanofluid also requires more pumping power to be circulated in the system because of the increased viscosity of the fluid [31]. In such a systematic approach, the rate of heat transfer or heat transfer coefficient and pumping power are compared in the form of some performance parameter. For example, Choi and Eastman [34] compared the pumping power ratio and the heat transfer coefficient ratio, both of the nanofluid over the base fluid, based on a reference state. However, due to lack of information, the authors considered all thermophysical properties of the nanofluid, except thermal conductivity, to be constant. In another work, an evaluation criterion for nanofluid heat transfer performance was proposed, which compared the rate of heat removal-to-pumping power ratio of the nanofluid to that of the base fluid [35]. Kirez [31] evaluated the nanofluid heat transfer performance by taking pumping power as equal for both the base fluid and the nanofluid for laminar flow in a pipe and observed the heat transfer rate ratio or wall temperature ratio of the nanofluid and base fluid for constant wall temperature or constant wall heat flux boundary condition, respectively. Other works were based on efficacy (achieving a predefined heat removal aim, such as maximum temperature or heat transfer rate, with a lower pumping power) and entropy generation (comparing the entropy generation rate of the nanofluid flow to the entropy generation rate of base fluid at the same mass flow rate). More recently, Liu et al. specifically investigated the impact of nanofluid heat transfer enhancement on the performance of HEX theoretically in laminar and turbulent flow regimes [36]. At first, water and ethylene glycol-based nanofluid thermophysical properties and resulting heat transfer coefficient values for the base fluids and nanofluids were presented. However, the Einstein viscosity model, which underestimates the viscosity, and conventional Nusselt number correlations (developed for pure fluids) were used to calculate the Nusselt number. The heat transfer enhancement was evaluated for the same nanofluid pumping power, and an improvement in the HEX performance was observed. Enhancement in number of transfer units (NTU, which represents the dimensionless heat transfer size of the HEX; see Section 4.2.2.1.2) and resulting enhancement on the heat transfer rate were explained through equations similar to the ones utilized in the effectiveness-NTU method (Section 2.1.2). It is important to note that the thermal resistance weight of the nanofluid side of the HEX (fraction of thermal resistance of nanofluid in the total thermal resistance of the HEX) is as important as the heat transfer coefficient enhancement [31]. The maximum heat transfer rate enhancement was presented as 7% when the heat transfer coefficient enhancement was 50% with the nanofluid.

4.2.2

Background/Fundamentals

In this section, the basic design methodology common to all types of HEX will be summarized first. Then, the background on plate HEX will be provided. Different types of HEX, such as double-pipe, shell-and-tube, and GPHEX have procedures – often computerized – additional to what is provided in this section, to complete the design process. The main focus of this chapter is the description and application of such an approach for GPHEX.

4.2.2.1

Design Methods for Heat Exchangers

Design of a HEX refers to either sizing or rating of the device [2]. Sizing refers to selecting a suitable HEX and determining the appropriate dimensions (length, number of tubes or plates, diameter, etc.) to satisfy the process requirements related to flow rates, inlet and outlet fluid temperatures, and pressure drops. On the other hand, in the rating of a HEX, which is also known as performance analysis, the HEX is either already available, or a predesign has been performed and an initial selection has been made. In this case, the task of the designer is to calculate the outlet temperatures, heat transfer rate, and pressure drop, and check whether the HEX satisfies the plant requirements. Often, modifications need to be made to satisfy other criteria mentioned in Section 4.2.1.2, and computerized procedures are utilized in the design modification process. In general, logarithmic mean temperature difference (LMTD) method is preferred in sizing and effectiveness-NTU (e-NTU) method is preferred in rating; however, both methods, which will be summarized in this section, are applicable to design. Table 2 summarizes these two HEX design tasks and the parameters that are available or to be determined in each task. Table 2

A summary of the two basic design tasks for a heat exchangers (HEX)

Preferred design method Known Find

Sizing

Rating

Logarithmic mean temperature difference (LMTD)

Effectiveness-number of transfer units (e-NTU)

Flow rates, inlet and outlet temperatures, available (allowable) pressure drop Dimensions – type and size of HEX, actual pressure drop

Flow rates, inlet temperatures, allowable pressure drop, dimensions (size) Heat transfer rate, fluid outlet temperatures, and actual pressure drop

50

Heat Exchangers

4.2.2.1.1

Log mean temperature difference method

In the basic design methodology, a two-fluid, indirect contact recuperator is considered. Although a complete design process involves cost and structural considerations, the basic equations are related to thermal analysis only. Taking the HEX as the control volume, assuming steady-state conditions with no work or stray heat transfer and negligible kinetic and potential energy effects, with constant thermodynamics properties, and no phase change, the rate of heat transfer may be written for either fluid as follows.
 _ p h ðTh1 Th2 Þ Q ¼ mc ð1Þ
 _ p c ðTc2 Tc1 Þ ð2Þ Q ¼ mc

_ is the mass flow rate, cp is the specific heat of the fluid, T is temperature and the subscripts h, c, 1, and 2 refer to hot fluid, here, m cold fluid, inlet, and outlet, respectively. The temperature difference between the fluids changes with position along the HEX, therefore an appropriate mean temperature difference must be established so that the heat transfer rate may also be evaluated through the following equation. Q ¼ UADTlm

ð3Þ

The temperature difference in Eq. (3) is known as the LMTD, and is the basis of the current design method. It is defined as follows. DTlm ¼

DT1 DT2 lnðDT1 =DT2 Þ

ð4Þ

Here, DT1 and DT2 are the difference between the two fluids at the two ends (inlet and outlet) of the HEX, which will be different for parallel-flow and counter-flow. If the HEX consists of another flow configuration, cross- or mixed-flow, LMTD must be corrected with a factor F, which is available through charts developed from experimental data for, for example, shell-and-tube HEX, or others. When the temperature differences at the two ends are equal, that is DT1 = DT2, DTlm will be indeterminate from Eq. (4). In that case, applying L’Hospital’s rule results in DTlm ¼ DT1 ¼ DT2. A very important revelation from this approach is that, for the same inlet and outlet temperatures, the LMTD for counter-flow is greater than that for parallel-flow. As a result, under ideal conditions (assuming the same UA), for a given heat transfer rate, the required heat transfer area for counter-flow is less than that for parallel-flow. In fact, counter-flow LMTD represents the maximum temperature potential for a HEX that could be obtained in counter-flow configuration only. Most HEX designs are based on counter-flow, unless the plant requirements dictate otherwise. In Eq. (3), A is the heat transfer surface area, specific to the HEX type, and U is known as the overall heat transfer coefficient. Overall heat transfer coefficient accounts for all the thermal resistances to heat transfer between the two fluids, including forced convection of both fluids, wall conduction, fouling of both fluids, if any, and fins, if any. The UA product, which is the inverse of the total thermal resistance, may be written in terms of these individual resistances as follows.   1 1 Rfi Rfo 1 þ Rw þ þ ð5Þ þ UA ¼ Uo Ao ¼ Ui Ai ¼ Ao ho Ao hi Ai Ai In this equation, the subscripts o and i refer to outside and inside, respectively, related to the main heat transfer surface. h is the convection heat transfer coefficient, Rw is the wall conduction resistance, and Rf is the fouling factor (resistance), which is available in tabulated form in many references, such as the Standards of the Tubular Exchanger Manufacturers Association (TEMA) [37]. Wall resistance depends on the HEX geometry, but for the two most common geometries, bare plane wall and bare tube wall, respectively, it is defined as below. t kw A

ð6Þ

lnðro =ri Þ 2pLkw

ð7Þ

Rw;plane ¼ Rw;tube ¼

Here, t, r, and L are the wall thickness, radius, and length of the tube, respectively, and kw is the thermal conductivity of the wall material. When fins are present on one or both sides of the HEX, fin efficiency must also be accounted for (see Ref. [2]). For an unfinned tubular HEX, the overall heat transfer coefficient based on the outside surface area of the tubes is expressed as below.   ro ro ro lnðro =ri Þ 1 1 þ Rfo þ ð8Þ þ Rfi þ Uo ¼ ri hi ri kw ho

4.2.2.1.2

Effectiveness-number of transfer units method

The LMTD method requires inlet and outlet temperatures of the two fluids to be known, which results in knowing the heat transfer rate, since flow rates are generally available. When outlet temperatures are unknown, the LMTD method requires a trial-and-error type of procedure, and may be tedious. An alternative is the effectiveness-NTU (e-NTU) method, which states that the inlet or exit temperature differences are a function of certain dimensionless quantities. The first of these is the heat capacity rate ratio. C ¼

Cmin Cmax

ð9Þ

Heat Exchangers

51

_ p , and min and max refer to the smaller and larger of the heat capacity rates of the hot Here, C is the heat capacity rate equal to mc and cold fluids. The range of heat capacity rate ratio is 0rC*r1, with 0 corresponding to one fluid evaporating or condensing, and 1 simply the rates being equal. The second dimensionless quantity is HEX effectiveness, defined as below. e¼

Q Qmax

ð10Þ

Effectiveness is the ratio of the actual heat transfer rate to the thermodynamically limited maximum heat transfer rate, which may be obtained if an infinitely large heat transfer area were available in a counter-flow HEX. In such a case, the fluid with the minimum heat capacity rate will theoretically go through the maximum temperature change and leave the HEX at the inlet temperature of the other fluid. Thus Qmax may also be expressed as below. Qmax ¼ Cmin ðTh1

Tc1 Þ

ð11Þ

Combining with Eqs. (1) and (2), e¼

Ch ðTh1 Th2 Þ Cc ðTc2 Tc1 Þ ¼ Cmin ðTh1 Tc1 Þ Cmin ðTh1 Tc1 Þ

ð12Þ

The value of e ranges between 0 and 1 also. The third and last dimensionless quantity is the NTU, which represents the dimensionless heat transfer size of the HEX. NTU ¼

UA Cmin

ð13Þ

Equations or graphs are available in many Refs. [2,38], for e as a function of NTU and C* for various flow arrangements and HEX types. These equations are useful when flow rates, size, and inlet temperatures are known and heat transfer rate and outlet temperatures are to be determined; thus for the rating of a HEX. Equations of NTU as a function of e and C* are also available, which may be used when temperatures and heat transfer rate are known and sizing is to be performed. The e-NTU method is more versatile and more commonly used by designers, but the LMTD method is much easier in sizing.

4.2.2.2

Background on Gasketed-Plate Heat Exchangers

The crucial part of plate HEX research is to be able to determine the characteristics of the plates correctly. The design of the plates is very important as well. Both the hydraulic and thermal performance need to be maximized. The corrugation pattern and other geometrical parameters of the plates need to be designed for maximum performance and this process requires tedious experimental and computational effort. It is possible to use a GPHEX selection software to design a thermal system that meets the technical requirements, once the properties and performance of the plates that are going to be used in the HEX of the thermal system are known. In literature, there are many correlations based on computational and experimental studies to predict the heat transfer coefficient for the hot and cold side flows across the plates, to be used in Eq. (5). These correlations are developed for specific plates and there would be errors in the plate selection process if the available correlations are used, unless the geometries match exactly [39,40]. Therefore, plate selection software needs to include different experimentally and/or computationally validated correlations for each plate that is in their database. These databases also need to be updated based on new plate designs. The correlations in literature that are utilized for plate selection do not necessarily work for every plate, since the plates are custom designed; however, they give an insight on the form, functional dependence, and efficiency of correlation development. Some of the studies in literature are on the design of new plates based on experiments, computations, or both; whereas some of the studies are on the determination of the thermal and hydraulic performances of plate HEX and correlation development process. Here, both types of studies are examined in detail to provide information on the state of the art of GPHEX design and development. Zahid [41] presents a detailed literature survey of heat transfer and flow characteristics of GPHEX with single and multiphase flows. Types and historical background of GPHEX are examined. There are over 30 different practical correlations for Fanning friction factor and convective heat transfer coefficient that are explained in their detailed review. Almost all of these correlations are in the form of a power law and some recent studies use an exponent of chevron angle, b, which is related to the shape of corrugations on the plates (see Fig. 8). Two-phase flow correlations are functions of various thermodynamic properties, such as heat flux, mass flux, local flow regimes, and film thickness, which increase the complexity of the correlation when compared to single-phase flow correlations. In Fig. 8, the geometry of a chevron-type plate and gaskets are shown. Mean channel spacing (channel gap) (b) is calculated through the following equation, using plate pitch (p) and thickness (t). b¼p

t

ð14Þ

Since the corrugated structure of the plates determines the heat transfer rate, the ratio of the actual effective (corrugated) area A1 to projected plate area A1,p is defined as the surface enlargement factor. F¼

A1 A1;p

ð15Þ

52

Heat Exchangers



Lp

Lv

Lh

Dp Fig. 8 Chevron-type plate geometry.

Projected plate area is calculated as: A1;p ¼ Lp  Lw

ð16Þ

Dp

ð17Þ

Lw ELh þ Dp

ð18Þ

Lp ELv

here, Lp, Lw, Lv, and Lh are the length of the plate between the ports, plate width, vertical port distance, and horizontal port distance, respectively, as shown in Fig. 8. Dp is the port diameter. Hydraulic diameter of the GPHEX is calculated from Eq. (19). Dh ¼

4bLw 2b D 2ðb þ Lw FÞ F

ð19Þ

To calculate the friction coefficient and Nusselt number, mostly hydraulic diameter or equivalent diameter is used. The latter is defined as below. De ¼ 2b

ð20Þ

Garcia-Cascales et al. [42] studied refrigeration cycles in which HEX are used as both evaporators and condensers. Fernandes et al. [43] analyzed fully developed laminar flows in double-sine chevron-type GPHEX, numerically. Several corrugation angle values are used in the analyses. chevron angle varies between 30 and 60 degrees in their computations. It is found that when the chevron angle decreases, surface enlargement factor increases, but it is weakly influenced by the channel aspect ratio. Three-dimensional (3D) flow and energy equations are integrated through the gap between the plates and they are transformed into two-dimensional (2D) equations in the study of Lyytikainen et al. [44]. The computational time is reduced considerably using this methodology. Computational fluid dynamics (CFD) analyses are performed using depth-averaged governing equations for five different corrugation angles. Pressure drop and temperature profiles are found compatible with 3D profiles. It is also found that pressure drop is not as sensitive to geometry as temperature. It was mentioned in Section 4.2.1 that GPHEX were initially developed for the food industry, mainly due to their ease of cleaning. One such numerical work was conducted by Afonso et al. [45], who obtained a correlation for the determination of heat transfer coefficients for stirred yogurt taking into account rheological features. Stirred yogurt is a non-Newtonian fluid with high viscosity and it has complex flow patterns due to temperature, shear rate, and elastic properties. HEX efficiency is adversely affected by these properties. The flow in corrugated GPHEX is simplified as laminar flow to minimize the computational complexity in their study. Correlations of convective heat transfer coefficient obtained for the GPHEX are combined with the corrugation pattern to obtain reliable results. Numerical results agree well with the experiments mainly because the high

Heat Exchangers

53

viscosity of yogurt makes corrugation pattern irrelevant; therefore the corrugated surfaces do not seem to cause significant perturbations in the flow. Prabhakara et al. [46] conducted experiments to determine the pressure differences from port to channel in plate HEX for a wide range of Reynolds numbers. Different number of channels is used in the experiments for corrugation angles of 20 to 80 degrees. Water is used as the working fluid for both hot and cold sides of the plate. Pressure probes are inserted into both the inlet and exit ports of the channel through the gasket. Total pressure drop is also measured for different flow rates. Total pressure drop is a function of the flow rate, the cross-sectional area, channel-to-port ratio, and number of channels per fluid. It is observed in the experiments that, when total pressure change increases, flow maldistribution increases in the HEX. Fanning friction factor correlations are also obtained in the study. GPHEX are also utilized in nonconventional applications. For instance, absorption phenomenon is investigated in the study of Cerezo et al. [47], where ammonia and water are used as working fluids for absorption chillers driven by low-temperature heat sources, where GPHEX is utilized for the absorption part and as the main component of the device. The objective of the study is to increase the efficiency of the chiller by managing the sources. It is shown that there is a reduction in energy consumption using solar energy and waste heat. Second law analysis of GPHEX has also been done in literature. Durmus et al. [48] investigated the performance of three GPHEX with different plates based on heat transfer, friction factor, and exergy losses. A standard plate HEX, a corrugated plate HEX, and an asterisk plate HEX are used in the experiments. The heat transfer characteristics of the corrugated plate HEX are better than the others according to the experimental findings. Galeazzo et al. [49] utilized CFD tools to simulate a four-channel GPHEX. Parallel and series configurations are also tested using an experimental setup. The discrepancy between the experimental and numerical findings is 8% for the series and 25% for the parallel configuration, respectively. Although the results are not in perfect agreement with the experimental findings, it is shown that CFD can be a powerful tool for the prediction of GPHEX performance. The computational time is the limiting parameter for the design when utilizing CFD techniques. Islamoglu and Parmaksizoglu [50] performed experiments to determine forced convection heat transfer coefficient and friction factor for GPHEX. They showed that increase in channel height decreases the hydraulic performance, while improving the thermal characteristics. The effects of flow maldistribution and fouling in GPHEX are also investigated by several researchers. For example, Kho and Müller-Steinhagen [51] worked on fouling of different GPHEX configurations. Calcium sulfate is utilized to provide fouling in the experiments. It is determined that, to prevent fouling, flow distributers need to be located at plate edges. Miura et al. [52] investigated the effects of flow arrangement on pressure drop. Pressure drop is found to be a function of air bubbles in the HEX, pressure, gasket, and plate arrangement in their experiments. They also developed a friction factor formulation in terms of Reynolds number and experimentally determined the constants of the equation. Muley and Manglik [81] also performed experiments to develop correlations for heat transfer and pressure drop in a U-type chevron plate HEX. Their resulting friction factor and Nusselt number correlations are given in Table 3. Rao et al. [54] studied flow maldistribution in single- and multipass GPHEX. Experiments indicate that Z-type plates cause more maldistribution than the Utype HEX [2]. Thonon and Mercier [55] developed a 3D computational analysis software based on experimental data. The effects of chevron angle, friction factor, and convective heat transfer coefficient for two different plates are investigated. A wide range of Reynolds numbers is used. Tsai et al. [67] examined the hydrodynamic characteristics of GPHEX with two cross-corrugated channel plates. According to their results, experimental findings for pressure drop are 20% higher than CFD predictions. In the study of Warnakulasuriya and Worek [56], heat transfer and pressure drop characteristics of an absorbent salt solution in a commercial GPHEX serving as a solution subcooler in the high loop of triple-effect absorption refrigeration cycle is investigated. In many of the studies summarized above, correlations to predict the heat transfer coefficient, in terms of the dimensionless Nusselt number, and friction factor, are proposed. In general, Nusselt number is a function of Reynolds and Prandtl numbers, while friction factor is a function of Reynolds number only. A comprehensive summary of available correlations is presented in Table 3 to provide a general understanding of the mathematical form of the correlations in literature. Shah and London [68] and Kays and Persins [69] propose the following mathematical form, which is frequently observed in Table 3, for complex geometries. This equation is helpful when the variation of the fluid viscosity with temperature is significant.  d m Nu ¼ aReb Pr c ð21Þ mw Here, m is the fluid viscosity, with w referring the value at wall temperature. The ranges for coefficients a, b, c, and d, determined from plate-specific experiments, are provided in Marriott’s work [57]. Alternatively, when viscosity is not a critical issue, Bounape et al. [70] and Cooper [71] propose the following form for Nusselt number for fluids with high Reynolds numbers. Nu ¼ aReb Pr c

ð22Þ

Fanning friction factor is generally proposed in the following form [72,73]: f ¼ a=Reb

ð23Þ

54

Heat Exchangers

Table 3

Gasketed-plate HEX (GPHEX) Nusselt number and friction factor correlations available in literature

Researcher

Year

Correlation

Notes

Afonso [45]

2008

Prabhakara et al. [46]

2006

For adiabatic wall condition: Nu¼ 1.69Re0.362Pr0.3 For constant heat flux condition: Nu¼ 1.694Re0.56Pr0.5 f¼1.059Re 0.145

Cerezo et al [47]

2009

Yogurt b¼30 degree F ¼1.105 t¼0.5 mm Water b¼60 degree 900oReo10,000 3.5oPro5.7 Lv ¼0.357 m t¼0.6 mm Ammonia–water mixture

Muley and Manglik [81] Rao et al [54]

1999 2005

Nu¼ 0.99Re0.53Pr0.33 66oReo400 Nu¼ 0.399Re0.703Pr0.33
400oReo900 0:09 Nu ¼ 0:295Re 0:64 Pr 0:32 p2 b Nu¼ 0.218Re0.65Pr1/3 Re4500

f¼21.41Re

0.301

100oReo7000

Nu¼C1RemPr1/3 Re 50oReo15,000 50oReo15,000 50oReo15,000 50oReo15,000 50oReo15,000 50oReo15,000 50oReo15,000 50oReo15,000  0:14 Reo400 Nu ¼ 0:292Re 0:705 Pr 0:35 mm

Thonon and Mercier [55]

1996

Warnakulsuriya and Worek [56]

2008

Marriot [57]

1971

Nu ¼ 0:37Re 0:668 Pr 0:333

Jackson and Troupe [58] Maslov and Kovalenko [59]

1964

Nu ¼ 0:742Re 0:38 Pr 0:333

1972

Okada et al. [60]

1972

1/3 Nu¼ 0.78Re0.5Pr   f ¼ 915Re 0:25 DLve

Rosenblad and Kullendroff [61]

1975

Nu¼ 0.1528Re0.66Pr0.4 b ¼60 degree Nu¼ 0.241Re0.64Pr0.4 b ¼45 degree Nu¼ 0.317Re0.65Pr0.4 b ¼30 degree Nu¼ 0.463Re0.62Pr0.4 b ¼15 degree Nu¼ 0.289Re0.697Pr1/3

Amooie-Foomey [62]

1977

f¼

Tovazhnyanski et al. [63]

1980

Nu ¼ 0:051eð0:64tancÞ Re 0:73 Pr 0:43

Leuliet et al. [64]

1991

f¼

Chisholm and Wanniarachchi [65] Roetzel et al. [66]

1992

 0:55 Nu ¼ 0:72Re 0:59 Pr 0:4 F0:41 ð9030 bÞ  3:5 f ¼ 0:8Re 0:25 F1:25 ð9030 bÞ

b 30 30 45 45 60 60 90 90

w

f ¼23.8Re

h

26ðFÞ Re

0.205

 0:14 m mw

 0:14

1000 Re

m mw

i þ 0:16 ðFÞ

f¼0.085exp[1.52tanc]Re

1994

Water–water applications

þ 2:60Re



Pr Prw

0:25

(0.25 0.06tanc)

0:10

Nu¼ 0.371Re0.703Pr1/3

C1 0.2946 0.2946 0.2998 0.2998 0.2267 0.2267 0.1 0.1

m 0.7 0.7 0.645 0.645 0.631 0.631 0.687 0.687

Water b¼60 degree 600oReo6000 3.5oPro5.7 b¼2.9 mm t ¼0.8 mm f¼C2Re p Re Reo1000 Re41000 Reo550 Re4550 Reo200 Re4200 Reo160 Re4160

p 0.9 0.39 0.83 0.22 0.68 0.17 0.67 0.17

ALFA-LAVAL model POI-VG counter-flow heat exchanger, 82oPro174 Water 4oDeo10 mm 10omo100 kg m 1 s 1 Water and corn syrup 0.03oReo500 b¼30 degree De/Lv ¼0.0192 50oReo20,000 700oReo20,000 Water A1 ¼0.044 m2 F¼1.294 Water 60oReo2415 b¼30 degree Water 0.8oReo2200 F¼1.25 b¼30, 45, 90 degrees c¼p 2b F ¼1.16 2000oReo25,000 Herringbone-type plates b¼30 degree Water 1000oReo4000 30 Degreeobo80 degree Water Chevron-type plates (Continued )

Heat Exchangers Table 3

55

Continued

Researcher

Year

Correlation

Notes b¼20 degree Lh ¼176.5 mm, Lw ¼71 mm t ¼0.5 mm, b¼2 mm 400oReo2000

Table 4 z¼(Tb

Equations of thermophysical properties and Prandtl number of water as a function of temperature 75.01)/43.734 2.2669  10 6z10 þ 7.1155  10 6z9 þ 1.0058  10 6z8 4.1379  10 5z7 þ 1.6368  10 4z6 6.7876  10 7z4 þ 3.0995  10 3z3 1.6647  10 2z2 þ 3.0108  10 2z þ 0.66676

Thermal conductivity (W m 1 K 1)

k¼

Dynamic viscosity (Pa s)

m ¼2.6903  10 7z10 þ 1.7807  10 5z4

Prandtl number

Pr ¼0.0027282z10 0.0062558z9 0.0048218z8 0.35652z3 þ 0.78974z2 1.4512z þ 2.3751

Density (kg m 3)

r ¼ 0.0007372z10 þ 0.00090699z9 þ 0.0022672z8 þ 0.0020248z7 þ 0.51852z3 4.9714z2 26.108z þ 974.81

Specific heat (J kg K 1)

cp ¼0.030245z10 0.043553z9 0.10675z8 þ 0.014169z7 þ 0.65741z6 þ 15.773z2 þ 29.045z þ 4193.3

6.4358  10 7z9 3.6488  10 7z8 þ 4.6227  10 7z7 þ 4.4261  10 6z6 4.4496  10 5z3 þ 1.0563  10 4z2 2.1620  10 4z þ 0.0003766 0.0063953z7 þ 0.042122z6

0.090944z5

3.9601  10 4z5 9.8916  10 6z5

0.15071z4

0.017614z6 þ 0.037529z5

0.15635z4

1.0414z5 þ 0.97217z4 þ 2.8177z3

here, a and b are experimentally determined based on the plate geometry [73]. As it is seen from the mentioned literature, there is no unique correlation, for neither Nusselt number nor friction factor that covers the entire range of Reynolds and Prandtl numbers and an extended range of plate geometries, which shows the necessity of the development of custom design correlations for every plate.

4.2.3

Specifics of Gasketed-Plate Heat Exchanger Design

Life frame of a HEX is the time between its design and disposal. The design process of a HEX using design and analysis software covers the identification of the problem by gathering information from the user, performing the necessary calculations (thermal, hydraulic, and mechanical), selecting the best alternative among several workable HEX, and the phases of production, assembly, and installation. The designer needs to gather enough information on several factors, such as the properties of the HEX, its intended purpose, operating, and environmental conditions. Permissible pressure drop, cost, volume, and weight are some of the specifications that affect the design. It is crucial to make sure that the requirements of the user are met since the specifications further affect the decisions of the designer in subsequent phases of the design [74]. The basic equations in design have been introduced in Section 4.2.1.2, as Eqs. (1–4). The area in Eq. (3) should be replaced by the effective plate area, Aeff, which is the total heat transfer area of a GPHEX. In addition, Eqs. (5) and (6) may be combined to obtain the overall heat transfer coefficient for a GPHEX.   1 t 1 1 U¼ þ Rfh þ þ Rfc þ ð24Þ hh kw hc As seen in Table 3, the heat transfer coefficient is provided in terms of the dimensionless Nusselt number in the correlations, and the two are related through the following relation. Nu ¼

hDh k

ð25Þ

Thermophysical properties, such as thermal conductivity, viscosity, density, and specific heat, and the fluid Prandtl number are functions of temperature. Through curve fitting of data from National Institute of Standards and Technology (NIST) database [75], these properties may be obtained, as provided in Table 4. In the design, it is often required to determine the number of plates to be used in the GPHEX. To calculate the number of plates for heat transfer, Ne, the total effective surface area for heat transfer appearing in Eq. (3) is divided by the actual effective area of a single plate. Ne ¼

Aeff A1

ð26Þ

Heat Exchangers

56

The total number of plates, Nt, is determined by considering the first and last plates of the GPHEX, which do not take part in the heat transfer process. Therefore Nt ¼ Ne þ 2

ð27Þ

Reynolds number is based on the mass flow rate of the fluid and may be obtained by first calculating the number of channels per pass, as in Eq. (28), then the mass velocity, as in Eq. (29). Here, Np is the number of passes of the HEX. Ncp ¼

Nt 1 2Np

ð28Þ

Gc ¼

_ m Ncp bLw

ð29Þ

Gc De m

ð30Þ

Finally, Reynolds number is obtained. Re ¼

Since the fouling coefficients used in Eq. (24) are obtained by cumbersome experimental studies, instead of using fouling factors, it is preferred to specify a value for over surface (OS), and to take the reduction in the performance resulting from the fouling in the HEX into account [76].   Af 1 ð31Þ OS ¼ 100 Ac Here, Af and Ac are the fouled and clean effective areas of the GPHEX. It is recommended not to use more than 30% OS. The pressure loss of the fluid that is in contact with the plate surface, connecting pipes, passages, and the other parts of the GPHEX is unavoidable. A pump is used in the system along with the HEX to compensate the pressure loss and to make the fluid flow at a constant rate. Selection of the pump affects both the total cost and the mechanical performance; therefore, it has a critical role in the design process. Pump power is directly related to the pressure drop _ _ p ¼ mDP W r  Zp

ð32Þ

here, Zp is the pump isentropic efficiency. According to Wang [76] and Shah [74], the total pressure drop in plate HEX is calculated through Eq. (33). DPt ¼ DPf þ DPg þ DPa þ DPp

ð33Þ

In Eq. (33), DPf, DPg, DPa, and DPp are pressure drop due to friction, gravity, fluid acceleration, and port inlet/outlet. Frictional losses are calculated through the following relation [76].   0:17 Lv Np G2c m ð34Þ DPf ¼ 4f Dh 2r mw Here, f is the friction factor determined through correlations, such as those provided in Table 3.

4.2.4

Experimental Work and Software Development

In this section, the details of the development of a GPHEX selection software are presented. The software is used to design several GPHEX alternatives using available plate geometries in the database of the code, utilizing the above design procedure. The database is formed using the Nusselt number and friction factor correlations obtained from experiments performed with various inlet and outlet temperature and flow rates, for several Reynolds numbers for different plates.

4.2.4.1

Experiments

The details of the experiments are available in the studies of Gulenoglu et al. and Akturk et al. [39,40]. The GPHEX shown in Fig. 9 is assembled using four different plates. Main design parameters for these plates are summarized in Table 5. For the four different plate types used in this research, Nusselt number and friction factor correlations are developed and presented in detail in Akturk et al. [40] and Gulenoglu et al. [39]. Correlations for the specific plates are shown in Table 6.

4.2.4.2

Computer Software Development

Microsoft Visual Basic 2010 Express edition [77] is preferred as the programming language. The software is composed of two main parts. The first part is the tab called design parameters where the user inputs the data. Thermal and hydraulic analysis and HEX design is performed and the most suitable plate type or types are summarized and presented to the user in the form of a table in the results tab.

Heat Exchangers

57

Fig. 9 Gasketed-plate HEX (GPHEX) used in this study. Table 5

Parameters for the plates used in the study

Design parameter

Unit

Plate 1

Plate 2

Plate 3

Plate 4

b Dp Lw Lv Lp b t A1 A1p F De Dh kw (AISI 316) Nt

degree m m m m mm mm m2 m2 – m m Wm 1K –

30 0.069 0.230 0.6058 0.5368 2.85 0.45 0.142 0.109 1.30 0.0057 0.0044 16.2 10, 15, 21

30 0.035 0.109 0.3700 0.3350 2.76 0.45 0.035 0.030 1.17 0.0055 0.0047 16.2 20, 31

30 0.035 0.109 0.665 0.63 2.76 0.45 0.073 0.062 1.17 0.0055 0.0047 16.2 21, 31

24 0.100 0.343 0.732 0.632 2.64 0.45 0.266 0.207 1.288 0.0053 0.0041 16.2 10

1

The user specifies the heat load for the GPHEX, fluid types for the primary and secondary circuits, inlet and outlet temperatures of these fluids, and the volumetric or mass flow rate. The fouling factors are also input. If there is a plate type that is preferred by the user in the predesign phase, the calculations are limited to this plate. In Fig. 10, the interface of the design parameters tab of the computer program is shown.

58

Heat Exchangers Table 6 Plate Plate 1 Plate 2 Plate 3 Plate 4

Nusselt number and friction factor correlations for the four plates used in this work Nusselt number

Friction factor  0:14

Nu ¼ 0:32643Re 0:6125 Pr 1=3 mmb  w0:14 Nu ¼ 0:32867Re 0:68 Pr 1=3 mmb  w 0:14 Nu ¼ 0:32774Re 0:675 Pr 1=3 mmb  w0:14 Nu ¼ 0:17422Re 0:7 Pr 1=3 mmb

f¼66055Re

1.72

f¼259.9Re

0.9227

f¼1371Re f¼

1.146

þ 0.4299 þ 1.246

þ 1.139

0.003743Re0.5981 þ 0.9132

w

Fig. 10 Design parameters tab of the software interface.

Three different materials, AISI 304, AISI 316, and titanium, which are among the most preferred material types in industrial applications, are presented to the user. Gasket material should also be determined considering the working temperature and pressure. The alternative designs with all the plate types are demonstrated to the user in the results tab as shown in Fig. 11. For each suitable plate type, data, such as number of plates, effective heat transfer area, overall heat transfer coefficient of the HEX, pressure drops that are expected to occur in the primary and secondary circuits, the Reynolds numbers, and flow rates are presented. The flow chart of the computer program is provided in Fig. 12.

4.2.5

Case Study

A model problem is solved using the developed design software [78]. In the problem, it is desired to use a GPHEX to heat city water. Hot water, for instance provided from a geothermal source or a solar power plant, enters the HEX at 901C and exits at 701C, meanwhile, cold fluid temperature increases from 15 to 451C. The total heat load of the HEX is 50 kW. The allowable total pressure drop for both fluids is 5 kPa. AISI 316 is chosen as the plate material, and the OS is designated as 15%. Mass flow rates and fluid properties are shown in Table 7. Computer software results for the alternative designs using four different plates are presented in Table 8. For the same operating conditions, when Plate 1 or Plate 4 are used, the actual effective area for one plate is higher when compared to Plate 2 and Plate 3. Therefore, using a total of 10 plates is enough with Plate 1 and Plate 4. However, when Plate 2 is in use, the number of plates used for the same heat load and pressure drop is 30, and for Plate 3, it is 38. When the total pressure

Heat Exchangers

59

Fig. 11 Results tab of the software interface.

drop values of different plates are examined, it can be seen that it is below the allowable 5 kPa user input. By evaluating the area for heat exchange and cost, the user makes the most suitable GPHEX selection according to the parameters that are prioritized. In the case study, although Plate 1 and Plate 4 seem advantageous for application, factors, such as limitations in the operating area of the HEX, costs, etc. may lead to the selection of Plate 2 or Plate 3 for a specific application.

4.2.6

Validation and Verification

In order for the software to be useful in GPHEX design, it should be validated through experimental data and verified through known correlations from literature. However, it was previously mentioned that, unless the exact plate geometries are utilized, it is not possible to match the software results with available correlations. This will be scrutinized below.

4.2.6.1

Validation of the Software Using Experimental Data

The developed computer software provides several GPHEX alternatives to the user based on the designs it performs using different plate types. Therefore, the number of plates used in each design is different. Fig. 13 shows the number of plates obtained for several designs for different heat loads if Plate 1 is used in the design. Scattered data on the figure are the experimental findings, whereas the lines represent the solutions of the software. As seen in the figure, the number of plates delivered through the software is greater than the number of plates used in the experiments for the same heat load, which provides a safer design. Fig. 14 shows the number of plates obtained for several designs for different heat loads using Plate 2. Scattered data is experimental results and the lines represent the software results. As the number of plates increases from 20 to 31, the predictions of the software start deviating from the experiments. This is probably because of the unreliability of the experimentally developed correlations due to flow maldistribution. Flow maldistribution is crucial in plate HEX. When more plates are used, the heat transfer for each plate surface is not the same since the flow cannot be uniformly distributed on each plate. When 10 plates are used, it is easier to distribute the flow uniformly; however, when 31 plates are utilized, as shown in the figure, because of nonuniformity of flow distribution (maldistribution of the flow on the plates), the experimental correlations that assume uniform flow become unreliable. The same trend is observed for Plate 3, as shown in Fig. 15. The results for friction factor obtained from experiments, the correlation developed using the experimental data, and the results of the computer software are presented in Fig. 16, for Plate 2. The agreement between the correlation and the software results is expected, since the computer software directly uses the results of the experimentally obtained correlations. The Nusselt number correlation that is obtained by processing the data gathered in the experiments with Plate 2 and values of Nusselt number obtained using the same operating conditions in the software are provided in Fig. 17; results for other plates are

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Heat Exchangers

Inputs: Inlet and outlet temperatures, flow rates, types, fouling factors, max. allowed pressure losses of hot and cold fluids, heat load, number of pass, working temperature and pressure, preffered type of plates

Heat balance calculation Taking an initial value for overall heat transfer coefficient using LMTD method, calculation of effective heat transfer area

Set calculated value to the initial value For every plate type

No

Calculation of number of plates, reynolds numbers, nu numbers, heat transfer coefficients and pressure drop values of hot and cold fluid, overall heat transfer coefficient, pressure losses

Increase the number of plates Yes

Change plate type

The calculated overall heat transfer coefficient is equal to the initial value?

Yes The calculated pressure drops are greater than the max. allowed pressure drop?

No

Outputs: name of the plate, number of plates, effective heat transfer area, pressure drops, channel reynolds numbers and channel flow rates of hot and cold fluids

Fig. 12 Flow chart of the gasketed-plate HEX (GPHEX) selection software. LMTD, logarithmic mean temperature difference.

similar and have not been included here for simplicity. Only Plate 2 data and results are presented. The program outputs agree with the experimental findings, as seen in number of plates, Nusselt number, and friction factor graphs given in Figs. 13–17, respectively.

4.2.6.2

Verification of the Computer Software With Correlations in Literature

Fig. 18 shows the Nusselt number values obtained using several correlations in literature suggested for chevron plates with a range of chevron angles and the computer software comparatively for Plate 2 geometry. Fig. 19 shows the trend for overall heat transfer coefficient calculated based on these Nusselt number values. As it is seen in these figures, although the same type of plate (chevron) is used, it is inevitable to have disagreement due to differences in plate geometries among these studies. It is always necessary to obtain correlations experimentally or computationally for the specific plates and use them in the design to avoid over design or the possibility of not satisfying the expectations for thermal and hydraulic performance. The characteristics of the distribution of total pressure drop values, which are calculated with the help of friction coefficients attained from different correlations for similar operating conditions by using the Plate 2 geometry, are demonstrated in Fig. 20 along with the correlation developed herein. Fig. 21 shows friction factors used in the calculation of the pressure drop values. Similar comments as those for overall heat transfer coefficient can be made in this case also. It is necessary to obtain

Heat Exchangers Table 7

Mass flow rates and fluid properties for the case study

Fluid properties

Unit

m_ c m_ h Prc Prh mc mh kc kh rc rh

Table 8

61

kg s 1 kg s 1 – – Pa s Pa s Wm 1K Wm 1K kg m 3 kg m 3

0.3932 0.5921 5.4356 2.2193 0.0007974 0.0003543 0.61545 0.66998 995.607 971.767

1

Design results for different plate types for the case study

Design parameter

Unit

Plate 1

Plate 2

Plate 3

Plate 4

Nt Ncp,h Ncp,c Gc,c Gc,h Rec Reh Nuc Nuh fc fh hc hh U DPt,c DPt,h

– – – kg m kg m – – – – – – Wm Wm Wm kPa kPa

10 5 4 150.0 180.7 1072 2906 43.9 53.6 0.84 0.50 6186 8215 3216 4.84 4.98

30 15 14 93.4 131.2 646 2,044 50.2 72.8 1.91 1.48 6,553 10,336 3611 2.54 4.64

38 19 18 72.6 103.6 503 1614 40.9 59.5 2.24 1.43 5339 8458 3002 3.2 4.9

10 5 4 108.6 130.8 719 1949 32.6 43.4 0.72 0.57 4899 7099 2684 2.82 3.78

2 2

2 2 2

s s

1

K K K

1

1

1 1

25 Number of plates

Nt = 21 20 Nt =15 15 Nt =10

10

5 0

20

40

60

80

100

120

140

160

180

200

220

240

Heat load (Q) (kW) Fig. 13 Number of plates using Plate 1 in gasketed-plate HEX (GPHEX) design changing with heat load.

plate-specific correlations either experimentally or computationally to make sure thermal and hydraulic performance requirements are satisfied. Figs. 22 and 23 show the friction factor and Nusselt number variations with changing Reynolds number for Plate 3. The figures show the friction factor and Nusselt numbers computed for this plate using the software, based on different correlations in literature and the plate-specific correlations developed for this plate. As seen in the distribution of the friction factor, the available

62

Heat Exchangers

40 Nt=20

Nt=31

35 Number of plates

Nt=31 30 25 Nt=20

20 15 10 0

20

40

60

80 100 120 Heat load (Q) (kW)

140

160

180

Fig. 14 Number of plates using Plate 2 in gasketed-plate HEX (GPHEX) design changing with heat load.

Nt=21

Nt=31

35 33 Nt=31

Number of plates

31 29 27 25 23

Nt=21

21 19 17 15 0

20

40

60 80 Head load (Q) (kW)

100

120

140

Fig. 15 Number of plates using Plate 3 in gasketed-plate HEX (GPHEX) design changing with heat load.

2.0 Experiment

Program output

Correlation

1.8

Friction factor (f )

1.6 1.4 1.2 1.0 0.8 0.6 0.4 0.2 0.0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

5000

Reynolds number (Re) Fig. 16 Friction factor distributions by experiments and developed correlation for Plate 2 for several Reynolds numbers.

Heat Exchangers

160

Experiment

63

Program output

Nusselt number (Nu)

140 120 100 80 60 40 20 0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

5000

Reynolds number (Re) Fig. 17 Nusselt number results using Plate 2 in the design.

Talik and Swanson (79)

Muley (53)

Muley and Manglik (81)

Tovazhnyanski et al. (63)

Focke and Olivier (84)

Maslov and Kovalenko (59)

Chisholm and Wanniarachchi (65)

Marriot (57)

Martin (83)

Thonon and Mercier (55)

Kumar (80)

Okada et al. (60)

Wang and Sundén (82)

Wang and Sundén (82)

Plate 2 correlation

200 180 160

Nusselt number (Nu)

140 120 100 80 60 40 20 0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

Reynolds number (Re) Fig. 18 Nusselt number predictions using different correlations for Plate 2 geometry changing with Reynolds number [53,79,80,82–84].

correlations in literature underpredict the friction factor for this plate. However, when the Nusselt number variations are examined, some of the correlations in literature underpredict, some of them overpredict the Nusselt numbers. There is no firm evidence to clearly state that the values are underpredicted and overpredicted, which also clearly shows that plate-specific correlations are inevitable to make sure that the predictions of any plate HEX selection software are reliable. Similar conclusions are obtained when the overall heat transfer coefficient and pressure drop values obtained using these friction factor and Nusselt numbers for Plate 3 are examined. Figs. 24 and 25 show the overall heat transfer coefficient and pressure drop variations for Plate 3. Pressure drop calculated based on friction factor is underpredicted with the available software in literature, which shows the necessity of custom correlations for each plate.

64

Heat Exchangers

Talik and Swanson (79)

Kumar (80)

Thonon and Mercier (55)

Focke and Olivier (84)

Maslov and Kovalenko (59)

Wang and Sundén (82)

Muley (53)

Muley and Manglik (81)

Tovazhnyanski et al. (63)

Chisholm and Wanniarachchi (65)

Plate 2 correlation

Marriot (57)

Overall heat transfer coefficient, U (W m−2 K−1)

8000 7000 6000 5000 4000 3000 2000 1000 0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

Reynolds number (Re) Fig. 19 Overall heat transfer coefficient obtained with different correlations for the same operation conditions [53,79,80,82–84].

Talik and Swanson (79)

Kumar (80)

Thonon and Mercier (55)

Martin (83)

Marriot (57)

Focke and Olivier (84)

Maslov and Kovalenko (59)

Wang and Sundén (82)

Muley (53)

Muley and Manglik (81)

Chisholm and Wanniararchi (65)

Plate 2 correlation

45 40

Total pressure drop (kPa)

35 30 25 20 15 10 5 0 0

500

1000

1500

2000 2500 3000 Reynolds number (Re)

3500

4000

4500

5000

Fig. 20 Change of total pressure drop values using different correlations for Plate 2 geometry with Reynolds number [53,79,80,82–84]

Heat Exchangers

Talik and Swanson (79)

Kumar (80)

Thonon and Mercier (55)

Martin (83)

Marriot (57)

Focke and Olivier (84)

Maslov and Kovalenko (59)

Wang and Sundén (82)

Muley (53)

Muley and Manglik (81)

Chisholm and Wanniarachchi (65)

Plate 2 correlation

65

3.2 2.8

Friction factor (f )

2.4 2.0 1.6 1.2 0.8 0.4 0.0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

Reynolds number (Re) Fig. 21 Friction factor predictions using different correlations for Plate 2 geometry changing with Reynolds number [53,79,80,82–84].

Talik and Swanson (79)

Kumar (80)

Thonon and Mercier (55)

Martin (83)

Marriot (57)

Focke and Olivier (84)

Maslov and Kovalenko (59)

Wang and Sundén (82)

Muley (53)

Muley and Manglik (81)

Chisholm and Wanniarachchi (65)

Plate 2 correlation

2.8 2.6 2.4 2.2

Friction factor (f)

2.0 1.8 1.6 1.4 1.2 1.0 0.8 0.6 0.4 0.2 0.0 0

500

1000

1500

2000

2500 3000 3500 Reynolds number (Re)

4000

4500

5000

5500

Fig. 22 Friction factor predictions using different correlations for Plate 3 geometry changing with Reynolds number [53,79,80,82–84].

4.2.7

Future Directions

This study provides information on classification, types, design methodology, and selection of HEX, focusing on GPHEX. Details of software developed based on correlations that are obtained using experimental data are presented. GPHEX literature is

66

Heat Exchangers

Tovazhnyanski et al. (63)

Talik and Swanson (79)

Muley (53)

Muley and Manglik (81)

Focke and Olivier (84)

Maslov and Kovalenko (59)

Chisholm and Wanniarachchi (65)

Marriot (57)

Martin (83)

Thonon and Mercier (55)

Kumar (80)

Okada et al. (60)

Wang and Sundén (82)

Wang and Sundén (82)

Plate 2 correlation

200 180 160

Nusselt number (Nu)

140 120 100 80 60 40 20 0 0

500

1000

1500

2000

2500 3000 3500 Reynolds number (Re)

4000

4500

5000

5500

Fig. 23 Nusselt number predictions using different correlations for Plate 3 geometry changing with Reynolds number [53,79,80,82–84].

Talik and Swanson (79)

Kumar (80)

Thonon and Mercier (55)

Focke and Olivier (84)

Maslov and Kovalenko (59)

Wang and Sundén (82)

Muley (53)

Muley and Manglik (81)

Tovazhnyanski et al. (63)

Chisholm and Wanniarachchi (65)

Plate 3 correlation

2500

3500

Marriot (57)

8000

Overall heat transfer coefficient, U (W m−2 K−1)

7000

6000

5000

4000

3000

2000

1000

0 0

500

1000

1500

2000

3000

4000

4500

Reynolds number (Re) Fig. 24 Overall heat transfer coefficient obtained with different correlations for the same operation conditions for Plate 3 geometry [53,79,80,82–84].

Heat Exchangers

Talik and Swanson (79)

Kumar (80)

Thonon and Mercier (55)

Martin (83)

Marriot (57)

Focke and Olivier (84)

Maslov and Kovalenko (59)

Wang and Sundén (82)

Muley (53)

Muley and Manglik (81)

Chisholm and Wanniararchi (65)

Plate 3 correlation

67

50 45

Total pressure drop (kPa)

40 35 30 25 20 15 10 5 0 0

500

1000

1500

2000

2500 3000 3500 Reynolds number (Re)

4000

4500

5000

5500

Fig. 25 Change of total pressure drop values using different correlations for Plate 3 geometry with Reynolds number [53,79,80,82–84].

explained in detail and the correlations in literature are used in the developed HEX selection software for comparison purposes with the new correlations. The analysis shows that specific correlations based on experimental data are necessary to be used with plates utilized in HEX selection, since the plate geometries are unique.

4.2.8

Closing Remarks

Although HEX are not a new subject, the developments continue, especially with the help of computer utilization in the design. Experiments are necessary for HEX design. However, the use of CFD for fluid mechanics and heat transfer simulations started to play an important role in their design process. Artificial neural networks and other intelligent systems started to be used instead of classical correlations to minimize the errors associated with generalized correlations. More recently, nanofluids have been offered as alternative high-performance heat transfer fluids, which are possible to be used in HEX applications. In addition, the advances in new materials will affect new designs in many aspects, from heat load to cost. For instance, polymer HEX have recently become popular mainly due to their enhanced resistance to fouling and corrosion, and advantages related to weight and cost. HEX are indispensable in all energy-related processes and they will probably continue to be one of the hot topics of energy in the future, as well.

Acknowledgments The software presented in this study is developed with the help of funding from Turkish Ministry of Science, Industry and Technology SANTEZ program under grant number STZ0347.2009-1. The authors would also like to thank Tanpera Inc. for their financial support and insight for the development of the software.

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69

[60] Okada K, Ono M, Tomimura T, Okuma T, Konno H, Ohtani S. Design and heat transfer characteristics of a new plate heat exchanger. Heat Tran Japanese Res 1972;1:90–5. [61] Rosenblad G, Kullendroff A. Estimating heat transfer from mass transfer studies on plate heat exchanger surfaces. Warme-und Stoffubertragung. 1975;8:187–91. [62] Amooie-Foomey M. Flow distribution in plate heat exchangers [Ph.D thesis]. Bradford: University of Bradford; 1977. [63] Tovazhnyanski LL, Kapustenko PA, Tsibulnik VA. Heat transfer and hydraulic resistance in channels of plate heat exchangers. Energetika 1980;6:123–5. [64] Leuliet JC, Mangonnat JF, Laiande M. Flow and heat transfer in plate heat exchangers treating viscous Newtonian and pseudoplastic products, modeling the variations of the hydraulic diameter. Canadian J.Chem Eng 1990;220–9. [65] Chisholm D, Wanniarachchi AS. Maldistribution in single-pass mixed-channel plate heat exchangers. In: Proceedings of the ASME HTD compact heat exchangers for power and process industries; 1992. pp. 95–9. [66] Roetzel W, Das SK, Luo X. Measurement of the heat transfer coefficient in plate heat exchangers using a temperature oscillation technique. Int J Heat Mass Tran 1994;37:325–31. [67] Tsai YC, Liu FB, Shen PT. Investigations of the pressure drop and flow distribution in a chevron-type plate heat exchanger. Int Commun Heat Mass Tran 2009;36:574–8. [68] Shah RK, London AL. Laminar flow forced convection in ducts. 1st ed. New York, NY: Academic Press; 1978. [69] Kays WM, Perkins HC, Rohsenow WM, Hartnett J JP, Ganic EN, editors. Handbook of heat transfer fundamentals New York, NY: McGraw Hill; 1985. [70] Buonopane RA, Troupe RA, Morgan JC. Heat transfer design methods for plate heat exchangers. Chem Eng 1963;42:57–61. [71] Cooper A. Recover more heat with plate heat exchangers. Chem Eng 1974;285:280–5. [72] Raju KSN, Bansal JC. Design of plate heat exchangers. In: Kakaç S, Shah RK, Bergles AE, editors. Low reynolds number flow heat exchangers. Washington, DC: Hemisphere; 1983. p. 913–32. [73] Raju KSN, Bansal JC. Plate heat exchangers and their performance. In: Kakaç S, Shah RK, Bergles AE, editors. Low reynolds number flow heat exchangers. Washington, DC: Hemisphere; 1983. p. 899–912. [74] Shah RK, Dusan PS. Fundamentals of heat exchanger design. New York, NY: John Wiley and Sons; 2003. [75] National Institute of Standards and Technology. Thermophysical properties of fluid systems. Available from: http://webbook.nist.gov/chemistry/fluid/; 2016 [accessed 29.09.16]. [76] Wang L, Sundén B, Manglik RM. Plate heat exchangers design, applications and performance. 1st ed. Boston, MA: WIT Press; 2007. [77] Visual Basic Developer Center. Visual basic 2010 express edition. Available from: http://msdn.microsoft.com/en-us/vbasic/default.aspx; 2010 [accessed 29.09.16]. [78] Gülben G. Development of a computer program for designing gasketed plate heat exchangers for various working conditions and verification of the computer program with experimental data [M.Sc. thesis]. Ankara: TOBB University of Economics and Technology; 2011. [79] Talik AC, Swanson LW. Heat transfer and pressure drop characteristics of a plate heat exchanger using a propylene-glycol/water mixture as the working fluid. In: Proceeding of the 30th national heat transfer conference, vol. 12, p. 83; 1995. [80] Kumar H. The plate heat exchanger: construction and design. In: 1st UK national conference on heat transfer, the institution of chemical engineers symposium series, vol. 86, 1275–86; 1984. [81] Muley A, Manglik RM. Experimental study of turbulent flow heat transfer and pressure drop in a plate heat exchanger with chevron plates. J Heat Tran 1999;121:110–7. [82] Wang L, Sundén B. Optimal design of plate heat exchangers with and without pressure drop specifications. Appl Therm Eng 2003;23:295–311. [83] Martin H. A theoretical approach to predict the performance of Chevron-type plate heat exchangers. Chem Eng Proc 1996;35:301–10. [84] Focke JZ, Olivier I. The effect of the corrugation inclination angle on the thermohydraulic performance of plate heat exchangers. J Heat Mass Tran 1985;28:1669–79.

Further Reading Edwards MF. 1983. Heat transfer in plate heat exchangers at low Reynolds numbers. In: Kakaç S, Shah RK, Bergles AE, editors. Low reynolds number flow heat exchangers. Washington, DC: Hemisphere; 1983. p. 933–47.

Relevant Websites http://webbook.nist.gov/chemistry/fluid/ National Institute of Standards and Technology – Thermophysical Properties of Fluid Systems. http://msdn.microsoft.com/en-us/vbasic/default.aspx Visual Basic Developer Center – Visual Basic 2010 Express Edition.

4.3 Heat Pipes Hussam Jouhara, Brunel University London, Uxbridge, United Kingdom r 2018 Elsevier Inc. All rights reserved.

4.3.1 Introduction 4.3.2 Fundamentals 4.3.2.1 Heat Pipe Casings 4.3.2.2 Heat Pipe Wick Structures 4.3.2.3 Heat Pipe Working Fluids 4.3.2.4 Heat Pipe Advantages 4.3.3 Heat Pipe Types 4.3.3.1 Classification Using the Operating Temperature 4.3.3.2 Classification According to Their Geometry/Operation 4.3.3.2.1 Variable conductance heat pipes – Gas loaded heat pipes 4.3.3.2.2 Thermal diodes and switches heat pipes 4.3.3.2.3 Oscillating heat pipes 4.3.3.2.4 Loop heat pipes 4.3.3.2.5 Micro-heat pipes 4.3.3.2.6 Rotating heat pipes 4.3.3.2.7 Controlled heat pipes – Variable conductance heat pipes 4.3.3.2.8 Gas loaded heat pipes 4.3.3.2.9 Excess liquid heat pipes 4.3.3.2.10 Vapor flow modulated heat pipes 4.3.3.2.11 Liquid flow modulated heat pipes 4.3.3.2.12 Flat heat pipes 4.3.3.2.13 Thermosyphons 4.3.4 Systems and Applications 4.3.4.1 Heat Pipe Heat Exchangers 4.3.4.2 Heating, Ventilation and Air Conditioning Applications 4.3.4.3 Air Handling Units 4.3.4.4 Energy Conservation and Renewable Energy 4.3.4.4.1 Solar energy collection 4.3.4.4.2 Thermal energy storage 4.3.5 Analysis and Assessment 4.3.5.1 Heat Pipe Modeling 4.3.5.2 Heat Pipe Based Heat Exchanger Modeling 4.3.6 Case Studies 4.3.6.1 Steam Generator, China National Offshore Oil Corp, China Sea, April 2016 4.3.6.2 Gas to Air Unit, Automotive, Aluminum Furnace, USA 2008 4.3.6.3 Exhaust to Coke Gas Unit, Steel Mill Blast Furnace, Czech Republic 2011 4.3.6.4 G2W, Shale Gas Well Head Fracking, Thermal Oxidizer, Canada 2012 4.3.6.5 Multi-Stage, in Series, Steam/Water, Natural Gas, Spirax Sarco, Italy 2012 4.3.6.6 Steam Condenser, Food, Dirty Steam, Ireland, 2010 4.3.6.7 G2W, in Line Through-Flow Recuperator, Biomass Incinerator, Bologna, Italy 4.3.7 Closing Remarks 4.3.8 Future Directions References Relevant Website

Nomenclature Symbol A Cp Csf

70

Surface area (m²) Specific heat of vapor (J/kg K) Constant, determined from experimental data

D Dd FP FPF

71 73 73 74 74 75 75 75 76 76 76 76 76 78 78 78 79 79 79 79 79 80 80 80 80 81 82 82 83 84 84 85 90 90 90 91 92 92 93 94 95 95 95 97

Diameter (m) Bubble detachment diameter (m) Pressure correction factor Pressure correction factor in Gorenflo correlation

Comprehensive Energy Systems, Volume 4

doi:10.1016/B978-0-12-809597-3.00403-X

Heat Pipes

g h hfg Ja k l m M N Nu P

Gravitational acceleration (m/s²) Heat transfer coefficient (W/m2 K) Latent heat of vaporization (J/kg) Jacob number Thermal conductivity of liquid (W/m K) Length (m) Intermediate variable Molecular weight (g/mol) Number of rows Nusselt number Saturation pressure (N/m²)

Pr pr q Q Re Rp T DpSAT

Prandtl number Reduced pressure Heat flux (W/m²) Heat transfer rate (W) Reynolds number Mean surface roughness (mm) Temperature (1C) The difference between the vapor pressures at the heating surface and liquid temperatures (N/m²)

Greek Symbols a Liquid thermal diffusivity (m²/s) y Bubble contact angle (degree) m Dynamic viscosity (kg/m s)

r s C

Density (kg/m3) Surface tension (N/m) Angle of condenser orientation with the vertical (degree)

Subscripts av c ci crit ei

f l r v x

Film Liquid reduced Vapor Excess

n

experimental constant that depends on fluid

Refers to average value Condenser section Corresponds to inner wall of condenser Critical Corresponds to inner wall of evaporator

Supercripts

4.3.1

71

Introduction

Heat pipes are recognized as one of the most efficient passive heat transfer technologies available. Their integration into heat exchangers and heat sinks offer key advantages over conventional heat distribution systems in terms of energy savings, manufacturing cost, thermal conductivity, temperature distribution, and reliable operation. A heat pipe is a structure with very high thermal conductivity that enables the transportation of heat, while maintaining almost uniform temperature along its heated and cooled sections. Heat pipes are considered to be “thermal superconductors” due to the high rates they transfer heat across small temperature difference across the heat pipe. In their simplest form heat pipes are called thermosyphons, their operation relies on gravity, and heat can only be transferred from the lower to the upper end of the pipe. Heat pipes that allow only bottom to top transfer of heat are also termed wickless. A condensed fluid at the bottom of the pipe evaporates, collecting heat, and moves as a vapor to the top of the pipe where it condenses, releasing heat, and then trickles back as a liquid down to the bottom of the pipe where the cycle repeats. Practically, heat pipes were first introduced by Jacob Perkins through a series of patents in 1836. The Perkins tubes, as they were known, were mostly wickless gravity-assisted heat pipes (thermosyphons), in which heat transfer was achieved by a fluid’s change of phase (latent heat of evaporation); quite similar to modern wickless heat pipe designs. The Perkins tube design, which is closest to the present heat pipe technologies, consisted of a closed loop tube containing a small quantity of water operating in a two-phase cycle [1] (Fig. 1). Richard Gaugler in 1944 enabled the full introduction of the heat pipe concept in modern scientific applications. At the time, Gaugler was working in refrigeration for General Motors Corporation, where he designed a device which could passively evaporate a liquid above the level where condensation would occur. To do this he incorporated an internal wick in the heat pipe, resulting in the first wicked heat pipe design. Gaugler’s refrigeration unit used the heat pipe to transfer heat from the interior of the fridge compartment to a pan of crushed ice below [3] (Fig. 2). However, Gaugler’s patent was not used by General Motors and it was only some years later that his invention received the necessary attention of the scientific community and the recognition of heat pipes as reliable thermal devices was established. The majority of current heat pipes are cylindrical in cross-section, but they can be manufactured in a wide variety of shapes, such as right-angle bends, S-turns, spirals, and even in flat configurations. Heat pipes and heat pipe-based heat exchangers find application in many industries from computer cooling to waste heat recovery and energy savings systems, while their operation has been the subject of significant research investigation.

72

Heat Pipes

Expansion tube

Heat

‘The Engineer’

Fig. 1 Perkins heat pipe boiler [2].

Fig. 2 Gaugler’s heat transfer device [3].

Interceptor

Swain Sc.

Heat Pipes

4.3.2

73

Fundamentals

A heat pipe is a passive thermal transfer device able to transport large amounts of heat over relatively long distances, with no moving parts, using phase-change processes and vapor diffusion. The main structure of heat pipes consists of an evacuated tube partially filled with a working fluid that exists in both liquid and vapor phases. Fig. 3 represents the basic steps of operation of heat pipes. The left part of the heat pipe is the evaporator and the right part is the condenser. When a high temperature is applied at the evaporator section of the heat pipe, the working fluid existing in the liquid phase evaporates and flows as a vapor with high velocity toward the cooler end of the pipe – the condenser. As soon as the vapor reaches the condenser section it condenses and gives up its latent heat. Then the condensate returns to the evaporator part of the pipe, by the influence of gravity (thermosyphons) or by some sort of capillary wicking structure (wicked heat pipes) (Fig. 4). The operational principle remains the same for all heat pipe applications, only the materials and the geometry change. The key materials that determine suitability for each application are the casing material, the wicking structure, and the working fluid.

4.3.2.1

Heat Pipe Casings

The casing provides structural stability and enclosure, maintaining the liquid and vapor under a seal at all working temperatures and transmitting heat between the exterior and the interior of the pipe. The selection of the casing material is based on its thermal, physical, kinetic, and chemical properties, as well as on economic criteria. Appropriate casing materials to consider would be those which are compatible with the materials used as the working fluid and the wicking structure; have mechanical properties able to Outside Evaporator

Inside

Condenser Condenser section

Qout

Insulated section

Qout

Condensate working fluid

Adiabatic section

Vapour flow Condensate working fluid

Insulated section Q (A)

Qin

Evaporator section

(B)

Wickless heat pipe (gravity assisted)

Q Wicked heat pipe

Fig. 3 Heat pipe working cycle [4].

Cooling Heating Vapour

Wick

Vapour

Liquid condensate

Liquid condensate

Heating

Liquid

(A)

Fig. 4 Structure of thermosyphons (A) and heat pipes (B) [5].

Cooling (B)

Qin

74

Heat Pipes

endure the pressure and temperature changes of the heat pipe structure during storage and nominal operation; are characterized by a high thermal conductivity facilitating heat transfer between the heat pipe and heat sources or sinks; and are load and corrosion resistant in their working environment. Finally, the manufacturing cost should be taken into account when selecting a casing material. Typical heat pipe casings are made from glass, ceramics, or metal. A common heat pipe casing material is copper due to its high thermal conductivity. However, in applications where strength and/or weight are primary requirements, alloys of aluminum, titanium, magnesium, or even stainless steel are used. Usually, metals are preferred as a construction material for the heat pipe casing due to their mechanical strength and high thermal conductivity, but recently silicon has been highlighted as the best material to replace metals due to its simple fabrication and compatibility with semiconductor devices and polymer-based casings are highly attractive due to their flexibility and low cost. However, by far, the most common in use container materials are copper, aluminium, and stainless steel. Copper is preferred for operating temperatures between 0 and 2001C, while aluminum is an ideal choice due to its weight advantages. Carbon steel casings have not been used when water is chosen as working fluid, due to hydrogen generation problems in the container which limits the operational performance of the heat pipe [6].

4.3.2.2

Heat Pipe Wick Structures

The wick structure consists of flow channels to connect the condenser and the evaporator ends of the heat pipe, through which the working fluid returns to the evaporator by capillary action. Any wick structure design has to satisfy two conditions: to provide a low resistance flow path for the working fluid and to increase the wick pumping pressure for faster flow. In order to meet both of these design conditions wick structures are usually made of several different materials arranged with certain geometrical characteristics. In the part of the heat pipe where the heat is lost (condenser) the design of the wick has large internal pore sizes, which reduces the flow resistance there, while the wick structure toward the evaporator side contains small pore sizes, which increases the capillary pressure and the thermal conductivity of the flow. The following two types of design are the typical wick structure configurations; either homogeneous wicks made of a single material or composite wicks containing two or more materials. Single material wicks Axially grooved Annular and crescent Artery wick Composite material wicks Screen-covered groove Slab

The highly conductive metal paths minimize the radial temperature drop – space applications Show small resistance to liquid flow but are vulnerable to liquids of low thermal conductivity Reduces the thickness of the radial heat flow path and provides a low-resistance path The fine mesh screen increases the capillary pressure, the axial grooves reduce the flow resistance and the metal structure reduces the radial temperature drop The fine mesh screen at the surface increases the capillary pressure, the coarse screen inside the slab assures the liquid flow, the threaded grooves provide uniform circumferential distribution of the flow, and enhance radial heat transfer

The selection of the appropriate wick is closely connected to the properties of the working fluid. Besides the casing material and the working fluid used, the heat pipe’s operation is governed by the pressure difference between the evaporator and condenser sections that the wick capillaries can provide. Moreover, the wick structure has to ensure the sufficient and uniform distribution of the liquid around the evaporator area. Finally, the thickness of the wick affects the total performance of the heat pipe. Increased wick thickness leads to increased heat transport capabilities; however, it also increases the thermal resistance across the wick, and thus reduces the overall thermal performance of the heat pipe.

4.3.2.3

Heat Pipe Working Fluids

Theoretically, a heat pipe can operate at any given operational temperature, as long as that temperature is between the triple state and the critical point of the working fluid utilized. Both of these state points refer to the pressure–temperature curve of a substance and are defined as follows: the triple point refers to the state (temperature and pressure) where the three phases (vapor, liquid, solid) of a substance coexist; while the critical point is the end point under which the liquid and the vapor phase of a substance can coexist (Table 1). After the choice of operating temperature range, the prime requirements for a suitable working fluid selection are: the material compatibilities with casing and wick; high thermal stability; the latent heat, surface tension and thermal conductivity; low liquid and vapor viscosities; the thermodynamic behavior of the fluid; and finally the acceptable pressure range and freezing points. The amount of working fluid used plays a significant role in the performance of a heat pipe. If the heat pipe is charged with an excess of working fluid its heat transfer coefficient is decreased, as the heat finds greater resistance traveling through the additional fluid. If the heat pipe is undercharged, the diminished liquid flow increases the heat losses due to friction forces, reducing the heat pipe’s performance. An important factor when selecting a working fluid is the Dunbar parameter, which determines the heat transfer capacity of the working fluid; higher Dunbar parameter value means higher heat transfer capacity. The Dunbar parameter is defined as a function

Heat Pipes

Table 1

75

Triple and critical points for common working fluids

Working fluid

Propane Neon Hydrogen Helium Oxygen Nitrogen Ethane Butane Pentane Methanol Toluene Acetone Ammonia Mercury Water Cesium Potassium Sodium Lithium

Triple point

Critical point

Temperature (1C)

Pressure (kPa)

187.68 248.6 259.31 270.973 218.79 209.97 182.78 185.35 128.69 97.7 94.99 94.3 77.7 39.0 0.01 – 63 801 –

1  10 7 43.37 7.04 5.048 0.152 12.6 11.3  10 7  10 4 76  10 3 19  10 5 4  10 5 2.32  10 6.076 1.65  10 0.62 – – 30  10 3 –

Temperature (1C) 96.672 228.66 239.95 267.96 118.6 147 32.17 146.14 196.7 240 318.64 235 132.4 1476.9 373.946 1664.85 1949.85 3600 2950

4

3

7

Pressure (MPa) 42.5 2.7 1.3 0.2 5.05 3.4 4.9 4 3.36 8 4.1 4.8 11.3 174 22.06 9.50 16 26 66.1

of the evaporative latent heat (l), the surface tension of the working fluid (s), the density (rn) and viscosity (mn) of the vapor of the working fluid and is given by the following formula [7]: Du ¼

4.3.2.4

l1:75 srn mn0:25

ð1Þ

Heat Pipe Advantages

Two phase systems such as heat pipes offer key advantages over conventional single-phase systems. First of all, they are capable of transporting the same amounts of energy as the single phase liquid or gas systems, but with considerably smaller mass flow rates, due to the high sensible and latent heat capacity of the working fluid. As a result, heat pipes offer smaller sized systems with much greater heat transfer coefficients than conventional single-phase systems. The amount of heat that can be transported through the use of latent heat is typically several orders of magnitude greater than transported by sensible heat for a geometrically equivalent system. Finally, heat pipe systems do not require the use of any external mechanical system, such as pumps or fans, to circulate the working fluid, increasing their reliability and minimizing maintenance requirements, and operating costs. The four main reasons for using heat pipes are:

• • • •

for heat transfer applications, where efficient heat transfer with small temperature differences is the primary purpose, for isothermal applications, where reduction of pre-existing temperature gradients of a body and operation with isothermal surfaces is the primary purpose, for temperature control applications, where the heat pipe controls the temperature of a body, for heat flux transformation applications, where the heat source and the heat sink require different heat fluxes.

4.3.3

Heat Pipe Types

Classification of heat pipe designs can be considered in terms of their operating temperature and therefore the type of working fluid utilized, their geometry/operation or their application.

4.3.3.1

Classification Using the Operating Temperature

Type

Comment

Cryogenic heat pipes

Any type of heat pipe design. Cryogenic heat pipes reported in the literature can be categorized into three types: thermosyphons

Temperature range 271 253 243 193

to to to to

2691C 2431C 2331C 1631C

Working fluid Helium Hydrogen Neon Nitrogen

76

Heat Pipes

[8–11], oscillating heat pipes (OHPs) [5,12–16] and loop heat pipes (LHPs) [7,17–24]. However, the investigation of the last two types has only been conducted theoretically and experimentally Low temperature heat pipes

Medium temperature heat pipes

High temperature heat pipes

4.3.3.2 4.3.3.2.1

1831C to 1331C 73 to 331C

Oxygen Propane

70 to 2701C

Short carbon chain organic fluids such as methanol, ethanol, ammonia, acetone or water Mercury, sulphur or long carbon chain organic fluids such as naphthalene and biphenyl Liquid metals such as potassium, sodium or silver

270 to 4201C

In industrial high-temperature processes, recovery and utilization of waste heat at these temperatures can have huge benefits for enhancing the system performance

420 up to 47001C

Classification According to Their Geometry/Operation Variable conductance heat pipes – Gas loaded heat pipes

VCHPs maintain the same operating temperature under varying heat inputs using gas-loaded control. This type of control relies on the effect of the presence of non-condensable gases on the condenser’s performance. By including these gases, they are able to block a part of the condenser’s surface and thereby reduce the heat flow capacity of the heat pipe. VCHPs affect this form of control by connecting the heat pipe to a reservoir containing the non-condensable gas. Regulating the pressure in the reservoir permits regulation of the operating temperature of the heat pipe condenser by controlling the heat transfer rate (Fig. 5).

4.3.3.2.2

Thermal diodes and switches heat pipes

Thermal switches are devices which permit or block the passage of heat depending on a set of conditions applied. Thermal diodes are devices that permit heat to flow in one direction only, such as thermosyphons. When heat is applied at the bottom of the heat pipe the heat is transferred to the top, but if the temperature distribution reverses the heat pipe does not work, as the working fluid cannot travel against gravity.

4.3.3.2.3

Oscillating heat pipes

OHPs operate by a pulsating motion of liquid slugs and vapor bubbles between the evaporator and the condenser created by pressure and temperature changes occurring during the phase change of the working fluid. The advantage of OPHs is that the liquid and vapor flow in the same direction without the need of a wick structure. The OHP generally consists of a series of tube bundles in serpentine configuration (Fig. 6). Parameters, such as the internal diameter, number of turns, working fluid, the length ratio of heating to cooling sections, and the inclination angle of the system, have been investigated to determine their effect on the OHPs thermal performance [27–30]. The latest development of the technology is the addition of nano- or microparticles into the base fluid, which seems to increase the heat transport capabilities of the system. This heat transfer enhancement appears to derive from the motion of the particles within the slugs (mixing them), rather than their thermal conductivity [31–33].

4.3.3.2.4

Loop heat pipes

The high heat transfer capacity, the ability to operate at any orientation, the ability to remove heat from several spatially separated sources, the distinct liquid/vapor transportation lines, but most importantly the controllability of LHPs distinguish them as attractive two phase heat transfer devices [34,35]. The LHP is a closed circuit, in which the vapor and liquid flows in different channels, and in addition to the evaporator and the condenser, a compensation chamber regulates the redistribution of the liquid, as shown in Fig. 10. In contrast with traditional heat pipes, LHPs separate the tubes and wicks (Fig. 7). During the operation of a LHP system part of the condenser is filled with vapor, which transmits the heat output, and the remaining part is filled with liquid. At the same time, the compensation chamber is partially filled with liquid and saturated vapor and their proportion changes according to the heat input. As heat is applied to the evaporator, the working fluid evaporates and the liquid is pumped by the capillary forces of vapor from the compensation chamber. As the temperature increases the vapor becomes superheated and due to the pressure drop at the exit of the evaporator, the vapor moves toward the condenser, where it condenses and gives up its latent heat. During condensation its temperature and pressure decrease to its boiling point, until all the

Heat Pipes

77

Gas reservoir No heat transfer

Cooled section Heat out

No heat transfer

Liquid flow

Adiabatic section

Vapour flow

Heated section

Heat in

Fig. 5 Variable-conductance heat pipe [25]. Reproduced from Sauciuc I, Akbarzadeh A, Johnson P. Temperature control using variable conductance closed two-phase heat pipe. Int Commun Heat Mass Transf 1996;23:427–33. doi:10.1016/0735-1933(96)00028-0.

(A)

(B)

(C)

Fig. 6 Types of oscillating heat pipes (OHPs): (A) a closed-loop oscillating heat pipe (CLOHP), (B) a closed-loop oscillating heat pipe with check valves (CLOHP/CVs), and (C) a closed-end oscillating heat pipe (CEOHP) [26].

78

Heat Pipes

Heat input Compensation chamber evaporator

Liquid flow

Vapour line Liquid line

Vapour flow

Condenser

Heat output Fig. 7 Loop heat pipe [26].

vapor changes phase into liquid. The liquid then flows toward the compensation chamber, where it is heated again and repeats the cycle. The two main functions of the compensation chamber are to regulate the amount of liquid in the loop during normal operation and to secure the constant wettability of the capillary wicks. Under steady state conditions the vapor temperature and pressure of the evaporator zone are higher than that of the vapor in the compensation chamber, which traps the liquid in the compensation chamber, allowing only the passage of the vapor, in which case the LHP is operating at constant conductance. At the same time, higher temperature vapor cannot diffuse back into the compensation chamber due to the capillary forces of the liquid/vapor interface of the chamber.

4.3.3.2.5

Micro-heat pipes

Originally micro heat pipes were developed as a microscale heat removal device for miniaturized applications, such as electronics cooling. The average length of micro heat pipes is a few centimeters and they have a hydraulic (internal) diameter between 50 and 600 mm [35,36]. Micro heat pipes are capillary-driven two phase heat transfer systems, which use micro wicks or channels to supply capillary force and small tubes to transfer heat to the condenser. The circulation of the condensate is achieved by the capillary pressure induced by sharp angle corners of the device. The heat transfer coefficient of the micro heat pipe depends on the rate of circulation of the working fluid, which can be regulated by the sharpness and the number of the corners. Heat transfer capacity increases with a decrease of the channel apex angle and the length of the pipe. The shape of the wick channel of the micro heat pipe can be triangular, square, hexagonal, or rectangular; the smaller the wick channel, the higher the heat transfer coefficient and the higher the heat transfer surface area per unit flow volume. In general, a micro heat pipe can be described as a closed microchannel filled with a working fluid and a long vapor bubble moving through the inner core of the channel. When a heat input is applied to the device the working fluid evaporates and the local increase of vapor pressure in the evaporator section creates a vapor flow toward the condenser. The vapor gives up its latent heat and condenses back to liquid. As the amount of liquid increases in the condenser section a pressure gradient is created, which drives the liquid back to the evaporator section, so the cycle can be repeated (Fig. 8). Despite the fact that micro heat pipes are a reliable and low maintenance passive cooling solution, the current technology is not very efficient in terms of heat transfer ratios compared to conventional heat pipes [38,39].

4.3.3.2.6

Rotating heat pipes

Rotating heat pipes (RHPs) are wickless heat pipes, divided into three sections, as with all heat pipes (condenser, adiabatic, evaporator), with the difference that the return of the condensate is driven by centrifugal forces. RHPs consist of a sealed hollow shaft filled with a fixed amount of working fluid. Rotation about the axis causes the condensate liquid to return to the evaporator [6]. RHPs can be classified into radial or axial, depending on the axis of rotation compared to flow. When the centrifugal force is parallel to the condensate flow from the condenser to the evaporator the RHP is called radial. When the centrifugal force is parallel to the axis of the heat pipe the RHP is called axial. Usual applications of RHPs are as heat dissipating devices in rotating machinery, heavily loaded bearings, and rollers for presses.

4.3.3.2.7

Controlled heat pipes – Variable conductance heat pipes

A type of control is usually necessary to assure the proper function of the heat pipe, as without it heat pipes tend to self-adjust their operating temperature in accordance to the temperatures developed at the evaporator and condenser ends. The control feature

Heat Pipes

Evaporator

Adiabatic section

Condenser

Lc

La

Lc

Heat input

79

Heat output

Solid wall

u1 uv

Liquid phase

u1

Vapor phase

tw

tw w tw w

tw Fig. 8 Micro-heat pipe [37].

gives the heat pipe the ability to maintain the evaporator at near constant temperature, regardless of the increase or decrease of the thermal input. These types of heat pipes are called variable conductance heat pipes and they can be categorized into four groups depending on the technique they use to moderate their thermal conductance [40].

4.3.3.2.8

Gas loaded heat pipes

The most common type of VCHP is the gas loaded heat pipe, which use an amount of non-condensable gases to eliminate the convective heat transfer, as described above. Gas loaded heat pipes can secure the isothermal operation of the condenser, despite any varying heat load inputs in the evaporator side. During the normal operation of the heat pipe, the buffer of the noncondensable gas is driven toward the condenser side, by the vapor movement. When it reaches the condenser, remains in a gas form, blocking partially the vapor flow and diminishing the condensation of the working fluid. As the heat input increases the pressure and the temperature of the vapor increases, thus the non-condensable buffer in the condenser section is compressed, allowing more vapor to condense. As a result, the condenser’s heat transfer coefficient is enhanced, while its temperature remains constant.

4.3.3.2.9

Excess liquid heat pipes

The operational principle of the excess liquid heat pipes is similar to the gas loaded heat pipes with the difference that the former use excess working fluid liquid to control the heat transfer capabilities of the condenser. This method utilizes a bellows and control fluid. By changing the heat load input, the pressure of the control fluid inside the bellows changes, which regulates the amount of working fluid to be driven toward the condenser.

4.3.3.2.10

Vapor flow modulated heat pipes

Vapor flow modulated heat pipes are equipped with bellows, control fluid, and a vapor throttling valve. As the heat load input conditions change, the pressure of the control fluid inside the bellows changes, which cause movement of the bellow and the opening/closing of the valve. Thus, the vapor flow is regulated and the thermal performance of the heat pipe is controlled.

4.3.3.2.11

Liquid flow modulated heat pipes

The operation principal of the liquid control is similar to the vapor flow control heat pipe, with the exception that in this case the flow regulation is conducted by a wick-lined reservoir located in the evaporator end (trap wick). Under normal operating conditions of the heat pipe the trap wick is dry, but when the orientation of the heat pipe changes or the heat input increases abnormally, condensation occurs in the trap. As the working fluid accumulates in the trap the heat pipe gradually stops functioning. The condensed liquid is ultimately returned to the evaporator section by the influence of gravity.

4.3.3.2.12

Flat heat pipes

Flat heat pipes are also known as vapor chambers. They rely on two-dimensional flow of vapor and liquid in the core of the casing and the wick, respectively. They are very similar to conventional cylindrical heat pipes, being capillary driven, but their

80

Heat Pipes

geometry is rectangular. They consist of an enclosed chamber whose inner surface is lined with a capillary wick structure, with the remaining volume containing the vapor. Heat sources and heat sinks are located on the chamber, with the other parts being thermally insulated. The working fluid is vaporized by the heat source at the evaporator, and the generated pressure difference drives vapor from the evaporator to the condenser, where it condenses and enters the wick structures [41–43].

4.3.3.2.13

Thermosyphons

A special heat pipe where the condensed liquid moves to the condenser by gravity is a thermosyphon. In this device no capillary structure is present and it works in a two-phase close cycle where latent heat of evaporation and condensation is used to transfer heat. In the design of thermosyphons, the condenser section has to be slightly elevated than the evaporator, so the return of the working fluid relies only on gravity. The thermal applications of thermosyphons are much wider than wicked heat pipes since the flow resistance is lower (due to the absence of wick structure) [44]. Thermosyphons find applications in various industries, including chemical, petroleum, gas, electronic cooling, telecommunications, energy storage, transportation, heating, cooling, air conditioning, cryogenics, renewable energy etc. However, thermosyphons have heat transfer limitations related to viscous, sonic, dryout, boiling, flooding, and entrainment factors [45]. Sonic limits mainly affect heat pipes which use liquid metals as working fluids. When the vapor velocity inside the thermosyphon reaches the speed of sound in that vapor, the core of the thermosyphons experiences a shock wave, which can cause mechanical failure of the device [6]. Viscous limits mainly affect heat pipes which are operating at low temperatures. When the vapor pressure difference between the evaporator and the condenser is lower than the viscous forces, the vapor cannot flow [46]. Dryout limits affect heat pipes where the filling ratio is very small. When the amount of working fluid inside the heat pipe is not sufficient to circulate the vapor and liquid films continuously, dry spots appear on the evaporator and it fails to operate optimally [46–48]. Boiling limits affect heat pipes with large filling ratios and high radial heat fluxes. When the transition from nucleation to film boiling occurs, a sharp decrease in the heat transfer rate occurs which can lead to mechanical damage to the device [49]. Flooding limits affect heat pipes with large filling ratios, high axial, and small radial heat fluxes. Flooding occurs when the vapor velocity is high enough to prevent the liquid film flowing back to the evaporator. Liquid drops are then carried by the vapor to the condenser creating dryout phenomena at the evaporator [49–51].

4.3.4

Systems and Applications

4.3.4.1

Heat Pipe Heat Exchangers

The high reliability and the passive high thermal conductivity of heat pipes have made them a very beneficial technology when combined with heat exchangers. Applications such as heating, ventilation, air conditioning, waste heat recovery, distillation, steam condensers, latent thermal energy storage (TES), district heating, and eco-friendly technologies like solar and geothermal energy have improved their efficiency when heat pipes were integrated into the conventional system designs. Heat energy can be reutilized and/or recycled by heat exchangers. Heat exchangers are devices that, as the name states, are able to extract heat from where it is unwanted and transfer it to where it may be usefully applied. They can be found essentially everywhere in modern industry due to current environmental policies, adopting different designs depending on their application including heat pipe-based designs. Heat transfer within a heat pipe heat exchanger is facilitated through the use of two phase closed thermosyphons. When equipping heat exchangers, thermosyphons are often preferred to wicked heat pipes mainly due to the reduction in costs from the absence of a wick structure. The advantages of thermosyphon heat exchangers over conventional heat exchangers include a good flow separation, no additional power input to the system, high reliability, lower initial investment and operating costs, the ability to work at lower temperature differences and a contingency plan design where if a pipe or number of pipes were to fail, the heat exchanger would remain operational. Heat pipe heat exchangers can be classified into two main groups based on the hot flow path through the heat exchanger: cross flow HP heat exchangers, standard and modular (Fig. 9), and through flow HP heat exchangers (Fig. 10).

4.3.4.2

Heating, Ventilation and Air Conditioning Applications

Building legislation, along with environmental and comfort concerns, are increasingly driving designers of building services and air conditioning equipment toward more energy efficient solutions. Heat pipe technology is emerging as a viable, efficient, and energy minimizing technology for applications in air handling unit designs. In the air conditioning industry, they provide a means for energy efficient treatment of the outside ventilation air, particularly in hot and humid climates where the cooling and dehumidifying of external ventilation air is a major contributor to high running costs of the building. Straight, wickless, heat pipes (thermosyphons) have been used in conventional heat pipe-based heat exchangers for energy recovery in ventilation systems. Here, heat recovery takes advantage of the level of energy (heating or cooling) in the extracted air from the building to either precool or preheat the outside ventilation air. Due to the orientation

Heat Pipes

81

Fig. 9 Cross flow HP unit (steam generator). Courtesy of Econotherm (UK) Limited.

Fig. 10 Through flow HP unit (gas to water). Courtesy of Econotherm (UK) Limited.

of these pipes they are limited to these applications. However, wraparound or LHPs can extend their design use. Air-to-air, air-to-water, water-to-water heat exchangers, and different working fluids such as R22, R134a, R407C and R410A have been intensively studied for these applications [42,52–56]. Wraparound heat pipes are designed, as their name suggests, to wraparound a conventional cooling coil. Gravity assistance is used to assure the return of condensed liquid by a unidirectional flow of liquid and vapor. This unidirectional flow overcomes previous obstacles to the use of long, small bore heat pipes where the performance has been limited by entrainment of the returning liquid in the high velocity vapors flowing in the opposite direction. When wrapped around the cooling coil, the heat pipe transfers heat from the warm outside air upstream of the cooling coil to the cold, dehumidified air downstream of the cooling coil. This technique is of great assistance to applications in which outside air is cooled and dehumidified prior to being supplied direct to the occupied space for ventilation. Its energy saving benefits are maximized when treating air that is both warm/hot and moisture laden. The benefits of this arrangement are twofold. Firstly, the precooling effect of the heat pipe serves to cool the air prior to it reaching the cooling coil; this not only allows more of the cooling coil surface to be used to actively remove moisture but also significantly reduces the cooling load. Secondly, the reheat effect of the heat pipe reduces the requirement for traditional electric reheat to bring the overcooled air back to its required supply temperature.

4.3.4.3

Air Handling Units

In ventilation systems, outside air entering buildings can be precooled or preheated using the return air from the building through the incorporation of a heat pipe heat exchanger. Wu applied a bundle of heat pipes in an actual air conditioning system and the

82

Heat Pipes

effects of preheating and precooling were examined; it was observed that the cooling capacity of the system was increased by almost 30% after the incorporation of the heat pipe bundle [57]. Various investigations have been carried out to understand the effect on the performance of heat pipe heat exchangers of different heat exchanger parameters and varying operating conditions. Konev et al. developed and tested theoretical models on the performance of an air-to-air heat exchanger on the basis of a collector heat pipe using Freon-113 as the working fluid [58]. Wadowski et al. investigated the effectiveness of gravity assisted air-to-air heat exchangers and concluded that a minimum temperature difference between two air streams is required in order to initiate operation.

4.3.4.4

Energy Conservation and Renewable Energy

The application of heat pipes is not limited to waste heat recovery and air handling units; heat pipes are relatively widespread for renewable energy applications as well. Solar collectors, geothermal heat pumps, and passive heating and cooling systems for buildings are some of the application that heat pipes can perform in renewable energy generation [26,42,53,54,59–62].

4.3.4.4.1

Solar energy collection

Solar thermal systems collect the sun’s heat most commonly for water heating purposes. A standard design for some time has been a copper thermosyphon heat pipe using water as a working fluid (operating at a reduced internal pressure) with a thin thermal collector bonded to it, enclosed within an evacuated tube, with the condenser end of the heat pipe (at the top of the collector) heating a water flow. The need to pump circulating water either entirely through the solar thermal panel (if not using a heat pipe design) on the roof or past the condenser of an evacuated tube heat pipe system is one of their major installation and running problems since this requires that water is circulated through the roof into the exterior panels and back into the building, with all the attendant problems of roof leaks, plumbing leaks, frozen pipes, etc. A further development of solar thermal panels is to integrate them with solar photovoltaics (PV/T). This has the advantage that the PV panels are kept cooler during peak solar intensity and can therefore generate more electricity, and, in some circumstances the solar thermal system can still generate hot water for the house. Again, most implementations of these approaches require pumping water out onto the roof and around the PV/T systems, with all the risks of potential failure and long-term maintenance issues. Instead of heat pipes being used to perform short distance heat transfer (from the evacuated tube to its end), the author has designed a long length (4 m), flat heat pipe which can be integrated with, or completely form, the roofing structure – the Solar Roof. A flat, internally finned, thermosyphon design is used, with aluminum as the casing material and ammonia as the working fluid, stretching the full height of the roof, with the condenser mounted at the top of the heat pipe, within the roof apex. A circulating water cooling manifold is secured to the back of the condenser section with a thermal interface material between them. The panel is sprayed with high emissivity paint to act as a thermal absorber. PV panels can be mounted on top of the heat pipe thermal panels to generate electricity, with their temperature moderated by the solar thermal heat pipe (Fig. 11).

The full scale testing of the solar roof at the Building Research Establishment (BRE), Watford, UK

Fig. 11 Flat heat pipe in assembly and integrated into roof with PV panels in some sections [61].

The back of the flat heat pipe panel

Heat Pipes

1200 Solar irradiation

1000

800

1000

800

Thermal mat

600

600

400

400 PV/T mat

200

200

0 9:00

Solar irradiation (W m−2)

Absorbed thermal energy (W m−2)

1200

83

10:00

11:00

12:00

13:00

14:00

15:00

0 16:00

Time of day Fig. 12 Absorbed solar energy for thermal heat pipe roofing system alone and integrated with PV panels [61].

Such an approach offers many potential advantages:

• • • • • •

isothermal surfaces, 100% utilization of the roof surface area, an efficient heat transfer process, ease of PV installation mounted on top of the heat pipe panels, PV panels deliver higher electrical output than uncooled mounted panels, cost effective compared to installation of solar thermal and PV panels separately, can form the skin of the roof to replace current roofing materials.

The reuse of the waste thermal energy produced during the cooling process of the solar PV panels remains a challenging aspect of hybrid collectors. To ensure high cooling efficiency for the PV, the temperature of the collector should be kept around 251C; the integration of a heat pump in the PV/T system is a potential method for achieving adequate hot water while sufficiently cooling the panels (Fig. 12).

4.3.4.4.2

Thermal energy storage

An efficient and cost-effective solution is a necessity if we wish to store industrial waste heat for later use to enhance the energy efficiency of an industrial process. TES has the ability to store energy for later use, improve the reliability and performance of the overall system, and reduce the mismatch between supply and demand of the available energy sources. TES has been used extensively in most buildings in the form of a hot water storage tank. There are many types of TES systems and many forms that energy can be stored in, such as sensible heat, latent heat, chemical, mechanical, and electrical TES systems. Recently, latent heat TES systems have begun to receive increased interest, due to their large energy storage density, low storage volume, and uniform temperature behavior. Latent TES systems utilize the storage capacity of phase change materials (PCMs); as a material changes phase, it can either release or store a large amount of energy in a small volume compared to sensible storage. These systems offer the possibility of storing greater amounts of energy, approximately 5–14 times more heat per unit volume, than sensible heat systems, and maintain nearly constant temperatures which depend on the phase change temperature of the PCM. Depending on the application, a TES heat exchanger will utilize a PCM that is designed to melt and solidify at a specific operating temperature. For a chosen temperature, the selected PCM should have a melting temperature that is slightly higher in value. The melted PCM in the heat exchanger has the ability to transfer its stored energy back into the circulating fluid, when solidifying, for the end-use application. However, PCMs are capable of only storing the thermal energy; in order to transfer this energy from the source to the PCM and from the PCM to the load a heat transfer medium combined with a heat exchanger is required. Therefore, the basic components of any latent TES systems are the appropriate PCM chosen according to the application, a suitable heat exchange device and a container compatible with the changes the PCM undergoes in different phases. Latent heat TES by PCMs is also an important technology for building energy conservation, solar energy utilization, and industrial heat recovery due to rising energy costs. While using PCM materials for TES is an attractive option, the low thermal conductivity of the PCMs is an issue. This is because, when charging or discharging heat to/from the PCM through the walls of the heat exchanger, a layer of the molten/solid PCM will

84

Heat Pipes

surround the heat transfer surface inhibiting the heat transfer between the PCM and the heat transfer fluid that facilitates the delivery/removal of the thermal energy. To overcome this low conductivity issue, various methods have been proposed, such as the attachment of fins to the heat transfer walls and the inclusion of metal particles, metal rings or carbon fibers of high conductivity into the PCMs. However, a more promising solution to deliver the heat effectively to within the bulk of the PCM is to use heat pipe technology there as well. The author developed a latent heat TES system using PCM to both store and release a large amount of energy in a small volume compared to the use of a sensible heat TES system. The low conductivity of PCMs was addressed using finned water-charged heat pipes embedded into the PCM bulk. Both heat pipes and the PCM tank were made of 316 L stainless steel and the PCM was PLUSICE S89, which has a melting temperature of 891C and crystallization point of 771C. The evaporator section of the heat pipe was heated by condensing a steam flow. The heat that was absorbed in the evaporator section was then discharged to the PCMs by the heat pipe multi-legged finned condenser. Tests were conducted for both charging (melting) and discharging (crystallization) of the PCM. It was observed that the thermal resistance posed by PCM during the discharging stage was higher compared to that during the charging process. Such systems should be able to reach a heat storage efficiency of around 50% if correctly optimized [63–72].

4.3.5

Analysis and Assessment

Heat pipe and heat pipe-based heat exchanger models have become increasingly sophisticated and considerable efforts have been made to develop analytical and computer simulation tools to enable developers to understand and predict the performance of their designs. The following section sets out the analytical approaches developed by researchers for modeling heat pipes and their incorporation into heat exchangers.

4.3.5.1

Heat Pipe Modeling

The heat pipe is a heat exchanger in its own right where the heat is transferred through it, passively, by latent means. Current advances in manufacturing techniques have led to a very cost effective production of wickless heat pipes (thermosiphons or gravity assisted heat pipe). The heat transfer through the heat pipe is done passively and relies on the highly efficient thermal transport process of evaporation/condensation cycles of the working fluid to transport heat from one end to the other where the heat can be dissipated through a heat sink. The efficiency of the heat pipe is usually linked to its overall thermal resistance. This resistance is a combination of the conduction resistances through the shell wall and the boiling and condensation resistances as per the following schematic:

The outside wall of the condenser

Pipe wall Tci Condenser

Rci

Tco Rcond_c

Tv

Th

Evaporator

The outside wall of the evaporator

Teo Reo

Rcond_e

Tei Rei

Tc Rco

Heat Pipes

85

As per the above schematic, the wickless heat pipe consists of a set of thermal resistances that are linked in series: Rhp ¼ Rcond_e þ Rei þ Rci þ Rcond_c

ð2Þ

The thermal performance is then the temperature difference divided by the total resistance: Teo Tco Rhp

Qhp ¼

ð3Þ

where Reo is the convection thermal resistance of evaporator section, Rcond_e is the conduction thermal resistance of the evaporator wall, Rei is the boiling thermal resistance, Rci is the condensation thermal resistance, Rcond_c is the conduction thermal resistance of the condenser wall, and Rco is the convection thermal resistance of condenser section. The heat pipe boiling resistance is calculated as 1 hei
Aei

Rei ¼

ð4Þ

where Aei is the heat pipe evaporator area (m²) and hei is the boiling heat transfer coefficient (W/m² 1C) [73–80]. The heat pipe condenser resistance is calculated as 1 hci
Aci

Rci ¼

ð5Þ

where Aci is the Heat pipe condenser area (m²) and hci is the condensation heat transfer coefficient (W/m² 1C) [43,81–86]. The forced convection resistances are calculated as Reo ¼

1 heo
Aeo

ð6Þ

Rco ¼

1 hco
Aco

ð7Þ

Rcond_e ¼

lnðDo =Di Þ 2pLe ks

ð8Þ

Rcond_c ¼

lnðDo =Di Þ 2pLc ks

ð9Þ

Conduction resistances are given as follows:

4.3.5.2

Heat Pipe Based Heat Exchanger Modeling

A heat pipe heat exchanger consists of a number of heat pipes that are linked, thermally, in parallel. The overall thermal resistance of the heat exchanger (here with five heat pipes) is given as Tc,in

Tc,out

Rhp,1

Rhp,2

Rhp,3

Rhp,4

Rhp,5

Th,in

1 1 1 1 1 1 ¼ þ þ þ þ ) Rhp;total ¼ Rhp;total Rhp;1 Rhp;2 Rhp;3 Rhp;4 Rhp;5

Th,out

1 Rhp;1

þ

1 Rhp;2

þ

1 1 Rhp;3

þ

1 Rhp;4

þ

1 Rhp;5

ð10Þ

86

The most common correlations to calculate the heat transfer coefficients (hei and hci) are listed in the following tables: Equation

Conditions

State

n ¼1 for water and n¼1.7 for other liquids

Nucleate boiling

Boiling Rohsenow [73]

" #3   cp
ðTei Tv Þ g
ðrl rv Þ 1=2
qei ¼ ml
hfg
s Csf
hfg
Pr nl

ð11Þ

where qei ¼ hei
DTv

Cooper [74]

0:12 0:4343 ln Rp

hei ¼ 55pr

0:55

ð log10 pr Þ

M

0:5 0:67 qei

ð12Þ

pr ¼

P Pcrit

Nucleate boiling

Rp the mean surface roughness Gorenflo [75]

hei ¼ h0 FPF ðqei =q0 Þn Rp =Rp0


0:133

ð13Þ

FPF ¼ 1:2pr0:27 þ 2:5pr þ 1 prpr n ¼ 0:9

Gorenflo [75]

ð14Þ

0:3p0:3 r

Nucleate boiling

Rp0 ¼0.4 mm and q0 ¼20,000 W/m2 All fluids except water and helium

ð15Þ

  0:68 2 p þ 6:1 þ FPF ¼ 1:73p0:27 r 1 pr r n ¼ 0:9

0:005rpr r0:95

For water and helium only where these are for water

Nucleate boiling

ð16Þ

0:3p0:15 r

ð17Þ

Pcritical ¼220.6 (bar) M¼18.02 h0 ¼5600 W/m² 1C

McNelly [76]

Forster and Zuber [77]

     qCp 0:69 Pkl 0:31 rl hei ¼ 0:225 s rv hf g qei ¼ 0:00122

C0:45 rl0:49 k0:79 p l 0:24 h0:24 s0:5 m0:29 l fg rv

where DTv ¼(Tei Mostinski [78]

Stephan and Abdelsalam [79]

Tv), Dpv ¼Pei " 

0:69 0:7 qei 1:8 hei ¼ 3:7  10 5 Pcrit

Nu ¼

!

P Pcrit

0:31 1

DTv1:24 Dp0:75 v

Nucleate boiling ð18Þ

ð19Þ

Dpv is the difference between the vapor pressures at the heating surface and liquid temperatures

Nucleate boiling

Pv and qei ¼hei DTv 0:17

    # P 1:2 P 10 þ4 þ 10 Pcrit Pcrit

hei Dd ¼ 0:24  107 X10:67 X4 1:58 X31:62 X85:22 kl

Nucleate boiling ð20Þ 10 4 o ð21Þ y¼451

P o0:88 Pcrit

Nucleate boiling – water

Heat Pipes

Author

     2  Cp Tv D2d qei Dd a rl ; ; X2 ¼ ; X3 ¼ 2 a kl Tv sDd       hfg D2d Cp ml r ; X5 ¼ v ; X6 ¼ ; X4 ¼ kl a2 rl     rl;ei cp;ei kl;ei r rv ; X8 ¼ l X7 ¼ rl rl Cp;l kl  12 s kl Dd ¼ 0:0208  y ; a¼ rl Cp g ðrl rv Þ

Nucleate boiling

X1 ¼

Nu ¼

hei Dd ¼ 0:054X50:34 X10:67 X40:248 X8 4:33 kl

ð22Þ

5  10 3 o ð23Þ

P o0:9 Pcrit

Nucleate boiling – hydrocarbons

P o0:97 Pcrit

Nucleate boiling – cryogenics

P o0:78 Pcrit

Nucleate boiling – refrigerants

y¼351 Nu ¼

hei Dd ¼ 4:82X70:117 X10:624 X30:374 X4 0:329 X50:257 kl

4  10 3 o ð24Þ y¼11

Nu ¼

hei Dd ¼ 207X10:745 X50:581 X60:533 kl

3  10 3 o ð25Þ y¼351

Simplified relations for boiling heat transfer coefficient to water at atmospheric pressure Jakob and Hawkins [80]

Q/A , (KW/m2)-h, (W/m2 1C)-Approximate range of DT, (1C)-Approximate range of h, (W/m2 1C) 1 Qei o16-hei ¼ 1042ðDTx Þ3 -DT ðrangeÞ ¼ ½0 7:76Š-hei ðrangeÞ ¼ ½0 2060Š A Qei o240-hei ¼ 5:56ðDTx Þ3 -DT ðrangeÞ ¼ ½7:32 14:4Š-hei ðrangeÞ ¼ ½2180 16o A

Horizontal surface ð26Þ

16600Š

ð27Þ

Q/A , (KW/m2)-h, (W/m2 1C)-Approximate range of DT, (1C)-Approximate range of h, (W/m2 1C) 1 Qei o3-hei ¼ 537ðDTx Þ7 -DT ðrangeÞ ¼ ½0 4:51Š-hei ðrangeÞ ¼ ½0 670Š A Qei o63-hei ¼ 7:96ðDTx Þ3 -DT ðrangeÞ ¼ ½4:41 9:43Š-hei ðrangeÞ ¼ ½680 3o A

where DTx ¼(Tei

Vertical ð28Þ

6680Š

ð29Þ

Tv)

Nusselt [81,82]

" rl ðrl

hci ¼ 0:943

3 rv Þgh fg kf

lc ml ðTv

Tci Þ

#14

ð30Þ

Properties in should be evaluated at the film temperature

87

ðTv þ Tci Þ Tf ¼ 2 0o Re o30

Laminar vertical

Heat Pipes

Condensation

_ ci 4Q 4Aci hci ðTv Tci Þ ¼ Pml h Pml h fg fg

ð33Þ

0

rv Þghfg k3f

lc ml ðTv

Tci Þ

0

Kutateladze [83]

hci;wavy ¼

Re kl 1:08Re1:22 "

Rever;wavy ¼ 4:81 þ Nusselt [81]

Chen [84]

ð34Þ 1 2 3

grl m2l

 3:7lc kl ðTv Tci Þ gr2l  m2l ml hfg

" rl ðrl

#14 3 rv Þgh fg kl cosC

" rl ðrl

3 rv Þgh fg kl

hci ¼ 0:943

Nusselt [81]

5:2

hci ¼ 0:725

lc ml ðTv

Dml ðTv

hci ¼ 0:728½1 þ 0:2ðN

1 #0:82

"

ð36Þ

Nusselt [81]

Nusselt [81]

3 rv Þgh fg kl

NDml ðTv

Tci Þ

#14

Tci Þ Inclined

Horizontal

3 rv Þgh fg kl

NDml ðTv

Tci Þ

#14

ð39Þ

The physical properties of the liquid film should be evaluated at an effective film temperature Tf Tf ¼ Tv þ 0.25(Tv

" #1 3 4 rl ðrl rv Þgh fg kl hci ¼ 0:729 Dml ðTv Tci Þ

hci ¼ 0:729

30o Re o1800 rn { rι ci Þ Tf ¼ ðTv þT 2  hfg ¼ hfg þ 0:68Cp ðTv

ð38Þ

rl ðrl

" #1 3 4 rl ðrl rv Þgh fg kl hci ¼ 0:725 NDml ðTv Tci Þ

" rl ðrl

0o Re o30 0o Re o30 rn { rι

for surface inclined by an angle C from the vertical

½ðN 1ÞJaŠo2 Ja ¼ Cp ðTv Tci Þ=hfg Nusselt [81]

Laminar- vertical

ð37Þ

#14

1ÞJaŠ

ð35Þ

3

Tci Þ

Tci Þ

hfg ¼ hfg þ 38 Cp ðTv Tci Þ 0o Re o30 Tf ¼Tv þ 0.25(Tv Tci) P: wetted perimeter

 2 13 1 grl hci D1:47kl Re ð3Þ m2l 

Equations may be used for vertical plates and cylinders and fluids with C ðT T Þ Pr40.5 and p hvfg ci r1

Heat Pipes

ð32Þ

" rl ðrl

hci ¼ 0:943

Re ¼

ð31Þ

#14

Nusselt [81]

Nusselt [81]

Tci Þ

88

h fg ¼ hfg þ 0:68Cp ðTv

ð40Þ

Horizontal - N rows

Tci)

Horizontal tube bank with N rows

Horizontal ð41Þ

ð42Þ

Horizontal tube bank with N rows

hci;N tubes ¼

Jouhara et al. [43]

 r exp 0:000067 l hci ¼ 0:85Re0:1 f rv where Re ¼

Labuntsov [85]

1 hci;1 tube N 1=4 " #1  3 4 rl ðrl rv Þgh fg kl 0:14  0:943 lc ml ðTv Tci Þ

Wavy laminar and turbulent

ð44Þ

_ ci 4Q pDml h fg

hci; vert; turbulent ¼

Re kl 8750 þ 58Pr l 0:5 ðRe0:75

" 0:0690lc k Pr 0:5 ðTv Revert; turbulent ¼ ml h fg

Chato [86]

ð43Þ

gr2l 253Þ m2l

 1 Tci Þ gr2l 3 m2l

  r ðr rv Þgk3l 3 hfg þ Cp ðTv hci; internal ¼ 0:555 l l ml DðTv Tci Þ 8 where Revapor ¼





13

Re 41800 ð45Þ #43

Turbulent

rn { rι

151Pr 0:5 þ 253

ð46Þ 1=4 Tci Þ

ð47Þ

Condensation inside horizontal tubes

 rv Vv D o35;000 mv inlet

Heat Pipes 89

90

Heat Pipes

4.3.6

Case Studies

The following case studies detail industrially installed and operating heat pipe-based systems the author has designed, and mainly cover heat pipe heat exchanger systems.

4.3.6.1

Steam Generator, China National Offshore Oil Corp, China Sea, April 2016

Gas to steam, cross flow heat exchanger:

• • • • •

GA 6400 smooth pipe 2 stage steam generator, on-site assembly, high reliability required for offshore location, low footprint required by space limitations, instant start up from gas turbine.

Exhaust temp in/out Water/steam temp in/out Exhaust flow/steam rate Energy recovered Recovered energy Project cost Payback period d/KW recovered

4.3.6.2

Gas to Air Unit, Automotive, Aluminum Furnace, USA 2008

4001C/2501C 501C/1801C 130,000/ 8000 kg/h 6.4 MW d2100k p/a d1200k 7 Months d328

Heat Pipes

Gas to Air, cross flow heat exchanger:

• • • • • •

GA 360 smooth pipe heat exchanger, 500 kW combustion air pre-heater, high particulate matter exhaust from furnace, low fouling, easy cleaning and maintenance, high reliability, unit positioned outside main factory premises, customer had been advised application was not feasible by consultants (due to acid corrosion, etc.)

Exhaust temp in/out Air temp in/out Exhaust/air mass flow Energy recovered Recovered energy Project cost Payback period $/KW recovered

4.3.6.3

4001C/2661C 301C/2931C 12,000/6374 kg/h 528 kW $155k p/a $150k 16 Months $123 (d76)

Exhaust to Coke Gas Unit, Steel Mill Blast Furnace, Czech Republic 2011

Gas to air, cross flow heat exchanger:

• • • • • •

heat pipe GPH, 12.6 MW duty, each unit consists of 1575  7.6 m helically finned, distilled water stainless steel heat pipes, unit performance increased significantly after upgrade, repeat order secured, full turnkey replacement delivered through local distributor.

Exhaust temp in/out Coke gas temp in/out Exhaust/air mass flow Energy recovered Recovered energy Project cost Payback period d/KW recovered a

2861C/1561C 62.21C/1971C 97,551/97,551 kg/h 12,626 kW d800k p/a d400,000a c. 6 Monthsa d31.67a

Estimated figures based on extrapolated installed costs.

91

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92 4.3.6.4

G2W, Shale Gas Well Head Fracking, Thermal Oxidizer, Canada 2012

Gas to Water, through flow heat exchanger:

• • • •

GW 2000 hybrid pipe heat exchanger, 2.2 MW fracking water heater: highly robust mobile unit for traveling around Canada, high particulate matter exhaust from furnace; removable panels incorporated for cleaning, low fouling, easy cleaning and maintenance, high reliability.

Exhaust temp in/out Water temp in/out Exhaust/water mass flow Weight of unit Exhaust pressure drop Energy recovered Recovered energy value Heat exchanger cost Payback period Price per kW recovered

4.3.6.5

8161C/1501C 51C/161C 11,016/180,000 kg/h 3600 kg 800 Pa 2260 kW d360k p/a d65k o3 months d27

Multi-Stage, in Series, Steam/Water, Natural Gas, Spirax Sarco, Italy 2012

Heat Pipes

Multi-stage steam generator and water pre-heater, cross flow heat exchanger:

• • • • •

anaerobic digester genset for the food industry, 1 stage of 1981C/12 bar steam generation, left-hand side, 2 stages of water heating delivered by 2 standard modules, visible on right of unit, pipes screwed in from underneath on standard modules; modules can be removed individually, hinged access doors for easy cleaning.

Exhaust temp in/out Water temp in/out Exhaust/water mass flow Weight of unit Exhaust pressure drop Energy recovered Recovered energy value Heat exchanger cost Payback period Price per kW recovered

4.3.6.6

4201C/1601C 1601C/1981C (12 bar) 11,484/ 900 kg/h 1852 kg 700 Pa 520 kW d19k p/a €19k 12 months d36

Steam Condenser, Food, Dirty Steam, Ireland, 2010

Steam condenser/hot water, through flow heat exchanger:

• • • • •

SC model 400 smooth/finned hybrid pipe ‘through-flow’ heat exchanger, 440 kW process water heater, contaminated steam; regulatory requirement to condense, system delivers fuel savings, stainless steel, eliminated existing air-cooled equipment.

93

Heat Pipes

94

Steam temp in/out Water temp in/out Exhaust/water mass flow Weight of unit Exhaust pressure drop Energy recovered Recovered energy value Heat exchanger cost Payback period Price per kW recovered

4.3.6.7

1051C/951C 101C/881C 844/8000 kg/h 300 Kg N/A 446 kW 2  d20k p/a 2  d10k 6 months d22

G2W, in Line Through-Flow Recuperator, Biomass Incinerator, Bologna, Italy

Gas to water, through flow heat exchanger:

• • • •

GW through flow heat exchanger, 2.1 MW waste water treatment plant biomass incinerator plant: highly robust low fouling unit, high organic particulate matter in exhaust; removable panels incorporated for cleaning, low fouling, easy cleaning and maintenance, high reliability.

Exhaust temp in/out Water temp in/out Exhaust/water mass flow Weight of unit Exhaust pressure drop Energy recovered Recovered energy value Heat exchanger cost Payback period Price per kW recovered

6101C/1501C 731C/901C 17,500 Kg/h 5000 kg 400 Pa 2100 kW €315k p/a €172,500 o18 months €140/kW

Heat Pipes

4.3.7

95

Closing Remarks

The application of heat pipes in standard and novel configurations and their integration with heat exchangers offers many benefits in managing the transfer and exchange of heat. The following lists the major advantages of the approach: Multiple redundancy:

•

each pipe operates independently so a whole unit configured from multiple heat pipes is not vulnerable to a single pipe failure. Prevention of cross contamination between fluid streams in heat exchangers:

•

each heat pipe acts as an additional buffer between the two fluids, significantly reduced the chance of failure states resulting in cross contamination between the fluid streams. Better fouling management:

•

use of smooth surfaced pipes allows exchangers to be used in high particulate or oily applications. Ease of cleaning and maintenance:

• •

can be maintained in situ with no uninstall, can be used with either manual or automated cleaning systems. Isothermal operation – no hot or cold spots:

• • •

heat pipes eliminate cold corners and condensation, allows greater energy recovery, thermal stability and control enables greater longevity for thermal oil. Robust materials and long life:

• •

good design practices allow pipes to freely expand and contract, exerting minimal thermal stress on the overall structure, so most thermal expansion occurs at the heat pipes which can be mounted so that this is fully accounted for, thick pipe walls resist erosion and corrosion. Intermediate pipe working temperature:

•

allows higher exhaust temperature limits on some applications. Highly scalable, customisable and configurable:

• •

modular design allows on site assembly, can be designed for future expansion, to meet specific application or operational needs. Reactivity:

•

fast reaction time, offers different control options and suitable for sensitive apparatus: does not require preheating. Passive devices:

•

no need for pumping energy to drive the heat transfer process through the heat pipe.

4.3.8

Future Directions

The use of heat pipe technology in heat exchange and thermal management of challenging scenarios is expanding fast due to their advantageous characteristics compared with conventional heat exchange and temperature control systems. Advances in the design and capabilities of heat pipes have led to the development of cost effective manufacturing techniques for both wicked and wickless heat pipes and this in turn is creating new areas of implementation for heat pipe based systems. The developments of the heat pipe based systems in new waste heat recovery, temperature control and thermal management applications are demonstrating relatively short Return On Investment timescales. In addition, with advances in automation and development in materials sciences, new heat pipe materials can be investigated to deal with challenging areas that have so far been out of reach for conventional solutions, particularly in dealing with high temperature and strongly contaminated flows.

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Relevant Website https://www.ashrae.org/home American Society of Heating, Refrigerating and Air-Conditioning Engineers (ASHRAE).

4.4 Heat Pumps Arif Hepbasli, Yasar University, Bornova, Izmir, Turkey r 2018 Elsevier Inc. All rights reserved.

4.4.1 Introduction 4.4.1.1 Importance 4.4.1.2 Contribution Objectives 4.4.2 Background 4.4.2.1 Historical Development 4.4.2.2 Classification of Heat Pumps 4.4.2.3 Working Principle of Heat Pumps 4.4.3 Energy Efficiency of Heat Pumps 4.4.4 Thermodynamic Analyses of Heat Pumps 4.4.4.1 Conventional Energy and Exergy Analyses 4.4.4.2 Enhanced (or Advanced) Exergy Analysis 4.4.4.3 Enhanced (or Advanced) Exergoeconomic Analysis 4.4.5 Energy, Exergy and Exergoeconomic Diagrams of Heat Pumps 4.4.5.1 Energy (Sankey) and Exergy Loss and Flow (Grassmann) Diagrams 4.4.5.2 Exergoeconomic-Based Flow Diagram 4.4.6 Illustrative Example(s) or Case Studies 4.4.6.1 Illustrative Example 1 4.4.6.1.1 System description 4.4.6.1.2 Analysis 4.4.6.1.3 Results and discussion 4.4.6.2 Illustrative Example 2 4.4.6.2.1 System description 4.4.6.2.2 Analysis 4.4.6.2.3 Results and discussion 4.4.7 Future Directions 4.4.8 Concluding Remarks Acknowledgments References Further Reading Relevant Websites

Nomenclature c C_ COP E E_ EER Ex _ Ex f h I IP_

Unit exergy cost rate ($GJ ) Exergy cost rate ($h 1) Coefficient of performance ( ) Amount of energy (kJ) Energy rate (kJs 1 or kW) Energy efficiency ratio (Btu W 1 h 1) Amount of exergy (kJ) Exergy rate (kJs 1 or kW) Exergoeconomic factor ( ) Specific enthalpy (kJ kg1) Current (A) Improvement potential rate (kJs 1 or kW)

Greek Letters c Specific exergy (kJ kg 1) Z Energy efficiency ( )

98

1

99 100 101 101 102 102 103 105 106 106 107 108 110 110 111 112 112 115 116 117 117 118 118 119 121 122 122 123 124 124

_ m PEF Q _ Q RI s SF SPF T V W _ W Z_

Mass flow rate (kgs 1) Primary energy factor Amount of heat (kJ) Heat transfer rate (kJs 1 or kW) Relative irreversibility ( ) Specific entropy (kJkg 1K 1) Sizing factor ( ) Seasonal performance factor ( ) Temperature (1C or K) Voltage (V) Work (kJ) Work rate or power (kJs 1 or kW) Hourly levelized cost rate of investment ($h 1)

Cos (j) e

Power factor ( ) Exergy (or second law) efficiency (

Comprehensive Energy Systems, Volume 4

)

doi:10.1016/B978-0-12-809597-3.00404-1

Heat Pumps

Subscripts act comp con D dest elec evap ex exp f GSHP HP in

int k max mech out p prt r s T U w ww 0

Internal kth component Maximum Mechanical Output, outlet Product Pressure relief tank Refrigerant Isentropic Total Useful Water Wastewater Reference (dead) state

Superscripts AV Avoidable AVEN Avoidable-endogenous AVEX Avoidable-exogenous EN Endogenous

EX UN UNEN UNEX

Exogenous Unavoidable Unavoidable-endogenous Unavoidable-exogenous

Abbreviations AC Alternative current ACHP Alternative current heat pump AHU Air handling unit ASHP Air source heat pump DC Direct current DCHP Direct current heat pump DH Domestic heating DHW Domestic hot water GSHP Ground source heat pump HE Heat exchanger

HP HPWH HTHP HRE ktoe RE RES SACHP WWHE WWSHP

Heat pump Heat pump water heater High temperature heat pump Heat Roadmap Europe kiloton of oil equivalent Renewable energy Renewable energy sources Solar assisted compression heat pump Wastewater heat exchanger Wastewater source heat pump

4.4.1

Actual Compressor Condenser Drive Destroyed or destructed Electrical Evaporator Exergetic, exergy Expansion valve Fuel Ground source heat pump Heat pump Input, inlet

99

Introduction

Heat pumps (HPs) can be considered part of the environmentally friendly technologies using renewable energy sources (RES) while they indicate a great opportunity for reaching the European Union (EU) target, which states a reliable, affordable, and sustainable energy supply. They have been utilized in the developed countries for years due to their higher energy utilization efficiencies and quoted in the European Directives on the use of renewable energy (RE), the Energy Performance of Buildings, Energy-related Products, and other relevant directives [1]. At present, HP technology has been quickly developed all over the world as a clean and energy efficient heating and air conditioning unit. It has been widely used in a number of places such as apartments, shops, hospitals, and office buildings. Based on the type of heat source, HP technology includes water-source, air-source, and ground-source HPs [2]. HPs may be defined in various ways. According to a definition based on the Directive 2010/31/EU of the European Parliament and of the Council dated May 19, 2010 on the energy performance of buildings (Article 2 point 18) “HP means a machine, a device or installation that transfers heat from natural surroundings such as air, water or ground to buildings or industrial applications by reversing the natural flow of heat such that it flows from a lower to a higher temperature. For reversible HPs, it may also move heat from the building to the natural surroundings.”

A broader and extended definition was also proposed by Borre as follows: “HP is a machine, device or installation using renewable natural energy sources from aerothermal, geothermal or hydrothermal heat or nonnatural processed wasted heat from water or air and which transfers it to buildings or industrial applications by reversing the natural flow of heat such that it flows from a low to a higher useful temperature” [1].

100

Heat Pumps

As far as some recent review studies on HPs are concerned, Hepbasli et al. [3] comprehensively reviewed wastewater source HP (WWSHP) systems in terms of applications and performance assessments including energetic, exergetic, environmental, and economic aspects. They concluded that the coefficient of performance (COP) values of the reviewed studies ranged from 1.77 to 10.63 for heating and 2.23 to 5.35 for cooling based on the experimental and simulated values. The performance assessments were mostly made using energy analysis methods while the number of exergetic evaluations was very low. Sarbu and Sebarchievici [4] provided a detailed literature review on the ground-source HP (GSHP) systems along with their recent advances. They reported that the GSHP technology may be used in both cold and hot weather areas and the energy saving potential is significant. Ni et al. [5] presented various applications and developments of HP technology in HVAC in China, while they reviewed the progress of researches, applications, and development in the field of HPs for building cooling/heating in China since the beginning of the 21st century. Gaigalis et al. [6] reviewed the HP simple implementation in Lithuania and European countries while they analyzed HPs technical segmentation in Lithuania and categories by heat collector, output (kW), and hot water production. Atam and Helsen [7,8] reviewed the state-of-the-art in modeling of ground-source HP (GSHP) systems and their optimal control along with the associated research challenges. The main focus is on optimal control but since design of an optimal controller may require a model, challenges in modeling and optimal control/optimal design of GSHP systems were also given. Arpagaus et al. [9] presented the major advantages and challenges of mechanically driven HPs and refrigeration systems with focus on multitemperature applications. They gave different design strategies, covering cycles with multistage compressors, ejectors, expansion valves, cascades, and separated gas coolers. The major part (about 70%) of multitemperature HP applications were in supermarket refrigeration, household refrigeration, air-conditioning, and for other refrigeration purposes. Fischer and Madani [10] reported HP systems in smart grids, and focused on their application and control approaches. They identified three main categories of applications using HPs in a smart grid context as follows: (1) economic operation of power grids, (2) the integration of RE sources, and (3) operation under variable electricity prices. Willem et al. [11] comprehensively reviewed HP water heater research on system energy efficiency and performance topics, focusing on laboratory and field (in-situ) experiments and measurements, modeling of energy use and efficiency, technological modifications or upgrades, and control operation strategies. It was concluded that most current HP water heater systems operated with the COP values between 1.8 and 2.5, while through some potential technological updates, COP values could reach a range of 2.8–5.5. Leonzio [12] performed a key literature review on absorption HPs integrated in solar systems and thermal energy storages to produce cool energy. It was highlighted that researches on the integration and control of various schemes for multiple uses (cooling, heating, water heating, and power generation) could produce synergic efficiency enhancement. The main objective of this chapter is to provide a comprehensive coverage of HP systems including historical developments and the technology classification and backgrounds. Some thermodynamic relations to assess their performances in terms of energetic and exergetic aspects are also included.

4.4.1.1

Importance

The synergy between the demand side energy savings and new supply options was analyzed within the framework of the Heat Roadmap Europe (HRE) studies, of which the first one covered the EU27 countries, with a focus on increasing domestic heating (DH) levels. In the second study, the potential for a combined retrofitting and DH strategy was investigated. Based on the HRE studies, it is estimated that the district heating (DH) share will potentially increase to 50% of the entire heat demand by 2050, with its approximately 25%–30% being supplied using large-scale electric HPs [11]. Fig. 1 illustrates European countries with the biggest projected HP thermal energy contribution for the period 2015–2020 in kiloton of oil equivalent (ktoe) (5) while Fig. 2 shows overall capacities, number of units and average capacities for HPs where the

3500 2020

3000

2015

ktoe

2500 2000 1500 1000 500

U

ni te

d

Ki

Ita l ng y do m Fr an c G er e m an Sw y ed en Fi nl an D en d m N a et he rk rla nd Be s lg iu m Au st ria

0

Fig. 1 European countries with the biggest projected HP thermal energy contribution for the period 2015–2020. Reproduced from Gaigalis V, Skema R, Marcinauskas K, Korsakiene I. A review on heat pumps implementation in Lithuaniain compliance with the National Energy Strategy and EU policy. Renew Sustain Energy Rev 2016;53:841–58.

400 300 200 100

Capacity (MW)

500

6 01 10 −2

20

06 −2

00

01

5 20

01 −2 20

19

96 −2

00

99

0

5

0 91 −1 19

19

86 −1

99

98

0

0

81 −1 19

101

600

40 35 30 25 20 15 10 5 0

5

Number of units

Heat Pumps

Years Number of units

Capacity (MW)

Fig. 2 Overall capacities, number of units, and average capacities for heat pumps (HPs). Data from David A, Vad Mathiesen B, Averfalk H, Werner S, Lund H. Heat roadmap Europe: large-scale electric heat pumps in district heating systems. Energies 2017;10:578.

establishment year is known [13]. All the values in Figs. 1 and 2 clearly indicate the importance of the HP sector. From 2010 to 2015 approximately 800,000 electrically driven HP units were sold in the EU (EU21) per year, adding up to more than 7.5 million units [10,14]. Based on a survey, which quantified the heat sources, refrigerants, efficiency, and types of operation of 149 units with 1580 MW of thermal output (operating at almost 80 locations across 11 European countries), the technical level of the existing HPs is mature enough to make them suitable for replication in other locations in Europe. As seen in Fig. 2, more HP units with smaller sizes were installed between 2010 and 2016 compared to the previous decade. The reasons for changing in size of HPs were generally due to the type of refrigerant used and the to the lack of surplus electricity to be integrated by HPs [13]. The technology is supported by increasing efficiency, the deployment of computing and communication technology, and increased renewable electricity generation. Considering that over a period between 2007 and 2015 the topic of HPs in smart grid constituted the focus of research, for a successful integration of HPs into a smart grid, it is very critical to use a holistic approach to the energy systems affected. Analyzing the smart grid barely from the electric perspective will lead to missing how HP system efficiency and indoor comfort will be affected by potential changes in HP control. Oppositely, if only HP efficiency becomes the main focus without considering the characteristics and expectations needed in the future electric system, this will cause considerable costs and waste of resources in the power system. Hence, a holistic perspective is needed for analyzing, designing, and operating the future energy system [10].

4.4.1.2

Contribution Objectives

The objective of this contribution is to provide a comprehensive coverage of HPs considering especially recently performed studies.

4.4.2

Background

Ever since the Stone Age, mankind has been able to produce heat by artificially sparked fires. The problem associated with the complex structure of artificial cooling was not solved until about 1850, when the first pioneers invented refrigeration machines. The same machines could be used in heating as HPs. The huge demand for cooling caused the rapid development of this technology [15,16]. The theoretical conception of the HP was described in a neglected book, published in 1824, and written by a young French army officer, Sadi Carnot. Its practical application on a large scale is attributable to designers J. Donald Kroeker and Ray C. Chewning; building engineer, Charles E. Graham; and architect Pietro Belluschi [17]. HPs have been supported by several countries since the 1970s as a strategy to improve energy efficiency, support energy security, reduce environmental degradation, and combat climate change [18]. Parallel to this development, the COP values of HPs have increased. Based on evaluation of over 800 HPs at nominal conditions listed in Ref. [19], COP values for market available HP units varied between 3.2 and 4.5 for air source HPs (ASHPs) and between 4.2 and 5.2 for ground source HPs (GSHPs) for testing conditions according to EN 14511 [9]. Besides, the world HP market increased 7.2% by volume in 2013 to almost 2 million units. The growth was due to the strong progression of sales of HP water heaters in the United States especially on the one hand, and the recovery of the European market on the other hand [20]. Although most of the installations occur in North American, Europe and China, the number of countries with GSHP installations increased from 26 in 2000, to 33 in 2005, to 43 in 2010, and to 48 in 2015. The equivalent number of installed 12 kW units (typical of United States and Western Europe homes) in 2015 was approximately 4.16 million [21].

102

Heat Pumps

4.4.2.1

Historical Development

Milestones of HPs are given below [11,16,17,22–24]:

• • • • • • • • • • • • • • • •

1748: William Cullen demonstrated artificial refrigeration. 1834: Jacob Perkins built the first practical vapor compression machine for producing ice. 1851: F. Carré designed the first commercially successful ammonia absorption cooling system. 1852: Lord Kelvin described the theory underlying HP. 1855: A.C. Twining presented the first commercial ice making plant using vapor compression refrigeration. 1855–1857: Peter von Rittinger developed and built the first HP. 1870: Van der Weyde invented the thermostatically controlled refrigeration system. 1912: GSHP was patented. 1937: Wilkes and Reed patented the first HP water heater (HPWH). 1945: John Summer built a full scale water source HP in Norwich. 1948: Carl Nielsen developed the first GSHP for use at his residence. 1950s: The industry began to manufacture domestic HPWH for the market. 1970s: The first investigations on gas engine driven HP (GEHP) were performed. 1977: The first GEHP was started up using in Dortmund–Wellinghofen open air swimming pool. 1985: The first merchandized GEHP was produced and introduced in the market. 1996: The world’s largest installation of GSHPs was completed at the U.S. Army’s Fort Polk military base in Leesville, Louisiana.

4.4.2.2

Classification of Heat Pumps

HPs may be classified in various ways. It is difficult to classify various types of HPs through a systematic approach because a number of elements, such as purpose of application, output, type of heat source, types of HP process, method, and technology, may be considered in the classification. Customarily, in the United States HPs are classified for the heating of buildings according to the type of heat source (first place) and type of heat carrier (second place). A distinction can be made between the terms as follows: HP, covering only the refrigeration machine aspect, and HP plant, which besides the HP itself also contains the heat source. This differentiation is due to heat from the heat source being transferred to the cold side of the HP by an intermediate circuit, the cold carrier. Another usual classification may be done as follows: (1) primary HPs utilizing a natural heat source present in the environment, such as external air, soil, ground water, and surface water; (2) secondary HPs reusing waste heat as heat source, i.e., already used heat, such as extract air, waste water, waste heat from rooms to be cooled; and (3) tertiary HPs, which are in series with a primary or secondary HP in order to raise the achieved, but still relatively low temperature further, for example, for hot water preparation. Furthermore, HPs generally classified by their respective heat sources and sinks. Depending on cooling requirements, it is possible to arrange HPs according to heat source and heat sink in practice as follows: (1) water-to-water, (2) water-to-air, (3) air-to-air, (4) air-to-water, (5) ground-to-water, and (6) ground-to-air [25]. Generally a residential HP system can be characterized by the type of heat source and sink, the technical features of the subsystems such as compressor type, refrigerant cycle properties and controls, and the heating system of the building. Fig. 3 shows main distinctive features of residential HP systems [10].

System characteristics

Heat source

Heat sink

Heat distribution

Heat storage

Heat supply concept

Unit capacity control

Air

Air

Ventilation

Tanks

Monovalent

Variable speed

Ground

Water

Underfloor

Building mass

Moneenergetic

On/Off controlled

Radiator

Borehole

Bivalent

Other

Fig. 3 Main distinctive features of residential heat pump (HP) systems. Adapted from Fischer D, Madani H. On heat pumps in smart grids: a review. Renew Sustain Energy Rev 2017;70:342–57.

Heat Pumps

103

System

Heat pumps

Driving energy

Model

Technology

Closed systems

Open systems

Compression heat pump

Sorption system

Mechnical vapour recompression

Absortion heat pump

Heat transformer

Process heat

Thermal vapour recompression

Steam

Direct gas/oil fired

Gas engine

Process heat

Electric engine

Steam

Gas engine

Electric engine

Steam turbine

Fig. 4 Classification of heat pumps (HPs) according to system, technology, model, and driving energy. Reproduced from Nellissen P, Wolf S. Heat pumps in non-domestic applications in Europe: potential for an energy revolution. In: Presentation given at the 8th EHPA European heat pump forum. 29.5.2015. Brussels. Belgium. Available from: http://www.ehpa.org/fileadmin/red/_EHPA_Archive_Forum/8th_Heat_Pump_Forum_ 2015/Presentations/Philippe_Nellissen_IHP_potential_EU.pdf; 2017 [accessed 01.08.17] and Nowak T. Heat pumps. Vision vs. reality. Available from: https://www.ee-ip.org/articles/detailed-article/?article=54&cHash=bd3568515e538ad350faf063915c6141; 2017 [accessed 01.08.17].

Compression work 1 2

Water condensation

IV

Industrial HP

4 III 3 Expansion

Heat sinks

Waste, exhaust heat

Space heating Condenser

Cooling tower

I Evaporator

Heat sources

Ground, air, water

Drying process Process heat

II

Fig. 5 Heat sources and sinks in industrial heat pump (HP) applications. Modified from Arpagaus C, Bless F, Schiffmann J, Bertsch SS. Multitemperature heat pumps: a literature review. Int J Refrig 2016;70:342–57.

Additionally, a classification may be made in terms of system, technology, model and driving energy, as illustrated in Fig. 4. A variety of HP technologies have been used in industrial and commercial applications. Besides this, the residential HP market is dominated by electric compression HPs [26,27]. HPs are highly attractive energy conversion devices for the industry due to offering efficient means for reducing primary energy consumption by utilizing heat recovery. Fig. 5 illustrates heat sources and sinks in industrial HP applications [9]. Most multitemperature HP cycles use two heat sources and one heat sink. Fig. 6 shows cycles applied for multitemperature HP applications [9].

4.4.2.3

Working Principle of Heat Pumps

A HP system mainly consists of three separate subcircuits, namely (1) a heat extraction circuit (e.g., the ground coupling circuit for a GSHP), (2) a HP unit (the refrigerant circuit or a reversible vapor compression cycle), and (3) a heat distribution system (the heating system of the building). In some applications HPs are integrated with RE sources, as shown in Fig. 7, which illustrates the working principles of a HP system consisting of six main parts, namely (1) a water-to-water HP unit, (2) a ground heat exchanger (HE) system consisting of two U-boreholes, (3) a solar collector system consisting of rooftop thermal solar collectors, (4) a domestic hot water (DHW) tank with an electrical supplementary heater, (5) a heating/cooling floor, and (6) circulating pumps [28].

104

Heat Pumps

Cycles for multi-temperature HPs

Multi-stage cycles

Multi-stage compressor Industry: Supermarket, heating, domestic hot water Challenge: Oil migtation TRL: Key technology in supermarkets

Cascade cycles

Separated gas cooler Industry: Water heating in residential buildings

Cascade Industry: Supermarket, high temperature HPs Challange: Temperature gap TRL: Established in industry

Challange: Energy efficiency TRL: under development, publications

Expansion valve Industry: Household refrigerator Challenge: Energy efficiency TRL: Established in industry

Secondary loop Industry: Supermarket Challange: Temperature gap and pumping losses

Ejector

TRL: Established in industry

Industry: Supermarket Challenge: Capacity control TRL: Prototype in industry, publications Multi-ejector Industry: Supermarket Challange: Capacity control TRL: Prototype status in industry Fig. 6 Cycles applied for multitemperature heat pumps (HPs). Adapted from Arpagaus C, Bless F, Schiffmann J, Bertsch SS. Multi-temperature heat pumps: a literature review. Int J Refrig 2016;70:342–57.

A HP provides heating, cooling, and domestic (sanitary) hot water for residential, commercial, and industrial applications while transforming the energy from the air, ground, and water to useful heat. This transformation is done via the refrigerant cycle, which consists of four main processes, namely evaporation, compression, condensation, and expansion, as explained below where a GSHP operation principle is explained. 1. Evaporation: a HP always has an outdoor heat source and an indoor outlet. The energy from outdoor sources, such as ambient air, exhaust air, ground-rock, groundwater, water, is infinite and hence renewable. This energy makes up about 75% of the energy delivered by HP. The fluid present in the pipes, buried in the ground, absorbs the heat from the ground, which has a stable temperature of around 10–121C throughout the year. This temperature is enough to heat the refrigerant due to its a very low boiling point, meaning that it is necessary for a very low temperature to heat up. The HE, the so-called evaporator, uses the thermal energy from the outdoor source to boil the refrigerant (the liquid in the HP) and turns it into a gaseous state. 2. Compression: the refrigerant then arrives at the compressor, which is considered the heart of the HP (or a refrigerator) where the refrigerant in gaseous state is compressed to a high pressure, that by consequence increases its temperature. For driving the compressor, additional energy is needed: from electricity, gas, or thermal. They make up 25% of the total energy needed to run the HP. If green electricity is used, for example, by means of photovoltaic, a HP then uses 100% renewables and is hence CO2 neutral. 3. Condensation: on the discharge side of the compressor, now hot and highly pressurized vapor passes through the HE, the socalled condenser, which allows the refrigerant to release the heat into the heating system for the house (air blower, floor heating, or radiators), and in connection with this the refrigerant is condensed, i.e., the refrigerant moves from gaseous into liquid state. The indoor outlet can be an air system (as the typical air conditioner units) or a hydraulic (water-based) system and is connected to a floor-heating system or radiators. For the provision of DHW, the indoor unit (also) exists of a hot water storage tank. 4. Expansion: the condensed refrigerant then passes through a pressure-lowering device, the so-called expansion valve. Lowpressure liquid refrigerant then enters another HE, the so-called evaporator, where the fluid absorbs heat and boils. From there on, the cycle starts again.

Heat Pumps

105

21 18

To load VII

VI

13

20

7

12

1

P4 19

2 I

10

IX

P IV

8

VIII

GSHP unit

P3

4 III

3

II

15

17 16

X

6 5

P2

2×90 m

Ground level

I II III IV V VI VII VIII IX-X

V Composed of two U-tubes

Compressor Condenser Capillary tube Evaporator Ground heat exchanger Solar collector Domestic hot water tank Floor heating system Pressure relief tanks

Fig. 7 Solar assisted domestic hot water (DHW) tank integrated ground-source heat pump (GSHP) system. Reproduced from Hepbasli A. Exergetic modeling and assessment of solar assisted domestic hot water tank integrated ground-source heat pump systems for residences. Energy Build 2007;39(12):1211–7.

4.4.3

Energy Efficiency of Heat Pumps

The COP is a widely used indicator for assessing the energy efficiency of HPs while it is defined as the ratio between useful effect produced (useful thermal energy EU) and energy consumed to obtain it (drive energy ED) as follows [4]: COP ¼ EU =ED

ð1Þ

If both usable energy and consumed energy are summed during a season (year), the seasonal COP (COPseason), which is also referred to as the seasonal performance factor (SPF), is used. The COP is given by the following relation in the heating operation: _ HP =W _ comp COPheating ¼ Q

ð2Þ

_ cooling =W _ comp EER ¼ Q

ð3Þ

_ HP is the thermal power of HP in W and W _ comp is the drive power of the HP (or the power input to the compressor). where Q In the cooling mode, the energy efficiency ratio (EER) is used in Btu/(Wh), as given below:

_ cooling is the cooling capacity of HP in in British thermal unit per hour (Btu/h) and W _ comp is the drive power of the HP (or where Q the power input to the compressor) or the COP for cooling is used as COPcooling. The COP of HP in cooling mode is calculated by: COP ¼ EER=3:413

ð4Þ

where the value 3.413 is the conversion factor for power to heat, namely 1 W ¼3.413 W/(Btu/h). _ HP Þ to the maximum heating demand Q _ Hmax as follows The sizing factor (SF) of the HP is defined as ratio of the HP capacity ðQ while it can be optimized in terms of energy and economics, depending on the source temperature and the used adjustment schedule [4]. _ HP =Q _ max SF ¼ Q

ð5Þ

106

Heat Pumps

The SPF of HPs is calculated by SPF ¼ QHP =Qelec

ð6Þ

where QHP is the annual produced heat and Welec is the annual used electric energy of the HP system [6]. The primary energy factor (PEF) of HPs (e.g., air, water, and ground source HPs) is defined as follows [6]: PEFHP ¼ PEFelec =SPF

ð7Þ

where PEFelec from various sources may be considered to be 2.49 for EU electricity mix, 3.5 for nuclear power, 1.14 for photovoltaic, 1.08 for hydroelectricity, and 1.03 for wind power.

4.4.4 4.4.4.1

Thermodynamic Analyses of Heat Pumps Conventional Energy and Exergy Analyses

The following section covers the relations on the system component basis illustrated in Fig. 4. Mass and energy balances as well as exergy destructions obtained from exergy balances for each of the main components (compressor, condenser, expansion valve, and evaporator) of the GSHP system illustrated in this figure are derived as follows [26]: Compressor (I): _ 2;s ¼ m _ act;s ¼ m _r _1¼m m

ð8aÞ

_ comp ¼ m _ r ðh2;act W

h1 Þ

ð8bÞ

_ comp;elec c2;act Þ þ W

ð8cÞ

_ dest;comp ¼ m _ r ðc1 Ex c ¼ ðh

h0 Þ

T0 ðs

s0 Þ

ð8dÞ

_ comp =ðZcomp;elec Zcomp;mech Þ _ comp;elec ¼ W W pffiffiffi _ com;elec ¼ 3 Vcomp Icomp Cos ðjÞ W

_ dest;comp;mech;elec ¼ W _ comp;elec ð1 Ex _ dest;comp;int ¼ Ex _ dest;comp Ex

Zcomp;elec Zcomp;mech Þ

_ dest;comp;mech;elec Ex

ð8eÞ ð8f Þ ð8gÞ ð8hÞ

where heat interactions with the environment are neglected. Condenser (II): _2¼m _3¼m _r m

ð9aÞ

_ 14 ¼ m _ 17 ¼ m _ w;fh m

ð9bÞ

_ con ¼ m _ r ðh2;act Q

h3 Þ

_ con ¼ m _ w;fh Cp;w;fh ðT14 Q _ dest;cond ¼ m _ r ðc2;act Ex

ð9cÞ

T17 Þ

_ w;fh ðc14 c3 Þ þ m

ð9dÞ c17 Þ

ð9eÞ

Expansion (throttling) valve (III): _4¼m _r _3¼m m

ð10aÞ

h3 ¼ h4

ð10bÞ

_ dest;exp ¼ m _ r ðc3 Ex

c4 Þ

ð10cÞ

Evaporator (IV): _1¼m _r _4¼m m

ð11aÞ

_6¼m _7 ¼m _ w;prt9 _5¼m m

ð11bÞ

_ evap ¼ m _ r ðh1 Q

h4 Þ

_ evap ¼ m _ w;prt9 Cw;prt9 ðT7 Q _ dest;evap ¼ m _ r ðc4 Ex

ð11cÞ T5 Þ

_ w;prt9 ðc7 c1 Þ þ m

ð11dÞ c5 Þ

ð11eÞ

Heat Pumps

107

Please note that exergy destruction may be calculated in two ways, of which the first one uses exergy balance equation while the second one is based on the entropy balance equation, through which the entropy generation is calculated. Multiplying the entropy generation by the temperature in Kelvin gives the exergy destruction. Exergy efficiencies of the system components and the whole system are evaluated as follows: Compressor (I): ecomp ¼

_ 1 _ 2;act Ex Ex _ comp;elec W

ð12Þ

Condenser (II): econ ¼

_ 17 _ 14 Ex _ w;fh ðc14 m Ex ¼ _ 2;act Ex _ 3 _ r ðc2;act m Ex

c17 Þ c3 Þ

ð13Þ

Expansion (throttling) valve (III): eexp ¼

_ 4 c Ex ¼ 4 _ 3 c3 Ex

ð14Þ

Evaporator (IV): eevap ¼

_ w;prt9 ðc5 c7 Þ m Ex_ 7 ¼ _ r ðc4 c1 Þ Ex_ 1 m

Ex_ 5 Ex_ 4

ð15Þ

GSHP unit (I–IV): eGSHP ¼

_ 17 _ heat _ 14 Ex Ex Ex ¼ _ _ W comp;elec W comp;elec

ð16Þ

_ 17 _ 14 Ex Ex _ pump;elec _ W comp;elec þ ðSÞW

ð17Þ

Overall GSHP system: eGSHP;sys ¼

The exergetic COP of the GSHP unit and whole system are as follows:   _ con 1 T0 Q Tcon COPex;GSHP ¼ _ comp;elec W

COPex;sys ¼

 _ con 1 Q

T0 Tcon



_ comp;elec þ ðSÞW _ pump;elec W

ð18aÞ

ð18bÞ

_ is expressible as: Van Gool [30] improvement potential on a rate basis, denoted IP, IP_ ¼ ð1

eÞ ðEx_ in

Ex_ out Þ

ð19Þ

Relative irreversibility is given by: RI ¼

4.4.4.2

_ i Ex _ExT

ð20Þ

Enhanced (or Advanced) Exergy Analysis

In the following, enhanced exergy analysis is applied to only the main components (I–IV) of the HP unit used in a food drying system, which consists of a GSHP system and a drying cabinet, as shown in Fig. 8 [31]. Before performing the enhanced exergy analysis, the balance equations for the conventional exergy analysis are written as listed in Table 1. After performing the detailed conventional exergetic analysis, the input results are used to conduct enhanced exergetic methods, of which methodology is outlined in Fig. 9 as a flow diagram of the processing steps. Exergy destructions of the system components are split into their avoidable (AV) and unavoidable (UN) and/or endogenous (EN) and exogenous (EX) parts. While current technological and economical design limitations should be determined to calculate the AV and UN parts, theoretical (ideal) operating conditions for each system component should be considered to figure out the EN and EX part [32]. Under the determined conditions, the system balances are achieved by arranging the mass flow rates by virtue of the mass and energy balances for the overall system while keeping the overall system exergetic product constant. Then, the conventional exergetic analyses are performed for each condition and splitting the exergy destructions is completed by using equations listed in Table 2. The modified exergy efficiency term, which focuses on the AVEN exergy destructions that indicate the realistic improvement potentials of the component by concentrating on the component itself, is computed [31].

108

Heat Pumps

VIII: Drying duct 10 8 II: Condenser 8a Fresh air

9

14 VII: Drying cabinet

2

3

I: Compressor III: Expansion valve 4

11

1 IV: Evaporator

7

5

V: Circulation pump 12

6

VI: GHE 13 Fig. 8 A schematic of the ground-source heat pump (GSHP) food drying system used in the analysis. Reproduced from Erbay Z. Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15.

Table 1 No. I

Conventional exergy balance equations for the main components of the GSHP system Control volume

1 11

8a

IV

2 3

2

II

III

Compressor

3

Condenser 9 Expansion valve

4 7

4 1

Evaporator

Balance equations   _ dest;comp _ 2 Ex _ 1 ¼ Ex E x_ 11 Ex  _ 2 Ex _ 3 Ex

ðE x_ 7

_ 3 Ex



 _ x9 Ex

 _ dest;con E x_ 8a ¼ Ex

_ dest;exp E x_ 4 ¼ Ex E x_ 5 Þ

 _ 1 Ex

Eq. no. (21)

(22)

(23)

 _ dest;evap _ 4 ¼ Ex Ex

(24)

5

Source: Reproduced from Erbay Z. Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15.

Afterwards, the four practical and beneficial terms, which are named as avoidable-endogenous (AVEN), unavoidableendogenous (UNEN), avoidable-exogenous (AVEX), and unavoidable-exogenous (UNEX), are determined with respect to the results of enhanced exergetic analysis by combining the two splitting concepts.

4.4.4.3

Enhanced (or Advanced) Exergoeconomic Analysis

Before performing the enhanced exergoeconomic analysis, the conventional exergoeconomic analysis, which gives the cost balance relations listed in Table 3 [31], is completed. Similar to the enhanced exergetic analysis, the important terms, such as cost rates associated with exergy destructions and hourly levelized investment costs, are split into their AV and UN and/or EN and EX parts in enhanced exergoeconomic analysis. The AV cost rate of the exergy destruction indicates the fuel costs used to comprise the AV destruction of the component when the overall exergetic

Heat Pumps

109

Conventional exergy analysis

Avoidable operating conditions

Theoretical operating conditions

System balance achieved

Systems balances achieved

Exergy analysis performed

Exergy analysis performed

Exergy destructions calculated

Exergy destructions calculated

Unavoidable/avoidable exergy destructions calculated

Endogenous/exogenous exergy destructions calculated

UN-EN, UN-EX, AV-EN, AV-EX exergy destruction rates calculated Fig. 9 Flow diagram for enhanced exergetic analyses. Reproduced from Erbay Z. Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15. Table 2

Relations used for enhanced exergy analysis

Parameters

Enhanced exergetic analysis

Eq. no.

UN

(25)

REAL UN AV E_ dest;k ¼ E_ dest;k þ E_ dest;k  EN EN REAL E_ E_ dest;k ¼ E_ p;k  E_dest;k

(26)

REAL EN EX E_ dest;k ¼ E_ dest;k þ E_ dest;k  UN UNEN EN E_ E_ dest;k ¼ E_ p;k  E_dest;k

(28)

UNEX

UN UNEN UNEX E_ dest;k ¼ E_ dest;k þ E_ dest;k

(30)

AVEN

EN E_ dest;k

AVEN E_ dest;k

(31)

AVEX

AVEX AVEN AV E_ dest;k ¼ E_ dest;k þ E_ dest;k

(32)

UN AV EN

UN REAL E_ dest;k ¼ E_ p;k 



E_ dest;k E_ p;k

(27)

p;k

EX UNEN

(29)

p;k

Performance parameters

UNEN ¼ E_ dest;k

emodified ¼

þ

REAL E_ p;k REAL UN AVEX E_ f ;k E_ dest;k E_ dest;k

 100

(33)

Source: Reproduced from Erbay Z, Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15.

product is arranged as constant. On the other hand, the EN part is associated only with the costs occurring within the kth component when the component being considered operates with its real conditions, but all other components operate ideally. Under these conditions, the overall exergetic product of the system is kept constant while all thermodynamic parameters are determined. The combination terms (UNEN, UNEX, AVEN, and AVEX) are then specified by combining the two splitting concepts. By calculating the AVEN costs associated with the exergy destructions and hourly levelized investment costs, the costs, which occurred in the analyzed system component and can be removed, can be determined. Moreover, by calculating the AVEX costs associated with the exergy destruction and hourly levelized investment costs, the costs, which can be decreased by structural improvements of

110

Table 3

Heat Pumps

Conventional exergoeconomic balance relations for the main components of the system

No.

Control volume

1

I

11

Compressor

2

8a III

3

Condenser

Expansion valve

7

2

Eq. no. C_ 1

(34)



  C_ 3 þ Z_ T ;con ¼ C_ 9

C_ 8a

c2 ¼ c3 ðF ruleÞ

9

(35)



C_ 3 þ Z_ T ;exp ¼ C_ 4

4

4

IV

 C_

3

2

II

Balance equations  C_ 11 þ Z_ T ;comp ¼ C_ 2

 C_ 7

1 Evaporator

(36)

  C_ 5 þ Z_ T ;evap ¼ C_ 1

C_ 4

c7 ¼ c5 ðF ruleÞ

5

(37)



Source: Reproduced from Erbay Z, Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15.

Table 4

Relations used for enhanced exergoeconomic analysis

Parameters UN AV EN

Eq. no.

Eq. no.

REAL

(48a)

UN UN C_ dest;k ¼ cf ;k  E_ dest;k

(38b)

REAL UN AV Z_ T ;k ¼ Z_ T ;k þ Z_ T ;k  REAL EN EN Z_ Z_ T ;k ¼ E_ p;k  E_ T ;k

(39a)

(39b)

(40a)

AV AV C_ dest;k ¼ cf ;k  E_ dest;k EN C_ dest;k

EN E_ dest;k

(40b)

(41a)

EX EX C_ dest;k ¼ cf ;k  E_ dest;k

(41b)

(42a)

UNEN C_ dest;k

UNEN E_ dest;k

(42b)

(43a)

UNEX UNEX C_ dest;k ¼ cf ;k  E_ dest;k

(43b)

AVEN E_ dest;k

(44b)

AVEX E_ dest;k

(45b)

EN EN Z_ T ;k ¼ E_ p;k 



Z_ T ;k E_ p;k

p;k

EX UNEN

EX EN REAL Z_ T ;k ¼ Z_ T ;k þ Z_ T ;k  UN UNEN EN Z_ Z_ T ;k ¼ E_ p;k  E_ T ;k p;k

UNEX

UNEX UNEN UN Z_ T ;k ¼ Z_ T ;k þ Z_ T ;k

AVEN

EN Z_ T ;k

UNEN ¼ Z_ T ;k

AVEX

AV Z_ T ;k

AVEN ¼ Z_ T ;k

Performance parameters

fkAVEN ¼

þ þ

AVEN Z_ T ;k AVEX Z_ T ;k

AVEN Z_ T ;k AVEN AVEN Z_ T ;k þC_ dest;k

 100

¼ cf ;k 

¼ cf ;k 

(44a)

AVEN C_ dest;k

(45a)

AVEX C_ dest;k

(46a)

AVEN AVEN AVEN C_ T ;k ¼ Z_ T ;k þ C_ dest;k

¼ cf ;k  ¼ cf ;k 

(46b)

Source: Reproduced from Erbay Z. Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15.

the whole system, or by improving the efficiency of the component being considered, and by improving the efficiencies of the AVEN remaining components, can be revealed. Finally, the modified exergoeconomic factor (fkAVEN ) and the total cost terms (C_ T;k ) are calculated to characterize the process or component more clearly by the help of the calculated enhanced exergoeconomic parameters. All equations used in enhanced exergoeconomic analysis are listed in Table 4.

4.4.5

Energy, Exergy and Exergoeconomic Diagrams of Heat Pumps

4.4.5.1

Energy (Sankey) and Exergy Loss and Flow (Grassmann) Diagrams

Energy or Sankey diagrams are used to visualize energy flows with proportional arrow magnitudes. Considering an air/water HP, _ heat , and W _ air , Q _ in ) on the external interfaces, as shown in Fig. 10 [33]. there are three energy flows (Q _ air ¼ Q _ evap ) in the evaporator is The energy and exergy flow diagram of the HP is illustrated in Fig. 11 where the heat flow (Q transmitted from ambient air with the temperature Tair over an average temperature gradient DTevap to the working fluid with the _ heat is transferred from the working fluid with a temperature of Tcon temperature Tevap. In a similar way, the generated heat flow Q via an average temperature gradient DTcon to the heating water with the average generated temperature Theat in the condenser.

Heat Pumps

111

Qheat T Theat

Tcon Condensation

∆Tcon ∆Theat ∆Troom

Theat Theat Troom

Compression

∆Tlift

Win

Expansion

∆Tlift ideal

Evaporation

Ambient temperature

Tamb Tamb

Condensation temperature Heating temperature generated Heating temperature required Room temperature

∆Tevap

Qair

Tevap

Evaporation temperature

Fig. 10 Heat Pumps (HPs) circuit along with energy flows and temperatures. Adapted from Gasser L, Wellig B, Hilfiker K. WEXA: exergy analysis for increasing the efficiency of air/water heat pumps. Final report. Forschungsprogramm UAW. Umgebungswärme. WKK. Kälte. Im Auftrag des Bundesamtes für Energie; 2008. 141 pages.

Qheat

Ex Qheat

Exloss,con

Wrev

HP

Wdrive Exloss,int

Exloss,evap

Qamb

Fig. 11 Energy and exergy flow diagram of a real-world heat pump (HP). Adapted from Gasser L, Wellig B, Hilfiker K. WEXA: exergy analysis for increasing the efficiency of air/water heat pumps. Final report. Forschungsprogramm UAW. Umgebungswärme. WKK. Kälte. Im Auftrag des Bundesamtes für Energie; 2008. 141 pages.

_ loss;evap and Ex _ loss;con . Other The temperature gradients in the evaporator and condenser DTevap and DTcon lead to the exergy losses Ex _ _ exergy losses occurring in the compressor and expansion valve are Exloss;comp and Exloss;exp valve , which are designated as the internal _ drive is not represented to scale. _ loss;int . Fig. 8 shows the exergy loss and flow diagram of the HP where the driving power W loss Ex The Sankey and the Grassmann diagrams of the HP system shown in Fig. 8 are drawn in Figs. 12 and 13, respectively. The energy flow (Sankey) diagram provides information about the energy transfers to or from the system while it gives no information about the changes in the quality of energy. Therefore, the Grassmann diagram, which gives quantitative information related to the share of the exergy input/output to/from the whole system, is drawn in Fig. 13.

4.4.5.2

Exergoeconomic-Based Flow Diagram

The results obtained from exergoeconomic analysis (a combination of both exergy and economic analysis methods) may be indicated through cost flow diagrams, which allow for the identification of the most important cost streams of the system with absolute and levelized costs.

112

Heat Pumps

1.84 kW

17,33 kW

VIII 0.13 kW 19,16 kW

Drying duct Drying cabinet

VIII-A

0.01 kW VII 19,30 kW

1.39 kW

15,93 kW

II

8.81 kW

I

Condenser

Compressor

Expansion valve

III

1.15 kW Evaporator 5.44 kW

7.66 kW

IV

16.76 kW 0.09 kW Circulation pump

V

18.98 kW

GHE

16.84 kW

VI

2.14 kW

Fig. 12 The Sankey (energy flow) diagram for the ground source heat pump (GSHP) system. Reproduced from Erbay Z, Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15.

In this regard, a cost flow diagram is shown in Fig. 14 where the highest cost rate streams are obtained in the drying part of the system. The cost rate of the internal streams increases through the addition of capital investment-related costs and the costs of exergy. Streams 8, 9, and 10 are air streams while one of the main reasons for these highest cost rates is the high destruction cost rates calculated in the condenser, the drying cabinet, and the drying duct. The other important reason is due to the high cost rates measured during the evaporation.

4.4.6 4.4.6.1

Illustrative Example(s) or Case Studies Illustrative Example 1

WW is considered a safe and locally available energy source. It is also a good heat sink because of the temperature level in the WW, being approximately 9–141C and 28–291C in winter and summer seasons, respectively [34]. The idea of heat recovery

Heat Pumps

113

Drying duct

0.116 kW

0.835 kW

VIII

0.719 kW

0.051 kW

VII

VIIIA Condenser

0.582 kW

Drying cabinet

0.886 kW

0.137 kW

1.722 kW 1.086 kW

II

0.332 kW

0.085 kW

Compressor

III

I

0.661 kW

Expansion valve

0.831 kW

0.269 kW

0.255 kW 1.145 kW

IV 0.219 kW

0.120 kW 0.087 kW

Circulation pump

Evaporator

V 0.167 kW

0.040 kW 0.056 kW

VI

0.108 kW

GHE

Fig. 13 The Grassmann (exergy loss and flow) diagram for the ground source heat pump (GSHP) system. Reproduced from Erbay Z, Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15.

114

Heat Pumps

Fig. 14 Exergoeconomic-based cost flow diagram for the ground source heat pump (GSHP) system. Reproduced from Erbay Z, Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15.

from WW through HPs is certainly not new. Since the 1980s, centralized systems in Germany, Switzerland, and in the Scandinavian countries have used the heat in WW in the sewage systems and the effluent of sewage treatment plants [35]. Especially in recent years, utilization of WWHPs has become very popular. The performance of a WWSHP system varies system by system according to operation conditions such as sewage temperature, clean water temperature, and HP equipment [34]. The illustrative example here covers a WWSHP system, which was designed, installed, and tested at Yasar University in Izmir, Turkey while the following analysis is based on the operation way in the heating mode [36].

Heat Pumps 4.4.6.1.1

115

System description

Fig. 15 shows a schematic view of the experimental WWSHP system while its pictures are illustrated in Fig. 16 [36]. There are three main subsystems: (1) a WW system, (2) a WWSHP, and (3) an end user system. The WW subsystem consists of three main parts: a PV/T system, wastewater heat exchanger (WWHEs), and WW tanks. A local WW drainage system, through which WW flows, is not utilized because the WWSHP system cannot be connected to it. Therefore, two 500 L tanks are used for simulation purposes. Water is stored in these tanks and circulated by pumps. An 8 kW resistance and cooling coil is located in one of the tanks for keeping the WW temperature constant, so that the system can reach steady state conditions. For transferring heat from/to the WW, two various WWHEs, a plate HE, and an immersed HE, connected in parallel, are utilized. In the HP subsystem, there exist two compressors (1 AC and 1 DC), three water source and one air source HEs, an electronic expansion valve, a four-way valve and some other auxiliary equipment, such as the drier, oil separator, etc. In the end user system, a fan-coil unit connected parallel to the air source HE and a DHW tank is in use. Valves 1, 3, 6, 8, 10, 11, 12, 13, and 16 are kept open for maintaining the heating mode. The air source HE is used as the condenser, while the DC compressor and plate type WWHE are considered to be the main units. On the other hand, the PV/T system is not included in the analysis (it is bypassed both electrically and thermally). The refrigerant is compressed to the condenser by the DC compressor, where it transfers heat to the indoor air. After that, it enters the electronic expansion valve and expands to the evaporator pressure. The refrigerant enters the evaporator in two-phase state and absorbs heat from the clean water in the intermedium cycle. The temperature of the water in the intermedium cycle is decreased in the evaporator and is then sent to the WWHE where WW is used to heat the intermedium water and pumped back to the WW tank. During this process, temperatures, pressures, flow rates, and power consumptions are continuously measured at the shown locations in the schematic and recorded in the data loggers.

I. AC Compressor II. Air Source Condenser III. Electronic Expansion Valve IV. Evaporator V. WWHE VI. Acumulator VII. WW Tank VIII. Oil Seperator IX. 4 Way Valve X. PV/T XI. Water Source Condenser XII. Immersed WWHE XIII. Drier XIV. Observation glass XV. WW Cooling Coil XVI. Plate Heat Exchanger XVII. Domestic Hot Water Tank XVIII. Fan-Coil Unit XIX. Electrical Resistance XX. DC compressor

Fig. 15 A schematic view of the wastewater source heat pump (WWSHP) considered. Reproduced from Araz M, Ekren, O, Biyik E, Gunerhan H, Hepbasli A. Experimental exergetic performance evaluation of a wastewater source heat pump system (WWSHP). In: Heiselberg PK. editor. CLIMA 2016 – proceedings of the 12th REHVA World congress, volume 3. Aalborg: Aalborg University. Department of Civil Engineering. 22–25 May. Aalborg. Denmark; 2016.

116

Heat Pumps

Fig. 16 Pictures of the wastewater source heat pump (WWSHP) system. Reproduced from Araz M, Ekren, O, Biyik E, Gunerhan H, Hepbasli A. Experimental exergetic exergetic Performance performance evaluation of a wastewater source heat pump system (WWSHP). In: Heiselberg PK. editor. CLIMA 2016 – proceedings of the 12th REHVA World congress, volume 3. Aalborg: Aalborg University. Department of Civil Engineering. 22–25 May. Aalborg. Denmark; 2016.

4.4.6.1.2

Analysis

The following assumptions are made in the analysis of the WWSHP system: 1. All processes are steady state and steady flow with negligible potential and kinetic energy effects and no chemical or nuclear reactions. 2. Water properties are used instead of wastewater. 3. The pressure losses in the pipelines and HEs are neglected. 4. Air is taken as an ideal gas at given conditions. 5. The mechanical and electrical efficiencies of the pumps are taken as 82% and 88%, respectively. 6. The values for the dead (reference) state and pressure are taken to be 21.341C and 101.325 kPa, respectively. Eqs. (8a)–(15) are used for analyzing the main components of the system. Additionally, the following relations may be written for the WWHE, pumps, and the entire system, respectively: WWHE (IX): _8¼m _ ww ; m _ 11 ¼ m _ w;int _7¼m _9 ¼m m _ wwhe ¼ m _ ww ðh7 Q

_ w;int ðh9 h8 Þ ¼ m

_ dest;con ¼ m _ rww ðex 7 Ex

ð47aÞ

h11 Þ

_ w;int ðex 11 ex8 Þ þ m

ex9 Þ

ð47bÞ ð47cÞ

_ 7 Ex _ 9 Ex

ð47dÞ

_ 14 ¼ m _ w;user _ 13 ¼ m m

ð48aÞ

_ 8 Ex _ 11 Ex

ewwhe ¼ Pumps (V–VIII):

_ pump ¼ m _ w;user ðh14 W

h13 Þ

_ pump ¼ W _ pump;elec Zpump;elec Zpump;mech W _ dest;pump ¼ m _ w;user ðex 13 Ex epump ¼

_ pump ex 14 Þ þ W

_ 13 _ 14 Ex Ex _ W pump;elec

ð48bÞ ð48cÞ ð48dÞ ð48eÞ

The functional exergy efficiency of the WWSHP system and overall system can be calculated using the following equations: _ 12 _ 14 Ex Ex _ comp;elec W

ð49Þ

_ 12 _ 14 Ex Ex P _ pump;elec _ W comp;elec þ W

ð50Þ

eWWSHP ¼ esystem ¼

Heat Pumps

The overall exergy efficiency based on product/fuel basis can be calculated from P P Exergetic product P_ i ¼P eoverall ¼ P Exergetic fuel F_ i

4.4.6.1.3

117

ð51Þ

Results and discussion

Using the equations presented above, one can evaluate energetic and exergetic performances of each component in the WWSHP system along with the entire system. Figs. 17 and 18 give the analysis results for the system components. It is clear from Fig. 17 that the greatest exergy destruction (irreversibility) occurred in the compressor approximately at 70 kW and is mainly related to the mechanical–electrical losses. This is followed by the expansion valve, the condenser, and the evaporator at values of 26 kW, 19 kW, and 13 kW, with exergy efficiency values of 83%, 72%, and 65%, respectively. The irreversibility associated with the expansion valve is due to the significant pressure drop, while that of the HEs is related to the temperature difference between the hot and cold fluids. The exergy efficiency of the WWHE is determined to be 80%, while that of the evaporator and condenser are 65.77% and 65.17%, respectively. Exergy efficiencies for WWSHP and the whole system are estimated to be 72.40% and 64.03%, on product/fuel basis, while their functional exergy efficiencies are obtained to be 25.1% and 18.6%, respectively. The highest improvement potential rate occurred in the compressor with a rating of 24 kW, while the WWHE has the minimum relative irreversibility value of 2.4%. In terms of relative irreversibility rates on the whole system basis, the highest value belongs to the WWSHP unit at about 64%, followed by the pumps at 34%, as seen in Fig. 18.

4.4.6.2

Illustrative Example 2

This illustrative example is related to a HP system of the energy efficient & cost competitive retrofitting solutions for shopping buildings (ECOSHOPPING) project, which is cofunded by the EC within the 7th Framework Program. It aims at producing a systematic methodology and cost-effective solutions for retrofitting commercial buildings. The “EcoShopping” platform integrates existing building services and interoperates with other ICT-based subsystems. The control and management of automation systems is based on advanced self-learning algorithms. The project cost is 4.10 million €, starting in September 2013 with a duration of 4 years [37]. The overall objectives of the project are to (1) reduce primary energy consumption down to less than 80 kWh/m2 per year, (2) increase the proportion of renewable energy sources (RES) to more than 50%, (3) investigate a retrofitting solution with innovative thermal insulation solutions and day lighting technologies, (4) develop and install a RES direct powered DC variable speed HP, (5) increase the building thermal mass with a view to reducing the energy consumption, (6) integrate the intelligent automation unit concept with a mobile robot, and (7) develop a solution for automatically identifying and predicting failures and inefficiencies in HVAC system performance [3–39]. The demo building IKVA Shopping Centre is a retail mall, located in the city of Sopron in Győr-Moson-Sopron County of Hungary and built in 1979. It has approximately 3700 m2 leasable area and two main sections, namely (1) the main building part Compressor (I)

Condenser (II)

Expansion valve (III)

Evaporator (IV)

250

(kW) or (%)

200 150 100 50

) (% it

(%

un

y nc

P

ie

y-

H

fic

ilit

ef

ib

gy

er s

er

ev

Ex

de

R

el

at

iv

Ex

e

er

irr

gy

er Ex

)

W ) (k te ra io n ru st

tic ge

pr tic ge er Ex

ct

fu

el

od uc

tr

ra

at

te

e

(k

(k

W

)

W

)

0

Fig. 17 Exergetic values for the heat pump (HP) unit components. Data from Araz M, Ekren, O, Biyik E, Gunerhan H, Hepbasli A. Experimental exergetic exergetic performance evaluation of a wastewater source heat pump system (WWSHP). In: Heiselberg PK. editor. CLIMA 2016 – proceedings of the 12th REHVA World congress, volume 3. Aalborg: Aalborg University. Department of Civil Engineering. 22–25 May. Aalborg. Denmark; 2016.

Heat Pumps

100 90 80 70 60 50 40 30 20 10 0

C on

C om

pr es so r( Ex d I) e pa ns nse r( io n II) va Ev Sy lv e a st In (II em po te ) ra rm w t o e a r Fi di t (I rs um er t-s pu V) Se w ta m co a p te nd ge (V rp se -s um ) w ta a ge p g (V se e p um I) w ag p e pu (VII ) m p (V W W W III) W SH HE (IX P un ) it Pu (I− m IV ps (V ) −V O ve III ) ra ll (I− IX )

Relative irreversibility (%)

118

Components Fig. 18 Relative irreversibility values for the whole wastewater source heat pump (WWSHP) system components. Data from Araz M, Ekren, O, Biyik E, Gunerhan H, Hepbasli A. Experimental exergetic performance evaluation of a wastewater source heat pump system (WWSHP). In: Heiselberg PK. editor. CLIMA 2016 – proceedings of the 12th REHVA World congress, volume 3. Aalborg: Aalborg University. Department of Civil Engineering. 22–25 May. Aalborg. Denmark; 2016.

(commercial area), which consists of two floors plus an open parking lot on the ground level, and (2) a service area with three floors and a basement [38,40].

4.4.6.2.1

System description

The total heat demand of the building, taking into account different functions, is 618.6 kW/year, of which 136.7 kW are used by the air heating unit. There are three HVAC systems: (1) a heating system (three condensing gas boilers), (2) a cooling system (local split air conditioners), and (3) a DHW (only some local, electrical water heater). The total thermal capacity of the boilers is 3x66 kW¼ 198 kW. The gas boilers supply only the radiators. In addition, there exists a ventilation system (2 air handling units (AHUs)), but these are operated only in summer, 2 h per day. The heating pipe for the HE of the AHU was cut off, so the AHU has no heating capability [37]. The HP system is integrated with a capillary tube radiant cooling/heating system with inlet water and outlet water temperature values of 321C and 281C, and 161C and 181C in the heating and cooling modes, respectively. Based on the simulation results, the total heating and cooling loads of the retrofitting area are determined to be 39.5 kW and 66.9 kW, respectively. The retrofitting surface area is 984.7 m2 with about a volume of 3250 m3 while the outdoor design temperatures for the heating and cooling seasons are 151C and 321C, respectively. Fig. 19 illustrates a schematic view of the HP system where there are 5 HPs, of which four are alternative current (AC) types while the remainder is a direct current (DC) type. A buffer tank with a volume of 1000 L is also used to improve the overall system operating efficiency by reducing unnecessary equipment short cycling [40].

4.4.6.2.2

Analysis

As part of the improvements to be made, it is aimed at assessing the performance of designing a RE powered HP system using energetic and exergetic analysis tools. For the exergetic assessment, relations given in Section 4.4.4.1 were used to calculate the exergy destructions and exergy efficiencies of the each component of the HP unit as well as the values for the relative irreversibility and the exergy efficiency of the whole system. The calculations were made on one state basis dated 25 May 2017 (time: 4:11) in the heating mode while the dead (reference) state temperature was taken to be 121C [41]. During the tests, the values for the ambient temperature, the compressor suction and discharge temperatures and pressures, the tank water temperature, the compressor power consumption, the coil (evaporator) temperature, the HP flow and return temperatures, the supply (entering) and return (leaving) temperatures of the capillary tube radiant cooling/heating system, and the mass flow rate of the HP water cycle were measured using the appropriate measurement devices. R-417A was used as the refrigerant (which is also an alternative to R-22 in medium temperature refrigeration and air conditioning). The mass flow rate of the refrigerant was not measured. The heating/cooling capacity on waterside of the HP is calculated first using Eq. (9d). The mass flow rate (kg/s) of the refrigerant flowing through the condenser/evaporator in the heating/cooling mode is then determined from the energy balance using Eq. (9c).

Heat Pumps

ACHP indoor unit

ACHP indoor unit

ACHP indoor unit

DCHP outdoor unit

ACHP indoor unit

ACHP indoor unit

119

DCHP indoor unit Separator

ACHP indoor unit

Buffer tank

Connection to cooling/ heating ceiling

Pump

system Fig. 19 A schematic view of the heat pump (HP) system. ACHP, alternative current heat pump; DCHP, direct current heat pump. Reproduced from Hepbasli A, Ekren O, Biyik E. Exergoeconomic evaluation of the IKVA shopping Center in the City of Sopron. Hungary. SBE 16. Smart Metropoles. Istanbul, Turkey; 2016.

COP for heating 3

COP

2 1 0 04:04

04:19

04:33

04:48

Time Fig. 20 Variation of coefficient of performance (COP) values with time in the heating mode (Date: March 25, 2017; All heat pumps (HPs) were in operation).

COP for cooling

EER

COP or EER

20 15 10 5 0 16:04

16:19

16:33

16:48

17:02

17:16

Time Fig. 21 Variation of coefficient of performance (COP) values with time in the cooling mode (Date: June 13, 2017; the direct current heat pump (DCHP) was in operation).

4.4.6.2.3

Results and discussion

Fig. 20 illustrates a variation of COP values with time-based experiments dated March 25, 2017 in the heating mode (at an average ambient temperature of 11.251C), when all HPs were in operation. It is obvious from the figure that the COP values range from about 1.6 to 2.6. Fig. 21 shows a variation of the COP for cooling (or EER) with time-based experiments dated June 13, 2017 in the cooling mode (at an average ambient temperature of 28.251C), when the DCHP was in operation. It is clear from the figure that the COP values for cooling vary between about 3 and 5. Fig. 22 indicates a variation of the water temperatures entering and leaving the HP and the capillary tube radiant cooling/ heating system. It is obvious from the figure that the difference between HP flow and return temperatures is almost 51C, being lower than the values dated March 5, 2017.

120

Heat Pumps

50 48 Temperature (°C)

46 HP water flow temperature

44 42

HP water return temperature

40 38

Water supply temperature

36 34

Water return temperature

32 30 04:04

04:19 04:33 Time

04:48

Fig. 22 Variation of heat pump (HP) water flow/return and water supply/return temperatures in the heating mode (Date: March 25, 2017; All HPs were in operation).

Compressor charactersitic item

70 60 Suction pressure

50 40

Discharge pressure

30

Suction temperature

20

Discharge temperature

10

Power consumption

0 04:04

04:19 04:33 Time

04:48

Fig. 23 Variation of compressor characteristic values with time in the heating mode (Date: March 25, 2017; All heat pumps (HPs) were in operation).

Fig. 23 shows a variation of the compressor characteristic values with time. It is clear from the figure that the HPs were operated at almost 4 bar, and between 14 and 19 bar for compressor suction and discharge pressures, respectively, being very close to the values dated March 5, 2017. Table 5 shows energy and exergy analysis results while Table 6 gives some exergetic factors such as exergetic fuel, exergetic rate, exergy destruction rate, and exergy efficiencies for each component of the HP unit. As far as exergetic assessment is concerned, considering Eq. (20), the highest irreversibility (RI ¼ 76%) occurs in the motorcompressor subassembly. These losses are due to the electrical, mechanical, and isentropic efficiencies and emphasize the need for paying close attention to the selection of this type of equipment because components of inferior performance can considerably reduce the overall performance of the system. The second largest irreversibility is due to the condenser (RI ¼ 9%). This is partly due to the large degree of superheat achieved at the end of the compression process, leading to large temperature differences associated with the initial phase of heat transfer. The third highest irreversibility is in the expansion valve (RI ¼ 8%) due to the pressure drop of the refrigerant passing through it. Besides this, the evaporator has the lowest irreversibility (RI ¼ 7%) on the basis of the HP cycle. The component irreversibility results indicate that the most potential for improvement is probably in the compressor, followed by the condenser and the capillary tube. Since compressor power depends strongly on the inlet and outlet pressures, any HE improvements that reduce the temperature difference will reduce compressor power by bringing the condensing and evaporating temperatures closer together. From a design standpoint, compressor irreversibility can be attacked independently. In recent years, it has been substantially reduced by improving motors, valves, lubrication, etc. The only way to eliminate throttling loss would be to replace the expansion valve with an isentropic turbine (an isentropic expander) and to recover some shaft work from the pressure drop. The findings were in good agreement with Ref. [42]. Considering Fig. 10, the values for Tc ¼531C, TH* ¼ 351C, TH ¼ 351C, TR ¼ 251C, TA ¼111C, and TE ¼7.11C were obtained in the heating mode while those of Tc ¼60.51C, TH* ¼13.081C, TH ¼101C, TR ¼ 251C, TA ¼ 291C, and TE ¼ 16.51C in the cooling mode.

Heat Pumps

Table 5

Energy and exergy analysis results Pressure Specific (P) enthalpy, (h) (kPa) (kJ/kg)

State no.

Description

Fluid

Phase

Temp (T) (1C)

0

–

Refrigerant (R-417a) Water Air Refrigerant

Dead state

12

101.325 385.76

Dead state Dead state Superheated vapor Superheated vapor Compressed liquid Mixture

12 12 6.8

101.325 385.76 101.325 490 375.18

Compressed liquid Compressed liquid Air Air

000 000 1

5

– – Evaporator outlet/ compressor inlet Condenser inlet/ compressor outlet Condenser outlet/ expansion valve inlet Expansion valve outlet/ evaporator inlet Condenser outlet

6

Condenser inlet

Water

7 8

Evaporator inlet Evaporator outlet

Air Air

2,act 3 4

121

Table 6

Refrigerant Refrigerant Refrigerant Water

Specific entropy, (s) (kJ/kg K)

x

Mass flow rate, m_ (kg/s)

Exergy rate, E_ x (kW)

1.7346

–

–

1.7346 1.6329

– – 18.42

– – 0.848

51.2

1490

400.3

1.645

40.09

1.845

36.11

1490

250.9

1.1718

25.62

1.179

490

250.9

1.1848

0.29245 21.92

1.009

35.8

149.98

0.51597

0

111.71

34.661

30.5

127.82

0.44364

0

110.18

34.19

0 0.000441

0 0.00502

2.778

12 11.5

101.325 101.325

Some exergetic values for the HP unit

Device no.

Device

Exergetic fuel rate F_ (kW)

Exergetic product rate P_ (kW )

Exergy destruction rate E x_ dest (kW)

Exergy efficiencye %

Exergetic improvement potential rate I P_ (kW)

RI (%)

I II III IV

Compressor Condenser Expansion valve Evaporator

2.7 0.679 1.179 0.161

0.997 0.476 1.009 0.005

1.703 0.190 0.171 0.156

0.369 0.715 0.855 0.031

1.074 0.054 0.025 0.151

76 9 8 7

Using RE on a HP can be realized with AC or DC compressor, but AC compressor usage requires conversion from DC power to AC by using a power inverter. This leads to increasing initial investment cost and also decreasing energy conversion efficiency. Therefore, RE powered HP using DC compressor enhances the efficiency and reduces the initial investment cost due to the elimination of power conversion from DC to AC. DC compressors present better efficiency than AC compressors. In addition, the DC compressor indicates a higher performance, giving a more stable performance and suffering a lower performance decrease in the range of compressor speeds. Therefore, in comparison to the conventional ACHPs, high efficiency AC/DC inverter HPs and DCHPs offer better compressor performances while they can regulate the compressor speed in order to be adjusted to the demand [41].

4.4.7

Future Directions

Related to the integration and management of HPs in the power system, the following points should be addressed [10]: 1. Focusing on the development of scalable control concepts and knowledge about the flexibility of a HP pool in contrast to single entities for utilization of a large number of HPs in a pool, and 2. Developing the business cases to build the foundation for integrating HPs in smart grids, and a techno-economic analysis of HSPs when operating on different electricity markets such as day-ahead, intraday, and the reserve markets. In terms of the integration and management of HPs in building energy systems, the following points should be addressed [10]: 1. A more clear design of optimal flexible systems for a given application, 2. Investigating the impact of various control approaches and smart grid applications on system cost and efficiency needs, and 3. Use of model predictive control in many studies for operating HPs in a smart grid context. Related to the impact of smart grid use on HP units, the focus should be on Ref. [10] 1. Use of variable speed compressors for enabling a continuous regulation of power consumption,

122

Heat Pumps

2. Investigating how to design the whole HP system for optimal adaption to the requirements from the electric system, and 3. Improving flexibility characteristics and lifetime of a HP unit. Based on a detailed review by Mohanraj et al. [43], the following research needs related to solar assisted compression HPs (SACHPs) have been reported. 1. 2. 3. 4. 5. 6. 7. 8. 9. 10.

Standardization of SACHP systems for drying and water heating applications, Possible use of modified compression cycles for drying and water heating applications, Possible regeneration in solar stills using SACHP systems, Considering HP assisted desalination systems, Use of transparent PV panels in HP assisted solar stills, Scope of SACHP systems for laundry applications, Utilization opportunity of SACHP systems for biomedical applications, Effective harvesting of solar energy using heat storage materials packed in evaporators, Identification of energy efficient and environment-friendly refrigerants for SACHP systems, Performance prediction of SACHP for drying, water heating, and space heating applications through artificial neural network models, and 11. Technoeconomical feasibility assessment of SACHP systems for different applications, and life cycle assessments of SACHP systems for drying, space heating, water heating, and desalination applications. Examining the integration and control of various schemes for multiple uses (cooling, heating, water heating, and power generation) may produce synergic efficiency enhancement [12]. Most of the investigations reported in open literature have used the halogenated group of refrigerants due to its good thermodynamic and thermophysical properties. However, such refrigerants have poor environmental properties due to their high global warming potential. Hence, further research investigations on refrigerants with low global warming potential are essential [44]. The use of high temperature heat pumps (HTHPs) operating with natural fluids indicates a potential environmentally friendly solution for increasing energy efficiency in industrial processes. There is a potential of using a vapor compression cycle and its variations in HTHP for heat delivery above 801C. However, many methods suggested in research still require technological advancement in certain components to be possible. Development of compressors operating at higher temperatures and pressures is also needed [45].

4.4.8

Concluding Remarks

HPs have been used for years for heating and cooling of buildings along with DHW production in developed countries. The number of applications has significantly increased in recent years. However, studies on the performance improvement of HPs are still in progress. The main concluding remarks may be listed as follows [9–13,29,46]: 1. The major part (about 70%) of multitemperature HP applications are found in supermarket refrigeration, household refrigeration, air-conditioning, and for other refrigeration purposes [9]. 2. Optimization and technology integration of HPWHs could increase the COP from a typical range of 1.8–2.5 to higher range of 2.8–5.5 based on current state of research [11]. 3. There is a huge potential for using sewage water and ambient water as the main heat sources for the future HPs due to their long-term stability, proximity to urban areas and not least of all temperatures, mainly for sewage water. These types of heat sources are considered as the main enablers for achieving the capacities estimated in the HRE of 40 GW thermal output [13]. 4. HPs are a suitable source of heat when combined with the DH systems of the fourth generation and in the case of using electricity from renewable sources [29]. 5. Phase change material energy storage in HPs for space cooling production may be used for enhancing the performance of HPs, shifting energy consumption to off-peak electricity periods, and preventing the oversizing of HPs in periods of high cooling loads [34]. 6. Exergy analysis has been widely used as a useful tool for determining the locations, types, and true magnitudes of energy losses in recent years while helping in the design of more efficient HP systems.

Acknowledgments Illustrative Example 1 was established within the framework of the project entitled Design, Construction and Experimental Investigation of a Novel Solar Photovoltaic/Thermal (PV/T)-Assisted Wastewater Heat Pump System (113M532), which was financially supported by the Scientific and Technological Research Council of Turkey (TUBITAK). Illustrative Example 2 was prepared within the framework of the project entitled “ECOSHOPPING-Energy Efficient & Cost Competitive Retrofitting Solutions

Heat Pumps

123

for Shopping Buildings,” cofunded by the European Commission (FP7-2013-NMP-ENV-EeB. grant agreement no: 609180). In this regard, the author would like to thank both the TUBITAK and EC.

References [1] Borre AV. Definition of heat pumps and their use of renewable energy sources. REHVA J 2011;38–9. [2] Chen H, Li D, Dai X. Economic analysis of a waste water resource heat pump air-conditioning system in north China. In: Proceedings of the sixth international conference for enhanced building operations. Shenzen, China; 2006. [3] Hepbasli A, Biyik E, Ekren O, Gunerhan H, Araz M. A key review of wastewater source heat pump (WWSHP) systems. Energy Convers Manag 2014;88:700–22. [4] Sarbu I, Sebarchievici C. General review of ground-source heat pump systems for heating and cooling of buildings. Energy Build 2014;70:441–54. [5] Ni L, Dong J, Yao Y, Shen C, Qv D, Zhang X. A review of heat pump systems for heating and cooling of buildings in China in the last decade. Renew Energy 2015;84:30–45. [6] Gaigalis V, Skema R, Marcinauskas K, Korsakiene I. A review on heat pumps implementation in Lithuaniain compliance with the National Energy Strategy and EU policy. Renew Sustain Energy Rev 2016;53:841–58. [7] Atam E, Helsen L. Ground-coupled heat pumps: part 1 – literature review and research challenges in modelling and optimal control. Renew Sustain Energy Rev 2016;54:1653–67. [8] Atam E, Helsen L. Ground-coupled heat pumps: part 2 – literature review and research challenges in optimal design. Renew Sustain Energy Rev 2016;54:1668–84. [9] Arpagaus C, Bless F, Schiffmann J, Bertsch SS. Multi-temperature heat pumps: a literature review. Inter J Refrig 2016;70:342–57. [10] Fischer D, Madani H. On heat pumps in smart grids: a review. Renew Sustain Energy Rev 2017;70:342–57. [11] Willem H, Lin Y, Lekov A. Review of energy efficiency and system performance of residential heat pump water heaters. Energy Build 2017;143:191–201. [12] Leonzio G. Solar systems integrated with absorption heat pumps and thermal energy storages: state of art. Renew Sustain Energy Rev 2017;70:492–505. [13] David A, Vad Mathiesen B, Averfalk H, Werner S, Lund H. Heat roadmap Europe: large-scale electric heat pumps in district heating systems. Energies 2017;10:578. [14] Thomas N, Pascal N. European heat pump market and statistics report 2015. Technical report. The European Heat Pump Association AISBL (EHPA), Brussels; 2015. [15] Zogg M. History of heat pumps-Swiss contributions and international milestones. In: 9th international IEA heat pump conference, 20–22 May, Zürich, Switzerland; 2008. [16] Zogg M., 2008. History of heat pumps-Swiss contributions and international milestones. Department of Environment. Transport. Energy and Communications DETEC. Swiss Federal Office of Energy SFOE. Section Energy Efficiency and Renewable Energies Process and Energy Engineering. CH-3414 Oberburg. Switzerland. 114 pages. Available from: http://zogg-engineering.ch/publi/HistoryHP.pdf [accessed 15.11.17]. [17] Hepbasli A, Kalinci Y. A review of heat pump water heating systems. Renew Sustain Energy Rev, 2009;13(6–7):1211–29. [18] Kiss B, Neij L, Jakob M. Heat pumps: a comparative assessment of innovation and diffusion policies in Sweden and Switzerland. Historical case studies of energy technology innovation. Chapter 24 In: Grubler A, Aguayo F, Gallagher KS, et al. The global energy assessment. Cambridge: Cambridge University Press; 2012. [19] Bundesamt für Wirtschaft und Ausfuhrkontrolle. Erneuerbare Energien – Wärmepumpen mit Prüfnachweis,technical report, Bundesamt für Wirtschaft und Ausfuhrkontrolle, Eschborn; 2016. [20] BSRIA. Growth in the World Heat Pump Market-August 2014. Available from: https://www.bsria.co.uk/news/article/growth-in-the-world-heat-pump-market/; 2017 [accessed 01.08.17]. [21] Lund JW, Boyd TL. Direct utilization of geothermal energy 2015 worldwide review. In: Proceedings world geothermal congress 2015. Melbourne, Australia. April 19–25, 2015; 2015. [22] FINN Geothermal. The history of heat pump technology. Available from: https://finn-geotherm.co.uk/the-history-of-heat-pumps/; 2017 [accessed 01.08.17]. [23] UMR. History of heat pumps. Available from: https://umrgeothermal.com/geothermal-101/how-it-works/heat-pumps/history-of-heat-pumps/; 2017 (accessed 01.08.17). [24] Hepbasli A, Erbay Z Z, Icier F, Colak N, Hancioglu E. A review of gas engine driven heat pumps (GEHPs) for residential and industrial applications. Renew Sustain Energy Rev 2009;13:85–99. [25] Refrigerator Troubleshooting Diagram. Classification of heat pumps. Available from: http://www.refrigeratordiagrams.com/refrigeration-systems-and-applications/refrigeratorheat-pumps/classification-heat-pumps.html; 2017 [accessed 01.08.17]. [26] Nellissen P, Wolf S. Heat pumps in non-domestic applications in Europe: potential for an energy revolution. In: Presentation given at the 8th EHPA European Heat Pump Forum, 29.5.2015. Brussels. Belgium. Available from: http://www.ehpa.org/fileadmin/red/_EHPA_Archive_Forum/8th_Heat_Pump_Forum_2015/Presentations/ Philippe_Nellissen_IHP_potential_EU.pdf; 2017 [accessed 01.08.17]. [27] Nowak T. Heat pumps. Vision vs. reality. Available from: https://www.ee-ip.org/articles/detailed-article/?article=54&cHash=bd3568515e538ad350faf063915c6141; 2017 [accessed 01.08.17]. [28] Hepbasli A. Exergetic modeling and assessment of solar assisted domestic hot water tank integrated ground-source heat pump systems for residences. Energy Build 2007;39(12):1211–7. [29] Poredoš P, Vidrih B, Duh T, Kitanovski A, Poredoš A. Eligibility of a heat pump based on the primary energy factor. In: 12th IEA heat pump conference. Rotterdam. 17 May 2017. Available from: http://hpc2017.org/wp-content/uploads/2017/06/o252.pdf; 2017 [accessed 01.08.17]. [30] Van Gool W. Energy policy: fairly tales and factualities. In: Soares ODD, Martins da Cruz A, Costa Pereira G, Soares IMRT, Reis AJPS, editors. Innovation and technology-strategies and policies. Dordrecht: Kluwer; 1997. p. 93–105. [31] Erbay Z, Hepbasli A. Assessment of cost sources and improvement potentials of a ground-source heat pump food drying system through advanced exergoeconomic analysis method. Energy 2017;127:502–15. [32] Tsatsaronis G, Morosuk T. Advanced exergetic analysis of a novel system for generating electricity and vaporizing liquefied natural gas. Energy 2010;35:820–9. [33] Gasser L, Wellig B, Hilfiker K WEXA: exergy analysis for increasing the efficiency of air/water heat pumps. Final report. Forschungsprogramm UAW. Umgebungswärme. WKK. Kälte. Im Auftrag des Bundesamtes für Energie; 2008. 141 pages. [34] Culha O, Gunerhan H, Biyik E, Ekren O, Hepbasli A. Heat exchanger applications in wastewater source heat pumps for buildings: a key review. Energy Build 2015;104:215–32. [35] Seybold C, Brunk MF. In-house waste water heat recovery. REHVA J 2013 18–21. Available from: http://www.rehva.eu/fileadmin/REHVA_Journal/REHVA_Journal_2013/ RJ_issue_6/P.18/18-21_Seybold_RJ1306.pdf; 2017 [accessed 01.08.17]. [36] Araz M, Ekren O, Biyik E, Gunerhan H, Hepbasli A Experimental exergetic performance evaluation of a wastewater source heat pump system (WWSHP). In: Heiselberg PK. editor. CLIMA 2016 – proceedings of the 12th REHVA World congress, volume 3. Aalborg: Aalborg University. Department of Civil Engineering. 22–25 May. Aalborg. Denmark; 2016. [37] Lewry A. ECOSHOPPING: energy efficient & cost competitive retrofitting solutions for retail buildings. In: CIBSE technical symposium. Available from: http://www.cibse. org/getmedia/196612be-03f5-4da5-806f-109f85bff732/118-Lewry-Slides.pdf.aspx; 2015. [38] ECOSHOPPING. Available from: http://www.ecoshopping-project.eu/; 2017 [accessed 20.02.16]. [39] ECOSHOPPING. Newsletter. Nr. 1. September 2014. Available from: http://ecoshopping-project.eu/document/EcoShopping_Newsletter_September%202014.pdf; 2017 [accessed 20.02.16]. 4 pages.

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[40] [41] [42] [43]

Hepbasli A, Ekren O, Biyik E. Exergoeconomic evaluation of the IKVA shopping Center in the City of Sopron. Hungary. SBE 16. Smart Metropoles. Istanbul, Turkey; 2016. ECOSHOPPING Reports. Deliverable 4.3 and 4.4. 2015-03-03 and 2017-09-30; 2015. Hepbasli A, Akdemir O. Energy and exergy analysis of a ground source (geothermal) heat pump system. Energy Convers Manag 2004;45(5):737–53. Mohanraj M, Belyayev Y, Jayaraj S, Kaltayev A. Research and developments on solar assisted compression heat pump systems – a comprehensive review (Part-B: applications). Renew Sustain Energy Rev 2017 Available from: https://doi.org/10.1016/j.rser.2017.08.086. [44] Mohanraj M, Belyayev Y, Jayaraj S, Kaltayev A. Research and developments on solar assisted compression heat pump systems – a comprehensive review (Part A: modeling and modifications). Renew Sustain Energy Rev 2017 Available from: https://doi.org/10.1016/j.rser.2017.08.022. [45] Bamigbetan O, Eikevik TM, Nekså P, Bantle M. Review of vapour compression heat pumps for high temperature heating using natural working fluids. Int J Refrig 2017;80:197–211. [46] Pardinas AA, Alonso MJ, Diz R, Kvalsvik KH, Fernández-Seara J. State-of-the-art for the use of phase-change materials in tanks coupled with heat pumps. Energy Build 2017;140(28-41):28–41.

Further Reading Arpagaus C, Bless F, Schiffmann J, Bertsch SS. Multi-temperature heat pumps: a literature review. Int J Refrig 2016;70:342–57. Atam E, Helsen L. Ground-coupled heat pumps: part 1 – literature review and research challenges in modeling and optimal control. Renew Sustain Energy Rev 2016;54:1653–67. Atam E, Helsen L. Ground-coupled heat pumps: part 2 – literature review and research challenges in optimal design. Renew Sustain Energy Rev 2016;54:1668–84. Bamigbetan O, Eikevik TM, Nekså P, Bantle M. Review of vapor compression heat pumps for high temperature heating using natural working fluids. Int J Refrig 2017;80:197–211. Cabeza LF, Sol A, Barreneche C. Review on sorption materials and technologies for heat pumps and thermal energy storage. Renew Energy 2017;110:3–39. Cantor J, Gavin DJ. Heat pumps for the home. Ramsbury, Marlborough: The Crowood Press Ltd; 2013. Fischer D, Madani H. On heat pumps in smart grids: a review. Renew Sustain Energy Rev 2017;70:342–57. GasTerra. Gas heat pumps – efficient heating and cooling with natural gas. Groningen: Castel International Publishers; 2010. Grassi W. Heat pumps: fundamentals and applications. Cham: Springer International Publishing AG; 2017. Mohanraj M, Belyayev Y, Jayaraj S, Kaltayev A. Research and developments on solar assisted compression heat pump systems – a comprehensive review (Part A: modeling and modifications). Renew Sustain Energy Rev 2017 Available from: https://doi.org/10.1016/j.rser.2017.08.022. Mohanraj M, Belyayev Y, Jayaraj S, Kaltayev A. Research and developments on solar assisted compression heat pump systems – a comprehensive review (Part-B: applications). Renew Sustain Energy Rev 2017 Available from: https://doi.org/10.1016/j.rser.2017.08.086. Reda F. Solar assisted ground source heat pump solutions: effective energy flows climate management. Cham: Springer International Publishing AG; 2016. Rees SJ. Advances in ground-source heat pump systems. Duxford: Woodhead Publishing; 2016. Sarbu I, Sebarchievici C. Ground-source heat pumps: fundamentals, experiments and applications. London: Academic Press; 2015.

Relevant Websites https://energy.gov/energysaver/heat-pump-systems. Energy.GOV. http://www.egshpa.com/. European Ground Source Heat Pump Association (EGSHPA). http://www.ehpa.org/. European Heat Pump Association (EHPA). http://www.geo-energy.org/gea_heat_pumps.aspx. Geothermal Energy Association (GEA). http://www.gshp.org.uk/. Ground Source Heat Pump Association. http://www.heatpumps.org.uk/. Heat Pump Association (HPA). www.heatpumpingtechnologies.org. International Energy Agency (IEA)-Heat Pump Technologies (HPT). https://igshpa.org/. International Ground Source Heat Pump Association (IGSHPA).

4.5 Heat Engines Ibrahim Dincer and Ahmed Hasan, University of Ontario Institute of Technology, Oshawa, ON, Canada r 2018 Elsevier Inc. All rights reserved.

4.5.1 Introduction 4.5.2 Classification 4.5.3 Carnot Cycle 4.5.3.1 Reversed Carnot Cycle 4.5.4 Vapor Cycles 4.5.4.1 Simple Ideal Rankine Cycle 4.5.4.2 Actual Rankine Cycle 4.5.4.3 Analysis of the Rankine Cycle 4.5.4.3.1 Pumps 4.5.4.3.2 Heat exchangers 4.5.4.3.3 Power plant efficiencies 4.5.4.3.4 Efficiencies of vapor power cycles 4.5.5 Case Study: Combined Cycle Fueled With Natural Gas 4.5.5.1 Systems Description 4.5.5.1.1 Energy and exergy analyses 4.5.6 Trilateral Cycle Assessment 4.5.7 Gas Cycles 4.5.7.1 Totally Reversible Gas Power Cycles 4.5.7.2 Otto and Diesel Cycle 4.5.7.3 Diesel Cycle 4.5.7.4 Gas Turbine (Brayton) Power Cycles 4.5.7.4.1 Air-standard Brayton cycle 4.5.7.4.2 Regenerative Brayton cycle 4.5.7.4.3 Reheat regenerative Brayton cycle 4.5.7.4.4 Brayton cycle with intercooler 4.5.7.5 Exergy Destructions in Brayton Cycle Power Plants 4.5.7.5.1 Efficiencies of gas power plants 4.5.7.5.2 Efficiencies of cogeneration plants 4.5.8 Future Directions 4.5.8.1 Pulse Detonation Engine 4.5.8.2 Wave Disk Engine 4.5.8.3 Single Atom Heat Engine 4.5.8.4 Using Renewable Fuels and Renewable Energy Sources 4.5.9 Conclusions References Further Reading Relevant Websites

Nomenclature cp cv COP E EER ex Ex Exdestroyed _ Ex g h hfg

Specific heat at constant pressure (kJ/kg
K) Specific heat at constant volume (kJ/kg
K) Coefficient of performance Energy (kJ) Energy efficiency ratio Specific exergy (kJ/kg) Amount of exergy (kJ) Exergy destruction (kJ) Rate of exergy (kW) Gravitational acceleration (m/s2) Enthalpy (kJ/kg) Enthalpy of vaporization (kJ/kg)

Comprehensive Energy Systems, Volume 4

HV HHV I k KE LHV m _ m n P PE PER q

doi:10.1016/B978-0-12-809597-3.00405-3

126 126 127 128 128 129 130 131 132 135 136 136 137 137 139 143 144 144 146 148 148 148 153 154 155 156 163 165 166 166 167 167 167 167 167 168 168

Heating value (kJ/kg) Higher heating value (kJ/kg) Current (amp) Specific heat ratio Kinetic energy (kJ) Lower heating value (kJ/kg) Mass (kg) Mass flow rate (kg/s) Polytropic constant Pressure (kPa) Potential energy (kJ) Primary energy ratio Specific heat transfer (kJ/kg)

125

126

Heat Engines

W _ W z

Temperature (K or 1C) Specific internal energy (kJ/kg) Total internal energy (kJ) Specific volume (m3/kg) Velocity (m/s) Voltage (V) Volume (m3) Volume flow rate (m3/s) Amount of work (kJ) Rate of work or power (kW) Elevation (m)

Greek letters Zex Exergy efficiency

Zen Zth

Energy efficiency Thermal efficiency

Subscripts 0 amb Comp Cond Comb Cogen CV e elect Evap Exp Valve H HP HTr in

isen L mech out ov P R regen rev s surr th Tr w WTr

Isentropic Low-temperature Mechanical Outlet Overall Pump Refrigerator Regenerator Reversible Source, isentropic Surroundings Thermal Turbine Water Water turbine

Q _ Q r rc rp R s S Sgen SEER t

4.5.1

Amount of heat transfer (kJ) Rate of heat transfer (kW) Compression ratio Cutoff ratio Pressure ratio Gas constant (kJ/kg
K) Specific entropy (kJ/kg
K) Total entropy (kJ/K) Entropy generation (kJ/K) Seasonal energy efficiency ratio Time (s)

Dead (environmental) state Ambient Compressor Condenser Combustion Cogeneration Control volume Electricity Electricity Evaporator Expansion valve High-temperature Heat pump Hydraulic turbine Inlet

T u U v V V V V_

Introduction

Transforming energy from one form to another takes its natural course, with no external forces needed. The first law of thermodynamics states that energy can be transformed from one form to another but can neither be created nor destroyed. However, analyzing a process within a closed or open system is not accurate in some cases, specifically in the cases where energy is being converted to work. For example, a ball could be dropped from an arbitrary height, as the ball is falling its gravitational potential energy is being transformed into kinetic energy. However, the reverse process could not naturally occur, as work needs to be done to lift the ball back to the original height. In accordance with the first law of thermodynamics, this processes energy is conserved therefore valid. This is why the second law of thermodynamics is needed to remove such discrepancies and false assumptions while correcting and making a more accurate insight from a thermodynamic view. Some systems allow for a reverse process such as a ball going back up the height it was dropped using a system that requires an input of energy. For example, supplying electricity to a motor that’s connected to a pulley system, which would reel the ball back to the original height. However, in the field of thermodynamics, most of the power generation systems are heat engines. Heat engines are systems that are designed for the efficient conversion of thermal or chemical energy to mechanical energy or electric energy. The Rankine and Brayton cycles, or both combined, are considered to be the main power generation cycles used in industry. The Rankine cycle is a steam-driven cycle, and the Brayton cycle is a gas-driven cycle. Both cycles convert thermal energy to mechanical energy and electricity. In the following chapter various internal and external combustion engines and power generation cycles are discussed.

4.5.2

Classification

The Rankine and Brayton cycles, or both combined, are considered to be the main power generation cycles used in industry. The Rankine cycle is a steam-driven cycle, and the Brayton cycle is a gas-driven cycle. Both cycles convert thermal energy to mechanical

Heat Engines

127

energy and electricity. The Carnot cycle results in the highest efficiency of any cycle of heat engines. In the following chapter various internal and external combustion engines and power generation cycles are discussed (Fig. 1).

Heat engines

Internal combustion

Rotary

Open Brayton cycle

External combustion

Reciprocating

Wankel engine

Otto cycle

Deisel cycle

Reciprocating

Steam engine

Rotary

Stirling engine

Steam turbine

Closed Brayton cycle

Fig. 1 Classification of various heat engines.

4.5.3

Carnot Cycle

The Carnot cycle results in the highest efficiency of any cycle of heat engines. This comes as the Carnot cycle neglects all irreversibilities such as friction, heat losses, etc. This makes the Carnot cycle a completely reversible cycle that consists of the following process: 1. Reversible isothermal expansion in which the gas absorbs thermal energy qin and expands at a constant temperature Th. As the gas expands, work is done on its surroundings, as the cylinder head is moving. Theoretically as the air expands, the gas should experience a temperature drop. However, a heat source is provided to maintain the gas’s isothermal state. 2. Reversible adiabatic expansion in which the gas expands and experiences a temperature drop from Th to Tl. The gas continues to expand. However, the thermal heat source is not provided anymore but is replaced by insulation. The piston in this process is assumed frictionless. Furthermore, the process is assumed to be quasi equilibrium. 3. Reversible isothermal compression, the piston is going back to its original position as the external forces from the surroundings are doing work on the gas. The insulation in this process is removed and a heat sink at a temperature of Tl is provided to the cylinder’s head. The heat sink ensures the gas’s isothermal state, as it is compressed. 4. Reversible adiabatic compression in this process insulation replaces the heat sink, and as the piston continues to compress the gas, the temperature of the gas rises to its original temperature of Th. Fig. 2 demonstrates the Carnot cycle using an ideal piston. The compression and expansion of the piston as previously mentioned is isothermal in process 1-2 and 3-4 and adiabatic in the remaining process. 1−2

qin Fig. 2 Carnot cycle demonstration using an ideal piston.

2−3

3−4

qout

4−1

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Heat Engines

A demonstrative P–V diagram for the Carnot cycle is shown below in Fig. 3. The temperatures at states 1 and 2 are assumed to be equal to Th and the temperatures at states 3 and 4 are assumed to be equal to Tl. The area inside the P–V diagram represents the

qout

1

Isothermal Adiabatic

2

P Wnet,out

4

Adiabatic 3

Isothermal qin v Fig. 3 Demonstration of the P–V diagram of the Carnot cycle.

network output of the system along with the quasi-static equilibrium boundary. During the processes, 1-2-3 work is being done by the gas onto its surroundings as the gas is expanding as seen in Fig. 2. As for process 3-4-1 work is being done on the gas as the gas is being compressed and the correlating decreasing volume can be seen in Fig. 3. The efficiency of the Carnot cycle, as mentioned previously, is the highest of any other cycle, as it’s assumed to consist of fully reversible processes. The thermal efficiency of a Carnot heat engine can then be defined as the heat transferred to the engine from the heat source Qin and the heat rejected by the system to the heat sink.

4.5.3.1

Reversed Carnot Cycle

As previously mentioned in this chapter the Carnot cycle is completely reversible, allowing all the process that it experiences to be reversed. If the processes are reversed the cycle becomes the Carnot refrigeration cycle. This is done by reversing when the heat is rejected and absorbed by the cycle. The P–V diagram for the Carnot refrigeration cycle is shown in Fig. 4; it is very similar to the Carnot cycle’s P–V diagram. However, the diagram has reversed heat rejection and absorption.

qout

1

Isothermal Adiabatic

2

P 4

Wnet,out

Adiabatic 3

Isothermal qin v Fig. 4 P–V diagram for the Carnot cycle.

4.5.4

Vapor Cycles

Vapor cycles have existed since the 1800s and revolutionized both the power production industry and the locomotive industry. Since then many developments have been created, especially creations such as the steam turbine (Dincer, Calin), which now provides 80% of the world’s electrical power. Vapor-based cycles are a crucial part of the power generation sector in the world, as most power plants are driven using the steam Rankine cycle. There are derivatives of the Rankine cycle such as the organic Rankine

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129

cycle that usually operate at lower temperatures. In most instances, the organic Rankine cycle is paired with a renewable source of energy such as solar, geothermal, etc. as it could operate at lower temperatures. Furthermore, some organic cycles are paired with the Brayton or Rankine cycle when there's enough excess heat to drive the cycle.

4.5.4.1

Simple Ideal Rankine Cycle

The vapor cycle’s most basic power cycle is the simple ideal Rankine cycle (Fig. 5). The ideal Rankine cycle is considered internally reversible. The components that compromise the ideal Rankine cycle include the following: a pump that works isentropically and

Steam turbine/ generator Wet

4

Steam generator

1 Condenser Pump 3

2

QCond Fig. 5 Simple Rankine cycle with all its components.

adiabatically, a vapor generator that operates with no pressure drops, a condenser with no pressure drops with ideal heat transfer properties, and a turbine that operates isentropically and adiabatically. The vapor generator has the working fluid go in, in saturated liquid state and it exits the vapor generator as a superheated gas. The vapor generator does the superheating process over three basic processes, preheating, boiling, and vapor superheating. The T–s diagram is shown in Fig. 6 for the ideal Rankine cycle. The process starts with a pump that works isentropically and adiabatically, which pressurizes a saturated liquid, and a vapor generator that operates with no pressure drops that changes the state of the saturated liquid to superheated vapor. Then the superheated gas enters a turbine that operates isentropically and

T

qi

n

3

Wturb,out

2 wpump 1

4

S Fig. 6 T–s diagram for the ideal simple Rankine cycle.

130

Heat Engines

adiabatically where the superheated vapor expands and turns into a mixture of liquid and gas, and finally the mixture enters the condenser with no pressure drops, with ideal heat transfer properties and a turbine that operates isentropically and adiabatically.

4.5.4.2

Actual Rankine Cycle

The ideal Rankine cycle is assumed internally reversible. However when it comes to the actual cycle all the irreversibilities is considered. The components that compromise the ideal Rankine cycle are the same in the actual cycle. The pumps are no longer isentropic and require a larger amount of work input to accommodate the pressure losses in other components. The vapor generator operates with a pressure drop as there are losses due to fluid friction and energy losses through the connecting pipe system to the turbine. The condenser experiences a slight pressure drop with the heat rejection happening slower than in the ideal cycle, and the turbine can no longer be assumed to operate isentropically and adiabatically. As for the vapor generator, the heat input needed should be higher to compensate for the heat losses inside the system. There are other energy losses that occur due to heat transfer from the working fluid to its surroundings. The T–s diagram is shown in Fig. 7 for the actual Rankine cycle. The pump needs more power input than the ideal cycle to compensate for the losses. As seen in the T–s diagram the boiler heats up the working fluid more than in the ideal cycle, which is to

T

qi

n

3

Wturb,out

2 wpump 4

1

S Fig. 7 T–s diagram for the actual Rankine cycle.

compensate for heat losses of the working fluid to its surroundings. Unlike in the ideal Rankine cycle, the boiler does experience pressure drops in the actual cycle due to frictional losses and heat losses. The working fluid has to enter the turbine as a superheated vapor since the turbine might corrode as moisture enters it (Fig. 8).

Inlet Steam turbine/ generator

Wt

Outlet Fig. 8 Steam turbine.

Heat Engines 4.5.4.3

131

Analysis of the Rankine Cycle

The simple Rankine cycle component mass, energy, entropy, and exergy balance equations are shown in Table 1. Table 1 is inclusive of all irreversibilities so it could be used as a simple actual Rankine cycle. The mass and energy balance equations for each of the Table 1

Balance equations

Component

Balance equations

Pump

_2 MBE: m_ 1 ¼ m EBE: m_ 1 h1 þ Wpump ¼ m_ 2 h2 _ 1 s1 þ S_ gen ¼ m _ 2 s2 ENBE: m _ Q þ Ex _ D m_ 1 ex1 þ Wpump ¼ m_ 2 ex2 þ Ex

Boiler

_3 MBE: m_ 2 ¼ m _ 2 h2 EBE: m_ 2 h2 þ Q__ boiler ¼ m _ _ 2 s2 þ Q boiler _ SBE: m Ts þ S gen ¼ m 3 s3 _ Q ¼m _ D _ 3 ex3 þ Ex EBE: m_ 2 ex2 þ Ex

Turbine

_4 MBE: m_ 3 ¼ m EBE: m_ 3 h3 ¼ m_ 4 h4 þ Wturbine þ_ Q_ Loss _ 2 s2 þ QTLoss _ 1 s1 þ S_ gen ¼ m ENBE: m _ Q þ Ex _ 0D þ Wturbine EBE: m_ 1 ex1 ¼ m_ 2 ex2 þ Ex

Condenser

_1 MBE: m_ 4 ¼ m EBE: m_ 4 h4 ¼ m_ 1 h1 þ Q_ cond _ _ 4 s4 þ S_ gen ¼ m _ 1 s1 þ QTcond ENBE: m Q _ _ 0D EBE: m_ 4 ex4 ¼ m_ 1 ex1 þ Ex þ Ex

components and the overall balances for the ideal cycle are given in Table 1. For the case of the ideal Rankine cycle, there are no exergy destructions in condenser, pump, or turbine. However, provided that the heat source is at a constant temperature, the externally irreversible Rankine cycle shows exergy destructions in the vapor generator that can be estimated based on the temperature difference between heat source and working fluid. Let us denote the temperature of the heat source Tso. When the ideal Rankine cycle is analyzed, the heat source can be considered at a constant temperature that is equal to the highest temperature of the working fluid. The heat source Tso provides heat input to the processes of preheating, boiling, and superheating. In other words, the vapor generator is assumed with an “infinite surface for heat transfer,” which ensures that the working fluid reaches the heat source temperature. Turbines are recognized as work producing devices that are commonly employed in many power generating systems and applications, ranging from steam Rankine cycles, which use steam turbines, to hydropower plants, which use a hydroturbine. There are also closed or open types of air-standard Brayton cycles where gas turbines are utilized. Furthermore, there are some types of organic Rankine cycles where expanders are utilized as work producing devices. The type of work produced by these turbines and expanders is essentially shaft work, unlike the type of boundary movement work produced by reciprocating type units. As the fluid is flowing through the turbine, the high pressure and temperature fluid passes around the blades of the turbine. The blades of the turbine have a cross-sectional profile similar to that of the airplanes (aerofoil profile). The high pressure and temperature fluid will flow around the blades, which will produce different pressures between the top and the bottom of the blade, which will have a net force acting in the direction from high pressure side to the low pressure side. The net force on each blade will rotate the hub of the blades and that is how mechanical work is produced. The energy that will rotate the hub of the blades is extracted from the high pressure and temperature fluid, which will result in reduction of the pressure and temperature of the flowing fluid. The turbines are considered open thermodynamic systems, since the mass can enter and leave the control volume of the turbines. Often turbines are treated as steady flow devices, and they are analyzed based on that. The performance of the turbines is measured through both energy and exergy efficiencies, which is illustrated in the following example. A schematic diagram of a steam turbine is shown in Fig. 8. Example 1: Consider an adiabatic steam turbine, as shown in Fig. 9, with the following inlet and exit states: P1 ¼ 12,000 kPa, T1 ¼ 6251C, P2 ¼ 10 kPa, x2 ¼ 0.95. Taking the dead-state temperature of steam as saturated liquid at 251C, determine isentropic efficiency and exergy efficiency of the steam turbine (Fig. 10). Solution: The balance equations are written as follows: _2 _1 ¼m MBE: m _ out _ 1 h1 ¼ m _ 2 h2 þ W EBE: m _ 1 s1 þ S_ gen ¼ m _ 2 s2 EnBE: m _ out þ Ex _ d _ 1 ex 1 ¼ m _ 2 ex 2 þ W ExBE: m

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T1 = 625 °C P1 = 12 MPa

Wt

X2 = 0.95 P2 = 10 kPa Fig. 9 Adiabatic steam turbine with the following inlet and exit states: P1 ¼12,000 kPa, T1 ¼6251C, P2 ¼10 kPa, x2 ¼0.95.

Discharge Pump

Win

In flow   Fig. 10 A schematic diagram: a pump, where the supplied liquid (in flow) is pressurized and discharged (discharge) using the supplied shaft work W_ in .

Using the EBE to calculate the specific power produced by the steam turbine as follows: h1 ¼ h2 þ wout wout ¼ 1206 kJ=kg Then using the ExBE the exergy destructed through the expansion process is calculated as follows: ex1 ¼ ex2 þ wout þ exd exd ¼ 268:2 kJ=kg Then by using the exergy efficiency definitions these are the resulting efficiencies: Zisen;Tr ¼ 0:708 ¼ 70:8% wout Zex;Tr ¼ ¼ 0:742 ¼ 74:2% ex in

4.5.4.3.1

Pumps

Pumps are considered mechanical devices that receive shaft work to operate and transfer the energy that they received in the form of work to the fluid as a mechanical energy. Since the mass can enter and leave their control volume of the pumps, the pumps are then considered and analyzed as open thermodynamic systems. The analysis of such a device based on the first and the second law of thermodynamics is shown in the following example. The pump revives the shaft work and then with the help of the mechanical design of the pump, the supplied liquid is pumped to a higher pressure using the supplied shaft work as shown in Fig. 11. Example 2: The pressure of water at an ambient condition is increased to a high pressure of 900 kPa at the exit of a pump. The temperature of the water entering the pump is 151C; the water enters the pump through an opening of 1 cm in diameter and the exit is 1.5 cm in diameter. If the mass flow rate of the water is 0.5 kg/s, calculate the velocity of the water at the inlet and the exit of the pump (Fig. 12). Solution: It is first necessary to write four balance equations as follows: _ out ¼ m _ MBE: min ¼ m


2 2 _ in ¼ m _ in hin þ m _ in Vin _ out hout þ m _ out Vout EBE: m =2000 þ W =2000

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133

exd=78.0

exheat =57.9

qin =75.0

qLHV=100

exfuel =103.6

q lost,boiler =25.0

wnet,out= 25.6

exsteam=33.2 Fig. 11 Combined energy and exergy diagram for the steam power plant considered.

P2 = 900 kPa Pump

Win P1 = 101 kPa T1 = 15°C Mass flow rate = 0.5 kg s−1

Fig. 12 A pump that increases the pressure of water at an ambient condition is increased to a high pressure of 900 kPa at the exit of a pump.

_ in sin þ S_ gen ¼ m _ out sout EnBE: m



2 2 _ d _ in ¼ m _ out exout þ m _ out Vout _ _ in Vin ExBE: min ex in þ m =2000 þ Ex =2000 þ W

From the mass balance equation, the velocity is calculated at the inlet when the specific volume is equal to 0.001001 m3/kg as follows: _  Vin ¼ m

vin 0:001001 ¼ 0:5  ¼ 6:37 m=s Ain p  0:012 C 4 Ain Vin rin ¼ Aout Vout rout

However since water is considered as an incompressible substance then the previous equation reduces to the following: Ain Vin ¼ Aout Vout Vout ¼

    Ain Vin Din 2 0:01 2 ¼ 6:37  ¼ Vin ¼ 2:83 m=s Aout Dout 0:015

Example 3: A numerical example is used to illustrate and contrast the various efficiencies defined in this section. We consider a simple steam power plant with a net power output of 10 MW and boiler and condenser pressures of 10,000 kPa and 10 kPa, respectively (Fig. 35). We assume a turbine inlet temperature of 5001C and isentropic efficiencies of 85% for both the turbine and the pump. In addition, we assume that the furnace–boiler system has an efficiency of 75%. That is, 75% of the lower heating value of the fuel is transferred to the steam flowing through the boiler while the remaining 25% is lost, mostly with the hot exhaust gases passing through the chimney. The source and sink temperatures are taken as 1300K and 298K, respectively. We consider methane as the fuel with a lower heating value of 50,050 kJ/kg and a chemical exergy of 51,840 kJ/kg. For the given values and assumptions, an analysis of this cycle yields wnet;out ¼ 1081 kJ=kg; qin ¼ 3172 kJ=kg; exin

1

¼ 1400 kJ=kg; exin

2

¼ 2444 kJ=kg

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Heat Engines

as well as the following efficiency values: Zen

1

¼ 34:1%; Zen

2

¼ 25:6%; Zex

1

¼ 77:2%; Zex

2

¼ 44:2%; Zex

3

¼ 24:7%:

When the energy and exergy losses in the furnace–boiler system are not considered, the thermal efficiency is 34.1% while the corresponding exergy efficiency is much higher (77.2%). However, when the losses in the furnace–boiler are considered, the exergy efficiency (24.7%) is lower than the thermal efficiency (25.6%). When teaching undergraduate thermodynamics, it is normally stated that the exergy efficiency is greater than the thermal efficiency for heat engines, referring to the first approach here. This point is made by emphasizing that thermal efficiency is the fraction of heat input that is converted to work while exergy efficiency is the fraction of the work potential of the heat (this work potential, i.e., exergy, is smaller than heat) that is converted to work. However, when one considers the effect of furnace–boiler losses, and uses the chemical exergy of the fuel in the exergy efficiency and the heating value of the fuel in the thermal efficiency, the exergy efficiency becomes smaller than the thermal efficiency. In thermodynamics, it is often misleading to make generalized statements as they may not always apply. For example, can we state that the exergy efficiency, based on Eq. (9) (Zex 3), is always lower than the thermal efficiency as defined by the Eq. (1) (Zth 2)? The answer will be yes only if the chemical exergy of the fuel is always greater than its heating value. According to data in Ref. [1], this is the case for methane but not for hydrogen (qLHV ¼ 119,950 kJ/kg, exfuel ¼117,120 kJ/kg). For a reversible heat engine cycle operating between a source at Ts and a sink at T0, the thermal efficiency is given by Zth;rev ¼ 1

T0 Ts

ð1Þ

The ratio of the actual thermal efficiency to the thermal efficiency of a reversible heat engine operating between the same temperature limits gives a type of exergy efficiency of the heat engine. For a furnace temperature of Ts ¼ 1300K and an environment temperature of T0 ¼298K, the reversible thermal efficiency found to be 77.1%. Dividing the actual thermal efficiency of 34.1% by this efficiency (0.341/0.771) gives 44.2%. Note that this is the same as the exergy efficiency obtained. The results of the numerical example considered in this section are shown in a combined energy and exergy diagram in Fig. 13. In many studies with energy and exergy analyses of power cycles, energy and exergy flow diagrams are given separately. The combined flow diagram approach used here appears to be useful in conveying energy and exergy results of the cycle in a scaled, compact, and comprehensive manner. The heating value of the fuel is normalized to 100 units of energy and other values are

P1 = 100 kPa T1 = 30 °C

P1 = 100 kPa T1 = 80 °C

Fig. 13 A schematic diagram of the heat exchange process presented in Example 4.

normalized accordingly. The thermal and exergy efficiencies discussed in this section can easily be obtained using the values in this diagram by taking the ratios of various terms. The total exergy destruction in this power plant is 78 kJ for a total exergy input of 103.6 kJ. The exergy destruction in the cycle based on an exergy input of 33.2 kJ is only 7.6 kJ (33.2 25.6), which is only 9.7% of total exergy destruction. That is, the exergy destructions in the furnace–boiler system account for the remaining 90.3% of the total exergy destruction. This significant exergy destruction is not considered in an exergy efficiency definition neglecting the destructions in the furnace–boiler system (see Eq. 5). One may question the value of exergy analysis as a tool for assessing a power plant because the thermal efficiency based on the heating value of the fuel and the exergy efficiency based on the exergy of the fuel are very close. Although the exergy efficiency in this case adds little new information for addressing cycle efficiency, we have to remember that a major use of exergy analysis is to analyze the system components separately and to identify and quantify the sites of exergy destruction. This information can then be used to improve the performance of the system by trying to minimize the exergy destructions in a prioritized manner. Note that the exergy efficiency defined earlier addresses the fact that only a fraction of the heat from combustion that is transferred to the steam in the boiler is available for work, and the exergy efficiency compares the actual work output to this available work (i.e., exergy). The exergy efficiencies in these cases become greater than the corresponding thermal efficiencies, providing more

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135

realistic measures of system performance compared to the corresponding thermal efficiencies. For a more comprehensive thermodynamic analysis of a power cycle, the various energy- and exergy-based efficiencies are best considered.

4.5.4.3.2

Heat exchangers

In a heat exchanger, two fluid streams exchange heat without mixing. One of the most commonly used heat exchangers is the shell and tube heat exchanger, where the high-pressure fluid flows through the pipes that are surrounded by the lower pressure fluid. Heat exchangers are usually given different names that reflect the purpose of exchanging the heat between two fluids. Steam generators are heat exchangers that exchange the heat between hot gases and water to produce high-pressure steam. The closed feed water heater is a heat exchanger that exchanges the heat between two different water streams in a Rankine cycle or a power plant. Boilers can also be a kind of a heat exchanger, having a similar function as the steam generator however the steam that exits the boiler usually is saturated vapor. The analysis of heat exchangers based on energy and exergy analyses is presented in the coming examples. Heat exchangers are presented in two main cases in energy systems, such as steam power plant, water heating, cooling space using chilled water, plus others. These two cases are presented next, the first case is considered in Examples 4 and 5, and the second case is presented in Example 6. Example 4: Water at a pressure of 100 kPa and flowing at a mass flow rate of 1.0 kg/s is heated from a temperature of 301C–801C, using a source of thermal energy that releases heat at a temperature of 1201C. Determine (1) the amount of thermal energy required to heat the water, and (2) the energy and exergy efficiencies of the heating process (Fig. 14). Steam turbine/ generator 4

Wet

Steam generator

1 Condenser Pump 3

2

QCond Fig. 14 A schematic diagram of the Rankine cycle.

Solution: We always write the balance equation for both the energy system as a whole or for each component or a device, and they are written as follows: _ in ¼ m _ out ¼ m _ water MBE: m _ in ¼ m _ in hin þ Q _ out hout EBE: m _ in =Ts þ S_ gen ¼ m _ in sin þ Q EnBE: m   _ out sout _ in 1 To ¼ m _ d _ in ex in þ Q _ out ex out þ Ex ExBE: m Ts

The properties of water at the inlet and the outlet conditions are presented next. The properties of the inlet and the exit streams:

State point

P (kPa)

T (oC)

h (kJ/kg)

s (kJ/kg K)

ex (kJ/kg)

Reference state In Out

101.3 101.3 101.3

25 30 80

104.8 125.8 335.0

0.367 0.437 1.075

0.000 0.1734 18.94

(a) Substituting the water properties at different streams in the EBE and the required thermal energy is then calculated. _ in ¼ ð1  335:0Þ ð1  125:8Þ þ Q _ Qin ¼ 209:2 kW

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Heat Engines

(b) The energy and the exergy efficiency of the heating process are calculated using the efficiency equations, which are derived from the energy and the exergy balance equations. The energy efficiency equation derived from the EBE as follows: _ in _ out hout m _ in hin Þ=Q Zen ¼ ðm Zen ¼ ð1:0  ð335:0 125:8ÞÞ=209:2 ¼ 1 ¼ 100% As we can see the energy efficiency is 100%, which will give no indication of whether it was compared to a similar process, and that is why considering the quality of the energy will give us another important measure for the performance of the energy processes and devices. The exergy efficiency equation derived from the ExBE as follows: Zex _ _ Ex Qin Zex

4.5.4.3.3

_ _ _ out ex out m _ in ex in Þ=Ex ¼ ðm Qin   298  209:2 ¼ 50:55 kW ¼ 1 120 þ 273:15 ¼ ð1:0  ð18:94 0:1734ÞÞ=ð50:55Þ ¼ 0:371 ¼ 37:1%

Power plant efficiencies

To assist in improving the efficiencies of power plants, their thermodynamic characteristics and performances are usually investigated. Power plants are normally examined using energy analysis but, as pointed out previously, a better understanding is attained when a more complete thermodynamic view is taken, which uses the second law of thermodynamics in conjunction with energy analysis via exergy methods. Although exergy analysis can be generally applied to energy and other systems, it appears to be a more powerful tool than energy analysis for power cycles due to the fact that it helps determine the true magnitudes of losses and their causes and locations, and improve the overall system and its components. In this chapter, we provide an overview of various energy- and exergy-based efficiencies used in the analysis of power cycles, including vapor and gas power, cogeneration, and geothermal power plants. Differences in design aspects are considered. The various approaches that can be used in defining efficiencies are identified and their implications discussed. Numerical examples are provided to illustrate the use of the different efficiencies, and the results include combined energy and exergy diagrams. Note that the emphasis in this chapter is to describe various energy- and exergy-based efficiencies used in power plants and discuss the implications associated with each definition. Therefore, simple cycles are selected to keep the complexity of the plants at a minimum level for gas and vapor cycles to better facilitate understanding of the efficiencies, which can be very useful for improved energy management in power plants. One can easily adapt the efficiencies discussed here to more complex power systems. Some efficiency definitions for gas cycles found in many thermodynamics textbooks are repeated so that the coverage in this chapter is comprehensive and can serve as a convenient and practical tool for students, engineers, and researchers. It is shown that a better understanding of energy and exergy efficiencies and successful use of them can help improve energy management in power plants.

4.5.4.3.4

Efficiencies of vapor power cycles

The thermal efficiency, also referred to as the energy efficiency or the first-law efficiency, of a power cycle is defined as Zth

1

¼

wnet;out ¼1 qin

qout qin

ð2Þ

where wnet,out is the specific network output, qout is the specific heat rejected from the cycle, and qin is the specific heat input to the cycle, which is usually taken to be the specific heat input to the steam in the boiler of a steam power plant. That is, qin ¼ h3

h2

ð3Þ

where h denotes specific enthalpy and the subscripts refer to state points in Fig. 14. This simple approach neglects the losses occurring in the furnace–boiler system due to the energy lost with hot exhaust gases, incomplete combustion, etc. To incorporate these losses, one can express the thermal efficiency of the cycle by a second approach as Zth

2

¼

_ net;out W _ fuel qHV m

ð4Þ

_ fuel is the mass flow rate of fuel and qHV is the heating value of the fuel, which can be chosen as the higher or lower heating where m value. For furnace–boiler systems where the water in the exhaust gases is not expected to condense, like in internal combustion engines, it is customary to use the lower heating value [1]. Some tend to use lower heating values to make a device appear more efficient. This is frequently done in manufacturer descriptions of commercial boilers. Often a claimed efficiency exceeds 100%. This is due to recovering some of the heat of condensation of steam in the exhaust gases while still defining boiler efficiency based on lower heating value. This is misleading and a thermodynamically improper use of efficiency. If there is any possibility of recovering some of the energy of condensing steam in exhaust gases, the efficiency should be based on the higher heating value.

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137

The second-law efficiency, also referred to as exergy efficiency, of a power producing cycle is defined as Zex ¼

wnet;out ¼1 Ex in

Exd Exin

ð5Þ

where Exin is the specific exergy input to the cycle and Exd is the specific total exergy destruction in the cycle. One can express the exergy input to the cycle as the exergy increase of the working fluid in the boiler of a steam power plant (Fig. 14) as Exin ¼ h3

h2

T0 ðs3

s2 Þ

ð6Þ

where T0 is the dead-state or environment temperature and s is the specific entropy. Zex;1 ¼

wnet;out h2 T0 ðs3

h3

s2 Þ

ð7Þ

In this definition, the irreversibilities during energy transfer from the furnace to the steam in the boiler are not accounted for. Alternatively, the exergy input to the cycle may be defined as the exergy input to the boiler accompanying the heat transfer. The exergy efficiency in this case becomes Zex;1 ¼

wnet;out qin ð1 T0 =Ts Þ

ð8Þ

where Ts is the source temperature. This efficiency definition incorporates the irreversibility during heat transfer to the steam in the boiler. We may also incorporate in the efficiency definition the exergy destruction associated with fuel combustion and the exergy lost with exhaust gases from the furnace. In this third approach, the exergy efficiency can be expressed as Zex;1 ¼

_ net;out W _ fuel exfuel m

ð9Þ

where exfuel is the specific exergy of the fuel. The exergy of a fuel may be obtained by writing the complete combustion reaction of the fuel and calculating the reversible work by assuming all products are at the state of the surroundings. Then the exergy of fuel is equivalent to the calculated reversible work. For fuels whose combustion reaction involves water in the products, the exergy of the fuel is different depending on the phase of water (vapor or liquid). The exergies of various fuels listed in Ref. [1] are based on the vapor phase of water in combustion gases. Different efficiency definitions are possible if one selects different system boundaries. Clearly defining the system boundary allows the efficiency to be defined unambiguously. For example, the exergy efficiencies in Eqs. (8) and (9) correspond to systems whose boundaries are given by the inner and outer dashed lines, respectively.

4.5.5

Case Study: Combined Cycle Fueled With Natural Gas

In the following case study, a combined cycle fueled with natural gas is analyzed using exergy and energy analysis to investigate the performance of the combination of the Brayton cycle and Rankine cycle. However, to show the benefits of combining the Brayton cycle and the steam cycle we are going first to assess the performance of a Rankine cycle, then the combined cycle performance is going to be assessed.

4.5.5.1

•

Systems Description

Rankine cycle In an actual Rankine cycle (Fig. 14) with an inlet pressure to the turbine of 7 MPa and a temperature of 5001C, the turbine exit pressure is 100 kPa. The isentropic efficiency of the turbine is 94%, and the pump, condenser, and the boiler pressure losses are neglected. If the temperature of the water leaving the condenser is 501C and the mas flow rate is 20 kg/s of water in the cycle calculate (1) the thermal energy sent to the boiler, (2) the net power produced, and (3) the energy and exergy efficiencies of the cycle. Writing the mass, energy, entropy, and exergy balance equations is the first step in calculating the vapor fraction and the exergy efficiency of the throttling process.

Condenser: _ out _ in ¼ m MBE: m _ cond _ in hin ¼ m _ out hout þ Q EBE: m _ cond =Tcond _ in sin þ S_ gen;cond ¼ m _ out sout þ Q EnBE: m _ _ _ in exin ¼ m _ out ex out þ ExQ_ cond þ Ex d;cond ExBm Pump: _ out _ in ¼ m MBE: m _ p¼m _ in hin þ W _ out hout EBE: m _ in sin þ S_ gen;p ¼ m _ out sout EnBE: m _ p¼m _ d;p _ in ex in þ W _ out ex out þ Ex ExBE: m

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Heat Engines

Boiler: _ out _ in ¼ m MBE: m _ boiler ¼ m _ in hin þ Q _ out hout EBE: m _ boiler =Tboiler ¼ m _ in sin þ S_ gen;boiler þ Q _ out sout EnBE: m _ _ _ d;boiler _ in ex in þ Ex _ out ex out þ Ex ExBE: m ¼m Qboiler

Steam turbine: _ out _ in ¼ m MBE: m _ st _ in hin ¼ m _ out hout þ W EBE: m _ _ _ EnBE: min sin þ Sgen;st ¼ mout sout _ d;st þ W _ st _ in ex in ¼ m _ out exout þ Ex ExBE: m Using the energy balance equation of the boiler, we can see that in order to calculate the thermal energy the boiler receives from external source we first need to find the enthalpy entering the boiler. The enthalpy of the stream entering the boiler is the same as the enthalpy of the stream exiting the pump. Then: wp ¼ vin;p Pout;p Then:


Pin;p ¼ 0:001012  ð7000

100Þ ¼ 6:984 kJ=kg

¼ hout;p hin;p -6:984 ¼ hout;p 209:4 wp hout;p ¼ 216:4 kJ=kg qboiler ¼ hout;boiler hin;boiler ¼ 3411 216:4 ¼ 3194 kJ=kg _ boiler ¼ m _  qboiler ¼ 20  3194 ¼ 63; 883 kW Q

In the cycle there is the pump that consumes power and there is the steam turbine that produces power, and the power produced by the turbine is calculated based on the energy balance over the turbine wst;is ¼ hin;st hout;st;is ¼ 3422 2466 ¼ 944:5 wst;a ¼ wst;is =Zis ¼ 944:5=0:94 ¼ 1005 kJ=kg wnet ¼ wst;a wp ¼ 1005 6:984 ¼ 944:5 kJ=kg _ net ¼ m _  wnet ¼ 20  944:5 ¼ 19; 957 kW W

The energy and the exergy efficiency equations for the overall cycle are presented next; since there is no mass entering or leaving the overall Rankine cycle, then it is treated as a closed energy system and based on that the efficiencies are derived. _ boiler ¼ 19; 957=63; 883 ¼ 31:24% _ net =Q Zen;RC ¼ W _ _ Zex;RC ¼ W net =Ex _ ¼ 19; 957=42; 076 ¼ 47:43% Qboiler

•

Natural gas fueled combined cycle In this combined cycle natural gas is the fuel that is fed to the combustion chamber. The combined cycle investigated is shown in Fig. 15. From state 1 to state 2 air is compressed using a compressor where the power is supplied by the gas turbine. The resulting compressed air exiting the compressor is used as oxidant in the combustion chamber fueled by natural gas, where state 3 presents the exhaust gases. The gas turbine is used to expand the high pressure exhausts to produce power. The exhaust gases leave the turbine then pass into the shell and tube heat exchanger where they release heat to the water entering the heat exchanger. The heat is released in the heat exchanger, and the exhaust gases are released to the environment. The heat gained by the water entering the heat exchanger leaves in a state of saturated vapor. The saturated vapor heads to the turbine where it expands to produce power. The water leaves the steam turbine in a state of saturated mixture, and is completely converted to saturated liquid through the condenser.

The bottoming cycle is a simple Rankine cycle operating between the pressure limits of 7 MPa and 5 kPa. The steam is heated to a temperature of 5001C in a heat exchanger by the heat supplied from the exhaust gases of a biomass combustion system. The exhaust gases leave the heat exchanger (state 5) at 450K. The ambient temperature and pressure are at 271C and 100 kPa. All the components are considered adiabatic. Biomass cycle is modeled on the basis of air-standard theorem. Air is treated as an ideal gas. Temperature and pressure losses in the pipes and connections are neglected. It is assumed that 20% of the power is lost due to the parasitic losses.

Heat Engines

139

Combustion chamber Qin

2

3 Gas turbine/ generator

Compressor

Wc

Wt

Power shaft

1

5

4

Heat exchanger

1 Wt

Condenser

7

9

6 Qcond

Fig. 15 Natural gas fueled combined cycle.

4.5.5.1.1

Energy and exergy analyses

The pressure at state 2 is calculated as Pr2 ¼ rp  Pr 1 where, Pr2 is the reduced pressure at state 2, rp is the pressure ratio, and Pr1 is the reduced pressure at state 1. The pressure at state 4 is calculated as 1 Pr 4 ¼  Pr3 rp where, Pr4 is the reduced pressure at state 4, rp is the pressure ratio, and Pr3 is the reduced pressure at state 3. First each component of the system is analyzed by writing the balance equation as follows: Air compressor: _2 _1 ¼m MBE: m _ comp ¼ m _ 1 h1 þ W _ 2 h2 EBE: m _ 1 s1 þ S_ gen;comp ¼ m _ 2 s2 EnBE: m _ comp ¼ m _ d;comp _ 1 ex 1 þ W _ 2 ex2 þ Ex ExBE: m Combustion chamber: _ bio ¼ m _3 MBE: m _2þm _ bio LHV bio ¼ m _ 3 h3 EBE: m _ 2 h2 þ m _ bio sbio ¼ m _ 3 s3 EnBE: m _ 2 s2 þ S_ gen;CC þ m _ d;CC _ bio ex bio ¼ m _ 3 ex 3 þ Ex ExBE: m _ 2 ex 2 þ m Gas turbine: _4 MBE: m _3¼m _ GT _ 4 h4 þ W EBE: m _ 3 h3 ¼ m

ð10Þ

ð11Þ

140

Heat Engines

_ 4 s4 EnBE: m _ 3 s3 þ S_ gen;GT ¼ m _ GT þ Ex _ d;GT _ 4 ex4 þ W ExBE: m _ 3 ex 3 ¼ m Heat exchanger: _7¼m _5þm _8 MBE: m _4þm _ 7 h7 ¼ m _ 5 h5 þ m _ 8 h8 EBE: m _ 4 h4 þ m _ 7 s7 þ S_ gen;HX ¼ m _ 5 s5 þ m _ 8 s8 EnBE: m _ 4 s4 þ m _ d;HX _ 7 ex 7 ¼ m _ 5 ex 5 þ m _ 8 ex 8 þ Ex ExBE: m _ 4 ex 4 þ m Steam turbine: _9 MBE: m _8¼m _ ST _ 9 h9 þ W EBE: m _ 8 h8 ¼ m _ _ 9 s9 EnBE: m _ 8 s8 þ Sgen;ST ¼ m _ ST þ Ex _ d;ST _ 9 ex9 þ W ExBE: m _ 8 ex 8 ¼ m Condenser: _ a;in ¼ m _6þm _ a;out MBE: m _9þm _ a;in ha;in ¼ m _ 6 h6 þ m _ a;out ha;out EBE: m _ 9 h9 þ m _ a;in sa;in _S_ gen;C ¼ m _ 6 s6 þ m _ a;out sa;out EnBE: m _ 9 s9 þ m _ d;C _ a;in exa;in ¼ m _ 6 ex 6 þ m _ a;out ex a;out þ Ex ExBE: m _ 9 ex 9 þ m Pump: _7 MBE: m _6¼m _ pump ¼ m _ 7 h7 EBE: m _ 6 h6 þ W _ 7 s7 EnBE: m _ 6 s6 þ S_ gen;pump ¼ m _ pump ¼ m _ d;pump _ 7 ex7 þ Ex ExBE: m _ 6 ex 6 þ W Overall system: _ bio ¼ m _5 MBE: m _1þm _ comp þ W _ pump ¼ m _ GT þ W _ ST _ bio LHV bio þ W _ 5 h5 þ W EBE: m _ 1 h1 þ m _ bio sbio þ S_ gen;ov ¼ m _ 5 s5 EnBE: m _ 1 s1 þ m _ comp þ W _ pump ¼ m _ GT þ W _ ST þ Ex _ d;ov _ bio ex bio þ W _ 5 ex5 þ W ExBE: m _ 1 ex 1 þ m From the balance equations we can rearrange them to calculate the required work, thermal energy, etc. The ideal specific work needed by the compressor of the biomass cycle to compress air from state 1 to state 2 is calculated as wcomp;g;s ¼ hs;2 h1 ð12Þ where wcomp,g,s is the ideal specific work needed by the compressor of the biomass cycle, hs,2 is the ideal specific enthalpy at state 2, and h1 is the specific enthalpy at state 1. The ideal specific power produced by the turbine of the biomass cycle is defined as wturb;g;s ¼ h3 hs;4 ð13Þ where wturb,g,s is the ideal specific work produced by the turbine of the biomass cycle, h3 is the specific enthalpy at state 3, and hs,4 is the ideal specific enthalpy at state 4. The ideal specific heat input to the biomass cycle is found using qin;g;s ¼ h3 hs;2 ð14Þ where qin,g,s is the ideal specific heat input to the combustion chamber of the biomass cycle, h3 is the specific enthalpy at state 3, and hs,2 is the ideal specific enthalpy at state 2. The ideal parasitic loss is calculated as
wparasitic;g;s ¼ 0:2  wturb;g;s wcomp;g;s ð15Þ where wparasitic,g,s is the ideal specific parasitic loss of the biomass cycle. The ideal specific net power produced by the biomass cycle is defined as wnet;g;s ¼ wturb;g;s

wcomp;g;s

wparasitic;g;s

where wnet,g,s is the ideal net specific power produced by the biomass cycle. The specific exergy at state 1 is found using ex 1 ¼ ðh1 h0 Þ T0  ðs1 s0 Þ

ð16Þ

ð17Þ

where ex1 represents specific exergy at state 1, h0 represents specific enthalpy at ambient state, s1 is specific entropy at state 1, and s0 is specific entropy at ambient state. The same formulation is used to calculate specific exergy at each state. The ideal specific thermal exergy of the biomass cycle is found using   T0 ð18Þ ex th;g;s ¼ 1  qin;g;s Tavg;g;s

Heat Engines

141

ðT3 þTs;2 Þ and exth,g,s is the ideal specific thermal exergy of the biomass cycle, and Tavg,g,s is the ideal average where Tavg;g;s ¼ 2 temperature. The ideal energy and exergy efficiencies of the biomass cycle are calculated using wnet;g;s ð19Þ Zen;g;s ¼ qin;g;s Zex;g;s ¼

wnet;g;s ex th;g;s

ð20Þ

The actual specific work needed by the compressor of the biomass cycle to compress air from state 1 to state 2 is calculated as wcomp;g;a ¼

hs;2

h1

ð21Þ

Z

where wcomp,g,a is the actual specific work needed by the compressor of the biomass cycle, hs,2 is the ideal specific enthalpy at state 2, h1 is the specific enthalpy at state 1, and Z is the isentropic efficiency, which is 85%. The actual specific power produced by the turbine of the biomass cycle is defined as wturb;g;a ¼ Z  ðh3

hs;4 Þ

ð22Þ

where wturb,g,a is the actual specific work produced by the turbine of the biomass cycle, h3 is the specific enthalpy at state 3, hs,4 is the ideal specific enthalpy at state 4, and Z is the isentropic efficiency, which is 85%. The actual specific heat input to the biomass cycle is found using qin;g;a ¼ h3

ð23Þ

h2

where, qin,g,a is the actual specific heat input to the combustion chamber of the biomass cycle, h3 is the specific enthalpy at state 3, and h2 is the actual specific enthalpy at state 2. The actual parasitic loss is calculated as
wparasitic;g;a ¼ 0:2  wturb;g;a wcomp;g;a ð24Þ where wparasitic,g,a is the actual specific parasitic loss of the biomass cycle. The actual specific net power produced by the biomass cycle is defined as wnet;g;a ¼ wturb;g;a

wcomp;g;a

wparasitic;g;a

ð25Þ

where wnet,g,a is the actual net specific power produced by the biomass cycle. The actual specific thermal exergy of the biomass cycle is found using   T0  qin;g;a ð26Þ exth;g;a ¼ 1 Tavg;g;a ðT3 þTa;2 Þ where Tavg;g;a ¼ and exth,g,a is the actual specific thermal exergy of the biomass cycle, and Tavg,g,a is the actual average 2 temperature. The actual energy and exergy efficiencies of the biomass cycle are calculated using wnet;g;a Zen;g;a ¼ ð27Þ qin;g;a Zex;g;s ¼

wnet;g;a ex th;g;a

The specific power consumed by pump of the steam cycle is found using   P7 P6 wp ¼ v6 Z

ð28Þ

ð29Þ

where wp is specific power consumed by the pump, v6 is specific volume at state 6, and Z is the isentropic efficiency of the pump. The mass ratio “y” is defined as mass flow rate of the steam divided by mass flow rate of the biomass gas and is calculated using y¼

h4 h8

h5 h7

ð30Þ

The specific power produced by the turbine of the steam cycle is defined as wturb;st ¼ ðh8

h9 Þ

ð31Þ

where wturb,st is the specific work produced by the turbine of the steam cycle, h8 is the specific enthalpy at state 8, and h9 is the specific enthalpy at state 9. The specific heat input to the steam cycle is found using qin;st ¼ h8

h7

ð32Þ

where qin,st is the specific heat input to the steam cycle, h8 is the specific enthalpy at state 8, and h7 is the specific enthalpy at state 7.

142

Heat Engines

The specific heat output from the condenser of the steam cycle is found using qcon;st ¼ h9

ð33Þ

h6

where qcon,st represents specific heat output from the condenser of the steam cycle. The parasitic loss is calculated as
wparasitic;st ¼ 0:2  wturb;st wp

ð34Þ

where wparasitic,st is the specific parasitic loss of the steam cycle. The specific net power produced by the steam cycle is defined as wnet;st ¼ wturb;st

wp

wparasitic;st

where wnet,g,a is the actual net specific power produced by the biomass cycle. The specific input thermal exergy of the steam cycle is found using   T0  qin;st ex th;in;st ¼ 1 Tavg;st

ð35Þ

ð36Þ

8Þ where Tavg;st ¼ ðT7 þT and exth,in,st is the specific input thermal exergy of the steam cycle, and Tavg,st is the average temperature. 2 The specific condenser thermal exergy of the steam cycle is found using   T0 ð37Þ ex th;con;st ¼ 1  qcon;st Tavg;con

6Þ where Tavg;con ¼ ðT9 þT and exth,con,st is the specific condenser thermal exergy of the steam cycle, and Tavg,con is the average 2 temperature. The energy and exergy efficiencies of the steam cycle are calculated using wnet;st ð38Þ Zen;st ¼ qin;st

Zex;g;s ¼

wnet;st ex th;st

ð39Þ

The net overall power output of the combined cycle is found using wnet;ov ¼ y  wnet;st

wnet;g;a

ð40Þ

The ideal specific exergy destruction in the compressor is found using ex dest;comp;g;s ¼ ex 1

exs;2 þ wcomp;g;s

ð41Þ

where exdest,comp,g,s represents ideal specific exergy destruction in the compressor. The actual specific exergy destruction in the compressor is found using ex dest;comp;g;a ¼ ex 1

ex 2 þ wcomp;g;a

ð42Þ

where exdest,comp,g,a represents actual specific exergy destruction in the compressor. The ideal specific exergy destruction in the combustion chamber is found using ex dest;cc;g;s ¼ ex s;2

ex 3 þ exth;g;s

ð43Þ

where exdest,cc,g,s represents ideal specific exergy destruction in the combustion chamber. The actual specific exergy destruction in the combustion chamber is found using ex dest;cc;g;a ¼ ex 2

ex3 þ exth;g;a

ð44Þ

where exdest,cc,g,a represents actual specific exergy destruction in the combustion chamber. The ideal specific exergy destruction in the turbine of the biomass cycle is found using ex dest;turb;g;s ¼ ex 3

exs;4

wturb;g;s

ð45Þ

where exdest,turb,g,s represents ideal specific exergy destruction in the turbine of the biomass cycle. The actual specific exergy destruction in the turbine of the biomass cycle is found using ex dest;turb;g;a ¼ ex 3

ex 4

wturb;g;a

ð46Þ

where exdest,turb,g,a represents actual specific exergy destruction in the turbine of the biomass cycle. The specific exergy destruction in the turbine of the steam cycle is found using ex dest;turb;st ¼ ex 8

ex 9

wturb;st

ð47Þ

where exdest,turb,st represents specific exergy destruction in the turbine of the steam cycle. The specific exergy destruction in the pump of the steam cycle is found using ex dest;p;st ¼ ex 6

ex7 þ wp

where exdest,p,st represents specific exergy destruction in the pump of the steam cycle.

ð48Þ

Heat Engines

143

The specific exergy destruction in the heat exchange is found using exdest;he ¼ ex 4 þ ex7

ex5

ð49Þ

ex 8

where exdest,he represents specific exergy destruction in the heat exchanger. The specific exergy destruction in the condenser of the steam cycle is found using ex dest;con;st ¼ ex 9

ex 6

ð50Þ

exth;con;st

where exdest,con,st represents specific exergy destruction in the condenser of the steam cycle. The overall specific exergy destruction is found using ex dest;ov ¼ ex1 þ ex th;g;a þ wp þ wcomp;g;a

ex5

wturb;g;a

wturb;st

exth;in;st

ð51Þ

where exdest,ov represents overall specific exergy destruction. The overall energy and exergy efficiencies are then defined as Zen;ov ¼

wnet;ov qin;g;a

ð52Þ

Zex;ov ¼

wnet;ov exth;g;a

ð53Þ

2000 1800 1600 1400 1200 1000 800 600 400 200

ra ve O

ns de on

ll

er

p C

xc H

ea

te

m

ng ha

rb tu m ea St

Pu

er

e in

in rb tu as G

C

om

bu

st

io

C

n

om

ch

pr

am

es

so

be

e

r

0

r

Specific exergy destruction (kJ/kg)

Results: As shown in Fig. 16 the component that contributes the most to the exergy destruction in the system is the heat exchanger or the boiler.

Fig. 16 The specific exergy destruction of the components in the cycle.

The cycle overall energy efficiency is 37.3% and the overall exergy efficiency is 54.53% for the same parameters used in Dincer and Ratlamwala [2].

4.5.6

Trilateral Cycle Assessment

According to a study by Dincer and Zam the performance of an ammonia water Rankine cycle was assessed. The cycle uses no boiler, but instead the saturated liquid was flashed with the use of a positive displacement expander (e.g., reciprocating, centrifugal, rotating vane, screw, or scroll type expander) for power generation. The cycle described has no pinch point and therefore the exergy of the heat source is utilized by matching the temperature profiles of the hot working fluids to improve the performance. Another feature of the ammonia water mixture is that it offers further opportunity to match that of the temperature profiles at the heat sink level. This feature brings a 7% improvement of the exergy efficiency with respect to the case with a single type of working fluid such as steam. The influence of the expander efficiency, in accordance with the ammonia concentration and the coolant flow rate, was investigated and reported in the case study. The optimized cycle was also compared to four different organic cycles and a kalian type cycle and showed the best performance. It was also shown that in order to determine the best cycle configuration, the energy efficiency had to be taken into account and had to take the heat recovered from the source. The efficiency of the proposed cycle with ammonia–water as a working fluid was 30%, and with 23% exergy efficiency, which is an increment of 7.0% from the

144

Heat Engines

same operating conditions. In the case of cogeneration the cycle effectiveness may go up to 70%. This cycle can be applied for low power/low temperature heat recovery from geothermal sources, ocean thermal energy conversion, solar energy or process waste heat, etc.

4.5.7

Gas Cycles

Gas cycles differ from vapor cycles as the working fluid in a gas cycle does not experience phase change, therefore is continuously in the gas phase. However, in vapor cycles the working fluid is sometimes a gas, liquid, or both depending on where the working fluid is in the cycle. Gas cycles include Otto, Diesel, Carnot, Stirling, and Brayton cycles. These gas cycles can be divided into two types: fully and internally reversible cycles. The ideal cycles that could be classified as internally reversible include Otto, Diesel, and Brayton cycle, as they involve irreversibilities that are external to the system such as heat transfer from and to the system to its surroundings. The ideal cycles simplify systems as they are assumed not to experience frictional losses, pressure drops; expansion, and compression takes place in a quasi-static equilibrium process; and the system is overall well insulated so it doesn’t experience any heat loss. As for totally reversible cycles they include Carnot, Stirling, and Ericson cycles, as they don’t experience any external irreversibilities such as heat transfer to their surroundings.

4.5.7.1

Totally Reversible Gas Power Cycles

As mentioned in the previous section totally reversible gas power cycles are classified under certain conditions. The heat transfer processes must take place isothermally, meaning the heat transfer rate is infinite. If heat transfer is experienced then this means there is a temperature difference causing heat transfer, making the system externally irreversible. The Carnot, Stirling, and Ericson cycles include isothermal heat addition processes. However, they differ in the following processes: the Carnot cycle experiences two adiabatic process apart from the heat addition ones, the Stirling cycle has two isochoric processes, and the Ericson cycle has two isobaric processes apart from the two heat addition processes these cycles experience. The efficiency of any totally reversible heat engine, such as the Stirling, Carnot, and Ericson, that is operating with a heat source and a heat sink can be defined with the Carnot factor. TL TH

Ztot;Carnot ¼ 1

The P–V diagram shown in the figure is for the Carnot cycle. The following Carnot cycle described in Fig. 17 can be visualized with a heat engine that works with an ideal gas and has a heat exchanger that’s attached to a heat sink, which is followed by a compressor that is both isentropic and adiabatic, followed by a heat exchanger that receives heat from a heat source, and finally followed by a turbine that works adiabatically and isentropically.

P

Qin Ex

pa

ns

ion

Ignition

Com

pre

Patm

ssio

n

Intake opens

V Fig. 17 P–V diagram for the Otto Cycle.

The Stirling cycle’s P–V diagram is shown in Fig. 18. The Stirling cycle can be visualized using a double effect piston and cylinder that includes a porous regenerator matrix placed inside the cylinder. As for the working medium, it could be considered

Heat Engines

145

T

qin 1

S = constant

2

S = constant

TH

TL

3

4 qout

(A)

s

P

1

qin

T

H

Re

ge

4

=c

on

ne

sta

nt

rat

ion

2

TL = co nstant

3

qout

V

(B) Fig. 18 Diagram for the Stirling cycle (A) T–s and (B) P–V diagram.

an ideal gas. The heat transfer processes in and out of the system are considered reversible. The Stirling cycle can be described by the following processes (Fig. 19): 1 TH TL 4 TH TL

2 TH TL 3 TH TL

Fig. 19 Piston system representing the Stirling cycle.

• •

The isothermal expansion of the working medium because of heat addition occurs in process 1-2. The heat addition is performed externally. As the cylinder is heated the left piston moves toward the left and it generates useful work output. The cooling is done historically along with heat regeneration; this is done in process 2-3. The pistons move toward the right side at the same rate, allowing the volume to remain constant. As the pistons move the gas passes over the porous regenerator matrix; as this process occurs the heat transfer occurs between the gas and the porous regenerator matrix. At the end of this process, the pistons reach the right side. This process occurs with no friction making it a reversible process. The gas’s temperature drops during this process from TH to TL.

146

•

•

Heat Engines

The isothermal compression and heat rejection processes occur in process 3-4. Work is applied to the piston on the right side of the cylinder; this, in turn, moves the piston, therefore compressing the ideal gas medium. While the piston is compressing the working gas, the gas starts to transfer the heat to the heat sink; this allows the process to be isothermal. Overall the work input required in process 3-4 is lower than the work output in process 1-2. Therefore the network generated is positive. Process 4-1 includes isochoric heating with heat regeneration. The pistons both move leftward simultaneously keeping the volume between the pistons constant during the process. As the gas moves toward the left side, heat is transferred from the matrix to the gas, therefore raising the temperature of the gas to TH.

There are many difficulties that make the implementation of the Stirling engine impossible. This is due to the unavoidable irreversibilities, especially for heat transfer experienced with finite temperature difference. For example, heat transfer must occur from the regenerator matrix to the gas under an infinitesimal temperature gradient, where the gas reaches its initial matrix temperature. The process described is deemed impossible to replicate or to achieve as there are many irreversibilities during these processes. Even if internal combustion is utilized, this process will still induce high irreversibilities to the regeneration process. This is highly due to the fast combustion process in comparison to the slow regeneration process. Practical Stirling engines currently being used, either use external combustion as at its heat source or use heat sources such as concentrated solar radiation or even waste heat recovery. Some issues with the Stirling engine are that the cylinders need both a heat sink and source, which need an additional temperature difference and thermal response time. With all the disadvantages listed the Stirling engine has been successfully and practically used using low temperature heat sources with the use of helium or air as the working fluid. However, designing a Stirling engine is no easy task as it is very complicated, due to the fact that the dynamic behavior of the mechanism and performance of the heat exchanger highly affect the efficiency of operation. The Ericson cycle includes two isothermal and two isobaric processes. The basic components of an Ericson cycle include a compressor, an expander, and one heat exchanger that are utilized as a regenerator. This makes the Ericson cycle similar to that of the closed-loop Brayton cycle. However, the compression and expansion processes for the Ericson cycle are not adiabatic, unlike the Brayton cycle. The Ericson cycle expansion and compression processes usually include a heat exchanger making the processes isentropic and isothermal instead of adiabatic.

• • • •

Process 1-2: this process includes the isothermal expansion of the working fluid and heat addition process. An external heat source is used to heat the working fluid. As the fluid is heated, it expands simultaneously with the heat addition process at a constant temperature. Process 2-3: this process includes the isobaric heat removal with regeneration process. As the working fluid passes through the regenerative material, the temperature of the working fluid decreases at a constant pressure making this process isobaric. The regenerator stores the thermal energy for process 1-4 where it transfers back to the working fluid. Process 3-4: this process includes isothermal compression and heat removal. The working fluid is compressed at a constant temperature, as the heat is rejected simultaneously. The compression process requires work input; part of the work input is supplied by the work generated by the expansion process. Process 4-1: this process includes isobaric heat absorption from the regenerator stored in process 2-3. After the compressed working fluid passes through the regenerator where heat is transferred to it at a constant pressure, its temperature increases. The reversible process of regeneration is assumed to have an effectiveness of 1.

There are some difficulties implementing an ideal Ericson cycle because of the irreversibilities that occur mainly in the isothermal compression, isothermal expansion, and regeneration processes. The isothermal compression process poses the problem of very fast heat rejection as the working fluid is compressed. To achieve an isothermal state an infinitely large heat transfer area of the cylinder is needed so that the temperature differences are minuscule. There are some obstacles to implementing an ideal Ericson cycle. The difficulties of achieving the ideal Ericson cycle arise from the irreversibilities that are caused during isothermal compression, isothermal expansion, and regeneration processes. Achieving a perfect isothermal state during compression imposes a slow and infinitely large heat transfer area in which the temperatures differences are minimized. Similar issues arise from the expander. As for the regenerator, a temperature gradient is needed for heat transfer to take place. Therefore the effectiveness cannot be 1. Having an effectiveness of 1 in the regenerator would be mean that the heat transfer area would have to be infinitely large, and this would cause a finite pressure drop along the flow, which causes an irreversible process.

4.5.7.2

Otto and Diesel Cycle

Internally reversible cycles as mentioned previously include the Brayton, Otto, and Diesel. These fundamental thermodynamic cycles are a good choice for internal combustion engines and engine-powered generators. The Otto cycle and Diesel cycle will be analyzed and explained in great detail in this section. These cycles as mentioned previously are great and common choices for engine driven generators, including an internal combustion engine installed on the same chassis as an electrical generator and a power regulation block. The applications could be stationary and portable for engine driven power generators. The Otto cycle is best suited for engines that include spark ignition systems. The Otto cycle was established in the 1870s after Nicolaus Otto successfully demonstrated a four-stroke spark ignition engine. The cycle and its processes are demonstrated in Fig. 19. The piston-cylinder operation is demonstrated along with the P–V diagram. The overall cycle includes two isentropic and two

Heat Engines

147

isochoric processes for the working fluid (gas). The example shown in the diagram provides the following relations for volume ratio: rv ¼

v1 v2

r¼

P3 P1

In thermodynamic state 1 there is a gas enclosed T1 ¼ T0 The processes are given as follows:

• •

• •

Process 1-2: this process includes the isentropic compression of the working gas as external work is being applied. The piston moves leftward in this process. The process is considered both isentropic and adiabatic. Process 2-3: this process includes isochoric heat addition. In this process, the piston remains in a fixed position as the heat is being added to the working fluid. In the case of internal combustion engines, this heat would be added by the combustion process of the fuel. The internal temperature and pressure of the cylinder increase considerably during this process. Process 3-4: this process includes isentropic expansion. During this process, the piston moves down and generates usable work. The expansion process continues until the maximum volume of the stroke is reached. v4 ¼v1 Process 4-1: this process includes isochoric heat removal from the cylinder. The piston remains fixed in its final position from the last process while heat is rejected from the working fluid to a heat sink in contact with the working fluid.

The Otto cycle is considered to be externally irreversible as its heat addition, and removal processes are not isothermal. The efficiency of the Otto cycle is lower than that of the Carnot factor. The cycle efficiency can be derived under the air assumption. As the isochoric processes take place, there are no work exchanges between the cylinder and its surroundings. The heat input and output can be described in the following relation: Qin ¼ Cv ðT3

T2 Þ and Qout ¼ Cv ðT4

T1 Þ

The isentropic equations can be derived from the equation above. The isentropic equations are as follows: T2 ¼ rvg 1 T1 and T3 ¼ rvg 1 T4 . By using these relations it could be derived that the efficiency of the internally reversible Otto cycle is:   Qin T4 T1 Z¼1 ¼1 ¼ 1 rv1 g Qout T3 T2 The exergy efficiency of the cycle is derived as follows as well: c¼

_ net _ Z W W  net  ¼   ¼ _Exin T 0 _n 1 Q 1 TTso0 Tso

The variation of energy and exergy efficiency of the Otto cycle in relation with the volume ratio is provided in three pressure ratios in Fig. 20. The volume ratios’ typical range for the compression process for spark ignition engines varies between 7 and 11.

 0.8

 for r = 70  for r = 90  for r = 120

 and 

0.6

0.4

0.2

Typical range 0 2

4

6

8

rv

10

12

14

16

Fig. 20 Efficiency variance with rv for ideal Otto cycle using the standard air assumption. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

148

Heat Engines

The energy efficiency range for the ideal Otto cycle is within the range of 40%–50%. The exergy efficiency of the Otto cycle depends on the additional parameters apart from rv : rv is described as the ratio between the maximum and minimum pressures of the Otto cycle. The exergy efficiency for the pressure ratio range of 7–11 and for r¼ 70–120 ranges from 44% to 68%. From the exergy efficiency range of the Otto cycle it could be seen the magnitude of the exergy destruction due to the heat transfer from the heat source and to the heat sink on the exergy efficiency. The ratio of the exergy destruction to the exergy output is within the range 32%–66%.

4.5.7.3

Diesel Cycle

The Diesel cycle was suggested by a German researcher Rudolf Diesel in 1895. The Diesel cycle is a widely used cycle that has many applications especially in industrial power generation as large engine generators can be installed. The engine generator uses the Diesel cycle. Therefore the engines are enlisted as compression ignition internal combustion engines. The combustion process in this type of engine occurs by compressing the air up to the auto ignition temperature of the fuel. The combustion process occurs when the air is compressed to the auto ignition temperature where the pressurized fuel is then sprayed and instantaneously combusts. The air temperature for the combustion process to occur will have to be over 800K; the volume ratio is also known as the compression ratio for the Diesel cycle is in the following range rv ¼12–24. With such compression ratio and high temperatures, a spark is no longer required to initiate the combustion process in a compression ignition engine. The rate of the reaction due to no spark plug causes the combustion to be relatively slow compared to that of the spark ignition engine. The pressure within the cylinder can be kept relatively steady contingent that the injection process of the fuel occurs at the top dead center of the cylinder. The volume during the combustion process increases and the temperature of the gas will also increase.

• • • •

Process 1-2: this process includes the isentropic compression of the working gas as external work is being applied. The piston moves leftward in this process. The process is considered both isentropic and adiabatic. Process 2-3: this process includes isobaric heat addition. In this process, the volume of the cylinder increases as the heat is being added. The temperature of the gas increases as the heat is being added, even though the volume is increasing. Process 3-4: this process includes isentropic expansion. During this process, the piston moves down and generates usable work. The expansion process continues until the maximum volume of the stroke is reached. v4 ¼ v1 Process 4-1: this process includes isochoric heat removal from the cylinder. The piston remains fixed in its final position from the last process while heat is rejected from the working fluid to a heat sink in contact with the working fluid. The efficiency of the Diesel cycle under the standard air assumption is given by the following equation:

ZDeisel ¼ 1 þ

4.5.7.4

QL Cv ðT1 ¼1þ QH Cp ðT3

T4 Þ T2 Þ

Gas Turbine (Brayton) Power Cycles

In the 1930s combustion turbine power plants started to be commercially developed, based on an open type simple Brayton cycle. The simple Brayton cycle includes a turbo compressor, combustion chamber, and a gas turbine. In the 1930s the combustion turbines were thought to be the best technology to convert chemical exergy of gaseous and liquid fuels in electric power. The combustion turbine could be used with various types of fuel such as natural gas, fuel oil, and gasified coal. The combustion turbine was further developed with changes and additions, which led to more advanced combustion turbine power generators. The advanced combustion turbine power generators allowed for a higher efficiency as they integrated regenerators, gas reheaters, compressor intercoolers, and turbine blade cooling (this allows for higher operating temperature). The combustion turbine also progressed in startup times relative to the steam Rankine power plant, in some cases startup times lower than 1 min. The short startup times allowed the combustion turbines to be recommended for use in regional grids to compensate for peak loads. The efficiency of a combustion turbine power plant is known to reach values over 40% when the expelled gases have elevated temperatures of over 625K. If the combustion turbine is combined with a Rankine cycle power plant, the efficiencies are known to reach and surpass 55%. This cycle is known as the combined cycle.

4.5.7.4.1

Air-standard Brayton cycle

The working fluid for an actual combustion turbine is represented by the pressurized combustion gases that are compressed, expanded, and generate power. Excess air is provided to the combustion chamber to increase fuel utilization. Excess air provided is in the range of 4–50 approximately. Therefore, using air as a combustion gas is a good approximation and is widely accepted. kJ at a temperature of 298K and When the working fluid is modeled as an ideal gas using air, with the specific heat value of 1.005 kgK g¼ 1.4 this combustion turbine cycle is named the air standard Brayton cycle (Fig. 21). If the cycle compromised of internally reversible processes then the cycle is denoted as an ideal standard air Brayton cycle. The simple Brayton cycle is compromised from four processes, which include isentropic compression, isobaric heat addition, isentropic expansion, and isobaric heat rejection. On a mechanical level, the processes are performed by the following components: the compressor, heater, turbine, and cooler.

Heat Engines

Qin

149

Combustion chamber

2

3 Gas turbine/ generator

Compressor

Wc

Wt

Power shaft

1

4

Heat exchanger Q out Fig. 21 Standard Brayton cycle with heat exchanger schematic.

The Brayton cycle has isobaric processes for heat addition and rejection, instead of isothermal processes. This makes the Brayton cycle an externally irreversible cycle. The cycles that can be called externally reversible as mentioned previously are the Carnot, Stirling, and Ericson cycles. Furthermore, the Brayton cycle could operate as an internal combustion system as well as an external combustion system; this is highly dependent on the manner at which the isobaric heat addition occurs from the heat source. The majority of combustion turbine power plants operate based on the open Brayton cycle shown in Fig. 22. The open Brayton cycle is directly connected with atmosphere. Therefore its intake and exhaust pressure are equal to the atmospheric pressure. These constraints represent a limitation regarding working conditions. This constraint can be overcome with the utilization of a heat 1300 Working fluid: standard air (perfect gas assumption)

3

1080 P = 10 P0

T (K)

860 4 640

2

420

P = P0 = 1.01325 bar 1

200 5.6

5.8

6.0

6.2

6.4

6.6

s (kJ/kg.K) Fig. 22 Ideal air Brayton cycle T–s diagram.

exchanger, which takes the role of the cooler. With the use of a heat exchanger, the turbine could be made to discharge in a vacuum, increasing the power output and the efficiency. The turbine discharge pressure could be adjusted by means of changing the heat sink temperature. However, the closed Brayton cycle must operate with an inert working fluid that could include air, helium, etc. The heat addition process can be achieved from external combustion by means of heat transfer to the heater process 2-3. The working fluid and the combustion gases are never in contact throughout the process, and with these changes, the Brayton cycle becomes an external combustion engine. The key point is that both the open and closed Brayton cycles are equivalent thermodynamically. However, the open cycle uses the atmosphere as its cooler whereas the closed Brayton cycle uses the heat exchanger as its cooler.

150

Heat Engines

The ideal air standard Brayton cycle’s T–s diagram is shown in Fig. 23. The energy and exergy balance equations can be noted for each of the cycle processes. They can be noted by assuming that the surrounding atmosphere is at standard conditions, T0 and 900 net

800

c 700

t

w (kJ/kg)

600 500 400 300 200 100 0 0

10

20

30 r

40

50

60

Fig. 23 Specific work relative to the compression ratio in standard Brayton cycle. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

P0, which define the reference states for the exergy analyses. The ideal Brayton cycle is internally reversible, meaning that both the compression and expansion processes occur isentropically. However, the heat addition and rejection processes occur isobarically, meaning the process is externally irreversible, which means that the boundary at the heat source is set to T3 (the highest temperature within the cycle) and the heat sink temperature is set to T1 (the lowest temperature within the cycle). Therefore, there will be entropy generation, which leads to exergy destruction within the cycle. The network generated from the cycle is defined by the following relation: wnet ¼ wt wc. The relation also defines the net power generated by the Rankine cycle. The Brayton cycle’s net power generated can also be defined by the back work ratio (BWR). The BWR represents the relation in the form of ratio of the work consumed by the compressor and the work output of the turbines. The BWR can be defined by the following equation: BWR ¼

wc wt

ð54Þ

Therefore, the network output can be defined as follows: wnet ¼ wt(1 BWR). As for the energy efficiency of the cycle it could be represented by the ratio between the network output and total rate heat input to the cycle.   qsi qso qsi Z  ¼ ; exergy efficiency can be defined c1 ¼ Z¼1 T0 4Z T0 qso 1 qso 1 T3 T3

The exergy efficiency above is defined as the ratio of the useful work output to the total exergy input at the source heat. An alternative way to define the exergy efficiency relation is to use the exergy the received by the cycle exin and the exergy output at the heat sink exout. The exergy efficiency can also be defined as follows: 0 1 wnet ex d   A ð55Þ ¼1 @  c2 ¼ ex in exout qso 1 T0 qsi 1 T0 T3

T1

The third way of defining the exergy efficiency of the overall cycle is by using the ratio of the energy efficiency relative to the reversible heat engine efficiency that is operating between T1 and T3. The exergy efficiency can also be defined as follows: c3 ¼

Z 1

T1 T3

¼1

1 1

qsi qso T1 T3

ð56Þ

Heat Engines

151

A remark could be given about all three exergy efficiency equations parameters in the following: If T1 ¼ T0 c1 ¼ c2 ¼ c3 This relation is valid for the open Brayton cycle, as it is assumed that the reference temperature is that of the atmospheric air. In individual cases with the closed Brayton cycle, T1 may differ from T0 hence c1 ¼c2 ¼ c3. As for the standard air Brayton cycle, a perfect gas equation of the specific state can be applied to derive analytical relations for the overall cycle efficiency. The ideal gas equation assumes that the gas has a constant specific heat regardless of what the temperature might be. The specific heats are estimated at 298.15K using the following relation Dh¼cpDT and the energy balance equation, the energy efficiency becomes: !     T4 1 cp ðT4 T1 Þ T1 T1 ¼1 Z¼1 ð57Þ T3 T2 cp ðT3 T2 Þ 1 T2 Another general relation that could be derived from a perfect gas that describes the relationship between temperature and pressure during an isentropic process (1-2) in accordance with T1 P2K ¼ T2 P1K is k¼

1

g g

ð58Þ

where g is the specific heat ratio. Given that processes 1-2 and 3-4 are isentropic and operate within the same pressure, P1 ¼P4 and P3 ¼ P2, it could be noted that  k  k T3 P3 T2 P2 ¼ ¼ ¼ ð59Þ T4 P4 T1 P1 T4 T3 ¼ T1 T2

ð60Þ

With the following relations the energy efficiency of the standard air Brayton cycle becomes the following: k

ð61Þ

P2 41 P1

ð62Þ

Z¼1 r¼

r

Therefore the energy efficiency of the standard air Brayton cycles becomes only dependent on the pressure ratio of the cycle. Meaning the cycle is not dependent on the highest (T3) and lowest (T1) operating temperatures of the cycle. However, some restrictions do apply to the operating temperatures due to the configuration of the cycle. The heat source temperature T3 must be higher than that of the compressor discharge temperature T2. In accordance with the relation T2 ¼ rkT1 the cycle must operate with T34rkT1. The exergy efficiency is highly affected unlike the energy efficiency, as exergy input is dependent on the heat sink temperature T1. As for the heat source temperature, T3, it is inversely proportional with the exergy efficiency, meaning as T3 increases the exergy efficiency decreases. Furthermore, the pressure ratio is fixed due to Z will not vary causing the Carnot factor to increase. However, there is a pressure and temperature limitation due to material and cooling issues for the turbine. The pressure at the turbine inlet is typically 20–30 bar. The temperature at the inlet is typically limited to 1200–1500K. These limitations constrain and justify the cycle design for the maximum temperature of the heat source. These justifications and constraints in the design cause tradeoffs that will be mentioned next. As seen in Fig. 24, when the temperature values of the heat source T3 and the heat sink T1 are restricted the net work generated is maximized at certain pressure ratios. It is to be noted that for small pressure ratios the specific turbine work generated is small relative to the power consumed by the compressor. In other extreme scenarios, if the pressure ratio has a high value, the compressor network input is higher than turbine network output. Therefore, in both scenarios (low and high pressure ratios) the network output from the power cycle that is defined by the difference of wt and wc is small. However, in between these two extremes pressure ratios lie pressure ratios that generate high network. In Fig. 24 it could be observed that the specific network output wnet is obtained when the pressure ratio is roughly 14. This is a highly considered value when designing a power plant because if the specific work is high, then the power plant could be more compact. To define the compactness of a power plant relative to the specific network output, one can introduce the specific size parameter (SSP), which is defined as the reciprocal value of the specific work output shown below:   1 kg ð63Þ SSP ¼ wnet kJ The SSP relation to the pressure ratio is plotted in Fig. 24. In Fig. 25 the energy and exergy efficiency curves could be found for the cycle. The minimum SSP that is the most compact system is defined as the “knee” of the efficiency curves of the system. If the pressure ratio value is higher than the optimal pressure ratio for the cycle then the efficiency decreases significantly and quickly. From an economic point of view it does not make sense to enhance the efficiency of the cycle as this will increase the plant’s overall size and cost.

Heat Engines

× 10−3 6.50

0.9 SSP 1 

6.00

0.8 0.7

SSP (kg/kJ)

5.50

0.6 5.00 0.5 4.50

1 and 

152

0.4 4.00

0.3

3.50

0.2

3.00 0

20

10

30 r

40

50

0.1 60

Fig. 24 Compression ratio relative to the specific size parameter (SSP) for a standard air Brayton cycle. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

Qin Combustion chamber Regenerator 3

2 Compressor

4

5

Gas turbine/ generator

Wc

Power shaft

Wt

Power shaft Regenerator 4

4

1

6

Heat exchanger Qout

Fig. 25 Brayton cycle with regenerator components.

The efficiency of the Brayton cycle can be increased with some modification that can be applied to the basic cycle, where the external irreversibilities experiences due to the finite temperature differences at the heat sink and source are reduced. The improvements can be achieved by the following modifications:

• •

•

Regeneration could be utilized by preheating air after it exits the compressor using the heat recovered from the cot gas exhausted by the turbine. Therefore, the heat input required is reduced for a similar network generated from the cycle, which raises the efficiency. Reheating in this case is performed by the means of installing an interstage reheated in between two or more turbines. In an actual combustion turbine, typically a high and low pressure turbine are installed on the same shaft, with a secondary combustion chamber installed in between the turbines to supply heat to the working gas. Therefore, the working fluid is reheated before entering the low pressure turbine. The reheating increases the network output, which increases the cycle efficiency. Intercooling is performed by performing the air compression process in stages with interstage cooling. The intercooling decreases the work consumed by the compressor, therefore decreasing the BWR, which increases the overall cycle efficiency.

Heat Engines 4.5.7.4.2

153

Regenerative Brayton cycle

The improved Brayton cycle with regeneration is shown in Fig. 24. The regenerator will preheat the working gas coming out of the compressor Eq. (2) prior to entering the main heat at the heat source Eq. (3). The stream of the hot low pressure gas exhausted by the turbine Eq. (5) is further cooled down by the regenerator, for the regenerator to heat the working fluid. After the regenerator the stream is either exhausted to the environment Eq. (6) or recycled by the means of a cooler, which is dependent on the cycle type (closed or open). The temperature at the temperature outlet must be higher than that at the compressor exit, as the heat transfer in the regenerator is performed from stream 5-6 to 2-3. Therefore, the constraints of the cycle are T54T2. The T–s diagram of the regenerative Brayton cycle is shown in Fig. 26. On the diagram it can be seen that the reduction of the heat input in conditions where the net work output is constant can be clearly observed (Fig. 26). 1300

4

Working fluid: Standard air (ideal gas) 1080 P = 10 P0 T (K)

860 3

5

Regeneration process

640 2

6

420

P = P0 = 1.01325 bar 1

200 5.6

5.8

6.0

6.2

6.4

6.6

s (kJ/kg.K) Fig. 26 T–s diagram for Brayton cycle with regenerator components.

One can derive a simple expression for the exergy efficiency of the Brayton cycle with a regenerator, using the air standard and no internal irreversibility assumptions. The no internal irreversibility assumption comes from the implication that the regenerator is well suited to match the cold and hot stream profiles. The outlet temperature of the regenerator at state 3 reaches a similar temperature of the turbine outlet at state 5. Therefore, it could be assumed that T3 ¼ T5 due to the standard air assumption with a constant specific heat. The energy balance for the regenerator unit includes the following relation: h3

h2 ¼ h5

h6 -T3

T2 ¼ T5

T6 -T2 ¼ T6

ð64Þ

The assumption of no internal irreversibilities derives the implication that the cycle experiences isentropic expansion and compression processes. These processes can be modeled based on the ideal gas equation of state, similar to the simple configuration indicated above.  K  K T2 P2 T4 P4 ¼ ¼ ¼ ð65Þ T1 P1 T5 P5 The energy efficiency is derived from the overall cycle energy balance, which includes that the network output has to be equal to the heat input minus the heat rejected. Therefore, the energy efficiency is considered to be one minus the heat rejected by the heat input. Using the two equations the energy efficiency becomes the following:   T1 ðT1 T2 Þ1 cp ðT6 T1 Þ T2 T   ¼ 1 тr k ; where т ¼ 1 ¼1 ð66Þ Z¼1 T1 T4 cp ðT4 T3 Þ ðT1 T2 Þ1 T2

The above energy efficiency equation includes the ratio т, which represents the relation between the highest and lowest temperatures in the cycles. Given that T44T2 ¼ rkT1 one should note that тor k the energy efficiency reduces to zero. Furthermore, if one predefines the condition that the overall cycle efficiency with the inclusion of regeneration must be higher than that without the modification, it must satisfy the following inequality: 1

r k o1

тr k or тor 2k

ð67Þ

In which this is more restrictive than the previous condition. It should also be noted that the cycle efficiency reaches the Carnot efficiency regardless of the pressure ratio r, if the heat source temperature is infinity, as т¼ 0 and hence Z ¼ 1: Furthermore, one can use the efficiency improvement factor (EIF) specifically for the regenerative Brayton cycle in comparison to the simple Brayton

Heat Engines

154

cycle, meaning it is an efficiency ratio. EIF ¼

1 1

Zregeneration тr k ¼ r k Zbasic

ð68Þ

The above equation illustrates the EIF’s dependence on the heat sink and source temperature ratio (т) and various pressure ratios r for the standard air assumption. Low pressure ratios and high heat source temperatures produce a better EIF.

4.5.7.4.3

Reheat regenerative Brayton cycle

Fig. 27 displays the diagram for the improved Brayton cycle, which includes a reheater along with a regenerator. The reheating process increases both the network generated and a higher temperature from the turbine’s exhaust. A higher exhaust temperature 7 Qin Combustion chamber Regenerator 3

2

4 Gasturbine/ generator

Compressor

Wc

Wt

Power shaft Reheater 5

6

1 Qin 6

Heat exchanger Qout

Fig. 27 Improved Brayton cycle that includes a reheater along with a regenerator.

means higher potential for heat recovery for the regenerator. The T–s diagram for the Brayton cycle with reheating and regeneration is shown in Fig. 28. The standard reference state in the Brayton cycle is assumed to be right in the inlet of the compressor. All the internal processes are considered reversible. The cycle is not externally reversible due to the temperature differences that are found throughout the cycle. The heat source temperature is assumed to be the highest temperature in the cycle (T4 ).

T 4

6

3

7

5 2 8

1

S

Fig. 28 T–s improved Brayton cycle that includes a reheater along with a regenerator.

The heat sink temperature is assumed to be that of the atmosphere’s: T0 ¼ T1 ¼298.15K for the open cycle type Brayton cycle. The exergy destruction occurring mainly at the heater (process 3-4), the reheater (process 5-6), and cooler (process 8-1) is calculated based on the exergy balance equations. The turbine’s work generated is maximized by selecting an ideal interstage pressure for expansion where P6 ¼ P5 must be chosen. Note that the overall pressure ratio r ¼ PP21 is an constraint, but the pressure ratio in the

Heat Engines

155

first stage, r ¼ PP45 , is taken as a variable, where T4 ¼ T6. Therefore, the specific work generated by both turbines is shown by: w ¼ kRT4 r1k þ r k r1 k


y 2 with k ¼

1 y

and y ¼

Cp CV

ð69Þ

r1opt ¼ r 0:5 : This kind of result is also valid for interstage compression processes and is applied to reduce the amount of work consumed by the compression process.

4.5.7.4.4

Brayton cycle with intercooler

With the use of interstage cooling in the compression process, the energy and exergy efficiency of the overall Brayton cycle increases for two main reasons: the compressor power work consumption is reduced and more heat can be recovered from regeneration. A Brayton cycle with reheating, regeneration, and intercooling is shown in Fig. 29. The T–s diagram for the cycle is shown in Fig. 30. 9 Qin Combustion chamber Regenerator 4

5

6 Gas turbine/ generator

Compressor

Wc

Wt

Power shaft 2

Reheater

Intercooler 3

7

Qout

8 Qin

Heat exchanger Q out

10 1

Fig. 29 Improved Brayton cycle that includes a reheater along with a regenerator and intercooler.

1600 1400

r = 15,  = 0.2  = 0.71  = 0.88

6

8

T (K)

1200 5

1000

7

9

800 600 4

2

400

10 3

200 150,000

1 160,000

170,000

180,000

190,000

200,000

210,000

s (J/kmol.K) Fig. 30 T–s improved Brayton cycle that includes a reheater along with a regenerator with intercooler. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

For the case study the compression was the same, r¼ 15 and т¼0.2; for the study above reheating and regeneration is only considered. It is seen that intercooling improves the energy and exergy efficacies of the cycle. The external irreversibilities cause exergy destruction as follows in the case study: 22% at the heater and reheater and 28% at the intercooler and cooler.

156

Heat Engines

4.5.7.5

Exergy Destructions in Brayton Cycle Power Plants

The estimation of exergy destruction for the Brayton cycle is very important as they affect the efficiency of the cycle significantly. The back work ratio of the Brayton cycle is usually relatively high. The isentropic efficiency of the turbomachinery also has drastic effects on the cycle performance. In the Brayton cycle the turbomachinery must have an isentropic efficiency of 80% for the power plant to be viable. The irreversibility experienced by the turbomachinery is one case, however there are irreversibilities in the Brayton cycle due to the heat transfer across finite temperature differences, pressure drops in flow and heat exchangers (Figs. 30 and 31). 2i

2s

2r 2ha

T (K)

600

Exergy distructions: Ideal gas (1−2i): 37.844 kJ/kg Real gas (1−2r): 38.388 kJ/kg Humid air (1−2hr): 38.970 kJ/kg T1 = 298.15K P1 = 101.325 kPa r = 15

400 Process 1−2s: s = ct, standard air Process 1−2i: s = 0.8, air (idal gas, Cp = ct.) Process 1−2r: s = 0.8, air (real gas) Process 1−2hr: s = 0.8, humid air (1 = 60%) 165

166

167 s (kJ/kmol.K)

168

169

Fig. 31 The T–s diagram displaying the actual compression process of three different working fluids. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

The isentropic efficiency is a key parameter in quantifying the irreversibilities in the Brayton cycle. The typical isentropic efficiency of a compressor in the Brayton cycle is within the following range: ZSC ¼ 0.8 0.85. The isentropic efficiency of a compressor could be defined as the ratio of the isentropic work divided by the actual work as the compressor is pressuring the working fluid. The isentropic efficiencies affect multiple types of irreversibilities as the operation of the compressor is proceeding. The irreversibilities experienced by the compressor are relatively higher than those experienced by the turbine, due to the dissipative phenomena, such as weak compression. The thermophysical properties of the working fluid have a significant impact on certain measures of the magnitude of dissipation, specifically the effects due to specific heat. If the specific heat is relatively higher, then the capacity of internal energy increases in the working fluid; this property directly affects the temperature and the exergy of the stream at the compressor outlet. This affects combustion turbine power plants specifically as the air excess ratio is high in such power plants, therefore the specific heat of the combustion products is very similar to that of air in similar pressure and temperature conditions. Based on the previous comment it can be seen why the working fluid in the Brayton cycle is considered air. However, to get an accurate estimate of the exergy destruction experienced in turbomachinery, the specific heat variation with temperature must be considered. In order to demonstrate the effect of the working fluid on the exergy destruction in the compressor, we can use a compressor with a compression ratio of 15 as an example under the standard air assumption. The compressor will use three types of air including standard air (constant specific heat), real air (variations of the specific heats are accounted for different states (various temperatures) and the gas is modeled as dry air), and humid air (it is assumed that the relative humidity is 60% at the compressor intake). In all three cases the isentropic efficiency is assumed to be 80%. Fig. 27 demonstrates the compression processes for all three cases in the T–s diagram. When the working fluid is assumed to follow the standard air assumption, the temperature after the isentropic compression process is found to be 646.3K. This value is obtained by using the following equation: TPk ¼ cons. with k ¼ 1 g g. The actual temperature of the compressor discharge could be determined by taking into account a constant specific heat value and including the isentropic efficiency value to derive the following relationship: Cp ðT2 ;S

T1 Þ ¼ Zs Cp ðT2

T1 Þ

ð70Þ

where the second state (compressor discharge) is considered to be isentropic hence the subscript 2,s is given. Subscript 2 is given to the actual discharge state. The compressor in the example given previously is assumed to follow the standard air condition; as a result the thermomechanical exergy at the intake is set to zero. The work consumption of the compressor is given by the following relation: wc ¼h2 h1 ¼Cp(T2 T1). The exergy balance equation for the compressor using the standard air assumption can be used to find the exergy destruction as follows: 0 þ wc ¼ h2

h1

T0 ðs2

s1 Þ þ exd -ex d ¼ T0 ðs2

s1 Þ

ð71Þ

Heat Engines

157

The compressor discharge temperature under the standard air assumption is T2,S ¼ 733.4K and the exergy destruction is kJ . The exergy destruction is calculated for process 1-2r as shown in Fig. 23. The figure also displays the actual valued at 34.88 kg compression of the real air working fluid, where the compressor discharge temperature is found to be T2,a ¼719.2K and the exergy kJ destruction is higher 38:97 kg . In the case with humid air the compressor discharge temperature is even lower at T2,h ¼719.1K. The standard air assumption is a good approximation for modeling the compression process, however the real gas model should be used for improved accuracy. The real gas model accounts for specific heat variations, allowing for a more accurate modeling (Fig. 32). 750 T2 EDF

7.6

7.4

700

EDF (%)

T2 (K)

725

7.2

675

7

650 Case I

Case II

Case III

Fig. 32 Comparison of the turbine outlet temperature and the exergy destruction fractions for the different three cases. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

As for turbines their isentropic efficiency lies between the following range: Zts ¼ 0:8 0:9. To demonstrate the influence of the various working fluid types that include standard air, ideal gas air, and combustion gases, they will be discussed in the form of a case study. The inlet temperature of the turbine is set as 1273K and the inlet pressure is considered to be equal to that of the atmospheric pressure multiplied by the compression ratio. The compression ratio is considered in this case to be 15. The isentropic kJ where the exergy efficiency of the turbine is fixed at 85% the final analysis shows that the molar specific exergy was exd ¼ 1:4mol kJ . A relation to relate exergy destruction to the exergy input can be used as a parameter called the exergy input was exin ¼ 18:4mol destruction fraction: ex d EDF ¼ ð72Þ ex in The second case in this example considers real air with varying specific heat depending on the conditions Cp ¼ Cp(T): A commonly used practice to calculate entropy and enthalpy as it changes as an ideal gas requires the tabulation of the Cp values versus temperature.R There are many common thermodynamic tables that provide the values of Cp(T) and provide the data for the T integral h0 h00 ¼ T0 Cp ðT ÞdT; where h0 is the corresponding enthalpy at the given temperature T and h00 is the enthalpy value at the corresponding reference state T0: The following integral could also be found in tabulated form in thermodynamic tables: Z T Cp ðT Þ dT ¼ s0 s00 ; where s00 ¼ sðT0 ; P0 Þ and s0 ¼ sðT; P0 Þ T T0 Therefore, the enthalpy value can be expressed at any temperature (T) where hðTÞ ¼ h0 ðTÞ and the entropy at any T and P is obtained by sðT; P Þ ¼ s0 ðT Þ Rln PP0 . Using the assumptions and considerations the calculations are presented in Table 2. As for the calculation of the temperature variation along an isentropic process, if the working fluid is an ideal gas with a varying specific heat the procedure in chapter 7 of Cengel and Boles (2010) is used. The method includes the use of reduced pressure, and can be derived as follows considering an isentropic process: s1 ¼ s2     P1 P2 s0 ðT1 Þ Rln ¼ s0 ðT2 Þ Rln ð73Þ P0 P0 When rearranged the equation can represent the following: 0  exp s ðRT2 Þ P2 P; 0  ¼ r 2 ¼ P1 Pr;1 exp s ðRT1 Þ 0  where the reduced pressure is represented as the following Pr ¼ exp s RðT Þ ; which is only a function of temperature.

ð74Þ

158

Heat Engines

Table 2 Calculation step

Assumptions

Equations

Calculated values

Reference state

T0 ¼ 298:15K P0 ¼ 101:325kPa kJ h0 ¼ 8649 kmol kJ s0 ¼ 165 kmolK

∅0 ¼ h T0 s0 ex0 ¼ 0 P0 v0 ¼ RT0

kJ ∅0 ¼ 40:5 mol m3 v0 ¼ 24:47 kmol

Parameters of state 1 (inlet)

T1 ¼ 1273K P1 ¼ rP0 r ¼ 15

m v1 ¼ 6:964 kmol kJ h1 ¼ 37:0 mol J s1 ¼ 185 molK

Standard air kJ Cp ¼ 29:101 molK

P1 v1 ¼ RT1 h1 h0 ¼ Cp ðT1  T0 Þ   s1 s0 ¼ Cp ln TT10 Rln PP10

ex1 ¼ h1

Parameters of state 2s

Process 1-2s: Isentropic expansion, g ¼1.4

s2s ¼ s1 ; P2s ¼ P0 ; g g P2s v2s ¼ P1s v1s ; g 1 v1 T1g 1 ¼ v2s T2s ; h2s h0 ¼ Cp ðT2s T0 Þ; ex2s ¼ h2s T0 s2s ∅0

m v2s ¼ 48:19 kmol T2s ¼ 587:2K kJ h2S ¼ 17:1 mol J s2S ¼ 185 molK kJ ex2S ¼ 2:5 mol

Parameters of state 2

Zs;t ¼ 0:85

P2 ¼ P0 ; ðh1 h2Þ ¼ Zs;t ðh1 ðT1 T2 Þ ¼ Zs;t ðT1 P2 v2 ¼ RT2 ;   s2 s0 ¼ Cp ln TT20

m v2 ¼ 56:6 kmol T2 ¼ 690:1K kJ h2 ¼ 20:0 mol J s2 ¼ 189 molK kJ ex2 ¼ 4:1 mol

Exergy balance

Process 1–2: Adiabiatic expansion

T0 s1

∅1

exd exom

¼1

kJ ex1 ¼ 22:5 mol

3

3

h2s Þ; T2s Þ; Rln

ex2 ¼ h2 T0 s2 ∅0 ; ex1 ¼ ex2 þ wt þ exd ; exin ¼ ex1 ex2 ; wt ¼ h2 h1 C ¼ exwint ; EDF ¼

3

  P2 P0

kJ exd ¼ 1:4 mol ; kJ ; exin ¼ 18:4 mol EDF ¼ 7:63%

C

The above equations show the relation of pressure and temperature variation in an isentropic process. As seen in Table 3 the exergy destruction fraction in case (ii) is 7.13%, hence it is slightly lower than that of standard air. In case (iii) is the complete combustion of CH4 using an excess air fraction of 5. The combustion reaction equation is given below: CH4 þ 2lðO2 þ 3:76N2 Þ-CO2 þ 2H2 O þ 7:52lN2 þ 2ðl

1ÞO2

ð75Þ

Therefore, the mole fraction of the compounds in the product gas results are dependent on the total number of moles/mole of reacted methane: 1 þ 2 þ 7.52l þ 2(l 1)¼48.6. The molar fractions product gas are yCO2 ¼ 2:0%; yH2 O ¼ 4:1%; yN2 ¼ 77:4%; and yO2 ¼ 16:5%. As for the temperature of the working fluid at the turbine inlet it is to be above T1 ¼ 1273K, since the adiabatic flame temperature at any given condition has to be estimated for the verification of AFT4T1. The AFT is the result of the energy balance of the adiabatic combustion chamber, where HR ¼ Hp, where H is the total enthalpy where the subscripts stand for (R) reactants, and (P) products. The equation provides an AFT¼ 1544K, which stratifies AFT4T1. The amount of exergy in combustion gases is highly affected by the chemical exergy of the components as given by the calculation steps given in Table 4; take this into account. However the chemical exergy of the components is not very high as there is no combustible material. A subject of interest is to calculate the entropy value of a mixture of an ideal gas. If one denotes yi, as the molar fraction of the components i ¼1…n then the partial pressures of the components can be given by Pi ¼yiP, where P is the P total pressure. Therefore, the term for pressure for the ideal has mixture entropy by accounting for ni¼ 1 yi ¼ 1 as follows: 3 2 n       P yi  n  yi  n X Pi yi P P n 7 6 P i¼1 n yi ln ð76Þ ¼ ln ∏ ¼ ln4 ∏ ðyi Þyi 5 ¼ ln ∏ ðyi Þyi P0 P0 P0 i ¼ 1 i ¼ 1 P0 i¼1 i¼1 Therefore, the entropy relation of a gas mixture at temperature T and pressure P becomes   n X P n yi yi s0i ðT Þ Rln sðT; P Þ ¼ ∏ yi P0 i ¼ 1 i¼1

ð77Þ

Heat Engines

159

Table 3 Case study of exergy destruction during real air expansion within Brayton cycle. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.) Calculation step

Assumptions

Equations

Calculated values

Reference state

T0 ¼ 298:15K P0 ¼ 101:325kPa

∅0 ¼ h T0 s0 ex0 ¼ 0 P0 v0 ¼ RT0

kJ ∅0 ¼ 40:5 mol m3 v0 ¼ 24:47 kmol kJ h0 ¼ 8649 kmol kJ s0 ¼ 165 kmolK

Parameters of state 1 (inlet)

T1 ¼ 1273K P1 ¼ rP0 r ¼ 15

P1 v1 ¼ RT1 h1 h0 ¼ Cp ðT1  T0 Þ   s1 s0 ¼ Cp ln TT10 Rln PP10

m v1 ¼ 6:964 kmol kJ h1 ¼ 39:5 mol J s1 ¼ 187:4 molK

s2s ¼ s1 ; P2s ¼ P0 ; g g P2s v2s ¼ P1s v1s ; g 1 g 1 ; v1 T1 ¼ v2s T2s h2s h0 ¼ Cp ðT2s ex2s ¼ h2s T0 s2s

m v2s ¼ 52:2 kmol kJ h2S ¼ 18:7 mol J s2S ¼ 187:4 molK kJ ex2S ¼ 3:36 mol

ex1 ¼ h1

Parameters of state 2s

Parameters of state 2

Exergy balance

CP ¼ f ðT ; onlyÞ

Zs;t ¼ 0:85

Process1

2 : Adiabiaticexpansion

T0 s1

∅1

P2 ¼ P0 ; ðh1 h2Þ ¼ Zs;t ðh1 ðT1 T2 Þ ¼ Zs;t ðT1 P2 v2 ¼ RT2 ;   s2 s0 ¼ Cp ln TT20

exd exom

¼1

kJ ex1 ¼ 24:2 mol 3

T0 Þ; ∅0

3

h2s Þ; T2s Þ;   Rln PP20

ex2 ¼ h2 T0 s2 ∅0 ; ex1 ¼ ex2 þ wt þ exd ; exin ¼ ex1 ex2 ; wt ¼ h2 h1 C ¼ exwint ; EDF ¼

3

m v2 ¼ 60:46 kmol kJ h2 ¼ 21:82 mol J s2 ¼ 191:98 molK kJ ex2 ¼ 5:12 mol kJ ; exd ¼ 1:36 mol kJ ; exin ¼ 19:0 mol EDF ¼ 7:13%

C

Another parameter and relation needed to be considered is the pressure variation along and isentropic process. This is done by assuming an isentropic process from process 1 to 2. Given that s1(T1,P2), the above equation could be utilized to find the following: "  n n X X P1 n P2 n 0 yi s0i ðT1 Þ Rln yi si ðT1 Þ Rln ð78Þ ∏ ðyi Þyi ¼ ∏ ðyi Þyi P P0 i ¼ 1 0 i ¼ 1 i¼1 i¼1 If the above equation were to be rearranged, one can obtain the relation between the initial pressure and final pressure (P1&P2) as reduced pressures. In the case where gas mixtures are introduced, the reduced pressure relation could be given as a single gas, as follows:   n P exp yi s0i ðT2 Þ=R P2 Pr i¼1  ð79Þ ¼ 2 ¼ n P P1 P r1 exp yi s0i ðT2 Þ=R i¼1

In Table 4 it can be
seen that at the end of the isentropic expansion T2s results from solving equation P Pr ;2s ¼ exp ni¼ 1 yi s0i ðT2s Þ=R iteratively. T2 is also found through an iterative process by solving the following equation: Xn y h0 ðT2 Þ h2 ¼ i¼1 i i

For T2. The results of the case study were analyzed comparatively in Figure 31. The temperature at the end of the expansion process is predicted within accepted accuracy when the varying specific heats of air are considered, as it is very close to the results found from the combustion gases. The standard air assumption is different from the mentioned method. As for exergy destruction, the standard air assumption is a reasonable prediction, for example, 7.6% under the standard air assumption and 7.05% with combustion gases. To conclude, this brief comparative study gives insight into how useful the standard assumption is, and if the varying specific heats are accounted for they give more accurate results. On another note turbine pressure ratio per stage in the standard air assumption where g¼ 1.4 is given by the following relation: P ¼ 0:528 ð80Þ Pin * where P is the pressure value in the throat, and Pin is the pressure inlet in each stage of expansion. As for the factor 0.528 it results from Eq. 80.

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Heat Engines

Table 4 Case study of exergy destruction during combustion gas expansion within Brayton cycle Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.) Calculation step

Assumptions

Equations

Calculated values

Reference state

T0 ¼ 298:15K P0 ¼ 101:325kPa

∅0 ¼ h T0 s0 h0 ¼ Syi hi0 ðT0 Þ s0 ¼ Syi si0 ðT0 Þ exch ¼ Syi exch i i ¼ CO2 ; H2 O; N2 ; O2 P0 v0 ¼ RT0

kJ ∅0 ¼ 77:7 mol m3 v0 ¼ 24:5 kmol kJ exch ¼ 1:998 kmol

Parameters of state 1 (inlet)

T1 ¼ 1273K P1 ¼ rP0

P1 v1 ¼ RT1  Q  s1 ¼ Syi si0 ðT1 Þ Rln PP10 yiyi     s1 ¼ Cp ln TT10 Rln PP10

m v1 ¼ 6:964 kmol kJ h1 ¼ 13:5 mol

r ¼ 15

ch

ex1 ¼ ex þ h1

Parameters of state 2s

CP ¼ f ðT ; onlyÞ

Zs;t ¼ 0:85

Process1

2 : adiabiatic expansion

kJ ex1 ¼ 26:5 mol 3

m v2s ¼ 52:2 kmol kJ h2S ¼ 18:7 mol J s2S ¼ 187:4 molK

P2 ¼ P0 ; ðh1 h2 Þ ¼ Zs;t ðh1 h2s Þ; h2 ¼ Syi hi0 ðT2 Þ;  Q  s2 ¼ Syi si0 ðT2 Þ Rln PP20 yiyi ;

m v2 ¼ 61:16 kmol kJ h2 ¼ 4:35 mol J s2 ¼ 227:8 molK kJ ex2 ¼ 7:37 mol

ex1 ¼ ex2 þ wt þ exd ; exin ¼ ex1 ex2 ; wt ¼ h2 h1 C ¼ exwint ;

kJ ; exd ¼ 1:35 mol kJ ; exin ¼ 19:17 mol EDF ¼ 7:03%

P2 v2 ¼ RT2 ; ex2 ¼ exch þ h2

Exergy balance

∅0

J s1 ¼ 223 molK

s2s ¼ s1 ; P2s ¼ P0 ; P2s v2s¼ RT2s ;   0  Sy s 0 ðT Þ Sy s ðT Þ P1 exp i Ri 1 ¼ P2 exp i iR 2s ; ex2s ¼ exch þ h2s

Parameters of state 2

T0 s1

3

EDF ¼

exd exom

¼1

T0 s2s

T0 s2

∅0

kJ ex2S ¼ 3:36 mol 3

∅0

C

Another reason for irreversibilities apart from the nonisentropic processes is that the turbomachinery experiences the existence in friction of bearings (Fig. 31). The mechanical energy conversion efficiency is within the same range as the steam turbine’s, which 0.9–0.95. The mechanical to electrical conversion efficiency is approximately 0.94–0.97. Heat exchangers located near the heat source such as heaters and reheaters have an irreversibility caused by the finite temperature difference that is occurring on the heat transfer surface. This situation occurs when an external source is used to heat up the heater and reheater. Let 1 and 2 be thermodynamic states of working fluid at the entrance and exit of the heat source heat exchanger. Where Tso is the average temperature of the heat source. With these notations the energy and entropy balances for standard air assumption are written in mass (or molar) specific form as follows: q þ sgen ¼ s2 Cp T1 þ q ¼ Cp T2 and s1 þ Tso Therefore, the entropy generation rate must be positive. The following constraint relationship must exist between the temperatures: T2 T1 ð81Þ Tso 4   ln TT21 A temperature difference must be available between the average heat source temperature and the working fluid where: DTso ¼ Tso

T2

ð82Þ

In most practical applications of the Brayton cycle the heat is provided internally and not externally. In such cases it is acceptable to assume that the source temperature is equal to that of the combustion gases at the exit stage. Therefore, Tso ¼ T2. If the case is Tso ¼ T2 it satisfies the above inequality of one take y¼ T2/T1 the condition sgenZ0 is equivalent to ylny þ 1Z0, which is true for any y40. Note that if y¼ 0 or T2 ¼ T1, sgen ¼ 0. The exergy destruction in the combustion chamber could be approximated with better accuracy when using varying specific heats and chemical exergy is obtained based on varying temperatures. Take Fig. 32 in consideration where state 1 is the preheated air at the air inlet, state 2 is the fuel inlet, state 3 is where the combustion has occurred and represents the temperature of the

Heat Engines

161

combustion gases. The molar fraction of combustion gases is denoted with yi . The total exergy of the fuel taken in gaseous form can be represented like so (Fig. 32):     P1 0 0 0 0 þ h ð T Þ h ð T Þ T s ð T Þ Rln ð T Þ ð83Þ s ex1 ¼ exch 0 1 0 0 f 1 f f f f P0 However, if the fuel is in liquid phase, then the pressure term is evaluated as zero. Therefore the total exergy of air in state 2 could be stated as follows:     P1 0 ex 2 ¼ ex ch h0air ðT0 Þ T0 s0air ðT1 Þ Rln ð84Þ s0air ðT0 Þ air þ hair ðT2 Þ P0 The total exergy expression for state 3 could be described as follows: X  X X   0 0 ex3 ¼ yi ex ch þ y h ð T Þ h ð T Þ T yi s0i ðT3 Þ i i 3 0 0 i i

s0i ðT0 Þ

Rln



P3 ∏yi P0



ð85Þ

The exergy destruction in the adiabatic combustion chamber is the result of the equation above. It can also be found from the _ d . The exergy destruction can be also defined based on entropy generation per mole balance equation n_ 1 ex1 þ n_ 2 ex 2 ¼ n_ 3 ex 3 þ Ex _ d ¼ T0 s_gen : Hence, the molar flow rate is denoted with n_ i , where i¼ 1,2,3. To the readers: note that the making the relation Ex W combustion gas temperature T3 ¼aft adiabatic flame temperature. Tatf can be found from the relation: X ð86Þ n_ 1 h0f ðT1 Þ þ n_ 2 h0air ðT2 Þ ¼ n_ 3 yi h0i ðTaft Þ Combustion chambers operate at high temperatures relative to their surroundings. This causes a large temperature gradient making thermal losses unavoidable. Because the actual combustion chamber don’t operate adiabatically the exergy destruction is higher than that in the ideal case. (Fig. 33(B)) illustrates a nonadiabatic combustion chamber.

1,Fuel

T3, P3

2,Air

1,Fuel 3,Combustion gases

T3, P3

2,Air

P3 = ΣyiPi

P3 = ΣyiPi Tdiss

Ex1 + Ex2 = Ex3 + Exd (A)

(B)

3,Combustion gases

QLoss = UA(T3 −T0)

Ex1 + Ex2 = Ex3 + QLoss (1−T0/Tdiss) + Exd

Fig. 33 Exergy destruction rate in a combustion chamber in two cases: (A) adiabatic combustion chamber and (B) nonadiabatic combustion chamber. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

The effectiveness parameter of a combustion chamber can be used to express the importance of the combustion gas temperature difference between adiabatic and nonadiabatic combustion chambers. If the adiabatic flame temperature is denoted as Tatf, T2 as the temperature at the combustion chambers inlet, and T3 as the temperature of the combustion gases at the exit. The effectiveness of the combustion chamber can be defined as the following: ecc ¼

T3 Taft

T2 T2

ð87Þ

Once the estimate of the effectiveness is available, the temperature of the combustion gases can be found from the above equation by inputting Tatf and T2 and solving for T3. As for the heat losses they can be determined: X _ loss yi h0i ðecc ðTaft T2 Þ þ T2 ÞÞ þ Q n_ 1 h0f ðT1 Þ þ n_ 2 h0air ðT2 Þ ¼ n_ 3

_ loss is calculated, the exergy destruction rate of the nonadiabatic combustion chamber can be determined with the use of After Q the exergy balance equation:   T0 _ loss 1 _ d n_ 1 ex 1 þ n_ 2 ex 2 ¼ n_ 3 ex3 þ Q þ Ex Tdiss

where Tdiss is the temperature at which heat is being dissipated from the component to the atmosphere. The temperature is similar to that of the casing of the gas turbine. The key in the thermodynamic analysis of a combustion chamber is assuming a reasonable value to define the effectiveness ecc. _ loss : The energy balance equation for an nonadiabatic system can be written as follows to find Q X 
_ loss ¼ n_ 3 Q yi h0i ðTaft Þ h0i ðecc ðTaft T2 Þ þ T2 Þ Š ð88Þ

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Heat Engines

The assumption is that at higher temperatures the combustion gases can be modeled as an ideal gas with a constant specific _ loss ¼ UAðT3 T0 Þ; where U represents the overall heat transfer coefficient and A is the heat heat. The heat loss can be written as Q transfer where leakages are occurring; to represent this relation the above equation becomes: UA½ecc ðTaft

T2 Þ þ T2

T0 Š ¼ n_ 3 CP ðTaft

ecc ðTaft

T2 Þ þ T2

T0 Þ

The above relation can be used to isolate and solve for ecc. This is applicable only if all the other parameters are available and specified. Furthermore, the effectiveness can be defined as a function of heat transfer units for the combustion chamber (NTU) and a dimensionless adiabatic flame temperature yaft. NTU yaft 1 ecc ¼ ð89Þ ðNTU þ 1Þðyaft 1Þ where the NTU and yaft are defined as follows: NTU ¼

n_ 3 Cp Taft and yaft ¼ T2 UA

T0 T0

ð90Þ

The effectiveness parameter of a combustion chamber can be used to express the importance of the combustion chamber’s effectiveness. With that the exergy destruction could be estimated. Therefore knowing the effectiveness operating range for a combustion chamber will allow for the analyses, parameter studies, and optimization of the combustion process. A simple analysis could be done to estimate the practical operation range as shown below. Assuming that atmosphere air as the reference temperature T0 ¼298K and the working fluid in the Brayton cycle analyzed is standard air. The typical operational compression ratio of the Brayton cycle is considered to be 16. If a stage compression process is used then r¼ 4/stage. the temperature at the end of the compression process can be determined by considering as an isentropic process where PTk ¼ ct. With k ¼ g g 1 and g¼ 14, r¼ 4…16 the air at the end of the compression process is considered to be 450–650K. The adiabatic flame temperature is considered to be a minimum of 1000K in this context and could range from 1500 to 2500K. Hence, the practical operational range for yaft can be estimated to be 4–10. Using the provided range for the combustion chamber effectiveness against NTU is plotted as shown in Fig. 34. 0.50

1.0

0.5

∝

(m)

0.10

aft = 4.8

aft = 6.8

aft = 8.8

0.2

0.01

0.1 0

1

10

100

0.00 1000

NTU Fig. 34 Exergy destruction rate in a combustion chamber in two cases: (A) adiabatic combustion chamber and (B) nonadiabatic combustion chamber. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

The number of thermal units (NTU) could be improved for the combustion chamber when the thermal insulation is thicker, therefore reducing heat transfer. The thermal conductivity ðk) of the insulation could be used to approximate the overall heat transfer coefficient by using U ¼ kδ, where δ is the thickness of the insulation layer used. Furthermore, going back to the NTU definition the thickness of the insulation can be found as follows: δ ¼ k An_ 3NTU Cp . The thermal capacity of the combustion chamber rises as the flow rate of the combustion gases rises. Furthermore, the size of the “exposed” area in the combustion chamber increases. The area can be found as follows: A ¼ ctn_ 3 and therefore the thickness becomes δ ¼ k NTU Cp . Using a varying range of NTU, k, Cp,δ¼ constant NTU with constant ¼ 0.015[m 1]. The insulation thickness is shown in Fig. 34. The figure displays typical values for thermal insulations (say from 1 cm to 20 cm assuming the combustion chamber’s power capacity ranges kilowatt to megawatt) the range of NTU is found to be ranging from 0.6 to 11 and the of the effectiveness ranging from 0.3 to 0.9. Taking the following case study for exergy destruction where these assumptions were made r ¼ PP30 ¼ 15, excess air has a value of 2.0, and methane as a fuel. From the previous analysis it is known that the combustion gases in state 3 have n1 ¼ 1 mol of natural

Heat Engines

163

gas preheated at T2 ¼ 800K. In state 3 there must be n3 ¼ [3 þ 7.52l þ 2(l 1)]n1 ¼ 48.6n1 amount of combustion gases available. The process in the combustion chamber is considered to be isobaric, where P1 ¼ P2 ¼ P3. The effectiveness has a large impact on the temperature of the combustion gases. The effectiveness of the combustion chamber ranges from 0.4 to 1.0. An important parameter while performing an exergy analysis is the temperature of the heat leakage from the combustion chamber to its surroundings. If the system boundary is set really far then Tdiss ¼ T0, otherwise the temperature of dissipation is much higher. The exergy destruction percentage can go up to 40%–50%, if the effectiveness is low (20%) when the operation is adiabatic. For ecc ¼ 0.9 the exergy destruction fraction is approximately 25%. Another important source that affects the exergy destruction in the combustion chamber is if the chemical reaction is not completed. This kind of irreversibility is highly dependent on the combustion chamber and some operational parameters that include excess air, combustion chamber effectiveness, and adiabatic flame temperature. The reaction when combustion is present is spontaneous and has high negative energy. A high negative energy is an indication that the reaction is evolving at nonequilibrium conditions. In a gas turbine, for example, fuel in the flue gas is always existent. This is one reason why heat recovery steam generators combined cycle power plants (CCPPs) in conjunction with practice gas turbines. The extent of reaction is a measure of the completion of the combustion reaction, the parameter varies from 0 to 1. Where 0 corresponds to no reaction and 1 is a complete combustion reaction. A general chemical reaction can described as: X X X R-ζ Ρ þ ð1 ζÞ R ð91Þ

where R represents the reactants and P represents the products, and ζA[0,1] represents the extent of the reaction. To the readers: note 1 represents a complete reaction. However, when the reaction is at equilibrium it will satisfy the inequality ζrζeqr1. The reaction’s free energy is given by DG¼ Gp GR, where GP represents Gibbs free energy of the products and GR represents the Gibbs free energy of the reactants. GP,R ¼ HP,R TSP,R where T represents the temperature of the reaction, which is usually taken as the product temperatures. If an adiabatic combustion reaction is taken into account then DG ¼DH TDS ¼ TDS where DH ¼HP HR ¼0. The adiabatic flame temperature and reaction entropy for an adiabatic combustion scenario can be found from the following relation: 8

P P P < ζÞ k nk hk Taft jR i ni hi ðTi ÞjR ¼ ζ j nj hj Taft jP þ ð1

P P P ð92Þ : DS ¼ ζ j nj sj Taft ; yj P jP þ ð1 ζÞ K nK sK Taft ; yk P jR i ni si ðTi ; yi P ÞjR If the reaction entropy is negative then the reaction is considered to be spontaneous. In most cases with combustion the entropy is negative. A numerical example is given next to illustrate how the adiabatic flame temperature is affected by the reaction effects and the reaction effects on exergy destruction. Assuming an adiabatic combustion chamber with methane as a fuel with the operating parameters similar to the previous example. With one parameter being different, which is the reaction completion is relaxed. Therefore, the reaction for the combustion reaction becomes: CH4 þ 2lðO2 þ 3:76N2 Þ-ζ ðCO2 þ 2H2 O þ 2ðl

1ÞO2 þ 7:52lN2 Þ þ ð1

ζÞðCH4 þ 2lðO2 þ 3:76ÞÞ

The total number of moles of products per mole of CH4 is highly dependent on the extent that the reaction took place and can be determined from the above equation: n ¼ ζ ð1 þ 3 þ 2ðl 1Þ þ 7:52lÞ þ ð1 ζÞð1 þ 2l þ 7:52lÞ ¼ 9:52l þ 1: The molar fractions exist in product gas as a function of the extent of reaction where: yCO2 ¼ 0:0499 ζ; yH2 O ¼ 0:0998 ζ; yCH4 ¼ 0:0499ð1 ζÞ; yO2 ¼ 0:1996 0:0998 ζ; and yN2 ¼ 0:7505: The reactants’ entropy and enthalpy can be calculated for the temperature of the methane T1 ¼ 400K and the temperature of air T2 ¼ 800K and the total pressure P ¼ rP0, with r¼ 15. The following is shown below: h i 8 > HR ¼ n1 h0f ðT1 Þ þ n2 0:21h0O2 ðT2 Þ þ 0:79h0N2 ðT2 Þ ¼ 9:76MJ=kmol < h i > : SR ¼ n1 sf ðT1 ; P Þ þ n2 0:21s0O2 ðT2 ; 0:21P Þ þ 0:79s0N2 ðT2 ; 0:79P Þ ¼ 998:3kJ=kmol K Using the combustion chamber effectiveness and the extent of reaction the enthalpy and entropy can be calculated. Fig. 34 displays a parametric study that shows the varying energy destruction depending on the combustion chamber effectiveness and the extent of reaction. The adiabatic flame temperature is at its maximum Tatf ¼ 1617K when the reaction is complete and it decreases to 822K when ζ ¼0.3 (Fig. 35). As for the exergy destruction fraction it is found to have a higher value when the effectiveness is lower. The exergy destruction calculated was also found to be lower when the extent of the reaction value was lower. These finding are logical, based on the second law of thermodynamics (Fig. 36).

4.5.7.5.1

Efficiencies of gas power plants

The schematic of an open-cycle gas turbine power plant is given in Fig. 36. The thermal efficiency of this plant may be expressed as _ net;out W ð93Þ Zth ¼ _ fuel qHV m

164

Heat Engines

0.4

1000

Taft (K)

EDF (%)

0.3

0.2 800 Taft cc = 0.5

0.1

cc = 0.75 cc = 1.0 0 0.3

0.5

0.4

0.6

0.7

0.8

0.9

1

 Fig. 35 An open-cycle gas-turbine engine. Reproduced from Dinçer I, Zamfirescu, C. Advanced power generation systems (n.d.).

Qin

Combustion chamber

2

3 Gas turbine/ generator

Compressor

Wc

Wt

Power shaft

Fig. 36 An open-cycle gas-turbine engine.

The exergy efficiency of this gas-turbine engine is Zex ¼

_ net;out W _ fuel ex fuel m

ð94Þ

This engine is sometimes modeled by a closed-cycle gas-turbine engine as shown in Fig. 37. The working fluid is assumed to be air and the combustion process is replaced by a heat addition process. In this cycle, the energy efficiency may be written as _ comp;in _ turb;out W _ net;out _ net;out W W W ð95Þ Zth ¼ ¼ ¼ _ air ðh3 h2 Þ _ fuel qHV _ air qin m m m Note that the heat added to the cycle is equal to the heat resulting from the combustion process. Eq. (95) is equivalent to Eq. (93). The exergy efficiency may be written by different approaches as
_ in ð1 T0 =Ts Þ _ net;out = Q ð96Þ Zex ¼ W _ net;out =ðm _ air ðh3 Zex ¼ W

h2

T0 ðs3

s2 ÞÞÞ

ð97Þ

Here, equations give different results for the exegetic efficiency values do not account for the exergy destruction during the combustion process while equations do not account for the exergy destructions during combustion and during the heat transfer to the working fluid in the cycle. For the idealized models of internal combustion engines (Otto, Diesel, and dual cycles), modified versions of may easily be obtained using the same principles. Simplified thermal efficiency relations of idealized cycles for internal combustion engines and gas-turbine cycles are available. When the Otto cycle is used to represent the operation of an internal combustion engine, the thermal efficiency under air-standard

Heat Engines

165

Qin Combustion chamber

2

3 Gas turbine/ generator

Compressor

Wc

Wt

Power shaft

1

4

Heat exchanger

Qout

Fig. 37 A closed-cycle gas-turbine engine.

assumptions (working fluid is air; air is an ideal gas with constant specific heats) is 1 Zth;Otto ¼ 1 rk 1

ð98Þ

where r is the compression ratio and k is the specific heat ratio. Similarly, the thermal efficiency of the Diesel cycle, which is the idealized model for compression ignition engines, is  k  1 rc 1 Zth;Diesel ¼ 1 ð99Þ k 1 r kðrc 1 where rc is the cutoff ratio, defined as the ratio of cylinder volumes after and before the combustion process. The efficiency relation for the dual cycle is   rp rck 1 1 ð100Þ Zth;Dual ¼ 1 k 1 krp ðrc 1Þ þ rp 1 r where rp is the ratio of pressures after and before the constant-volume heat addition process. The thermal efficiency of the simple Brayton cycle, which is the idealized model for gas-turbine engines, is expressed using the air-standard assumption as 1 ð101Þ Zth;Brayton ¼ 1 ðk 1Þ=k rp where rp is the ratio of maximum and minimum pressures in the cycle. For the idealized regenerative Brayton cycle, the efficiency relation is   T1 ðk 1Þ=k Zth;Brayton;regen ¼ 1 ð102Þ r T3 p where T1 and T3 are the temperatures at the inlets of the compressor and the turbine, respectively. The operational description of these idealized cycles may be found in most thermodynamics textbooks [3]. The above equations are only applicable to the idealized cycles considered, and they should not be used to determine the thermal efficiencies of actual internal combustion engines or gas-turbine cycles. The equations are useful in that they illustrate the effects of some key design parameters such as compression ratio, cutoff ratio, and pressure ratio on cycle efficiency.

4.5.7.5.2

Efficiencies of cogeneration plants

Cogeneration refers to the simultaneous generation of more than one form of energy product. For a cogeneration plant producing electric _ process , a first-law based efficiency is defined as the ratio of useful energy output to energy input: _ net;out and process heating Q power W Zcogen ¼

_ process _ net;out þ Q W ¼1 _ in Q

_ loss Q _ in Q

ð103Þ

_ process is the output rate of process heat and Q _ loss is the heat lost in the condenser. This relation is referred to as the utilization where Q efficiency to differentiate it from the thermal efficiency, which is used for a power plant where the single output is power. Students are consistently taught not to compare apples and oranges, which usually refers to two commodities that are different. Work and heat have the same units but are fundamentally difficult to add because they are different, with work being a more valuable commodity than heat.

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Heat Engines

We can overcome this situation by defining the efficiency of a cogeneration plant based on exergy, as the ratio of total exergy output to exergy input: Zex;cogen ¼

_ process _ net;out þ Ex _ out Ex Ex ¼ ¼1 _ in _ in Ex Ex

_ d Ex _ in Ex

_ process is the exergy transfer rate associated with the transfer of process heat, expressible as where Ex Z _ process ð1 T0 =T Þ _ process ¼ δQ Ex

ð104Þ

ð105Þ

where T is the instantaneous source temperature from which the process heat is transferred. This relation is of little practical value _ process and temperature T is known. In many cases, the process unless the functional relationship between the process heat rate Q heat is utilized by the transfer of heat from a working fluid exiting the heat producing device (e.g., a turbine or an internal combustion engine) to a secondary fluid in a heat exchanger. One can express the exergy rate of process heat as the exergy decrease of the hot fluid in the heat exchanger as _ process Ex

1

¼

_ hot ¼ m _ hot ½h1 DEx

h2

T0 ðs1

s2 ފhot

ð106Þ

h3

T0 ðs4

s4 ފcold

ð107Þ

or by the increase of the exergy of the cold fluid in the heat exchanger _ process Ex

2

¼

_ cold ¼ m _ cold ½h4 DEx

where the subscripts refer to state points in Fig. 27. The difference between these two exergies is the exergy destruction in the heat exchanger. Then, from Eq. (107), the exergy efficiencies based on these two approaches become Zex;cogen

1

¼

_ net;out þ m _ hot ½h1 W

h2 _ in Ex

T0 ðs1

s2 ފhot

ð108Þ

_ net;out þ m _ cold ½h4 W

h3 _ in Ex

T0 ðs4

s4 ފcold

ð109Þ

and Zex;cogen

2

¼

The exergy input in these relations can be expressed differently using various inputs as in the denominators of Eqs. (5, 6, and 7), yielding different exergy efficiencies.

4.5.8

Future Directions

Research and development are very important to increase the efficiency of the current heat engine technologies, which will have a great influence in reducing fuel consumption and reducing the overall operational cost of any power plant project that is operating using heat engines. Efforts are also made toward designing the heat engines in the simplest possible way to make them cheaper and lighter. In this section, advances in the heat engine technologies represented by the pulse detonation engine, which is mainly used as rocket engine; wave disk engine, which has the potential to substitute internal combustion engines in vehicles and provide a substantial improvement in hybrid electric vehicles; and the single atomic heat engine and utilizing renewable fuels as a source of chemical energy, along with renewable energy sources such as geothermal and solar energy in heat engines, will be discussed.

4.5.8.1

Pulse Detonation Engine

In this kind of engine, the detonation wave started by an ignition source is used to combust the supplied fuel-oxidizer mixture at which the detonation wave moves through the gas mixture and high pressure gas fills the detonation chamber. The detonation wave then exits the combustion chamber generating the required thrust for propulsion and air is drawn into the combustion chamber due to the pressure decrease that has occurred in the combustor. The mixture must be provided in the combustion chamber between each detonation wave and the next. Hypothetically, a pulse detonation engine can achieve higher thermodynamic limits compared to other internal combustion devices such as turbojets and turbofans; this is because of the fact that the detonation wave promptly pressurizes the supplied fuel–air mixture causing heat upsurge at constant volume. This technique can also lead to a decrease in the total weight and cost of the engine since some part such as the turbo-pump, which is an expensive part of conventional rocket engines, must deliver the fuel-oxidizer mixture into the combustion chamber at a tremendously elevated pressure reaching 140 bar unless the fuel is blown back out [2]. Numerous test bed engines have been constructed; a successful one was integrated into a lowspeed demonstration aircraft that was able to fly in 2008. Attempts have been made to use this type of engine in the hypersonic blackswift by the Defense Advanced Research Projects Agency, but the project was canceled in late 2008. The world’s first successful test for this type of engine was carried out by Russian scientists in 2016 using a clean fuel represented in the oxygen–kerosene pairing. The researchers announced that this type of engine achieved high thermodynamic efficiency and has the ability to carry a heavier payload and reduce the cost of shipping cargo into orbit. The experimental studies validated the hypothetical feasibility of using detonation engines in missile technology [4]. The possible future direction of this type of heat engine is carrying out extra experiments, enhancing the technology of this engine and making it more commercialized.

Heat Engines 4.5.8.2

167

Wave Disk Engine

This type of engine is a pistonless rotary engine that was constructed and developed at Michigan State University and Warsaw Institute of Technology. The engine is comprised of a rotating disk with curved blades. When supplying the fuel and air mixture into the engine, the revolution of the disk generates shockwaves that pressurize the air and fuel mixture. Combusting the mixture leads to its expansion and consequently forces the disk blades to rotate. The body of the disk will function as a gate that opens and closes the intake and exhaust ports. This concept is defined as a radial internal combustion wave rotor. The advantage of this engine is its design simplicity compared to conventional piston internal combustion engines. This engine is also predicted to use five times less fuel compared to conventional internal combustion engines due to its high efficiency, as it operates based on the most efficient, practical thermodynamic cycle, the Humphrey cycle, comprising the constant volume combustion and complete expansion in an ideal cycle achieving high efficiency compared to Otto, Diesel, and Brayton (gas turbine) cycles. Therefore, using this type of engine to charge the battery in hybrid vehicles would be very promising. This project also drew the attention of the U.S. Department of Energy’s Advanced Research Projects Agency–Energy (ARPA-e) and they offered a fund of $2.5 million for the research and development of this engine; the project is targeting to generate a 25-kW wave disk engine. The optimized versions are believed to achieve an energy efficiency of 65% [5].

4.5.8.3

Single Atom Heat Engine

Rossnagel et al. [3] introduced an experimental realization of a single-atom heat engine, at which an ion is confined in a linear Paul trap with tapered geometry and operated thermally by connecting it interchangeably with hot and cold reservoirs. The produced power of this engine is utilized to operate a harmonic oscillation. Moreover, from direct measurements of the ion dynamics, they were capable of defining the thermodynamic cycles for different temperature alterations of the reservoirs. These cycles were used to assess the power and the efficiency of the single atom engine. The maximum values of the power were recorded at 3.4  10–22 J/s and Z¼ 0.28%; these results were consistent with their analytical model. This type of system would allow the investigation and analysis of the performance of small quantum machines, the investigation of genuine quantum effects in thermodynamics, such as quantum coherences and correlations. Moreover, the estimation of quantum resource theory can be assessed, this research was carried out for better understanding of thermodynamics of single particles without any plans to use it in any applications. However, this understanding can result in a next generation of experiments that allows to apply this technique in future devices in different applications such as robotic parts or in single atom refrigerators.

4.5.8.4

Using Renewable Fuels and Renewable Energy Sources

Renewable fuels such as biofuels, bioethanol, and biodiesel are sustainable and can be used as a potential source of heat by combusting them and harnessing the heat to be used for power generation purposes. Biomass and hydrogen that is particularly produced from renewable sources can also be utilized in obtaining heat. Renewable energy sources such as geothermal and solar are very promising and globally growing at a promising rate and can also be utilized as a source of heat for power production. For instance using the geothermal source in single, double flash, and binary cycles to generate electricity. Solar heat can also be utilized in generating power in solar tower plants and in the power plants that are using heat from concentrated solar collectors to generate steam that operates steam turbines, which are coupled to electric generators. Other technologies are also under development and investigation such as osmotic heat engines for energy production such as the project funded by the U.S. Department of Energy [6]. This technology is able to convert the low heat temperature sources 40–1001C into power, which allow a beneficial use of the low temperature geothermal wells and enhanced geothermal system (EGS). The input to this kind of system is slow-grade heat and the output is the heat released to a cold reservoir and electric power from hydropower turbine. Applying this system will ensure the utilization of the geothermal low temperature sources in the best economic way and integrating with current high-grade conversion systems will definitely enhance the energy recovery, efficiency of the system, and reduce outputs costs.

4.5.9

Conclusions

In this chapter various heat engines have been discussed. The heat engines that use internal and external combustion were discussed thoroughly in terms of effectiveness, exergy destruction, and energy/exergy efficiency. Vapor compression cycles were also discussed and analyzed thermodynamically. The components of each of those cycles were analyzed individually on a thermodynamic level. The importance of these heat engines was emphasized throughout the chapter as they are crucial to the success of future societies. The future directions of heat engines were discussed, as new and upcoming technologies are arising.

References [1] [2] [3] [4]

DiPippo R. Second law analysis of flash-binary and multilevel binary geothermal power plants. Geotherm Resour Counc Trans 1994;18:505–10. Dincer I, Rosen MA. Exergy, energy, and sustainable development. 2nd ed. Oxford: Elsevier; 2013. Bejan A. Advanced engineering thermodynamics. 3rd ed. New York, NY: Wiley; 2006. Kanoglu M, Cengel YA, Dincer I. Efficiency evaluation of energy systems. New York, NY: Springer; 2012.

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Heat Engines

[5] Cengel YA, Boles MA. Thermodynamics: an engineering approach. 7th ed. New York, NY: McGraw-Hill; 2015. [6] Joshi AS, Dincer I. Solar based hydrogen production systems. New York, NY: Springer; 2013.

Further Reading Al Ali M, Dincer I. Performance assessment of integrated energy systems for HVAC applications. Int J Green Energy 2016;13:1342–51. Barron R. Cryogenic systems. New York, NY: Oxford University Press; 1997. Brzustowski TA, Brena A. Second law analyses of energy processes. IV – the exergy of hydrocarbon fuels. Trans Can Soc Mech Eng 1986;10(3):121–8. Buchmann I. Batteries in a portable world. 2nd ed. Richmond, BC: Cadex Electronics Inc.; 2001. Carmo M, Fritz DL, Mergel J, Stolten D. A comprehensive review on PEM water electrolysis. Int J Hydrogen Energy 2013;38:4901–34. Connelly L, Koshland CP. Two aspects of consumption: using an exergy-based measure of degradation to advance the theory and implementation of industrial ecology. Resourc, Conser Recycl 1997;19:199–217. Cornelissen RL. Thermodynamics and sustainable development: the use of exergy analysis and the reduction of irreversibility [PhD thesis]. University of Twente; 1997. Daily Mail. Russia reveals world’s first test of radical pulse-detonation super-rocket. Available from: http://www.dailymail.co.uk/sciencetech/article-3764051/Russia-reveals-worlds-test-radical-pulse-detonation-super-rocket.html; 2017 (accessed 20.10.17). Dincer I, Hamut HS, Javani N. Thermal management of electric vehicle battery systems. Hoboken, NJ: Wiley; 2017. Dincer I, Rosen MA. Exergy. 2nd ed. London: Elsevier Science; 2013. Dincer I, Zamfirescu C. Advanced power generation systems. New York, NY: Elsevier; 2014. Dincer I, Kanoglu M. Refrigeration systems and applications. 2nd ed. New York, NY: Wiley; 2010. Dincer I, Ratlamwala TAH. Importance of exergy for analysis, improvement, design, and assessment. WIREs Energy Env 2013;2.335–49. doi:10.1002/wene.63. Holland FA, Watson FA, Devotta S. Thermodynamic design data for heat pump systems. Oxford: Pergamon Press; 1982. Kestin J. Available work in geothermal energy. In: Di Pippo R, Khalifa HE, Ryley DJ, editors. Sourcebook on the production of electricity from geothermal energy. Washington, D.C.: U.S. Dept. of Energy; 1980. Klein SA. Engineering equation solver (EES), F-Chart Software. Available from: www.fChart.com.; 2006. Kotas TJ. The exergy method in thermal plant analysis. 2nd ed. Malabar: Krieger; 1995. Logan E. Handbook of turbomachinery. New York, NY: Marcel Dekker, Inc; 1995. Namisnyk AM. A survey of electrochemical supercapacitor technology. Sydney: University of Technology; 2003. Pulkrabek W. Engineering fundamentals of the internal combustion engine. 2nd ed. New York, NY: Prentice Hall; 2004. Pulse Detonation. Rocket engines 2005:2005–2005. Available from: https://www.nasa.gov/centers/marshall/pdf/173616main_pulse_detonate.pdf; 2017 (accessed 20.10.17). Rossnagel J, Dawkins ST, Tolazzi KN, et al. A single-atom heat engine. Science 2016;352(80-):325–9. doi:10.1126/science.aad6320. Szargut J, Morris DR, Steward FR. Exergy analysis of thermal, chemical, and metallurgical processes. New York, NY: Hemisphere Publishing Corp.; 1988. Wepfer WJ, Gaggioli RA, Obert EF. Proper evaluation of available energy for HVAC. ASHRAE Trans 1979;85(Part I):214–30.

Relevant Websites https://www.e-education.psu.edu/egee102/node/1941 e-Education Institute. https://energy.gov/eere/vehicles/advanced-combustion-engines Energy.GOV. http://physics.bu.edu/Bduffy/py105/Heatengines.html Heat Engines. http://hyperphysics.phy-astr.gsu.edu/hbase/thermo/heaeng.html HyperPhysics: Heat Engine Cycle. http://www.uni-mainz.de/presse/20212_ENG_HTML.php JGU Mainz. https://www.mahamasenergiindonesia.com/otec-power-plant-ocean-thermal-energy-conversion/ Mahamasenergiindonesia. http://www.nuclear-power.net/nuclear-engineering/thermodynamics/thermodynamic-cycles/rankine-cycle-steam-turbine-cycle/ Nuclear Power. http://www.renewableenergyworld.com/articles/2014/07/solar-fuels-how-close-how-real.html Renewable Energy World. http://www.solarheatengines.com/ Solar Heat Engines.

4.6 Stirling Engines Ramla Gheith and Houda Hachem, University of Monastir, Monastir, Tunisia Fethi Aloui, University of Valenciennes (UVHC), Valenciennes, France Sassi Ben Nasrallah, University of Monastir, Monastir, Tunisia r 2018 Elsevier Inc. All rights reserved.

4.6.1 4.6.2 4.6.2.1 4.6.2.2 4.6.2.2.1 4.6.2.2.2 4.6.2.2.3 4.6.2.2.4 4.6.2.2.5 4.6.2.3 4.6.2.3.1 4.6.2.3.2 4.6.2.4 4.6.2.4.1 4.6.2.4.2 4.6.2.4.3 4.6.2.4.4 4.6.2.4.5 4.6.2.4.6 4.6.2.4.7 4.6.2.5 4.6.2.5.1 4.6.2.5.2 4.6.2.5.3 4.6.2.5.4 4.6.2.5.5 4.6.2.5.6 4.6.3 4.6.3.1 4.6.3.2 4.6.3.3 4.6.3.4 4.6.3.5 4.6.3.6 4.6.3.7 4.6.3.8 4.6.3.9 4.6.4 4.6.4.1 4.6.4.2 4.6.4.2.1 4.6.4.2.2 4.6.4.2.3 4.6.4.3 4.6.4.4 4.6.4.4.1 4.6.4.4.2 4.6.4.4.3 4.6.5 4.6.5.1 4.6.5.2

Introduction Background and Fundamentals Invention Evolution 1936 – the Philips Stirling engine The Philips 1-98 engine From 1945 to 1975: Sleeping technology 1962 – Free piston engine Since 1975 – The rebirth Stirling Cycle & Thermodynamic Assessment Stirling cycle Thermodynamic assessment Stirling Engine Classification Simple or double acting Stirling engine Mono or multiphasic Stirling engine Resonant or not resonant Classification according to the cylinders arrangement Classification by pistons coupling Gas coupling (free-piston engines) Liquid coupling (Fluidyne) Applications Electricity production from solar energy Electricity generation from biomass Electricity generation from nuclear energy Residual electricity generation from natural gas Car propulsion Cold calories production Regenerator Global Optimization Characteristics Constituting Material Porosity Anisotropy Conception Effectiveness Effective Thermal Conductivity Permeability Pressure Drops Temperature Distribution Numerical and Theoretical Simulation of Stirling Engine Regenerator Global Theoretical Modeling Losses in Regenerator Internal conduction loss Loss due to regenerator imperfection er _v Friction loss in heat exchangers δQ Model Comparison Computational Fluid Dynamics Simulations Turbulent models Case 1: Computational fluid dynamics simulation of a 25 W Beta type Stirling engine Case 2: Computational fluid dynamics simulation of a 1 kW double acting type Stirling engine Experimental Studies of Stirling Engine Regenerators Figure of Merit Formulation One-Variate Regenerators Experimentation

Comprehensive Energy Systems, Volume 4

doi:10.1016/B978-0-12-809597-3.00409-0

170 171 171 171 172 172 172 172 172 172 172 173 176 176 176 176 176 177 180 181 181 181 181 181 182 182 184 184 184 187 187 187 188 188 189 189 189 190 190 190 190 191 191 191 191 192 192 194 200 200 202

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170

4.6.5.3 Experimental Design Methodology 4.6.5.4 Comparison Between Methods 4.6.6 Main Results and Discussion 4.6.6.1 Stirling Engines Performance Investigations 4.6.6.2 Stirling Engine Regenerator Investigation 4.6.7 Future Directions 4.6.7.1 Description of the Recovery Process Using the Stirling Engine 4.6.8 Conclusions Acknowledgments In Memory of Our Colleague, Professor Sassi BEN NASRALLAH References Further Reading Relevant Websites

Nomenclature A

N

Fluid passage section through heat exchanger (m2) Heat capacity of the cooling water (J kg 1 K 1) Hydraulic diameter (m) Diameter (m) Cycle time (s) Internal conduction loss in regenerator (W) Viscous friction loss in regenerator (W) Pressure drop (bar) Temperature difference between inlet and outlet cooling water (K) Electric efficiency (%) Mechanical efficiency (%) Thermodynamic efficiency (%) Efficiency (%) Darcy friction factor throw a porous medium Frequency (Hz) Friction factor through a porous medium Height (m) Permeability of a porous medium (m2) Dynamic viscosity of the working fluid (Pa. s) Mass flow of the cooling water (kg s 1) Mass flow rate through the heat exchanger (kg s 1) Rotational speed (rpm)

Indices g h hyd k k-r1 r

Working gas Heater Hydraulic Cooler Interface between cooler and regenerator Regenerator

Cp dhyd d dt _ Cdr δQ _ vr δQ Dp DT Zel Zm Zth e f f fr h K m _w m _ m

4.6.1

NUT Nu NUTr Nk P f Pr Q _ irr Q QL QH Re r rm T T0 TH TL u V

m r2-h r1-r2 wh wk

202 203 203 203 204 204 204 205 205 206 206 208 208

Number of transfer units Nusselt number Number of transfer units of regenerator Effective gas conductivity due to thermal dispersion related to of molecular conductivity Pressure (bar) Porosity Prandtl number Flow rate (m3 s 1) Effective thermal power exchanged in regenerator(W) Removed heat from the cold reservoir (J) Supplied heat to the hot reservoir (J) Reynolds number Density (kg m 3) Density of matrix solid material (kg m 3) Temperature (K) Temperature reference which is equal to ambient temperature (K) Temperature of the high-temperature reservoir (K) Absolute temperature of the low-temperature reservoir (K) Axial velocity (m s 1) Volume of the heat exchanger (m3)

Regenerator constituting material Interface between regenerator and heater Interface between first and second section of the regenerator Heater wall temperature Cooler wall temperature

Introduction

The industrial revolution in the 19th century led to both a great necessity to produce thermal energy and several environmental accidents. After the oil booms and the increase of oil price many industries closed and the industrial areas become brownfields (polluted areas that cannot be used for housing or for agriculture). All these events led to an increase in awareness about

Stirling Engines

171

environmental problems. Also, governments became more involved in climate change problems. Hence the resumption of interest in alternative solutions. The Stirling engines were invented in 1816, by Robert Stirling (see Relevant Websites section). The regenerator is considered as an added value of Stirling engine performances but they are the seat of an important part of thermal losses recorded in such engines. The regenerator is a porous medium used to economize heat for/from the working fluid [1,2]. All the exchanged heat energy passes through it and is proportional to its performances [3]. A Stirling engine without a regenerator needs five times more energy to produce the same performances as an engine including a regenerator [4]. According to Gheith et al. [3], a performant regenerator needs to have high thermal capacity [5] and conductivity [6], large surface area [7], small dead volume with dense matrix [8,9], and highly porous matrix with minimum resistance to flow [10]. Several studies treated numerically the heat exchange, the design, and the losses inside Stirling engine regenerator [10–13], but few experiments have been made. Stirling engine regenerators are very complex to model and design. It will require a large number of equations to describe their thermodynamic behavior (material side and working fluid side). The engine performances are more sensitive to a change in the regenerator efficiency and its ability to accommodate the high heat flow. The regenerator efficiency increase leads to an increase of the exchanged thermal energy through it and consequently an amelioration of engine brake power [14]. Tlili et al. [15] showed that the energy lost in the regenerator represents 86% of the total energy lost in the engine. Hachem et al. [14] showed that the regenerator is the set of 44% of viscous loss, 33% of internal conduction loss, and 22% of imperfection loss respectively from the total losses inside a Stirling engine. The determination of adequate material for Stirling engine regenerator is widely studied. Several materials have been used as Stirling regenerator [15–17], new material tested [18], and parameters influence detected [19–23].

4.6.2 4.6.2.1

Background and Fundamentals Invention

In the 19th century, the industrial revolution needs a great thermal energy production. The used technologies (especially steam boilers) have several technical problems causing harm to humans and damaging materials. A safe external combustion engine was the invention proposed by Robert Stirling to save human life and materials [24]. The initial installation is represented in Fig. 1. Robert Stirling proposed quickly an amelioration of his installation by adding an economizer, which enhances the heat exchange inside the installation [25,26].

4.6.2.2

Evolution

Initially, in 1816, the Stirling engine was used to pump water from a stone quarry. In 1820, James Stirling (Robert’s younger brother) introduced the working fluid at a pressure higher than atmospheric pressure in order to increase the output power of the engine. Until 1922, more than 25,000 Stirling engines were used either to pump water or to train electrical generators (see Relevant Websites section).

Q2 > 0

Q1 < 0 Engine

Hot source TH

Cold source TL

Wout < 0

(A)

Counter clockwise

Q1 < 0

Q2 > 0

Driven engine

Refrigerator

(B)

Win < 0

Heat pump

Clockwise

Fig. 1 Schematic view of Stirling machine working principle. (A) Hot air engine and (B) receiving machine.

172

Stirling Engines

Table 1

Maximum brake efficiencies for various Stirling engines

Engine designation

Prototype

4-235 Prototype

40 HP Prototype

Anal. Ph. I Prototype

4-400 Prototype

Manufacturer Working fluid Mean pressure (MPa) Heater temperature (1C) Cooler temperature (1C) Power (kW) Rotational speed (rpm) Thermodynamic efficiency (%) Carnot efficiency (%) Engine type Number of cylinder

United Stirling H2 14.5 691 71 35 2000 30 47 2 pistons 4

Philips He 22.1 683 43 175 1800 31 46 Piston-Displacer 4

Philips H2 14.2 649 16 23 725 38 55 Piston-Displacer 4

United Stirling H2 14.5 719 71 76 1200 35 54 2 pistons 8

MAN technology He 10.8 633 41 88 1000 32 49 Piston-Displacer 4

4.6.2.2.1

1936 – the Philips Stirling engine

The N. V. Philips Society of The Netherlands was the first great society interested in Stirling engine [27,28]. Table 1 below recapitulates geometric and functioning characteristics, power and efficiency of same engine realized by Philips society. The Philips society designed and manufactured a Stirling engine generator of radio. These radios can be used in isolation habitation, which is not connected to the public electricity network. Philips Electronics Group developed a Stirling engine to power portable radios. These radios were designed for use in remote areas without access to electricity.

4.6.2.2.2

The Philips 1-98 engine

Only 30 engines of this type were manufactured. The 1-98 engine includes three heat exchangers, a single power piston and displacer, and a Rhombic drive system. The major problem of this installation was its heat exchangers. A second version of the installation was proposed with ameliorated heat exchanger.

4.6.2.2.3

From 1945 to 1975: Sleeping technology

During this period, the Stirling engine lost its competitiveness compared to systems with the same power. It was used in some military applications including ships and submarines.

4.6.2.2.4

1962 – Free piston engine

William Beale invented the free piston Stirling engine, which has no mechanical losses compared to classical installation.

4.6.2.2.5

Since 1975 – The rebirth

Given the energy and environmental problems many companies from various sectors such as automotive, marine propulsion, or heating systems are rediscovering technology somewhat forgotten and adapting it to their own process. New Stirling installations are patterned and commercialized such as microgen installation (see Relevant Websites section), Stirling dish, etc.

4.6.2.3 4.6.2.3.1

Stirling Cycle & Thermodynamic Assessment Stirling cycle

The Stirling engine has the ability to work either in receiving mode or in driven mode (Fig. 1). The reversibility of the Stirling engine is an additional advantage. It can operate either in driven machine (produced work) or on receiving machine (heat transfer). The hot source can be from renewable energy (solar, recovered heat, biomass, etc.) and the cold source can be ambient air, water, or frigorific fluid (see Relevant Websites section). The Stirling engine is working in a closed cycle (Fig. 2). The working fluid trapped inside the engine can be air, helium, CO2, nitrogen, hydrogen, etc. It takes periodically four transformations, representing the Stirling cycle. 1. Isothermal compression process: the compression piston compresses the working fluid, so, the pressure increases. The temperature is maintained constant because of the heat flow from cooler to surroundings. 2. Isochoric regeneration (heat addition): the working fluid is transferred from compression space to expansion space through porous media regenerator. The working fluid is preheated in the regenerator. 3. Isothermal expansion: the expansion piston moves away from the regenerator. The pressure decreases as the volume increases. The temperature remains constant by adding heat to the system from the heater. 4. Isochoric regeneration (heat removal): both pistons move simultaneously to transfer working fluid from expansion space to compression space through regenerator at constant volume. The heat is transferred from the working fluid to the regenerator matrix.

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173

1 Q12 2

1

P

T

Q12 2 4

QR 4 3

Q34

3 Q34 V

S

Fig. 2 Clapeyron diagram (P–V) and entropic diagram (T–S). Theoretical cycle

P

***

Pmax

Experimental cycle

Pmin Vmin

Vmax

V

Fig. 3 Theoretical and experimental P–V diagrams.

The Stirling cycle has same advantage over the Carnot cycle [29,30]: (1) The replacement of two isentropic and two isochoric transformations, which increases the area of the P-V diagram. (2) Heating and cooling processes, which are improved by a regenerative porous media (see Relevant Websites section). The experimental Stirling cycle is different from the theoretical one (Fig. 3) [31]. An important difference between both cycles (experimental and theoretical) can be noted. These differences are essentially explained by:

• • • • •

Losses by friction and by singular pressure drop when the working fluid goes through heat exchangers. The real movements of the pistons are different from the theoretical movements. The great heterogeneity of the instantaneous temperatures in the thermal machine and to the irreversibility presents in the Stirling machine [32]. The regenerator cannot follow the temperatures variations from the hot side to the cold one, and vice-versa. Indeed, the temperature of the regenerator reaches that of the hot source and does not fall anymore because of its big thermal inertia. The loss of the pressure drops recorded in the regenerator is important. It is due to its geometric aspects (weak diameter) and to its physical characteristic (weak porosity).

Hachem et al. [31] studied a Beta Stirling engine working in receiving mode. Knowing the volumes and the pressure evolutions they represented the PV diagrams of the refrigerator (Fig. 4) modes. The network transmitted to the machine is about 3.513 J at 126 rpm of rotational speed. Stirling heat pump has COP ranging from 2 to 3.4, implying that they deliver 2 to 3.4 times more energy than it consumes. The best COP is obtained at the less rotational speed inputted in the machine (103 rpm), the heat has an energetic COP of about 3.4 and an exergetic COP of 0.75.

4.6.2.3.2

Thermodynamic assessment

1. Engine mode The working fluid (the thermodynamic system) provides output work δWo0. So, an amount of heat is added to the cycle by an external heat source and assigned to the cold source [33–35].

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1.9

× 105 (Refrigerating machine) Cycle n° 1

1.8 1.7

Pressure (Pa)

1.6 1.5 1.4 1.3 1.2 1.1 1 0.9 −20

0

20

40

60 80 Volume (cm3)

100

120

140

Fig. 4 P–V diagram at 126 rpm (refrigerating mode).

The theoretical thermal efficiency Zth of the Stirling engine is defined by: Zth ¼

Produced work δW ¼ amount of heat received by the working fluid δQ

ð1Þ

In practice, due to the various heat losses (imperfect insulation, radiation, etc.), an addition among of heat δQʹ is dissipated by the engine. With consideration of these losses the thermodynamic efficiency can be formulated as: Z0th ¼

amount of heat received by the working fluid δQ ¼ amount of heat gived to the working fluid δQ0

ð2Þ

Similarly, knowing the mechanical losses (especially mechanical friction), the mechanical work δWʹ effectively recovered is less than δW, the mechanical efficiency (or organic) can be defined as: Zmec ¼

Produced work δW 0 ¼ δW indicated work

ð3Þ

Consequently, a Stirling engine global efficiency can be estimated as follows: Zgol ¼

Produced work δW 0 ¼ ¼ δQ0 dissipates amount of heat

δW 0 δW δQ   ¼ Zmec :Zth :Zcal δW δQ δQ0

ð4Þ

Thermodynamic imperfections can be quantified as exergy destructions, which represent losses in energy quality or usefulness [31]. The exergy efficiency is presented to judge the ideality of a system. The exergy efficiency, also called the second law efficiency, is defined as the ratio of the actual thermal efficiency to the maximum possible (reversible) thermal efficiency under the same conditions: Z ZEx ¼ th ð5Þ Zc It can also be expressed as the ratio of the exergy recovered (as useful input) and the exergy supplied (as total input): ZEx ¼

Energy recovered ¼1 Energy supplied

Energy destroyed Energy supplied

ð6Þ

For a heat engine, the exergy supplied is the decrease in the exergy of the heat transferred to the engine, which is the difference between the exergy of the heat supplied and the exergy of the heat rejected. The network output is the recovered exergy. Finally,

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the exergy efficiency is calculated as follows: ZEx ¼

Wout Ex QH

ð7Þ

The exergy efficiency ZEx frequently gives a finer understanding of performance than the energy efficiency Zth. In evaluating efficiency, the same weight is assigned to energy whether it is shaft work or a stream of low-temperature fluid. The parameter ZEx points out that there is a need to deal with all external and internal irreversibilities to improve the machine performance [31]. Finally, the overall efficiency of the Stirling heat engine is calculated as follows: Zg ¼

Output work W ¼ ¼ Electric energy supplied Qin

W QH : ¼ Zth :Zcal QH Qin

ð8Þ

where Zcal is the calorific efficiencies of the hot heat exchanger (HEX). 2. Refrigerator mode When an external work is applied to a Stirling engine, this latter will work a refrigerator. It can provide very low temperature. The coefficient of performance (COP) of the refrigerator can be estimated as follows: COPR ¼

Desired output jQL j jQL j 1 ¼ ¼ ¼ Required input Win QH QL jQH =QL j

1

ð9Þ

The amount of heat received from the low-temperature reservoir QL during one cycle is defined as the sum of the inputted work and the amount of heat rejected to the high-temperature: QL ¼ Win þ QH

ð10Þ

The amount of heat rejected to the high-temperature reservoir QH during one cycle is calculated as follows: _ w Cp DTdt QH ¼ m

ð11Þ

The input work Win is described as follows: Win ¼

Wel Zel Zm Zth

ð12Þ

Additionally, the Carnot COP of a refrigerating cycle is defined as follows:
 QL TL ¼ ðCOPÞC ¼ QH QL C ðTH TL Þ

ð13Þ

The COPc is always greater than the COP of an irreversible refrigeration cycle when each operates at the same conditions. The thermal exergy for the refrigeration mode is defined as follows:
 T0 jQL j ð14Þ Ex QR ¼ 1 TL The exergetic coefficient of performance COPEx is calculated as follows: COPExR ¼

ExQR Win

ð15Þ

Referring to the exergy balance: X

Ex in ¼

X

Ex out þ

X

Ex D

ð16Þ

Thus, the amount of exergy destruction ExD in the Stirling refrigerator can be estimated as follows: ExD QR ¼ Win

Ex QR

ð17Þ

The exergy coefficient of performance COPEx frequently gives a finer understanding of performance than the energy COP. In evaluating COP, the same weight is assigned to energy whether it is shaft work or a stream of low-temperature fluid. The parameter COPEx points out that both external and internal irreversibilities need to be dealt with to improve the machine performance [36].

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3. Heat pump mode When an external work is provided to the Stirling engine and the flywheel is driven counterclockwise, the engine operates as a heat pump. The COP of the Stirling heat pump can be expressed as follows: COPHP ¼

QH QH ¼ ¼ Win QH QL 1

1 ðQL =QH Þ

The real heat pump COP is always less than the reversible Carnot efficiency given as follows:
 QH TH ¼ ðCOPÞC ¼ QH QL C TH TL The thermal exergy for heat pump mode is defined as follows:
 T0 Ex QHP ¼ 1 QH TH

ð18Þ

ð19Þ

ð20Þ

Then the exergy COP for heat pump can be calculated as: COPExHP ¼

Ex QHP Win

ð21Þ

Finally, energy and exergy formulations even for heat pump and refrigerating machine are applied to the experimental operating data obtained from the Beta Stirling machine.

4.6.2.4

Stirling Engine Classification

Since their invention by Robert Stirling, the Stirling engine has undergone several transformations. Three levels of categorization are generally used:

• • •

the cylinder composition the cylinders arrangement the pistons coupling In addition to these classifications, several particular Stirling engines are available.

4.6.2.4.1

Simple or double acting Stirling engine

The single-acting Stirling engines are constituted by a piston and a displacer, which can be included in the same cylinder or separated into two separate cylinders. The compression and expansion fields are in communication through a regenerator. The double-acting engines are essentially constituted of two or more pistons. In these engines, each piston acts as a displacer for the piston in its vicinity. The great advantage of this kind of arrangement is that the number of pistons is halved, which significantly reduces the cost of such engine. A schematization of the Stirling engine simple and double effect are presented in Fig. 5.

4.6.2.4.2

Mono or multiphasic Stirling engine

Mono or multiphase Stirling engines classification is obtained according to the thermodynamic state of the fluid, which may be two phases, for example, in the Fluidyne [37]. Fluidyne is the name given to a class of Stirling engines in which the pistons are actually columns of liquid (usually water) moving up and down in a set of U-tubes [37].

4.6.2.4.3

Resonant or not resonant

This distinction is only relevant for the free piston engines and Fluidyne. The resonant mode of operation corresponds to machines with a displacer and a piston moving continuously and, in most cases, sinusoidal. Or nonresonant mode (“over driven”) corresponds to operation in which the movement of the displacer and/or the piston is discontinuous. An ameliorated version of FSE was proposed by Boucher et al. [38]. They proposed a study of DFPSE working with helium. The DFPSE produces a mechanical power of 1 kW and it has a design operating point of 1.4 MPa corresponding to the frequency about 22 Hz.

4.6.2.4.4

Classification according to the cylinders arrangement

According to the kinematic arrangement of its different compartments, the Stirling engine can be classified in three major configurations: Alpha, Beta, and Gamma (Fig. 6). The Alpha configuration is the simplest building Stirling engine, characterized by its compactness. The Beta configuration is mainly composed of a simple piston placed coaxially with a displacer. This configuration presents serious sealing problems. The Gamma configuration is the oldest and the most cumbersome one. It is composed of two separate working spaces, which causes an important dead volume. This latter compromises the effective expansion temperature and hence the engine efficiency. The diversity of Stirling engine configurations is a constraint to their development and specifically to their standardization. However, this diversity has extended its applications to several areas: dish Stirling, wood boiler, electricity production, engine propulsion, etc. Three configurations can be mainly listed [11].

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Heater

Expansion space

Regenerator

Cooler Compression space

(A)

(B)

Fig. 5 Simple (A) and double (B) acting Stirling engine. Reproduced from Descombes G, Magnet JL. Moteur non conventionnels. Techniques de l’ingénieur, BM 2 593; 1997. p. 1–34.

C R H

H

H DP

R

C

C PP PP (A)

Alpha

PP (B)

DP

R

Beta

PP (C)

Gamma

Fig. 6 The three main configurations of Stirling engines, such as: C, Cooler; DP, Displacer; H, Heater; PP, Power piston; R, Regenerator. Reproduced from Wang K, Sanders SR, Dubey S, Choo FK, Duan F. Stirling cycle engines for recovering low and moderate temperature heat: a review. Renew Sustain Energy Rev 2016;62:89–108.

The Beta engine [39,40] includes the set of Stirling engines with a single cylinder pact and where the displacer and the power piston are linked in tandem. It is the most complicated configuration. The power is generated by the action of the pistons jointly. This type of configuration has sealing problems but has the advantage of compactness, lower dead space involved, and overlapping of volume. The Gamma engine [41] is very comparable to beta-group since the power output is generated in the same way as in Beta engines. The only difference is that the two pistons move in separate cylinders. This configuration has an important dead volume and produces a lower compression ratio, but it is usually simpler mechanically and is often used in multicylinder Stirling engines (Figs. 7 and 8).

4.6.2.4.5

Classification by pistons coupling

Three different types of pistons coupling are distinguished for Stirling engine:

• • •

the rigid coupling (kinematic engines) the gas connection (free-piston engines) the fluid coupling

1. The rigid coupling (kinematic engines): engines’ rigid coupling (or kinematic) uses a mechanical link between pistons. The main considerations for the choice of a coupling mechanism are the following: • Look for simple systems that are thus inexpensive in manufacture and maintenance. • Search systems allowing the highest possible seal, because one of the characteristics of the Stirling engine is operating at relatively high pressures.

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Hot volume Displacer

Cold volume Piston

Displacer rod (IR)

Piston rod ()

Crank case

6.

66

m

m

r A 2

B

1

r 3

4

Fig. 7 Schematic illustration of the Rhombic driven Stirling engine.

Compression space

Cooler

Fig. 8 Stirling engine with a crank and rod connection system. Reproduced from Gheith R, Aloui F, Tazerout M, Ben Nasrallah S. Experimental investigations of a Gamma Stirling engine. Energy Res 2012;36:1175–82.

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2. Crank and rod system: the most commonly used mechanism is the drive crank device. It is often used for transmitting rotary motion from a reciprocation one and allows a 90 degree phase shift between the pistons. This type of mechanism is used for smaller engines but has the major disadvantage of not allowing the dynamic balancing of a single cylinder engine [42]. Using the Ross yoke linkage (Fig. 9) is well known because of the high power-to-volume ratio. However, the thermodynamic analysis of this engine was so restricted according to a literature overview since it is reported to necessitate high temperature, but this problem was solved after the latest technological advances [42] (Fig. 10).

Expansion space Hot space piston Regenerator Fin cooler Compression space Cold space piston

Connecting rod Yoke mechanism Crankshaft

Fig. 9 The Yoke Ross driven mechanism. Reproduced from Tlili I, Musmar SA. Thermodynamic evaluation of a second order simulation for Yoke Ross Stirling engine. Energy Convers Manag 2013;68:149–60.

Ad Ve VC Ap



b2 b2

b2 sin x

 yC = x + b2 sin

yC = x.b2 sin

Fig. 10 Geometric derivation of the Yoke Ross drive. Reproduced from Tlili I, Musmar SA. Thermodynamic evaluation of a second order simulation for Yoke Ross Stirling engine. Energy Convers Manag 2013;68:149–60.

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Vd−2 Vs−2 A1 D2 N2

D1 L1

Dp

L2 A2 Vc



N1 Vs−1

Vh Vd−1 Regenerator

Fig.11 Schematic representation of the sliding disk mechanism (“swashplate”). Reproduced from Campos MC, Vargas JVC, Ordonez JC. Thermodynamic optimization of a Stirling engine. Energy 2012;44:902–10.

Fig. 12 Free piston Stirling engine.

3. The rhombic drive mechanism: the evolution of the Stirling engine for a wide range of power has led to isolate the cylinder crankcase and avoid pressurizing the entire crankcase. Rhombic command (Fig. 7) was developed by Philips for converting a linear motion into a rotary motion. This system must be designed and made with great care and good precision for satisfactory mechanical efficiency. It has the advantage of being dynamically balanced even for a single cylinder machine. The disadvantage of this movement is due to the large number of moving parts. 4. The splash plate: the “Swashplate” (Fig. 11) is a device used in mechanical engineering to translate the rotary motion of a shaft into a reciprocating, translating, or reciprocating motion to a rotating crankshaft to replace the engine to drawings. In this system each drive rod is connected to a rotating inclined plate, which generates the reciprocating movement of the pistons. The angle of inclination of the plate allows varying the stroke of the pistons and thus varying the power output. This system provides a good balance of the engine [43]. Meijer [44] proposed a method to control the engine performances by adjusting the swashplate angle.

4.6.2.4.6

Gas coupling (free-piston engines)

The free piston Stirling engine (Fig. 12) [45] has the particularity to convert the thermal energy in a directly usable energy. This configuration incorporates within a single, hermetically sealed transfer piston, while the free piston acts directly on the driven

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engine. Its efficiency is greater compared to the kinematic engines because of the absence of driving mechanism (crankshaft, crank rod) [46]. Free-piston machines are therefore characterized by lower cost, longer service life and minimal maintenance compared to kinematic Stirling engines. The rapid movement of the displacement piston has the advantage of bringing the actual cycle of the motor of the theoretical cycle of Stirling into effect. There are different types of free-piston machines among which the Martini machine, the Ringbom machine, and the Stirling free-displacement (FPFD) machine (see Relevant Websites section).

4.6.2.4.7

Liquid coupling (Fluidyne)

For this Stirling engine class, the pistons are connected by means of a liquid. In most cases, the pistons are themselves liquids. An interesting example is the motor called “Fluidyne,” which consists of oscillating liquid column [47]. It is a “pump-motor” based on the simple Stirling cycle. It is rustic, inexpensive, and easy to implement. The Fluidyne water pump is a machine for converting low-temperature thermal energy by bringing external heat (solar, rejects, etc.). Even if its thermal efficiency remains relatively low, this type of machine can be used as an inexpensive means of irrigation using an extremely simple technology with a minimum of moving parts [37].

4.6.2.5 4.6.2.5.1

Applications Electricity production from solar energy

Many companies are trying to revive the Stirling engine as a source of electrical energy (see Relevant Websites section). Some industries have developed Stirling engines to exploit solar energy reflected by a parabolic collector (Table 2). Among these contributions we find the system infinea [52] or solar dish [53]. Other precommercial prototypes (see Relevant Websites section) have been developed in the field of cogeneration of electricity and heat (CHP), the cogeneration units Cleanenergy [51], Whispergen [54], Stirling DK, and the “Sunmachine” gas or biomass [55]. All solutions are based on high-temperature sources. In the field of low-temperature differences, it is proved experimentally that the Stirling engine can operate with only a few degrees of temperature difference. But all the experienced engines remain miniature models [56–58]. Actually, only about 10% of the world’s electricity is generated by cogeneration. Exceptions are some European countries, such as Denmark and Finland, which have successfully expanded the use of cogeneration to 30%–50% of total electricity production in recent years [59]. Microcogeneration systems are reducing CO2 emissions, reducing the need for electricity transmission and distribution networks, and making beneficial use of local energy resources (e.g., through the use of waste, biomass, solar energy, thermal heat releases, and others).

4.6.2.5.2

Electricity generation from biomass

Stirling engine installation can be used to generate electricity from biomass. This installation is mainly for domestic use since the installation is silent and safe (no risk to have vibrations due to explosion). The CHP “Stirling DK” cogeneration unit (Fig. 13(A)) is studied by Obernberger et al. [60], who have designated a heat exchanger for a Stirling engine coupled to a biomass boiler. The heat exchangers consist of a series of U-shaped tubes surrounding the combustion chamber to favorize convective and radiative transfer. The CHP cogeneration unit of “Sunmachine” (Fig. 13(B)) is a wood pellet boiler that produces hot water and electricity for domestic use. It has an overall efficiency of about 90%, thermal power 4.5–10.5 kW and electrical power 1.5–3 kW. The Stirling wood boiler, shown in Fig. 11(C), is already sold in a thousand copies in the Northern European countries. It is a boiler that heats a house with wood pellets and produces the electricity of the house. This model has a power of 15 kW thermal (hot water) and 1 kW electric. However, the problem of particles emerging from the combustion of wood remains.

4.6.2.5.3

Electricity generation from nuclear energy

NASA and other space agencies use the Stirling engine to provide electrical power to satellites and space probes in addition to solar panels, which it helps to orient to maximize efficiency. Currently, NASA is conducting studies regarding the installation of a permanent base on the moon. To generate electricity at this station, NASA is considering a small nuclear power plant that generates heat for a free-piston Stirling engine. The cold source would consist of large radiators. The Stirling engine must have a good efficiency and especially a very good power/mass ratio. This last point is crucial in the space domain to limit the mass of fuel embarked by the launching rocket. Table 2

Performance and operating conditions of some Stirling engines used for thermoelectric conversion of solar energy

Stirling engine model

Power (kW)

Efficiency (%)

Working gas

Displacement (cc)

Rotational speed (rpm)

Number of cylinders

Refs.

Solo 161 STM 4-120 SES 4-95

10 25 25

– – Between 38% and 40%

He H2 H2

160 480 380

1500 2200 1800

2 4 4

[48,49] [50] [51]

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Fig. 13 Stirling engine cogeneration unit working with biomass. (A) Cogeneration unit “Stirling DK” ayant une puissance de 35 kW au Denmark (biomass, gasification) (Reproduced from Obernberger I, Carlsen H, Biedermann F. State-of-the art and future developments regarding small-scale biomass CHP systems with a special focus on ORC AND Stirling engine technologies. In: International Nordic bioenergy conference; 2003.). (B) Microcogeneration unit “Sunmachine” (biomass, pellets) (Reproduced from Crema L. et al., Development of a pellet boiler with Stirling engine for m-CHP domestic application. Energy, Sustain Soc 2011;15.). (C) Wood Stirling micro CHP boiler (Reproduced from Stirlingpowermodule. Available from: www.stirlingpowermodule.com.).

4.6.2.5.4

Residual electricity generation from natural gas

According to De Paepe et al. [61], cogeneration systems are attractive for residential use. Several industries have produced microcogeneration systems with Stirling engines fueled by heat due to the combustion of natural gas. Among them is the WhisperGen technology, which is a residual energy system. The WhisperGen unit (Fig. 14(B)) has been chosen as the “world’s largest residential combined heat and power facility.” It is a natural gas boiler that produces hot water for domestic use and electricity A Stirling double-acting engine is coupled to an alternator. The WhisperGen unit has been studied by several researchers such as Cacabelos et al. [62] who propose a dynamic model of the commercial microcogeneration unit. They analyze its dynamic behavior when the engine runs at different mass flows. The transient behavior of the cogeneration unit is studied in references [63,64]. They propose a thermodynamic model that provides both electricity and energy produced. The Stirling Thermal Motors industry, recently renamed STM Power Inc., manufactured a four-cylinder Stirling generator for cogeneration applications (Fig. 14(C)). The drive system of the Stirling engine is a swash plate. The Stirling engine’s HEX is adjacent to a natural gas combustion chamber with a power of 55 kW.

4.6.2.5.5

Car propulsion

The Stirling engine used as a means of propelling an automobile is part of the past (but maybe also part of the future). Indeed, the Philips company studied during the years 1940 to 1980 various applications of the Stirling engine. One of these was to equip a Ford Torino, but this test was not transformed and the project abandoned. The reasons are probably related to the difficulty of having an engine capable of rapidly varying its power and speed. In 1986, a report by NASA summarized the characteristics of the Stirling engines for the automotive industry that had been developed for this purpose, namely MOD I and MOD II, which correspond to two stages of development of a single engine. The

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Fig. 14 Micro-CHP units using biogas or natural gas. (A) Clean energy micro-CHP unit (Biogas) (Clean Enegy. Gasbox 901 data sheet. Published online.). (B) WhisperGen micro-CHP unit “natural gas” (Reproduced from Whisper Tech Ltd. Whispergen. Product specifications. Published online.). (C) STM Power Inc. Stirling generator (Reproduced from Conroy G, Duffy A, Ayompe L. Validated dynamic energy model for a Stirling engine m-CHP unit using field trial data from a domestic dwelling. Energy Build 2013;62:18–26.).

engine used pressurized hydrogen as the working gas. It was developed and manufactured as part of collaboration between NASA and Mechanical Technology Incorporated (MTI). The MOD II engine (Fig. 15) could achieve a performance of 38.5% (much higher than the internal combustion engine (ICE): currently 20%–25% on the road and 33% in the laboratory) for comparable power to the ICE (83.5 hp ¼ 62.3 kW). It consumes fuel with less emission than an ICE because it burns fuel outside the engine and continuously without explosion. It thus generates much less noise in operation. As a result, it is not necessary to use a catalytic converter or silencer on the exhaust line. Kockums developed its AIP engine during the 1980s. It has proved its worth on board the French submarine SAGA. The engine was subsequently installed on Swedish military submarines. With this system, the submarine is also able to recharge its batteries while remaining immersed. Indeed, the gases after combustion are at a pressure higher than that of the water. Contrary to what is required of an automobile engine, an aircraft engine operates almost constantly at constant power. In this case, the Stirling engine is truly in its prime. Its silence, compared to a traditional engine, can be an asset for both the passengers of the plane and for the residents. The low vibratory level of the Stirling engine also pleads in its favor. When you take altitude, the outside air lowers in temperature. This air is the cold source of the Stirling engine. There is therefore no loss of power when one rises in amplitude. This would make it possible to fly faster than with a traditional engine. The choice of fuel being broader, one could imagine a less volatile, less explosive, and less polluting engine.

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Fig. 15 The MOD II: Stirling engine used for General Motors’ car and Opel prototypes. Reproduced from Autoblog. Available from: https://www. autoblog.com/2009/07/08/blast-from-the-past-nasas-stirling-powered-amc-spirit/.

4.6.2.5.6

Cold calories production

The first domestic Stirling cycle refrigeration system was studied by Finkelstein and Polonski [65]. The Stirling type V refrigerator (VISR) was developed and tested by Le’an et al. [66]. Parameters such as energy consumption and COP are studied at different load rotational speeds and pressures of the Stirling machine. Ataer et al. [67] studied numerically a Type V Stirling refrigerator using air as a working fluid. They found that when the load pressure of the machine exceeds 2 bar, the COP of the refrigerator decreases. Similarly, Giannetti et al. [68] studied a Stirling machine using air as a working fluid. They found that for regenerator efficiency er ¼ 0.95 the COP would increase to 0.77 for the top and 0.81 for the low-pressure cycle. Otaka et al. [69] designed and tested a Stirling type Beta machine with 100 W capacity. They studied the effect of many parameters such as the ratio of the dead volumes, the working fluid, the ratio of the volume of compression to the volume of expansion, and the phase difference between the two pistons. And they found that the refrigeration produced by nitrogen was 28% lower than that produced by helium. Recently, Formosa et al. [70] studied the main losses in a free-piston Stirling machine as a function of geometric and operational parameters. According to their studies, the use of the diaphragm in place of the working piston and the displacer could be a solution to the gas leak. However, the membrane will provide high thermal insulation (Fig. 16).

4.6.3

Regenerator Global Optimization Characteristics

Despite its small size compared to the dimensions of the Stirling engine, the regenerator porous structure has an important impact on the overall efficiency and output power of the Stirling engine. However, regenerators are very complex to model. Thus, before the modeling step, it is very important to:

• • • • • • • • •

check thermophysical proprieties of the material constituting the regenerator determine optimal porosity value check the anisotropic propriety of the porous media choose the better conception identify its efficiency determine the effective thermal conductivity determine permeability quantify pressure drop check temperature inhomogeneity around the circumference of the regenerator

4.6.3.1

Constituting Material

The constituting material has a direct influence on the efficiency of the regenerator to store heat. Different materials are used as regenerator material such as stainless steel, copper, aluminum, Monel, graphite, ceramic, carbon fiber, etc. Many researches have

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Specific heat capacity (J/kg.K)

Fig. 16 250 We Philips Stirling Refrigerating Machine. Reproduced from WikiVisually. Available from: https://wikivisually.com/wiki/ Stirling_heat_engine.

Graphite

2000

Aluminum

1000 Thermoplastics 500 Glass

200

0.1

1 10 Thermal conductivity (W/m.K)

100

Fig. 17 Thermal conductivity and specific heat capacity of the materials used as regenerator in Stirling engines.

been made in order to identify the best regenerator material. Hofacker et al. [71] identified the best materials that can be used as regenerator in a Stirling engine, and they have classified them according to their thermal conductivity and heat capacity as shown in Fig. 17. The graphite was used for the first time as a regenerator in a Stirling engine by Hofacker et al. [71]. Graphite can exchange twice the amount of heat exchanged with the conventional materials used as regenerator in Stirling engines. The carbon fiber is used for the first time as regenerator material in Stirling engine by Bin Wan [18]. Carbon fiber has the characteristics of high temperature resistance, high thermal conductivity, and high corrosion resistance. The manufacturing costs are very low with long service life and the contact area between the air and the carbon fiber body is large. So, the heat transfer efficiency is high. Abduljalil et al. [9] developed an experimental study to identify a low-cost material for the regenerator with performance similar to those of the materials typically used. To do this, they used a porous ceramic support, a steel sponge, a kind of stainless

186

Stirling Engines

Fig. 18 Different regenerator materials: (A) ceramic catalyst support; (B) steel “scourers”; (C) stainless steel “wool,” and (D) wire mesh screens. Reprinted from Abduljalil AS, Zhibin Y, Jaworski AJ. Selection and experimental evaluation of low-cost porous materials for regenerator applications in thermo-acoustic engines. Mater Des 2011;32:217–28.

Porosity of 75%

Porosity of 85%

Porosity of 95%

Fig. 19 Photography of copper regenerator after 15 h of experimentation.

steel wool, and finally a wire mesh screen as shown in Fig. 18. They tested the performance of these regenerators according to the average pressure between 0 and 10 bar. They demonstrated that the increase of the pressure drops and the flow resistance depends on the porosity, or the regularity of the porous material and the way of disposition of the solid material. Their results show that the cellular ceramics may offer an alternative to traditional regenerator materials to reduce the overall system costs. Formosa et al. [20] presented theoretical and experimental studies of a GPU-3 Stirling engine. They showed that the conductivity of the material constituting the regenerator matrix has strong effect on Stirling engine performances. Gheith et al. [21] studied different regenerator materials and demonstrated that the Stirling engine regenerator performances are very sensitive to its material characteristics. They experiment four different regenerator materials [22,23]. These materials are: stainless steel, copper, aluminum, and Monel 400. The regenerator in Monel 400, stainless steel, and copper present the highest thermal efficiency and engine brake power. The presence of oxygen in the working fluid is a great handicap leading to the rapid oxidation of the material, and after the deterioration of its thermophysical properties and consequently the mechanical power of the Stirling engine. A photo of copper regenerator after 15 h of experimentation is presented in Fig. 19 and marks of oxidation can be clearly seen. Fig. 20 presents regenerators made with Monel 400 and aluminum. The aluminum regenerator has an acceptable thermal efficiency, and

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187

Fig. 20 Photography of the regenerator of (A) Monel 400 and (B) aluminum after 15 h of test (heating temperature under 5001C).

does not oxidize. However, its use is limited by its melting point. Gheith et al. [21] concluded that the stainless steel is the best material that can be used as Stirling engine regenerator working in this range of heating temperature (3001C to 5501C). According to Sadrameli [50] the use of classical material (aluminum, stainless steel, etc.) is not recommended for Stirling installation using high temperature (higher than 10001C). At this temperature level the regenerator material must be made from ceramic with very low thermal conductivity. This dictated the effects of radial conduction in the matrix to be considerable. The dimensionless parameter that reflects this effect in packed bed heat exchangers is the Biot number.

4.6.3.2

Porosity

The porosity of the regenerator, which represents the proportion of the void in the porous part, considerably affects the performance of the regenerator to store heat. In this context, Timoumi et al. [72] studied the effect of porosity on thermal energy losses. They demonstrated that the reduction of porosity increases internal conduction loss and viscous friction loss as well as the dissipations by shuttle effect. But, on the other hand, it reduces losses by external convection. According to Abduljalil et al. [9], the high porosity of the regenerator leads to a higher thermal relaxation loss. According to the experimental studies of Gheith et al. [21], the optimum porosity of a regenerator constructed of stainless steel is equal to 85%. This value was considered as the most suitable matrix porosity for maximizing the Stirling engine performances and minimizing heat and friction losses. Using finite volume numerical resolution, Costa et al. [73] studied the effect of various structures and porosities of metal grids on the alternating flow loss. The porosity of the regenerator is the key parameter for optimizing performance of a Stirling engine. It affects the hydraulic diameter of the regenerator, the total dead volume, the energy dissipation by pressure drops, and certainly the thermal efficiency of the regenerator and its specific exchange surface. According to Chen et al. [74], the use of porous material inside the Stirling engine leads to system stability. It becomes more stable to possible disturbances in the mass flow rate supply to hot fluid. The decrease in mesh porosity leads to the highest friction factor and pressure drop. In spite of the higher pressure drop we have better power and thermal efficiency because we have better heat transfer. Based on experimental studies of five stainless steel regenerators for different porosities it was found that the porosity of 85% provides the highest engine performances.

4.6.3.3

Anisotropy

The porous matrices are generally anisotropic. Several studies aim to determine the influence of this anisotropy on the hydrodynamic parameters of the flux flowing through the regenerator. Clearman et al. [75] experimented several types of annular porous matrices to confirm the importance of the anisotropy and the average pressure in such structures. They tested five porous structures: stacked 325 mesh screens, 400 mesh stacked screens, 400 mesh sintered screens, metal foam, alignment of micromachined nickel disks of 36–40 mm hole diameter. Recently, Tao et al. [76] studied numerically heat transfer in regenerator for different mesh types, materials and lengths of the porous layer. They were able to determine the anisotropic characteristics of a porous medium as regenerator in a cryogenic cooling system (PTR). The regenerator design was also deeply studied.

4.6.3.4

Conception

In order to slightly reduce the space and the weight of the engine, Eid [77] changed the classical regenerator by an alternative moving one installed inside a Beta engine cylinder. In this case, the displacer acts as a displacer and a regenerator at the same time. He concluded that a regenerator piston having a square mesh gives 20% more work and 10% more efficiency than the GPU3 engine. Andersen et al. [78] proposed a new conception of the regeneration matrix, divided into 3 sections. The end sections have a greater porosity and a larger wire diameter than the center section. Their new design has improved engine efficiency from 32.9% to 33.2%. Tlili et al. [79] studied a Stirling engine with a driving type Ross Yoke system designed for solar applications. They proposed a detailed analysis of the influence of regenerator parameters on the performance of a Stirling engine. According to them, increased porosity

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Stirling Engines

leads to the decrease of the friction coefficient and the pressure drop. The effectiveness of regenerator can be changed by changing the wire diameter and length, which in turn changes the wetted surface. Xiao et al. [80] assumed that pressure drop per unit length decreases as the length of regenerator increases. Dietrich [81] changed the design of the regenerator to overcome the negative effects of recirculation flow and in order to improve cross thermal conductivity of the regenerator material.

4.6.3.5

Effectiveness

Referring to the major researches about the regenerator heat exchanger in the Stirling machine, it can be concluded that the most important variation of temperature, pressure, losses, and flow in the Stirling engine are observed in the regenerator (porous media, complicated heat exchange, heated and cooled at the same time). The average air temperature of the working fluid inside the regenerator is lower than its constituting material wall temperatures. Thus, many researches aim to quantify heat transfer rate through it and to determine its efficiency as shown in Table 3. The regenerator effectiveness was one of the decision variables used in the multiobjective evolutionary algorithms proposed by Ahmadi et al. [85–87]. They proposed an optimal Pareto frontier in objective space. It was found that a final optimal solution with higher values of objectives (thermal efficiency and output power) could be achieved if the volumetric ratio and the effectiveness of the regenerator were increased. The regenerator is an imperfect heat exchanger. Number of wires and wire diameter of regenerator matrix are taken as decision variables used in parallel multialgorithms optimization by Luo et al. [88] to maximization of thermal efficiency and output power and minimization of power loss.

4.6.3.6

Effective Thermal Conductivity

To characterize the porous medium, the equivalent or effective thermal conductivity is an essential parameter to be determined. It depends on thermal conductivity of the constituting regenerator material kw, thermal conductivity of the working gas kg, Table 3

Researches from literature on the regenerator effectiveness

Author(s)

Engine type

Results

Ataer et al.

Free-piston Stirling engine

•

Cheng et al.

300 W Beta-type Stirling engine with cam-drive mechanism

• •

Chen et al.

A c-type twin power piston Stirling engine charged with helium

• • •

Gheith et al. Andersen et al.

500-W Gamma type Stirling engine 9-kW Stirling engine

• • •

Hachem et al.

Beta type Stirling refrigerator

•

Kato and Baba

Low temperature differential Stirling engine (LTDSE)

• •

Glushenkov et al.

Single-piston alternative to Stirling engines

•

The most important heat energy is transferred in the regenerator, which underlines its importance The regenerator effectiveness depends on porosity, permeability, wire mesh number and material. It is also dependent on the rotational speed When using a 120 wires mesh regenerator, the shaft power of the engine reaches 390 W at 1400 rpm with 1.21 kW input heat transfer rate (32.2% thermal efficiency) The efficiency of a moving regenerator depends on its material, matrices arrangement, matrix wire diameter, and regenerator filling factor The regenerator efficiency had the most prominent effect on global engine efficiency, while engine rotational speed had the greatest effect on the engine output power. The operating temperature ratio dependent on the thermal resistance of heating and cooling spaces, the effectiveness of regenerator, the working fluid mass, and the engine rotational speed The selection of the appropriate regenerator constituting material is based on the recorded efficiency Regenerator heat transfer rate is a very crucial factor to regenerator effectiveness especially at high engine rotational speed when the time for heat transfer is very short in each cycle A regenerator with low heat transfer rate will perform poorly at high engine rotational speed even though it has large heat capacity The regenerator porosity increases beyond a critical point the Stirling refrigerator performance decreases, due to increased external conduction and lack of thermal transfer with the working fluid The efficiency of the regenerator depends on the fluctuation of pressure The temperature distribution in the regenerator is a determinant parameter of its effectiveness The energy efficiency is highly sensitive to regenerator performance

Refs. [67] [82]

[17]

[21] [8]

[83]

[19]

[84]

Stirling Engines

189

regenerator porosity e, and the structure of the solid matrix. There are several calculation models of the effective thermal conductivity of a porous medium such as the effective conductivity is given by Zahi et al. [89] as follows: lr ¼ lg
e þ ð1

eÞ
lw

ð22Þ

and the regenerator equivalent thermal conductivity is given as follows: lg e þ lw ð2 eÞ lg ð2 eÞ þ lw e

lr ¼ lg

4.6.3.7

ð23Þ

Permeability

The permeability of a porous medium (denoted by K) characterizes the ability of the medium to allow flow there through of the fluid (liquid or gas). It depends on both the porosity and the geometry of the solid matrix. The first known permeability experiments were produced by Darcy [119] when introducing the viscosity m of an incompressible fluid in a tube containing a porous and homogeneous medium. By measuring the rotational speed and the two pressures P1 and P2 respectively upstream and downstream of the porous block, Darcy has shown that there is a relationship between these pressures, Section S, height h of the tube, and the flow rate Q, which is written: Q¼S

K P2 P1 h m

ð24Þ

Permeability can be calculated using Darcy Eq. (3). However, this law is valid only for very low Reynolds number. For high speeds, inertial effects are manifested by the appearance of the term patch Forchheimer. On the other hand, it is possible to evaluate the permeability through specific geometries of the medium, means porosity, and solid matrix characteristic dimension. Such as: Kozeny–Carman relation [90] for a simple porous medium geometry consisting of identical elements: K¼

d2 e3 36 C0 ð1 eÞ2

ð25Þ

where d is the diameter of the constituting particles of the porous matrix and C0 is a constant depending on the shape of particles (3.6oC0o5). Ergun relation [91] is established by considering a unidirectional flow within a porous column consisting of spherical particles of a diameter d. The column is subjected to a pressure gradient. This relation is similar to the previous relation in which C0 ¼ 4.16: K¼

d2 e3 150ð1 eÞ2

ð26Þ

When the porous matrix is modeled as a bundle of parallel capillary tubes, the expression of the permeability is: K ¼e

d2 n p d2 with e ¼ 32 4

ð27Þ

where n is the number of tubes per unit area perpendicular to the direction of flow and d is the diameter of the tube. When the permeability varies from one direction to another, the porous medium is said to be anisotropic. The permeability is characterized, in this case, by a permeability tensor.

4.6.3.8

Pressure Drops

Stirling engines are known to work under high working pressures that can reach 200 bar depending on the machine’s dimensions [92]. Given its porous structure, the regenerator is the set of an important pressure drop. Many researches are made to characterize flow through a porous medium and calculate pressure drop through it. Some friction factor correlations under different working conditions are shown in Tables 4–6.

4.6.3.9

Temperature Distribution

Temperature distribution in both sides of the regenerator matrix is not uniform due to nonconstant fluid temperatures distribution on both the cold and hot sides. Thus, the exchanged heat in the regenerator (porous medium) is not symmetrical. Several numerical and experimental studies deal with this problem of temperature inhomogeneity on the circumference of the regenerator. Experimentally, Gheith et al. [21] recorded temperature distribution in the regenerator of a 500 W Gamma type Stirling engine, using eight thermocouples located in different positions of the porous matrix of the regenerator and arranged in a symmetrical manner. Their results show the axisymmetric evolution of temperature reduces the efficiency of regenerator to store calories. Dietrich [81] assumes a temperature inhomogeneity around the circumference of the regenerator of a Stirling pulse tube (PTC). He concluded that regenerator temperature inhomogeneity depends on the temperature gradient, on the mass flow, and on the transverse thermal conductivity of the material constituting the regenerator.

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190

Table 4

Summary of friction factor correlations for monophasic flow through a porous medium

Author(s)

Friction factor correlation

Conditions

1

Ergun [91] Macdonald [93] Hicks [94]

6:8Re

Tallmadge [96] Lee and Ogawa [97]

6:25

• •

ð1

eÞ1:2 Re e3

0:2

ð1 e3

eÞ2 

1

þ 1:56Re

•

The gas in the expansion space and the heater is at the highest temperature, and the gas in the compression space and the cooler is in the lowest temperature The total mass of the gas engine is constant. The fluid used is assumed to be perfect for this we apply the equation ideal gas For the assumed linear temperature distribution in the regenerator, the effective regenerator temperature Tr is given by:

n

 þ 0:1

b¼1.8 (Smooth particles) b¼4 (rugged particles) 500oReo6000 1000oReo6000

ð32Þ

0.35oeo0.88 0.1oReo105

ð33Þ

n¼0.352 þ 0.1e þ 0.275e2 1oReo105

Adiabatic model

Quasi-steady model

The working spaces are assumed adiabatic The conditional temperatures caused by the discontinuity of those between workspaces is introduced, are also considered nonideal

• • • • •

Each compartment is considered a homogeneous entity represented by the gas mass m, its absolute temperature T, its volume v, and pressure P In this ideal model, the losses are neglected: while the pressure P is the same in all compartments

4.6.4

ð29Þ

ð31Þ

h T kÞ Tr ¼ ðT T ln Th k

•

eo0.8 Reo3000

ð30Þ

1

29:32Re

ð28Þ

Theoretical assumptions for models

Isothermal model

•

0:2

þ 125Re 0:5 þ 14
 1 e 150 4:2 þ e3 Re Re1=6

1000Re

Rose and Rizk [95]

Table 5


 e 150 þ 1:75 3 e Re
 1 e 150 þ b e3 Re

The heater and cooler wall temperatures are maintained constants at: TWh and TWk The working fluid temperature is different from those of the heat exchanger wall The heat exchanger temperature isn’t constant, those of the compartments interfaces too The regenerator is subdivided into two cells r1 and r2 The outflow direction is arbitrarily defined from the compression space toward the expansion one. Fig. 1 schematizes the distribution to be theoretically studied

The dynamic model takes also into consideration the frictional drag force, which gives rise to a pressure drop across each heat exchanger component as well as a corresponding flow dissipation internal heat generation effect

Numerical and Theoretical Simulation of Stirling Engine Regenerator

4.6.4.1

Global Theoretical Modeling

Three theoretical models are usually used. The isotherm analysis is the easiest to develop. It was proposed by Gustave Schmidt. The Stirling engine power is based on an ideal analysis. A simple correction factor is then used to deduce the real mechanical power from the ideal one. The second model is the adiabatic model that was developed by Finkelstein [65]. The third model is the quasistationary model. The major difference between this model and the adiabatic model is that the gas temperature is not equal to the temperature of the walls of the associated heat exchanger [98,99].

4.6.4.2 4.6.4.2.1

Losses in Regenerator Internal conduction loss

An important gradient of temperature between Stirling engine interfaces leads to an internal conduction loss through the regenerator. This letter is divided in two sections (hot one: near the heater and cold one: near the heater). This loss is calculated

Stirling Engines

Table 6

191

Comparison between theoretical models and experimental results

Engine

Isothermal model

Adiabatic model

Quasi-steady model

Quasi-steady model þ losses

Experimental results

Brake power (W) Thermal efficiency (%) Remarks

632.97

438.42

284.70

272.73

273.90

62.8

56.9

37.2

22.8

23

• • •

High power and efficiency Nonacceptable results poor model

• • •

High power and efficiency Nonacceptable results Poor model

Low power and efficiency Acceptable results Good model

• • •

• • •

Low power and efficiency Acceptable results Very good model

•

Low power and efficiency

within the two sections of the porous matrix as follows [100]: _ Cdr2 ¼ δQ

lm Ar2 ðTr2 h Lr2

Tr1

r2 Þ

_ Cdr1 ¼ δQ

lm Ar1 ðTr1 r2 Lr1

Tk

r1 Þ

ð34Þ

ð35Þ

The regenerator thermal conductivity depends on the porosity f, on the thermal conductivity of the working fluid lg and of the material lm. It can be calculated as [92]: lr1;r2 ¼ lg

4.6.4.2.2

lg f þ lm ð2 fÞ lg ð2 fÞ þ lm f

ð36Þ

Loss due to regenerator imperfection er

The regenerator material stores energy from the hot gas and restores it during its passage in the opposite direction toward the cooler. The amount of heat stored or restored by the regenerator matrix depends on the characteristics of the material. This loss can be illustrated by the regenerator thermal efficiency er, expressed as follows: er ¼

NTUr 1 þ NTUr

ð37Þ

_ irr , calculated as follows: So, the effective thermal power exchanged in the Stirling engine regenerator is Q _ irr ¼ ð1 Q

4.6.4.2.3

_r er ÞQ

ð38Þ

_v Friction loss in heat exchangers δQ

The fluid friction associated with the flow through the porous media will result in a pressure drop, Dp, across the three heat exchangers, which greatly reduces efficiency and power output of the engine. To evaluate pressure, drop through heat exchangers, we use this correlation: Dp ¼

2f muV Dd2hyd

ð39Þ

_ vr ) through the porous media due to pressure drop in the heater regenerator is expressed as follows: The energy power loss (δQ _ vr ¼ δQ

4.6.4.3

_ r Dp m rm

ð40Þ

Model Comparison

For the same conditions: Th ¼ 5001C, TK ¼ 303K, N¼ 390 rpm, and Pi ¼ 5 bar the following performances for a Gamma type Stirling engine are obtained.

4.6.4.4

Computational Fluid Dynamics Simulations

CFD simulation has gained important interest to treat engineering problems. CFD is concerned with numerical solution of differential equations governing transport of mass, momentum, and energy in moving fluids. In Stirling engine, CFD can be used to study flow, heat transfer, pressure, and losses inside the whole engine. For this, the geometry and the mesh model representing

192

Stirling Engines

the real installation are created and then initial and boundary conditions are defined. Finally, the movement of hot and cold room pistons are added to the simulation. CFD approach allows the heat transfer characteristics to be investigated in great detail. Results describe the behaviors of temperature, pressure and velocity in different spaces and especially in the regenerator. The effect of regenerator properties is investigated. The impact of the porosity on convective heat transfer inside the regenerator is deduced. A comparison between three regenerator materials is presented and the better material assuming good temperature gap between hot and cold rooms is deduced. Recently, studies based on computational fluid dynamics (CFD) optimize the Stirling engine performances and tested new installations (cogeneration, recovery systems, etc.). Zhigang Li et al. [101] simulated flow and heat transfer in a compact poroussheets heat exchanger. They found that the porous-sheets regenerator has 38%–51% lower total entropy generation rate, thus leading to less available work loss, contributing to higher power output and thermal efficiency compared to the conventional wire mesh regenerator. Jose Leon et al. [102] studied, with a very simple design and geometry, heat transfer characteristics of a b-type Stirling engine. It is found that impingement is the major heat transfer mechanism in the expansion and compression chamber and temperature distribution is highly nonuniform across the engine at any given moment. Wen-Lih Chen et al. [103] studied heat transfer characteristics of a twin-power piston g-type Stirling engine. They proved that using a moving regenerator can enhance the performance of a b-type Stirling engine. The moving regenerator acts as an effective thermal barrier between both working spaces causing the reduction in rates of heat input and output and increasing of the engine’s indicated power. Using three-dimensional numerical simulations, Costa et al. [104] characterize heat transfer and pressure drop phenomena through both stacked and wound woven wire matrix regenerator under different porosity and flow boundary conditions. They present new Nusselt number correlation equations applied to characterize and hence to optimize stacked and wound woven wire Stirling regenerator.

4.6.4.4.1

Turbulent models

The flow inside the Stirling engine is turbulent, hence the choice of a turbulence model. The turbulence modeling consists in representing either the influence of turbulence on the average flow (statistical approaches of the RANS type), or the influence of the unresolved scales on the solved scales (filtered approaches of the large eddy simulation: LES). RANS stationary models are used to obtain a good approximation of average values in industrial flows. The main RANS models available are listed in Table 7.

4.6.4.4.2

Case 1: Computational fluid dynamics simulation of a 25 W Beta type Stirling engine

A CFD model is developed to simulate flow and heat transfer across Beta type Stirling engine singularities. CFD approach allows the heat transfer characteristics to be investigated in great detail. Results describe the behaviors of temperature, pressure, and velocity in different spaces. The effect of regenerator proprieties is investigated. The impact of the porosity on convective heat transfer inside the regenerator is deduced. A comparison between three regenerator materials is presented and the better material assuming good temperature gap between hot and cold rooms is deduced. The geometry of a Beta type Stirling engine is represented in Fig. 21. Its geometric characteristics are given in Table 8. The Stirling engine is composed of three parts: the warm space, the porous medium, and cold space; the geometry is devised into three distinct parts connected when creating interfaces between the porous medium and hot and cold rooms. Air (ideal gas) is selected as working fluid and copper as regenerator material. Operating pressure is chosen to be 101,325 Pa. As boundary conditions, the regenerator is chosen to be porous region with adiabatic walls having two interfaces, one hot side and

Table 7

Statistical models of turbulence type RANS

Models

Advantages

Inconvenient

Spalart-Allmaras

Economic (Eq. (1)). Good for free flows, and flows on profile without detachments or large pressure gradients Robust, economical and relatively precise. Suitable for flows with large Reynolds number Derived by a rigorous statistical method. Good for moderately complex flows (jet impact, flow separation, and recirculation) Respect a physical constraint that can violate the model k-epsilon. Offers the same benefits as the RNG. Recommended in the case of turbomachines Model recommended for turbine engine problems (compare to realizable k-epsilon). The SST k-omega version consists of a transition between the standard k-omega model (developed for moderate Reynolds numbers and boundary layers) and the high Re version of k-epsilon when far from the walls The most physically complete model (transport and anisotropy of turbulence are taken into account)

Requires a higher resolution of the mesh at the borders (no laws on the walls) Mediocre results for complex flows (strong pressure gradient, rotation, and swirl) Limited by the hypothesis of isotropic turbulent viscosity Limited by the hypothesis of isotropic turbulent viscosity

Standard k-epsilon RNG k-epsilon Realizable k-epsilon

SST and Standard k-Omega

Reynolds stress model (RSM)

Requires a higher resolution of the mesh at the borders (no laws on the walls)

Requires more CPU time. The equations of momentum and transport of turbulence are closely related

Stirling Engines

193

36

1.1 1.2

2.5 11 30

(A)

(B)

(C)

Fig. 21 Beta type Stirling engine. (A) Experimental prototype. (B) Geometry. (C) Dimensions.

Table 8

Geometrical properties of the experimental engine

Hot space diameter [mm] Hot space height [mm] Displacer con rod length [mm] Displacer stroke [mm] Cold space diameter [mm] Cold space height [mm] Power piston con rod length [mm] Power piston stroke [mm] Regenerator inner diameter [mm] Regenerator outer diameter [mm] Regenerator height [mm] Regenerator porosity Maximum machine volume [cm3] Minimum machine volume [cm3]

60 49.2 100 48 60 35.7 197 48 22 60 59 0.79 300 160

one cold side. Hot space is chosen to have isothermal walls, maintaining at a constant temperature of 700K. The cold space walls are maintained at a constant temperature of 300K. And hot and cold pistons are chosen to be adiabatic. Fig. 22 represents the repartition of temperature obtained by the CFD simulation. Isothermal curves for two different crank angle values can be seen. The highest temperature is observed at the expansion space while the lowest temperature is recorded at the compression space. The temperature inside the regenerator increases linearly as we are closer to the hot space. During compression isotherm curves are entrained by the direction of the moving pistons. Temperature distribution inside the Stirling engine does not have a symmetric distribution due to turbulence. It is clearly seen that during expansion, the diffusion of the hot heat flux becomes less important than the distribution of cold heat flow. This can be explained by the fact that the velocity of the flow during the compression is much greater than during expansion. Instantaneous variation of temperature and pressure can be obtained by CFD simulation too (Figs. 23 and 24). Very precise variation can be observed. The pressure difference among the two spaces can be attributed to the internal pressure drop of the gas passage through the regenerator. Despite its ability to transfer heat, the porous structure of the regenerator also acts as a barrier to the gas following the compression and expansion spaces. This can be seen from the variation of pressure in Fig. 8. Here the pressure difference is defined as Dp¼ Pcompression Pexpansion. In the case with 10 Hz of operating frequency, the average frictional pressure drop in the regenerator is about 0.1 bar. Fig. 25 shows the 2D velocity vectors at different crank angles. It can be seen that velocity of the working fluid through the regenerator is higher than other compartments. At the beginning of the compression phase (y ¼ 36 degree), the velocity vectors are parallel to the symmetry axis. When the fluid particles are accelerated (y ¼ 180 degree), there is a strong deviation of the velocity vector. The behavior of working fluid particles

194

Stirling Engines

0 degree

300.00

403.00

90 degree

180 degree

Temperature (K) 506.00 609.00

712.00

270 degree

360 degree

815.00

Fig. 22 Temperature contour of Beta type Stirling engine during a computational fluid dynamics (CFD) simulation at different crank angle.

550

T hot space T regenerator T cold space

Temperature (K)

500 450 400 350 300 39.5

40

40.5

41

Time (s) Fig. 23 Average temperature evolution in each space vs. time.

can be seen. For y ¼216 degree, two recirculation zones in the regenerator are observed. This is explained by the separation phenomenon of the boundary layer and by the pressure drop generated by the porous block, which causes the formation of a vortex. The most important parameters of a Stirling engine regenerator are its material and its porosity. Both parameters can be investigated through a CFD simulation. The Nusselt number evolutions versus operation time for two regenerator porosities are presented in Fig. 26. The porosity increase leads to the working fluid velocity inside the porous region, which ameliorates convective heat transfer and consequently increases the Nusselt number. The copper, aluminum, and stainless steel temperature coefficients versus operation time are presented in Fig. 27. It can be seen that the aluminum is heating faster and the cooler is heating lower. So, aluminum is the faster material that attains the quasisteady state but it is the worst material that can separate between hot and cold spaces, which deteriorate the regenerator efficiency.

4.6.4.4.3

Case 2: Computational fluid dynamics simulation of a 1 kW double acting type Stirling engine

The double acting Stirling engine of the WhisperGen boiler was considered to simulate the flow and heat transfer inside the Stirling engine. Geometric characteristics of this engine are shown in Table 9. CFD simulation results show the variation of pressure, temperature, and velocity during a cycle. The effect of the regenerator porosity on the mechanical power was investigated.

Stirling Engines

195

P hot space P cold space

2.9

Pressure (bar)

2.7

Pressure drop

2.5 2.3 2.1 1.9 1.7 1.5 0.2

0.4 Time (s)

0.6

Fig. 24 Periodic evolution of the average pressure at hot and cold spaces at steady state when f¼10 Hz.

0 degree

y z x

90 degree

180 degree

270 degree

360 degree

Velocity (m/s) 0.00000

0.20000

0.40000

0.60000

0.80000

1.0000

Fig. 25 Velocity vector diagram of the fluid through the regenerator at different crank angle.

The double acting Stirling engine consists of four cylinders. The geometry of one cylinder is considered in the CFD simulation. The simulated geometry is devised into five different regions as shown in Fig. 28: Hot region (expansion space þ hot canals), regenerator and finally cold region (compression space þ cold canals). 1. Pressure, temperature, and velocity: the pressure in the regenerator remains as the mean pressure between the hot and cold chamber and it is represented by the green color (Fig. 29(A)). According to Fig. 29(B), it is clearly seen that the maximum temperature is in the hot chamber domain. The red color represents temperature around 1000K, the yellow color in the interface between the regenerator and the hot canals shows that the temperature rises in that side from 785K to 900K and on the other hand the interfaces between the regenerator and the cooler shows

196

Stirling Engines

Average Nusselt number in regenerator

6E+5

Poro=0.5 Poro=0.3

5E+5 4E+5 3E+5 2E+5 1E+5 0E+0 37

37.1

−1E+5

37.2

37.3

37.4

Time (s)

Fig. 26 Average Nusselt number evolution vs. time for different regenerator porosity.

Aluminium

75

Stainless steel

Temperature coefficient

70

Copper

65 60 55 50 45 40 35 30 9

9.5

10

Time (s) Fig. 27 Temperature coefficient evolution versus time for different regenerator material.

Table 9

Geometry values of input data for the simulation

Geometry parameters

Values

Engine piston number Piston diameter (cm) Displacement volume (cm3) Expansion space volume (cm3) Compression space volume (cm3) Regenerator volume (cm3) Expansion space phase angle advance (degree) Rotational speed (Hz) Wire matrix diameter (m) Matrix screen width (m)

4 4.34 64 33.85 30.11 22.11 180 25 6E-5 2.16E-04

the existence of light blue color, which is in the middle between the dark blue color of the cooler temperature and the green color of the regenerator, which shows that the heat exchange between the regenerator and the cooler is taking place in the simulation, thus the regenerator works as a heat exchanger between the two chambers not only to prevent thermal shock but also to fasten the heat process of the hot chamber to get hotter and of the cold chamber to get cooler, which will end up raising the engine efficiency.

Stirling Engines

197

Fig. 28 Geometry investigated for the numerical computational fluid dynamics (CFD) simulations. (A) Expansion space boundary. (B) Hot canals boundary. (C) Regenerator boundary. (D) Compression space boundary. (E) Cold canals boundary. Reproduced from Gheith R, Frikha M, Hachem H, Aloui et F, Ben Nasrallah S. Simulation CFD des échangeurs de chaleur dans un moteur Stirling à Double effet, Proceeding Journées Internationales de la Thermique, JITH 2017, Monastir, Tunisia; 2017.

Velocity (m/s) Pressure (bar) 33.129

4.3136

Temperature (K) 1100.0

3.4509

942.60

31.907

2.5882

785.20

30.686

627.80

29.464

1.7255

470.40 28.243

0.86273 313.00

27.021 (A)

0.00000 (B)

(C)

Fig. 29 Distribution of (A) pressure, (B) temperature and (C) velocity. Reproduced from Gheith R, Frikha M, Hachem H, Aloui et F, Ben Nasrallah S. Simulation CFD des échangeurs de chaleur dans un moteur Stirling à Double effet, Proceeding Journées Internationales de la Thermique, JITH 2017, Monastir, Tunisia; 2017.

The working fluid velocity in the canals are high, presented by the red color, because the canal is too small and the heating process takes less time so the temperature increases, which increases the pressure and results in increasing the working fluid velocity in the canals. The working fluid velocity decreased from 3 m/s in the canals part to 1 m/s in the regenerator pores and 0 m/s in other point of the regenerator (Fig. 29(C)). Four vortexes were formed due to the incoming fluids from the canals with high velocity and temperature spread in the expansion hot chamber, in fact the hot fluid (1100K) is mixed with the coming cold fluid from canals. However, in the compression cold chamber, there are only two vortexes, in fact the fluid leaves the regenerator at about 670K of temperature and starts the cooling process. The change of surface between canals and chambers causes not only pressure drop but also turbulence in the fluid flow, which leads to the formation of vortexes. The vortex apparition enhances the heat transfer between walls and fluids.

Stirling Engines

8

600

7

Temperature (K)

550

6 500

5 4

450

3

400

2 350

Pressure drop (bar)

198

1

300 0

60

120 180 240 Crank shaft angle in degree

300

0 360

Fig. 30 Regenerator velocity and pressure drop evolution with crank shaft angle (∆f¼100 degree). Reproduced from Gheith R, Frikha M, Hachem H, Aloui et F, Ben Nasrallah S. Simulation CFD des échangeurs de chaleur dans un moteur Stirling à Double effet, Proceeding Journées Internationales de la Thermique, JITH 2017, Monastir, Tunisia; 2017.

Temperature 600

12

540

10 8

480

6

420

4

Temperature (K)

Velocity in the regenerator (m/s)

W velocity vector 14

360

2 0 0

60

120 180 240 Crank shaft angle in degree

300

300 360

Fig. 31 Regenerator velocity and temperature evolution with crank shaft angle (∆T¼ 126 degree). Reproduced from Gheith R, Frikha M, Hachem H, Aloui et F, Ben Nasrallah S. Simulation CFD des échangeurs de chaleur dans un moteur Stirling à Double effet, Proceeding Journées Internationales de la Thermique, JITH 2017, Monastir, Tunisia; 2017.

2. Regenerator temperature and pressure drop: Fig. 30 shows that the maximum regenerator temperature corresponds to the minimum pressure drop. In fact, pressure drop is highly related with energy dissipation, between expansion and compression phases the flow changes its directions, velocity in that moment equals zero, and hence the pressure drops equal zero, at that moment the viscous dissipation related to the flow movement is equal to zero, which means the temperature rises as the velocity decreases. The pressure drop is more important in the middle of expansion and in the middle of compression phase, at these two moments the regenerator pressure reaches a maximum value. 3. Regenerator temperature, pressure drop and velocity: during the expansion phase between 0 and 180 degrees the pressure in the hot chamber increases, and its volume decreases. Thus, the regenerator starts stocking the heat coming from the fluid going through it from hot chamber through canals to the regenerator. During the compression phase, as the piston start to move from 180 to 360 degrees the regenerator exchanges heat with both sides, with cold chamber in order to prevent thermal shock of the engine and with the hot side preparing for the next cycle in order to save energy, and that’s why the regenerator temperature decreases (Fig. 31). This phenomenon is similar to the electrical behaviors of capacitors while charging and discharging due to presence of electrical resistances in the electrical circuit. In the Stirling engine a similar phenomenon occurred in the regenerator due to presence of two resistances: the viscous resistance and the inertial resistance (Fig. 32).

14

8

12

7 6

10

5

8

4 6

3

4

2

2 0

199

Pressure drop (bar)

Velocity in the regenerator (m/s)

Stirling Engines

1 0 0

60

120

180

240

300

360

Crank shaft angle in degree Fig. 32 Regenerator velocity and pressure drop evolution with crank shaft angle (∆f¼6 degree). Reproduced from Gheith R, Frikha M, Hachem H, Aloui et F, Ben Nasrallah S. Simulation CFD des échangeurs de chaleur dans un moteur Stirling à Double effet, Proceeding Journées Internationales de la Thermique, JITH 2017, Monastir, Tunisia; 2017.

Table 10 porosity

Table 11

Regenerator inertial resistance and viscous resistance at different

Porosity

Inertial resistance Rvz

Viscous resistance Riz

0.65 0.725 0.85 0.95

1.8586e þ 10 8.2688e þ 09 1.5266e þ 09 1.2149e þ 08

7.4344e þ 04 4.2095e þ 04 1.4248e þ 04 3.4019e þ 03

Comparison between figure of merit (FOM) formulations from literature

FM1 (Eq. (43))

FM2 (Eq. (44))

FMSC (Eq. (45))

Includes the heat transfer and the thermal dispersion (denominator terms)

Includes the thermal energy transport (enthalpy þ dispersion) and the pumping power

Does not include the regenerator flow area. Inclusion of the dead volume ration as defined below:

4. Effect of regenerator porosity: based on Ergun equation (Eq. (28)), we are able to introduce the two principal parameters for the porous media simulation, which are the viscous resistance and inertial resistance as follows:

Ri ¼

Rv ¼

ð1

eÞ2 150 e3 Dp2

ð1 eÞ 1:75 e3 Dp

ð41Þ

ð42Þ

When varying only porosity and making all other parameters constant, the inertial resistance and the viscous resistance of the regenerator are calculated according to Eqs. (41) and (42) as shown in Tables 10 and 11. Effect of regenerator porosity on output mechanical power of the double acting Stirling engine is investigated. Fig. 33 shows that the highest output mechanical power (about 1.4 kW) is obtained when the regenerator porosity is about 0.775. The regenerator porosity and matrix wire diameter are critical parameters for the Stirling engine performance. High matrix porosity values lead to high regenerator effectiveness but low engine performance. An increase of the matrix heat capacity leads to the reduction of the engine’s performance, the torque as well as the work output. The regenerator temperature profile of the working fluid is related to position and time. The nonideal behavior of the regenerator is demonstrated and the remaining heat in the regenerator grid is recorded when the thermal equilibrium is achieved.

200

Stirling Engines

1600

Engine power (W)

1400 1200 1000 800 600 400 200 0 0.6

0.7

0.8

0.9

1

Porosity φ Fig. 33 Evolution of the power output with the porosity for Pi ¼20 bar, Th ¼1100K, and Tk ¼313K.

4.6.5 4.6.5.1

Experimental Studies of Stirling Engine Regenerators Figure of Merit Formulation

The figure of merit (FOM) is considered as a powerful tool to determine adequate Stirling engine regenerator. The FOM allows comparing different ratio for the tested regenerators at the same time. A regenerator with a FM ¼ 1 is an ideal regenerator (no losses). Ibrahim et al. [105] proposed a regenerator steady based on FOM formulations. They found that the microfabricated regenerator has a higher FM than woven screen or random fiber matrices. The FM is in the range of 0.07–0.42 when the Reynolds number is 10–1000. The FOM is useful to compare several regenerators for the same Stirling engine and does not track overall engine efficiency. Initially the formulation of the FOM was proposed by Gedeom [106–108] (Eq. (43)), then it was ameliorated by Ibrahim et al. [105] (Eq. (44)) FM1 ¼

F M2 ¼

1 Nk þ Re:Pr

ð43Þ

1 fDðRePrþ Nk Þ

ð44Þ

RePr 4Nu

4Nu

Re:Pr

Recently Costa et al. [109] reconsidered these expressions and proposed a new formulation including the dead volume ratio as defined below: FMSC ¼ 

1 _t DP Vdr Q _r Pm Ve Q



ð45Þ

The FOM does not track overall engine efficiency so closely when comparing regenerators of different matrix structures. The Nasa/Sunpower installation tested several materials as regenerator (see Relevant Websites section). The highest FOM was given by the microfabricated regenerator (Fig. 34). According to authors the FOM is the largest and most comprehensive method to judge the regenerator performances. For a Reynolds number between 10 and 100, the FOM is in the range of 0.07–0.42 for a microfabricated regenerator. The highest value can be rich for Reynolds number about 400. Gheith et al. [110] studied a Gamma type Stirling regenerator (Fig. 35) based on FMSC formulation. They proposed a first study to determine the adequate material (aluminum, copper, Monel 400, and stainless steel) and then to determine the adequate porosity (75%, 80%, 85%, 90%, and 95%). The regenerators with 80% and 85% of porosity has, respectively, a FMSC of 0.3 and 0.29. These two matrices present the best porosities of a Gamma Stirling engine regenerator. The FMSC was estimated for the stainless steel matrices of 80% of porosity (Figs. 36 and 37). Different stainless steel regenerators are manufactured with different porosity. The volume ratio increases with porosity due to the increase of its corresponding dead volume. This ratio is a geometrical parameter depending on the volumetric porosity of the regenerator. In our case the expansion space volume (VE) is constant for all experiments. The losses ratio (Qloss/Qtot) decreases initially with porosity until a minimal value for a porosity of 85% the increases with the porosity. The following Table 12 summarize the effect of some losses.

Stirling Engines

201

Parallel plates theoretical

0.5

0.4

Randomfibers 90% porosity

Figure of merit FM

Mezzo involutemicrofab

Etched foil 2A sample

0.3

Randomfibers 96% porosity

0.2 Screens 70% porosity

0.1 Packedspheres 39% porosity

0 10

100 Reynolds number

1000

Fig. 34 Figure of merit (FOM) of different regenerator materials, porosities, and Reynolds number. Reproduced from Timoumi Y, Tlili I, Ben Nasrallah S. Performance optimization of Stirling engines. Renew Energy 2008;33:2134–44.

TR1

TR5

TR2

TR6

Porous

TR7

TR3

TR8

TR4

Fig. 35 Gheith et al. Stirling regenerators. Reproduced from Gheith R, Aloui F, Ben Nasrallah S. Investigation of regenerator matrix through figure of merit analysis, AJKFluids2015-22087, pp. V001T22A001. doi:10.1115/AJKFluids2015-22087; 2015.

0.7 0.6 0.5 0.4 0.3 0.2 0.1 0

Stainless steel Dp/Pm

Cooper Vdr/VE

Aluminum Qloss/Qtot

Monel 400 FMSC/1000

Fig. 36 Figure of merit (FOM) and its different ratios estimated for different regenerator materials. Reproduced from Gheith R, Aloui F, Ben Nasrallah S. Investigation of regenerator matrix through figure of merit analysis, AJKFluids2015-22087, pp. V001T22A001. doi:10.1115/ AJKFluids2015-22087; 2015.

202

Stirling Engines

0.35

0.3041

0.2736

0.2968

0.2563

0.30

0.2083

0.25 0.20 0.15 0.10 0.05 0 75 Dp/Pm

80 85 90 Regenerator porosity Qloss/Qtot

FMSC/1000

95 Vdr/5*VE

Fig. 37 Figure of merit (FOM) (FMSC) and its constituting ratios estimated for different regenerator porosities: stainless steel material. Reproduced from Gheith R, Aloui F, Ben Nasrallah S. Investigation of regenerator matrix through figure of merit analysis, AJKFluids2015-22087, pp. V001T22A001. doi:10.1115/AJKFluids2015-22087; 2015. Table 12

Influence of regenerator porosity in thermal losses through regenerator

Parameter

Porosity o85%

Porosity 485%

Thermal losses by external conduction Amount of heat exchanged between the working fluid and the regenerator Heat quantity stored in the regenerator Friction losses

Low Low High High

High High Low Low

Table 13

Effect of operation parameters

Parameters

Range

Influence on asymmetry of temperature between both regenerators' sides

Cooling flow rates (l/min) Charge pressure (bar) Heating temperature (1C) Operating time (s)

0.26–8.16 3–8 300–500 4–20

Significant effect Significant effect The most significant parameter No effect recorded

4.6.5.2

One-Variate Regenerators Experimentation

The one-variate experimentation was proposed by Gheith et al. [3], in order to determine the adequate regenerator for a gamma Stirling engine installation. They found the same result obtained by the FOM formulation [110]. This method needs long time to explore the whole domain of experimentation, a significant experimental time, and does not ultimately represent the complete effect of each parameter on the desired response.

4.6.5.3

Experimental Design Methodology

Experimental design ensures good organization of experimental tests in scientific research or industrial study [111]. This method has several advantages over the one-variant method:

• • • • • •

reduction in the number of test to be carried out to scan the entire field of study; possibility to study a large number of factors at the same time; detection of single or double interactions between the studied factors; modeling of the studied answers; optimal precision of the results; and optimization of the answers.

Gheith et al. [112] applied an experimental design plan to study the effect four operation parameters (Table 13) on the asymmetry of temperature between both regenerators’ sides. Based on this study, authors proposed an empirical model for predicting the value of an asymmetry of temperature between both regenerator sides’ function of operation parameters.

Stirling Engines

Table 14

Comparison between experimental methods One variate

Number of tests

Results precision

Results and perspectives

4.6.5.4

203

• • •

Very important number of tests Do not scan the entire field of study Applied to all parameters

Acceptable precision

–

Figure of merit (FOM)

Experimental design

Limited

Reduction in the number of test to be carried out Scan the entire field of study Applied for only independent parameters Good precision based on isosurface plot Consider the interaction between studied parameters Gives an empirical model for predicting results function of studied factors

Good precision Including losses and Reynolds number –

Comparison Between Methods

The presented experimental method to study the Stirling engine presents different advantages and disadvantages, which are summarized in Table 14. The choice of the adequate method depends on needed results.

4.6.6

Main Results and Discussion

4.6.6.1

Stirling Engines Performance Investigations

1. Compared to Ericson engine:

•

•

At nearly the same working conditions, the Stirling engine presents higher global performances (specific indicated work, thermodynamic and exergetic efficiencies) compared with the Ericsson engine [113], due to the presence of a regenerator. The gap between these performances (about 24.18% of global exergetic efficiency and 15.53% of global thermodynamic efficiency) might be reduced using a preheater in the Ericsson engine. According to Hachem et al. [113], the proportion of total exergy destruction compared with the exergy flux from the hot source is similar for both engines (respectively 44% and 47% of the exergy flux from the hot source for the Stirling and Ericsson engines). The largest exergy destruction occurs in the compression cylinder, mainly due to generated entropy in the case of the Ericsson engine and due to a similar proportion of generated entropy and of heat loss toward the cold source for the Stirling engine. The regenerator (or preheater for the Ericsson engine) presents an important role to supply the expansion cylinder with a high exergy flux. The exergy recovered reaches about 28% of the destroyed exergy.

2. Effect of operation parameters:

•

•

•

The rotation speed is a very determining parameter. The increase of the speed has double effects. On one hand, it favors the exchanges of heat by convection and on the other hand, it increases the losses by viscous friction through the singularities of the thermal machine. Thus, an optimum value of the speed must always be respected to ensure the proper functioning of the Stirling machine [114]. The increase of initial charge pressure leads to an increase of working fluid mass, which increases the Stirling engine brake power [14]. However, the load pressure is limited on the one hand by the capacity of the motor to withstand the high pressure (resistance of the materials) and on the other, by the realization of a perfect seal (to reduce the leaks of working gas). The increase of hot end temperature leads to an increase of the thermal exchanged energy. Thus, the increase of Stirling engine brake power. However, the temperature of the hot end should be moderate. It is limited by the melting temperature of the material of the HEX.

3. The efficiency of heater:

•

The exchanged heat and the efficiency in the heater present periodic evolution [115]. All operation parameters (heating temperature, initial filling pressure, and cooling water flow rate) have significant influence. The amount of heat absorbed by the working fluid in the heater increases with the heating temperature, the difference of temperature between both Stirling engine working spaces increases with the cooling water flow rate and the mass for working fluid involved to the process of heat exchange in the heater increases with the initial filling pressure. All this parameters contribute to the amelioration of a Stirling engine heater [115].

4. The CFD modeling:

•

The CFD model gives the trends and the correct values of the influence of the different operating parameters on the performance of the Stirling engine.

204

Stirling Engines

4.6.6.2

Stirling Engine Regenerator Investigation

The performance of Stirling engine is closely related to its regenerator’s properties. Thus, the regenerator is the most studied compartment of the Stirling engine. The main results are summarized as follows: 1. Thermal losses inside regenerator: • The regenerator is the seat of significant losses by internal conduction, by imperfection and pressure drops (caused by friction of working gas with the internal walls of the porous matrix) [14]. • The maximum losses are recorded in regenerator. According to Ref. [14], it is the set of 44% of viscous loss, 33% of internal conduction loss, and 22% of imperfection loss respectively from the total losses inside it. The Shuttle loss is only recorded in compression and expansion spaces. The mechanical loss is recorded in the engine driving mechanism. They decrease the output brake power of the Stirling engine. • These losses depend on the geometrical and physical properties of the regenerator’s material [14]. • The Stirling engine is exposed to continuous internal and external perturbations while operating. However, it cannot change speed quickly [116]. Irreversibilities inside the Stirling engine increase when changing its functioning regime. Thus, thermal losses inside the regenerator increase. 2. Average entropy generation rate in the Stirling engine regenerator: • Entropy generation in the regenerator is generated from the irreversibility owing to heat transfer with finite temperature gradients and the friction of fluid flow [117]. • Entropy generation in the regenerator is associated with a number of parameters including the characteristics of the regenerator (geometry, porosity, and material), the working conditions (temperature ratio between hot and cold source, initial pressure, and rotational engine speed) and the thermophysical properties of the working fluid [117]. Thus, the best regenerator qualities are those corresponding to the minimum entropy generation (f¼0.85, Lr/Dr ¼ 1.3 and stainless steel as matrix material). • Compared to air as working fluid, the average entropy generation rate in the regenerator is reduced when using helium as working fluid [117]. • The average entropy generation rate in the regenerator increases with rotational engine speed, hot end temperature, and initial pressure [117]. 3. Regenerators proprieties: • The stainless steel material is the most suitable material that can be used as a regenerator. An optimal porosity of 80% can be considered for the porous regenerator. It is recommended to use a high pressure and a low initial filling pressure to maximize regenerator performances [110]. • The use of copper for air as working fluid must be avoided because of oxidation problem [110]. • An optimum value of porosity corresponds to the maximum mechanical power [118]. • According to Ref. [83], thermal losses inside the regenerator are function of regenerator length and diameter (regenerator design). For A 25 W Beta type Stirling prototype, the optimal values for length and diameter are respectively about 60 mm and 22 mm. 4. Optimization methods: • The FOM is a rapid tool to evaluate Stirling engine regenerator. To avoid the multiobjective complicated studies the FOM can be used since it considers the major characteristics of porous media regenerators [110]. After, the FMSC evaluation tool was calculated for different regenerators having different material (stainless steel, copper, • aluminum, Monel) and different porosities (75%, 80%, 85%, 90%, and 95%). The stainless steel and the copper material present the highest FMSC for respectively 0.273 and 0.274.

4.6.7

Future Directions

In order to design and size less polluting and less energy consuming systems, the main goal of the next researches will be the study, using experimental and numerical CFD simulation, the evaluation of an innovate micro-CHP unit that will couple a 1.2-kW double acting Stirling engine to the exhaust gas of an ICE. In the automotive field, usually the efficiency of current engines does not exceed 40% on their best operating point. In most cases, the yield is well below 20%. A significant part of the losses is in the form of heat evacuated by the cooling circuit and by the exhaust gases. Several technical devices are possible to recover this energy, such as external combustion engines and Stirling engines in particular.

4.6.7.1

Description of the Recovery Process Using the Stirling Engine

In the proposed recovery process, the Stirling engine will be coupled to a generator to power the car’s battery and the other car systems consume electrical energy, instead of using electrical power produced directly from the alternator coupled to ICE drive shaft while it is running. It can enhance the amount of power available for the car propulsion. The innovative recovery system will be composed of a gasoline engine and a double acting Stirling engine that is composed of heater, cooler, and motion transmission

Stirling Engines

205

Output work

Motion transmitting mechanism Exhaust gas Hot heat exchanger (HEX)

Gasoline Combustion chamber

Catalyst

Air

Internal combustion engine (ICE)

Expansion space (Hot)

Regenerator

Output work

Exhausts out

Exhaust gas

Stirling engine

Compression space (cold)

Motion transmitting mechanism

T0, P0

Fig. 38 Equipment of the recovery process.

system coupled to an alternator. The mechanical power produced by the Stirling engine will be used to drive the alternator, thus providing the required electrical power for the different electronic devices inside the car (air conditioning system, electric power steering, electrical windows motors, built-in satellite navigation, etc.). In order to achieve this project, different targets must be accomplished. Fig. 38 presents all these equipments and the main bond between them. The hot exhaust gases from the ICE go through the hot heat exchanger (HEX) in contact with the working gas (helium or nitrogen in this case) in the Stirling engine (SE). On the other side, cold water will go through the cold heat exchanger (CEX) in contact with the working fluid (helium or nitrogen) inside the Stirling engine.

4.6.8

Conclusions

A review of Stirling engine regenerators was proposed in this chapter. The most significant regenerator parameters were presented with their main influences on Stirling engine output and discussed. The constituting material and the porosity are the most influencing parameters for regenerator performances, which is the main key of the Stirling engine. New material, such as graphite and carbon fiber, can give high thermal efficiency. The quasi-steady model including thermal losses can give satisfying results compared to experimental ones. The CFD simulations can provide detailed evolution of working fluid particles and help avoiding recirculation zones when designing and sizing new Stirling engine installations. The FOM is presented as a powerful tool to test several regenerators for the same installation. It provided fast and precise results, with a minimum of experimental investigation measurements series. The experimental design methodology scans the entire field of the study and determine the influence of the interaction between studied factors on the response. This method gives an empirical model (equation) that can link the desired result(s) to the studied factors.

Acknowledgments This work was supported by the laboratory LAMIH CNRS UMR 8201 (University of Valenciennes), the laboratory LESTE (ENIM, Monastir, Tunisia), the regional Council of Hauts-de-France (province of north of France) and the European Commission within

206

Stirling Engines

the International Research Staff Exchange Scheme (IRSES) in the 7th Framework Program FP7/2014-2017/ under REA grant agreement no. 612230. This support is gratefully acknowledged.

In Memory of Our Colleague, Professor Sassi BEN NASRALLAH This research topic, on the boundary Stirling engine, was initially initiated at the LAMIH laboratory (UMR CNRS 8201) of the University of Valenciennes by Professor Fethi ALOUI and LESTE laboratory of the University of Monastir (Tunisia) and by our colleague, Professor Sassi BEN NASRALLAH, coauthor of this current study, who left us suddenly on June 27, 2017 before the final submission of this paper, after a myocardial heart attack. We would like firstly to pay a very great tribute to him. Professor Sassi BEN NASRALLAH was an excellent colleague, very serious, scientifically very curious, and very human. With him, we lost more than just a colleague, but a very dear friend and brother.

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Further Reading NASA. A microfabricated segmented-involute-foil regenerator for enhancing reliability and performance of Stirling engines Phase II Final Report for the radioisotope power conversion technology NRA Contract NAS3-03124, NASA/CR-215006; 2007.

Relevant Websites https://www.stirlingengine.com/ American Striling Company. http://www.boehm-stirling.com/en/engines.html Boehm. http://cleanergy.com/ CleanEregy. http://diystirlingengine.com/stirling-engine-generator/ DIY Stirling Engine. http://www.microgen-engine.com/ Microgen. https://www.grc.nasa.gov/www/tmsb/stirling.html NASA. https://www.ohio.edu/mechanical/stirling/intro.html Ohio education: Background and Introduction. http://www.robertstirlingengine.com/ Stirling engine. https://en.wikipedia.org/wiki/Stirling_engine Stirling engine. http://sunpowerinc.com/1kw-stirling-engine/ Sunpower.

4.7 Gas Turbine Cycles Ibrahim Dincer and Murat E Demir, University of Ontario Institute of Technology, Oshawa, ON, Canada r 2018 Elsevier Inc. All rights reserved.

4.7.1 Introduction 4.7.2 Historical Development 4.7.3 Classification of Gas Power Cycles 4.7.3.1 Totally Reversible Gas Power Cycles 4.7.3.2 Internally Reversible Gas Power Cycles 4.7.4 Gas Turbine Cycles 4.7.5 Classification of Gas Turbine Cycles 4.7.5.1 Type of Combustion Process 4.7.5.2 Type of Gas Turbine Cycles According to Working Fluid Path 4.7.5.3 Type of Purpose of the Cycle 4.7.5.4 Classification of Gas Turbine Cycle According to Number of Cycles 4.7.6 Air – Standard Brayton Cycle 4.7.6.1 Actual Brayton Cycle 4.7.6.2 Regenerative Brayton Cycle 4.7.6.3 Reheat-Regenerative Brayton Cycle 4.7.6.4 Brayton Cycle With Intercooler 4.7.6.5 Exergy Destructions in Brayton Cycle Power Plants Estimation 4.7.7 Jet Propulsion Cycles 4.7.7.1 Exergy Analysis of a Turbojet 4.7.7.2 Modifications to Turbojet Engines 4.7.8 Combined Cycle Power Plants 4.7.8.1 The Brayton–Kalina Cycle 4.7.8.2 The Brayton–Brayton Cycle 4.7.8.3 The Brayton–Diesel Cycle 4.7.8.4 The Brayton–Stirling Cycle 4.7.8.5 The Brayton-Fuel Cell Cycle 4.7.8.6 The Chemical Recuperation Cycle 4.7.8.7 Integrated Gasification Combined Cycle 4.7.9 Case Studies 4.7.9.1 Case Study 1 4.7.9.1.1 Thermodynamic analysis 4.7.9.1.2 Results and discussion 4.7.9.1.3 Conclusions 4.7.9.2 Case Study 2 4.7.9.2.1 Thermodynamic analysis 4.7.9.2.2 Solar receiver 4.7.9.2.3 Motor work 4.7.9.2.4 Case study results and discussion 4.7.9.2.5 Parametric studies 4.7.9.2.6 Effects of direct normal irradiance 4.7.9.2.7 Effects of ambient temperature 4.7.9.2.8 Effects of wind speed 4.7.9.2.9 Effects of tank insulation 4.7.9.2.10 Conclusions 4.7.10 Future Directions 4.7.11 Concluding Remarks Acknowledgment References Further Reading Relevant Websites

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Nomenclature

N P _ Q r R s T U v V w _ W

Total number Pressure (kPa) Heat rate (kW) Compression ratio Universal gas constant (J/K mol) Specific entropy (kj/kg K) Temperature (k) Overall heat transfer coefficient (W/m2 K) Specific volume (m3) Velocity (m/s) Specific work (kJ/kg) Work rate (kW)

Acronyms AF ASTM

HEX HRSG GTC IDGTE IMechE LHV LNG LPG PCM RAF NG SI UPR WHR

Heat exchanger Heat recovery steam generation Gas turbine cycles Institution of Diesel and Gas Turbine Engineers Institution of Mechanical Engineers Lower heating value (kJ/kg) Liquefied natural gas Liquefied petroleum gas Phase change material British Royal Air Force Natural gas Spark-ignition Union Pacific Railroad Waste heat recovery

Greek letters Zen Energy efficiency c Exergy efficiency Stefan–Boltzmann constant (5.76 10 8/m2K4) sSB

l ζ Ε

Air–fuel equivalence ratio Specific exergy function Surface emissivity

Subscripts ac C d field GT HEX i In ins ise Out ov

p P Ph reg s Sys si sr T 0 1 1, 2, … i

Propulsion Pump Physical Regeneration Shaft power System Heat sink Heat source Turbine, thrust power Ambient condition Free stream State points

C cp

Compressor Specific heat capacity at constant pressure (kJ/kg K) Specific heat capacity at constant volume (kJ/kg K) Exergy rate (kW) Specific exergy (kJ/kg) Specific enthalpy (kJ/kg), convective heat transfer coefficient (W/m2 K) Thermal conductivity (W/m-K), Specific heat ratio Mass flow rate (kg/s)

cv Ex_ ex h k _ m

Air–fuel ratio American Section of the International Association for Testing Materials ATS Advanced turbine systems CAES Compressed air energy storage CCPP Combined cycle power plants CEGB Central electricity generating board CI Compression-ignition CPC-GTC Constant pressure combustion gas turbine cycle CVC-GTC Constant volume combustion gas turbine cycle DNI Direct normal irradiance (W/m2) EIF Efficiency improvement factor EES Engineering equation solve

4.7.1

and superscripts Aircraft Compressor Destruction Heliostat field Gas turbine Heat exchanger State number Inlet of a component Insulation Isentropic Outlet of a component Overall

Introduction

Power generation has been a critical subject in meeting the electricity demands (i.e., electrification) in almost every part of the world. Although there are two crucial cycles, namely Rankine cycle (using water as working fluid) and Brayton cycle (using air as the working fluid), Brayton cycle (i.e., air standard Brayton cycle or gas turbine cycle) offers some more advantages over steam Rankine cycle, such as efficiency, effectiveness, flexibility, and utilization diversity.

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Gas turbine power systems have been finding applications for commercial purposes in various sectors, ranging from power generation to propulsion in aircrafts since the early 19th century. Usually, these systems consist of a simple closed or open air standard Brayton cycle in which a compressor takes the air at the atmospheric conditions and pressurizes it. A combustion chamber is where the fuel is burned at constant temperature with air and gas turbine, where the air is expanded to generate electricity [1]. The gas turbine is a useful device to convert heat energy into mechanical energy by rotating shaft with higher capacities and efficiencies. In addition, another advantage of the gas turbine systems is that these systems possess the ability of using various types of fuels in the combustion chamber. While natural gas (NG), liquefied natural gas (LNG), liquefied petroleum gas (LPG), refinery gas, coke oven gas, coal gas, and hydrogen are considered some of the most common fuels used in gas turbines, no. 2 diesel, kerosene, jet A fuel, naphtha, ethanol and methanol, heavy residual-grade oils, and crude oils are the most commonly used liquid fuels in gas turbines [2]. The efficiency of a single gas turbine cycle can exceed 50% by means of using combined or integrated cycle options and/or heat recovery (and regenerative) subsystems. Moreover, reheating is employed to improve the performance of the system. Inexpensive and readily available working fluids (such as air and water), along with well-developed technologies (such as gas turbine, heat recovery steam generator, or steam turbine units), all within a short construction period and a high overall efficiency, have helped achieve greater acceptance of these systems. Note that combined cycle plants are able to reach efficiencies well above 58% with plant capacities in the range between 350 and 500 MWe [3,4]. In this chapter, we primarily focus on gas turbine cycles (GTCs) to discuss their historical developments in several sectors (including utility and aviation sectors) and classifications, to present their thermodynamic studies through energy and exergy approaches, to compare theoretical (ideal) versus actual cycles and their performances, and to develop analyses and assessments for integrated and combined systems. There are also sample problems and case studies presented to illustrate the importance of gas turbine systems and their critical role in achieving the generation of multiple commodities. Furthermore, future directions for the GTCs are discussed.

4.7.2

Historical Development

The modern gas turbines used in industrial applications have evolved to their current outlook after going through a series of technological research, improvements, and developments that emerged in the 18th century. Hunt [5] explained the technological development of the gas turbines in six stages as illustrated in Fig. 1. Additionally, Hunt along with other scientists [5,6] classified and explained some of the milestones for the gas turbine technologies. Here, we provide a brief historical summary to outline the progress [5,6]:

• • • • • • • • • •

• •

1791: The English coalmaster and inventor Josh Barber, who also had several patented devices, patented the first gas turbine entitled “A Method of Rising Inflammable Air for the Purposes of Procuring Motion [5].” 1808: The English engineer John Dumbell patented his work for a vertical axis and multirotor steam turbine [6–8]. 1837: The French inventor M. Bresson came up with the idea of compressing the working fluid before the combustion process [7]. 1872: The American mechanical engineer George Brayton initiated a cycle for gas turbine plants and registered a patent under his name, also known as the Brayton cycle [6]. 1872: The German inventor, photographer, stenographer, and writer Franz Stolze patented his first gas turbine model, which included a multistage reaction turbine and a multistage axial flow compressor. Later on, the model he introduced was tested several times between 1900 and 1904, however, none of them was successful [9]. 1899: The American engineer, inventor, and patent attorney patented the first American gas turbine. Later on, he sold the rights of his design to General Electrics in 1901 [5]. 1898–1981: The Swiss aeronautical engineer Jakob Ackeret, who was considered as the pioneer of modern aerodynamics, made several significant types of research in the field including airfoil theory, axial flow compressors, and high-speed propulsion problems [5,10]. 1903: The Slovak engineer, physicist, and inventor Aurel Boleslav Stodola who was also known as the pioneer in technical thermodynamics published the book Steam Turbines and allocated an appendix on gas turbines [5]. 1903: Two bright engineers (Charels Lamale and René Armengaud) presented the first successful commercial combustion gas turbine at the Société Turbos Moteurs in Paris. The system that they built consisted of a multistage compressor, a combustion chamber where the liquid fuel was combusted, and a gas turbine. The system cooling was succeeded by the water injection [6]. 1903: The Norwegian researcher, inventor, and one of the pioneers of the gas turbine technology Ægidius Elling was able to propose and build the first gas turbine with an excess power, which included both rotary turbine and compressors. This innovative design produced 8.2 kW. Afterwards, he constructed his second model and quadrupled the power output of the first design to produce 32.8 kW [5,6]. 1905: The German engineer Dr. Hans Holzwarth suggested a propulsion gas turbine in which expansion phase of the cycle was extended until the atmospheric pressure. The proposed system was constructed, and many experiments were conducted based on it between 1909 and 1913; as a consequence, the maximum overall thermal efficiency achieved by them was about 13% [6,9]. 1930: The British engineer air commodore Sir Frank Whittle, who was also an engineering officer in the British Royal Air Force (RAF), registered for the first patent of the jet engine. Independently from Whittle in 1934, the German physicist Dr. Hans von Ohain patented his jet engine design with a different internal configuration. Both models were tested with attainment in 1937 [5,6].

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The brief history of industrial gas turbines 1750

1800

1850

1900

1950

2000

The early stage The engineers and technology

Gas turbine concepts

Experiments

Development stages of gas turbines

SE England black outs

Compressors New York black out

Trials

Natural gas

Early machines

Technological developments Industrial gas turbines introduced Combined cycle power plants

Packaged plants introduced

1750

1800

1850

1900

1950

2000

Fig. 1 The historical timeline of the development of gas turbines. Modified from Hunt RJ. The history of the industrial gas turbine (Part 1 the first fifty years 1940–1990). Indep Tech Forum Power Gener 2011;15(2):1–48.

• • • • • • •

1939: Dr. Adolf Meyer [9], who was a member of the Brown Boveri Company, introduced his remarkable publication “The Combustion Gas Turbine: Its History, Developments and Prospects” in London and this paper was followed by the first practical industrial gas turbine production by the same company in 1939. 1948: Bowden and Jefferson [11] from CA Parsons and his company introduced their study “The Design and Operation of the Parsons Experimental Gas Turbine” at a conference for the Institution of Mechanical Engineers (IMechE) in Newcastle upon Tyne. In their paper, they explained their experimental study results as well as their gas turbine in detail. 1948: R.J. Welsh from the English Electric Company presented his study, which was about improvement and applications of a gas turbine. The paper was published by the Institution of Diesel and Gas Turbine Engineers (IDGTE), which was previously known as Diesel Engine Users Association (DEUA). Later on, the same institution presented many studies in the same field [5]. 1949: Oklahoma Gas and Electric Company constructed the first gas turbine facility for electricity production in Belle Isle. The design consisted of a 15-stage compressor and two-stage expansion turbine, which was driven by NG. The turbine operated the alternator by 3600 revolutions per minute (rpm) and resulted in 3.5 MW electricity generation. 1953: The first propane-fueled locomotive, which was driven by the gas turbine, was constructed and commissioned into service by the Union Pacific Railroad (UPR). The locomotive was able to reach 3.6 kW power and carried the freight from Los Angeles to Las Vegas and vice versa [6]. 1965: In the mid-1960s, Dr. Seippel announced the combined cycle model, which consisted of a gas turbine system and a steam turbine subsystem; discussion over the economic feasibility of the gas turbines in large scale industrial use clouded his efforts [5]. 1992: In order to improve the industrial gas turbine facilities and build a partnership between academia and industry, the Advanced Turbine Systems (ATS) Program was introduced by the US Department of Energy [6].

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•

• •

213

2003: A large-scale combined cycle turbine was installed by General Electric in Cardiff, Wales. The device had a turbine speed of 50 Hz, general dimensions of 11.9 m length and 4.9 m diameter, and a weight of 368 t. The gas turbine system had the privilege to be the first system that passed the threshold of 60% thermal efficiency because of its innovative steam cooling system that allowed more than 90oC increase in operation temperature over prior designs. 2009: An integrated solar combined cycle was constructed and operated in Iran. The plant consisted of three turbines as two gas turbines and a steam turbine, and a parabolic solar collector plant. The solar subsystem was utilized for a supplementary steam generation for the steam turbine [6]. 2011: Several airline companies started using biofuels after American Section of the International Association for Testing Materials (ASTM) International, which is an international standards organization that develops and publishes technical standards, agreed on the utilization of the renewable fuel with conventional fuel in a mixture [6].

Throughout the historical evolution of the gas turbine system, its application areas have been expanded with new technologies. At present, gas turbines have a versatile type of function that can be listed and briefly explained as follows (see Ref. [5] for details):

• • •

• •

• •

•

Marine propulsion: with the installment of the Beryl engine to the Motor Gun Boat MGB2009, gas turbines were first utilized in marine vessels. Today, gas turbines have an important role in sea vessels since marine applications have around 7% share of the total gas turbine market. Road vehicle engine: gas turbines were first used in land vehicles after Centrax installed a gas turbine in its new truck design in 1948. The constructed engine had a power capacity of 119 kW. Locomotive engines: gas turbines found application opportunity in the railway industry in the 1940s. The commercial company Brown Boveri introduced its first gas turbine driven locomotive in 1941. Following this development, a couple of manufacturers made their designs for applications. However, the practical application of the gas turbines in locomotives was interrupted later on due to higher oil prices. Today, only a few gas turbine locomotives are actively in service. Power station standby and peak lopping: due to insufficient electricity generation over to peak loads, serious electricity blackouts happened in the 1960s in some regions of England. Therefore, the Central Electricity Generating Board (CEGB) designed a fast response gas turbine service for the peak demands. Mechanical drive: even though the majority of industrial gas turbine applications have been focused on either power production or the marine industry, gas turbines have been used for mechanical drives since the early period of this technology. For instance, they were also utilized to drive devices such as compressors or pumps. Now, the mechanical drive applications have around 30% share of the gas turbine market. Cogeneration plants: in the mid-20th century, heat and electricity cogeneration plants gained increased popularity. Those facilities were usually utilized for increasing the efficiency by recovering the waste heat of power production systems. Today, it is possible to add cooling as an output beside heat and electricity, which leads to trigeneration. Combined cycle: a combined cycle power plant uses both steam and gas turbines together to produce more power than a single plant. The waste heat of the gas turbine is used for steam production for the steam turbine. Dr. Meyer is considered an early proponent of this technology and published a study on the waste heat recovery (WHR) of gas turbine systems. Nevertheless, those systems did not get the attention that they deserved up until the mid-1960s. Afterwards, because of their higher performances, CCPPs attracted the attention for applications over the traditional single cycle plants. Educational purposes: gas turbines are also utilized in the educational sector. From 1955 to 1965, various companies, for educational and training purposes sold over 250 small-scale gas turbines (44.7 kW) to universities/colleges throughout the world.

4.7.3

Classification of Gas Power Cycles

Unlike steam power cycles, no phase change through condensation or boiling is observed in gas turbine power cycles. Thus, in the gas turbine power cycles, the working fluid remains in a gaseous state during the entire cycle process. Gas power cycles can be classified into two primary categories as shown in Fig. 2: internally reversible and totally reversible cycles. The Otto, Diesel, and Brayton cycles are considered the three common types of internally reversible gas power cycles whereas Carnot, Stirling, and Ericson are considered the main three types of totally reversible gas power cycles [1].

4.7.3.1

Totally Reversible Gas Power Cycles

A reversible power cycle can only be achieved if both heat addition and rejection processes occur under isothermal conditions. If such processes are not isothermally reversible, there would be some irreversibilities due to a finite temperature difference between thermal reservoir (at fixed temperature) and working fluid [1]. Both isothermal heat addition and rejection are very challenging to accomplish, as they require very large heat exchangers (HEXs) and heat transfer time. However, in practice, heat engines cannot have both of those conditions. Hence, those cycles indicate the ideal conditions and show the limits of the systems. The efficiency of any totally reversible heat engine such as Carnot, Stirling, between a heat source at the temperature TH and a heat sink
or Ericson  at the temperature TL is described by the Carnot factor Zcar ¼ TTHL . Any other power cycle that works between those temperatures

214

Gas Turbine Cycles

Gas power cycles

Internally reversible cycles

Totally reversible cycles

Carnot cycle

Ericson cycle

Stirling cycle

Otto cycle

Diesel cycle

Brayton cycle

Fig. 2 Classification of gas power cycles.

P

T qin

1

1

2

TH

qin 2 4 qout

3

TL V

3

4 qout

S

Fig. 3 P-v and T-s diagrams of the Carnot cycle.

cannot exceed the Carnot efficiency. P-v and T-s diagrams of the Carnot cycle are shown in Fig. 3 and explained as follows:

• • • •

Process 1-2: the working fluid expands isothermally while the space is heated externally by the heat source, QH. During the expansion, the system produces useful work output. Process 2-3: at this step, heat addition to the cycle stops and air continues to expand in the thermally isolated media isentropically till the temperature drops from TH to the temperature of the heat sink, TL. During the expansion, the system also produces the useful work output. Process 3-4: the working fluid is compressed isothermally while the space discharges the heat to the heat sink, QL. During the compression, the system requires work input. Process 4-1: heat rejection from the cycle stops and air continues to be compressed into the thermally isolated media isentropically. The temperature of the air reaches to the temperature of the heat source, TH. During the compression, the system requires the work input again.

Regarding the Stirling cycle, a double-effect piston and cylinder set-up can achieve this. A regenerator porous matrix should be installed inside the set-up. The working fluid can be an ideal gas and all the processes should take place under reversible conditions as expected. P-v and T-s diagrams of the Stirling cycle are shown in Fig. 4. The processes can be described as follows [1]:

• • • •

Process 1-2: similar to Process 1-2 in the Carnot cycle, the working fluid expands isothermally while the space is heated externally by the heat source QH. During the expansion, the system produces useful work output. Process 2-3: here, a regenerator works with 100% effectiveness and cooling with regeneration achieved at constant volume. At the end of the process, the temperature of the working fluid drops from TH to TL. Process 3-4: similar to Process 3-4 in the Carnot cycle, the working fluid is compressed isothermally while the space discharges the heat to the heat sink, QL. During the compression, the system requires work input. Process 4-1: heating occurs while the regenerator works with 100% effectiveness at constant volume. At the end of the process, the temperature of the working fluid increases from TL to TH.

As it can be observed in Figs. 3 and 4, the Carnot and Stirling cycles have identical heat addition and rejection processes. The isochoric processes possessed in the Stirling cycle differ from each other. Similarly, the Ericsson cycle has the same isothermal heat addition and discharge processes (see Fig. 5). In the Ericsson cycle, heating and cooling processes occur at isobaric conditions instead of isochoric conditions in the Stirling engine.

Gas Turbine Cycles

P

215

T

1

qin

1

TH

2

qin Regeneration

2

Regeneration

4

qout TL

3

4

3

qout

V

S

Fig. 4 P-v and T-s diagrams of the Stirling cycle.

T

P

qin 1

qout

Regeneration

4

1

TH

qin

Regeneration

qout TL

3

2

2

4 V

3 S

Fig. 5 P-v and T-s diagrams of the Ericsson cycle.

4.7.3.2

Internally Reversible Gas Power Cycles

The main three internally reversible gas power cycles are the Otto, Diesel, and Brayton cycles. All three cycles are suitable for internal combustion engines and therefore engine-driven power generation applications. Otto and Diesel cycles are operated by a piston-cylinder mechanism while the Brayton cycle drives a gas turbine. Hence, in this chapter, the Brayton cycle and its derivatives are explained in detail. P-v and T-s diagrams of the ideal Otto cycle are presented in Fig. 6. The Otto cycle is the ideal cycle for spark-ignition (SI) engines. This cycle was presented in the late 19th century after Nikolaus Otto demonstrated the four-stroke SI engine successfully [1]. Description of the processes is given as follows:

• • • •

Process 1-2: isentropic compression of the working fluid occurs as the piston moves from bottom dead center (BDC) to top dead center (TDC). Process 2-3: here, with the SI, a rapid burning occurs inside the piston; therefore, the heat addition takes place at the constant volume. Process 3-4: in this process, the working fluid expands isentropically and produces the useful work for the cycle. Process 4-1: heat removal at constant volume. In practical applications, the heat is removed by expelling the exhaust gas to the atmosphere.

The Diesel cycle is used in numerous large-scale industrial applications for power generation. The Diesel engine is also known as a compression ignition (CI) engine, which was proposed in Germany by Rudolf Diesel in the late 19th century. In the CI engine, the ignition of the combustion process is attained by compression of the working fluid up to a larger temperature than the fuel’s autoignition temperature. When the temperature of the air increases adequately, high-pressure liquid fuel is sprayed into the combustion chamber. Thus, the ignition process occurs rapidly [1]. P-v and T-s diagrams of the ideal Diesel cycle are presented in Fig. 7. A description of the Diesel cycle processes is given as follows:

•

Process 1-2: isentropic compression of the working fluid occurs as the piston moves from BDC to TDC. This process requires an external work input.

216

Gas Turbine Cycles

P

T 3

3 qin

qin

4

4 2 2 qout

1

qout 1 S

V TDC

BDC

Fig. 6 P-v and T-s diagrams of the Otto cycle. BDC, bottom dead center; TDC, top dead center.

P

T qin 2

3

qin

2

3 4

4 qout

qout 1

1 S

TDC

BDC

Fig. 7 P-v and T-s diagrams of the Diesel cycle. BDC, bottom dead center; TDC, top dead center.

• • •

Process 2-3: here, with the rising temperature of the compressed gas and fuel injection, ignition occurs instantly, which results in growth in the temperature. Since the piston moves during the process, the heat addition takes place at constant pressure. Process 3-4: in this process, the working fluid expands isentropically and produces the useful work for the cycle. Process 4-1: heat removal at constant volume by expelling the exhaust gas to the atmosphere.

4.7.4

Gas Turbine Cycles

Commercial power plants with combustion turbines have been developing since early 20th century. Conventional power combustion turbine power plants operate on a simple open Brayton cycle consisting of a compressor, a combustion chamber, and a gas turbine. The combustion turbine is considered as the most effective way to convert the chemical exergy of fuels in the gaseous or fluid state to electric power. GTCs are widely accepted systems due to their ability to utilize various kinds of fuel as the heat source. Moreover, the working fluid of those cycles is usually air, which is highly available in the environment. Another reason for their popularity is their quick response time. A combustion turbine can start up in less than a minute. With the developing technology, efficiency of those systems has been increasing. By using reheating, regeneration, compressor intercooler, and turbine blade cooling systems, it is possible to enhance the performance of the GTCs. However, even after installation of those systems, the exhaust temperature of the cycle is still considerably high. Therefore, conventional systems use the exhaust of the gas turbine cycle as the heat source of a bottoming Rankine cycle. The efficiency of these combined cycles are able to exceed 55% [1]. For the following sections, simple and combined Brayton cycles are classified, explained, and thermodynamically investigated.

4.7.5

Classification of Gas Turbine Cycles

GTC can be structurally and functionally classified in many ways. In this section, GTCs are classified into four main categories as shown in Fig. 8. Descriptions of the categories are briefly explained in the scope of this chapter as follows.

Gas Turbine Cycles

217

Gas turbine cycles

Type of combustion process

Type of working fluid

Purpose of cycle

Number of cycles

Constant pressure combustion gas turbine cycles

Open gas turbine cycles

Aircraft jets

Simple cycle

Constant volume combustion gas turbine cycles

Close gas turbine cycles

Ground-based applications

Combined cycle

Fig. 8 Classification of gas turbine cycles (GTCs).

Fuel intake

Compressor

Intake valve

Exhaust valve Combustion chamber

Power turbine

Generator

Turbine Air intake

Exhaust

Fig. 9 A simple layout of a constant volume combustion gas turbine cycle (CVC-GTC).

4.7.5.1

Type of Combustion Process

GTCs can be classified according to combustion processes they have: constant pressure combustion gas turbine cycle (CPC-GTC) and constant volume combustion gas turbine cycle (CVC-GTC). CPC-GTC is the most common gas turbine system cycles, which mainly follow the Brayton cycle (see Fig. 12). A simple open CPC-GTC cycle takes the working fluid with a compressor at isentropic conditions and pressurizes the air to combustion pressure. As it is an open flow, the air is burned into the combustion chamber at isobaric conditions. After, high-pressure and high-temperature working fluid flows through and expands in the gas turbine isentropically. Unlike CPC-GTC, in a constant volume gas turbine cycle, combustion occurs in an enclosed media (constant volume). Therefore, pressure increase takes place in two steps. First, the compressor takes the air from ambient and increases its pressure slightly with a low-pressure ratio. Then, the working fluid burns into the combustion chamber at constant volume, which results in massive pressure increase at the elevated temperature. This way, once the highest pressure of the cycle is fixed, the power adsorbed for compression is reduced. Consequently, the generated power from the turbine and its efficiency are greater. On the other hand, as the blades of the turbine, combustion chamber, and control valves are exposed to the high gas temperature, the performance of this system becomes limited. Thus, continued development of this type of turbines ended. These problems could be overcome by using the automation technology for valve control systems [12]. In Figs. 9 and 10, a simple schematic and T-s diagram of a CVC-GTC are presented respectively. First, air is taken by the compressor from the atmospheric conditions (state point 1) and then the pressure and temperature rise to the conditions of the state point 2. Afterwards, the working fluid flows through the intake valve and enters the combustion chamber. During air intake to the combustion chamber, fuel is also directed to the chamber by a nozzle as well. When the chamber is filled with the air–fuel mixture thoroughly, the intake valve closes, and an igniter starts combustion. As the combustion process occurs in an enclosed medium, both temperature and pressure increase, and reach their maximum values (state point 3). When the combustion process finishes, the exhaust valve is opened, which allows the working fluid to flow through the turbine and expand (state point 4). Therefore, the power gas turbine provides the output power (state point 5). When the exhaust gases are discharged from the combustion chamber, the exhaust valve is closed, the intake valve is opened, and the cycle repeats the same procedures.

4.7.5.2

Type of Gas Turbine Cycles According to Working Fluid Path

GTCs can also be categorized another way: closed and open cycles. In closed cycles, the working fluid is returned to the initial state at the end of the cycle, and then it is recirculated. A HEX performs the heat addition to the cycle. On the other hand, in open cycles,

218

Gas Turbine Cycles

3

T

First gas turbine 4 Combustion chamber Power gas turbine

5 2 Compressor 1 s

Fig. 10 T-s diagram of a constant volume combustion gas turbine cycle (CVC-GTC).

the working fluid is discharged at the end of each cycle instead of being recirculated, and the compressor takes fresh air continuously. Heat addition is supplied by the combustion of the fuel. Even both cycles are thermodynamically equivalent; they have strong sides to each other. For example, operation conditions of the open cycles are limited by the ambient conditions. The air can only expand to the atmospheric pressure. This problem can be overcome by using closed cycle configuration [1,13]. In closed cycles, it is possible to expand the air below atmospheric pressure hence more power could be generated by the gas turbine. Additionally, exhaust gas corrosion is not seen in a closed cycle as the only clean air is circulated in the cycle. However, closed cycles require two additional HEXs, which cause flexibility problems.

4.7.5.3

Type of Purpose of the Cycle

GTCs can also be grouped by their duties. A conventional Brayton cycle is generally used to generate a shaft work, which could be mounted on a generator or it could supply the work for a propeller of a marine vessel. A relatively small portion of the shaft work is used for the compressor in these applications. Jet engines differ from the conventional gas turbine cycle by their exit turbine pressures. Exhaust gases leave the gas turbine at a considerably higher pressure than the atmospheric pressure and expand through the nozzle at high speed and discharge. The high-velocity changes create the propulsive force [13]. In an ideal jet engine application as shown in Fig. 11, the turbine generates the work only enough to run the compressor. In other words, the net work of the jet engines is zero. The main working principle of a jet engine is to increase the momentum of the incoming air at the combustion chamber and maintain the highest air velocity through the gas turbine. As the air discharges with a high speed, the imbalance in the momentum creates the thrust.

4.7.5.4

Classification of Gas Turbine Cycle According to Number of Cycles

One of the methods to categorize the GTCs is a classification with respect to the number of power cycles they have. Simple GTCs consist of only one power cycle in which the working fluid is discharged or recirculated at the end of the cycle, instead of being utilized as a heat source for a bottoming cycle. Therefore, they are relatively less efficient. Despite their low efficiency, they are still favorable due to their compact structure, comparatively low costs, and operational advantages. Moreover, as the cycle can be initiated quickly, it can reach the peak load quicker as well. Besides all these benefits, they have low efficiencies, which result in more fuel consumption per generated power and limits the applications of these cycles. Exhaust temperatures of simple gas turbines are considerably higher than the atmospheric temperature. The discharging processes occur at atmospheric pressure for the open Brayton cycle. Therefore, it is impossible to expand the working fluid below atmospheric pressure. The high-temperature and low-pressure stream still have high potential and it can be used by the integration of the WHR units to the Brayton cycle. The high energy can be recovered if the expelled gases can be utilized in a Rankine cycle (or other bottoming cycle options) [1]. CCPP mainly work on the principle of recovering the waste heat of the topping cycle, which results in the increase of the overall thermal efficiency of the cycle. This leads to less fuel consumption per generated power. On the other hand, investment costs of these cycles and relatively poor response time to meet the power demand can be considered as the weak points of the cycle. In the scope of this chapter, comprehensive energy and exergy analysis for both simple and combined GTCs are conducted.

Gas Turbine Cycles

219

Fig. 11 A Turbofan engine (LEAP engine manufactured by CFM international). Adapted from LEAP Engines – CFM international jet engines CFM international. Available from: https://www.cfmaeroengines.com/engines/leap/; 2017 [accessed: 11.02.17].

Heat rejection Heat exchanger 4

1

1 – 2: isentropic compression 2 – 3: isobaric heat addition 3 – 4: isentropic expansion 4 – 1: heat rejection

Power shaft Heat addition 3

2 Compressor

Type of the cycle Open: 1 – 2 – 3 – 4 Closed: 1 – 2 – 3 – 4 – 1 Note: thermodynamically open Gas turbine and closed cycles are equivalent

Heat exchanger or combustion chamber Fig. 12 A simple layout of a Brayton cycle.

4.7.6

Air – Standard Brayton Cycle

George Brayton initially suggested the Brayton cycle in the 19th century for the reciprocating oil-burning engine, which was also developed by him. However, today the Brayton cycle is only used for gas turbines where both compression and expansion processes occur in a rotating device [13]. The gas turbine employed in the cycle can be considered as a heat engine as it takes heat from the heat source and rejects it to a heat sink and generates useful work. In an actual combustion turbine, pressurized combustion gases are used as working fluid of the system. In these systems, to increase the combustion efficiency, excess air is supplied even more than 50 times the stoichiometric air rate. Therefore, assuming the properties of combustion gas as the properties of the air corresponding temperature and pressure would be a decent permissible approximation. For the cases when the air is modeled, the working fluid is modeled as ideal gas called “air standard” cycles. In the air standard cycle, the specific heat of air is assumed constant, which is determined at 251C [1]. A simple layout and T-s diagram of an air standard cycle can be seen in Figs. 12 and 13, respectively. Air is taken from ambient conditions at state 1 and pressurized to turbine inlet pressure. Ideally, that

220

Gas Turbine Cycles

3

T

4 2

1 s Fig. 13 T-s diagram of a Brayton cycle.

process is isentropic, which means it is both adiabatic and reversible. In state 2, high-pressure air enters the combustion chamber/ HEX and heats up at constant pressure. In state 3, combustion gases with high temperature and pressure expand in the gas turbine and the gas turbine generates mechanical work. The expansion is also isentropic in ideal conditions. For the closed systems, heat removal after the expansion is provided by a HEX at a constant pressure where the open systems discharge the combustion gases to the atmosphere after the turbine. Burning the working fluid causes the feed air requirement in the system. Hence, heat addition should also be done by a HEX in the closed systems. In terms of thermodynamic aspect, closed and open Brayton cycles are equivalent. To be precise, in the open Brayton cycle, the atmosphere acts as a cooler, where the cooling process 4-1 occurs at constant pressure. In the closed Brayton cycle, the same process is conducted in a HEX. Closed systems are usually preferred for the cases when the work output of the cycle is needed to be enhanced. Atmospheric conditions limit the outlet pressure of the turbine. In order to overcome that limitation, the cycle should be closed, so that the turbine can be expanded up to the vacuum [1]. T-s of an ideal air standard Brayton cycle under steady state conditions can be seen in Fig. 13. When the kinetic and potential energy changes in the working fluid are neglected and the ideal gas laws applied, balance equations can be expressed for a unit mass basis as follows [1]: Mass balance equations for compression, heat addition, expansion and heat rejection processes respectively: _ 2; MBE : m _1¼m

_2¼m _ 3; m

_3¼m _ 4; m

_4¼m _ 1; m

ð1Þ

here, m_ indicates the mass flow rate and the numbers used as indices show the state number in Fig. 13. Since the system is closed and steady, there is no external mass addition or rejection into the system. So all the processes have the same masses. Energy balance equations for compression, heat addition, expansion, heat rejection, and overall system, respectively: EnBE : h1 þ wc ¼ h2 ;

h2 þ qsr ¼ h3 ;

h3 ¼ h4 þ wt ;

h4 ¼ h1 þ qsi

wc þ qsr ¼ wt þ qsi

ð2Þ

here, h, w, and q represent the specific enthalpy, specific work, and specific heat values. si, sr, c, and t indicate the heat sink, heat source, compressor, and turbine, respectively. While the compressor work and the heat addition to the system are the energy inlet of the cycle, the turbine work and the heat rejection are the energy outlet of the cycle. According to the first law of thermodynamics, heat and work inlet should be equal to the heat and work exit of the system. Net useful work done by the system can be expressed as the difference between produced work by the turbine and consumed work by compressor: wnet ¼ wt

ð3Þ

wc

Heat inlet from the heat source and heat removal to the heat sink can be defined as follows: qsr ¼ h3

h2 ¼ cp ðT3

T2 Þ;

qsi ¼ h4

h1 ¼ cp ðT4

T1 Þ

ð4Þ

here h3 and h2 are the specific enthalpies of the stream after and before the heat addition respectively, and h4 and h1 represent the specific enthalpies of the air at the inlet and exit of the heat rejection. Entropy balance equations for compression, heat addition, expansion, heat rejection, and overall system respectively: EnBE: s3 þ sg;t ¼ s4 ;

s4 þ sg;si ¼ s1 þ

qsi ; T1

s1 þ sg;c ¼ s2 ;

Gas Turbine Cycles s2 þ

qso þ sg;sr ¼ s3 ; T3

qso qsi þ sg ¼ T3 T1

221 ð5Þ

here, “s” is for specific entropy and subindex “g” represents entropy generation. As the compression and expansion processes in the ideal Brayton cycle are isentropic, entropy generation for those processes equals to zero (sg,c ¼ sg,t ¼ 0 kj kg 1 K 1). On the other hand, heat transfer processes at both the heat sink and heat sources are internally reversible but externally irreversible for the ideal cycle. In other words, heat transfer from the system boundaries to the heat sink and heat source take place. Therefore, entropy production and exergy destruction occur in the cycle. Exergy balance equations for compression, heat addition, expansion, heat rejection, and overall system respectively:   T0 ExBE: ex 1 þ wc ¼ ex2 þ exd;c ; ex2 þ qsr 1 ¼ ex3 þ ex d;sr ; ex 3 ¼ ex 4 þ wt þ exd;t T3       T0 T0 T0 ð6Þ ex 4 ¼ ex 1 þ qsi 1 þ ex d;si ; wc þ qsr 1 ¼ wt þ qsi 1 þ exd T1 T3 T1 here, ex represents the specific exergy rate, T0 is reference temperature, and d indicates the exergy destruction. In the ideal Brayton cycle, exergy destruction for compression at the compressor and expansion at the turbine processes are also zero as those processes are adiabatic and reversible (exd, c ¼ exd,t ¼0 kj kg 1). As the compression process at the compressor and the expansion process at the turbine occur isentropically and the heat addition and the heat rejection processes are conducted at isobaric conditions. Those relations can be written under ideal gas conditions [13].  k 1=k  k 1=k T2 P2 T3 P3 c¼ ¼ ¼ ¼ ð7Þ T1 P1 T4 P4 here k is the specific heat ratio for the air, which shows the ratio between specific heat of the air at constant volume and constant pressure. The thermal efficiency of the Brayton cycle is the ratio between useful net work generated by the system and total energy input to the cycle. It can be shown as follows [13]: Z¼

wnet ¼1 qsr

qsi ¼1 qsr

cp ðT4 cp ðT 3

T1 Þ ¼1 T2 Þ

1Þ 1Þ

T1 ðT4 =T1 T2 ðT3 =T2

As the unit heat, addition, and rejection processes occur at constant volume, Eq. (10) can be rewritten as follows [13]:  k 1  k 1   T1 v2 v3 T4 ¼ ¼ ¼ T2 v1 v4 T3

ð8Þ

ð9Þ

By substituting the Eq. (10) into the Eq. (9), this equation can be simplified [13]. 1 rk 1

Z¼1

ð10Þ

here r is the compression ratio, which indicates the ratio between maximum and minimum pressures at the system. As it can be seen from the equation above, efficiency for an ideal air standard Brayton cycle is the only function of compression ratio. On the other hand, because of the cycle structure, some limitations should be defined on the heat sink and heat source temperatures for the cycle. Therefore, the temperature of the heat source (T3) must be always higher than the temperature at the compressor discharge (T2). As T2 ¼ T1(r)k 1/k , it turns out that for the cycle to function, one must satisfy the T34T1(r)k 1/k relation. Hence, for a Brayton cycle in which the compressor input and the maximum temperature conditions are known, it is possible to calculate the maximum compression ratio as a result of the following steps [14]: From Eq. (3) wnet ¼ wt

wc ¼ ðh3

h4 Þ

As it is an air standard cycle, we can rewrite the same equation as wnet ¼ cp ðT3

T4 Þ

ðT2

ðh2

h1 Þ

ð11Þ

T1 Þ

ð12Þ

After applying algebraic manipulations wnet ¼ cp T1

"

T3 T1

  T3 P1 ðk T1 P2

1Þ=k

 ðk P2 P1

1Þ=k

þ1

#

ð13Þ

From the equation above, it can be seen that, if the maximum, minimum temperature values in the cycle, and the cp are known, the power capacity of the cycle depends only on the compression ratio. In order to calculate the pressure ratio that maximizes the net work output, the net energy equation is partially differentiated according to the compression ratio and this value is equalized to zero. This point is the maximum point for net power. The expression giving the maximum point can be shown as follows [14]:    ∂wnet ∂ T3 T3 ð1 kÞ=k ¼ cp T1 r r ðk 1Þ=k þ 1 ð14Þ ∂r T1 T1 ∂r

222

Gas Turbine Cycles

cp T1

k

 1 T3 1=k r T1 k

r¼

r

2

 k½2ðk P2 T3 ¼ P1 T1

r

1=k



1ފ

ð15Þ

Exergy efficiency of the system can be introduced in different ways. The first expression of the exergy efficiency can be done as the ratio between useful work output to the exergy input at the heat source [1]. qsr qsi  ð16Þ c1 ¼
qsr 1 TT03 As the energy efficiency of the system can also be written as: wnet qsi qsr qsi Z¼ ¼1 ¼ qsr qsr qsr

ð17Þ

Eq. (16) can be rewritten as follows [1]: c1 ¼ 1

1

Z
 4Z

ð18Þ

T0 T3

As it can be seen in Eq. (18), exergy efficiency of the system is always higher than the energy efficiency. The second expression of the exergy efficiency of the cycle can be done regarding the received exergy to the cycle [1]. wnet ex
 d
 ð19Þ c2 ¼ ¼1 T0 ex in exout qsi 1 T0 qsr 1 T3

T1

here exin is the exergy input to the system at the heat source and ex out is the exergy output at the heat sink. A third way to define the exergy efficiency of the Brayton cycle is the ratio of the actual energy efficiency of the Brayton cycle and the efficiency of a reversible heat engine that runs between the heat sink and the heat source [1]: c3 ¼


Z 1

T1 T3

¼

1 1

qsi qsr T1 T3

ð20Þ

It should be noted that for all three expressions of the exergy efficiencies are equal to each other if the inlet temperature is equal to the ambient (reference) temperature: c1 ¼c2 ¼ c3 if T1 ¼ T0 [1]. Example 1: Assume a Brayton cycle in which air goes to the compressor at 100 kPa and 25oC temperature and is compressed by a compression ratio r¼ 15. Compression at the compressor occurs under isentropic conditions. After heat addition at constant pressure, the temperature of the stream reaches 800oC, then hot and pressurized air expands isentropically at the gas turbine and produces useful work. The air again enters the compressor after it rejects heat. Calculate the: a. b. c. d.

Thermodynamic properties of the air for each state. System overall energy efficiency. System overall exergy efficiency. Find the compression ratio value, which maximizes the net work for given conditions. Solution

a. Reference conditions T0 ¼251C, P0 ¼100 kPa State 1 (before the compression) Air at 100 kPa and 251C s ¼6.862 (kJ/kg K) h¼ 298.6 (kJ/kg) Initial conditions equal to conditions at State 1, therefore: ex1 ¼ 0(kj/kg) State 2 (after the compression at compressor) r¼

P2 ¼ 15-P2 ¼ 15  100ðkPaÞ ¼ 1500ðkPaÞ ¼ 1:5ðMPaÞ P1

From the equation for isentropic compression Eq. (8)    k 1=k T2 P2 T2 1500 ðkPaÞ 1:4 ¼ ¼ ¼ T1 P1 298:15 ðKÞ 100 ðkPaÞ

1=1:4

T2 ¼ 646:3 ðKÞ

Under ideal gas assumption at T¼ 643.3K specific enthalpy of the air h2 ¼ 656.3 kJ/kg Since the specific entropy of the stream remains constant at isentropic processes

Gas Turbine Cycles

223

s1 ¼ s2 ¼ 6:862 ðkJ=kg KÞ Compression work for unit basis analysis h1 ¼ 357:7 ðkJ=kgÞ

wc ¼ h2 Specific exergy ex 2 ¼ ðh2

ex2 ¼ 656:3 ðkj=kgÞ

298:6 ðkj=kgÞ

h0

T0  ðs2

s0 ÞÞ-ex2

298:2 ðKÞ  6:862 ðkj=kg KÞ

6:862ðkj=kg KÞ ¼ 357:7ðkJ=kgÞ

State 3 (heat addition at constant pressure) Temperature of the stream reaches 8001C after heat addition T3 ¼ 1073.15K Heat addition process occurs at constant pressure P3 ¼ P2 ¼ 1500 kPa Under ideal gas assumption at T3 ¼1073.15K specific enthalpy of the air h3 ¼ 1130 (kJ/kg) Air at 1500 kPa and 1073.15K s3 ¼7.439 (kJ/kg K) Heat addition at unit basis qsr ¼ cp ðT3

T2 Þ ¼ h3

h2 ¼ 474ðkJ=kgÞ

Specific exergy ex 3 ¼ ðh3 ex3 ¼ ð1130 ðkj=kgÞ

298:6 ðkJ=kgÞ

h0

T0  ðs3

s0 ÞÞ-ex3

298:2 ðKÞ  ð7:439 ðkJ=kgKÞ

6:862 ðkj=kgKÞÞ ¼ 660 kJ=kg

Here heat addition is assumed done by a HEX. Since no chemical reaction takes place, chemical exergies can be neglected for the system. State 4 (expansion at the turbine) Air expands to pressure level it has at State 1 P1 ¼ P4 ¼ 100 kPa From Eq. (8), which shows the temperature and pressure relation for isentropic expansion    k 1=k T3 P3 1073:15 ðKÞ 1500 ðkPaÞ ð1:4 1=1:4Þ ¼ ¼ ¼ ¼ T4 ¼ 495 ðKÞ T4 P4 T4 100 ðkPaÞ Under ideal gas assumption at T4 ¼495K specific enthalpy of the air h4 ¼498.3 kJ/kg Since the specific entropy of the stream remains constant at isentropic processes s3 ¼ s4 ¼ 7:434 ðkJ=kg KÞ Specific exergy ex 4 ¼ ðh4 ex4 ¼ ð498:3ðkJ=kgÞ

298:6 ðkJ=kgÞ

h0

T0  ðs4

s0 ÞÞ-ex4

298:2 ðKÞ  ð7:439 ðkJ=kgKÞ

6:862 ðkJ=kgKÞÞ ¼ 27:9 kJ=kg

Heat rejection at the HEX for unit basis analysis qsi ¼ h4

h1 ¼ cp ðT4

T1 Þ ¼ 498:3 ðkJ=kgÞ

298:6 ðkJ=kgÞ ¼ 199:7 ðkJ=kgÞ

The specific work done by the gas turbine: wt ¼ h3

h4 ¼ 1130 ðkJ=kgÞ

498:3 ðkJ=kgÞ ¼ 632:1 ðkJ=kgÞ

Specific net work performed by the cycle wnet ¼ wt

wc ¼ 632:1 ðkJ=kgÞ

357:7 ðkJ=kgÞ ¼ 274:4 ðkJ=kgÞ

224

Gas Turbine Cycles

Table 1

Thermodynamic properties of the air in Example 1 for each state point

State #

T (K)

P (kPa)

h (kJ/kg)

s (kJ/kg-K)

ex (kJ/kg)

1 2 3 4

298.2 646.3 1073 495

100 1500 1500 100

298.6 656.3 1130 498.3

6.862 6.862 7.439 7.439

0 357.7 660 27.9

700 650

Wnet Wt Wc

600 550 500

kJ/kg

450 400 350 300 250 200 150 100 50 0 0

5

10 r

15

20

Fig. 14 Change in net work of the cycle with respect to variation of the compression.

b. The overall efficiency of the system can be defined as: Z¼

wnet ¼1 qsr

qsi qsr qsi 474 ðkJ=kgÞ 199:7 ðkJ=kgÞ ¼ ¼ ¼ 0:5788-%57:88 qsr qsr 474 ðkJ=kgÞ

c. Overall exergy efficiency c1 ¼ 1

1

Z
 ¼1 T0 T3

1

0:5788
 ¼ 0:8014-%80:14 298:2 ðKÞ 1073 ðKÞ

d. The relation shows the maximum compression ratio for the cycle (Eq. (15))

r¼

r¼

 k½2ðk P2 T3 ¼ P1 T1

  P2 1073:15 ðKÞ 1:4½2ð1;4 ¼ 100 ðkPaÞ 298:15 ðKÞ

1ފ

1ފ

¼ 9:41-P2 ¼ 941 ðkPaÞ

The thermodynamic properties of the air in Example 1 for each state point are tabulated in Table 1. In Fig. 14, effect of the compression ratio of the power capacity of the cycle is presented. As can be clearly seen in the figure, slight changes in the compression at lower pressures affect drastically the net power. By the increase of the compression ratio, the heat addition and heat removal processes have increased also. This increase in wnet is continued up to the compression ratio of 9.4. After r ¼9.4 rise in the heat rejection causes a reduction in net work capacity of the system. The parametric study for the question can be seen in Fig. 15. All parameters of the question were kept constant and the impact of the compression ratio (r) on the energy and exergy efficiencies were investigated. As the compression ratio increases, the power capacity of the turbine increases. Moreover, since the turbine inlet temperature is constant, the turbine outlet temperature has also decreased with increasing compression ratio (see Eq. (8)). It should be noted that the compressor outlet temperature does not exceed 8001C and the turbine outlet temperature is not lower than 251C for the parametric study.

Gas Turbine Cycles

225

100 90 ψ 80

η

70

%

60 50 40 30 20 10 0

0

2

4

6

8

10

12

14

16

18

20

r Fig. 15 Change in exergy and energy efficiency with respect to variation of the compression ratio.

4.7.6.1

Actual Brayton Cycle

The actual GTCs do not take place as idealized in the Brayton cycle. Compression and expansion processes are not isentropic, and pressure and friction losses are inevitable. Moreover, actual compressor work is higher than the ideal compressor work where ideal turbine work is lesser than the ideal turbine work. So for a real gas turbine, there is no process that exists without entropy change. Various definitions were developed to evaluate the performance of each element in the system. For example, the performance of the compressor and the gas turbine is assessed by isentropic efficiency. For the compressor, isentropic efficiency is the ratio between actual work and isentropic work (adiabatic and reversible work in other words). Similarly, isentropic efficiency for the compressor is the ratio of the isentropic work required by the compressor to the actual work needed by the compressor. For HEXs, which provide heat input and heat rejection in a closed Brayton cycle, performance evaluation is done with effectiveness or thermal ratio. The “effectiveness” is defined as the ratio of the temperature rise in the cold stream to the temperature difference between the inlet temperatures of the hot and cold streams in a HEX [1,13,15]. The performance criteria of the system elements can be calculated as follows: Isentropic efficiency of the compressor Zc ¼ ðh2s

h1 Þ=ðh2

h1 Þ

ð21Þ

where h2s represents the specific enthalpy of the air at the exit of the compressor under isentropic conditions and h2 is the specific enthalpy of it for the actual case. Isentropic efficiency of the gas turbine Zt ¼ ðh3

h4 Þ=ðh3

h4s Þ

ð22Þ

here h4s is the specific enthalpy of the air at the exit of the gas turbine if the gas turbine expands isentropically and h2 is the specific enthalpy of it for the actual case and h4 is the actual specific enthalpy for it. Effectiveness of the HEXs ðTce

Tci Þ=ðThi

Tci Þ

ð23Þ

here Tce and Tci indicates the inlet and exit of the cold side of the HEX, where Thi represents the hot side in the HEX. One of the irreversibilities in actual gas turbine cycle applications is pressure losses in HEXs. However, the effects of these losses on the entire system are negligible. Therefore, the pressure losses in the HEX are neglected in the sample questions in this section. “T-s” diagrams are shown in Figs. 16 and 17 for the situations where pressure losses are neglected and taken into account respectively [14]. For the cases where pressure losses are too small compare to the pressure of the stream, the diagram in Fig. 17 can be followed for rough calculations. In order to get more accurate results, all the losses through the system should be taken into account and the diagram in Fig. 16 should be followed. Example 2: Reconsider the Brayton cycle in Example 1. Assume these time compression and expansion processes in the compressor occur irreversible adiabatically and their inlet conditions for both gas turbine and compressor are the same as in Example 1. Isentropic efficiencies of the turbine and compressor are 80% and 85%, respectively. Calculate:

226

Gas Turbine Cycles

T

3

P = constant 2

4

P = constant 1 s Fig. 16 T-s diagram of an actual closed Brayton cycle.

T

3 P = constant 2s 2

4s

1

4

P = constant

s

Fig. 17 T-s diagram of a closed Brayton cycle, which neglects the pressure losses in the heat exchangers (HEXs).

a. Net work produced by the cycle. b. Energy efficiency of the cycle. Solution a. Actual compressor work can be determined from the definition of the isentropic efficiency of the compressor Zc ¼ ðh2s

h1 Þ=ðh2

h1 Þ ¼ wcs =wc

here wcs indicates the compressor work under isentropic conditions and wc represents the actual compressor work.

Zc ¼ 0:85 ¼

357 ðkJ=kgÞ -wc ¼ 420:8 ðkJ=kgÞ wc

Gas Turbine Cycles

Heat rejection

227

Heat exchanger 5 Gas turbine

1 6

Power shaft

Regenerator

2

3

Compressor

Generator

Heat addition 4

Heat exchanger or combustion chamber Fig. 18 Regenerative Brayton cycle.

wc ¼ h2

h1 ¼ 420:8 ðkJ=kgÞ ¼ h2

298:6 ðkJ=kgÞ-h2 ¼ 719:4 ðkJ=kgÞ

Similarly, to calculate the actual turbine work Zt ¼ 0:8 ¼ ðh3

h4 Þ=ðh3

h4s Þ ¼ ðwt =wts Þ

here wts is the turbine work under isentropic conditions and wt indicates the actual turbine work. Zt ¼ 0:8 ¼ ð1130 ðkJ=kgÞ

h4 Þ=ð1130 ðkJ=kgÞ

wt ¼ 505 ðkJ=kgÞ wc ¼ 505 ðkJ=kgÞ

wnet ¼ wt

498:3 ðkJ=kgÞÞ ¼ ðwt =wts Þ

h4 ¼ 624:7 ðkJ=kgÞ 420:8 ðkJ=kgÞ ¼ 63:12ðkJ=kgÞ

b. As the energy efficiency is the ratio between net work produced by the system to the heat addition from the heat source, heat addition to the system at constant pressure should be calculated. h2 ¼ 1130 ðkJ=kgÞ

qsr ¼ h3

Z¼

4.7.6.2

719:4 ðkJ=kgÞ ¼ 410:9 ðkJ=kgÞ

wnet 63:12ðkJ=kgÞ ¼ ¼ 0:2064-20:64% qsr 410:0 ðkJ=kgÞ

Regenerative Brayton Cycle

Usually, the temperature of the air that leaves the turbine is considerably higher than the air temperature taken from the compressor. That means a significant amount of energy is rejected to the ambient. Energy losses cause lower efficiency for the system. In order to overcome that problem, some methods are developed. Regeneration is one of the methods for the heat recovery. It utilizes the turbine exhaust as a preheater for the compressed air. The cycle layout and T-s diagram of the regenerative Brayton cycle is presented in Figs. 18 and 19, respectively. The only additional component to the basic Brayton cycle is the regenerator after the compressor. For the reversible air standard model, heat addition in the regenerator takes place at constant pressure as in Fig. 18. Hence the following calculations can be done to determine the temperature of the streams in the regenerator [1]. Energy balance for the regenerator EnBE: (h3–h2 ¼ h5–h6). Since it is air standard model and heat addition takes place at constant pressures h3

h2 ¼ h5

h6 ¼ cp ðT3

T2 Þ ¼ cp ðT5

T6 Þ

ð24Þ

Under the assumption of no internal irreversibility for the regenerator, it could be said that the temperature profile of the streams that flow through the regenerator should have a perfect match. In other words, the temperature after that regenerator outlet should be equal to the temperature of the turbine outlet (T3 ¼ T5). The following implication can be done from Eq. (24). T3

T2 ¼ T5

T6 ;

T3 ¼ T5 -T2 ¼ T6

228

Gas Turbine Cycles

4

T

P = r P0

3

5

Regeneration

2

6 P = r P0

1 s Fig. 19 T-s diagram of a reversible air standard regenerative Brayton cycle.

As in the air standard Brayton cycle, energy efficiency of the cycle can be expressed as: Z¼

wnet ¼1 qsr

qsi ¼1 qsr

h6 h4

h1 ¼1 h3

cp ðT6 cp ðT4

T1 Þ T3 Þ

Since it is an internally reversible cycle all the expansions and compression processes are isentropic. Final form of the relation for the efficiency turns into [1]:

 T1   T1 T2 1 cp ðT6 T1 Þ T2 T1 k 1

 ¼ 1 ¼1 ð25Þ rk Z¼1 T5 cp ðT4 T3 Þ T4 T4 T1 1 T4

In the real cases, turbine outlet temperature T5 is always higher than the regenerator outlet T3, T54T3. In order to evaluate the performance of a regenerator, its effectiveness “A” is used. The effectiveness of a regenerator is the ratio between actual heat transfer from the exhaust to the compressed air and ideal heat transfer from the exhaust to the compressed air. A¼

qreg ðh3 ¼ qreg; id ðh5

cp ðT3 h2 Þ ¼ cp ðT5 h2 Þ

T2 Þ T2 Þ

ð26Þ

A system with higher regeneration effectiveness naturally has greater efficiency. However, making a product with high effectiveness results in increased cost of the component and higher pressure drop in the regenerator. Therefore, in real applications, regenerator efficiency is below 0.85 [13]. In order to evaluate the impact of the regeneration system on the Brayton cycle, an Z efficiency improvement factor can be introduced as EIF ¼ regeneration [1]. Zbasic Here, Zregeneration and Zbasic show the efficiency of the Brayton cycle with the basic configuration and regeneration, respectively. Example 3: Consider a regenerative Brayton cycle in which air goes to the compressor at 100 kPa and 251C temperature and is compressed by a compression ratio r¼10. Compression at compressor occurs under adiabatic conditions with Zc ¼ 0.85. The compressed air flows through the regenerator whose effectiveness is A¼ 0.9. After the heat addition at constant pressure, temperature of the stream reaches 8001C. Afterward hot and pressurized air expands adiabatically at the gas turbine with Zt ¼0.80. Calculate the a. b. c. d.

Thermodynamic properties of the air for each state. System overall energy efficiency. System overall exergy efficiency . Efficiency improvement factor (EIF). Modeling Assumptions

• • • •

Steady-state, steady flow. Compressor and gas turbine work adiabatically. No pressure drop through the HEXs. Air standard model is implemented for the working fluid.

Gas Turbine Cycles

229

3

T

4a

5

Regeneration

2a

1 s Fig. 20 T-s diagram of the regenerative air standard regenerative Brayton cycle described in Example 3.

• •

Kinetic and potential energy effects are neglected. Reference conditions T0 ¼ 251C, P0 ¼ 100 kPa. Solution T-s diagram of the air in Example 3 is shown in Fig. 20.

a. State 1 (before the compression) Air at 100 kPa and 25oC s1 ¼ 6:862 ðkJ=kg KÞ-h1 ¼ 298:6 ðkJ=kgÞ Initial conditions equals to conditions at State 1, therefore: ex1 ¼ 0 ðkJ=kgÞ State 2 (after the compression at compressor) r¼

P2 ¼ 10-P2 ¼ 10  100 ðkPaÞ ¼ 1000 ðkPaÞ ¼ 1:0 ðMPaÞ P1

From the equation for isentropic compression Eq. (8)    k 1=k T2s P2 T2s 1000 ðkPaÞ 1:4 ¼ ¼ ¼ T1 P1 298:15 ðKÞ 100 ðkPaÞ

1=1:4

T2s ¼ 575:6 ðKÞ

Under ideal gas assumption at T2s ¼575.6K specific enthalpy of the air h2s ¼581.8 kJ/kg. Actual state properties of air after compression can be determined from the definition of the isentropic efficiency of the compressor Zc ¼ ðh2s 0:85 ¼ ð581:8 ðkJ=kgÞ

h1 Þ=ðh2

h1 Þ

298:6 ðkJ=kgÞÞ=ðh2

298:6 ðkJ=kgÞÞ

h2 ¼ 631:8 ðkJ=kgÞ Under ideal gas assumption at h2 ¼ 631.8 (kJ/kg) temperature of the air T2 ¼623.2K Specific entropy of air at T2 ¼623.2(K), P2 ¼ 10,000 (kPa), s2 ¼ 6.954 (kJ/kg K) Compression work for unit basis analysis wc ¼ h2

h1 ¼ 631:8 ðkJ=kgÞ

298:6 ðkJ=kgÞ ¼ 333:2 ðkJ=kgÞ

Specific exergy ex 2 ¼ ðh2

h0

T0  ðs2

s0 ÞÞ-ex2

230

Gas Turbine Cycles ex 2 ¼ ð631:8 ðkJ=kgÞ

298:6 ðkJ=kgÞ

298:2 ðKÞ  ð6:954 ðkJ=kgKÞ

6:862 ðkJ=kgKÞÞ

¼ 305:8 ðkJ=kgÞ State 3 (heat addition at constant pressure) Temperature of the stream reaches 8001C after heat addition T3 ¼1073.15K Heat addition at constant pressure P3 ¼ P2 ¼ 10,000 kPa Under ideal gas assumption at T3 ¼1073.15K specific enthalpy of the air h3 ¼1130 (kJ/kg) Air at 10,000 kPa and 1073.15K s3 ¼7.555 (kJ/kg K) Specific exergy ex 3 ¼ ðh3 ex 3 ¼ ð1130 ðkj=kgÞ

298:6 ðkJ=kgÞ

State 4 (expansion at the turbine) Air expands to pressure level of State 1

T0  ðs3

h0

s0 ÞÞ-ex3

298:2 ðKÞ  7:555 ðkJ=kgKÞ

6:862 ðkj=kgKÞÞ ¼ 625:2 ðkJ=kgÞ

P1 ¼ P4 ¼ 100 kPa

From Eq. (8), which shows the temperature and pressure relation for isentropic expansion    k 1=k T3 P3 1073:15 ðKÞ 1000 ðkPaÞ ð1:4 1=1:4Þ ¼ ¼ ¼ ¼ T4s ¼ 556 ðKÞ T4s P4 T4s 100 ðkPaÞ

Under ideal gas assumption at T4s ¼ 556 (K), specific enthalpy of the air h4s ¼ 561.2 (kJ/kg) T4s and h4s represent the state properties of the air under isentropic conditions. Actual state properties of the air after expansion can be determined from the definition of the isentropic efficiency of the gas turbine Zt ¼ ðh3

h4 Þ=ðh3

h4s Þ

h4 Þ=ð1130 ðkJ=kgÞ

0:8 ¼ ð1130 ðkJ=kgÞ

561:2 ðkJ=kgÞÞ

h4 ¼ 646:5 ðkJ=kgÞ

Under ideal gas assumption at h4 ¼ 646.5 (kJ/kg) temperature of the air T4 ¼ 637.1K Specific entropy of air at T4 ¼ 637.1 (K), P4 ¼ 1000 (kPa), s4 ¼ 7.64 (kJ/kg K) Specific exergy ex 4 ¼ ðh4 ex 4 ¼ ð646:5 ðkJ=kgÞ

298:6 ðkJ=kgÞ

The specific work conducted by the gas turbine: wt ¼ h3

T0  ðs4

h0

298:2 ðKÞ  ð7:64 ðkJ=kgKÞ

h4 ¼ 1130 ðkJ=kgÞ

Specific net work performed by the cycle: wnet ¼ wt

s0 ÞÞ-ex4

wc ¼ 483:8 ðkJ=kgÞ

6:862 ðkJ=kgKÞÞ ¼ 116:2 kJ=kgÞ

646:5 ðkJ=kgÞ ¼ 483:8ðkJ=kgÞ 333:2ðkJ=kgÞ ¼ 150:6 ðkJ=kgÞ

State 5 (regeneration) Specific enthalpy of the air after regeneration can be determined by the definition of effectiveness A¼ A¼

h5 h4

h2 ¼ 0:9 h2

h5 631:8 ðkJ=kgÞ ¼ 0:9 646:5 ðkJ=kgÞ 631:8 ðkJ=kgÞ h5 ¼ 645:1 ðkJ=kgÞ

Under ideal gas assumption at h5 ¼ 645.1 (kJ/kg) temperature of the air T5 ¼ 635.8K Specific entropy of air at T5 635.8 (K), P5 ¼ 10,000 (kPa), s5 ¼6.976 (kJ/kg K) Specific exergy ex 5 ¼ ðh5 ex 5 ¼ ð645:1 ðkJ=kgÞ

298:6 ðkJ=kgÞ

h0

T0  ðs5

s0 ÞÞ

298:2ðKÞ  6:976 ðkJ=kg KÞ ¼ 312:7 kJ=kg

6:862 ðkJ=kg KÞÞ

The thermodynamic properties of the air in Example 3 for each state point are given in Table 2.

Gas Turbine Cycles

Table 2

231

Thermodynamic properties of the air in Example 3 for each state point

State #

T (K)

P (kPa)

h (kJ/kg)

s (kJ/kg-K)

ex (kJ/kg)

1 2s 2 3 4s 4 5

298.2 575.6 623.2 1073 555.8 637.1 635.8

100 1000 1000 1000 100 100 1000

298.6 581.8 631.8 1130 561.2 646.5 645.1

6.862 6.862 6.954 7.555 7.555 7.64 6.976

0 283.2 305.8 625.2 56.06 116.2 312.7

b. Energy efficiency of the cycle can be expressed as: wnet qsr

Z¼

The heat input of the cycle can be determined to calculate the efficiency. qsr ¼ h3 qsr ¼ 1130 ðkJ=kgÞ

c. Exergy efficiency of the cycle: c1 ¼ 1 d. EIF ¼

Zregeneration Zbasic

h5

645:1 ðkJ=kgÞ ¼ 485:3 ðkJ=kgÞ

Z¼

wnet 150:6 ¼ ¼ 0:31-31% qsr 485:3

1

Z
 ¼1 T0 T3

0:31


298:2 ðKÞ 1073 ðKÞ

1

 ¼ 0:43-%43

Zbasic is the efficiency of the cycle without regeneration. Zbasic ¼

wnet;b qsr;b

Installing a regenerator does not affect the work output of the cycle. The regenerator leads to a decrease in heat input. Hence, net work produced by the cycle is the same for both cases. Heat addition for the basic configuration: qsr;b ¼ h3 qsr;b ¼ 1130 ðkJ=kgÞ Zbasic ¼ EIF ¼

h2

631:8 ðkJ=kgÞ ¼ 498:5ðkJ=kgÞ

150:6 ðkJ=kgÞ ¼ 0:30-30% 498:5 ðkJ=kgÞ Zregeneration 0:30 ¼ 1:027 ¼ Zbasic 0:31

The change of the efficiency improvement factor (EIF) for the cycle with the compression ratio can be observed in Fig. 21. As the compression ratio decreases, the temperature following compression also decreases. Higher temperature differences in the regenerator result in a greater heat transfer rate as well. More heat transfer in the regenerator leads to a rise in EIF.

4.7.6.3

Reheat-Regenerative Brayton Cycle

As stated in equation, the net work of a gas turbine cycle is the difference between turbine and compressor work. To improve the net work output, the turbine work output should be increased, or the compressor work input should be decreased. In this chapter, increasing the net power production with reheat is investigated. Reheat does not only lead to rising in turbine work, but it also causes the higher temperature in turbine exhaust. Hence, utilizing reheat with regeneration could improve the system capacity and efficiency altogether [1,13]. Simple layout and T-s diagram of a Brayton cycle with reheating and regeneration are presented in Figs. 22 and 23, respectively. The unique change in this system is the multistage turbine power output. Additional heat input to the working fluid takes place between the gas turbine stages. As a result, the expansion of the air in the turbine occurs under pseudoisothermal conditions. Increasing the number of stages makes the expansion approach the isothermal expansion profile. The complete cycle in Fig. 23 consists of eight processes. First, air is taken to the compressor. In an open cycle, air enters to the compressor at atmospheric conditions, which is usually close to 1 bar for many cases. However, for the closed-cycle, the pressure of

Gas Turbine Cycles

232

3.5

3.0

EIF

2.5

2.0

1.5

1.0

0.5 2

3

4

5

6

7

8

9

10

rc Fig. 21 Variation of the efficiency improvement factor (EIF) with the compression ratio for the cycle described in Example 3.

Heat rejection Heat exchanger Gas turbine

5

1

7

Gas turbine

8 Power shaft

2

Heat addition

Heat addition

Regenerator

Generator 3

4

3

Compressor

6 Heat exchanger or combustion chamber

Reheater

Fig. 22 Brayton cycle with reheating and regeneration.

the inlet air depends on the pressure of the gas turbine’s exit pressure of the previous stage. After the compression, compressed air flows through the regenerator for preheating and passes through another HEX/combustion chamber that supplies external heat at isobaric conditions. The combustion chamber option is only valid for the open cycle. Gas at high pressure and temperature then enters the first gas turbine stage and generates useful work. Partially expanded air enters another HEX/combustion chamber where its temperature reaches the same level as the inlet temperature of the first gas turbine (T6 ¼ T4). Reheated stream runs the second turbine and the air expands till the pressure reaches the compressor inlet pressure. The final stage is isobaric heat rejection in which relatively hot air rejects its heat at constant pressure and ends up with air reaching the same temperature as the compressor inlet. In brief, the descriptions and state points for ideal standard air reheating regeneration cycle can be defined as follows: 1–2 Compression at isentropic conditions ds ¼ 0-s1 ¼ s2 k

Pv ¼ constant; h2

P1 v1 ¼ RT1 ; h1 ¼ cp ðT2

P2 ¼ rP1 -v1k ¼ rv2k ; T1 Þ;

T1 v1k

EnBE : h1 þ wc ¼ h2

1

¼ T2 v2k

1

Gas Turbine Cycles

T

4

6

3

7

5 8 2

1

s

Fig. 23 T-s diagram of a Brayton cycle with reheating and regeneration.

2–3 Heat addition at isobaric conditions dP ¼ 0-P3 ¼ P2 ;

h3

h2 ¼ cp ðT3 ds ¼ cp

P3 v3 ¼ RT3 ;

EnBE : h3

T2 Þ;

dT -s3 T

s2 ¼ cP ln

h2 ¼ h7

h8

  T3 T2

3–4 Heat addition at isobaric conditions dP ¼ 0-P4 ¼ P3 ;

T4 ¼ T1 =t; ds ¼ cp

P4 v4 ¼ RT4 ;

h2 ¼ cp ðT4

h4

dT -s4 T

s2 ¼ cP ln

T2 Þ

  T4 T2

4–5 Expansion at isentropic conditions ds ¼ 0-s5 ¼ s4 P4 ¼ P5 r 0:5 ; h5 5–6 Heat addition at isobaric conditions dP ¼ 0-P6 ¼ P5 ;

P4 v4k ¼ P5 v5k ;

h1 ¼ cp ðT5

T4 v4k

1

¼ T5 v5k

1

EnBE-h5 þ wt;1 ¼ h4

T1 Þ;

assuming that temperature reaches maximum after reheat T6 ¼ T4

P6 v6 ¼ RT6 ;

h6

h5 ¼ cp ðT6

T5 Þ;

s6

s5 ¼ cp ln

  T6 T5

6–7 Expansion at isentropic conditions ds ¼ 0-s7 ¼ s6 P7 ¼ P1 ; h7 3–4 Heat rejection at isobaric conditions

P7 v7k ¼ P6 v6k

h1 ¼ cp ðT7

dP ¼ 0-P8 ¼ P7 ; h7

h8 ¼ cp ðT7

T1 Þ;

T7 v7k

1

¼ T6 v6k

EnBE-h6 ¼ h7 þ wt;2

T8 ¼ T2 ðassumedÞ; T8 Þ

1

s7

s8 ¼ cp ln

P8 v8 ¼ RT8   T7 T8

233

234

Gas Turbine Cycles

T 6

8

qin

4

9

7

5

2

10

qout 3

1

s Fig. 24 Brayton cycle with an intercooler.

8–1 Heat rejection at isobaric conditions dP ¼ 0-P8 ¼ P1 ;

EnBE-qout ¼ h8

h1

One should note that it is very crucial to select the interstage pressure (P5 ¼P6) for expansion to maximize the overall turbine work output. The specific work output by the turbines of 2 stage cycles can be given as [1]:
k1  k 1 1 k 2 w ¼ k R T4 r k þ r k r1k

where

r¼

P2 P4 and r1 ¼ P1 P5

If the equation above were differentiated with respect to r1 (i.e., the result is equal to zero), then the relationship between r and r1, which provides the maximum work output, would be found. In other words, the r1 value that satisfies the equation k 1 k 1 k 1 r1k þ r k r1k ¼ 0, gives the highest overall generated work, which is r1 ¼ r0.5. This condition is also valid for intercooling stage compression [1].

4.7.6.4

Brayton Cycle With Intercooler

Applications of the intercooling stage for compression processes positively influence the overall energy and exergy efficiencies of the cycle. Two main points highlight this idea: firstly, the work inlet for the compressor reduces. Secondly, the capacity of the heat recovery increases with regeneration [1]. A simple scheme of a Brayton cycle with regeneration, reheat is presented in Fig. 24 and the T-s diagram of the cycle is provided in Fig. 25. In Fig. 25, it can be seen that after the intercooling process, the temperature of the working fluid drops to the same level as in state 1. This way, the cycle’s working fluid exits from the compressor at relatively lower temperature and by the help of reheating, it exits from the turbine at higher temperature. Large temperature differences between the compressor and turbine outlets lead to high heat transfer rate through the regenerator, which increases the cycle’s efficiency. As explained in the previous section, pressure ratios through the stages at reheating and intercooling should be optimized according to the r1 ¼ r0.5 relationship. Increasing the reheating and intercooling stages will improve the performance of the system, but it would not be commercially practical after a certain number of stages. Let us consider a hypothetical Brayton cycle, which has an infinite number of intercooling and regeneration stages. In that case, the compression and expansion of the working fluid occur under pseudoisothermal conditions, and the cycle becomes similar to the Ericsson cycle (see Fig. 25). The efficiency of the cycle approaches Carnot limit since all the heat is supplied at the highest temperature and all the heat rejection occurs at the lowest temperature. Example 4: Consider a reheat-intercooling Brayton cycle, which has two stages for both expansion and compression and is operated at ideal conditions. (All compression and expansion processes are reversible and adiabatic; pressure losses at the heat additions and rejections are neglected. After reheat, air reaches maximum temperature and the regenerator effectiveness, A¼ 1.) Suppose that air enters the compressor at atmospheric conditions (251C, 101.325 kPa), compressed by an overall pressure ratio r¼ 10 and reaches its maximum temperature after isobaric heat addition (temperature ratio, t¼ 0.2Þ Find: a. The back work ratio of the cycle.

Gas Turbine Cycles

235

T

Tavg

Tavg s Fig. 25 Brayton cycle with several intercoolers.

b. The thermal efficiency of the cycle. Modeling assumptions

• • • • • • •

Steady-state, steady flow. Compressor and gas turbine work isentropically. No pressure drop through the HEXs with effectiveness (A¼1). Air standard model is implemented for the working fluid. Kinetic and potential energy effects are neglected. Reference conditions T0 ¼251C, P0 ¼ 101.325 kPa. Regenerator effectiveness (A¼1).

a. The back work ratio of the cycle As the cycle operates under ideal conditions, pressure ratios between stages should be r1 ¼r0.5 in order to maximize the turbine work output and minimize the compressor work inlet. Hence, P2 P4 P6 P8 ¼ ¼ ¼ ¼ r1 ¼ r 0:5 ¼ 90:5 ¼ 3 P1 P3 P7 P9 The working fluid enters the second stage of the compressor and gas turbine at the temperature that it has at the first stage of those components and leaves at the same exit temperature, which means: T1 ¼ T3 ;

T2 ¼ T4 ;

T6 ¼ T8 ;

T7 ¼ T9

In addition, since the compressor and gas turbines work isentropically, specific enthalpies of the state points should be identical: h1 ¼ h3 ; P1 ¼ 101:325 kPa;

h2 ¼ h4 ;

h6 ¼ h8 ;

h7 ¼ h9 ;

T1 ¼ 25 þ 273:15 ¼ 298:15 K-h1 ¼ 298:6 ðkJ=kgÞ

P2 ¼ r1 P1 -P2 ¼ 3  101:325 ðkPaÞ ¼ 303:975 ðkPaÞ Isentropic compression of the air: T2 P2 ¼ T1 P1

k 1 k

-

T2 303:975 ðkPaÞ ¼ 298:15 ðKÞ 101:325 ðkPaÞ

For ideal gas assumption at T2 ¼ 408:1 ðKÞ; h2 ¼ h4 ¼ 409:5 ðkJ=kgÞ

1:4 1 1:4

-T2 ¼ 408:1 ðKÞ

236

Gas Turbine Cycles

After the second stage, the compression and the heat addition are at constant pressure while the temperature of the working fluid reaches the highest value: t ¼ 0:2 ¼

T1 298:15 ðKÞ ¼ -1491 ðKÞ-Pr6 ¼ 585:4 T6 T6 h6 ¼ 1625 ðkJ=kgÞ

After the expansion of the working fluid at the first stage of the gas turbine: Pr7 ¼

P7 1 Pr6 ¼ 585:4; 3 P6

 wcompi ¼ 2 wcomp; i c ¼ 2ðh2

 wturo ¼ 2 wtur; rh ¼ 2ðh6 wnet ¼ wturo

h7 ¼ h9 ¼ 1211 ðkJ=kgÞ h1 Þ ¼ 2ð409:5

h7 Þ ¼ 2ð1625

wcompi ¼ 828

298:6Þ ¼ 222 ðkJ=kgÞ 1211Þ ¼ 828 ðkJ=kgÞ

222 ¼ 606 ðkJ=kgÞ

Then the back work ratio becomes: 222 ¼ 0:268 828

rbw ¼ b. The overall energy efficiency of the cycle

Z¼

wnet qin

Since the effectiveness of the regenerator of the cycle is A¼ 1, the temperatures and specific enthalpies of the state points 5, 7, and 9 are identical. Thus, the heat input to the cycle becomes: qsr ¼ ðh6

h4 Þ þ ðh8

h7 Þ ¼ ð1625 Z¼

4.7.6.5

409:5Þ þ ð1625

1211Þ ¼ 1630 ðkJ=kgÞ

wnet 606 ¼ 0:372 ¼ 37:2% ¼ qin 1630

Exergy Destructions in Brayton Cycle Power Plants Estimation

The assessment of the second law analysis of a power cycle is crucial to see the power output and efficiency of the system. The ideal Ericsson, Carnot, and Stirling cycles are completely reversible. Hence, there is no exergy destruction that is accounted to them. However, for the Brayton cycle, external irreversibilities are unavoidable even for the ideal case. A comprehensive exergy analysis of the cycle discloses where the major irreversibilities take place and where to start system enhancements [1,13]. The isentropic efficiencies of turbines and compressors have a large influence on the system performance because of the poor back work ratios of the Brayton cycle. Thus, an economically feasible Brayton cycle should consist of turbomachineries with at least 80% isentropic efficiency. Other system elements that are important to the irreversibilities are HEXs and ducts among the cycle. The temperature of the surfaces where heat transfer occurs, heat and pressure losses through the components turn into irreversibilities [1]. Therefore, this section aims to cover the irreversibilities of the Brayton cycle and explain related case studies. As stated in the previous sections, isentropic efficiency of a compressor indicates the ratio between the reversible (ideal/ isentropic) work inlet of the compressor and the actual work inlet of the compressor and it takes into account several irreversibilities that take place in the compressor. Exergy destructions are higher in compressors compared to turbines due to natural dissipative phenomena, which also leads to lower weak compression waves in compressors [1]. The properties of the working fluid also have an effect on the quantity of dissipations, especially the dissipations caused by the specific heat. As the capacity of energy and the specific heat are higher, the capacity of accumulation of energy as internal energy in the working fluid is higher as well. This situation causes a change in the exergy and the temperature of the exit stream of the compressor. Therefore, the determination of the specific heat is critical in order to calculate irreversibilities. In a typical Brayton cycle, air–fuel ratio (AF) of the cycle is higher than the stoichiometric conditions, which means there is excess air in the stream. Hence, it makes choosing the properties of air as the properties of the stream a useful assumption for the calculations of the Brayton cycle. However, for more accurate results, one should also consider the variation of the specific heat with the temperature [1]. Standard air approach assumes the specific heat of air to be constant and that it does not change with

Gas Turbine Cycles

Table 3

237

Variation of the thermodynamic properties and exergy destruction of the working fluid with compression ratio for air standard model

rp

exd,c (kJ/kg)

T2 (K)

T2i (K)

cp (kJ/kg K)

cp (kJ/kg K)

cv,1 (kJ/kg K)

cv,2 (kJ/kg K)

k1

k2

1.5 2 3 4 5 6 7 8 9 10 11 12 13 14 15

5.726 9.351 13.93 16.84 18.94 20.55 21.85 22.93 23.86 24.66 25.37 26 26.57 27.09 27.56

341.2 374.8 426.9 467.4 500.9 529.7 555.1 578 598.7 617.7 635.4 651.8 667.2 681.8 695.5

334.8 363.4 407.7 442.3 470.9 495.6 517.4 536.9 554.7 571.1 586.2 600.4 613.6 626.1 638

1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006

1.009 1.012 1.019 1.026 1.032 1.038 1.043 1.049 1.053 1.058 1.062 1.066 1.07 1.074 1.077

0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179

0.7206 0.7237 0.7302 0.7367 0.7429 0.7488 0.7542 0.7593 0.7641 0.7686 0.7728 0.7767 0.7805 0.784 0.7874

1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402

1.4 1.399 1.395 1.392 1.389 1.386 1.383 1.381 1.379 1.377 1.375 1.373 1.371 1.369 1.368

temperature. In order to compare the standard air approach with real gas air, a case study is considered. A compression process is explored under both standard air and real air while exergy destructions are determined for both cases. Reference conditions for both cases are taken at 251C and 101.325 kPa and isentropic efficiencies of the compressors are assumed to be 85%. For the air standard model, the relationship between isentropic compression and actual compression can be given as:

 h2;i h1 ¼ cp T2;i T1 ¼ Zise cp ðT2 T1 Þ ¼ h2 h1 ð27Þ

here, subscript “2,i” indicates the exit temperature of the compressor at isentropic conditions and “2” refers to the actual exit temperature. Zise refers to the isentropic efficiency of the compressor. The temperature of the compressor outlet at isentropic conditions (T2,i) can be determined as:  k 1=k T2;i P2 ¼ ð28Þ T1 P1 The exergy destruction of the compressor can be determined from the air standard model as follows: wc ¼ h2 h1 ex d;c ¼ T0 ðs2

T0 ðs2 s1 Þ

s1 Þ þ exd;c

ð29Þ

The actual case exergy destruction of the compressor can be expressed as: exd:c ¼ ex1 þ wc

ex 2

here, ex1 and ex2 are the specific exergies of the inlet and streams, respectively, and they can be calculated as: ex 1 ¼ ðh1 ex 2 ¼ ðh2

h0 h0

T0 ðs1 T0 ðs2

ð30Þ

s0 ÞÞ s 0 ÞÞ

ð31Þ

Since no chemical reaction occurs in the compressor, only physical exergies are considered. The variation of the thermodynamic properties and the exergy destructions of the compressor for the air standard assumption and the real air are shown in Tables 3 and 4, respectively. As seen in Fig. 26 and Tables 3 and 4, the properties and specific exergy destructions of the compressor have similar quantities for both cases. As the compression ratio increases, the gap between the actual specific heat ratio and the specific heat ratio from the air standard model widens. The specific heat ratio of the working fluid at the compressor exit decreases from 1.4 at 1.5 compression ratio until it reaches a minimum value of 1.368 at 15 compression ratio. In addition, the temperatures of the stream at the exit of the compressor are affected significantly by the compression ratio for both standard air model and the actual air model. The exit temperatures of the compressor are identical at 1.5 compression ratio. However, as the compression ratio increases, the 12.31C temperature difference is noted between models. This situation can be explained by the change in the specific heat ratio. The second approach, which takes into account the variation of the thermodynamic properties of the air by temperature, can be modeled by using ideal gas assumptions and the data fromRthermodynamic properties tables. For instance, the air’s enthalpy at the T inlet of the compressor can be calculated as h1 h0 ¼ T 0I cp ðT Þ dT. Here h0 is the reference specific enthalpy at reference 0 0 temperature T and the function of temperature (h (T)). The specific heat values corresponding to the temperature can be taken from the thermodynamic tables.

238

Table 4

Gas Turbine Cycles

Variation of the thermodynamic properties and exergy destruction of the working fluid with compression ratio for actual process exd,c (kJ/kg)

T2 (K)

T2i (K)

cp (kJ/kg K)

cv (kJ/kg K)

k

1.5 2 3 4 5 6 7 8 9 10 11 12 13 14 15

5.814 9.535 14.38 17.64 20.14 22.19 23.94 25.48 26.86 28.11 29.27 30.34 31.34 32.28 33.17

341.2 375 427.5 468.6 502.9 532.6 559 582.8 604.5 624.6 643.3 660.8 677.3 692.9 707.8

334.8 363.4 408.1 443 472.2 497.5 519.9 540.1 558.6 575.6 591.5 606.4 620.4 633.7 646.3

1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005

0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176

1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4

35

1.405

30

1.4 1.395

25

1.39 20 1.385 15 1.38 10

Air standard (exergy destruction)

Specific heat ratio

Specific exergy destruction (kJ/kg)

r

1.375

Actual (ex destruction)

5

1.37

Air standard (k) Actual (k2)

0

1.365 1

3

5

7

9

11

13

15

Compression ratio Fig. 26 Variation of the specific heat ratio and exergy destruction of the working fluid with compression ratio for both cases.

Table 5

Variation of the thermodynamic properties and exergy destruction of the working fluid with compression ratio for air standard model

Zise

exd,c (kJ/kg)

T2 (K)

T2i (K)

cp (kJ/kg K)

cv (kJ/kg K)

k

0.75 0.7671 0.7843 0.8014 0.8186 0.8357 0.8529 0.87 0.8871 0.9043 0.9214 0.9386 0.9557 0.9729 0.99

49.44 45.48 41.66 37.96 34.38 30.92 27.56 24.31 21.15 18.09 15.12 12.23 9.43 6.703 4.052

668.1 659.9 651.9 644.4 637.1 630.2 623.5 617.1 610.9 605 599.3 593.8 588.5 583.4 578.4

575.6 575.6 575.6 575.6 575.6 575.6 575.6 575.6 575.6 575.6 575.6 575.6 575.6 575.6 575.6

1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005 1.005

0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176 0.7176

1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4 1.4

Gas Turbine Cycles

239

Table 6

Variation of the thermodynamic properties and exergy destruction of the working fluid with compression ratio for actual process

Zise

exd,c (kJ/kg)

T2 (K)

T2i (K)

cp (kJ/kg K)

cp (kJ/kg K)

cv,1 (kJ/kg K)

cv,2 (kJ/kg K)

k1

k2

0.75 0.7671 0.7843 0.8014 0.8186 0.8357 0.8529 0.87 0.8871 0.9043 0.9214 0.9386 0.9557 0.9729 0.99

45.07 41.29 37.64 34.1 30.67 27.35 24.13 21 17.97 15.02 12.16 9.373 6.665 4.03 1.465

658.8 651.1 643.6 636.5 629.6 623 616.7 610.6 604.7 599.1 593.7 588.4 583.4 578.5 573.7

571.1 571.1 571.1 571.1 571.1 571.1 571.1 571.1 571.1 571.1 571.1 571.1 571.1 571.1 571.1

1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006

1.067 1.066 1.064 1.062 1.061 1.059 1.058 1.056 1.055 1.054 1.053 1.052 1.05 1.049 1.048

0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179 0.7179

0.7783 0.7765 0.7747 0.773 0.7714 0.7698 0.7683 0.7669 0.7656 0.7643 0.763 0.7618 0.7607 0.7596 0.7585

1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402 1.402

1.371 1.372 1.373 1.374 1.375 1.376 1.377 1.377 1.378 1.379 1.38 1.38 1.381 1.382 1.382

1.405 1.4

50 Air standard (exergy destruction) Actual (ex destruction)

40

Air standard (k)

1.395 1.39

Actual (k2)

1.385

30

1.38

20

1.375 10

Specific heat ratio

Specific exergy destruction (kJ/kg)

60

1.37

0 70

75

80

85

90

95

1.365 100

Isentropic efficiency of the compressor (%) Fig. 27 Variation of the specific heat ratio and exergy destruction of the working fluid with isentropic efficiencies for both cases.

R T c ðT Þ The specific entropies can be derived similarly: s1 s0 ¼ T01 p T dT. Here s1 shows air’s specific entropy at the inlet of the compressor and s0 entropy refers to the reference specific entropy at reference temperature T0. Different from the reference specific enthalpy (h0), s0 is the function of the both temperature and pressure (s0 (T, P)). The change of the thermodynamic properties and the exergy destructions of the compressor for the air standard assumption and the real air are illustrated in Tables 5 and 6, respectively. As is presented in Fig. 27 and Tables 5 and 6, the properties and the specific exergy destructions of the compressor have similar quantities for both cases. As the compressor’s isentropic efficiency increases, the gap between the actual specific heat ratio and the specific heat ratio from the air standard model is narrowed. The specific heat ratio of the working fluid at the compressor exit increases from 1.371 at 75% isentropic efficiency until it reaches a maximum value of 1.382 at 99% isentropic efficiency. Moreover, the temperatures of the stream at the compressor exit are influenced considerably by the isentropic efficiency for both the standard air model and the actual air model. The difference between exit temperatures of the compressor is obtained as 9.31C at 75% isentropic efficiency, and it decreases to the 4.71C temperature when the isentropic efficiency reaches 99%. The following expressions can be driven for the state points at the inlet and the exit of the compressor under isentropic conditions (s1 ¼ s2) [1,13]:

s0 ðT1 Þ

R ln

  P1 ¼ s0 ðT2 Þ P0

R ln

  P2 P0

ð32Þ

240

Gas Turbine Cycles

By rearranging the equation above, the following result can be obtained:
0  s ðT2 Þ   R P2 P  ¼ R ln r;2 ¼ exp
0 s ðT1 Þ P1 Pr;1

ð33Þ

R

here Pr,1
and Pr,2 refer to reduced pressures for the state points. Reduced pressures are defined as a function of temperature, Pr ¼ exp s0 ðTRÞ . Another reason for the slight difference between the actual and standard air cycle results is that the air standard model is based on the assumption that the exhaust gases are only air. However, an actual complete combustion reaction of methane (CH4) with an excess air has other substances as products. ðO2 þ 3:76 N2 Þ2l þ CH4 -CO2 þ H2 O þ ð2l AFR . l ¼ AFR stoich

2ÞO2 þ 7:52 l ðN2 Þ

ð34Þ

where l shows air–fuel equivalence ratio, An exergy analysis of the expansion of a real combustion gas is shown in the following example: Suppose an open Brayton cycle combusts methane as the main fuel. AF of the cycle (AFR) is 85.8 and complete combustion occurs into the combustion chamber. After the combustion, 51.44 kg exhaust gas is formed as products. High temperature and pressure gas enters the turbine and expands with an isentropic efficiency Zise ¼ 0.85. Calculate the exergy destruction during the expansion. Stoichiometric AF of the methane is 17.16, which makes the excess air ratio of the reaction l ¼ 5. Complete combustion of the fuel can be illustrated as Reference state conditions

10 O2 þ 37:6 N2 þ CH4 -CO2 þ H2 O þ 8 O2 þ 37:6 N2

ð35Þ

T0 ¼ 298:15 K; P0 ¼ 101:325 kPa

Properties of the exhaust gas are based on the properties of the combustion products and their mole fractions. P P f0 ¼ h0 T0 s0 ; s0 ¼ yi s0i ðT0 Þ; h0 ¼ yi h0i ðT0 Þ P ex ch ¼ yi exch;i ðT0 Þ

ð36Þ

where i represents combustion products (O 2, H2O, CO2, N2), y represents the mole fractions. The volume of the exhaust gas can be determined by using the ideal gas assumption P0 V0 ¼ n R T0

Molar mass of the exhaust product

M¼

X

yi Mi

At reference conditions F0 ¼ 3961 kJ; V0 ¼ 44:15 m3 ; Ex ch ¼ 3:596 MJ State 1 (gas turbine inlet) The temperature of the gas combustion products is assumed as 1273K. One should note that the exhaust temperature must always be lower than the adiabatic flame temperature. The adiabatic flame temperature of the methane, in this case, is TAFT ¼1544K, which satisfies the T1oTAFT relation   n X

P1 n yi ∏ yi yi s0i ðT1 Þ R ln P1 ¼ r P0 ; P1 v1 ¼ RT1 ; s1 ¼ P0 i ¼ 1 i¼1 h1 ¼

X

yi h0i ðT1 Þ

ex 1 ¼ ex ch þ h1

T0 s1

ð37Þ

f0

The quantity of the exergy in exhaust gases is influenced by the chemical exergy of the exhaust components. Therefore, the chemical exergies are also taken into account. Nevertheless, because of the inexistence of the flammable substances in the exhaust gas, the chemical exergy has relatively fewer values. The specific entropy of the gas mixture can be determined as follows 3 2 n    n  yi    P yi n   n  X Pi yi P P n yi 6 P i¼1 y7 yi ln ¼ ln ∏ ¼ ln4 ∏ yi ∏ yi i 5 ¼ ln P0 P0 P0 i ¼ 1 i ¼ 1 P0 i¼1 i¼1 sðT; P Þ ¼

n X

i¼1

yi s0i ðT Þ

R ln



P n yi ∏ y P0 i ¼ 1 i



ð38Þ

Gas Turbine Cycles

241

here Pi refers to the partial pressure of the substance in the exhaust gas and P shows the total pressure of the exhaust gas. As the mixture is assumed as an ideal gas, the molar fraction of each component equals to their pressure fraction as well (Pi ¼yi P). H1 ¼ 24:3 MJ;

V1 ¼ 12:6 m3

S1 ¼ 401:KkJ=K

P1 ¼ 1520 kPa

Ex 1 ¼ 47:7 MJ

State 2s (gas turbine exit at isentropic conditions) s2s ¼ s1 ; P2s ¼ P0 ; P1 exp

h2s ¼

X

yi h0i ðT1 Þ;

P2s v2s ¼ RT2s

 X  yi s0i ðT1 Þ yi s0i ðT2s Þ ¼ P2s exp R R

X

ex 2s ¼ ex ch þ h2s

T0 s2s

f0

ð39Þ

here, the relationship between pressure at the inlet and the exit of the gas turbine (P1 and P2) is defined as the functions of Pn the y s0 i¼ 1 i i . In order to calculate the temperature after isentropic expansion (T2s), reduced pressure of the reduced pressure Pr ¼ exp R reduced pressure and the specific enthalpy of the state 2s should be obtained. Iterative methods can be used to calculate: H2s ¼

13:5 MJ;

V2s ¼ 95:5 m3

S2s ¼ 401 kJ=K

P2s ¼ 101:325 kPa

Ex2s ¼ 10:0 MJ

State 2 (exit of the gas turbine at actual conditions) Pressure of the gas turbine is identical for the actual and isentropic conditions P2 ¼ P2s ¼ P0 ¼ 101:325 kPa The enthalpy of the actual exit conditions can be determined by the definition of the isentropic efficiency. P 0 ðH1 H2 Þ ¼ Zise;t ðH1 H2s Þ h2 ¼ yi hi ðT2 Þ; Zise;t ¼ 0:85 H2 ¼ 7:8 MJ Similar to the isentropic condition, the entropy and volume can be obtained as   n X

P2 n yi s2 ¼ yi s0i ðT2 Þ R ln P2 v2 ¼ RT2 ∏ yi ; P0 i ¼ 1 i¼1 ex 2 ¼ ex ch þ h2 H2 ¼ 7:8 MJ; P2 ¼ 101:325 kPa

T0 s2

ð40Þ

f0 V2 ¼ 110:1 m3

S2 ¼ 410 kJ=K Ex2 ¼ 13:2 MJ

The exergy destruction during the expansion of the combustion can be calculated as Exd ¼ Ex in Wt -Ex in ¼ Ex1 Ex 2 ; Exin ¼ 34:5 Mj; Wt ¼ 32:1 M;

Wt ¼ H1 H2 Exd ¼ 2:4 MJ

Exergy efficiency of the gas turbine c¼

Wt ¼ 0:93-93% Ex in

The exergy destruction fraction of the gas turbine EFD ¼

Exd ¼1 Ex in

c¼1

0:93 ¼ 0:07-7%

In this section, the second law analysis for a gas turbine and a compressor is conducted by three different models. The first model assumes the specific heat of the air is constant, and properties of the combustion products are identical with the properties of air (air standard model). The second model takes into account the variation of the specific heat with the temperature but still assumes the exhaust gases as air. Lastly, the third model includes the combustion reaction and change of the thermodynamic properties with the temperature. Air standard model can be useful if the cycle has lower pressure ratios with higher isentropic efficiencies. Otherwise, it supplies slightly incorrect results. It can be considered as a tool for rough estimations. The second model provides the correct results if the cycle is a closed cycle, which utilizes a single stream throughout the cycle. It can still reflect the correct results if the excess air of the combustion chamber is too large lc1. With the increasing excess air ratio, the exhaust gases’ properties approach the properties of air. The third model should be used for the more detailed solutions especially for the cases where the chemical exergies have a significant role in the system.

Gas Turbine Cycles

242

3

T Combustors

Compressor

Turbine p=

c

4 Air

Product

in

gases out

2 5 1

p=

c

a a

1

2

Diffuser

3

Gas generator

4

5 Nozzle

(A)

S

(B)

Fig. 28 Turbojet engine diagram and ideal T-s diagram. (A) Turbojet engine; (B) T-s diagram. Adapted from Moran MJ, Shapiro HN, Boettner DD, Bailey MB. Fundamentals of engineering thermodynamics. John Wiley & Sons; 2010.

Air intake

Combustion chamber

Exhaust nozzle

Propeller

Gear box Compressor

Turbine

Fig. 29 Schematic diagram of a turboprop engine. Modified from Boldmethod. Available from: http://www.boldmethod.com/; 2017 [accessed 05.02.17].

4.7.7

Jet Propulsion Cycles

As a result of the development of the aircraft industry, researchers are facing different challenges, and new technological improvements are required for various fields. One of the most significant features is aircraft performance, which has a vital role in the financial prosperities of air transport companies, aircraft engineers, and manufacturers. To meet the demands of the aircraft industry, gas turbine engines were developed to be used as a part of the propulsion system in airplanes. Due to their light structure and high power capacity, they are commonly used in this field. A typical jet engine utilizes an open Brayton cycle. However, they differ from the conventional gas turbine cycle by their exit turbine pressures. Flue gases leave the gas turbine at a considerably higher pressure than the atmospheric pressure and expand through the nozzle at a high speed and then discharged. The large velocity changes create the propulsive force. Moreover, airplane gas turbines run on larger pressure ratios, which usually lie between 10 and 25 [13]. A typical schematic of a turbojet engine is illustrated in Fig. 28. First, the pressure of the incoming air increases slightly due to the velocity drop (in ideal conditions pressure increase in deceleration occurs isentropically through this component) at the diffuser, then it moves into the compressor and pressurizes. The acceleration of the masses in the reverse direction generates the propulsion for the airplane. The momentum generally can be created in three ways, by:

• • •

Turboprop engine: accelerates the large mass of air considerably (see Fig. 29). Turbojet engine: accelerates the relatively lower amount of air significantly. Propeller-driven engine: slightly accelerates the large amount of air.

After the compression processes, high-pressure air is directed to the combustion chamber where the fuel is mixed with the working fluid and combustion befalls at constant pressure. Afterward, high-temperature and high-pressure exhaust gases partly expand in the gas turbine. Nevertheless, this expansion does not extend to the atmospheric conditions. It only expands to meet the power required by the compressor and auxiliary equipment. Eventually, hot and pressurized flue gases accelerate and reach even

Gas Turbine Cycles

243

higher velocities in the nozzle and discharge to the atmosphere. However, unlike the ideal processes, actual jet engine cycles have irreversibilities due to dissipation losses at the expansion and compression processes at the turbine and compressor, acceleration at the nozzle. Therefore, the real propulsion is less than the ideal thrust. The inlet and exit velocities of the air are different, so it causes a momentum change hence the propulsion is developed and can be obtained by Newton’s second law. Since there is no pressure difference between the inlet and outlet of the turbojet engine (they are both at atmospheric pressure) the net thrust generated by the cycle can be obtained as [13]: _ V Þex F ¼ ðm

_ V Þin ¼ m _ ðVex ðm

Vin Þ

ð41Þ

here Vex represents the flue gases’ velocity at the exit, and Vin indicates the inlet velocity of the air. Both velocities are relative velocities to the airplane. Therefore, for an airplane that is cruising in constant air, Vin shows the airplane velocity. In fact, because of the fuel addition at the combustion chamber, the gases’ mass flow rates are different at the engine’s exit and inlet. Nevertheless, as the AF used in the cycle is quite higher, the effect of the mass fuel is negligible. Therefore, mass flow rates are assumed constant in Eq. (41) as the inlet mass of the air. An airplane traveling at constant speed faces drag force due to pressure differences through the plane and friction forces as a result of shear stresses. For such cases, the drag force is equal to the thrust, which makes the net _ p , can be obtained as follows [13]: force acting on the airplane equal to zero. The propulsive power developed by the thrust, W _ P ¼ FVaircraft ¼ m _ ðVex W

Vin ÞVac

ð42Þ

here F and Vac represent the propulsive force and aircraft velocity, respectively. Since the primary object of the turbojet engine is not producing work as in the other GTCs, a new definition for the efficiency is required. In this regard, the general definition of the efficiency can be used. The propulsive ratio is the ratio between desired outputs of the system to required inputs. Here propulsive power is the coveted output for an airplane, and the released heat rate of the combustion process is the necessary input for the output. Therefore, the propulsive efficiency can be calculated as [13]: Zp ¼

_p W Propulsive power ¼ _ Energy input rate Qin

ð43Þ

Propulsive efficiency is an indicator of the effect of the released heat at the combustion chamber on the propulsive energy. The rest of the released heat from the combustion chamber demonstrates as the kinetic energy of the exhaust gases and as an enthalpy growth of the air leaving the engine [13].

4.7.7.1

Exergy Analysis of a Turbojet

In order to properly conduct the evaluation of an aircraft, an engineer should understand aircraft performance issues. Furthermore, to build solid approaches for industrial applications, a decent knowledge of performance characteristics and design limitations of an airplane is needed to be defined. Exergy analysis is a useful tool that can help provide information for performance enhancement of the aircraft. The exergy analysis of a jet engine cycle differs from the conventional ground basis systems in two significant ways [16]:

• •

The aircraft engine is usually based on the open Brayton cycle, in which thrust is developed by the ejection of the high temperature and pressure exhaust gas at higher velocities. Hence, large exergy losses occur in the exhaust [17]. These losses are special for jet engine cycles and not caused by irreversibilities of the components. The operating environment varies continually, which is quite different from the ground-based systems as their environment relatively remains constant during the operation.

The exergy analysis of a turbojet engine, at a constant operating condition and during the flight, is explained in this section. The theoretical approach closely follows the analysis of Clarke and Horlock [18] and Dincer and Rosen [16]. For a material flow, the specific exergy function ζ can be shown as follows: ζ ¼ h0

0

ð44Þ

T1 s

here T1 , s and h represent the free steam temperature, specific entropy, and stagnation enthalpy, respectively. For the component “I” of a flow, the specific exergy ex can be defined as the difference between the specific exergy function at the initial state and the specific exergy function at reference state: ex i ¼ ζin

ð46Þ

ζ1

Under ideal gas assumption ex i ¼ cpi ðTi

T1 Þ

   Ti T1 cpi ln T1

  Ti c2 Ri ln þ i 2 T1

ð46Þ

where ci, Ri, cpi, pi, and Ti represents the temperature, pressure, specific heat at constant pressure, gas constant, and absolute velocity to a fixed reference environment of constituent i respectively.

244

Gas Turbine Cycles

Duct fan

Combustion chamber

Bypass air

Exhaust nozzle

Air intake

Compressor

Turbine

Fig. 30 Schematic diagram of a turbofan engine. Adapted from Boldmethod. Available from: http://www.boldmethod.com/; 2017 [accessed 05.02.17].

T1 and P1 denote the free stream temperature and pressure, which are also the same as the reference temperature and pressure. Hence, an exergy balance for a general control volume in motion can be defined as follows: X Ti T1 X X _ i exi g _ i ex i g m m qi ð47Þ  Ps þ PT þ þ ExBE : Ti i i i outgoing

incoming

where PT and Ps indicate the thrust power and shaft power, respectively. Similar to the energy efficiency, exergy efficiency of the aircraft can be described as the ratio between desired output to required input: c¼

PT _ incoming Ex

ð48Þ

where the incoming exergy is the summation of the exergies of air and fuel _ incoming ¼ m _ fuel exfuel þ m _ air ex air Ex

4.7.7.2

ð49Þ

Modifications to Turbojet Engines

Aircraft technology has been evolving through the ages. These improvements are defined by Cengel and Boles [13], and this section follows the steps explained by them. First generation airplanes were run on propellers, which are powered by engines which are similar to the car engines. The first important step for the aircraft industry was taken by the introduction of the turbojet engine in 1952. Both systems have advantages and disadvantages. Therefore, to promote the strengths of the propeller-driven and jet propulsion systems, various approaches have been tested. Designing a model that consists of the useful sides of the both engines has been the primary object. In this regard, the propjet engine and the turbofan (see Fig. 30) engine were developed. The fanjet (or turbofan) is the most commonly used thrust system of the aircraft industry, and consists of a large fan run on the turbine, which forces a large mass of air through the duct surrounding the engine [13]. The exhaust of the fan leaves the duct at a relatively greater velocity, which results in a rise in the momentum and improves the propulsive force of the engine substantially. Consider the same power applied to two systems. The first system is a large volume of slower-moving air, and the second one is a small volume of fast-moving air. The first one develops more thrust than the second one, and it is the principle that turbofan engine was based on. The first commercial turbofan engine was tested successfully in 1955 [13]. The main difference between a turbofan and a turbojet is the cowling of turbofan, which covers the entire fan. In a turbojet, all the propulsion is developed by the ejection of the high-speed exhaust gas; on the other hand, in a turbofan, the exhaust gas is mixed with the slower air, which results in a significant decrease of the noise. By the introduction of new burner technologies, the temperature of the combustion products has increased up to 15001C, which has led to substantial efficiency growths [13]. Also, increase in bypass ratio (BRP), which is the ratio of the mass flow rate of the bypass stream to the mass flow rate entering the core, resulted in an increase in the propulsion. Therefore, removing the cowl from the fan increases BRP and thrust. For instance, a typical turbofan has a bypass ratio by 5–6 where propjets have higher bypass ratios by 100 [13]. Another common adjustment for military airplanes is the installment of an afterburner unit between the nozzle and the turbine. An afterburner is simply a reheating unit in which additional fuel is injected into the exhaust stream of the turbine and combusted (since it is a product of lean-combustion exhaust gas containing O2). Thus, the nozzle inlet gas has a higher temperature than the case without the afterburner. As a result, a higher nozzle exit velocity is achieved, resulting in increased propulsion [13,14]. A ramjet engine is a well-shaped duct without turbine or compressor. A ramjet is occasionally used for high-speed thrusts of aircrafts and even for missiles. The ram effect of the incoming high-speed air leads the pressure rise in the engine. Hence, a ramjet engine requires being brought to an adequately high velocity by an external source prior. The ramjet has the best performance in

Gas Turbine Cycles

245

Fuel intake Combustion chamber Compressor 3

2

Gas turbine 1

4

Generator

Steam generator 5 6

7

Steam turbine

Generator 8

Condenser

Pump

9

Fig. 31 A simple layout of the air-steam combined cycle power plant.

an aircraft, which flies over Mach 2–3. In a ramjet, the air is decelerated and the speed reduced to about Mach 0.2, after then fuel is added to the air and combusted at low velocity. Afterwards, the exhaust gases are directed to the nozzle, which results in an acceleration. A ramjet with a speed of air over supersonic level is called scramjet engine [13]. Lastly, a rocket is a device where pressurized combustion products are expanded in a nozzle and in which combustion occurs by the reaction of a solid or liquid fuel with an oxidizer in the combustion chamber. The exhaust gases exit the rocket at higher velocities, which causes the thrust for propulsion [13].

4.7.8

Combined Cycle Power Plants

Exhaust temperatures of the gas turbines are considerably higher than the atmosphere temperature as is explained in the examples of the previous sections. The discharging processes occur at atmospheric pressure for the open Brayton cycle. Therefore, it is impossible to expand the working fluid below the atmospheric conditions. The high-temperature and low-pressure stream still has great potential, and it can be used by the integration of the WHR units to the Brayton cycle. The high energy can be recovered if the expelled gases can be utilized in a Rankine cycle [1]. CCPPs are smart power production options because of their higher thermal efficiencies than the simple steam or GTCs. In these structures, the rejected low-pressure hot flue gases from the gas turbine are sent to a heat recovery steam generation (HRSG) unit where additional fuel is supplied. There is enough O2 in the stream to sustain a burning process, which leads to temperature growth. Moreover, hotter gases are utilized with the arrangement of HEXs to produce steam for a Rankine cycle. The key task in modeling a combined cycle is the proper usage of the gas turbine waste heat in the steam cycle in order to attain the best turbine production. Combined cycles have greater thermal efficiency besides their higher power outputs compared to the gas and steam turbine cycles. Higher performances of CCPPs compared to Brayton and Rankine cycles have made them quite attractive for power generation. Based on these advantages and less specific emissions, CCPPs have widely been used all around the world. The main design parameters of these plants are pressure ratio (r), compressor isentropic efficiency, gas turbine isentropic efficiency, steam turbine isentropic efficiency, and gas turbine inlet temperature [1]. A simple diagram of an air-steam combined cycle power plant is illustrated in Fig. 31. In this diagram, a basic configuration of Brayton cycle is considered. The expelled hot gases have a relatively large flow rate and are directed into the HEX, which works as a steam boiler for the Rankine cycle, and helps to recover the part of the waste heat. The only heat addition to the system occurs at the combustion chamber at a constant pressure, and heat rejections take place after the steam generator for the Brayton cycle part and condenser at the Rankine cycle part. In this section, various types of combined cycles are explained and future directions are discussed. Dr. Andreas Poullikkas’s review article “An overview of current and future sustainable gas turbine technologies” [4] is followed closely in the next sections.

246

Gas Turbine Cycles

Steam turbine

Generator

Seperator Heater

Condenser HEX – steam generator

Fuel intake

Combustion chamber

Pump

Absorber

Compressor

Gas turbine

Generator

Fig. 32 The Brayton–Kalina cycle.

4.7.8.1

The Brayton–Kalina Cycle

The Kalina cycle is an innovative bottoming cycle developed by Dr. Alexander Kalina, which uses a zeotropic mixture of two fluids with different boiling points as the working fluid (ammonia and water). Its features are such that temperature of the cycle tracks the temperature of the turbine exit in the waste heat boiler. Nevertheless, at the condensing processes, the thermodynamic gain of the relatively small boiler temperature difference compared to a steam cycle would be gone due to the cooling temperature of condenser cooling medium. In order to solve these problems, the Kalina cycle was developed [4]. A basic Kalina cycle consists of a WHR vapor generator (in this case it is the HEX where the exhaust gas of the gas turbine is directed to supply heat to the bottoming Kalina cycle), a turbine works with steam-ammonia, and the distillation condensation system, as it is presented in Fig. 32. In the distillation condensation subsystem, first the flow incoming from the turbine is cooled by the heater (recuperator), and then the stream is mixed with a lean solution of NH3 to increase the condensation temperature of the working fluid. Finally, the basic solution is condensed in the absorber. The condensed solution is brought to the heater under pressure. A portion of the stream is directed to dilute the ammonia-rich stream coming from the separator. The primary flow passes the recuperator, then it is flashed in the separator. The vapor is mixed with the basic solution. The vapor is condensed, then pressurized by the pump before it flows to the vapor generator [19]. Ten to thirty percent more energy can be generated by the Kalina cycle when it is compared to a Rankine cycle [19]. The main reason for this situation is the exhaust pressure of the Kalina cycle, which is above the atmospheric conditions. The starting time of a Kalina cycle is much less because sustaining of vacuumed medium is not a necessity for the condenser during the operation of the cycle. The working fluid mixture can easily be varied to acquire the best performance with respect to changes in load or ambient conditions. Moreover, those systems are favorable due to their compact configuration. The projected area of a Kalina plant is about 40% smaller than a Rankine plant design when they are compared [4]. In order to commercialize the Brayton–Kalina cycle plants, a special global licensing contract has been signed with an agreement between Exergy Inc., who owned the rights to the technology, and General Electric in 1993. Afterward, GE designed a Brayton–Kalina cycle by 260 MW capacity, which was planned to be in service by 1998, but the project was later suspended [4,20].

4.7.8.2

The Brayton–Brayton Cycle

Two Brayton cycles can be combined by an air–gas HEX as shown in Fig. 33. The exhaust stream of the main cycle enters a HEX where it supplies the required heat by the working fluid of the second Brayton cycle. The working fluid is expanded in the secondary turbine to produce extra energy. When this scheme is compared with the traditional combined cycle, this layout does not need a large mass of steam equipment such as a condenser, boiler, and steam turbine. Moreover, it does not require a water treating component and makes possible the unmanned processes [4]. The current studies have presented the viability of this structure [19]. It is possible to increase the overall efficiency by 10% and the power capacity of the system 18–30% depending on the design parameters such as the number of intercoolers. For instance, the Allison 571K topping gas turbine can be considered as a good example for this layout [21]. By the addition of the air bottoming cycle, which consists of two intercoolers, power production of the cycle is increased from 5.9 to 7.5 MWe.

Gas Turbine Cycles

Fuel intake

247

Combustion chamber

Compressor

Gas turbine

Generator

Discharge

Compressor

Gas turbine

Generator

Fig. 33 The Brayton–Brayton cycle.

Moreover, the overall thermal efficiency is inclined from 33.9 to 43.2%. Similar outcomes were found with the General Electric’s LM2500 topping turbine [22]. The system can also be used for cogeneration. Instead of wasting the exhaust stream, which is discharged by the cycle at 473–523K, it can be utilized for the cycle essentials that need to be heated until reaching the desired temperatures [4].

4.7.8.3

The Brayton–Diesel Cycle

The performance of a Diesel cycle can be enhanced by preheating the working fluid sufficiently. Therefore, the exhaust stream of the gas turbine can be utilized in a HEX, which can preheat the incoming air of the Diesel engine. A simple layout of the system is presented in Fig. 34. An alternative application can be the direct usage of the exhaust stream of the gas turbine in the Diesel engine. As the gas turbine consists of an oxygen-rich mixture, there would be still sufficient amount of O2 (14–16%) in the combustion products, which can also be burned in the Diesel engine [19].

4.7.8.4

The Brayton–Stirling Cycle

In a combination of a gas turbine and a Stirling engine, the heat source of the Stirling engine can be located either in the combustion chamber of the turbine or next to the expander in the exhaust stream, as it is illustrated in Fig. 35. The configuration is set by the ideal operation conditions of the cycle. Another restriction for the cycle’s configuration is the type of the materials used in the head of the Stirling heater. Up to 9 MWe can be recovered by the utilization of the waste heat of a Rolls–Royce RB211 gas turbine of 27.5 MWe in the bottoming Stirling cycle [21]. The thermal efficiency of a similar power plant can have an efficiency by 47.7%. Similar to the Brayton–Brayton cycle, this integration makes available the small scale and easy to operate WHR application [4].

4.7.8.5

The Brayton-Fuel Cell Cycle

A fuel cell system can obtain higher efficiency (60%) and can be run at greater pressures and temperatures. This situation allows integrating a gas turbine to the fuel cell system, therefore the performance of the system is improved [23]. The diagram of the integrated system is illustrated in Fig. 36. The utilization of the fuel cells that are combined with the combustion chambers makes it possible for efficiency to reach almost 70% [24]. The Brayton-fuel cell cycle is appealing because it has the greatest efficiency compared with all the advanced cycles. Thus, It can be considered as one of the most promising choices for future power plants [4,25].

4.7.8.6

The Chemical Recuperation Cycle

The chemical recuperated gas turbine cycle converts CH4, H2O, and occasionally CO2 into an H2 and CO2 fuel mixture that can be burned in the combustion chamber by using a reforming process (see Fig. 37). This process increases heating values of the fuel as

248

Gas Turbine Cycles

Fuel Intake

Combustion chamber

Compressor

Gas turbine

Generator

Discharge

Diesel engine

Generator Fig. 34 The Brayton–Diesel cycle.

Fuel intake

Combustion chamber

Compressor

Gas turbine

Stirling engine

Generator

Regenerator

Generator Fig. 35 The Brayton–Stirling cycle.

the endothermic reaction absorbs the heat at a temperature that is below the burning temperature. Recuperation method, which utilizes both thermal and chemical reactions, has greater WHR rate than the conventional recuperation system. Furthermore, the H2-rich fuel is more flammable than methane and allows ignition at a lower flame temperature that theoretically decreases NOX formation. Conversely, the gas turbine flue gas temperature is not sufficient for a complete reforming reaction. At 823K, only 20% of the entire fuel can be reformed. To raise the temperature, some additional combustion processes can be used [4].

Gas Turbine Cycles

249

Water intake

Anode

C a Cathode

Combustion chamber Compressor

Gas turbine

Generator

Gas turbine

Generator

Fig. 36 The Brayton-fuel cell cycle.

Compressor

Motor

Fuel

Reformer

Compressor

Combustion chamber

Fig. 37 The chemical recuperation cycle with flue gas recycling.

4.7.8.7

Integrated Gasification Combined Cycle

The emissions of fossil fuels have led to environmental concerns to the point where coal usage for power production has been threatening. Parallel to the installation of flue gas scrubbers in conventional coal-fired power plants, development of the integrated gasification combined cycle is proceeding on groundbreaking power plant models, which are not only more suitable from an

250

Gas Turbine Cycles

Particulate Sulfur Mercury removal

Gas cleaning Compressor

Combustion chamber Gas turbine

Generator

Discharge

Steam generator (HEX)

Coal

Steam turbine

Generator

O2

Air separation unit

Compressor

Air intake

Ash

Pump Depleted O2 air Condenser Feed water Fig. 38 Integrated gasification combined cycle.

environmental aspect but also feature higher efficiencies. In order to achieve the combined cycle, the coal needs to be converted to a gaseous fuel via a gasifier firstly [26]. The simple schematic of an integrated gasification combined cycle is illustrated in Fig. 38. In the gasifier, gasification is attained as the result of the controlled combustion of coal/biomass with oxygen and steam. Then syngas (synthesis gas) and solid waste is produced. Oxygen is supplied to the gasifier by the companion air separation unit. The syngas produced by the gasifier is essentially composed of CO and H2. After the syngas is formed, it is cleaned from pollutants and later on combusted in the combustion chamber and drives the gas turbine. In integrated gasification combined cycles, pollutants such as sulfur and mercury are removed before the combustion process, unlike the conventional coal power cycles. Even though integrated gasification combined cycles emit less SO2, NO, Hg, and particulate emissions (sulfur dioxide, nitric oxide, mercury, and particulate emissions) than similar conventional coal plants, abundant solid waste still needs to be managed carefully [14]. Likewise, to the Brayton–Rankine combined cycle, the exhaust heat of the cycle supplies the heat for the steam generator of the bottoming cycle consequently the steam turbine runs and drives the generator. The amount of H2O going to the gasifier from the Rankine cycle is compensated by the feed water intake, which takes place after the condenser.

4.7.9

Case Studies

In this section, to illustrate the industrial combined power plant and integrated gas turbine systems, two case studies are presented.

4.7.9.1

Case Study 1

In this case study, a combined cycle power plant is proposed. A Rankine cycle is used to generate electricity from waste heat and enhance overall system efficiency. The system is designed as the schematic presented in Fig. 31. In the suggested system, the only output from the system is electricity, while the main inputs are air and NG. The system consists of an open Brayton cycle driven by NG and a Rankine cycle utilizes the waste heat of the Brayton cycle for the steam generation. In the proposed system, the compressor compresses air from 100 to 1410 kPa. The compressed air flows into the combustion chamber; the NG supplies the heat by burning in the combustion chamber. The combustion chamber is the place where the only chemical reaction occurs in the system. Hence, chemical exergies are taken into account only in the combustion chamber. The exhaust gas of the Brayton cycle flows to the HEX to generate steam for the Rankine cycle. After the flue gas passes through the HEX, the exhaust is discharged to the ambient.

Gas Turbine Cycles

Table 7

251

Design parameters of the system in the case study 1

Parameters

Value

Pressure ratio of compressor Compressor isentropic efficiency Atmospheric conditions Turbine isentropic efficiency Chamber combustion efficiency Generator efficiencies Air mass flow used Air–fuel ratio of the combustion (AF) Pressure of exhaust fumes

14.1 80% 100 kPa, 251C 77% 95% 100% 41.99 kg/s 76.75 100 kPa

In parallel, the Rankine cycle also generates electricity. Steam is generated by the exhaust flue of the gas turbine and enters the steam turbine to generate useful work. After the steam exits from the steam turbine as wet steam, it passes in the condenser and rejects its heat to the environment.

4.7.9.1.1

Thermodynamic analysis

The energy and exergy efficiencies, exergy destructions, temperature, pressure, specific enthalpy and entropy are calculated for all system components. The combustion reaction is modeled on ASPEN Plus software package and the rest of the elements of the system are constructed in the Engineering Equation Solver (EES). The general design parameters and assumptions made for the analysis are listed in Table 7. Besides the parameters in Table 7, the following assumptions are made for the study:

• • • • • •

The changes in potential and kinetic energies are negligible. The only chemical reaction occurs in the combustion chamber. The compression and expansion processes are adiabatic (externally reversible, internally irreversible). No heat losses through the HEX. Ideal gas principles apply for the gases. Steady state, steady flow conditions are applied.

The energy balance, according to the first law of thermodynamics, is applied to each component of the system. Energy balance equation for any component can be shown as follows: X X _ W _ þ _ in hin _ out hout ¼ 0 EnBE : Q m m ð50Þ

_ and W _ represent the heat transfer and work rate flows through the system boundaries, h represents the specific enthalpy of here Q the working fluid, and m_ represents the mass flow rate of the working fluid. Also, exergy balance is applied to each element of the system. The exergy balance equation for any component can be written as follows [27]: X X _ Qi Ex _ Wi þ _ di ¼ 0 _ in ex in _ out ex out Ex m m ExBE : Ex ð51Þ

_ Qi is the exergy transfer rate by heat, Ex _ Wi shows the exergy rate by work, and Ex _ di is the exergy destruction rate by the here Ex component. Specific exergies for each component are also determined for each element as follows: ex i ¼ hi

h0

T0 ðsi

s0 Þ þ exch

ð52Þ

here exi is the specific exergy for the ith state point. h0, T0, and s0 show the specific enthalpy, temperature, and specific entropy, respectively, for the reference point, which is the atmospheric conditions T¼251C and P ¼100 kPa. exch represents the specific chemical exergy. The specific chemical exergy is determined only for the state points where the chemical reaction occurs [28]. In this study, the only chemical reaction occurs in the combustion chamber. The calculation of specific chemical exergy can be shown as follows [27]. X X ð53Þ yi ex ch;i þ R T0 exch ¼ yi ln yi here, yi is the molar fraction of the gas species “i” in the gas mixture. For instance, NG used in this study consists of 15% ethane and 85% methane, which makes yi 0.15 and 0.85 for ethane and methane, respectively. exch,i represents standard chemical exergy of the gas species. Detailed exergy efficiency and exergy destruction rate definitions for each element are given in Table 8. Considering the main inputs and outputs of the system, the overall system energy and exergy efficiencies can be obtained as follows: P _ Q Zen;ov;sys ¼ 1 P out ð54Þ _ in Q

252

Gas Turbine Cycles

Table 8

Table 9

Exergy efficiency and exergy destruction rate definitions for the system components

Component

Exergy destruction rate definition

Steam turbine

_ d;ST ¼ m_ 7 ex7 Ex

Pump

_ d;PI ¼ m_ 9 ex9 þ W_ in;P Ex

Heat exchanger (HEX) 1

_ d;HXI ¼ m_ 4 ex4 þ m_ 6 ex6 Ex

Gas turbine

_ _ 10 ex10 Ex d;GT ¼ m

Compressor

_ _ 1 ex1 þ W_ in;C Ex d;C ¼ m

m_ 8 ex8

Exergy efficiency definition

W_ out;ST

c;ST ¼

m_ 6 ex6 m_ 7 ex7

m_ 11 ex11

W_ out;ST _ 7 ex7 m_ 8 ex8 Þ ðm _ Ex d;P _ Ex in;P

c;P ¼ 1 _ 5 ex5 m

W_ out;GT

m_ 2 ex2

c;HX1 ¼

_ 7 ex7 m _ 6 ex6 Þ ðm _ 4 ex4 m _ 5 ex5 Þ ðm

c;GT ¼

W_ out;GT _ 10 ex10 m _ 11 ex11 Þ ðm

c;C4 ¼

_ 2 ex2 m_ 1 ex1 Þ ðm W_ in;C

Thermodynamic data for all state points

State No.

Stream

T (K)

P (kPa)

m (kg/s)

h (kJ/kg)

s (kJ/kg
K)

Ex (kW)

0 0` 1 2 3 4 5 6 7 8 9

Air Water Air Air Combustion gases Combustion gases Combustion gases Water Water Water Water

298.2 298.2 298.2 706.3 1207 753.6 400 334.2 630 333.2 333.2

100 100 100 1410 1410 100 100 800 800 19.93 19.93

– – 41.99 41.99 42.54 42.54 42.54 5.397 5.397 5.397 5.397

104.8 298.4 298.4 720.6 1287 771.6 401.2 256 3176 2373 251.2

0.3669 6.864 6.862 6.989 7.593 7.82 7.159 0.8433 7.431 7.431 0.8312

– – 0 16,141 33,394 7986 611.9 49.2 5206 873.8 42.6

cov;sys ¼ 1

P Ex P d Ex in

ð55Þ

In the combustion chamber, the NG is burned. The heat input by combustion can be determined as follows _c ¼m _ fuel  LHV NG  Zc Q

ð56Þ

_ fuel is the mass flow rate of natural gas. In this study, LHVNG is the lower heating value of NG and Zc is combustion efficiency, m LHV is taken as 47,141 MJ for the NG [29].

4.7.9.1.2

Results and discussion

Thermodynamic calculations of each system component are performed by using EES software. The values of temperature (K), pressure (kPa), mass flow rate (kg/s), specific enthalpy (kJ/kg), specific entropy (kJ/kg K), and exergy (kJ/kg) are determined for each state point of the system as listed in Table 9. As seen in Table 9 and Fig. 39, the highest exergy destruction rates are determined in the combustion chamber and gas turbine. The summation of exergy destruction rates of two components has more than 50% of the total destruction rates in the system. It is because of the irreversibilities of these elements. As presented in Table 9 and Fig. 39, the exergy destruction rate of the pump has the lowest share in the overall system (less than 1%). The main reason for this situation is that water enters the pump at low temperature and its temperature slightly increases inside the pump. Thus, even the pump has a poor exergy efficiency (see Fig. 40), along with the lowest exergy destruction rate of the cycle. The compressor has the highest exergy efficiency by 91%. It has the highest efficiency as the heat losses through the surface of the compressor are neglected. The produced powers by the system elements are comparatively illustrated in Fig. 41. As it is expected, the highest power is generated by the gas turbine with 21,937 kW and followed by the steam turbine with 3726 kW power production. The power generation by the steam turbine presents the recovered power by the implementation of the bottoming cycle.

4.7.9.1.3

Conclusions

In this case study, a Brayton–Rankine combined system for electric generation is investigated and analyzed. The thermophysical properties, energy, and energy efficiencies for each component are calculated. The following concluding remarks can be stated:

• • •

The overall exergy efficiency is obtained as 38%. The overall energy efficiency is obtained as 35%. The energy efficiency of the system without the presence of the bottoming cycle is obtained as 18%.

Gas Turbine Cycles

253

100 90 Exergy destruction ratio (%)

80 70 60 50 40 30 20 10 0 Pump

Condenser

Steam turbine

Compressor

HEX boiler

Gas turbine

Combustion chamber Fig. 39 Exergy destruction ratios in the system.

100 90

Exergy efficiency (%)

80 70 60 50 40 30 20 10 0 Compressor

Gas turbine

Pump

HEX boiler

Steam turbine

Fig. 40 Exergy efficiencies of the system components.

• •

The system has 25.6 MW of generating electricity capacity. Integration of the Rankine cycle improves the overall system efficiency and power capacity significantly.

4.7.9.2

Case Study 2

In this case study, a combined system for compressed air energy storage (CAES), a sensible heat energy storage, and latent heat energy storage system are suggested. To increase the CAES system efficiency phase change materials (PCMs) are used. The system is designed as the presented schematic in Fig. 42. The proposed system consists of a compressor, motor-generator, gas turbine, solar tower, and energy storage systems. During the nighttime, when the energy demand is low, and electricity prices are relatively lower, the compressor works with the power it supplies from the grid and compresses the air from 100 to 7500 kPa. The temperature of the air reaches 1129K after compression. The compressed air flows to three different latent heat storage systems, which use MgCl2, LiBr, and a PCM, respectively. Thermophysical properties of PCM materials and molten salt can be seen in Table 10. After air

254

Gas Turbine Cycles

Generated power by components (kW)

25,000

20,000

15,000

10,000

5000

0 Gas turbine

Steam turbine

Fig. 41 Generated power by system components.

Compressor

Heliostat field 1

Gas Clutch turbine

Clutch

11

Motor / generator Reci ever

10

2

3

`

12

13 `

4

`

5

PCM 1

PCM 2

PCM 3

9

8

7

Cavern

15

14 ` Hot molten salt tank

HEX

6

Cold molten salt tank Fig. 42 Sketch of the developed system.

Table 10

Properties of phase change material (PCM) used in latent heat storage

Properties

MgCl2[30]

LiBr [30]

KNO3 [30]

Enthalpy of fusion [kJ/kg] Melting temperature [1C]

452 714

203 550

266 333

releases its heat to the PCMs, the temperature of air decreases to 651K. Pressurized air is stored in the cavern. Since the cavern is in underground and has even lower temperature than the ground, the temperature of the stored air decreases to 465K. During the daytime, these steps work in the reverse direction; pressurized air passes through PCMs and heats up. Heat losses arising from storage are compensated by the HEX. Finally, air at elevated temperature and pressure runs the turbine and generator to produce electricity. The HEX in the system supplies its energy from the solar system. When the solar normal irradiance is utilizable, heliostats reflect the sunlight to the receiver in the solar tower and heats up the molten salt (52NaCl–48MgCl2). In this study, it is assumed that solar energy can be useful for 12 h per day and maximum DNI is 900 W/m2. For the sake of simplicity variation of solar irradiance during daytime is assumed as a sinusoidal function. High-temperature molten salt is used at the HEX, then stored in a well-insulated tank.

Gas Turbine Cycles

Table 11

255

Design parameters of the system

Parameters

Value

Pressure ratio of compressor Compressor isentropic efficiency Atmospheric conditions Turbine isentropic efficiency Electromechanical efficiency Air mass flow used Molten salt mass flow used Thickness of insulator on molten salt tank Thermal conductivity of insulation (fiberglass) Pressure of turbine exhaust Maximum direct normal irradiation (I) Heliostat reflective area Number of heliostats Surface area of solar receiver area Wind speed

75 80% 100 kPa, 251C 77% 100% 30 kg/s 50 kg/s 5 cm 0.045 W/m-K 100 kPa 900 W/m2 10  5 m2 1000 50 m2 5 m/s

Table 12 Exergy efficiency and exergy loss plus destruction rate definitions for the system components

4.7.9.2.1

Component

Exergy loss þ exergy destruction rate definition

Exergy efficiency definition

Molten salt tank

_ _ 12 ex12 Ex dl;MST ¼ m

Zex;MST ¼

m_ 12 ex12 m_ 13 ex13

Cavern

_ _ 5 ex5 Ex dl;Cav ¼ m

Zex;MST ¼

m_ 6 ex6 m_ 5 ex5

PCM 1

_ _ 2 ex2 þ m_ 8 ex8 Ex dl;PCM1 ¼ m

m_ 3 ex3

_ 9 ex9 m

Zex;PCM1 ¼

_ 8 ex8 Þ ðm_ 9 ex9 m _ 3 ex3 Þ ðm_ 2 ex2 m

PCM 2

_ dl;PCM2 ¼ m_ 3 ex3 þ m_ 7 ex7 Ex

m_ 4 ex4

_ 8 ex8 m

Zex;PCM2 ¼

_ 7 ex7 Þ ðm_ 8 ex8 m _ 4 ex4 Þ ðm_ 3 ex3 m

PCM 3

_ _ 4 ex4 þ m_ 6 ex6 Ex dl;PCM3 ¼ m

m_ 5 ex5

_ 8 ex8 m

Zex;PCM1 ¼

_ 6 ex6 Þ ðm_ 7 ex7 m _ 5 ex5 Þ ðm_ 4 ex4 m

Gas turbine

_ _ 10 ex10 Ex d;GT ¼ m

Compressor

_ d;C ¼ m_ 1 ex1 þ W_ in;C Ex

Solar receiver

_ _ 13 ex13 Ex d;SR ¼ m

m_ 13 ex13 _ 6 ex6 m

m_ 11 ex11

W_ out;GT

m_ 2 ex2


m_ 14 ex14 þ Q_ SR 1

T0 Tsun



Zex;GT ¼

W_ out;GT ðm_ 10 ex10 m_ 11 ex11 Þ

Zex;C4 ¼

ðm_ 2 ex2 m_ 1 ex1 Þ W_ in;C

Zex;SR ¼

_ 14 ex14 Þ ðm_ 13 ex

13 m  T Q_ 1 0 SR

Tsun

Thermodynamic analysis

The energy and exergy efficiencies, exergy destructions, temperature, pressure, specific enthalpy and entropy are calculated for all system components. All elements of the system are constructed in the EES. The general design parameters and assumptions made for the analysis are listed in Table 11. Besides parameters in Table 11, the following assumptions are made for the study:

• • • •

Potential and kinetic energies are negligible. Ideal gas principles apply for the air. All pump works are negligible. No chemical reaction occurs throughout the system.

The balance equations are applied to each component of the integrated system and exergy losses and exergy efficiencies of the system elements are tabulated in Table 12. Regarding the inputs and outputs of the system, the overall system energy and exergy efficiencies can be obtained as follows: P _ Q ð57Þ Zen;ov;sys ¼ 1 P out _ in Q c;ov;sys ¼ 1

P Ex P d Ex in

ð58Þ

256

Gas Turbine Cycles

4.7.9.2.2

Solar receiver

The heat transfer rate between the solar receiver and the fluid flow can be calculated as follows [31]. _r ¼Q _h Q

_ loss ¼ m _ ðh3 Q

_ cp;air ðT3 T2 Þ h2 Þ ¼ m ð59Þ _ loss are the receiver losses due to convection and radiation. Q _h _ h is the thermal power received by the heliostat field and Q here, Q can be calculated as follows [31]. _ h ¼ Zh  Ah  N  I Q

ð60Þ

here, Zfield is solar effectiveness of the heliostat field, I is the direct normal irradiance (DNI), Ah is surface area of a heliostat sensor, and N is the number of heliostats. _ loss can be calculated as follows [31]. Q

 _ loss ¼ ha Ar ðT2 T0 Þ Ar s A T 4 T 4 Q ð61Þ r 0 here ha is heat transfer coefficient of air, Ar is the surface area of the receiver, s is Stefan–Boltzmann constant, and e is absorber emissivity. ha is calculated by using the following correlation [32].  pffiffiffiffi

ha ¼ 10:45 va þ 10 va ð62Þ W=m2 K here va shows the wind speed in terms of (m/s).

4.7.9.2.3

Motor work

The compressor operates on the power supplied from the grid and compresses the air. Since the motor-generator losses are neglected, power input by grid equals to compressor work: _ comp ¼ m _ 2 h2 W

_ 1 h1 m

ð63Þ

here, h1 and h2 are the specific enthalpies of the air at the exit and inlet of the compressor respectively. Energy losses at molten salt tank To calculate energy loss during storage, first, temperature differences should be calculated: rf Vcf

dy ¼ dt

UAy

ð64Þ

here rf is the density of the molten salt; V and cf present the volume and specific heat of the molten salt. t denotes time, y shows temperature decrease, A is heat transfer area, and U is overall heat transfer coefficient. The calculation of U is as follows: 1=U ¼ dins =kins þ 1=ha

ð65Þ

here kins is thermal conductivity of insulation material, which is fiberglass; dins is the thickness of insulation; and ha is heat transfer coefficient of air whose calculation is presented in Eq. (62) When Eq. (64) is integrated with respect to time t [33]: y T ðt Þ TO ¼ ¼e Ti TO yi

ðUAt Þ=ðrf cf V Þ

ð66Þ

here subscript i denotes the initial values. After calculation of y, Eq. (66) can be solved and heat loss during storage can be determined easily.

4.7.9.2.4

Case study results and discussion

Thermodynamic calculations of all system components are performed by using EES software. The values of temperature (K), pressure (kPa), mass flow rate (kg/s), specific enthalpy (kJ/kg), specific entropy (kJ/kg K), and specific exergy (kJ/kg) are determined for each state point of the system as listed in Table 13. The values of exergy destruction þ exergy loss rate (kW) and exergy efficiency (%) of each system component are presented in Table 13. In this section results are presented when the average direct irradiance DNI ¼ 450 W/m2 and time, t¼6. As seen in Table 14 and Fig. 43, the maximum exergy destruction rates are determined in the cavern and gas turbine. It is because of the high irreversibility of the components. The cavern lost 20% of its heat. To avoid this situation, the number of PCM units can be increased. The exergy efficiency of the gas turbine is observed as 85.15%. As presented in Table 14 and Fig. 43, the exergy destruction rates of PCM 1 have the lowest share in the overall system. Fig. 44 presents exergy efficiencies of each component of the thermal energy storage system when DNI is equal to 450 W/m2. The molten salt tank has the highest exergy efficiency by 96% and PCM 3 has the lowest exergy efficiency by 54.8%. The main reason for the high exergy efficiency of the molten salt tank is that heat losses at the tank are very low due to its good insulation. The main reason for the low exergy efficiency of the solar receiver is heat losses at higher temperatures, which increase the exergy losses. Exergy efficiency of PCM 3 is also lower due to the heat losses at high temperature. The heat losses at the solar receiver are taken into account in this study. Hence, it causes significant exergy destruction rate on this component. Another reason for this situation can also be explained as the heat transfer through the solar receiver occurs at high-temperature differences. To decrease the exergy destruction rate of this component, the heat losses need to be decreased.

Gas Turbine Cycles

Table 13

257

Thermodynamic data for all state points

State No.

Stream

0 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15

Air Molten Air Air Air Air Air Air Air Air Air Molten Molten Molten Molten Molten Molten

salt

salt salt salt salt salt salt

Table 14

T (K)

P (kPa)

m_ (kg/s)

298.2 298.2 298.2 1129 1042 828.2 651.2 465.3 601.2 778.2 857.2 1277 624.3 1282 1253 1552 1552

100 100 100 7500 7500 7500 7500 7350 7350 7350 7350 7350 100 100 100 100 100

– – 30 30 30 30 30 30 30 30 30 30 30 50 50 50 50

h (kJ/kg)

s (kJ/kg
K)

298.40 453.50 298.4 1201 1100 856.6 662.2 463.5 608.3 801 888.9 1374 633.1 609.1 578.6 900.3 900.3

6.86 0.95 6.864 7.036 6.943 6.682 6.418 6.065 6.338 6.619 6.726 7.187 7.62 0.626 0.6021 0.8325 0.8325

– – 0 851 778 612 497 403 467 576 632 980 109 593 570 823 823

Thermodynamic analysis results of the developed system

Component

Exergy destruction rate (kW)

Exergy efficiency (%)

Compressor Cavern Phase change material (PCM) 3 PCM 2 PCM 1 Heat exchanger (HEX) Gas turbine Molten salt tank Solar receiver

1538 2802 1569 1690 524 1038 3879 1166 2630

94.32 81.20 54.80 65.92 76.17 90.96 85.15 96.07 74.12

100 Exergy loss plus destruction ratio (%)

ex (kJ/kg)

Gas turbine Cavern

80 Solar reciever 60

PCM2 PCM3

40

20

Compressor Molten salt tank HEX

0

PCM1

Fig. 43 Exergy destruction ratios in the system.

Fig. 45 shows energy efficiencies of each component of the thermal energy storage system when DNI is equal to 450 W/m2. The molten salt tank has the highest energy efficiency by 95% and PCM 3 has the minimum energy efficiency by 74.5%. The key reason for the minimum and maximum energy efficiencies of the components are the same as the reasons for the exergy efficiencies. Since the heat transfer rates and melting temperature differences are not evenly distributed between PCM units, some heat losses are

258

Gas Turbine Cycles

100 90 80

Exergy efficiency (%)

70 60 50 40 30 20 10

r ie ec

M

ol

C

te

n

R

sa

lt

ta

ve

nk

T G

EX H

1 M PC

2 M PC

3 M PC

er av C

om

pr

es

so

n

r

0

Fig. 44 Exergy efficiencies.

100 90

Energy efficiency (%)

80 70 60 50 40 30 20 10 0 PCM3

PCM2

PCM1

Molten salt tank

Fig. 45 Exergy efficiencies of system components.

essential at the surface of the PCM units. Otherwise, PCMs melt completely. In order to avoid this situation, the number of PCM units can be increased, and selection of PCM materials can be made according to their melting temperature as explained above.

4.7.9.2.5

Parametric studies

Here, some key parameters are varied to investigate the impacts on the system performance.

4.7.9.2.6

Effects of direct normal irradiance

Fig. 46 represents the change of the overall exergy efficiency and exergy destruction plus exergy loss with average DNI. As it can be seen in Fig. 46, the exergy is highly dependent on solar contribution. The reason for that, as shown in Fig. 47, is that the exergy destruction rate of the solar receiver has larger relative values. As the DNI increases, temperature of the receiver surface rises as well. That situation causes high heat transfer between the receiver and molten salt. This is the main reason for increasing exergy of molten salt and solar receiver’s efficiency. On the other hand, higher temperature gradient leads to higher heat losses, especially by

19,000

100 95

Exergy efficiency (%)

18,000

Ex

ex

90

17,000

85 80

16,000

75 15,000

70 65

14,000

60 13,000

55 50 100

12,000 300

500

700

259

Exergy destruction plus exergy loss rate (kW)

Gas Turbine Cycles

900

Direct normal irradiance (W/m2)

100

5000

90

4500

80

4000

70

3500

60

3000

50

2500

40

2000

30

1500

20

Ex

ex

10

500

0 100

1000

Exergy destruction plus exergy loss (kW)

Exergy efficiency (%)

Fig. 46 Variation of overall exergy efficiency and exergy destruction plus exergy loss with average direct normal irradiance (DNI) level.

300

500

700

0 900

Direct normal irradiance (W/m2) Fig. 47 Variation of exergy efficiency of the solar receiver and exergy destruction plus exergy loss with the average direct normal irradiance (DNI) level.

the radiation. Also, heat losses at higher temperature gradient cause larger exergy destruction. That situation can be seen in Fig. 47. As the average DNI increases, the efficiency of the solar receiver increases from 40 to 73% and exergy destruction rate of hot molten salt increases almost by 4700 kW. High quantity of exergy input and destruction rate affects overall efficiency as expected.

4.7.9.2.7

Effects of ambient temperature

In Fig. 48, the change of overall exergy efficiency and exergy destruction plus exergy loss regarding ambient temperature is presented. As the ambient temperature increases, the quantity of usable energy decreases. For instance when the ambient temperature increases from 5 to 351C, exergy destruction rate at the compressor rises by almost 400 kW. Therefore overall exergy efficiency of the system slightly lowers down from 64.4 to 62.3%.

4.7.9.2.8

Effects of wind speed

Fig. 49 shows the change of the exergy efficiency of the solar receiver and temperature of molten salt with respect to the wind speed. Exergy destruction is at its lowest level when there is no wind. As the losses related to convection increase by the wind speed, there are more exergy losses observed on the surface of the receiver. Therefore, exergy efficiency of the solar receiver drops from

Gas Turbine Cycles

16,000

65

15,900 Exergy efficiency (%)

64.5

15,800 15,700

64

15,600 63.5

15,500 15,400

63 Exergy efficiency 62.5

15,300

Exergy destruction plus exergy loss

62 275

280

285

290

295

300

305

15,200

Exergy destruction plus exergy loss (kW)

260

15,100 310

Ambient temperature (K) Fig. 48 Variation of overall exergy efficiency and exergy destruction plus exergy loss with the ambient temperature.

77 1559 76 ex

1555

76

1553

75

1551

75

1549

74

1547

74

Exergy efficiency (%)

Temperature (K)

77

Temperature

1557

73

1545 1

2

3

4

5

6

7

8

9

10

Wind speed (m/s)

100.0

1256

99.5

1255 Temperature

ex

99.0

1255

98.5

1254

98.0

1254

97.5

1253

97.0

1253

96.5

1252

96.0

Temperature of molten salt

Exergy efficiency (%)

Fig. 49 Variation of exergy efficiency of the solar receiver and temperature of molten salt with the wind speed.

1252 0

2

4

6

8

10

Wind spreed (m/s) Fig. 50 Variation of exergy efficiency of molten salt tank and temperature of molten salt with the wind speed.

75.5 to 73% as the wind speed and temperature of the molten salt is decreased from 1549 to 1530K. Moreover, the effects of wind speed also can be seen on the molten salt tank. In Fig. 50 it can be seen that effect of convection has very little weight on temperature change and exergy efficiency of the storage tank. The key factor about that situation is the fact that the tank is wellinsulated so that thermal resistance due to conductive resistance dominates the overall thermal resistance of the tank.

1400

90.0

1300

Exergy efficiency (%)

100.0

80.0

1200

Temperature

ex

1100

70.0

1000 60.0

900

50.0

800

40.0

700

30.0

261

Temperature of molten salt

Gas Turbine Cycles

600 0

5

10

15

20

25

30

Thickness of insulation (cm) Fig. 51 Variation of exergy efficiency of molten salt tank and temperature of molten salt with respect to insulation thickness.

4.7.9.2.9

Effects of tank insulation

As explained in the paragraph above, the contribution of conduction dominates the overall heat transfer coefficient. Here, one should set the thickness of insulation very carefully. As the lateral area of the molten salt tank is enormous, even 1 cm of thickness increase can affect insulation cost considerably. In Fig. 51, it can be clearly seen that insulation thickness has a massive effect for the first 3 cm. If there is no insulation over the tank, the temperature of molten tank drops below 700K. However, for the case with 3 cm thickness insulation, the temperature drops to 1350K at the end of the storage period. After 5 cm insulation thickness, the temperature only slightly changes with the increasing insulation. In this parametric study, growing of the heat transfer surface with the increase of insulation material’s thickness is neglected since the parametric study is conducted in a small range. The error of this analysis will be larger for higher thicknesses.

4.7.9.2.10

Conclusions

A combined system for CAES, sensible heat energy storage, and latent heat energy storage system is suggested and investigated. The thermophysical properties, energy, and energy efficiencies for each component are calculated. The performance of the system for 450 W/m2 average DIN during the day is determined. The following concluding remarks can be stated:

• • • • • • • •

The overall exergy efficiency is obtained as 64.16% at 450 W/m2 direct solar irradiance solar contribution. The overall energy efficiency is obtained as 61.84 at 50% solar contribution. Maximum exergy destruction rates are determined in a gas turbine, cavern, and solar receiver. Losses related to convection have less impact on the molten salt tank. Temperature of the molten salt drops 28.71C in 12 h. Exergy efficiency of the latent heat storage is obtained as 79.26%, 74.49%, and 86.98%, respectively. By arranging the configuration of the HEXs and selecting the PCMs regarding their melting temperature differences as being evenly distributed between one PCM and another, the efficiencies can be increased. It is possible to recover 23.15 MW power from the gas turbine. Energy and exergy losses at the cavern can be decreased by installing more PCM units.

4.7.10

Future Directions

Gas turbine technology has been developing through the centuries. In the 18th century fundamentals of gas turbines were introduced by the pioneers of the technology. Afterward in the early 1900s, applications of the simple cycle gas turbines spread to many industries. Power plants have long used gas turbines in a simple-cycle configuration for a restricted maximum power generation. Furthermore, commercialized facilities utilize gas turbine components for ground basis power production, typically in arrangement with a heat generation process, such as steam generation, hot water, or air heating. Currently, the performance of industrial gas turbines has improved because of the significant incentives in research and development regarding fuel to electricity conversion efficiency, the capacity of the power plant, viability, and dependability [4]. Ongoing progress and the recent developments of innovative gas turbines will enhance the performance of the simple cycle facilities over 40%. The integration of the gas turbine cycle, such as the Brayton cycle, with a moderate or low temperature bottoming cycle such as the Kalina cycle or Rankine cycle, which is known as the typical combined cycle, is the best operational way to raise the thermal efficiency of a gas turbine cycle. Heavy duty NG combusted gas turbines together with WHR vapor generators and vapor turbines indicate the main idea of this approach [34].

262

Gas Turbine Cycles

For near the future, the fuel flexibility and intermediate capacity peak power generation are considered as the trends for gas turbine system development, where the rise of combustion temperature and challenges to new cycles are considered as mediumand long-term trends for the GTC [35]. New technologies toward the fuel range will direct the near future developments. Claire Soares [36] listed these developments, which are briefly explained as follows: Coal fuel combustion related: two innovative developments regarding the coal being utilized as a fuel in gas turbine systems appear to shape the future progresses: oxycombustion and hydrogen turbines. Theoretically, oxyfuel combustion systems can be used in both conventional and advanced power plants. Around 30% nominal efficiency can be attained by today’s systems with steam turbines when fueled with NG and capturing the carbon dioxide. Researchers focus their studies on improving the performance of this system by reducing the emissions to nearly zero and increasing the efficiency around 60% [36]. Hydrogen turbines: by the development of the hydrogen turbines, turbine systems and system elements are aimed to be improved; these include cooling and combustor technology, as well as materials and coating research. Those concepts appear to be the essential technologies for the next generations. A power plant system including an advanced turbine could be built to operate the world’s first near-zero emissions power plant to generate electricity and H2 from coal while capturing and storing carbon dioxide through sequestration [36]. Beside the improvement on fuel flexibility, thermodynamic cycle technologies like recuperation, aftercooling, intercooling, and cycle integration can be considered as possible ways to enhance the performance of gas turbine based power plants at feasible costs within the near future [4].

4.7.11

Concluding Remarks

The discipline of GTCs is a result of the accumulation of centuries’ worth of experience and knowledge. From John Barber’s invention of the first real gas turbine to the modern gas turbines, many enhancements are made by technology. They are some of the most reliable options especially for the situations where continuous high power output is required. Therefore, they have been used for many purposes from marine–train–aircraft propulsion to power generation. High availability of the working fluid of the cycle (like air) makes their application more practical. Moreover, the compact structure of the simple GTCs allows high power generation with a quick response yet with a low thermal efficiency (around 40%). To overcome this problem, various types of CCPPs have been proposed, which utilize the exhaust stream of the topping cycle. Today, it is possible to reach around 60% thermal efficiency with combined cycles.

Acknowledgment The authors acknowledge the support provided by the Turkish Academy of Sciences (TÜBA) and the Natural Sciences and Engineering Research Council of Canada.

References [1] Dincer I, Zamfirescu C. Advanced power generation systems. 1st ed. Oshawa: Elsevier; 2014. [2] Meher-Homji CB, Zachary J, Bromley AF. Gas turbine fuel system design. Combustion and operability. In: Proceedings of the thirty-ninth turbomachinery symposium. Texas; 2010. [3] Poullikkas A, Technology A. Selection algorithm for independent power producers. Electr J 2001;14:80–4. [4] Poullikkas A. An overview of current and future sustainable gas turbine technologies. Renew Sustain Energy Rev 2005;9:409–43. [5] Hunt RJ. The history of the industrial gas turbine (Part 1 the first fifty years 1940–1990). Indep Tech Forum Power Gener 2011;15(2):1–48. [6] Cleveland CJ, Morris C. Handbook of energy, Volume II chronologies, top ten lists, and word clouds. 2. Waltham, MA: Elsevier; 2014. p. 323–32. [7] Davey N. The gas turbine – development and engineering. New York, NY: Watchmaker Publishing; 2003. [8] Lincoln JW. New steam age, the magazine of modern steam power. Connecticut: Stonington Publishing Co.; 1942. [9] Meyer A. The combustion gas turbine: its history, development and prospects. Brown Boveri Company. London: The Institution of Mechanical Engineers; 1939. [10] Ackeret J. A pioneer of modern aerodynamics, Founder IfA – Institute of Fluid Dynamics, Zurich; 2009. [11] Bowden AT, Jefferson JL. The design and operation of the parsons experimental gas turbine. In: IMechE proceedings, Newcastle upon Tyne; 1948. [12] Andriani R, Ghezzi U, Pasini S. Thermodynamic study of gas turbine engine with constant volume combustion. In: 8th annual international energy conversion engineering conference. Nashville; 2010. [13] Cengel YA, Boles MA. Thermodynamics: an engineering approach. 8th ed. New York, NY: Mcgraw-Hiill Education; 2015. [14] Moran MJ, Shapiro HN, Boettner DD, Bailey MB. Fundamentals of engineering thermodynamics. John Wiley & Sons; 2010. [15] Horlock JH. Basic gas turbine cycles. In: Advanced gas turbine cycles. Elsevier Ltd; 2003: pp. 27–46. [16] Dincer I, Rosen MA. Exergy. 2nd ed Amsterdam: Elsevier; 2013. [17] Etele J, Rosen MA. Exergy losses for aerospace engines: effect of reference-environments on assessment accuracy. AIAA Paper; 2000. [18] Clarke JM, Horlock JH. Availability and propulsion. J Mech Eng Sci 1975;17:223–32. [19] Korobitsyn MA. New and advanced conversion technologies: analysis of cogeneration, Combined and integrated cycles [Ph.D. thesis]. University of Twente; 1998. [20] California Energy Commission. Available from: http://www.energy.ca.gov. [21] Rolls-Royce. Available from: http://www.rolls-royce.com. [22] Power GE. Available from: http://www.gepower.com; 2004 [accessed 01.01.04]. [23] U.S. Department of Energy. Available from: http://www.fe.doe.gov.

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[24] Rao AD, Yi Y, Samuelsen GS. Gas turbine based high efficiency, ‘Vision 21’ natural gas and coal central plants. In: Proceedings of the first international conference on industrial gas turbine technologies; 2005. [25] Lund University. Available from: http://www.vok.lth.se. [26] Giampaolo T. The gas turbine handbook: principles and practice. 2nd ed. New York, NY: The Fairmont and Marcel Dekker; 2003. [27] Dincer I, Rosen MA, editors. Exergy: exergy and energy analyses Amsterdam: Elsevier; 2013. [28] Bicer Y, Dincer I. Energy and exergy analyses of an integrated underground coal gasification with SOFC fuel cell system for multigeneration including hydrogen production. Int J Hydrogen Energy 2015;40:13323–37. [29] Boundy B, Diegel S, Wright L, Davis SC. Biomass energy data book. 4th ed. 4; 2011. p. 254. [30] Cardenas B, Leon N. High temperature latent heat thermal energy storage: Phase change materials, design considerations and performance enhancement techniques. Renew Sustain Energy Rev 2013;27:724–37. [31] Meriche IE, Beghidja A, Boumedjirek M. Energetic and exergetic analysis of solar gas turbine power plant in South Algeria. In: IREC 2014 – 5th international renewable energy congress; 2014. [32] European Regional Development Fund and European Neighbourhood and Partnership Instrument, Radiant heating. Convection heating systems and wall tempering, WP5 Education Econonic Promotion (CO2OL Bricks). 1; 2014. [33] Nag PK. Power plant engineering. 1st ed. New York, NY: Tata McGraw-Hill Education; 2002. [34] Poullikkas A. Parametric study for the penetration of combined cycle technologies into Cyprus power system. Appl Therm Eng 2004;24:1697–707. [35] Fukuizumi Y. Gas turbine technology: the future for gas turbines – power engineering international. Available from: http://www.powerengineeringint.com; 2017 [accessed 26.02.17]. [36] Soares C, editor. Future trends in the gas turbine industry: gas turbines, 2nd ed. Elsevier; 2015.

Further Reading Dincer I, Zamfirescu C. Advanced power generation systems. 1st ed. Oshawa: Elsevier; 2014. Dincer I, Rosen MA. Exergy: energy, environment and sustainable development. 2nd ed. Elsevier; 2013. Cengel YA, Boles MA. Thermodynamics: an engineering approach. 8th ed. Mcgraw-Hiill Education; 2015. Horlock JH, Bathie WW. Advanced gas turbine cycles. 1st ed. Elsevier Science; 2003. Giampaolo T. The gas turbine handbook: principles and practice. Cambridge: Cambridge University Press; 2003. Han J-C, Dutta S, Ekkad S. Gas turbine heat transfer and cooling technology. 2nd ed. CRC Press; 2012. Bejan A. Advanced engineering thermodynamics. 3rd ed. John: Wiley and Sons; 2006.

Relevant Websites http://www.alstom.com Alstom Power. http://www.gasturbineworld.com/ Gas Turbine World. https://www.geaviation.com/ GE Aviation. https://powergen.gepower.com/applications.html GE Power. http://www.globalenergyobservatory.org/list.php?db=PowerPlants&type=Gas Global Energy Observatory (GEO). https://www.iea.org/ International Energy Agency. https://global.kawasaki.com/en/corp/profile/division/gasturbine/ Kawasaki Heavy Industries. http://turbomachinery.man.eu/products/gas-turbines MAN Diesel & Turbo. http://www.mhps.com/en/products/thermal_power_plant/gas_turbin/index.html Mitsubishi Hitachi Power Systems. http://web.mit.edu/16.unified/www/SPRING/propulsion/notes/node27.html MIT Thermodynamics and Propulsion Lecture Notes. https://www.netl.doe.gov National Energy Technology Laboratory. http://nptel.ac.in/courses/112105123/24 National Programme on Technology Enhanced Learning (NPTEL). https://www.siemens.com/global/en/home/products/energy/power-generation/gas-turbines.html#!/ Siemens. https://energy.gov/ U.S. Department of Energy.

4.8 Steam and Organic Rankine Cycles Ibrahim Dincer and Murat E Demir, University of Ontario Institute of Technology, Oshawa, ON, Canada r 2018 Elsevier Inc. All rights reserved.

4.8.1 4.8.2 4.8.3 4.8.3.1 4.8.3.1.1 4.8.3.1.2 4.8.3.2 4.8.4 4.8.4.1 4.8.4.1.1 4.8.4.1.2 4.8.4.1.3 4.8.4.1.3.1 4.8.4.1.3.2 4.8.4.1.4 4.8.4.1.5 4.8.4.1.6 4.8.4.1.7 4.8.4.1.8 4.8.4.1.9 4.8.4.2 4.8.4.2.1 4.8.4.2.2 4.8.4.2.3 4.8.4.2.3.1 4.8.4.2.4 4.8.5 4.8.5.1 4.8.5.1.1 4.8.5.1.2 4.8.5.1.2.1 4.8.5.1.2.2 4.8.5.1.2.3 4.8.5.1.3 4.8.5.2 4.8.5.2.1 4.8.5.2.1.1 4.8.5.2.1.2 4.8.5.2.1.3 4.8.5.2.1.4 4.8.5.2.2 4.8.5.2.2.1 4.8.5.2.2.2 4.8.5.2.3 4.8.5.3 4.8.5.3.1 4.8.5.3.1.1 4.8.6 4.8.7 Acknowledgment References

264

Introduction Historical Development Classification of Vapor Power Cycles Steam Power Cycle The Carnot vapor cycle Rankine cycle Organic Rankine Cycles Thermodynamic Analysis of Vapor Power Cycles The Rankine Cycle Actual Rankine cycle Ideal reheat Rankine cycle Regenerative Rankine cycle Open feedwater heaters Closed feedwater heaters Ammonia–water Rankine cycle Kalina cycle Ammonia–water trilateral flash-Rankine cycle Supercritical Rankine cycle with inorganic fluids Combined cycle power plants Exergy destructions in Rankine cycle power plants Organic Rankine Cycles Concentrated solar collector Organic Rankine cycle Gas turbine cycle Vapor absorption chiller Working fluids for organic Rankine cycle Case Studies Case Study 1 Energy and exergy analyses Results and discussion Influence of varying ambient pressure and temperature on the subsystems and the overall system efficiencies Influence of varying the global solar irradiance on solar system and overall efficiencies Variation of exergy destruction rate and exergy efficiency of the overall system and subsystems by changing the ambient temperature Final remarks Case Study 2 Energy and exergy analyses Heliostat field and solar receiver Molten salt tank Multistage flash distillation Overall energy and exergy efficiencies Results and discussion Effects of ambient temperature Effects of ambient wind speed Final remarks Case Study 3 Energy and exergy analyses Effect of steam flow rate in SRSPR on overall efficiencies, SRSPR efficiency and combined cycle efficiencies Future Directions Concluding Remarks

Comprehensive Energy Systems, Volume 4

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doi:10.1016/B978-0-12-809597-3.00410-7

Steam and Organic Rankine Cycles Further Reading Relevant Websites

Nomenclature cp cv DNI EIF _ Ex ex h k

Specific heat capacity at constant pressure (kJ/kg K) Specific heat capacity at constant volume (kJ/kg K) Direct normal irradiance (W/m2) Efficiency improvement factor Exergy rate (kW) Specific exergy (kJ/kg) Specific enthalpy (kJ/kg), Convective heat transfer coefficient (W/m2 K) Thermal conductivity (W/m-K), Specific heat ratio

Acronyms AF ASTM

311 311

LHV _ m P _ Q r R s T U v V w _ W

IMechE LNG LPG N NG PCM RAF SI UPR WHR

Greek letters Zen Energy efficiency c Exergy efficiency Stefan Boltzmann constant sSB (5.76  10 8/m2 K4)

l ζ Ε

Air–fuel equivalence ratio Specific exergy function Surface emissivity

Subscripts ac C d field GT HEX i In ins ise Out ov

p P Ph reg s si sr Sys T 0 1 1, 2, … a

Propulsion Pump Physical Regeneration Shaft power Heat sink Heat source System Turbine, thrust power Ambient condition Freestream State points

and superscripts Aircraft Compressor Destruction Heliostat field Gas turbine Heat exchanger State number Inlet of a component Insulation Isentropic Outlet of a component Overall

HEX HRSG IDGTE

Lower heating value (kJ/kg) Mass flow rate (kg/s) Pressure (kPa) Heat rate (kW) Compression ratio Universal gas constant (J/K mol) Specific entropy (kj/kg K) Temperature (k) Overall heat transfer coefficient (W/m2 K) Specific volume (m3) Velocity (m/s) Specific work (kJ/kg) Work rate (kW)

Heat exchanger Heat recovery steam generation Institution of Diesel and Gas Turbine Engineers Institution of Mechanical Engineers Liquefied natural gas Liquefied petroleum gas Total number Natural gas Phase change material British Royal Air Force Spark-ignition Union Pacific Railroad Waste heat recovery

Air–fuel ratio American Section of the International Association for Testing Materials ATS Advanced turbine systems C Compressor CAES Compressed air energy storage CCPP Combined cycle power plants CEGB Central electricity generating board CI Compression-ignition CPC-GTC Constant pressure combustion gas turbine cycle CVC-GTC Constant volume combustion gas turbine cycle EES Engineering equation solve GTCs Gas turbine cycles

265

266

4.8.1

Steam and Organic Rankine Cycles

Introduction

Power generation has been an important subject in meeting the electricity demands (so-called electrification) in almost every part of the world. Even though there are two critical cycles, namely, the Rankine cycle (mostly using water as working fluid) and Brayton cycle (using air as the working fluid), the Rankine cycle offers some advantages over air Brayton cycle, such as being able to operate at lower temperatures. Steam turbine power systems have been finding applications for commercial purposes in various sectors, ranging from energy production to propulsion in marine vessels since early the 19th century. Usually, these systems consist of a simple closed standard Rankine cycle in which a pump takes the water at the vacuum conditions and pressurizes it, a boiler where the water changes its phase at constant temperature and is then superheated, and a steam turbine, where the steam is expanded to generate electricity [1]. The steam turbine is a useful device to convert heat energy into mechanical energy by rotating shaft with higher capacities and efficiencies. Also, another benefit of the steam turbine systems is that these systems possess the ability to use various types of fuels in the boiler and also various fluids can be used as the working fluid. While natural gas (NG), liquefied natural gas (LNG), liquefied petroleum gas (LPG), refinery gas, coke oven gas, coal gas, and hydrogen are considered some of the fuels used in steam turbines, secondly diesel, kerosene, jet A fuel, naphtha, ethanol and methanol, heavy residual-grade oils, and crude oils can also be used as the liquid fuels in the vapor generator [2]. The energy efficiency of a modern coal-fired Rankine cycle power plant can reach over 40 and it can be increased by means of using combined or integrated cycle options and/or heat recovery (and regenerative) subsystems. Moreover, reheating is employed to improve the performance of the system. Inexpensive and readily available working fluids (such as air and water), along with well-developed technologies (such as gas turbine (GT), heat recovery steam generator, or steam turbine units), all within a short construction period and higher overall efficiencies have helped to achieve greater acceptance of these systems. Note that combined cycle plants can reach up to the efficiencies well above 58% with plant capacities in the range between 350,000 and 500,000 kWe [3,4]. In this chapter, we generally deal with steam Rankine and organic Rankine cycles (ORCs) to discuss their historical developments in several sectors (including utility and marine sectors) and classifications, to present their thermodynamic studies through energy and exergy approaches, to compare theoretical (ideal) versus actual cycles and their performances, and to develop analyses and assessments for integrated and combined systems. There are also sample problems and case studies presented to illustrate the importance of steam and ORC systems and their critical role in achieving the generation of multiple commodities. Furthermore, the future direction for the steam and ORCs is discussed.

4.8.2

Historical Development

The modern steam turbines utilized in industrial applications have evolved to their present formation after going through a series of technological research, improvements, and developments that emerged in the 18th century. As well, they became an essential element for power generation during the industrial revolution. Cleveland and Morris [5] chronologically investigated the technological development of the energy conversion technologies and stated some significant milestones and provide a brief historical summary of the early period of the steam turbines [5].

• • •

• • • • •

1859: the Scotsman William Rankine introduced a closed thermodynamic cycle, which converts heat to work. The Rankine cycle systems generally utilize water as the working fluid and even in the 21st century, they are responsible for much of the global electricity generation. 1883: the Swedish engineer and inventor Karl Gustaf Patrik de Laval introduced a high-speed steam turbine by developing an innovative reduction gear that allows the high-speed rotor to drive the relatively lower speed propeller. This principle has been used in the marine industry for the ensuing years. 1884: the English engineer and inventor Charles Parson, who also has been credited for developments in optical devices, patented the first compound steam turbine, which converts the steam power directly into the electricity. The invention of the steam turbine has a significant influence on the naval propulsion systems and the power generation stations hence it is considered one of the most significant inventions of the 19th century. 1890: Karl Gustaf Patrik de Laval enhanced the steam turbines designs with a convergent–divergent nozzle that effectively accelerates the stream. The design later on was not only used in steam turbines but also it played a vital role in modern rocket engine and jet engine designs. 1894: Charles Parson took out a patent for “propelling a vessel using a steam turbine, which turbine actuates the propeller or paddle shaft directly or through gearing.” This allows marine vessels to trail with higher speed more efficiently. 1894: Swedish engineer, technical designer, and industrialist brothers Fredrik and Birger Ljungström together received a patent for their statorless centrifugal steam turbine. 1896: the American engineer, inventor, and patent attorney Charles Gordon Curtis developed the first velocity-compound steam engine, which reduced the size and weight of the engines previously used around 90%. 1896: the French engineer Auguste Rateau obtained the patent for a multistage turbine where the steam stream passes through from higher pressure to lower pressure stages consequently increasing the volume of the stream, while there is a decline in the pressure.

Steam and Organic Rankine Cycles

• • • • • • • •

267

1897: Turbinia, which was the first steamship driven by a steam turbine, was constructed in the United Kingdom. The promising performance of the vessel increased the popularity of the steam turbines as a viable alternative to the reciprocating steam engines in both marine industry and power stations. 1899: Charles Gordon Curtis patented his vertical steam turbine design, which led to a significant rate of electricity generation for the following decades of the century. 1901: the first merchant vessel propelled by steam turbines, TS King Edward, was built and launched in Dumbarton, Scotland. 1906: the British Royal Navy vessel HMS Dreadnought, which was the first battleship powered by steam turbines, was launched. The steam turbines were replacing the expansion reciprocating engines and it resulted in a considerable increase in speed of the battleships. 1907: the British passenger ships Lusitania and Mauretania were constructed, which were the first ocean liners driven by steam turbines of their kind. 1925: the first high-pressure steam turbine that can operate under 8274 kPa was introduced by Boston Edison in the United States. 1944: Pennsylvania Railroad used the highest capacity direct-drive steam turbine locomotive, which had maximum power output over 5 MW and could reach over 160 km/h. 1961: physicist Harry Zvi Tabor and engineer Lucien Bronicki developed an ORC in which there was an organic-based working fluid with high molecular mass and relatively lower boiling point. The application of ORCs allows power generation at lower temperatures.

4.8.3

Classification of Vapor Power Cycles

The steam-powered systems were the leading technologies of the 18th and 19th centuries for power production and marine industry. The Rankine cycle applications are used for only steam up until the introduction of ORC, which has been used in commercial level in the last few decades particularly with renewables. In this section classification of the vapor powered cycles is done considering the processes and working fluids as presented in Fig. 1.

4.8.3.1

Steam Power Cycle

Steam power cycles are utilizing water in alternatively vaporized and condensed form as the working fluid. Due to the promising characteristics of the steam, such as being economically viable, abundant in nature, and high heat of evaporation, it is the most preferred working fluid within the vapor power generation systems.

4.8.3.1.1

The Carnot vapor cycle

The Carnot cycle is a theoretical power cycle and is used for showing the theoretical maximum of a heat engine. Both isothermal heat addition and rejection processes are very challenging to achieve, as they need large heat exchangers (HEXs) and very long heat transfer time. Nevertheless, in practice, actual heat engines cannot have either of those conditions. Therefore, those cycles show the ideal conditions and indicate the theoretical limits of the systems. The energy efficiency of any totally reversible heat engine, such as Stirling, Carnot, or Ericson, between a heat source at the temperature TH and a heat sink at the temperature TL is described by the
 Carnot factor Zcar ¼ TTHL . Any other power cycle that operates between those temperatures cannot go beyond the Carnot efficiency. Hence, in this section, a Carnot cycle, which is using steam as the working fluid, investigated and briefly explained on a T–s diagram. Consider a steady-flow Carnot cycle operating within the limits of saturation dome as presented in Fig. 2 [6].

•

Process 1–2: the steam is heated isothermally, while the space is heated externally by the heat source, qin. Heat transfer at the constant temperature of a two-phase system is easy to attain in actual applications since phase changing occurs at the constant temperature in fixed pressure systems. The main restriction in this process is the temperature should not exceed the critical point value, which is 3741C for water. Constraining the temperature comes with the limitation of the cycle efficiency.

Steam rankine cycle Organic rankine cycles (ORC)

Rankine cycle Vapor power cycles

No-steam rankine cycles Carnot vapor cycle

Ammonia−water cycles Nonorganic rankine cycles Carbon dioxide cycles

Fig. 1 Classification of vapor power cycles.

268

Steam and Organic Rankine Cycles

T

qin

2

1

TH

TL

4

3

qout

s Fig. 2 T–s diagram of the Carnot vapor cycle.

T

TH

1

qin

2

qout

3

TL 4

s Fig. 3 T–s diagram of the alternative Carnot vapor cycle.

•

• •

Process 2–3: at this step, the heat addition to the cycle stops and the steam starts to expand in the thermally isolated medium at a constant entropy until the temperature declines from TH to the temperature of the heat sink, TL. During the expansion, the system also produces the useful work output. However, it is impractical in actual applications. The quality of the steam drops during the expansion and leads increase in the moisture content of the stream. Malfunctions due to erosion occur as the liquid droplets impinge the turbine blades. Therefore, steam with qualities less than around 0.9 cannot be tolerated in the power plants. This issue could be overcome by using a working fluid with a very steep saturated vapor line. Process 3–4: the steam condenses isothermally, while the space discharges the heat to the heat sink, qout. This can also be approached closely as process 1–2 in actual condensers. Process 4–1: heat rejection from the cycle stops and steam starts to be compressed into the thermally isolated media isentropically. The temperature of the steam reaches the temperature of the heat source, TH. During the compression, the system requires the work input. However, this impractical in two ways: first, controlling the condensation process, which finishes exactly at state point 4, is very hard to handle. Secondly, to design a compressor that works with two-phase is not easy to achieve.

Some of the aforementioned issues can be eliminated by performing the Carnot vapor cycle in an alternative way as presented in Fig. 3. Nevertheless, the alternative Carnot vapor cycle comes with other problems such as isothermal heat transfer at variable pressures and isentropic compression to extremely high pressures. Therefore, it is stated that the Carnot vapor cycle cannot be approximated in actual vapor driven systems [6].

Steam and Organic Rankine Cycles

269

T 3

qin 3a

3b

2 1

4 qout s

Fig. 4 T–s diagram of the ideal Rankine cycle.

4.8.3.1.2

Rankine cycle

Many of the impracticalities related to the Carnot vapor cycle can be overcome by superheating the steam in a boiler and condensing it entirely in a condenser [6]. Such cycle adhering these ideal processes is named as Rankine cycle and can be seen in Fig. 4. Rankine cycle is considered as the most basic vapor power cycle in which there are no losses associated with frictional pressure drops in both condenser and boiler and the working fluid passes through those components at a constant pressure. The ideal Rankine cycle can be executed by with a system consisting of:

• • • •

an ideal pump (reversible and adiabatic) an ideal HEX, boiler (with no heat losses and pressure drops) an ideal steam turbine (reversible and adiabatic) an ideal condenser (with no heat losses, pressure drops and an infinite heat transfer coefficient without a temperature difference between condensing steam and heat sink) The ideal Rankine cycle does not involve any internal irreversibilities and comprises four processes as follows:

• • • •

1–2 2–3 3–4 4–1

isentropic pressurization in a pump isobaric heat addition in a boiler isentropic expansion of the superheated steam in a turbine isobaric heat rejection in a condenser

For the following sections, simple and combined Rankine cycles are classified, explained, and thermodynamically investigated in detail.

4.8.3.2

Organic Rankine Cycles

Unlike the widely used steam Rankine cycles, an ORC employs organic substances as the working fluids, such as mixtures of hydrocarbons, commonly used refrigerants, ammonia, silicon oil, and pentane [7]. ORC technology is more suitable for the small-scale applications, renewable energy systems, and relatively low-temperature waste heat recovery (WHR) applications. Lowtemperature sources for the ORC systems can be exemplified, such as solar irradiance, biomass combustion systems, heat recovery systems from engine exhaust gases, geothermal energy, and ocean thermal energy [1]. The main difference between the steam and ORCs is the type of working fluid used and hence the cycle’s operating range. The paths of the processes are similar for both systems. For the following sections, various ORCs with various working fluids are explained and thermodynamically investigated in detail.

4.8.4

Thermodynamic Analysis of Vapor Power Cycles

In this section simple, reheat, regenerative, and combined Rankine cycles are explained and thermodynamically examined. Ideal and actual systems are comparatively analyzed to see the effects of irreversibilities associated with heat losses, pressure drops, and HEX effectiveness. Moreover, various working fluids are presented for the ORC applications to show the working range of those systems.

4.8.4.1

The Rankine Cycle

Commercial power plants with steam power have been developing since the 18th century. The steam Rankine cycles are the most widely used thermodynamic cycle for power production among the all other thermodynamic cycles [1]. A conventional power plant operates on a simple Rankine cycle consisting of a pump, a steam generator (boiler), a steam turbine, and a condenser. The simple Rankine cycle is the most basic steam cycle, which consists of four components and processes as presented in Figs. 4 and 5.

270

Steam and Organic Rankine Cycles

3 Qin

WST

Steam generator

Steam turbine

4

Condenser 2 Pump

1 Qout

WP Fig. 5 A simple layout of the Rankine cycle.

Initially, the water is pressurized by a pump and reaches to the boiler pressure (process 1–2). Then there is heat addition at constant pressure in the boiler. The working fluid is heated until it entirely changes its phase from water to gas (process 2–3b) and then is superheated (process 3b–3). The superheated steam generates useful work when it expands in the steam turbine (process 3–4) and afterward, a condensation process takes place until the working fluid reach saturated liquid state (process 4–1). In the ideal Rankine cycle, pump and turbine should operate isentropically (reversible adiabatic) to ensure no heat losses and irreversibility. The heat transfer surface of the condenser should be infinite (or the heat transfer coefficient of the component should be so large to keep the temperature difference between working fluid and the heat sink zero). Furthermore, pressure drops in the boiler and condenser should be zero. Even though the ideal Rankine cycle is internally reversible, due to the temperature change in the heat addition process it has external irreversibilities. Since the cycle is not totally reversible, the exergetic efficiency of a Rankine cycle must be determined smaller than 1 [1]. Mass balance equations (MBE) for pressurization, heat addition, expansion, and condensation processes, respectively: _ 2; m _2¼m _ 3; m _3¼m _ 4; m _4 ¼m _1 MBE: m _1 ¼m

ð1Þ

_ indicates the mass flow rate and the numbers used as indices show the state number in Fig. 4. Since the system is closed and here m steady, there is no external mass addition or rejection into the system. So all the processes have the same masses. Energy balance equations for pressurization, heat addition, expansion, condensation, and overall system, respectively: EBE: h1 þ wp ¼ h2 ; h2 þ qsr ¼ h3 ; h3 ¼ h4 þ wt h4 ¼ h1 þ qc ; wp þ qsr ¼ wt þ qc

ð2Þ

Here, h, w, and q represent the specific enthalpy, specific work, and specific heat values, and p, sr, c, and t indicate the pump, heat source, condenser, and turbine, respectively. While the pump work and the heat addition to the system are considered as the energy inlet of the cycle, the turbine work and the heat rejection from the condenser are the energy outlet of the cycle. According to the first law of thermodynamics (energy balance), heat and work inlet should be equal to the heat and work exit of the system. Net useful work done by the system can be expressed as the difference between produced work by the turbine and consumed work by compressor wnet ¼ wt

ð3Þ

wp

Heat inlet from the heat source and heat removal to the heat sink can be defined as follows: qsr ¼ h3

h2 ; qsi ¼ qc ¼ h4

ð4Þ

h1

here h3 and h2 are the specific enthalpies of the stream after and before the heat addition at the steam generator, respectively, and h4 and h1 represent the specific enthalpies of the saturated vapor and saturated liquid at the inlet and exit of the condenser, respectively. Entropy balance equations for expansion, condensation, pressurization, heat addition, and overall system, respectively: Z 1 dqsi ; s1 þ sg;p ¼ s2 EnBE: s3 þ sg;t ¼ s4 ; s4 þ sg;si ¼ s1 þ 4 Tsi s2 þ

Z

3 2

dqsr þ sg;sr ¼ s3 ; Tsr

Z

2

3

dqsr þ sg ¼ Tsr

Z

4

1

dqsi Tsi

ð5Þ

Here, s for specific entropy and subindex “g” represents entropy generation. As the pressurization at the pump and expansion through the steam turbine processes in the ideal Rankine cycle are isentropic, entropy generation for those processes equals to zero (sg,p ¼sg,t ¼ 0 kJ/kg/K). On the other hand, heat transfer processes at both the heat sink and heat sources are internally reversible but externally irreversible for the ideal cycle. In other words, heat transfer from the system boundaries to the heat sink and heat source takes place. Therefore, entropy production and exergy destruction occur in the cycle.

Steam and Organic Rankine Cycles

Exergy balance equations for pressurization, heat addition, expansion, condensation, and overall system, respectively:  Z 3 T0 1 ExBE: ex1 þ wp ¼ ex 2 þ ex d;p ; ex 2 þ dqsr ¼ ex 3 þ exd;sr Tsr 2  Z 1 T0 ex 3 ¼ ex 4 þ wt þ exd;t ; ex4 ¼ ex1 þ 1 dqsi þ ex d;si Tsi 4   Z 3 Z 1 T0 T0 wp þ 1 1 dqsr ¼ wt þ dqsi þ ex d Tsr Tsi 2 4

271

ð6Þ

here ex indicates the specific exergy rate; T0 is reference temperature, and d represents the exergy destruction. In the ideal Rankine cycle, exergy destruction for pressurization at the pump and expansion at the steam turbine processes are also zero as those processes are adiabatic and reversible (exd,c ¼ exd,t ¼0 kJ/kg). It should be noted that, due to the infinite heat transfer area approach in the thermally isolated steam generator and condenser, in the ideal Rankine cycle the temperature of the heat sink Tsi is equal to the condensation temperature (T2 and T4). Furthermore, the temperature of the heat source Tsr is considered as equal to the highest temperature of the stream (T3). The thermal efficiency of the Rankine cycle is the ratio between useful network generated by the system and total energy input to the cycle. It can be shown as follows: wnet qsi ¼1 ð7Þ Z¼ qsr qsr The overall energy efficiency can also be interpreted as the ratio between the enclosed area by the cycle of the T–s diagram and the area under the heat-addition process [6]. The exergy efficiency of the system can be defined either by the ratio between the net delivered exergy and the actual consumed exergy [1] or the ratio between the real work to the reversible work [6] as follows:   _ actual Ex useful Ex dov W ¼1 ; c2 ¼ ð8Þ c¼ _ rev Exinput Ex input W The exergy efficiency of the cycle can be expressed for the cases in which the heat source has a constant temperature: c¼

1

_ W
net T0 Tsr

_ sr Q

¼

Z ZC

ð9Þ

where ZC indicates the Carnot factor ZC ¼ 1 TTsr0 : Back work ratio (BWR) is another performance parameter of the system, which shows the fraction of the produced work by the steam turbine that is consumed by the pump wp BWR ¼ ð10Þ wt Pressure ratio (PR) and expansion ratio (ER) are also definitions to assess the performance of the system components. While PR is defined as the absolute outlet pressure of the pump divided by the absolute inlet pressure, ER shows the ratio quantifies the expansion of the flow in the turbine. PR ¼

P2 v4 ; ER ¼ P1 v3

ð11Þ

These definitions on the Rankine cycle are aimed to be illustrated in the following sample example. Example 1: Assume an ideal Rankine cycle as in Fig. 5. First, the saturated water at 251C is pumped to the boiler pressure and boiling occurs at Tb ¼ 2501C and it is superheated until the temperature of the stream reaches 5001C. Through the steam turbine, the stream is expanded to the condensation pressure. By considering the reference temperature is equal to the reference temperature, calculate the following:

• • • • •

thermodynamic properties of the working fluid for each state; system overall energy efficiency; system overall exergy efficiency; PR, ER, and BWR; and find the PR value, which maximizes the network for given conditions.

Solution: In order to complete the calculations, we need to make some assumptions as follows:

• • • • •

the system operates on steady-state and steady flow conditions; pump and GT work adiabatically; no pressure drop occurs through the HEXs; kinetic and potential energy effects are negligible; and reference conditions T0 ¼ 251C, P0 ¼ 100 kPa.

272

Steam and Organic Rankine Cycles

The properties of the working fluid at the first state point, which is the exit state of the condenser and inlet state for the pump, can be calculated as follows: Saturated water at 251C. T1 ¼ 298.2K, P1 ¼ 3.169 kPa. Since the two properties of the fluid are known, rest of the thermodynamic quantities can be found in tabular steam data in many thermodynamic textbooks. s1 ¼ 0.367 (kJ/kg K); h1 ¼ 104.8 (kJ/kg); v1 ¼0.001003 (m3/kg). The quality of the stream equals to x¼ 0 as it is saturated water. Initial conditions equal to conditions at State 1, therefore ex1 ¼ 0 ðkJ=kgÞ The second state point, which is after the pressurization can be determined as follows: P2 ¼ P3a ¼ P3b ¼ P3 ¼ boiling pressure of the water at 2501C P2 ¼ 3974 kPa Since the pump works under isentropic conditions s2 ¼ s1 ¼0.367 (kJ/kg K). Moreover, specific work done to operate the pump can be calculated as wp ¼ v1 ðP2

P1 Þ ¼ h2

h1 ¼ 3:979ðkJ=kgÞ-h2 ¼ 108:7 ðkJ=kgÞ

Specific exergy ex 2 ¼ ðh2 ex2 ¼ ð108:7ðkJ=kgÞ

104:8ðkJ=kgÞ

h0

T0  ðs2

s0 ÞÞ-ex2

298:2ðKÞ  ð0:367ðkJ=kg KÞ

0:367ðkJ=kg KÞÞ ¼ 3:979ðkJ=kgÞ

State 3a is the state point where the working fluid is heated at constant pressure and it turns into a saturated liquid. Temperature of the stream reaches 2501C after heat addition T3a ¼ 523.2K. Heat addition process occurs at constant pressure P3a ¼ P2 ¼ 3974 kPa. At this state point the stream is saturated liquid at 3974 kPa hence x3a ¼ 0. Water at 3974 kPa and x¼ 0, s3a ¼ 2.793 (kJ/kg K), h3a ¼ 1085 (kJ/kg), v3a ¼ 0.001251 (m3/kg). Heat addition at unit basis   qsr;3a ¼ cp T3;a T2 ¼ h3a h2 ¼ 976:6 ðkJ=kgÞ Specific exergy

ex 3 ¼ ðh3a ex 3 ¼ ð1085 ðkj=kgÞ

104:8 ðkJ=kgÞ

h0

T0  ðs3a

s0 ÞÞ-ex3a

298:2 ðKÞ  ð2:793 ðkJ=kg KÞ

0:367 ðkJ=kg KÞÞ ¼ 257:3 ðkJ=kgÞ

State 3b shows the state point where the working fluid changes its phase entirely in the boiler and becomes saturated vapor. Temperature of the stream is still 2501C, since the phase change occurs at a fixed pressure T3b ¼523.2K and P3a ¼ P3b ¼ 3974 kPa. At this state point the stream is saturated vapor at 3974 kPa, hence x3b ¼ 1. Water at 3974 kPa and x¼ 1 s3b ¼ 6.072 (kJ/kg K), h3b ¼ 2801 (kJ/kg), v3b ¼ 0.05011 (m3/kg). Heat addition at unit basis qsr;3b ¼ hfg;water@3974 kPa ¼ h3b

h3a ¼ 2801 ðkJ=kgÞ

Specific exergy ex 3b ¼ ðh3b ex 3b ¼ ð2801 ðkj=kgÞ

104:8 ðkJ=kgÞ

h0

T0  ðs3b

s0 ÞÞ-ex3b

298:2 ðKÞ  ð6:072 ðkJ=kg KÞ

0:367 ðkj=kg KÞÞ ¼ 995:1 ðkJ=kgÞ

State 3 represents the state point where the heat addition at constant pressure ends in the boiler and the working fluid reaches the turbine inlet conditions as the superheated vapor. Temperature of the superheated stream reaches 5001C at a fixed pressure T3 ¼ 773.2 K and P3b ¼ P3 ¼3974 kPa. Water at 3974 kPa and T3 ¼ 773.2K, s3 ¼7.094 (kJ/kg K), h3 ¼ 3446 (kJ/kg) v3b ¼ 0:08701 ðkJ=kgÞ

Heat addition at unit basis qsr;3 ¼ cp ðT3

T3b Þ ¼ h3

h3b ¼ 976:6 ðkJ=kgÞ

Specific exergy ex 3 ¼ ðh3 ex 3 ¼ ð3446 ðkj=kgÞ

104:8 ðkJ=kgÞ

h0

T0 ðs3

s0 ÞÞ-ex3

298:2 ðKÞ  ð7:094 ðkJ=kg KÞ

0:367 ðkj=kg KÞÞ ¼ 1335 ðkJ=kgÞ

Steam and Organic Rankine Cycles

273

Overall specific heat addition from the boiler qsr ¼ qsr;3a þ qsr;3b þ qsr;3 ¼ 3337 ðkJ=kgÞ The fourth state point represents the working fluid’s state after the isentropic expansion of steam in the turbine. The stream expands from the maximum pressure to the condensation pressure P1 P4 ¼P1 ¼ 3.169 kPa Since the specific entropy of the stream remains constant at isentropic processes s4 ¼ s3 ¼ 7:094ðkJ=kg KÞ The stream in this state point should be vapor–liquid mixture at T4 ¼ 298.2K. Quality of the vapor–liquid mixture x4 ¼ s4sfgsf ¼ 0:8214 this calculation step also proves the statement since 0oxo1 h4 ¼ hf þ x4 hfg ¼ 2110 ðkJ=kgÞ

3

Similarly v4 ¼ vf þ x4vfg ¼35.61 (m /kg) Specific exergy ex 4 ¼ ðh4 ex 4 ¼ ð2110 ðkj=kgÞ

104:8 ðkJ=kgÞ

T0  ðs4

h0

s0 ÞÞ-ex4

298:2 ðKÞ  ð7:094 ðkJ=kg KÞ

0:367 ðkj=kg KÞÞ ¼ 0:0001435 ðkJ=kgÞ

The useful specific work output from the steam turbine wt ¼ h3

h4 ¼ 1335 ðkJ=kgÞ

qc ¼ h4

h1 ¼ 2006 ðkJ=kgÞ

Heat rejection from condenser

•

The overall efficiency of the system can be defined as Z¼

•

wnet ¼1 qsr

Overall exergy efficiency c¼1

•

qsi qsr qsi 3337ðkJ=kgÞ 2006ðkJ=kgÞ ¼ 0:399-%39:9 ¼ ¼ qsr qsr 3337ðkJ=kgÞ Z
 ¼1 T0 T3

1

1

0:399


 ¼ 0:6494-64:94%

298:2 ðKÞ 773:2 ðKÞ

PR, BWR, and ER PR ¼

P2 3974 3:979 v4 35:61 ¼ 1254; BWR ¼ ¼ 0:00298; ER ¼ ¼ 409:3 ¼ ¼ P1 v3 3:169 1335 0:08701

The thermodynamic properties of the stream through the cycle are tabulated in Table 1. As is seen in the table, the pressure of the water increases drastically in the pump. PR quantifies the fraction of highest and lowest pressure of the cycle. In Rankine cycles, PR is usually high as in this example (1254) since the cycle is a closed cycle so the steam can be expanded to the vacuum. It also can be seen that the BWR is very low (0.3%). This is an auspicious feature of Rankine cycles for higher efficiencies. The work required to pressurize water (pump work) is negligible concerning the work produced by expansion of the steam (turbine work). In other thermodynamic cycles using a gas as a working fluid, BWR is usually higher than 30% [1]. The ER, which gives the proportion between the vapor specific volume at turbine exit and the volume at turbine inlet, is found to be 409.3, which means the volume of the expanded stream is more than 400 times greater than the steam volume at the inlet of the steam turbine. The ER is a crucial parameter for steam turbine design as it is an indicator of the type and size of the steam pipes at turbine inlet and outlet. The variations of the system performance parameters, such as BWR, PR, ER, Carnot factor, energy and exergy efficiencies, with the condensation temperature are given in Table 2. BWR witnesses an increase from 0.28% to 0.34% with the increasing Table 1

Thermodynamic properties of the air in Example 1 for each state point

State #

T (K)

P (kPa)

x(

)

0 1 2 3a 3b 3 4

298.2 298.2 298.2 523.2 523.2 773.2 298.2

3.169 3.169 3974 3974 3974 3974 3.169

0 0 Subcooled liquid 0 1 Superheated vapor 0.8214

v (m3/kg)

h (kJ/kg)

s (kJ/kg K)

ex (kJ/kg)

0.001003 0.001003 0.001001 0.001251 0.05011 0.08701 35.61

104.8 104.8 108.7 1085 2801 3446 2110

0.367 0.367 0.367 2.793 6.072 7.094 7.094

– 0 3.979 257.3 995.1 1335 0.0001435

274

Steam and Organic Rankine Cycles

Table 2

The variation of the system performance parameters with the condensation temperature

T1 (K)

BWR (%)

PR (

278.2 282.6 287.1 291.5 296 300.4 304.9 309.3 313.8

0.28 0.29 0.30 0.30 0.31 0.32 0.33 0.33 0.34

4538 3347 2495 1879 1429 1096 848.7 662.5 521.3

Table 3

)

Zc (

)

ER (

0.62 0.61 0.60 0.60 0.59 0.58 0.58 0.57 0.57

)

1395 1056 808 624 486 382 303 242 195

wp (kJ/kg)

wt (kJ/kg)

wn (kJ/kg)

qin (kJ/kg)

qsr (kJ/kg)

Z (%)

c (%)

3.969 3.97 3.971 3.974 3.977 3.981 3.985 3.99 3.995

1400 1370 1340 1310 1281 1251 1222 1194 1165

1396 1366 1336 1306 1277 1247 1218 1190 1161

1909 1921 1932 1944 1954 1965 1975 1986 1995

3306 3287 3268 3250 3231 3212 3194 3175 3157

42.2 41.6 40.9 40.2 39.5 38.8 38.2 37.5 36.8

68.7 68.2 67.8 67.3 66.9 66.4 66.0 65.5 65.0

The variation of the system performance parameters with the superheating temperature

T3 (K)

BWR (%)

PR (

673.2 710.7 748.2 785.7 823.2 860.7 898.2 935.7 973.2

0.33 0.32 0.31 0.29 0.28 0.27 0.26 0.25 0.24

1254 1254 1254 1254 1254 1254 1254 1254 1254

)

Zc (

)

ER (

0.56 0.58 0.60 0.62 0.64 0.65 0.67 0.68 0.69

)

459 438 420 404 390 377 365 354 344

wp (kJ/kg)

wt (kJ/kg)

wn (kJ/kg)

qin (kJ/kg)

qsr (kJ/kg)

Z (%)

c (%)

3.979 3.979 3.979 3.979 3.979 3.979 3.979 3.979 3.979

1199 1249 1301 1353 1407 1463 1519 1578 1638

1195 1245 1297 1349 1403 1459 1515 1574 1634

1910 1948 1983 2017 2048 2079 2108 2136 2163

3105 3193 3280 3366 3451 3537 3623 3710 3797

38.5 39.0 39.5 40.1 40.7 41.2 41.8 42.4 43.0

69.1 67.2 65.7 64.6 63.7 63.1 62.6 62.3 62.0

T Actual Rankine cycle Ideal Rankine cycle

3

qin 3a

3b

2 1

4 qout s

Fig. 6 T–s diagram of the actual Rankine cycle.

temperature from 5 to 401C. This can be explained by either the growth in the pump work or decrease in the turbine work. The pump work is not affected significantly. However, increasing the condensation temperature by 351C reduces the specific turbine work from 1400 to 1165 kJ/kg. As the stream condensates at higher temperatures, the expansion of the turbine cuts off at relatively higher pressures and it results in less useful work output with less ER. Energy and exergy efficiencies are declined by 5.4% and 3.7%, respectively. Table 3 demonstrates the influence of the superheating temperature on the system performance parameters. Increase in the superheating temperature from 500 to 7001C is accompanied by an increase in the energy efficiency and the network output. The energy efficiency increases from 42.2% to 36.8% and network output rises from 1195 to 1634 kJ/kg. While the pump work remains the same, the turbine work increases and it leads to decline in the BWR. The less percentage of the turbine work is consumed to operate the pump with the increasing temperature. The turbine work increases from 1199 to 1638 kJ/kg with the growing temperature.

4.8.4.1.1

Actual Rankine cycle

The actual steam turbine cycles do not take place as idealized in the Rankine cycle. Pressurization and expansion processes are not isentropic, and pressure and friction losses are unavoidable. Moreover, actual pump work is higher than the ideal pump work where ideal steam turbine work is lesser than the actual turbine work (see Fig. 6). So for a real steam turbine, there is no process

Steam and Organic Rankine Cycles

275

existing at constant entropy. Several definitions were developed to evaluate the performance of each element in the system. For instance, the performance of the pump and the steam turbine are assessed by isentropic efficiency. For the steam turbine, isentropic efficiency is the fraction of actual turbine work and isentropic turbine work. Likewise, the isentropic efficiency of the pump is the ratio of the isentropic work required by the pump and the real work needed by the pump [1,6,8]. The performance criteria of the pump and steam turbine can be calculated as follows: Isentropic efficiency of the pump Zc ¼ ðh2s

h1 Þ=ðh2

h1 Þ

ð12Þ

where h2s represents the specific enthalpy of the water at the exit of the pump under isentropic conditions and h2 is the specific enthalpy of it for the actual case. Isentropic efficiency of the steam turbine Zt ¼ ðh3

h4 Þ=ðh3

h4s Þ

ð13Þ

where h4s is the specific enthalpy of the steam–water mixture at the exit of the steam turbine for the cases where the steam expands isentropically and h4 is the specific enthalpy of the same stream for the actual case. Other reasons causing irreversibilities in a cycle can be considered as pressure drops in the condenser and boiler. Due to the friction on bearings between moving parts of the system components also additional losses form up. Possible leakage of the steam through the cycle and air that leaks into the condensation chamber also need to be taken into account as losses when designing a power plant. Also, it should be noted for the actual Rankine cycle water is usually sub-cooled to avoid cavitation on the pump propeller due to rapid vaporization and condensation of the working fluid. Lastly, auxiliary components in the cycle should be considered in the assessment of a power production plant operating on Rankine cycle [6]. Irreversibilities of the turbine and pump due to friction and pressure drops are considered as the internal irreversibilities and experienced by the working fluid flowing through the cycle. Even though they have a significant effect on the cycle, the most dominant irreversibilities are mostly associated with the combustion of the fuel and subsequent heat transfer at elevated temperatures to the working fluid in the steam generator [7]. The exergy destruction values of the components are investigated in the following sections (Table 4). Example 2: Reconsider the Rankine cycle in Example 1. Assume this time pressurization and expansion processes in the pump and turbine occur irreversible adiabatically and their inlet conditions for both steam turbine and pump are the same as in Example 1. Isentropic efficiencies of the turbine and pump are 80% and 85%, respectively. Calculate:

• • • • •

thermodynamic properties of the working fluid for each state; system overall energy efficiency; system overall exergy efficiency; PR, ER, and BWR; and find the mass flow rate of steam required by 50 MW power production plant.

Solution: In order to complete the calculation, we need to make some assumptions as follows:

• • • • • •

the system operates on steady-state and steady flow conditions; pump and GT work adiabatically; no pressure drop occurs through the HEXs; kinetic and potential energy effects are negligible; reference conditions T0 ¼ 251C, P0 ¼ 100 kPa; and State point 1 is identical for both cases in Example 1 and Example 2. The properties of the working fluid in state point 2 can be obtained from actual specific pump work, which can be determined from the definition of the isentropic efficiency of the pump. Zp ¼ ðh2s

Table 4

h1 Þ=ðh2

h1 Þ ¼ wps =wp

Thermodynamic properties of the air in Example 2 for each state point

State #

T (K)

P (kPa)

x(

0 1 2s 2 3a 3b 3 4s 4

298.2 298.2 298.2 298.5 523.2 523.2 773.2 298.2 298.2

3.169 3.169 3974 3974 3974 3974 3974 3.169 3.169

0 0

)

0 1 0.8214 0.9035

v (m3/kg)

h (kJ/kg)

s (kJ/kg K)

ex (kJ/kg)

0.001003 0.001003 0.001001 0.001001 0.001251 0.05011 0.08701 35.61 39.17

104.8 104.8 108.7 109.7 1085 2801 3446 2110 2311

0.367 0.367 0.367 0.3703 2.793 6.072 7.094 7.094 7.765

– 0 3.979 3.979 257.3 995.1 1335 0.000144 0.000158

276

Steam and Organic Rankine Cycles

here wps indicates the pump work under isentropic conditions and wp represents the actual pump work. Zp ¼ 0:85 ¼

3:979 ðkJ=kgÞ -wp ¼ 4:974 ðkJ=kgÞ wc

wp ¼ h2 2h1 ¼ 4:974 ðkJ=kgÞ ¼ h2

104:8 ðkJ=kgÞ-h2 ¼ 109:7 ðkJ=kgÞ

P2 ¼ P2s boiling pressure of the water at 2501C P2 ¼ 3974 kPa. Temperature and specific entropy of the working fluid at State 2 is slightly higher than the ideal state T2s at P2 ¼3974 kPa and h2 ¼ 109.7 kJ/kg-T2 ¼ 298.5K s2 ¼ 0.3703 (kJ/kg K). Specific exergy ex 2 ¼ ðh2 ex2 ¼ ð109:7ðkJ=kgÞ

104:8ðkJ=kgÞ

T0  ðs2

h0

s0 ÞÞ-ex2

298:2ðKÞ  ð0:3703ðkJ=kg KÞ

0:367ðkJ=kg KÞÞ ¼ 3:979 ðkJ=kgÞ

As it is stated in the question turbine inlet conditions, State point 3 is identical for both cases in Example 1 and Example 2. Similar to calculation of the actual pump work, actual turbine work can be determined as follows: Zt ¼ 0:8 ¼ ðh3

h4 Þ=ðh3

h4s Þ ¼ ðwt =wts Þ

here wts is the turbine work under isentropic conditions and wt indicates the actual turbine work. Zt ¼ 0:8 ¼ ð3446 ðkJ=kgÞ

wnet ¼ wt

h4 Þ=ð3446 ðkJ=kgÞ

2110 ðkJ=kgÞÞ ¼ ðwt =wts Þ

wt ¼ 1135 ðkJ=kgÞ

h4 ¼ 2311 ðkJ=kgÞ

wp ¼ 1135 ðkJ=kgÞ

4:974 ðkJ=kgÞ ¼ 1130 ðkJ=kgÞ

P4 ¼ P4s condensation pressure of the water at 251C P2 ¼ 3.169 kPa. Specific entropy of the working fluid at State 4 is slightly higher than the ideal state s4s at P2 ¼ 3.169 kPa and h2 ¼ 2311 kJ/kg-s2 ¼ 7.765 (kJ/kg K). x4 ¼ s4sfgsf ¼ 0:9035 this calculation step also proves the statement since 0oxo1 Specific exergy ex4 ¼ ðh4 ex4 ¼ ð2311 ðkJ=kgÞ

•

104:8ðkJ=kgÞ

Z¼

298:2ðKÞ  ð7:765ðkJ=kg KÞ

0:367ðkJ=kg KÞÞ ¼ 0:0001578 ðkJ=kgÞ

h2 ¼ 3446 ðkJ=kgÞ

109:7ðkJ=kgÞ ¼ 3336 ðkJ=kgÞ

wnet 1130 ðkJ=kgÞ ¼ 0:3388-33:88% ¼ 3336 ðkJ=kgÞ qsr

Overall exergy efficiency c¼1

•

s0 ÞÞ-ex 4

As the energy efficiency is the ratio between network produced by the system to the heat addition from the heat source, heat addition to the system at constant pressure should be calculated. qsr ¼ h3

•

T 0  ðs4

h0

1

Z
 ¼1 T0 T3

1

0:3388
 ¼ 0:5514-%55:14 298:2ðKÞ 773:2ðKÞ

PR, BWR, and ER PR ¼

P2 3974 4:974 v4 39:17 ¼ 1254; BWR ¼ ¼ 0:004382; ER ¼ ¼ 450:2 ¼ ¼ P1 v3 3:169 1335 0:08701

As it is observed from the results with the internal irreversibilities more work is consumed into the cycle. While in the ideal case the BWR is 0.298%, it is 0.438% in the actual cycle, which means 0.14% more of the turbine work is used to operate pumps. Moreover, energy and exergy efficiencies are declined from 39.9% and 64.9% to 33.9% and 55.1%, respectively.

4.8.4.1.2

Ideal reheat Rankine cycle

As stated in the previous section, the network of a Rankine cycle is the difference between turbine and pump work. To improve the network output, the turbine work output should be increased, or the pump work input should be decreased. Increasing steam generator pressure leads to an increase in the turbine work. However, it brings the issues related to moisture. The higher moisture fraction in the turbine may cause malfunctioning. This problem could be overcome by either superheating the vapor to very high temperatures or expanding the steam in two stages and placing a reheater in between them [6].

Steam and Organic Rankine Cycles

Low pressure turbine

High pressure turbine 3

Qin

WHP

277

6

WLP

Steam generator

Qin 2 WP

4

5

Pump Reheater

Condenser 1 Qout Fig. 7 Rankine cycle with reheating.

T 3 qin

qin

5

4 3a

3b

2 1

6 qout s

Fig. 8 T–s diagram of a Rankine cycle with reheating.

In this chapter, increasing the net power production with reheat is investigated. Reheat does not only lead to rising in turbine work, but it also causes a higher temperature value in the turbine inlet. Hence, utilizing reheat with regeneration could improve the system capacity and efficiency altogether [1,6]. Simple layout and T–s diagram of a Rankine cycle with reheating are presented in Figs. 7 and 8, respectively. The unique change in this system is the multistage turbine power output. Additional heat input to the working fluid takes place between the steam turbine stages. As a result, the expansion of the steam in the turbine occurs under pseudoisothermal conditions. Increasing the number of stages makes the expansion approach the isothermal expansion profile. The complete cycle in Fig. 7 consists of six processes. First, water is taken to the pump and pressurized to the boiler pressure. After the pressurization, the water flows through the steam generator that supplies external heat at isobaric conditions. The superheated steam at greater pressures then enters the first steam turbine stage (high-pressure turbine) and generates useful work. Partially, expanded air enters another HEX (reheater) where its temperature reaches the same level as the inlet temperature of the first steam turbine (T5 ¼ T3). Reheating processes is quite similar to reheating of GTs. The reheated superheated steam drives the second turbine (low-pressure turbine) and the air expands till the pressure reaches to the compressor inlet pressure. The final stage is isobaric heat rejection (condensation) in which liquid–vapor mixture rejects the heat at constant pressure and ends up with steam reaching the same temperature as the pump inlet. In brief, the descriptions for ideal standard air reheating regeneration cycle can be defined as follows: Overall turbine work wt;ov ¼ wt;hp þ wt;lp ¼ ðh3

h4 Þ þ ðh5

h6 Þ

ð14Þ

Overall heat inlet qin ¼ qsteam gen þ qreheat ¼ ðh3

4.8.4.1.3

h2 Þ þ ðh5

h4 Þ

ð15Þ

Regenerative Rankine cycle

A careful assessment of the T–s diagram of the Rankine cycle as in Fig. 4 discloses that heat is transferred to the water during process 2–3a at a relatively low temperature, which results in a reduction in the average heat addition temperature and hence the

278

Steam and Organic Rankine Cycles

efficiency of the cycle [6]. Regenerative Rankine cycle remedies this shortcoming by internal regeneration process within the cycle and preheating the water after the pressurization at the pump. Consequently, the regeneration process raises the average heat addition temperature and makes the ideal Rankine cycle approach the Carnot cycle more [1]. A practical regeneration process in a steam power plant is attained by extracting the steam from the turbine at intermediate pressures and condensing it to extract more heat for water preheating (see Fig. 9). Even though this process diminishes some of the turbine work, as it reduces the amount of mass enters the steam turbine, it increases the overall cycle efficiency. The performance of the cycle improves since the average temperature of the heat source goes up. The component in which heat addition to the feed water takes place is called a regenerator. A regenerator or feedwater heater (FWH) could transfer the heat either mixing the two streams (open feedwater systems) or without mixing them (closed feedwater systems) [6]. Both systems are investigated in the following sections.

Steam generator

Qin 4

5

3 Steam turbine WST

Q

2

6

Condenser

Pump 1 WP

Qout

Fig. 9 Regenerative Rankine cycle.

4.8.4.1.3.1 Open feedwater heaters An open (or direct-contact) feedwater heater (OFWH) is mainly a mixing chamber, in which the vapor extracted from the steam turbine blends with the incoming feedwater leaving the pump. In the ideal case, the mixture should leave the heater as a saturated water at the heater pressure. The simple layout of a single-stage regenerative cycle with an open feedwater is given in Fig. 10 and the T–s diagram of it is presented in Fig. 11. For the regenerative Rankine cycle with one OFW, the working fluid passes reversible adiabatically (isentropically) through the steam turbine stages and pumps. The stream flows through the boiler, condenser, and FWH without any pressure drop. However, there are still irreversibilities due to steam–water mixture into the FWH [7]. In this cycle, the saturated water leaving the condenser is sent to the first pump and pressurized similar to the Rankine cycle (process 1–2). Steam generator 5

Qin

Steam turbine

4

WST

WP Pump 2

3

6

Open FWH

2

Pump 1 Condenser 1 WP

Qout

7

Fig. 10 The simple layout of the ideal regenerative Rankine cycle with an open feedwater heater (OFW).

Steam and Organic Rankine Cycles

279

T 5

6

4 3 2 1

7 s

Fig. 11 T–s diagram of the ideal regenerative Rankine cycle with an open feedwater heater (FWH).

Then the stream enters a mixing chamber where it combines with the part of the steam extracted from the steam turbine in an intermediate stage (process 2–3). Ideally, this mixture leaves the chamber as a saturated liquid at turbine extraction pressure. Afterwards, the saturated liquid stream at the extraction pressure is sent to the second pump to be pressurized to the boiler temperature (process 3–4). Into the boiler similar to the simple Rankine cycle the water is superheated till the temperature reaches the turbine inlet temperature (process 4–5). The superheated steam expands into the cycle to the bled pressure (process 5–6). Then a part of the stream leaves the turbine from the intermediate stage and enters the FWH (process 6–3) the rest of it continues to expand through the turbine and develops useful work (process 6–7). The remaining stream enters the condenser where it rejects the heat and returns the initial state conditions of the cycle (process 7–1). The basic thermodynamic definitions, such as heat and work interactions of the regenerator and the turbine per unit mass can be expressed as follows: Overall heat inlet to the system qsr ¼ h5

ð16Þ

h4

The rejected heat through the condenser qsi ¼ ð1

yÞðh7

h1 Þ

where y shows the mass fraction of the extracted steam in the intermediate stage _6þm _7 _5 ¼m m

ð17Þ


 _6 m _5 m

_5 _6 m _7m _6 _7 _6þm _ 7Þ m m m ðm ¼ þ ¼y ¼y -1 ¼ _5 _5 _5 m _5m _5 _5 m m m m

ð18Þ 1

ð19Þ

The turbine work wt ¼ ðh5

h 6 Þ þ ð1

yÞðh6

h7 Þ

ð20Þ

The pump works The first pump wp;1 ¼ ðh2

h1 Þð1

yÞ

ð21Þ

The second pump wp;2 ¼ ðh4

h3 Þ

ð22Þ

The overall energy efficiency of the Rankine cycle improves because of the regeneration process. This is mainly because the regeneration increases the average temperature of the stream before it enters into the steam generator. The cycle efficiency could be increased further by installing more FWHs. Several active large-scale plants today use as many as eight regenerators [6]. However, it should be noted that the optimum number of FWHs should be obtained regarding the economical point of view. The installment of the excessive regenerators brings more capital, operational, and maintenance costs as well. 4.8.4.1.3.2 Closed feedwater heaters Another way to attain regenerative heating is using the extracted heat in a closed HEX system. Closed heaters are mainly shell-andtube-type recuperators where the temperature of the feedwater rises as the extracted steam condenses on the outside of the tubes carrying the feedwater. As the steam and water do not mix into a chamber, two streams can have different pressures. The layout of a basic regenerative Rankine cycle system with a closed FWH is depicted in Fig. 12 and its T–s diagram for the ideal case is presented in Fig. 13. In the ideal regenerative Rankine cycle with a closed feedwater heater, the feedwater is heated to the exit temperature of the extracted steam, which ideally leaves the regenerator as the saturated liquid at the extraction pressure [6]. However, in the real

280

Steam and Organic Rankine Cycles

Steam turbine

Steam generator

WST

6

Qin

7

8

5 Mixing chamber

9

Closed FWH 3

4

Condenser

Pump 1 1

2

WP

Qout

Pump 2 WP Fig. 12 The simple layout of the ideal regenerative Rankine cycle with a closed feedwater heater (FWH).

T 6

54 9 3 2 1

7

8 s

Fig. 13 T–s diagram of the ideal regenerative Rankine cycle with an open feedwater heater (FWH).

cases, the feedwater leaves the regenerator at a lower temperature than the extracted steam temperature since the infinite heat surface area or heat transfer coefficient is practically impossible. The main difference in this configuration compared to the OFWs system is that the streams mix after the heater and for the ideal systems, those streams leave the FWH at the same temperature.

4.8.4.1.4

Ammonia–water Rankine cycle

At low heat source temperatures, it is not appropriate to utilize pure steam as the working fluid for generating power due to its thermodynamic properties. Water boils at 1001C at the atmospheric pressure of 101.25 kPa. However, in order to achieve the boiling of water below 1001C, the operating pressure of the steam power engine needs to be vacuum pressure; that is, below the atmospheric pressure. Technical and thermodynamic challenges hinder the usage of vacuum pressure in steam engines. A zeotropic mixture can be formed by mixing ammonia with water; this increases the boiling point [9]. The zeotropic mixture of ammonia and water is incorporated with the characteristic of varying temperature during the phase change process from the liquid to vapor phase. This characteristic allows obtaining source and sink temperature profiles with an exceptional match. Moreover, matching of source and sink temperature profiles are significant at the low-temperature differential operation of the heat engine. One example of a suitable application of the ammonia–water Rankine cycle is the OTEC. In OTEC, the temperature of the heat sources vary between 25 and 301C, and the temperature of the heat sink is approximately 41C. [9] Some studies are reported on the ammonia–water Rankine cycle development in the literature. These include Desideri and Bidini [10], Roy et al. [11], Wagar et al. [12], and Pouraghaie et al. [13]. Through exergy and energy analyses, and evaluation of the surfaces of HEXs, the studies elucidated that with increasing source inlet temperature and ammonia concentration in the HEX, the evaporation pressure increases. This increase in evaporation pressure increases the thermal efficiency as well as the exergy efficiencies. The configuration of the ammonia–water Rankine cycle can either be basic or with regenerative configuration. This is discussed in the later sections. Furthermore, a positive displacement expander or a turbine can be utilized [9].

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281

The expander has the advantage of allowing expansion even when two fluid phases exist. This provides more flexibility in the power generation cycle. Furthermore, another remarkable characteristic of the ammonia–water cycle includes the flexibility of adjusting the concentration of ammonia during thin later operation of the heat engine. This allows the power generation cycle to adapt to the possible heat sink or heat source temperatures fluctuations. The possibility of varying the ammonia concentration exists from no ammonia usage (0%), as in the pure steam cycles, to complete 100% usage of ammonia in the power cycle. This makes it a pure ammonia Rankine cycle. Hence, the ammonia–water Rankine cycle can be a viable option for utilizing variable sources of heat. This includes natural solar radiation. Wagar et al. [12] have reported that the cycle energy efficiency can reach nearly 30% even when the heat source is at a low temperature of 2501C. A discussion about the different configurations of the ammonia–water Rankine cycle follows [9].

4.8.4.1.5

Kalina cycle

The Kalina cycle is an innovative thermodynamic cycle that can be utilized for the conversion of thermal energy from a comparatively low heat source temperature to mechanical power. Aleksandr Kalina developed the thermodynamic Kalina cycle in the late 1970s and early 1980s [14]. There have been various proposed modifications that can be made to the cycle, depending on the specific application considered. The Kalina cycle is essentially a modified Rankine cycle that utilizes the mixture of two different compounds as the working fluid: water and ammonia. The basic Rankine cycle does not utilize a mixture of two components; however, pure water is utilized. Kalina et al. [15] reported a possible increase of 10% to 20% in the exergy efficiency as compared to the conventional Rankine cycle. The concentration of ammonia in the working fluid provides a property to enhance the thermodynamic reversibility. As the thermodynamic reversibility is increased and the irreversibilities are decreased, higher thermodynamic efficiencies are achieved. Fig. 15 depicts the basic configuration of the Kalina cycle [9]. As can be observed from Fig. 14, the mixture of ammonia and water is passed through the heat recovery unit before reaching the generator. A concentrated vapor of ammonia is formed in the generator at high pressure and temperature. The concentrated vapor is then passed through the superheater, where the temperature of ammonia vapor is increased. After leaving the superheater, the vapor is passed through the turbine. As it expands in the turbine, thermal energy is converted into mechanical. After leaving the turbine, the vapor temperature, as well as pressure, is low and it is passed through the absorber. In the absorber, the weak solution and the strong solution streams are mixed at a low temperature. Hence, a concentrated solution is formed at low temperature, and it is again pumped through the heat recovery unit to the generator and the cycle continues in this manner [9]. The Kalina cycle has a high heat recovery potential as compared to conventional Rankine steam cycles. This was also specified by Corman et al. [16]. Furthermore, Park and Sontag [17] presented a case study showing that the Kalina cycle exergy efficiency is 15% higher as compared to the steam power cycle. In addition, Hettiarachchi [18] demonstrated the viability of utilizing the Kalina cycle with low-temperature heat sources [9].

4.8.4.1.6

Ammonia–water trilateral flash-Rankine cycle

One of the configurations of the ammonia–water flash Rankine cycle is another configuration of this type of power cycle. This kind of cycle can utilize geothermal energy resources for generating power. Furthermore, it can be utilized where the heat at low temperature is provided by only a heat transfer fluid, which only exchanges sensible heat. This cycle allows for the best match of temperature profiles of the heat source and sink. A boiler is not utilized for boiling. Rather, the working fluid is first heated to saturation; then a positive displacement expander is utilized to conduct expansion by flashing into the two-phase region and

Vapor generator

Qin Ammonia vapor

Ammonia vapor WT

Superheater

Heat recovery

Ammonia− water

Weak solution

WP Pump Fig. 14 Basic configuration of a Kalina cycle.

Absorber

282

Steam and Organic Rankine Cycles

NH3−H2O

140

T (°C)

Source/sink 100

60

20 0

100

200

300

400

500

Q (kW) Fig. 15 Cycle diagram of the trilateral ammonia–water flash Rankine cycle. Adapted from Zamfirescu C, Dincer I. Thermodynamic analysis of a novel ammonia–water trilateral Rankine cycle. Thermochim Acta 2008;477:7–15.

generating power. Expanders that can handle two-phase flow are preferred for this operation, Scroll or screw expanders are such expanders that can handle a two-phase flow [9]. The temperature versus heat rate diagram of the trilateral ammonia–water flash Rankine cycle is given in Fig. 15. In the diagram shown, the heat rate represents the rate of heat transfer in the HEXs. The figure is useful to observe the exceptional match between the sink and source temperature profiles. This can be attributed to the high cycle exergy efficiency. Zamfirescu and Dincer [14] have reported an exergy efficiency of 30% at a brine temperature of 1501C for this cycle, whereas, under the same operating conditions the exergy efficiency of the Kalina cycle was reported to be 13% [9].

4.8.4.1.7

Supercritical Rankine cycle with inorganic fluids

One way of avoiding the problem with the pinch point is to increase the pressure of the working fluid above its critical point. At supercritical pressures, the heating process does not pass through a constant-temperature boiling region. In this case, sensible heat is exchanged between the fluids in the HEX. Care needs to be exercised in the selection of the working fluid for supercritical operation, owing to the high pressure [9]. As materials have mechanical limitations, steam Rankine cycles operating under supercritical conditions are gaining attention and are considered as an emerging technology. The supercritical water heater provides an example of this type of application, as has been demonstrated by Boehm et al. [19] and Tsiklauri et al. [20]. The water pressure is increased to the critical pressure (22 MPa) or above the critical pressure and it is then heated to a temperature of over 4001C. This significantly high amount of pressure as well as temperature makes such cycles suitable for large-scale power plants. However, for small power plants, supercritical conditions are not suitable as the power output is low [9]. However, there are two working fluids that are being considered suitable to be utilized in low power supercritical Rankine cycles with low power outputs; these include organic fluids and carbon dioxide. The operating pressures for the supercritical Rankine cycle need to be more than 75 bar. Although operation at such high pressures has various safety concerns, supercritical carbon dioxide has improved heat transfer characteristics. Yamaguchi et al. [21] illustrated a Rankine cycle based on the solar energy resource utilizing supercritical carbon dioxide. The efficiency was reported to be 25%. Schuster et al. [22] elucidated that several different types of organic fluids are being tested as the working fluid of the transcritical or supercritical Rankine cycles [9].

4.8.4.1.8

Combined cycle power plants

Exhaust temperatures of the GTs are significantly higher than the ambient temperature. The turbine expands to the atmospheric pressure in the open Brayton cycle. Hence, it is impossible to expand the working fluid under the ambient conditions. The hightemperature and low-pressure exhaust gas still have great potential, and can be utilized by the integration of the WHR subsystems to the Brayton cycle. The high energy can be partially used again if the expelled gases can be used in a Rankine cycle [1]. Combined cycle power plants (CCPPs) are smart power production alternatives because of their greater energy efficiencies over the simple Rankine or Brayton cycles. In these systems, the rejected low-pressure hot exhaust gases from the GT are sent to a heat recovery steam generation (HRSG) unit where additional fuel is added. There is sufficient oxygen in the stream to sustain a burning process, which leads to the system reaching higher temperatures. Furthermore, hotter gases are utilized with the arrangement of HEXs to generate vapor for a Rankine cycle. One of the most important points in the modeling a combined cycle is the proper usage of the GT waste heat in the Rankine cycle in order to achieve the highest turbine production. Higher efficiencies of CCPPs compared to Brayton and Rankine cycles have made them quite promising for power generation. Based on these advantages and less specific emissions, CCPP have widely been used in the power production applications all around the world. The key design parameters of these systems are PR (r), compressor isentropic efficiency, gas and steam turbines isentropic efficiencies, and inlet temperature of the GT [1]. A simple diagram of an air–steam CCPP is illustrated in Fig. 16. In this schematic, a basic configuration of a gas turbine cycle (GTC) is considered. The discharged hot exhaust has a relatively large flow rate and is directed to the steam generator, which is

Steam and Organic Rankine Cycles

283

Fuel intake Combustion chamber

Compressor 2

3 WGT

WC

Gas turbine 4

1

5 6

7

Heat exchanger

WST Steam turbine 8

Condenser

Pump 9 WP

Q Fig. 16 A simple layout of the air–steam combined cycle power plant (CCPP).

mainly a HEX and helps to recover a part of the waste heat. The only heat addition to the system takes place in the combustion chamber at a constant pressure, and the heat rejection occurs at after the HEX for the Brayton cycle part and condenser at the Rankine cycle subsystem.

4.8.4.1.9

Exergy destructions in Rankine cycle power plants

Performing of the second law analysis of a power cycle is critical to obtain the power output and efficiency of the system. As it is mentioned in the previous sections in the Rankine cycle, there are irreversibilities since the heat addition to the cycle does not completely occur at the constant temperature. These external irreversibilities are unavoidable even for the ideal case. A comprehensive exergy analysis of the cycle diagnoses where the major irreversibilities take place and where to start system enhancements [1,6]. Even each element of the system participates in the overall exergy destruction, the most critical component of a power plant operates on Rankine cycle is the steam turbine. Besides the steam turbine, there are also other elements have majorly responsible for the irreversibilities, such as the steam generator, the pump, and the condenser. Vapor–liquid separators, steam traps, drains, valves, etc. can be considered as relatively minor sources of the exergy destruction. As stated by Dincer and Zamfirescu [1] the common approach in which irreversibilities are taken into account in a power plant analysis can be built by the use of some parameters, as follows:

• • • • • • • • • •

Steam turbine’s isentropic efficiency, which expresses the ratio between the actual turbine work to the reversible turbine work. Pumps’ isentropic efficiency, which expresses the ratio between the reversible pump work to the actual pump work. The mechanical efficiency, which indicates the difference between the net amount of work developed by the steam and the mechanical work developed by the shaft of the steam turbine. The electrical efficiency, which represents the conversion ratio of the steam turbine’s shaft power to the delivered electrical energy delivered to the end user. The temperature difference between the steam generator and the heat source, which is an indicator of the fraction of the extracted exergy from the heat source to the working fluid. Heat losses, which induce exergy losses and cause a reduction of net exergy delivered from the heat source to the working fluid. Temperature difference at condenser, which is an indicator of the difference between rejected exergy by the working fluid and received exergy by the heat sink. Subcooling degree, which is a technical requirement at pump inlet to avoid cavitation due to rapid vaporization. However, subcooling also leads to additional exergy losses. Pressure drop in pipes, which quantifies the exergy losses due to the friction forces between the working fluid and the pipe walls. High-pressure steam leakages, which are inevitable losses since no perfectly sealed turbine exists. Therefore extra energy is required to maintain the working fluid level in the cycle.

284

• •

Steam and Organic Rankine Cycles

Air penetration rates at condenser, which cannot be prevented as the condenser runs in a vacuum. As the air enters into the condenser, pressure increases, which results in a reduction of the turbine work. Furthermore, more energy consumption is needed to deaerate the condenser to maintain the lower pressure. Energy consumption for auxiliary equipment – several parts of auxiliary equipment, such as conveyors, fuel injectors, fans, pumps, blowers, electric motors, and also lighting diminish the net useful power provided to the grid by a power plant.

4.8.4.2

Organic Rankine Cycles

The primary difference between the conventional steam Rankine cycles and the ORC is the usage of different working fluids. In the ORC, the working fluid comprises organic substances, which include various types of refrigerants, mixtures of hydrocarbons, silicon oil, pentane, and ammonia [7]. However, the technology of the ORC is appropriate for small-scale applications. Also, energy systems based on renewable energy resources, as well as for applications involving comparatively low-temperature WHR, ORC is suitable. A few examples of the low-temperature heat sources that can be utilized include solar irradiance, exhaust engine gases, geothermal resources of energy, biomass combustion systems, and ocean thermal energy [1]. As the primary difference between the organic and steam Rankine cycles is the working fluid type, the operating temperatures are also different. However, the process paths for both systems are similar. In the examples and case studies that follow, different ORCs with various working fluids are elucidated and investigated thermodynamically. The operating principles of the ORC resemble that of the steam Rankine cycle. A pump is utilized to pump the working fluid in the liquid phase to a boiler. In the boiler, the working fluid is heated and the state changes to vapor. After leaving the boiler, mechanical work is obtained from the vapor by passing it through an expansion device. As the working fluid exits the expansion device, the temperature and pressure are lowered. The fluid is then passed through a condenser, where it condenses back to the liquid state [9]. In the ORC, water is not utilized as the working fluid; rather an organic fluid is used. The usage of the organic fluid rather than water is due to the lower boiling temperatures of the organic fluid. This is the primary criterion that needs to be considered in the usage of low-grade heat sources. Furthermore, organic fluids have low liquid to vapor volume ratios; this allows the usage of a single stage expansion for converting the thermal energy into mechanical work in the expansion device. On the other hand, the conventional steam Rankine cycles require the usage of a robust turbine, and the working fluid needs to be superheated before entering the turbine to prevent the formation of any condensation droplets during the expansion process. This is essential to avoid any damage to the turbine blades. However, in an ORC, it is not necessary to superheat the working fluid. This is because a compact low-speed expansion device is utilized. Such expansion devices do not make it necessary to superheat the working fluid. Superheating of the working fluid can deteriorate the overall system efficiencies. Hence, this poses an advantage on the usage of low-grade sources of heat [9]. A lab-scale ORC setup can be seen in Fig. 17. The system was built by University of Ontario Institute of Technology (UOIT) researchers investigating a scroll-based ORC focusing on the expansion process [9]. The system is a closed loop structure consisting of an expander, an air-cooled condenser, a vapor generator, and auxiliary components. Unlike the conventional power generation systems, the work output is provided by the scroll expander instead turbine expander. The scroll expander used in the study is a

Manometers

Boiler

Expander

Ampermeter Thermocouples Fig. 17 A lab-scale scroll based organic Rankine cycle (ORC) system. Adapted from Tarique MA. Experimental investigation of scroll based organic Rankine systems [Master’s thesis]. Oshawa, ON: University of Ontario Institute of Technology; 2011.

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modified refrigeration scroll compressor operating in reverse. The mathematical model of the refrigeration scroll compressor as the expander is analyzed by Oralli [23] with respect to rolling angle in order to obtain the appropriate built-in volume ratio, which assures better efficiency. R134a is selected as the working fluid of the system in Fig. 17 [9]. The compressor is a reciprocating type for refrigeration application and is suitable for operating under higher PRs and higher discharge pressures. The heater depicted in Fig. 18 is a radiant electric heater and consists of six heating elements connected in parallel in order to run separately through switches. The heater temperature is set according to the working fluid temperature, which is desired to be maximum 2001C. The operating temperature is selected as relatively lower temperatures as the designed ORC system is considered for driven by renewables or waste heat. The working fluid is condensed by an air cooling based condenser unit with a fan as presented in Fig. 20. As a result of the experimental measurements, the optimum parameters for an ORC cycle operating with an expander– generator unit are determined. The overall cycle energy and exergy efficiencies are obtained 5% and 30%, respectively, for the heat source temperature 1201C [9]. A similar ORC test bench system with a couple of modifications can be seen in Fig. 20 [24]. The setup uses an air duct with a shell-and-tube HEX and a blower instead of a boiler with the radial heater. The air duct is designed and built to fit the blower, used to circulate the heated air through the duct; the heater aims to supply heat in the ORC cycle, and the evaporator to transfer the heat to the working fluid [24] (Fig. 19).

Fig. 18 Inside view of the boiler with a radial heater. Adapted from Tarique MA. Experimental investigation of scroll based organic Rankine systems [Master’s thesis]. Oshawa, ON: University of Ontario Institute of Technology; 2011.

Fig. 19 Condensing unit. Adapted from Tarique MA. Experimental investigation of scroll based organic Rankine systems [Master’s thesis]. Oshawa, ON: University of Ontario Institute of Technology; 2011.

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Steam and Organic Rankine Cycles

Fig. 20 A lab-scale scroll based organic Rankine cycle (ORC) system using hot air as the heat source. Adapted from Hoque SME. Experimental investigation of an R134a based organic rankine cycle. Available from: http://hdl.handle.net/10155/194; 2011.

Liquid injection line Boiler Scroll expander

Hot vapor line

Upper flange

Lower flange

Condenser

Process fluid pump

Liquid receiver

Fig. 21 The computer aided design (CAD) drawing and the photograph of the trigeneration system. Adapted from Khalid F. Development and analysis of new integrated energy systems for sustainable buildings. Available from: http://hdl.handle.net/10155/491; 2014.

The evaporated gas from the evaporator passes to the expander to create motion and generate electric power [23]. The hot air, in this case, can be considered as an exhaust gas of any power cycle. The system can be operated relatively low temperatures. The exergy efficiency of the system in Fig. 20 is found to be 22% at a hot air temperature of 1051C [24]. Another ORC system built at UOIT for multigeneration including useful commodities, such as power, water/space heating, and cooling effect is presented in Fig. 21. The system was developed and analyzed as the part of the PhD study conducted by Tarique [25]. The developed integrated system combines a power and cooling cycle where the source heat is utilized to generate power through a scroll expander and a portion of the heat is used in an ejector cooling system. Similar to the described systems above this setup also using a modified scroll compressor as an expander. Due to its nonazeotropic properties, ammonia–water is selected as the working fluid for both power and cooling cycles. The released heat by the power cycle is captured and utilized for water and space heating. The system integrates a regenerative Rankine cycle with an ejector cooling cycle and it is developed for low scale power production in the range of kilowatts. Moreover, it produces heating of service water of tens of kilowatts and generates 1–2 kW of refrigeration at around 51C. The temperature of the heat source of this system is also set as 100–1601C since it is considered as a renewable energy based system [25]. The power cycle part of the system in Fig. 21 consists of a vapor generator (ammonia desorber), an expander (reverse scroll compressor), a pump, and a condenser (ammonia resorber). The heat is provided to the vapor generator by a heat transfer fluid. After vapor is generated in the ammonia desorber it drives the expander to produce useful energy. The expanded exhaust stream enters to the condenser where it rejects its heat and changes its phase. This process is accompanied by ammonia resorption into liquid. The condensate working fluid is sent to the pump and is pressurized to the vapor generator pressure. The cooling system mainly comprises an ejector, an evaporator coil, and some piping equipment. The high-pressure ammonia–water vapor channeled from the vapor generator flows through the ejector. An evaporator coil is placed to the ejector in the throat part, the other end being linked to the throttling valve provided with liquid from the receiver located at the bottom of the condenser. Steam flows through the ejector develops high velocity in the throat area and drops pressure significantly to induce

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the evaporation of the working fluid in the evaporator. The main vapor stream and the secondary vapor from the ejector, therefore, mix and enter to the diffuser part to increase pressure. Finally, the mixed flows are directed to the condenser and complete the cycle [25]. The residual heat, which is usually released to the environment in this type of power cycle, is captured as much for hot water heating or space heating. Ammonia–water is used for both the power cycle and the cooling cycle. The experimental investigation of this thesis work analyzed the performance of a scroll based heat engine working with low-temperature heat source within the trigeneration facility. Coupling trigeneration and renewable make a robust scheme to supply not only low-carbon electricity and low-carbon heat, but also cooling, in an integrated system with high utilization factor. The first part of the work is to study the performance of the heat engine that consists of the custom built expander, boiler, and condenser in an ammonia–water Rankine cycle system. The second part is to determine the utilization of the heat engine driven by the low grade heat source in a small scale trigeneration unit. Ammonia–water is suitable for low-temperature Rankine power cycle because of its nonazeotropic properties. The source of heat can be at relatively low and intermediate temperatures in the range of 80–2001C. This heat can be derived from a multitude of sources including solar panel collectors, biomass combustion, biofuel, recovered waste heat, etc. Due to the trigeneration feature, the utilization factor of the heat source or fuel energy with this system exceeds 90%. The experimental result shows a maximum isentropic efficiency of 67% and an overall energy efficiency of maximum 7% at 1201C source temperature, while the exergy efficiency is about 30%. The experimental result also shows that the concentration of ammonia is a dominant factor in determining the optimum efficiency with a range of ammonia–water mixture. Forty percent ammonia concentration is found optimum; however, higher concentration drastically reduces the work output. In the trigeneration facility, the cooling and heating can be adjusted without affecting power generation. Finally, it can be concluded that from the optimization results, if mass production is put in place, the system shows a high economic competitiveness with respect to conventional power generation methods, which require multiple individual systems to provide the same results [25]. There are a variety of working fluids that can be used in the ORC. The choice of the fluid highly depends on the operating conditions as well as the temperature difference between the heat sink and source. Nevertheless, the objective of an ORC is to utilize low-grade heat energy. Hence, the organic fluid needs to have the following characteristics [9]:

• • • • • • •

low boiling temperature low freezing temperature high density high latent heat of vaporization stable at high temperatures safe to use low environmental impact

The ORC as the bottoming cycle integrated with an internal combustion engine (ICE) is a model of externally supplied heat engine application. Bombarda et al. [26] investigated the usage of the ORC, as well as the Kalina type of heat engines to be utilized as the bottoming cycle integrated with a diesel ICE. The coupled systems showed improvements in the efficiencies. Miller et al. [27] utilized the ORC as the bottoming cycle with a thermoelectric power generator. Furthermore, Brasz et al. [28] designed a moderate temperature geothermal energy resource coupled with a heat engine. Kane et al. [29] proposed a heat engine based on solar energy resource. Moreover, Saitoh et al. [30] developed solar-driven heat engine that utilized a scroll expander technology. Furthermore, Manolakos et al. [31,32] explicated the advantages of heat engines that are utilized in multigeneration systems for providing other useful products along with power. They investigated a system that comprised of an integrated ORC and a desalination plant. Moreover, Gu et al. [9] investigated an ORC experimentally, as well as theoretically with low and moderate sources of heat. The temperatures of the heat source varied from 60 to 2001C. They found that the power cycle was comparatively insensitive to the temperature of the heat source. However, it was susceptible to the evaporating pressure. In addition, Mago et al. [33] investigated a regenerative ORC. The regenerative system was found to have an enhanced performance as compared to the simple ORC. A renewable energy driven system (solar and biomass) comprises two ORC to provide electricity, heating and cooling of a building can be seen in Fig. 22. The multigeneration system also includes a solar cycle, an open Brayton cycle, and a vapor absorption cycle. The system is proposed and analyzed by Khalid [25]. The cycle is explained by subsystems below.

4.8.4.2.1

Concentrated solar collector

A concentrated solar collector is used to harvest the solar irradiance. At State 33, Duratherm oil enters the collector and flows through it. Afterward, the hot oil flows a HEX to heat the isopentane that goes to the storage tank at State 12 to provide the energy need of ORC 1. The relatively low-temperature oil then enters to the HEX that is utilized to heat stream 23 (isopentane) to stream 19 to preheat the stream to be used in ORC 2. Lastly, the oil returns to the solar collector.

4.8.4.2.2

Organic Rankine cycle

In the developed system, there are two ORCs. The ORC is driven only by solar power, while the second one supported by both solar and biomass, depending on the availability. In ORC 1 isopentane flows through the storage tank in which its temperature increases at high pressures then the working fluid expands in the ORC turbine 1 to generate electricity in this cycle.

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Steam and Organic Rankine Cycles

12 13 ORCT 1 HEX 2

ORC 1

14 17

Air in

Condenser 1 11

18 Hot air out

16 15

33

Biomass HEX 3

26

Pump 1

19 29 HEX 4

28

GT

Combustion chamber

27

32 C

Air

31

30 Generator

20

7 ORCT 2

4

Condenser 3

3 8

ORC 2 23

To community

21 24

Condenser 2 22

Pump 2

Water in

HEX 1 EV 1

5

2

EV 2 25 Hot water out 6

Pump 3

9

1

Chilled water from building

Absorber

Evaporator 1 e

10 f Chilled water to building

Fig. 22 Schematic illustration of a multigeneration system with two organic Rankine cycles (ORCs). HEX, heat exchanger. Adapted from Khalid F. Development and analysis of new integrated energy systems for sustainable buildings. Available from: http://hdl.handle.net/10155/491; 2014.

After leaving the turbine, the stream is sent to a condenser where it rejects heat at constant pressure and provides heating for residential purposes. Afterward, stream 15 is pressurized at the pump and returns to the HEX to be heated by the oil from the solar cycle. Both ORCs work in similar fashion. In addition to the mentioned system elements in ORC 1, an extra HEX is utilized in the second ORC. The heat is supplied by the biomass GTC.

4.8.4.2.3

Gas turbine cycle

In this cycle, the air enters at atmospheric conditions to the compressor and then pressurized air enters the combustion chamber in which biomass is used as the main fuel. After the combustion, the high pressure and temperature gases pass through the GT to produce power. The waste heat of the exhaust is utilized in the HEX to provide heat to ORC 2. After the stream leaves the HEX, it is further utilized to supply heat to the generator in the vapor absorption cycle. 4.8.4.2.3.1 Vapor absorption chiller The heat is transferred from the hot air at state 30 to the boiler of the absorption chiller and the air leaves the generator at state 31. After the heat addition, a portion of the water in the boiler evaporates and flows to the condenser where the steam cools down and condenses by a cooling source, and is then throttled in an expansion valve. Since the water enters the evaporator its temperature declines. Afterward, absorbing the cooling load in the evaporator, the water is evaporated and flows to the absorber, where it mixes with the lean mixture of LiBr–H2O coming from the generator passing through a HEX and expansion valve and turns into a rich

Steam and Organic Rankine Cycles

289

mixture of LiBr–H2O. Then, the mixture from the absorber is sent to the generator through a solution HEX. Moreover, the evaporator in the vapor absorption chiller also supplies a cooling effect for the residential area (see Ref. [25] for details).

4.8.4.2.4

Working fluids for organic Rankine cycle

It is important to consider various factors, while selecting the working fluid of the ORC to obtain heat from low-temperature heat sources. The prime objective is to attain higher cycle efficiency. However, various other factors including safety issues, cost, environmental impact, and availability also need to be considered. The significant parameters that need to be considered are as follows [9]:

• • • • • • •

As the ORC needs to operate with a heat source at low temperature, an organic fluid that has a low boiling point is preferable. Nevertheless, if the boiling temperature is very low at given atmospheric pressure, a low condensation temperature is required. The freezing point of the organic fluid needs to be below the heat sink temperature, at the same time, freezing of the working fluid also needs to be avoided. A working fluid that has a comparatively low specific heat capacity can be used to mitigate the risk of overloading in the condenser. The chemical stability of the fluid is also important. Chemical deterioration or decomposition of the organic fluid may occur at high operating pressure and temperatures. Hence, the selected working fluid should be stable at the operating temperature and pressures. A working fluid that has a high latent heat of vaporization needs to be utilized, as a greater amount of heat can be absorbed during the evaporation. Hence, a fluid that has a high latent heat of vaporization can be chosen. Such fluids will aid in enhancing the recovery of heat and hence will enhance the overall efficiency. The working fluid needs to have low environmental impacts. Thus, the ozone depletion potential (ODP) and global warming potential (GWP) of the selected fluids should be appropriate. The selected fluid should have a low level of toxicity.

Tchanche et al. [34] conducted a study on the possible low-temperature ORC working fluids and has provided 20 suitable fluids; they are listed in Table 5. Furthermore, Fraas [35] elucidated that the expander selection or sizing is highly related to the working fluid properties. Furthermore, Mago et al. [33] conducted studies that showed that as the retrograde working fluid superheating increased, the cycle efficiency decreased [9]. In addition, care needs to be taken, while selecting the working fluid equation of state to be utilized in designing, modeling, and optimizing the ORC. This is attributed to the equation accuracy to predict the fluid properties in the two-phase region and in the vicinity of the two-phase region. Nonlinear gas-dynamic phenomena can occur in the vicinity of the vapor saturation curve during expansion of the complex molecular fluid. Expansion waves and shock formation are examples of such behavior. Table 5

Physical, safety, and environmental data of working fluids

Working fluid

Mol. mass (kg/Kmol)

Tbp (1C)

Tcrt (1C)

Pcrt (MPa)

Safety data ASHRAE

Atm life time (y)

ODP

GWP (100 y)

RC318 R600a R114 R600 R601 R113 Cyclohexane R290 R407c R32 R500 R152a R717 Ethanol Methanol R718 R134a R12 R123 R141b

200.0 58.1 170.9 58.1 72.2 187.4 84.2 44.12 86.2 52.0 99.3 66.1 17.0 46.1 32.0 10.2 102.0 120.9 152.9 116.9

6.0 11.7 3.6 0.5 36.1 47.6 80.7 42.1 43.6 51.7 33.6 24.0 33.3 78.4 64.4 100 26.1 29.8 27.8 32.0

115.2 135 145.7 152 196.5 214.1 280.5 96.68 86.79 78.11 105.5 113.3 132.3 240.8 240.2 374 101 112 183.7 204.2

2.778 3.647 3.289 3.796 3.364 3.439 4.075 4.247 4.597 5.784 4.455 4.520 11.333 6.148 8.104 22.064 4.059 4.114 3.668 4.249

A1 A3 A1 A3 – A1 A3 A3 A1 A2 A1 A2 B2 N/A N/A A1 A1 A1 B1 N/A

3200 0.019 300 0.018 0.01 85 N/A 0.041 N/A 4.9 N/A 1.4 0.01 N/A N/A N/A 14.0 100 1.3 9.3

0 0 1.000 0 0 1.000 N/A 0 0 0 0.738 0 0 N/A N/A 0 0 1.000 0.020 0.120

10,250 B20 10,040 B20 B20 6130 N/A B20 1800 675 8100 124 o1 N/A N/A o1 1430 10,890 77 725

Abbreviations: GWP, global warming potential; N/A, not available; ODP, ozone depletion potential; Pcrt, critical pressure; relative to R11; Tbp, normal boiling point; Tcrt, critical temperature. Source: Adapted from Bertrand TF, Papadakis G, Lambrinos G, Frangoudakis A. 2008. Criteria for working fluids selection in low-temperature solar organic Rankine cycles. In: Proceedings Eurosun 2008, Lisbon, Portugal. p. 1–8.

Steam and Organic Rankine Cycles

290

Zamfirescu et al. [36] illustrated that there exists a special gas-dynamic region comprising of dense gases from some organic working fluids. This occurs at locations where weak shocks may occur in expanding flows. Such occurring processes can affect the expander design as well as the selection of the operating conditions for the process of expansion as explained in Colonna et al. [37]. Example 3: For the system depicted in Fig. 23, saturated liquid at 483K is provided to a geothermal plant with a flow rate of 100 kg/s from a production well. The stream is first flashed by a flash chamber as shown in Fig. 24 to a pressure of 4.5 bar through an isenthalpic flashing process where the flashing chamber is well insulated. The resulting saturated steam is separated from the liquid part in a separator and sent to the steam turbine. The liquid–vapor mixture stream leaves the steam turbine at 0.5 bar with a quality of x¼ 0.9. Then the stream enters the condenser where the stream turns into all saturated liquid water. The liquid water exiting from the separator can be used as the heat source in an ORC with isobutane as the working fluid. The geothermal liquid water exits from the HEX at a temperature of 348K, while the isobutane enters the turbine at 33 bar and 428K and exits at 348K and 3.5 bar. The air-cooled condenser condenses the liquid–vapor mixture and then pressurized the saturated liquid to the boiling pressure. The isentropic efficiency of the pump is given as Zp ¼ 0.85. Calculate: The mass flow rate of isobutane in the ORC. The net power outputs of both the flashing and the binary sections of the plant. The energy and exergy efficiencies of the binary cycle and the overall combined plant. Solution: In order to complete the calculations, we need to make some assumptions as follows:

• • • • •

The system operates on steady-state and steady flow conditions. Pump and GT work adiabatically. No pressure drop occurs through the HEXs. Kinetic and potential energy effects are negligible. Reference conditions T0 ¼ 298K, P0 ¼101.3 kPa.

Analysis: Utilizing the properties of the water for geothermal water, we can obtain the following data: kJ T1 ¼ 483K, X1 ¼ 0, h1 ¼ 897kg Electric generator

Steam turbine

3 Electric generator

Condenser 2

Separator

9 Isobutane turbine

4

8

2

10 Condenser 1

6 Expansion valve

5

7 Pump Heater 11

1

From production well

To reinjection well

Fig. 23 Geothermal power driven combined system in Example 3.

To reinjection well

Steam and Organic Rankine Cycles

291

Condenser 2

32

33

Expansian valve 1

Evaporator 2

34

41

r ye

g He a

Cooling with dehumidification

G

28

Expansian valve 3

16

Heat pump (R134a)

19

17

22

Fan

21

18

Air 20

15 Comp.

Evaporator (HX) 1 3

14 Geothermal water

6

Heat exchanger (HEX) 1

P1

13

12

Evaporator 3

P3

9

Thermal energy storage

11 10

TES

Mixing chamber

7

4

5

1

Production well

d

e at or k p a il Ev m

23

P2

29

27

2

26

tin

36

24

25

Solar Photovoltaic/ thermal We

35

Expansian valve 2

37

M

Absorber

38

Floor heating

P4

40 Heat exchanger 2

Electricity

31

Generator

Turbine

Condenser 4

39

Condenser 3

Organic rankine cycle (Isobutane)

30

ilk

Cooling

Dr

Absorption cooling system (LiBr)

Hot water

8

Hot water storage tank

Reinjection

Fig. 24 The layout of the proposed system. TES, thermal energy storage. Adapted from Bicer Y, Dincer I. Analysis and performance evaluation of a renewable energy based multigeneration system. Energy 2016;94:623–32. kJ ¼ (expansion at constant enthalpy), P2 ¼4.5 bar h2 ¼ h1 ¼ 897kg

x2 ¼

h2

897 623:4 ¼ 0:1293 2120 kg _2¼m _ 1 ¼ 100 m s

hf hfg

¼

The mass flow rate of the isobutene ORC can be determined as follows: kg kg ð1 0:1293Þ ¼ 191:5 s s _ 6  h6 þ m _ 7  h7 ¼ m _ iso  ðh8 h11 Þ m _  ðh6 h7 Þ kg m _ iso ¼ 6 ¼ 51:9 m s ðh8 h11 Þ

_ 2 ð1 _6 ¼m Mass flow rate at State 6 ¼ m

X2 Þ ¼ 100

The net power outputs of both the flashing and the binary sections of the plant can be determined as follows: _ 3  ðh3 w_ st ¼ m

h4 Þ ¼ 4254 kW

_ 8  ðh8 w_ bi ¼ m

h9 Þ ¼ 5304 kW

The energy and exergy efficiencies of the binary cycle and the overall combined plant can be calculated as follows: The binary cycle: WPu ¼

v10  ðP11 Zis

P10 Þ

¼

v10  ðP11 P10 Þ kJ ¼ 6:31 0:85 kg

292

Table 6

Steam and Organic Rankine Cycles

Thermodynamic properties at each state point for the combined plant obtained from the engineering equation solve (EES) software

State

Fluid/phase

m_ (kg/s)

0 1 2 3 4 5 6 7 8 9 10 11

Water/liquid Water/liquid Water/mixture Water/steam Water/mixture Water/liquid Water/liquid Water/liquid Isobutene/superheated vapor Isobutene/superheated vapor Isobutene/liquid Isobutene/liquid

100 100 100 100 100 100 100 100 51.9 51.9 51.9 51.9

E_ x (kW) 0 17,930 16,510 70,410 37,510 1,977 187,600 1,574 8,321 4,605 22,892 2,876

H_ (kW)

P (kPa)

S_ (kW/K)

T (K)

10,480 89,770 89,770 274,400 241,500 34,050 62,340 31,400 40,733 35,428 13,460 13,792

101 1906 450 450 50 50 450 450 3300 350 350 3300

37 243 247 686 686 109 182 102 135 135 123 63

298 483 421 421 354 354 421 348 428 348 307 300

kJ kg _ P ¼ WPu  m _ iso ¼ 0:72  51:9 ¼ 327 kW W kg s _ bi _ net ¼ W W Zbi ¼

_ P ¼ 5304 kW W

327 kW ¼ 4977 kW

w_ net ¼ 0:184  100 ¼ 18:4% _ 8 ðh8 h11 Þ m

The exergy efficiency of the binary cycle can be determined as follows: cbi ¼

w_ net ¼ 0:91  100 ¼ 91% _ 8 ðex8 ex 11 Þ m

The energy efficiency of the overall combined plant can be calculated as follows: Since streams from States 5 and 7 are not used, therefore the total energy and exergy entering the combined power plant can be obtained as follows: Energy inlet to the system: _ 1  h1 Ein ¼ m

_ 5  h5 þ m _ 7  h7 Þ ¼ 58 MW ðm

Exergy inlet to the system: _ 1  ex1 Exin ¼ m

_ 5  ex 5 þ m _ 7  ex 7 Þ ¼ 16:3 MW ðm

Energy efficiency of the system: Zsys ¼

w_ st þ w_ net 4:254 MW þ 4:977 MW ¼ 0:159  100 ¼ 15:9% ¼ Ein 58 MW

Exergy efficiency of the system:

csys ¼

4:254 MW þ 4:977 MW w_ st þ w_ net ¼ ¼ 0:5663  100 ¼ 56:63% Exin 16:3 MW

The thermodynamic properties of each State in the example are tabulated in Table 6.

4.8.5

Case Studies

In this section, to illustrate the industrial combined power plant and integrated steam and organic fluid turbine systems, case studies are presented.

4.8.5.1

Case Study 1

Multigeneration systems are getting more popular since they are offering more useful commodities, while increasing the efficiency of the overall system. Bicer and Dincer [38] have integrated solar thermal/photovoltaic (T/PV) systems with geothermal energy as shown in Fig. 24. In their multigeneration system, the main energy sources are set as solar and geothermal energy. The developed system mainly consists of an ORC, a thermal energy storage (TES), a heat pump, a drying system, and an absorption cooling subsystem.

Steam and Organic Rankine Cycles

293

The proposed system by the authors is aiming to serve a local dairy farm area that is sited on a geothermal region. The layout of the integrated system is presented in Fig. 24. Isobutane (methylpropane) is selected as the working fluid of the ORC in which a medium-high temperature geothermal water drives the system by heating up isobutane at Evaporator 1. The temperature of the stream enters the Evaporator 1 is 475.15K and the temperature at the outlet is 423.15K, which is used by HEX 1 to the hot water storage medium. The temperature of the stream enters the turbine is 418K and the temperature of the exhaust is 373.15K. The expelled steam transfers its heat to the absorption cooling subsystem’s generator. The temperature of the exhaust is in the acceptable range of a lithium bromide–water absorption cooling system. The steam turbine is connected to an electric generator to meet the power demand of the dairy farm. The condensers of the proposed plant, except for absorption cooling subsystem, are supplying the cooling effect from the groundwater, which is pumping into the cold water tanks. The underground water’s temperature is around 283K. The exit stream of the Evaporator 2, located in absorption cooling subsystem, is supplying the required cooling effect by the dairy product storage (around 283K). Due to its lower phase change temperature, refrigerant R134a is selected as the working fluid of the heat pump that is used for dairy farm office heating. The TES system, whose temperature is around 331K, supplies the energy for the heat pump. The hot water of the hot water storage tank and the relatively warm water, which is the exit stream of the Evaporator 3, is mixed into a mixing chamber to maintain the temperature of the TES in the higher levels. A portion of the hot water is sent to the hot water storage medium for later use. The temperature of the working fluid (R134a) reaches 282.6K at the exit of the Evaporator 3. The stream passes through Condenser 3 after it leaves the compressor. The temperature of the inlet and exit streams of the Condenser are 348 and 308K, respectively. The rejected heat through Condenser 3 is used for space heating (offices of the dairy farm). The refrigerant R134 a expands in the expansion valve and hence the temperature of the stream drops to 275K. The cold stream passes through the cooling coils in order to dehumidify the air. Afterward, the dehumidified air is used in the drying process. The required air for the drying process is heated by a fan and duct system located under the solar PV/T subsystem. Sixty percent of the moisture of the milk is extracted by the drying air whose temperature is 328K. After the drying process, the evaporated milk becomes ready to be stored for long distance transportation. A combined PV/T system is used in this case study in which the PV part of the system converts the sunlight into the electricity and the thermal collectors installed adjacent to the PV modules recover the waste heat of the PV system. This combined system enhances the efficiency in two ways. First, as the PV modules cool down, the efficiency of it increases and second by installing the thermal collectors it allows the system to benefit from the sun also in thermal energy form. According to the literature, the temperatures of the PV/T system may reach up to 363K [39,40]. The solar thermal subsystem in this case study is operating under 348K and maintaining continuous operation of the heat pump; the produced hot water is stored in the TES system. Renewable sources usually bring intermittency problems. However, the places like dairy farms where the protection of the cold chain of the products is vital requires uninterrupted services. Therefore, coupling the renewables with a TES system offers 24 h operation in this study. The only exception is the drying process as the evaporated milk production is scheduled between 11:00 am and 3:00 pm in the daytime. According to a recent research published by the Government of Alberta, 111,000 kWh energy is consumed by a conventional dairy form consisting of 100 cows [41]. The power need of a dairy farm can be investigated through the subsystems. The following processes can be considered as the main power consumption areas of the farm: cleaning, ventilation, waste management, livestock keeping, milking, farmhouse, lighting, and water pumping [42]. The cooling process of the milk is held all night. About 8 h of cooling is assumed for the cooling time of the milk during the day. The exit stream of the absorption cooling subsystem is utilized in the cooling and then storage processes of the milk. During the daytime, the providers carry the produced milk to the local markets. The ORC supplies the electricity of such processes as waste management, milking, cleaning, illumination, etc. On the other hand, the PV thermal system is supplying the electricity for the fan and circulating pump of the TES. The fan is operated only during the drying process, which is held in the daytime. As Pump 3 is required to be working for 24 h, the pump is driven by the electric power generated by the ORC during the nighttime. The shaft of the turbine runs the generator, which is connected to the grid to work parallel with it. Forty PV modules are utilized in the proposed system that corresponds to 9.5 kW installed solar power [43]. The underground water in the cold-water storage tank with a temperature of 283K provides the cooling water for condensers. For further information regarding the study, Ref. [38] can be further examined.

4.8.5.1.1

Energy and exergy analyses

Energy and exergy analyses are executed for the designed integrated system, in order to give the related information about system performance, efficiency, and exergy destructions. The analysis is performed under the following assumptions as provided below:

• • • • • • • • •

No heat losses through the expansion valves, turbines, and compressors. Ideal gas assumption is made for the air. Reference conditions T0 ¼298.15K, P0 ¼ 100.00 kPa. Steady state and steady flow. No pressure drop through the HEXs. Kinetic and potential energy effects are neglected. No chemical reaction occurs within the system processes. Hence, the chemical exergy is neglected. The psychometric process cooling with dehumidification and simple heating is taking place at isobaric conditions. The back surface temperature of the solar PV/T is considered as 338K [44–46].

294

Steam and Organic Rankine Cycles

The energy balance of the system is applied all the subsystems and components. The general form of the energy balance equation for any element in the system can be defined as follows: X X _ W _ þ _ out hout _ in hin m ð23Þ DE ¼ Q m

_ and W _ show the heat transfer and here DE shows change in overall energy that is zero for the time-independent, steady state. Q _ and h show the mass flow rate and the specific enthalpy of the streams work energy passing through the element boundaries and m of the system working fluid. Assessment of the performance based on exergy analysis is one of the most significant aspects of the design and analysis of any system. The exergy analysis is mainly based on the second law of thermodynamics [6]. For the each state point, the flow exergy can be obtained as follows: exi ¼ hi

h0

T0 ðsi

s0 Þ

ð24Þ

The destructed exergy by the system elements can be determined through the exergy balance equations as follows: X X _ Qi Ex _ Wi þ _ di ¼ Ex _ in exin _ out exout m m Ex

ð25Þ

_ Wi and Ex _ Qi shows the exergy _ di shows the rate of the exergy destruction, which induces in the system component i, and Ex here Ex rate of work and heat transfer, respectively. The indices in and out indicate the flow inlet and exit from the designed system. Here the exergy of work is equal to the energy of work. The subsystems exergy efficiencies can be defined as follows: The exergy efficiency of the ORC is the ratio between the network generated by the turbine and total exergy input: Zex;ORC ¼

_ out;T W _ 2 ex 2 ðm

_ in;P4 W _ 3 ex 3 Þ m

ð26Þ

The energetic coefficient of performance (COP) of the absorption cooling system can be defined as follows: COPen;AC ¼

_ AC Q _ 29 h29 m _ 30 h30 Þ ðm

_ AC shows the absorbed heat by Evaporator 2. here Q The ideal COP can also be determined as follows [47]:    T32 T0 10 þ 273 COPreversible ¼ T32 þ 273 T0 10

ð27Þ

ð28Þ

The exergetic COP of absorption cooling system is defined as the fraction between the energetic COP and reversible COP: COPen;AC ð29Þ COPreversible For the heat pump system, useful commodity is considered as the rejected heat from Condenser 3. So overall energetic and exergetic COPs of the heat pump subsystem can be defined as COPex;AC ¼

COPen;HP ¼

COPex; HP ¼

_ out;Cond3 Q _ in;C W


_ out; Cond3 1 Q _ in; C W

ð30Þ T0 Ts



ð31Þ

By taking into account all the useful outputs and whole inputs, overall energy and exergy efficiencies of the system can be written as follows: P _ out; Cond3 þ Q _ cooling þ Q _ Heating þ Q _ AC þ mp; 26 hp; 26 _ PV _ Pump W _ in; Fan W _ in; C þ m _ out; T þ W _ 5 h5 m _ 6 h6 þ Q W W Zen; overall ¼ _ Solar _ 1 h1 þ Q m ð32Þ

Zex; overall ¼

4.8.5.1.2

_ PV _ out; T þ W W

P

_ Pump W

_ in;Fan W

_ in;C þ m _ QConda þ Ex _ QCooling þ Ex _ QHeating þ Ex _ QAC þ mp; 26 exp; 26 _ 5 ex5 m _ 6 ex 6 þ Ex W Q _ Solar _ 1 ex1 þ Ex m ð33Þ

Results and discussion

The thermodynamic performance assessment of the system is conducted based on energy and exergy efficiencies. The exergy destructions, temperature, pressure, specific enthalpy, and entropy are calculated for all system components. Thermodynamic values are obtained by using the Engineering Equation Solver (EES) software and tabulated in Table 7. Solar PV thermal system is obtained as accomplishing of electricity generation and hot water production sufficiently. The nominal power of the selected PV thermal module is found to be 0.235 kW. Daily exergy destruction values, energy consumption or generation values, exergy efficiency values of some key system elements are tabulated in Table 8.

Steam and Organic Rankine Cycles

Table 7

295

Thermodynamic data for all state points

State no.

Fluid type

T (K)

P (kPa)

m_ (kg/s)

h (kJ/kg)

s (kJ/kg
K)

ex (kJ/kg)

0 00 000 0000 00000 000000 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41

H2 O R134a LiBr-H2O AirH2O Air Methylpropane H2 O H2O H2O H2O H2O H2O H2O H2O H2O H2O H2O H2O R134a R134a R134a R134a R134a H2O H2O Air Air Air Air Air Product-H2O Product-H2O Methylpropane Methylpropane Methylpropane Methylpropane Methylpropane H2 O H2O H2O H2O LiBr-H2O LiBr-H2O LiBr-H2O LiBr-H2O LiBr-H2O LiBr-H2O

298 298 298 298 298 298 473 475 423 358 338 318 323 290 288 319 331 336 279 283 348 308 275 348 333 303 285 288 328 308 298 298 350 418 373 358 348 363 313 278 278 308 311 326 363 333 328

100 100 100 100 100 100 2500 2500 2500 2500 150 150 150 150 230 150 150 230 300 300 950 950 300 150 150 101.3 101.3 105 105 105 101.3 101.3 2000 2000 250 250 250 5.48 5.48 0.8 0.8 0.8 5.48 5.48 5.48 5.48 0.8

– – – – – – 54.0 54.0 18.0 18.0 0.7 0.7 0.4 0.4 0.4 0.4 0.4 0.4 0.6 0.6 0.6 0.6 0.6 0.3 0.3 0.3 0.3 0.3 0.3 0.8 1.4 0.9 72.0 72.0 72.0 72.0 72.0 0.6 0.6 0.6 0.6 6.1 6.1 6.1 5.4 5.5 5.5

105 276 51 51 299 599 853 862 634 358 272 189 209 71 63 193 243 264 133 259 310 101 101 314 251 96 34 40 81 93 – – 396.6 795.5 735.1 704.9 391.4 2669 167.5 40.73 2510 88.12 94.09 124.2 229.1 173.8 164.7

0.367 1.106 0.177 5.793 5.699 2.515 2.329 2.348 1.840 1.133 0.893 0.639 0.704 0.253 0.224 0.652 0.806 0.869 0.553 0.960 1.037 0.371 0.436 1.015 0.831 5.942 5.732 5.741 5.874 5.921 – – 1.620 2.677 2.794 2.711 1.610 8.674 0.5723 0.1475 9.064 0.2075 0.2269 0.3217 0.4861 0.3271 0.2996

– – – – – – 162.9 166.2 89.6 24.7 10.4 2.7 4.2 0.5 0.8 3.0 7.2 9.5 21.7 26.1 54.1 43.6 24.2 15.9 8.0 0.8 1.6 4.4 5.7 4.0 – – 64.5 148.3 53.1 47.5 62.3 86.9 1.434 1.319 88.1 28.32 28.53 30.39 86.2 78.35 77.48

Source: Adapted from Bicer Y, Dincer I. Analysis and performance evaluation of a renewable energy based multigeneration system. Energy 2016;94:623–32.

As seen in the table, the highest exergy destruction is observed in the turbine, which is about 16,697 kWh/day. Due to the greater temperature gradients and the large mass flow in the steam turbine, the highest exergy destruction obtained this component within the system elements. The other main exergy destructions in the system can be listed namely the vapor generator losses, the turbine, pump losses, and condenser losses. The increase of the exergy destruction is closely linked with the growth of the temperature differences through the turbine and evaporator. The exergy destructions of the first evaporator and HEX come after the turbine by the values of 8870.4 and 7749.6 kWh/day, respectively, as presented in Table 8. The maximum exergy destructions of the components highly depend on the operating parameters of the geothermal fluid condenser, turbines, and evaporator in the cycle. The highest exergy efficiency (c) values are obtained for Expansion Valve 1 and 2, as 0.92 and 0.99, respectively. In the ORC, the turbine produces about 1200 kW power with an exergy efficiency of c¼ 0.63. The turbine exergy destruction is 16,697 kWh/day. TES energy and exergy efficiencies

296

Table 8

Steam and Organic Rankine Cycles

Daily overall thermodynamic analysis data of the multigeneration system components

Component/process

Exergy destruction (kWh)

Exergy efficiency (

Absorber Compressor Condenser 2 Condenser 3 Condenser 4 Cooling with dehumidification Drying Evaporator 1 Evaporator 2 Evaporator 3 Expansion Valve 1 Expansion Valve 2 Expansion Valve 3 Fan Generator HEX 1 HEX 2 Hot water storage tank Mixing chamber Pump 1 Pump 2 Pump 3 Pump 4 Simple heating Solar photovoltaic/thermal (PV/T) Thermal energy storage Turbine

2484.00 88.51 71.09 35.33 861.12 8.06 134.18 8870.40 51.14 4.06 0.50 32.02 74.38 5.99 443.28 7749.60 168.82 27.72 3.69 2032.08 235.37 117.98 6988.80 10.98 87.65 5.04 16696.80

0.02 0.55 0.81 0.12 0.36 0.04 0.18 0.82 0.95 0.8 0.92 0.99 0.56 0.51 0.83 0.01 0.31 0.45 0.7 0.37 0.03 0.1 0.13 0.2 0.12 0.66 0.63

)

Power or heat transfer (kWh) 10,804.80 802.32 150,480.00 – – – – – 3232.80 243.84 130.63 8032.80 28,992.00 12.20 196.01 – 13,255.20 – – – 691.20 70.22 20.25 1.19 88.75 225.96/346.56 10.70

Source: Adapted from Bicer Y, Dincer I. Analysis and performance evaluation of a renewable energy based multigeneration system. Energy 2016;94:623–32. Abbreviation: HEX, heat exchanger.

are 0.857 and 0.766, respectively. Solar PV/T system’s exergy efficiency is 0.12, while power conversion efficiency of the PV/T system is calculated to be 0.107. The energy efficiency of solar PV/T system is 0.165. The overall energy and exergy efficiencies of the renewable energy based multigeneration system with the ORC are obtained to be 0.11 and 0.28, respectively. In the literature, it is observed that the energy efficiency values of geothermal-based multigeneration systems usually differs between 10% and 30%, while the respective exergy efficiencies change between 30% and 48%. The exergy efficiency yields are lower compared to literature efficiencies because exergy efficiency of the solar system is low, around 12%. The energetic and exergetic COPs of the absorption cooling system are found to be 0.73 and 0.21, respectively. Energetic COP of the heat pump system is 4.1 and exergetic COP is 0.03. In addition, the energy and exergy efficiencies, respectively, are calculated to be 9% and 42% for the ORC. In order to understand the effect of ambient conditions on subsystem and overall efficiencies, various parametric studies are performed. 4.8.5.1.2.1 Influence of varying ambient pressure and temperature on the subsystems and the overall system efficiencies The influence of ambient temperature on the overall efficiencies of the system was not significant. The parametric study also reveals that surrounding temperature has no influence on the performance of the system. In fact the ambient temperature is vital to be known and included in the thermodynamic analysis of the system as it might enhance or deteriorate the performance of the examined system. The influence of changing the ambient temperature on the main parts included in the system and the total efficiency of the proposed system are presented in Fig. 25. Raising the ambient temperature did not have any influence on the efficiencies that are related to energy analysis, for example, the COP of both absorption cooling system and heat pump system remains constant. Nevertheless, the exergetic COP of absorption cooling cycle augmented considerably from 0.1 to 0.9 when varying the surrounding temperature from 15 to 25˚C. While a reduction in the exergetic COP from 4.5 to 3.7 is detected when varying the surrounding temperature in the same range mentioned previously. The exergetic efficiency of the of the ORC rises until it reaches 60% by varying the surrounding temperature from 15 to 50˚C. Moreover, the exergetic efficiency of the process of cooling with dehumidification rises and the overall exergetic efficiency of the proposed reached a maximum value of 15% and 37%, respectively. The overall exergetic efficiency of the proposed system along with the exergy efficiency of each subsystem displayed a typical aggregating trend by raising the temperature of the surrounding environment; this can be interpreted by the high influence of the surrounding temperature on the amount of exergy of the heat that can be transferred to the absorption generator and HEX 1. Determining the exergy of this transferred thermal energy leads to an increase in the ORC exergetic performance. This trend

Steam and Organic Rankine Cycles

0.125

0.6

ex,ov,sys en,ov,sys

ex,ORC en,ORC

ex,cooling en,cooling

0.5

0.115

0.4

0.110

0.3

0.105

0.2

0.100

0.1

0.095 15

20

25

30 35 Ambient temperature (°C)

40

45

Exergy efficiency (−)

Energy efficiency (−)

0.120

297

0 50

Fig. 25 The effect of ambient temperature on absorption cooling system coefficient of performances (COPs). Adapted from Bicer Y, Dincer I. Analysis and performance evaluation of a renewable energy based multigeneration system. Energy 2016;94:623–32.

0.645

760

740

0.635 720 0.630 700

Exd,T (kW)

Exergy efficiency (−)

0.640

0.625 Exd,T ex,T

0.620 0.615 15

20

25

30 35 40 Ambient temperature (°C)

680

45

660 50

Fig. 26 Variation of turbine exergy destruction rate and exergy efficiency with changing ambient temperature. Adapted from Bicer Y, Dincer I. Analysis and performance evaluation of a renewable energy based multigeneration system. Energy 2016;94:623–32.

behavior exists due to a rise in the surrounding temperature that leads to a greater value of exergy of the required output. Increasing the value of exergy for the obtained system products will result in increasing the system exergy efficiency. Fig. 25 proved that the proposed system will be exergetically efficient when functioning at high ambient temperatures. 4.8.5.1.2.2 Influence of varying the global solar irradiance on solar system and overall efficiencies Fig. 26 demonstrates that the solar irradiance mitigates the energetic efficiency of the solar energy system because increase in the solar irradiance is higher than the increase in the output energy of the solar module. The parametric study is carried out considering a variation in solar irradiance from 600 to 1000 W/m2 to match the recorded solar irradiance in the countries that are situated near the equator. The solar irradiance variation did not have a significant influence on the overall system performance since the obtained power from the PV/T system has less effect on the total outputs of the system. 4.8.5.1.2.3

Variation of exergy destruction rate and exergy efficiency of the overall system and subsystems by changing the ambient temperature Increasing the ambient temperature 351C resulted in an upsurge in the exergy destruction rate of the turbine from 670 to 750 kW with a reduction in the exergy efficiency of the turbine by around 3%. The same behavior is observed in the compressor exergy destruction for the same increase in the ambient temperature but here, the exergetic efficiency is reduced by about 5%. The

298

Steam and Organic Rankine Cycles

exergetic efficiency of the PV/T system is mitigated from 13.7% to 12% by raising the ambient temperature accompanied by an increase in the PV/T system exergy destruction. Exergy destruction rate associated with air heating process is found to be very small, with an approximate value of 0.45 kW. Raising the ambient temperature from 15 to 25˚C mitigates air heating process exergetic efficiency to 21% and the exergetic efficiency of the dryer to 18%, verifying that the process of drying is significantly affected by the change in the ambient temperature. ORC evaporator and turbine exhibited the highest exergy destruction rate because of the vast geothermal mass flow rate entering the system and the high temperature change in the two parts.

4.8.5.1.3

Final remarks

In this study, a new energy system for multigeneration purposes and operating utilizing renewable energies is developed and assessed. The system aims to produce hot domestic water, hot air for residential applications, cooling effect, food drying, and electric power. The electric power generation is provided by the organic Ranking cycle, which is operated using geothermal energy and the PV/T solar system. The cooling effect is provided by the absorption cooling system that is operating using the heat from the stream leaving the turbine. A duct system below the solar PV/T system is utilized to adjust the temperature of the ambient air so that it becomes appropriate for the food drying process. The influences of varying some of the system operating parameters on the main system components and the overall performance of the proposed system are examined, such as changing ambient temperature, pressure, solar irradiance, and mass flow rate values. The study showed that varying the ambient pressure has no significant influence on the overall system performance and ORC efficiency. Raising the surrounding temperature of the proposed system affects positively the exergetic COP of the absorption cooling cycle, and exergetic performance of the ORC, cooling with dehumidification process, and total system performance. Hence, it is recommended to operate this system in locations that have high temperatures due to the decrease in the exergy destruction and the increase in the exergetic performance of the system. A significant influence of mass flow rates of the subsystems on the system performance has been detected. The energetic and exergetic efficiencies of the proposed system are found to be 11% and 28%, respectively. The greatest exergy destruction rate values are found in the evaporator and the turbine in the ORC as they share around 57% of the system overall exergy destruction rate.

4.8.5.2

Case Study 2

In this case study, the system proposed by Demir and Dincer [48] is presented, in which an integrated solar-energy based system for fresh water and electricity production is proposed and thermodynamically analyzed. The suggested system is depicted in Fig. 27. The system consists of a solar thermal field, a Rankine cycle that is driven by solar power, a molten salt thermal storage subsystem, and a multistage flash distillation (MFD) subsystem. In the suggested system, solar tower charges the molten salt (consisting of 59.5% LiCl and 40.5% KCl), which passes through a HEX to generate vapor for the Rankine cycle. A portion of the molten salt directly goes to the hot storage tank after the solar tower supplies the heat addition process. In order to maintain the produced energy at the same level, the molten salt in the hot storage tank compensates the deficient energy when direct normal irradiance (DNI) level is not sufficient. After the daytime, only the molten salt from the storage provides energy to the cycle. The MFD yields the desired amount of fresh water from seawater. The seawater used for the desalination is heated by the saturated steam–water mixture from the steam turbine. Using the output fluid as a heat source for the MFD also eliminates the external device for condensation accordingly. All system elements of the proposed system are analyzed in the EES. The overall energy and exergy efficiencies are calculated for each system element. The capacity of the power generation and fresh water production of the proposed system are also calculated. Moreover, a parametric study is performed to see the effects of ambient conditions. For further information regarding the study, Ref. [48] can be examined.

4.8.5.2.1

Energy and exergy analyses

The analysis of the proposed system is carried out for four subunits, includes solar heliostat field and solar receiver, thermal heat storage system, Rankine cycle, and MFD unit. The calculations of all the system components are described in this section. First, the reflected radiation from the heliostat field to the solar receiver is determined, and the transferred heat from the receiver to the molten salt is calculated for the solar part. Then, the heat losses through the molten salt tanks are calculated during the storage period. The steam generator fed by the hot molten salt tank provides the heat for the Rankine cycle. The balance equations are shown for the all the components and produced electricity is obtained. A mathematical model is applied to obtain the distilled water production from the seawater by utilizing waste heat of the Rankine cycle. The daily average DNI level is selected as 7.43 kWh/m2/day [49] and it is assumed that daylight is available 12 h daily. The design parameters of the system are tabulated in Table 1. Beside the parameters in Table 9 the following assumptions are used for thermodynamic calculations:

• • • •

No chemical reaction takes place in the system. The changes in potential and kinetic energies are negligible. Pump works are insignificant, except the one in Rankine cycle. The heat losses only occur through the surface of the molten salt tank and solar receiver.

The EES is used for thermodynamic calculations. The overall energy and exergy efficiencies of each system element are determined, and H2, freshwater, and electric power generation capacity of the system are then calculated. Furthermore, the fresh

Steam and Organic Rankine Cycles

299

Heliostat field

Receiver

4

3

2

1 Cold TES

`

Hot TES

HEX

5

AC output

Vacuum pump Feed sea water 13

8

9

10

6 Demister

Distillate product

12

Brine pool

Brine pool

Brine pool

7 Pump

11

Fig. 27 Layout of the system integrated system. HEX, heat exchanger; TES, thermal energy storage. Adapted from Demir ME, Dincer I. Development and analysis of a new integrated solar energy system with thermal storage for fresh water and power production. Int J Energy Res 2017. doi:10.1002/er.3846.

water production capability of the system is assessed by the performance ratio. Finally, some parametric studies are conducted via the EES software. 4.8.5.2.1.1 Heliostat field and solar receiver The heat transfer rate between the solar receiver and the molten salt can be calculated as follows [50]:

_r ¼Q _h Q

_ loss ¼ m _ ðh2 Q

_ p;ms ðT2 h1 Þ ¼ mc

T1 Þ

ð34Þ

300

Steam and Organic Rankine Cycles

Table 9

System parameters

Parameters

Value

Compressor and gas turbine isentropic efficiency Atmospheric conditions Mass flow rate of steam Turbine outlet pressure Daily average direct normal irradiance (DNI) level (Iave) Area of the heliostat mirror Total number of heliostats Surface area of the solar receiver Wind speed Salinity of the feed seawater Specific heat of the molten salt (at 5001C) Heat transfer surface of the molten salt tank (at void fraction 50%) Thermal conductivity of the insulation material Insulation material thickness

80% 1 atm, 251C 20 kg/s 1 atm 7.43 kWh/m2/day 11  11 m2 2200 50 m2 5 m/s 42,000 ppm 1.202 kJ/kg K 958 m2 0.045 W/m K 5 cm

Source: Adapted from Demir ME, Dincer I. Development and analysis of a new integrated solar energy system with thermal storage for fresh water and power production. Int J Energy Res 2017. doi:10.1002/er.3846.

_ h and Q _ loss show the heat transfer rate received by the heliostat field and receiver losses by convection and radiation, where Q _ respectively. Qh can be defined as follows [50]: _ h ¼ Zh  Ah  N  I Q

ð35Þ

where Zh represents the efficiency of the heliostat field. Ah and N indicate the reflective area of a single heliostat mirror and the total number (N) of heliostats, respectively. I is the DNI. In this study, since the DNI level varies during the day, the main results are presented on the daily average of the DNI level. The _ loss can be calculated as follows [50]: calculation of the solar receiver’s heat loss rate Q    _ loss ¼ Ar ha ðT2 T0 Þ se T 4 T 4 Q ð36Þ r 0

where ha is the convective heat transfer coefficient of air, s is Stefan–Boltzmann constant, e is the absorber emissivity, and Ar is the surface area of the solar receiver. The convective heat transfer coefficient of the air is obtained based on the empirical formula in the literature [51]  pffiffiffiffi ha ¼ 10:45 va þ 10 va W=m2 K ð37Þ where va indicates the ambient wind speed (m/s).

4.8.5.2.1.2 Molten salt tank The temperature drop (y) of the molten salt due to the heat losses through the storage tank surface is also taken into account in this study. The temperature drop by time can be obtained as follows: dy ¼ UAt y ð38Þ dt where rmf, Vmf, and cmf indicate the density, volume, and specific heat of the molten salt, respectively. t is time, At is the heat transfer surface area, and U is the overall heat transfer coefficient. The overall heat transfer coefficient U (W/m2 K) can be calculated as follows: rms Vms cms

1=U ¼ dins =kins þ 1=ha

ð39Þ

here kins and dins represent the thermal conductivity and thickness of the fiberglass-based insulation material, respectively. After Eq. (38) is integrated with respect to time t, and rewritten by using the relations in Eqs. (5) and (6), the following relation can be obtained [52]: y T ðt Þ To ¼e ¼ Ti To yi

ðUAt t Þ=ðrms cms Vms Þ

ð40Þ

where subscript i indicates the initial values. After the calculation of the temperature drop, heat losses through the molten salt tank can be obtained by using Eq. (38). In this mathematical model, since the volume of molten salt inside the tank is changing continuously the void fraction of the tank selected as an average value (et ¼0.5) and initial temperature of the hot tank is obtained for the temperature corresponding to the average DNI level in a day.

Steam and Organic Rankine Cycles

Table 10

Exergy efficiency definitions of the system components

Component

Exergy losses and destructions

Solar receiver

_ dl;SR ¼ m_ 2 ex2 Ex

Hot molten salt tank

_ dl;HT ¼ m_ 3 ex3 m_ 2 ex2 Ex _ dl;CT ¼ m_ 4 ex4 m_ 1 ex1 Ex _ _ 3 ex3 þ m_ 8 ex8 Ex dl;SG ¼ m _Ex _ 5 ex5 m_ 6 ex6 dl;GT ¼ m Exdl;BH ¼ m_ 6 ex6 þ m_ 9 ex9

Cold molten salt tank Steam generator Steam turbine Brine heater

301


m_ 1 ex1 þ Q_ SR 1

Exergy efficiency T0 Tsun

_ m_ 1 ex 1Þ T Q_ SR 1 T 0 sun _ 3 ex3 m cHT ¼ ðm _ 2 ex2 Þ m_ 1 ex1 cCT ¼ ðm_ 4 ex4 Þ _ 5 ex5 m_ 8 ex8 Þ m cSG ¼ ððm _ 3 ex3 m_ 4 ex4 Þ W_ cST ¼ ðm_ 5 ex5out;ST m_ 6 ex6 Þ _ _ 9 ex9 Þ 11 m cBH ¼ ððmm_116 ex _ 7 ex7 Þ ex6 m

cSR ¼ ðm2 ex
2



m_ 4 ex4 m_ 5 ex5 W_ out;ST m_ 7 ex7 m_ 11 ex11

Source: Adapted from Demir ME, Dincer I. Development and analysis of a new integrated solar energy system with thermal storage for fresh water and power production. Int J Energy Res 2017. doi:10.1002/er.3846.

4.8.5.2.1.3 Multistage flash distillation The multistage flash distillation unit utilized in this study is demonstrated in Fig. 27. The distillation unit consists of 20 tank (n¼20) stages whose operating pressures are gradually declining at each tank to maintain the phase change. The seawater enters the multistage distillation unit from the last stage and preheated as it is passing through the stages. Then the relatively hot stream enters the HEX, which is the evacuated saturated steam of the steam turbine. The heated water flows through the stages of MFD and condensates at lower pressures supplied by vacuum pumps. Then distilled products are discharged and stored in a tank for domestic usage (Table 10). The mathematical model of desalination process is developed by considering the following assumptions [53]:

• • • • • •

The temperature drop of feed water at all stages is assumed constant. The temperature drop of flashing brine through stages is assumed constant. The specific heat of the water at all stages is assumed constant as 4.18 (kJ/kg K). The latent Theat of vaporization of the water assumed constant for all stages and taken as the average of the hfg 20 hfg ¼ hfg at 0 þT . 2 The noncondensable gases have a negligible effect on the heat transfer. The distilled product is salt-free. Considering the assumptions stated above, the mass balance can be defined as follows: _f ¼m _dþm _b m

ð41Þ

_ f, m _ b show the mass flow rates of the feed seawater, distilled product, and rejected brine, respectively. Another _ d and m here m balance equation can be defined for overall salt as follows: _ f ¼ Xb m _b Xf m ð42Þ where Xf and Xb denote the salt concentration of feed seawater and brine, respectively. As the distilled water is assumed salt-free, it is canceled from the equation (Xd ¼ 0). The temperature decrease at each stage can be written as follows: DT ¼ ðT11

Ts12 Þ=20

ð43Þ

here T11 represents the temperature of the top brine and Ts12 represents the temperature of the last stage. A general expression can be defined for the stage temperature “Ti” at stage “i”: Ti ¼ T11 ði  DTÞ ð44Þ The general expression of energy balance at stage “i" can be rewritten as     _ d;i þ m _ b;i cp Ti _ d;iþ1 þ m _ b;iþ1 cp Tiþ1 ¼ m _ f cp ðti m m Ti Tiþ1 ¼ ti tiþ1

tiþ1 Þ

ð45Þ

where ti is the feed seawater’s temperature. This relation refers to the temperature rise of the feed seawater (Dt) and also shows that the temperature rise of the cooling feed seawater in each stage equals to the temperature rise of the brine (DT). In addition, the temperature of flashing vapor and brine is considered to be equal since the temperature difference between them is too small and has an insignificant influence on the energy balance equation. The quantity of the freshwater production per stage can be obtained by using general energy balance for stages. For instance, the amount of flashing vapor in the first stage can be written as _f _ d;1 ¼ ym ð46Þ m where y¼ cpDT/hfg,av. The total amount of the distilled water by the system can be driven by rearranging the equations above as the summation of all stages [53]:   _ f 1 ð1 yÞ20 _d¼m ð47Þ m For further information regarding the used mathematical model in this study, Ref. [53] can be examined.

302

Steam and Organic Rankine Cycles

4.8.5.2.1.4 Overall energy and exergy efficiencies In order to evaluate the performance of all the system components, the exergy destruction and loss rates and exergy efficiencies of the system elements are defined and tabulated in Table 2. Moreover, to evaluate the effectiveness of the MFD unit, the definition of performance ratio is used, which refers to the distilled water production per used steam as follows: m10 Performance ratio ¼ ð48Þ m6 The overall energy and exergy efficiencies of the system can be obtained regarding the main inputs and outputs. While the useful outputs of the system are fresh water and electricity, the only energy input is the solar energy. Energy efficiency can be obtained as _ net;e þ m _ H2 O ðDhÞ W Zov ¼ ð49Þ _ solar Q Unlike the energy efficiency, exergy efficiency takes into account both waste emissions (or external irreversibilities) and internal irreversibilities. Hence, exergy destruction plus exergy losses is used to define exergy efficiency and it needs to be dealt with to increase to improve performance. The overall exergy efficiency of the system can be defined as follows [54]: P _ dl cov ¼ 1 Ex ð50Þ _ _ Ex Qsolar

4.8.5.2.2

Results and discussion

All the thermodynamic calculations of the system elements are conducted by EES software. The main thermodynamic findings of the working fluids at each state point are obtained and tabulated in Table 11. The overall energy and exergy efficiencies for the integrated system are found to be 19.9% and 46.5%, respectively. Moreover, the integrated system’s power and freshwater production capacities are found as 13.5 MW and 46 kg/s, respectively. It is observed that the performance ratio of the distillation process is 1.95. The overall exergy loss and destruction rate of the system is obtained as 44,108 kW. The system elements’ contribution to the overall exergy loss and destruction rate is presented in Fig. 28. The lowest exergy destruction plus loss rates (E_ dl ) are obtained for the storage Table 11

Process data for the integrated system

State #

m (kg)

T (K)

P (kPa)

h (kJ/kg)

s (kJ/kg-K)

E_ x (kW)

1 2 3 4 5 6 7 8 9 10 11 12 13

197 197 197 197 24 24 24 24 809 46 809 763 809

649 922 920 650 600 351 351 352 331 313 346 313 298

101.3 101.3 101.3 101.3 3000 43.6 43.6 3000 101.3 7.4 101.3 7.4 101.3

25.02 353.7 351.1 26.51 3059 2484 326.3 332.9 242.7 167.5 305.4 167.5 104.8

0.038 0.462 0.459 0.040 6.651 7.197 1.051 1.061 0.805 0.572 0.991 0.572 0.367

28,240 68,197 67,844 28,399 25,400 8061 412.7 496.7 5740 65.64 11,799 1094 0

Source: Adapted from Demir ME, Dincer I. Development and analysis of a new integrated solar energy system with thermal storage for fresh water and power production. Int J Energy Res 2017. doi:10.1002/er.3846.

Exergy destruction + loss ratios of the system components (%)

100 90 80 70

Cold storage tank Hot storage tank

60

Brine heater

50

Steam turbine

40

MFD

30

Steam generator Solar receiver

20 10 0

Fig. 28 Exergy destruction plus loss ratios of the components. MFD, multistage flash distillation. Adapted from Demir ME, Dincer I. Development and analysis of a new integrated solar energy system with thermal storage for fresh water and power production. Int J Energy Res 2017. doi:10.1002/er.3846.

Steam and Organic Rankine Cycles

303

tanks where the highest rates are calculated for the solar receiver and steam generator, respectively. The exergy destruction rates of the storage tanks can be explained by their decent insulation, which results in less temperature drop during the storage period. Moreover, the highest exergy destruction plus loss rates can be interpreted by the heat losses through the components and heat transfer rates at higher temperatures. The solar receiver and steam generator together comprise 73% of the overall exergy destruction and losses. The receiver surface has the highest temperatures among all the system components, since the surface of the receiver cannot be blocked or insulated, which causes massive heat losses through the receiver surfaces. These losses can be considered as the inevitable reversibility of the component. Around 20% of the exergy destruction plus loss at that part is caused by the large mass of rejected brine whose temperature (401C) is slightly higher than the environmental conditions (251C). Exergy efficiencies of the integrated system components are also obtained. Due to the high exergy destruction rate of the steam generator, it has the lowest efficiency by 63% and is followed by the solar receiver with 65% exergetic efficiency. Exergy analysis obviously provides more reliable results as it takes into account ambient conditions. In order to interpret the influence of the ambient temperature on the overall system, parametric studies are executed. Furthermore, the effect of wind speed and insulation material thickness of the molten tanks are obtained and discussed in the following section. 4.8.5.2.2.1 Effects of ambient temperature In this case study, the influence of the ambient temperature on the components has also been investigated. Increasing the ambient temperature results in decline in the temperature difference between the components (storage tanks and solar receiver) and the surroundings. Therefore, fewer heat losses are encountered by these elements. A parametric study has been performed to see the effect of increasing temperature on the heat losses through the surface of the solar receiver. The lost heat rate to the ambient is declined from 1056 to 1013 kW with the temperature increase from 280 to 310K. However, 43-kW loss corresponds an insignificant value compared to the overall heat rate reached to the receiver. Less than 0.1% drop is observed in the temperature of the storage tank. The temperature drop on the storage tanks has also been examined during the time. 30oC increase in the ambient temperature leads around 0.1oC less temperature drop in the storage tanks. The effect of increasing ambient temperature on the exergy efficiency of some major components was investigated as well. For instance, solar receiver exergy efficiency decreases from 67.2% to 63.5%, steam turbine from 79.0 to 77.3, steam generator from 66.7% to 60.1%. Since the molten salt tanks are well-insulated, the temperature on the outer surface of the insulation is relatively low and it leads to less heat loss from the surfaces to the ambient for both tanks. Hence, storage tanks show similar characteristics for exergy efficiencies. No significant amount of change is observed within the range of parametric study for both tanks. 4.8.5.2.2.2 Effects of ambient wind speed The wind is another atmospheric parameter as the temperature, which could have a vital importance on thermal systems. The variations of increasing the speed of ambient air with the convective heat transfer coefficient of air and the heat losses through the surface of solar receiver was investigated. Increase in the wind speed leads also to increase in Reynolds number (Re). For the forced convection, the Nusselt number can be considered as a function of the Reynolds number and the Prandtl number (Pr). Hence, heat losses due to convection increase with the increment in the wind speed. The heat loss rate through the receiver surface increases from 935.1 (at 3 m/s wind) to 1269 kW (at 24 m/s wind). The heat losses on both storage tanks with the variation of wind speed follow similar trends. 86 and 49 kWh heat losses are observed for cold and hot storage tanks, respectively, with the rise of the wind speed from 2 to 20 m/s. Since the surface of the hot storage tank has a higher temperature, it is affected by the increase in convective heat transfer coefficient more than the cold storage tank.

4.8.5.2.3

Final remarks

In this case study, an integrated system for electricity and freshwater production is proposed and analyzed thermodynamically using energy and exergy approaches. The thermodynamic quantities, energy, and energy efficiencies for each component are calculated. The performance ratio of the MFD unit is obtained. Furthermore, a comprehensive parametric study is applied to the integrated system to observe the influence of ambient conditions on the system elements. The following concluding remarks can be extracted from this study:

• • • • •

The overall energy and exergy efficiencies for the integrated system are found to be 19.9% and 46.5%, respectively. The electricity generation capacity of the integrated system is obtained as 13.5 MW. The MFD unit daily produces 3958 t fresh water. 1.95 kg fresh water can be produced in the MFD unit by using 1 kg wet steam leaving the steam turbine. Solar receiver and steam generator represent 73% of the overall exergy destruction.

For further studies, optimization, life cycle analysis, and exergoeconomic should be performed based on the methods in the literature [55]. The environmental part of the analysis should include the influence of emissions arising from materials processing and manufacture. The capital and operational costs of the system should be considered with the projected lifetime of the suggested system. Moreover, a multiobjective optimization should be performed to obtain the optimum design parameters, accounting for exergetic, economic, and environmental factors.

304

Steam and Organic Rankine Cycles

4.8.5.3

Case Study 3

To achieve higher turbine inlet temperatures, a supercritical Rankine cycle can be utilized. Al-Zareer et al. [56] proposed an integrated system comprising of a coal gasification unit integrated with the supercritical Rankine cycle. The system utilizes hot syngas exhaust to fuel the Brayton cycle as shown in Fig. 29. The multigeneration integrated system proposed comprises of integrated gasification combined cycle (IGCC) along with carbon dioxide capture and a water gas shift membrane reactor (WGSMR). In addition, a direct splitter of H2S to H2 and sulfur is utilized. The primary inputs to the integrated system include coal feed with water, as well as atmospheric condition air. Water and air are not considered as thermodynamic energy inputs to the system. The schematic of the system is depicted in Figs. 29–31. They highlight major system components including the gasification unit, the WGSMR unit, direct splitting of

Fig. 29 An integrated the super critical Rankine cycle with coal gasification. High temperature water gas shift membrane reactor (HTWGSMR); HPT, high pressure turbine; IPT, intermediate pressure turbine; LPT, low pressure turbine; OFW, open feedwater.

Medium pressure steam turbines High pressure steam turbines C2

Reheat

C8

C4

STEAMG1

C10

STEAMSPL HPC1

HPC2

C7

STEAMG2

IPCYL1

C12 IPCYL2

SPLITTOH

C14

IPCYL3

IPCYL4

C9

C3

C15

SPLIT2 C1

MIX2FLO

C6

C25

C28 SPL6

C25-C C5

SPLIT4

SPLIT3

C35

C67

MIX3

H6

H7

H8

DEAERATO C30

C67-C

C70 C68

C27 C37

MIX-GT C38

C29

C71 C69

C100

C42

C36 SPTO C39

C34

CP C32 TDR C61

Fig. 30 First part of the single reheat supercritical pressure Rankine cycle, showing the high-pressure turbine stages and the intermediate pressure turbine (IPT) stages. HPC, high pressure cycle; IPCYL, intermediate pressure cycle; seperator (SPTO); TDR, turbine driver. Adapted from Al-zareer M, Dincer I, Rosen MA. Development and analysis of an integrated system with direct splitting of hydrogen sulfide for hydrogen production. Int J Hydrogen Energy 2016;41:20036–62.

Steam and Organic Rankine Cycles

C16

C15-V 6PL7

C21-V

SPLIT-8 C19

LPCYL2

PL2

SPLMS

C20-2

SPL

C17-V LPCYL1

C22

SPL2

C20

C18 SPLIT-6

C17

LPCYL3

C45 C17-L

SPLIT-9

C21

LPCYL4

C23

C20-1

C48

C41

C49

C46

H3

C21+51

C21-L

C15-L

C31

MTO-H4

305

H2-OWFH

C51

C50

MX-COND

MX

C52

COND-2

C47 C44

HXC43

HX1

HX4 C56

C60

COND

TO

C54

C58 COND-P2

C57

H5+H3SPL

CC

SPLIT-7 C53

COND-P1

C61

Fig. 31 Second part of the single reheat supercritical pressure Rankine cycle, showing the low-pressure turbine (LPT) stages and the condenser. LPCYL, low pressure cycle; OFWH, open feedwater heater Adapted from Al-zareer M, Dincer I, Rosen MA. Development and analysis of an integrated system with direct splitting of hydrogen sulfide for hydrogen production. Int J Hydrogen Energy 2016;41:20036–62.

Table 12

Simulation model input data

Single reheat Rankine cycle with supercritical water as working fluid High pressure steam turbine (HPST) Total number of turbine stages First stage’s discharging pressure Second stage’s discharging pressure Medium pressure steam turbine (MPST) Total number of turbine stages First stage’s discharging pressure Second stage’s discharging pressure Third stage’s discharging pressure Fourth stage’s discharging pressure Low pressure steam turbine (LPST) Total number of turbine stages First stage’s discharging pressure Second stage’s discharging pressure Third stage’s discharging pressure Fourth stage’s discharging pressure

– – 2 80 50

– – stages bar bar

4 28 11.5 5.2 3.3

bar bar bar bar

4 1 0.43 0.18 0.04

bar bar bar bar

Source: Adapted from Al-zareer M, Dincer I, Rosen MA. Development and analysis of an integrated system with direct splitting of hydrogen sulfide for hydrogen production. Int J Hydrogen Energy 2016;41:20036–62.

H2S into H2 and S unit, cryogenic air separation unit, and single reheat supercritical pressure Rankine (SRSPR) cycle and Brayton cycle. Table 12 lists the primary system inputs along with the primary operating parameters of the proposed integrated system. The products of the combustion pass through the GT. After leaving the GT, they are sent to the HRSG unit, where steam is produced that is passed to the SRSPR cycle. SRSPR system comprises two high pressure cycle (HPC) stages and four intermediate pressure cycle (IPCYL) and low pressure cycle (LPCYL) stages. The finally produced exhaust gases are sent to the CO2 capture unit. The latter system is placed at the end of the proposed integrated system; this is plausible only due to the usage of the membrane reactor and high CO2 concentration, which is produced during the shift reaction and has almost no effect on the reaction rate. This is due to the continuous removal of the produced hydrogen through the highly selective membrane walls. As the main focus of the chapter is vapor powered cycles, we are going to discuss only the supercritical Rankine cycle part of the study. For the further details about system refer to Ref. [56].

4.8.5.3.1

Energy and exergy analyses

A thermodynamic analysis is conducted through energy and exergy approaches. Considering only the energy analysis can be misrepresentative, as the performance of the system relative to the reversible case is not considered. The maximum work that can be received from the system is when the system is reversible. The analysis is conducted on the novel hydrogen and power

306

Steam and Organic Rankine Cycles

producing integrated system to evaluate the rates of exergy destruction, energy efficiencies, and exergy efficiencies of all system components. Various assumptions are made for the analysis that are listed below:

• • •

Steady state operation. Negligible changes in the potential and kinetic energy occur in the system components. All gases are considered as real gases in the Aspen Plus simulation, however, an ideal gas behavior is assumed in the exergy analysis. Other assumptions specific to some system components are also made for the simulation. Assumptions for the steam turbines:

• •

The steam turbines operate adiabatically. The isentropic efficiency of the steam turbines is 72%. The general steady state thermodynamic mass, energy, entropy, and exergy balance equations are written as follows: X X _ out _ in ¼ m m

ð51Þ

_ denotes the mass flow rate, and the subscripts in and out represent flow entering and exiting the system boundary. An where m energy balance for steady-state operation can be expressed as follows: X X _ in þ W _ out þ W _ out þ _ in þ _ _ ¼Q Q mh ð52Þ mh out

in

_ represents the heat rate, W _ denotes the work rate, and h is the specific enthalpy. The steady-state exergy balance equation where Q is written as follows: X X _ _ þ _ d _ wþ _ out ex out þ Ex _ in ex in ¼ Ex Ex ð53Þ m m Q in

out

_ across the component system boundary, it is expressed as _ _ denotes the exergy rate due to Q where Ex Q   _ 1 T0 _ _ ¼Q Ex Q Ts

ð54Þ

_ is the heat transfer rate, Ts is the temperature of the control volume boundary where heat transfer occurs. Also, ex denotes where Q the specific exergy of a flowing fluid, which can be expressed for fluid i at an ambient temperature of T0 as exi ¼ ex ph;i þ exch;i ¼ ðhi

h0 Þ

T0 ðsi

s0 Þ þ exch;i

ð55Þ

where si and hi denote the entropy and enthalpy of the fluid flow at state i, whereas, s0 and h0 are the entropy and enthalpy of the flow at the dead state. Also, the specific chemical exergy is represented as exch, for the flow of a mixture of ideal gases, that can be expressed as X X   xj ex 0ch þ RT0 ex ch;i ¼ ð56Þ xj ln xj where xj denotes the mole fraction of component j in the flow, and ex 0ch represents the standard chemical exergy of the component j in the flow. The efficiencies of the proposed system can be expressed as ZSRSPR ¼

_ HPC1 þ W _ HPC2 þ W _ IPC1 þ W _ IPC2 þ W _ IPC3 þ W _ IPC4 þ W _ LPC1 þ W _ LPC2 þ W _ LPC3 þ W _ LPC4 þ W _ TDR W _ HX P1 Q

_ CONDP1 W

_ CONDP2 W

_ CP W

cSRSPR ¼

_ HPC1 þ Ex _ HPC2 þ Ex _ IPC1 þ Ex _ IPC2 þ Ex _ IPC3 þ Ex _ IPC4 þ Ex _ LPC1 þ Ex _ LPC2 þ Ex _ LPC3 þ Ex _ IPC4 þ Ex _ TDR Ex w w w w w w w w w w w _ HX P1 Ex Q

_ CONDP1 Ex w

_ CONDP2 Ex w

_ CP Ex w

ð57Þ ð58Þ

4.8.5.3.1.1 Effect of steam flow rate in SRSPR on overall efficiencies, SRSPR efficiency and combined cycle efficiencies If the water flow rate in the SRSPR cycle is increased, a decrease in the steam temperature that is input to the first stage of the high pressure steam turbine occurs. As the energy content of the combusted gases that exit the GT is unchanged, even at different steam flow rates in the SRSPR cycle. Hence, as the steam flow rate in the SRSPR is lowered, the energy and exergy efficiencies of the SRSPR decrease. This lowers the efficiencies of the combined cycle as well, since the inlet steam turbine temperature is related directly to the steam turbine efficiency. Hence, as the turbine steam inlet temperature decreases, the energy efficiency of the steam turbine also decreases. When the turbine efficiency is reduced, their work output will also reduce. And hence, based on the efficiency equations, the SRSPR efficiency will reduce if the network reduces. Nevertheless, the change in the overall system efficiency is comparatively small, although the system is operating on a multigeneration system, the prime output is the hydrogen, and hence the hydrogen flow rate has the greatest effect on the overall system (Tables 13 and 14).

Steam and Organic Rankine Cycles

Table 13

307

Properties of state points of water in the single reheat supercritical pressure Rankine cycle

State point

m_ (kg/s)

m_ (t/h)

T (K)

P (kPa)

C-1 C-10 C-100 C-11 C-12 C-13 C-14 C-15 C-15-V C-16 C-17 C-17-V C-18 C-19 C-2 C-20 C-20-1 C-20-2 C-21 C-21 þ 51 C-21-V C-22 C-23 C-25 C-25-C C-27 C-28 C-29 C-3 C-30 C-31 C-32 C-34 C-35 C-36 C-37 C-38 C-39 C-4 C-41 C-42 C-43 C-44 C-45 C-46 C-47 C-48 C-5 C-50 C-51 C-53 C-54 C-56 C-57 C-58 C-6 C-60 C-61 C-67 C-67-C C-68

9.7 7.4 9.7 7.4 6.3 6.3 6.1 6.1 6.1 6.0 6.0 6.0 5.8 5.8 9.7 5.8 0.1 5.7 5.7 0.1 5.6 5.7 5.6 0.6 0.6 2.3 1.1 0.7 9.7 0.4 7.0 9.7 9.7 9.0 9.0 9.0 9.0 0.7 8.8 0.2 0.2 0.2 0.4 0.1 7.0 0.4 6.5 0.9 6.4 0.1 6.4 6.4 6.2 0.8 6.2 8.8 6.2 0.7 1.7 1.7 0.9

34.92 26.64 34.92 26.64 22.68 22.68 21.96 21.96 21.96 21.6 21.6 21.6 20.88 20.88 34.92 20.88 0.36 20.52 20.52 0.36 20.16 20.52 20.16 2.16 2.16 8.28 3.96 2.52 34.92 1.44 25.2 34.92 34.92 32.4 32.4 32.4 32.4 2.52 31.68 0.72 0.72 0.72 1.44 0.36 25.2 1.44 23.4 3.24 23.04 0.36 23.04 23.04 22.32 2.88 22.32 31.68 22.32 2.52 6.12 6.12 3.24

523 849 541 740 740 652 652 606 606 606 502 502 502 433 876 433 433 433 374 301 374 374 302 850 503 503 740 740 689 740 370 439 447 447 484 517 548 447 689 606 652 413 408 502 335 343 321 689 310 374 302 302 335 335 355 632 375 310 537 537 568

290,000.00 2,800.00 29,000.00 1,150.00 1,150.00 520 520 330 330 330 104 104 104 43 29,000.00 43 43 43 18 18 18 18 4 2,800.00 2,800.00 2,800.00 1,150.00 1,150.00 8,000.00 1,150.00 1,470.00 1,060.00 29,000.00 29,000.00 29,000.00 29,000.00 29,000.00 29,000.00 8,000.00 330 520 520 310 104 99 310 40 8,000.00 490 18 4 490 1,470.00 1,470.00 1,470.00 5,000.00 1,470.00 6 5,000.00 5,000.00 8,000.00

E_ x (kW) 9,940 14,370 8,140 12,420 10,560 9,240 8,950 8,270 8,270 8,060 6,560 6,560 6,410 5,420 20,600 5,420 100 5,320 4,470 50 4,390 4,470 3,140 1,170 450 1,790 1,860 1,170 17,430 690 3,930 6,160 6,530 6,060 6,570 7,110 7,680 470 15,860 210 290 120 320 150 3,760 190 3,470 1,570 3,390 70 3,380 3,390 3,320 450 3,390 14,900 3,490 430 1,350 1,350 780 (Continued )

Steam and Organic Rankine Cycles

308

Table 13

Continued

State point

_ (kg/s) m

m_ (t/h)

T (K)

P (kPa)

E_ x (kW)

C-69 C-7 C-70 C-71 C-8 C-9 C-51-C (CC)

0.8 8.0 0.8 0.9 8.0 8.0 0.1

2.88 28.8 2.88 3.24 28.8 28.8 0.36

632 632 537 538 929 850 303

5,000.00 5,000.00 5,000.00 8,000.00 5,000.00 2,800.00 18

1,330 13,580 640 720 17,050 15,530 50

Source: Adapted from Al-zareer M, Dincer I, Rosen MA. Development and analysis of an integrated system with direct splitting of hydrogen sulfide for hydrogen production. Int J Hydrogen Energy 2016;41:20036–62.

Table 14 Exergy efficiency and exergy destruction rate all components in the proposed IGCC system with WGSMR, direct splitting of H2S, and carbon capture Component (name of block on flow sheet)

ccomponent

E_ xd (kW)

High pressure cylinder 1 High pressure cylinder 2 HEX 1 HEX 2 HEX 3 HEX 4 HEX 6 HEX 7 HEX 8 HEX-P1 Intermediate pressure CYL 1 Intermediate pressure cylinder 2 Intermediate pressure cylinder 3 Intermediate pressure cylinder 4 Low pressure cylinder 1 Low pressure cylinder 2 Low pressure cylinder 3 Low pressure cylinder 4 M HEX Mixer 1 Pump 2 Pump 3 Pump1 STG1 STMG2 HEX 8

84 80 25 43 59 53 72 71 72 91 83 82 80 80 77 74 72 69 98 99 52 57 41 67 76 72

0.450 0.190 0.020 0.030 0.040 0.060 0.200 0.210 0.210 1.660 0.250 0.350 0.260 0.140 0.350 0.250 0.240 0.390 8.190 0.040 0.000 0.010 0.010 1.800 0.780 0.220

Abbreviation: HEX, heat exchanger. Source: Adapted from Al-zareer M, Dincer I, Rosen MA. Development and analysis of an integrated system with direct splitting of hydrogen sulfide for hydrogen production. Int J Hydrogen Energy 2016;41:20036–62.

4.8.6

Future Directions

The conventional Rankine cycle as introduced by William John Macquorn Rankine worked with water as its working fluid. However, in 1883 the advent of utilizing an organic fluid in the Rankine cycle began. Since then, several studies and investigations have been conducted on ORCs. Multigeneration systems provide an opportunity to fulfill multiple useful commodities. ORCs incorporated in multigeneration systems are the upcoming technologies, which are being investigated to be implemented in the near future. Various studies have been conducted on multigeneration systems with ORCs. Suleman et al. [57] integrated geothermal and solar energy resources with ORCs for electricity production. Furthermore, the integrated system also comprised of a cooling system based on absorption cooling and a system for drying wet products. The energy efficiency of the system was found to be 54.7%, and the exergy efficiency was obtained as 76.4%. Furthermore, Islam and Dincer [58] proposed a new multigeneration

Steam and Organic Rankine Cycles

309

system based on solar and geothermal sources of energy. ORCs were utilized for power generation, an absorption cooling system was incorporated to provide cooling, a drying process was included to dry wet products, and a heat pump was utilized to provide space heating. The overall system energy and exergy efficiencies were obtained as 51% and 62%. However, for a single generation system, the energy and exergy efficiencies were evaluated as 28% and 54%, respectively. Various types of multigeneration systems incorporating ORCs can be implemented to achieve higher system energy and exergy efficiencies. For instance, Al-Sulaiman et al. [59] proposed and analyzed three different types of multigeneration systems with ORCs. The evaporator of the ORC was provided the required heat from solid oxide fuel cell after burner, biomass combustor and parabolic trough solar collectors. The exhaust stream of the ORC turbine was utilized to provide heating and cooling through an absorption cooling system. The energy and exergy efficiencies of single generation systems were found to increase considerably when trigeneration was utilized. In addition, the CO2 emissions were also found to decrease with trigeneration. The working fluid plays in important role in the performance of ORCs. Due to the low operating temperature, inefficiencies accompanying the heat transfer processes are high. The fluid thermodynamic properties highly influence these inefficiencies. In order to be compatible with the ORC, the working fluid needs to have low boiling points to be able to utilize the low-grade heat. Compatible fluids include chlorofluorocarbons, as well as hydrocarbons, such as isobutane. New organic fluids to be utilized in these power cycles are being investigated. The working fluid needs to have suitable properties to be utilized in these power cycles. The fluid should have high critical temperatures; this allows more heat to be absorbed by the fluid without reaching the critical temperature. Furthermore, fluids with high specific enthalpies need to be investigated. In order to increase the efficiency of ORCs, it is essential to introduce new organic fluids that can absorb a large amount heat from the heat source without reaching the critical temperature and have high specific enthalpies. Moreover, at high temperatures, the organic fluids deteriorate chemically. Hence, the temperature of the heat source also has an upper limit, beyond which the organic fluids might become unstable. The latent heat of vaporization of the organic fluids is also an important fluid property that determines the required flow rates and the specific heat absorbing capacity. In addition, some organic fluids have high ODPs and GWPs. Henceforth, organic fluids with the appropriate properties and low environmental impacts need to be investigated to enhance the overall performance of ORCs.

4.8.7

Concluding Remarks

This chapter provides a comprehensive presentation of steam and ORCs and systems and their energy and exergy analyses with emphasis on irreversibility quantifying through exergy destruction. The discipline of Rankine and ORCs is a result of the accumulation of the experiences and knowledge for centuries. From William Rankine’s invention of the closed thermodynamic cycle to the modern power plants, many enhancements are experienced by technology. The majority of the modern power stations in use today are still based on the steam Rankine cycle. Over 80% of the electricity generation comes from steam turbine driven rotary generators. Moreover, simple structure of the Rankine cycle allows high power generation yet with a low thermal efficiency (around 40%). To overcome this problem, either the average temperature of the working fluid can be increased or the temperature of the heat sink can be decreased. Therefore, the methods as reheating, regeneration, and superheating are developed to enhance the system efficiency. Besides all the advantages of the steam Rankine cycles, its operating range is limited by the thermodynamic properties of water. Therefore, alternative organic working fluids are introduced that have high molecular mass with a boiling point at a lower temperature than the water–steam phase change. These cycles are called ORCs and mainly used for power production from renewables, such as solar, geothermal, or biomass, and WHR applications. Today, geothermal based ORCs contribute the power production by ORC power stations by around 75%. The overall energy efficiency of a power plant can be increased also by using a combined cycle. Various types of CCPP have been proposed that mainly utilize the exhaust stream of the topping cycle. Gas–steam combined cycles are the most popular combined cycle applications. Also, the low-operating temperatures of the ORCs make them a suitable option as the bottoming cycle. Today, it is possible to reach around 60% thermal efficiency with combined cycles. The efficiency and other relevant indicators are presented for each type of steam and ORCs in this chapter. Moreover, examples and case studies are presented to provide illustrative explanations to the topic.

Acknowledgment The authors acknowledge the support provided by the Natural Sciences and Engineering Research Council of Canada.

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Temperature coefficients for PV modules and arrays: measurement methods, difficulties, and results. In: Conf. Rec. Twenty Sixth IEEE Photovolt. Spec. Conf., IEEE; 1997. p. 1183–6. doi:10.1109/PVSC.1997.654300. Aman J, Ting DSK, Henshaw P. Residential solar air conditioning: energy and exergy analyses of an ammonia-water absorption cooling system. Appl Therm Eng 2014;62: 424–432. Demir ME, Dincer I. Development and analysis of a new integrated solar energy system with thermal storage for fresh water and power production. Int J Energy Res 2017; doi:10.1002/er.3846. Torresol Energy. Gemasolar thermosolar plant. Available from: http://www.torresolenergy.com/TORRESOL/gemasolar-plant/en; 2017 [accessed 14.03.17]. Meriche IE, Beghidja A, Boumedjirek M. Energetic and exergetic analysis of solar gas turbine power plant in South Algeria. In: IREC 2014 – 5th international renewable energy congress; 2014. doi:10.1109/IREC.2014.6826902. European Commission. 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[55] Ahmadi P, Dincer I, Rosen MA. Exergy, exergoeconomic and environmental analyses and evolutionary algorithm based multi-objective optimization of combined cycle power plants. Energy 2011;36:5886–98. [56] Al-zareer M, Dincer I, Rosen MA. Development and analysis of an integrated system with direct splitting of hydrogen sulfide for hydrogen production. Int J Hydrogen Energy 2016;41:20036–62. [57] Suleman F, Dincer I, Agelin-Chaab M. Development of an integrated renewable energy system for multigeneration. Energy 2014;78:196–204. [58] Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147:328–43. [59] Al-Sulaiman FA, Hamdullahpur F, Dincer I. Performance comparison of three trigeneration systems using organic rankine cycles. Energy 2011;36:5741–54.

Further Reading Dincer I. Refrigeration systems and applications. 3rd ed. London: John Wiley & Sons, Ltd.; 2017. Dincer I, Rosen MA, Ahmadi P. Optimization of energy systems. London: John Wiley & Sons, Ltd.; 2016. International Energy Agency. Energy efficiency indicators highlights. Paris: IEA; 2016. International Energy Agency. World energy investments 2017. Paris: IEA; 2017. International Energy Agency. World energy outlook 2017. Paris: IEA; 2017. Kanoglu M, Cengel YA, Dincer I. Efficiency evaluation of energy systems. New York, NY: Springer Verlag; 2012. Quoilin S, Van Den Broek M, Declaye S, Dewallef P, Lemort V. Techno-economic survey of organic rankine cycle (ORC) systems. Renew Sustain Energy Rev 2013;22:168–86.

Relevant Websites http://www.alstom.com Alstom Power. https://www.geaviation.com/ GE Aviation. https://powergen.gepower.com/applications.html GE Power. https://www.gepower.com/steam/steam-turbines GE Steam Power. http://www.globalenergyobservatory.org/list.php?db=PowerPlants&type=Gas Global Energy Observatory. https://www.iea.org International Energy Agency. https://global.kawasaki.com/ Kawasaki Heavy Industries. http://turbomachinery.man.eu/ MAN Diesel & Turbo. http://web.mit.edu/16.unified/www/SPRING/propulsion/notes/node27.html MIT Thermodynamics and Propulsion lecture notes. http://www.mhps.com/en/products/thermal_power_plant/ Mitsubishi Hitachi Power Systems. https://www.netl.doe.gov National Energy Technology Laboratory. http://nptel.ac.in/courses/112105123/24 National Program on Technology Enhanced Learning (NPTEL). https://www.siemens.com/global/en/home/products/energy/power-generation/ Siemens. https://energy.gov/ US Department of Energy.

4.9 Combined Energy Conversion Systems Ibrahim Dincer and Murat E Demir, University of Ontario Institute of Technology, Oshawa, ON, Canada r 2018 Elsevier Inc. All rights reserved.

4.9.1 Introduction 4.9.2 Combined Cycle Power Plants 4.9.2.1 Brayton–Rankine Combined Cycle 4.9.2.1.1 Case Study 1 4.9.2.1.1.1 Single steam Rankine cycle 4.9.2.1.1.2 Natural gas fueled combined cycle 4.9.2.1.1.3 Energy and exergy analyses 4.9.2.1.1.4 Results 4.9.2.1.1.5 Closure 4.9.2.1.2 Case Study 2 4.9.2.1.2.1 Closure 4.9.2.2 The Brayton–Kalina Cycle 4.9.2.3 The Brayton–Brayton Cycle 4.9.2.4 The Brayton–Diesel Cycle 4.9.2.5 The Brayton–Stirling Cycle 4.9.2.6 The Brayton–Fuel Cell Cycle 4.9.2.6.1 Case Study 3 4.9.2.6.1.1 System description 4.9.2.6.1.2 Thermodynamic analysis 4.9.2.6.1.3 Compressor 4.9.2.6.1.4 Recuperator 4.9.2.6.1.5 Solid oxide fuel cell 4.9.2.6.1.6 Combustor 4.9.2.6.1.7 Gas turbine 4.9.2.6.1.8 Power turbine 4.9.2.6.1.9 Results and discussion 4.9.2.6.1.10 Final remarks 4.9.2.7 The Chemical Recuperation Cycle 4.9.2.8 The Rankine–Rankine Cycle 4.9.2.8.1 Case Study 4 4.9.2.8.1.1 System description 4.9.2.8.1.2 Analyses 4.9.2.8.1.3 Energy and exergy efficiencies 4.9.2.8.1.4 Results and discussion 4.9.2.8.1.5 Closing remarks 4.9.2.9 Brayton–Organic Rankine Combined Cycle 4.9.2.9.1 Case Study 5 4.9.2.10 Integrated Gasification Combined Cycle 4.9.2.11 Thermoelectric Generator Combined Cycles 4.9.2.11.1 Case Study 6 4.9.2.11.1.1 System description 4.9.2.11.1.2 Analysis 4.9.2.11.1.3 Results and discussion 4.9.2.11.1.4 Conclusions 4.9.3 Cascade Refrigeration Cycle 4.9.3.1 Case Study 7: Two-Stage Cascade Refrigeration System 4.9.3.1.1 System description 4.9.3.1.1.1 Analysis 4.9.3.1.1.2 Results and discussion 4.9.3.1.1.3 Concluding remarks 4.9.4 Future Directions 4.9.5 Concluding Remarks Acknowledgment

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314 314 315 318 318 319 319 324 324 324 328 328 329 329 330 331 331 331 333 333 333 333 334 334 335 335 337 337 338 338 338 338 341 341 342 343 344 347 347 348 348 349 352 355 355 355 356 358 358 359 360 361 361

doi:10.1016/B978-0-12-809597-3.00411-9

Combined Energy Conversion Systems References Further Reading Relevant Websites

361 362 362

ZT

Limiting current density (mA/cm2) Thermal conductivity (W/m K); turbulent kinetic energy (m2/s2); specific heat ratio Lower heating value (kJ/kg) Negatively doped Mass flow rate (kg/s) Positively doped Pressure (kPa); power (kW) Heat rate (kW) Compression ratio Universal gas constant (J/ K mol) Seebeck coefficient (V/K) Specific entropy (kj/kg K) Temperature (K) Overall heat transfer coefficient (W/m2 K) Fuel utilization factor Specific volume (m3) Velocity (m/s) Specific work (kJ/kg) Work rate (kW) Modified dimensionless figure of merit

EIF GHG GTC HEX HRSG LNG LPG N NG ORC PCM TEG WHR

Efficiency improvement factor Greenhouse gases Gas turbine cycles Heat exchanger Heat recovery steam generation Liquefied natural gas Liquefied petroleum gas Total number Natural gas Organic Rankine cycle Phase change material Thermoelectric generator Waste heat recovery

Greek letters Z Energy efficiency c Exergy efficiency sSB Stefan Boltzmann constant (5.76  10 8/m2 K 4)

l ζ e

Air–fuel equivalence ratio Specific exergy function Surface emissivity

Subscripts ac act c conc C d Fc

field Ge GT i In ins ise ohm

Heliostat field Generator Gas turbine State number Inlet of a component Insulation Isentropic Ohmic

Nomenclature A An B cp cv DNI E E1 Eẋ ex F h I j j0

Acronyms AF ASTM ATS C C CAES CCPP CEGB CI CO2 EES

Constant Ratio between inlet and exit cross-sectional area of the nozzle Constant Specific heat capacity at constant pressure (kJ/ kg K) Specific heat capacity at constant volume (kJ/ kg K) Direct normal irradiance (W/m2) Nernst potential or open circuit voltage (V) Ideal cell voltage at standard conditions (V) Exergy rate (kW) Specific exergy (kJ/kg) Body forces (N m); Faraday constant (96,485 C/ mole) Specific enthalpy (kJ/kg), convective heat transfer coefficient (W/m2 K) Current (mA) Current density (mA/cm2) Exchange current density (mA/cm2)

Air–fuel ratio American Section of the International Association for Testing Materials Advanced turbine systems Compressor Coefficient Compressed air energy storage Combined cycle power plants Central electricity generating board Compression-ignition Carbon dioxide Engineering Equation Solver

and superscripts Aircraft Activation Cell Concentration Compressor Destruction Fuel cell

313

j1 k LHV _ m n p P _ Q r R S s T U Uf v V w _ W

314

Combined Energy Conversion Systems

Out ov p P Ph rec reg s

4.9.1

Outlet of a component Overall Propulsion Pump Physical Recuperator Regeneration Shaft power

Sys si sr ST T 0 1 1, 2, … i

System Heat sink Heat source Steam turbine Turbine, thrust power Ambient condition Freestream State points

Introduction

There are various sources of energy, which need to be converted into numerous forms of energy for direct or indirect utilization. Energy conversion is defined as the transformation of energy from forms supplied from nature to forms that can be utilized by humans. In nature, numerous energy conversion processes occur inherently, such as photosynthesis. In addition, starting from the first ages, humans have also been able to design and develop new devices for this purpose. The energy conversion systems have a large variety including simple applications such as the basic windmill, which converts wind energy to the kinetic energy, or more complex systems such as nuclear power plants, which convert nuclear fuels to thermal and then electrical energy. Some basic conversion samples are listed in Table 1, which shows the initial and final form of energy through the transformation process. Such systems consist of multiple stages and processes where energy undergoes a whole series of conversions via several intermediate forms. Some of these paths are depicted in Fig. 1, which indicates pathways of the energy conversion. The performance of these systems is determined and limited by the thermodynamic laws and other specific laws. In recent years, direct conversion systems have excited the attention of researchers and investors due to the significant attention that has been devoted to certain direct energy conversion devices, particularly solar cells and thermoelectric generators (TEGs), that bypass the intermediate step of conversion to heat energy in electrical power generation [1]. However, efficiencies of simple cycles have not reached the desired levels. Especially the exhaust streams of the heat engines cause extensive losses to the environment. Combined cycles were introduced in order to utilize the part of the rejected heat into a bottoming cycle. The combined cycles consist of two different systems providing the input energies from the same source and aim to produce the same useful output. Brayton–Rankine combined cycle is the most well-known application of the combined energy conversion systems. In this chapter, a survey of conversion methods of various types of the combined energy transformation cycles take place, numerous systems are thermodynamically analyzed and discussed, and the performance assessments of such systems are performed in the following sections.

4.9.2

Combined Cycle Power Plants

A combined cycle power plant usually uses both a gas and a steam turbine (ST) together. Exhaust temperatures of gas turbines (GTs) are considerably higher than the atmospheric temperature. The discharging process occurs at atmospheric pressure for the open Brayton cycle. Therefore, it is impossible to expand the working fluid below. The high temperature and low-pressure stream still have high potential and it can be used by the integration of the waste heat recovery (WHR) units to the Brayton cycle. The high energy can be recovered if the expelled gases can be utilized in a Rankine cycle [2]. CCPPs are smart power production selections because of their higher thermal efficiencies than that presented in individual steam or gas turbine cycles (GTCs). In these structures, the rejected low-pressure hot flue gases from the GT are sent to a heat recovery steam generation (HRSG) unit where Table 1 Examples of energy conversion processes listed according to the initial energy form and one particular converted energy form (the one primarily wanted) Initial energy forms

Converted energy forms Chemical

Chemical Electrical

Heat

Fuel cell, battery discharge

Burner, boiler Resistance, heat pump Convector, radiator, heat pipe

Electrolysis, battery charging

Heat

Mechanical Nuclear Radiant

Electrical

Thermionic and thermoelectric generators (TEGs) Electric generator, MHD Photolysis

Photovoltaic cell

Friction, churning Reactor Absorber

Mechanical

Radiant

Electric motor

Lamp, laser

Thermodynamic engines

Turbines

Source: Adapted from Sørensen B. Renewable energy conversion, transmission, and storage. Burlington, MA; London: Elsevier/Academic Press; 2007.

Combined Energy Conversion Systems

315

Energy sources

Fossil fuel energy

Green energy Energy recovery (industrial waste heat, landfill gas, etc.)

Thermoelectrical energy

Renewable energy

Solar

Biomass

Geothermal

Wind

Photonic energy

Biochemical energy

Thermal energy

Electrical energy

Hydro Energy

Ocean thermal

Tides and waves

Fig. 1 Pathways of energy conversion.

supplementary fuel is supplied. There is enough O2 in the gases to sustain a burning process, which leads to temperature growth. Moreover, the hotter gases are utilized in an arrangement of heat exchangers (HEXs) to produce steam for a Rankine cycle. The key task in modeling a combined cycle is the proper usage of the GT waste heat in the steam cycle to attain best ST production. According to the advantages of combined cycle power plants (CCPP), the quantity and output power of such cycles are higher than before today. Combined cycles have greater thermal efficiency, besides their higher output power when compared to GT and steam cycles. Higher performances of CCPP compared to Brayton and Rankine cycles have made them quite attractive for power generation. Based on these advantages and less specific emissions, CCPP have widely been used all around the world. The main design parameters of CCPP are pressure ratio (r), compressor isentropic efficiency, GT isentropic efficiency, ST isentropic efficiency, and gas turbine inlet temperature (TIT) [2]. A simple diagram of an air–steam CCPP is illustrated in Fig. 2. In this diagram, a Brayton cycle of basic configuration is considered. The hot gases expelled have a relatively large flow rate and are directed into the HEX, working as a steam boiler for the Rankine cycle, and help to recover the part of the waste heat. The only heat addition to the system occurs in the combustion chamber at constant pressure and heat rejections take place at after steam generator for the Brayton cycle part and condenser at Rankine cycle part.

4.9.2.1

Brayton–Rankine Combined Cycle

The gas and STs employed in the combined cycle can be considered as heat engines as they take heat from the heat source (solar, biomass, fossil fuels) and reject it to a heat sink (usually atmosphere) and generate useful work. The topping cycle of the gas–steam combined systems is mainly a simple Brayton cycle where in its combustion turbine, pressurized combustion gases are utilized as working fluid of the system. In such systems, to enhance the combustion efficiency, excess air is introduced to the combustion chamber even more than 50 times of the stoichiometric air rate. Thus, assuming the properties of combustion gas as the properties of the air corresponding temperature and pressure would be a decent permissible approximation. For the cases when the air is modeled, the working fluid is modeled as ideal gas called “air-standard” cycles. In the air-standard topping cycle, the specific heat of air is assumed constant, which is determined at 251C [2]. The air is taken from ambient conditions at state 1 and pressurized to GT inlet pressure. Ideally, that process is isentropic, which means it is both adiabatic and reversible. In state 2, high-pressure air flows into the combustion chamber/HEX and heats up at constant pressure. In state 3, combustion gases with greater temperature and pressure expand in the GT and the GT produces mechanical work. In the ideal cycle, the expansion is also isentropic in ideal conditions. For the closed systems, heat removal after the expansion is provided by a HEX at a constant pressure where the open systems discharge the combustion gases to the atmosphere after the turbine. Burning the working fluid causes the feed air requirement in the system. Therefore, heat addition should also be attained by a closed HEX in the closed systems. As it is stated in the GTCs chapter, in terms of thermodynamic aspect, closed and open Brayton cycles are equivalent. To be precise in the open Brayton cycle, the atmosphere acts as a cooler, where the cooling process 4–1 takes place at constant pressure. On the other hand, in the closed Brayton cycle, the same process is conducted in a HEX. Closed systems are usually favored for the cases when the work output of the cycle is required to be enhanced. Atmospheric conditions limit the outlet pressure of the turbine. In order to overcome that limitation, the cycle should be closed, so that the turbine can be expanded up to the vacuum [2]. Moreover, in the closed systems, the HEX for the heat rejection can be used for the

316

Combined Energy Conversion Systems

Fuel intake

Combustion chamber

Compressor 2

3 WGT

WC

Gas turbine 1

4

5 6

7

Heat exchanger

WST Steam turbine 8

Condenser

Pump 9 WP

Q Fig. 2 A simple layout of the air–steam combined cycle power plant (CCPP).

boiler of the Rankine (bottoming) cycle. A conventional bottoming cycle of a steam–air combined cycle operates on a Rankine cycle consists of a pump, a steam generator (boiler), a ST, and a condenser. The subsystem of the combined cycle is the most basic steam cycle, which consists of four subsystem components and processes. Initially, the water is pressurized by a pump and reach to the boiler pressure (process 9–6). Then the waste heat of the Brayton cycle is added at constant pressure in the boiler (process 6–7). The water is heated until it entirely changes its phase from water to steam and then superheated. The superheated steam produces useful work as it expands in the ST (process 7–8) and subsequently, a condensation process occurs up until the steam reaches saturated liquid state (process 8–9). In the ideal air–team cycle, pump, compressor and steam and GTs should operate isentropically (reversible adiabatic) in order to avoid internal and external irreversibilities. The heat transfer surface of the condenser should be infinite (or the heat transfer coefficient of the element should be so large to maintain the temperature difference between working fluids and the heat sink zero). Furthermore, pressure drops in the pipelines, connections, boiler, and condenser should be zero. Even though the ideal combined cycle is internally reversible, because of the temperature changes in the heat addition process it has external irreversibilities. Therefore, the exergetic efficiency of such cycles must be determined smaller than 1 [2]. When the kinetic and potential energy changes in the working fluid are neglected and the ideal gas laws applied balance equations can be expressed for a unit mass basis as follows [2]: Mass balance equations for compression, heat addition, expansion, and heat rejection processes, respectively: _ 2; m _2¼m _ 3; m _3¼m _ 4; m _4¼m _5 MBE : m _1¼m _6¼m _ 7; m _7¼m _ 8; m _8¼m _ 9; m _9 ¼m _6 m

ð1Þ

_ indicates the mass flow rate and the numbers used as indices show the state number in Fig. 2. There is no external mass here, m addition or rejection to the subsystems. So all the subsystem processes have the same mass. Energy balance equations for compression, heat addition, expansion, heat rejection of the Brayton cycle and pressurization, heat addition, expansion, condensation of the Rankine cycle can be written as: EnBE : h1 þ wc ¼ h2 ; h2 þ qsr ¼ h3 ; h3 ¼ h4 þ wgt _ 4 h4 ¼ m _ 5 h5 þ m _ 7 h7 wc þ qsr ¼ wt þ qsi h9 þ wp ¼ h6 _ 6 h6 þ m m h7 ¼ h8 þ wst ; h8 ¼ h9 þ qc

ð2Þ

here, h, w, and q show the specific enthalpy, specific work, and specific heat values. si, sr, c, p, gt, and st indicate the heat sink, heat source, compressor, pump, gas turbine (GT) and steam turbine (ST), respectively. While the compressor work, pump work, and the heat addition to the system are the energy inlet of the cycle, the GT and ST's work and the heat rejection are the energy outlet of the cycle. According to the first law of thermodynamics, heat and work inlet should be equal to the heat and work exit of the system. Net useful work done by the combined system can be expressed as the difference between produced work by the

Combined Energy Conversion Systems

317

turbine and consumed work by compressor _ gt þ W _ st _ net ¼ W W

_c W

_p W

ð3Þ

Heat inlet from the heat source and heat removal to the heat sink can be defined as follows: _ sr ¼ m _ 3 ðh3 Q

_ 3 cp ðT3 h2 Þ ¼ m

T2 Þ;

_ si ¼ m _ 9 ðh8 Q

h9 Þ

ð4Þ

here h3 and h2 are the specific enthalpies of the stream after and before the heat addition, respectively, and h8 and h9 represent the specific enthalpies of the air at the inlet and exit of the heat rejection. Entropy balance equations for compression, heat addition, expansion, heat rejection of the Brayton cycle and pressurization, heat addition, expansion, condensation of the Rankine cycle can be written as: _ 4 s4 þ S_ g;si þ m _ 6 s6 ¼ m _ 5 s5 þ m _ 7 s7 ; s1 þ sg;c ¼ s2 EnBE : s3 þ sg;gt ¼ s4 ; m Z 9 qso dqsi þ sg;sr ¼ s3 ; s8 þ sg;si ¼ s9 þ ; s9 þ sg;p ¼ s6 s2 þ T3 8 Tsi

ð5Þ

Here, s stands for specific entropy and subindex “g” represents entropy generation. As the compression and expansion processes in the ideal Brayton–Rankine combine cycle are isentropic, entropy generation for those processes equals to zero (sg,c ¼ sg,t ¼ 0 kJ/kg/ K). On the other hand, heat transfer processes at both the heat sink and heat sources are internally reversible but externally irreversible for the ideal cycle. In other words, heat transfer from the system boundaries to the heat sink and heat source take place. Therefore, entropy production and exergy destruction occur in the cycle. Exergy balance equations for compression, heat addition, expansion, heat rejection of the Brayton cycle and pressurization, heat addition, expansion, condensation of the Rankine cycle can be written as:   T0 ¼ ex 3 þ exd;sr ; ex3 ¼ ex4 þ wgt þ ex d;t ExBE : ex1 þ wc ¼ ex 2 þ ex d;c ; ex 2 þ qsr 1 T3 _ 6 ex 6 ¼ m _ 5 ex 5 þ m _ 7 ex 7 þ ex d ex7 ¼ ex8 þ wt þ exd;st _ 4 ex 4 þ m m ð6Þ   T0 ex 8 ¼ qsr 1 þ ex 9 þ exd;si ; ex 9 þ wp ¼ ex6 þ ex d;p T9 Here, ex represents the specific exergy rate, T0 is reference temperature, and d indicates the exergy destruction. In the ideal Brayton–Rankine combine cycle, exergy destruction for compression at the compressor, pressurization at the pump and expansion at the steam and GTs processes are also zero as those processes are adiabatic and reversible (exd,c ¼exd,t ¼ 0 kJ/kg). As the those processes occur isentropically and the heat addition and the heat rejection processes are conducted at isobaric conditions, those relations can be written under ideal gas conditions [3].  k 1=k T2 P2 ¼ ð7Þ c¼ T1 P1 Here k is the specific heat ratio for the air, which shows the ratio between specific heat of the air at constant volume and constant pressure. It should be noted that, due to the infinite heat transfer area approach in the thermally isolated steam generator and condenser, in the ideal Brayton–Rankine combine cycle the temperature of the heat sink Tsi is equal to the condensation temperature (T6 and T8). Furthermore, the temperature of the heat source Tsr is considered as equal to the highest temperature of the stream (T3). The thermal efficiency of the Brayton–Rankine combined cycle is the ratio between useful network generated by the system and GTCs and total energy input to the cycle. It can be shown as follows: Z¼

wnet ¼1 qsr

qsi qsr

ð8Þ

The exergy efficiency of the system can be defined either by the ratio between the net delivered exergy and the actual consumed exergy [2] or the ratio between the real work to the reversible work [3] as follows:   _ actual Ex useful Ex dov W c¼ ¼1 ð9Þ ; c2 ¼ _ rev Ex input Exinput W The exergy efficiency of the cycle can be expressed for the cases in which the heat source has a constant temperature: c¼ 1

_ W  net T0 Tsr

_ sr Q

¼

Z ZC

ð10Þ

where ZC indicates the Carnot factor ZC ¼ 1 TTsr0 . In the following case study, a combined cycle fueled with natural gas (NG) is analyzed using exergy and energy analysis to investigate the performance of the combination of the Brayton cycle and Rankine cycle. However, to show the benefits of combining the Brayton cycle and the steam cycle we are going first to assess the performance of a Rankine cycle then the combined cycle performance is going to be evaluated.

318

Combined Energy Conversion Systems

4.9.2.1.1

Case Study 1

In the following case study, a combined cycle fueled with NG is analyzed using exergy and energy analysis to investigate the performance of the combination of the Brayton cycle and Rankine cycle. However, to show the benefits of combining the Brayton cycle and the steam cycle we are going first to assess the performance of a Rankine cycle then the combined cycle performance is going to be assessed. 4.9.2.1.1.1 Single steam Rankine cycle An actual Rankine cycle with an inlet pressure to the turbine of 7 MPa and a temperature of 5001C, the turbine exit pressure is 100 kPa (see Fig. 3). The isentropic efficiency of the turbine is 94%, and the pump, condenser, and the boiler pressure losses are neglected. If the temperature of the water leaving the condenser is 501C and the mass flow rate is 20 kg/s of water in the cycle calculate (1) the thermal energy sent to the boiler, (2) the net power produced, and (3) the energy and exergy efficiencies of the cycle. We first write the mass, energy, entropy, and exergy balance equations as a first step in calculating the vapor fraction and the exergy efficiency of the throttling process. Condenser:

_ out MBE : m _ in ¼ m _ cond _ out hout þ Q EBE : m _ in hin ¼ m _ cond =Tcond _ out sout þ Q EnBE : m _ in sin þ S_ gen;cond ¼ m _ _ d;cond _ out ex out þ Ex _ ExBE : m _ in exin ¼ m þ Ex Qcond

Pump:

_ out MBE : m _ in ¼ m _ p ¼m _ out hout EBE : m _ in hin þ W _ out sout EnBE : m _ in sin þ S_ gen;p ¼ m _ p ¼m _ d;p _ out ex out þ Ex ExBE : m _ in ex in þ W

Qin

Boiler P = 7.0 MPa T = 500°C Steam turbine

Pump

Isentropic efficiency is 94%

Win

P = 100 kPa Condenser

T = 50°C mass flow rate 20 kg/s Fig. 3 A schematic diagram of the Rankine cycle.

Qout

Wout

Combined Energy Conversion Systems

319

Boiler: _ out MBE : m _ in ¼ m _ boiler ¼ m _ out hout EBE : m _ in hin þ Q _ _ boiler =Tboiler ¼ m _ out sout EnBE : m _ in sin þ Sgen;boiler þ Q _ _ _ d;boiler _ out exout þ Ex ExBE : m _ in exin þ Ex ¼m Qboiler

ST: _ out MBE : m _ in ¼ m _ st _ out hout þ W EBE : m _ in hin ¼ m _ _ out sout EnBE : m _ in sin þ Sgen;st ¼ m _ d;st þ W _ st _ out exout þ Ex ExBE : m _ in ex in ¼ m Going back to the EBE of the boiler, we can see that in order to calculate the thermal energy the boiler receives from the external source we first need to find the enthalpy entering the boiler. The enthalpy of the stream entering the boiler is the same enthalpy of the flow exiting the pump. Then:   wp ¼ vin;p Pout;p Pin;p ¼ 0:001012  ð7000 100Þ ¼ 6:984 kJ=kg Then:

wp ¼ hout;p

hin;p -6:984 ¼ hout;p

209:4

hout;p ¼ 216:4 kJ=kg qboiler ¼ hout;boiler hin;boiler ¼ 3411 216:4 ¼ 3194 kJ=kg _ boiler ¼ m _  qboiler ¼ 20  3194 ¼ 63883 kW Q In the cycle there is the pump that consumes power and there is the ST that produces power, and the power produced by the turbine is calculated based on the energy balance over the turbine wst;is ¼ hin;st hout;st;is ¼ 3422 2466 ¼ 944:5 wst;a ¼ wst;is =Zis ¼ 944:5=0:94 ¼ 1005 kJ=kg wnet ¼ wst;a wp ¼ 1005 6:984 ¼ 944:5 kJ=kg _ net ¼ m _  wnet ¼ 20  944:5 ¼ 19957 kW W The energy and the exergy efficiency equations for the overall cycle are presented next since there is no mass entering or leaving the overall Rankine cycle, then it is treated as a closed energy system and based on that the efficiencies are derived. _ boiler ¼ 19957=63883 ¼ 31:24% _ net =Q Zen;RC ¼ W _ net =Ex _ _ ¼ 19957=42076 ¼ 47:43% Zex;RC ¼ W Qboiler

4.9.2.1.1.2 Natural gas fueled combined cycle In this combined cycle NG is the fuel that is fed to the combustion chamber. The combined cycle investigated is shown in Fig. 4. From state 1 to state 2 air is compressed using the compressor where the power is supplied by the GT. The resulting compressed air exiting the compressor is used as an oxidant in the combustion chamber fueled by NG, where state 3 presents the exhaust gases. The GT is used to expand the high-pressure exhausts to produce power. The exhaust gases leave the turbine then pass into the shell and tube HEX where they release heat to the water entering the HEX. The heat is released in the HEX, and the exhaust gases are released to the environment. The heat gained by the water entering the HEX leaves in a state of saturated vapor. The saturated vapor heads to the turbine where it expands to produce power. The water leaves the ST in a state of saturated mixture and is completely converted to saturated liquid through the condenser. The bottoming cycle is a simple Rankine cycle operating between the pressure limits of 7 MPa and 5 kPa. The steam is heated to a temperature of 5001C in a HEX by the heat supplied from the exhaust gases of a NG combustion system. The exhaust gases leave the HEX (state 5) at 450K. The ambient temperature and pressure are at 271C and 100 kPa. All the components are considered adiabatic. NG cycle is modeled on the basis of the airstandard theorem. Air is treated as an ideal gas. Temperature and pressure losses in the pipes and connections are neglected. It is assumed that 20% of the power is lost due to the parasitic losses. 4.9.2.1.1.3 Energy and exergy analyses The pressure at state 2 is calculated as Pr2 ¼ rp  Pr 1 where, Pr2 is the reduced pressure at state 2, rp is the pressure ratio, and Pr1 is the reduced pressure at state 1.

ð11Þ

320

Combined Energy Conversion Systems

Natural gas

3

Combustion chamber 2

Gas turbine

Compressor Brayton cycle

Wout

Wout 4 Steam generator

1

8

Steam turbine 5 Steam cycle

Wout

7 9 Pump

Win

Condenser

6

Qout Fig. 4 Natural gas (NG) fueled combined cycle.

The pressure at state 4 is calculated as Pr 4 ¼

1  Pr3 rp

where Pr4 is the reduced pressure at state 4, rp is the pressure ratio, and Pr3 is the reduced pressure at state 3. First each component of the system is analyzed by writing the balance equation as follows: Air compressor: _2 MBE : m _1¼m _ comp ¼ m _ 2 h2 EBE : m _ 1 h1 þ W _ 2 s2 EnBE : m _ 1 s1 þ S_ gen;comp ¼ m _ comp ¼ m _ d;comp _ 2 ex2 þ Ex ExBE : m _ 1 ex 1 þ W

ð12Þ

Combined Energy Conversion Systems

321

Combustion chamber: _ NG ¼ m _3 MBE : m _2þm _ NG LHV NG ¼ m _ 3 h3 EBE : m _ 2 h2 þ m _ NG sNG ¼ m _ 3 s3 EnBE : m _ 2 s2 þ S_ gen;CC þ m _ d;CC _ NG exNG ¼ m _ 3 ex 3 þ Ex ExBE : m _ 2 ex 2 þ m GT: _4 MBE : m _3¼m _ GT _ 4 h4 þ W EBE : m _ 3 h3 ¼ m _ 4 s4 EnBE : m _ 3 s3 þ S_ gen;GT ¼ m _ GT þ Ex _ d;GT _ ExBE : m _ 3 ex 3 ¼ m4 ex4 þ W HEX: _7 ¼m _5þm _8 MBE : m _4þm _ 7 h7 ¼ m _ 5 h5 þ m _ 8 h8 EBE : m _ 4 h4 þ m _ 7 s7 þ S_ gen;HX ¼ m _ 5 s5 þ m _ 8 s8 EnBE : m _ 4 s4 þ m _ d;HX _ _ _ ExBE : m _ 4 ex4 þ m7 ex 7 ¼ m5 ex5 þ m8 ex 8 þ Ex ST: _9 MBE : m _8¼m _ ST _ 9 h9 þ W EBE : m _ 8 h8 ¼ m _ 9 s9 EnBE : m _ 8 s8 þ S_ gen;ST ¼ m _ ST þ Ex _ d;ST _ ExBE : m _ 8 ex 8 ¼ m9 ex9 þ W Condenser: _ a:in ¼ m _6þm _ a;out MBE : m _9þm _ a:in ha;in ¼ m _ 6 h6 þ m _ a;out ha;out EBE : m _ 9 h9 þ m _ a:in sa;in S_ gen;C ¼ m _ 6 s6 þ m _ a;out sa;out EnBE : m _ 9 s9 þ m _ d;C _ a:in exa;in ¼ m _ 6 ex6 þ m _ a;out ex a;out þ Ex ExBE : m _ 9 ex 9 þ m Pump: _7 MBE : m _6 ¼m _ pump ¼ m _ 7 h7 EBE : m _ 6 h6 þ W _ _ 7 s7 EnBE : m _ 6 s6 þ Sgen;pump ¼ m _ pump ¼ m _ d;pump _ 7 ex7 þ Ex ExBE : m _ 6 ex 6 þ W Overall system: _ NG ¼ m _5 MBE : m _1þm _ comp þ W _ pump ¼ m _ GT þ W _ ST _ NG LHV NG þ W _ 5 h5 þ W EBE : m _ 1 h1 þ m _ NG sNG þ S_ gen;ov ¼ m _ 5 s5 EnBE : m _ 1 s1 þ m _ comp þ W _ pump ¼ m _ GT þ W _ ST þ Ex _ d;ov _ NG ex NG þ W _ 5 ex5 þ W ExBE : m _ 1 ex1 þ m From the balance equations, we can rearrange them to calculate the required work, thermal energy, etc. The ideal specific work needed by compressor of NG cycle to compress air from state 1 to state 2 is calculated as wcomp;g;s ¼ hs;2

h1

ð13Þ

where wcomp,g,s is the ideal specific work needed by the compressor of the NG cycle, hs,2 is the ideal specific enthalpy at state 2, and h1 is the specific enthalpy at state 1. The ideal specific power produced by turbine of the NG cycle is defined as wturb;g;s ¼ h3

hs;4

ð14Þ

where wturb,g,s is the ideal specific work produced by the turbine of the NG cycle, h3 is the specific enthalpy at state 3, and hs,4 is the ideal specific enthalpy at state 4. The ideal specific heat input to the NG cycle is found using qin;g;s ¼ h3

hs;2

ð15Þ

322

Combined Energy Conversion Systems

where qin,g,s is the ideal specific heat input to the combustion chamber of the NG cycle, h3 is the specific enthalpy at state 3, and hs,2 is the ideal specific enthalpy at state 2. The ideal parasitic loss is calculated as wparasitic;g;s ¼ 0:2  ðwturb;g;s

wcomp;g;s Þ

ð16Þ

wparasitic;g;s

ð17Þ

where wparasitic,g,s is the ideal specific parasitic loss of the NG cycle. The ideal specific net power produced by the NG cycle is defined as wnet;g;s ¼ wturb;g;s

wcomp;g;s

where wnet,g,s is the ideal net specific power produced by the NG cycle. The specific exergy at state 1 is found using ex 1 ¼ ðh1

h0 Þ

T0  ðs1

s0 Þ

ð18Þ

where ex1 represent specific exergy at state 1, h0 represents specific enthalpy at ambient state, s1 is specific entropy at state 1, and s0 is specific entropy at the ambient state. The same formulation is used to calculate specific exergy at each state. The ideal specific thermal exergy of the NG cycle is found using  T0 ex th;g;s ¼ 1 ð19Þ qin;g;s Tavg;g;s ðT3 þTs;2 Þ and exth,g,s is the ideal specific thermal exergy of the NG cycle, and Tavg,g,s is the ideal average temperature. where Tavg;g;s ¼ 2 The ideal energy and exergy efficiencies of the NG cycle are calculated using Zen;g;s ¼

wnet;g;s qin;g;s

ð20Þ

Zex;g;s ¼

wnet;g;s ex th;g;s

ð21Þ

The actual specific work needed by the compressor of the NG cycle to compress air from state 1 to state 2 is calculated as wcomp;g;a ¼

hs;2

h1

ð22Þ

Z

where wcomp,g,a is the actual specific work needed by the compressor of the NG cycle, hs,2 is the ideal specific enthalpy at state 2, h1 is the specific enthalpy at state 1, and Z is the isentropic efficiency, which is 85%. The actual specific power produced by turbine of the NG cycle is defined as wturb;g;a ¼ Zðh3

hs;4 Þ

ð23Þ

where wturb,g,a is the actual specific work produced by the turbine of the NG cycle, h3 is the specific enthalpy at state 3, hs,4 is the ideal specific enthalpy at state 4, and Z is the isentropic efficiency, which is 85%. The actual specific heat input to the NG cycle is found using qin;g;a ¼ h3

ð24Þ

h2

where qin,g,a is the actual specific heat input to the combustion chamber of the NG cycle, h3 is the specific enthalpy at state 3, and h2 is the actual specific enthalpy at state 2. The actual parasitic loss is calculated as wparasitic;g;a ¼ 0:2  ðwturb;g;a

wcomp;g;a Þ

ð25Þ

wparasitic;g;a

ð26Þ

where wparasitic,g,a is the actual specific parasitic loss of the NG cycle. The actual specific net power produced by the NG cycle is defined as wnet;g;a ¼ wturb;g;a

wcomp;g;a

where wnet,g,a is the actual net specific power produced by the NG cycle. The actual specific thermal exergy of the NG cycle is found using  ex th;g;a ¼ 1

T0 Tavg;g;a



qin;g;a

ð27Þ

Combined Energy Conversion Systems

323

ðT3 þTa;2 Þ and exth,g,a is the actual specific thermal exergy of the NG cycle, and Tavg,g,a is the actual average where Tavg;g;a ¼ 2 temperature. The actual energy and exergy efficiencies of the NG cycle are calculated using wnet;g;a Zen;g;a ¼ ð28Þ qin;g;a Zex;g;s ¼

wnet;g;a ex th;g;a

ð29Þ

The specific power consumed by pump of the steam cycle is found using  P7 P6 wp ¼ v6 Z

ð30Þ

where wp is specific power consumed by the pump, v6 is specific volume at state 6, and Z is the isentropic efficiency of the pump. The mass ratio “y” is defined as mass flow rate of the steam divided by mass flow rate of the NG and is calculated using y¼

h4 h8

h5 h7

ð31Þ

The specific power produced by turbine of the steam cycle is defined as wturb;st ¼ ðh8

h9 Þ

ð32Þ

where wturb,st is the specific work produced by the turbine of the steam cycle, h8 is the specific enthalpy at state 8, and h9 is the specific enthalpy at state 9. The specific heat input to the steam cycle is found using qin;st ¼ h8

ð33Þ

h7

where qin,st is the specific heat input to the steam cycle, h8 is the specific enthalpy at state 8, and h7 is the specific enthalpy at state 7. The specific heat output from the condenser of the steam cycle is found using qcon;st ¼ h9

ð34Þ

h6

where qcon,st represents specific heat output from the condenser of the steam cycle. The parasitic loss is calculated as wparasitic;st ¼ 0:2  ðwturb;st

wp Þ

ð35Þ

wparasitic;st

ð36Þ

where wparasitic,st is the specific parasitic loss of the steam cycle. The specific net power produced by the steam cycle is defined as wnet;st ¼ wturb;st

wp

where wnet,g,a is the actual net specific power produced by the NG cycle. The specific input thermal exergy of the steam cycle is found using  T0 exth;in;st ¼ 1 qin;st Tavg;st

ð37Þ

8Þ and exth,in,st is the specific input thermal exergy of the steam cycle, and Tavg,st is the average temperature. where Tavg;st ¼ ðT7 þT 2 The specific condenser thermal exergy of the steam cycle is found using  T0 ð38Þ ex th;con;st ¼ 1 qcon;st Tavg;con

6Þ and exth,con,st is the specific condenser thermal exergy of the steam cycle, and Tavg,con is the average where Tavg;con ¼ ðT9 þT 2 temperature. The energy and exergy efficiencies of the steam cycle are calculated using wnet;st Zen;st ¼ ð39Þ qin;st

Zex;g;s ¼

wnet;st ex th;st

ð40Þ

The net overall power output of the combined cycle is found using wnet;ov ¼ y  wnet;st

wnet;g;a

ð41Þ

The ideal specific exergy destruction in the compressor is found using ex dest;comp;g;s ¼ ex 1

exs;2 þ wcomp;g;s

where exdest,comp,g,s represents ideal specific exergy destruction in the compressor.

ð42Þ

324

Combined Energy Conversion Systems

The actual specific exergy destruction in the compressor is found using ex dest;comp;g;a ¼ ex 1

ex 2 þ wcomp;g;a

ð43Þ

where exdest,comp,g,a represents actual specific exergy destruction in the compressor. The ideal specific exergy destruction in the combustion chamber is found using ex dest;cc;g;s ¼ ex s;2

ex 3 þ exth;g;s

ð44Þ

where exdest,cc,g,s represents ideal specific exergy destruction in the combustion chamber. The actual specific exergy destruction in the combustion chamber is found using ex dest;cc;g;a ¼ ex 2

ex3 þ exth;g;a

ð45Þ

where exdest,cc,g,a represents actual specific exergy destruction in the combustion chamber. The ideal specific exergy destruction in the turbine of the NG cycle is found using ex dest;turb;g;s ¼ ex 3

exs;4

ð46Þ

wturb;g;s

where exdest,turb,g,s represents ideal specific exergy destruction in the turbine of the NG cycle. The actual specific exergy destruction in the turbine of the NG cycle is found using ex dest;turb;g;a ¼ ex 3

ex 4

ð47Þ

wturb;g;a

where exdest,turb,g,a represents actual specific exergy destruction in the turbine of the NG cycle. The specific exergy destruction in the turbine of the steam cycle is found using ex dest;turb;st ¼ ex 8

ex 9

ð48Þ

wturb;st

where exdest,turb,st represents specific exergy destruction in the turbine of the steam cycle. The specific exergy destruction in the pump of the steam cycle is found using ex dest;p;st ¼ ex 6

ex7 þ wp

ð49Þ

where exdest,p,st represents specific exergy destruction in the pump of the steam cycle. The specific exergy destruction in the heat exchange is found using exdest;he ¼ ex 4 þ ex7

ex5

ð50Þ

ex 8

where exdest,he represents specific exergy destruction in the HEX. The specific exergy destruction in the condenser of the steam cycle is found using ex dest;con;st ¼ ex 9

ex 6

ð51Þ

exth;con;st

where exdest,con,st represents specific exergy destruction in the condenser of the steam cycle. The overall specific exergy destruction is found as follows: ex dest;ov ¼ ex1 þ ex th;g;a þ wp þ wcomp;g;a

ex5

where exdest,ov represents overall specific exergy destruction. The overall energy and exergy efficiencies are then defined as wnet;ov Zov ¼ qin;g;a Zex;ov ¼

wnet;ov exth;g;a

wturb;g;a

wturb;st

exth;in;st

ð52Þ

ð53Þ

ð54Þ

4.9.2.1.1.4 Results As shown in Fig. 5, the component that contributes the most to the exergy destruction of the system is the HEX/boiler. The variation of the energy efficiency and the exergy efficiency of the combined cycle with the variation of the ambient temperature is presented in Fig. 6. The cycle overall energy efficiency is 37.3% and the overall exergy efficiency is 54.53% for the same parameters used in Dincer and Ratlamwala [4]. For further readings, please see Ref. [4]. 4.9.2.1.1.5 Closure It can be concluded from the study that combining the two cycles resulted in increasing the overall performance of the two cycles in terms of energy and exergy efficiencies, which can be easily noticed through the difference between the energy and exergy efficiencies of the combined cycle and the Rankine cycle presented earlier.

4.9.2.1.2

Case Study 2

Two Rankine cycles can be combined in order to increase the percentage utilization of the thermal energy that is delivered to boil the water and then superheat the vapor produced. A possible configuration of the combined cycles is presented in Fig. 7. As shown

Combined Energy Conversion Systems

325

Specific exergy destruction (kJ/kg)

2000 1800 1600 1400 1200 1000 800 600 400 200

l ve ra l

de on

C

om

H

ea

C

O

ns er

p Pu m

ng xc ha

te

m St ea

as G

er

ne tu rb i

ne tu rb i

be m ch a

bu

st io

C

n

om

pr es so r

r

0

40

56.0

38

55.5

36

55.0

34

54.5

32

54.0

30 290

294

298

302

306

ov (%)

ov (%)

Fig. 5 The specific exergy destruction of the components in the cycle.

53.5 310

Ambient temperature (K) Fig. 6 Variation of the energy efficiency and the exergy efficiency of the combined cycle with the variation of the ambient temperature.

in Fig. 7, the upper Rankine cycle uses water as the working fluid. However, the bottoming Rankine cycle uses an organic fluid such as ammonia that has a very low boiling point at atmospheric conditions. The evaluation of the performance of the combined Rankine is done by writing the mass, energy, entropy, and exergy balance equations. The operating pressure of the condenser 1 is 500 kPa. The steam enters the turbine with a pressure of 7 MPa and temperature of 5001C. where the ST is assumed to be isentropic. The exit pressure of the ammonia coming out of the pump 2 is 7.583 MPa, which corresponds to the saturation temperature of ammonia at 1101C. The mass flow rate of the ammonia in the organic Rankine cycle (ORC) equals to 54.75 kg/s. For the steam Rankine cycle (SRC) (upper Rankine cycle): Condenser: _4 MBE : m _3 ¼m _ cond _ 4 h4 þ Q EBE : m _ 3 h3 ¼ m _ cond =Tcond _ 4 s4 þ Q EnBE : m _ 3 s3 þ S_ gen;cond ¼ m _ d;cond _ _ 4 ex 4 þ Ex _ ExBE : m _ 3 ex3 ¼ m þ Ex Qcond

326

Combined Energy Conversion Systems

Qin Boiler 2

1

Steam turbine WST

Heat exchanger WP1

3 Pump 1

4 6 5 Ammonia expander WAE 7

WP2

Pump 2

8

Condenser

Qout Fig. 7 Combined Rankine cycle.

Pump: _1 MBE : m _4¼m _ p ¼m _ 1 h1 m _ 4 h4 þ W _ _ 1 s1 EnBE : m _ 4 s4 þ Sgen;p ¼ m _ p ¼m _ d;p _ 1 ex1 þ Ex ExBE : m _ 4 ex 4 þ W EBE :

Boiler: _2 MBE : m _1¼m _ boiler ¼ m _ 2 h2 EBE : m _ 1 h1 þ Q _ _ boiler =Tboiler ¼ m _ 2 s2 EnBE : m _ 1 s1 þ Sgen;boiler þ Q _ _ _ 2 ex2 þ Ex d;boiler ExBE : m _ 1 ex 1 þ Ex _ ¼m Qboiler

ST: _3 MBE : m _2¼m _ st _ out hout þ W EBE : m _ 2 h2 ¼ m _ _ EnBE : m _ 2 s2 þ Sgen;st ¼ m3 s3 _ d;st þ W _ st _ 3 ex3 þ Ex ExBE : m _ 2 ex 2 ¼ m For the ORC (bottoming Rankine cycle): Condenser: _8 MBE : m _7 ¼m _ cond _ 8 h8 þ Q EBE : m _ 7 h7 ¼ m _ cond =Tcond _ 8 s8 þ Q EnBE : m _ 7 s7 þ S_ gen;cond ¼ m _ _ _ _ 8 ex 8 þ Ex ExBE : m _ 7 ex7 ¼ m Qcond þ Ex d;cond

Efficiency (%) and mass flow rate (kg/s)

Combined Energy Conversion Systems

327

65 60 55 Exergy efficiency (%) Ammonia mass flow rate in the bottoming ORC (kg/s) Energy efficiency (%)

50 45 40 35 500

600

700

800

900

1000

Operating pressure of condenser 1 (kPa) Fig. 8 The variation of the energy efficiency, exergy efficiency, and the mass flow rate of ammonia with the variation of the operating pressure of condenser 1. ORC, organic Rankine cycle.

Pump: _5 MBE : m _8¼m _ p ¼m _ 5 h5 EBE : m _ 8 h8 þ W _ 5 s5 EnBE : m _ 8 s8 þ S_ gen;p ¼ m _ p ¼m _ d;p _ 5 ex5 þ Ex ExBE : m _ 8 ex 8 þ W Boiler: _6 MBE : m _5¼m _ boiler ¼ m _ 6 h6 EBE : m _ 5 h5 þ Q _ boiler =Tboiler ¼ m _ 6 s6 EnBE : m _ 5 s5 þ S_ gen;boiler þ Q _ _ _ 6 ex6 þ Ex d;boiler ¼m ExBE : m _ 5 ex 5 þ Ex _ Qboiler

Organic turbine: _7 MBE : m _6¼m _ ot _ 7 h7 þ W EBE : m _ 6 h6 ¼ m _ 7 s7 EnBE : m _ 6 s6 þ S_ gen;st ¼ m _ d;st þ W _ ot _ 7 ex 7 þ Ex ExBE : m _ 6 ex 6 ¼ m By solving the balance equations, the overall performance of the cycle can be summarized as follows: _ st ¼ 13439 kW W _ ot ¼ 13367 kW W _ boiler ¼ m _  qboiler ¼ 20  3194 ¼ 63883 kW Q _ boiler ¼ 26806=63883 ¼ 41:96% _ net =Q Zen;RC ¼ W Zex;RC ¼

_ net 26806 W ¼ ¼ 63:70% _ _ 42076 Ex Qboiler

The performance of the combined cycle is investigated through varying the operating pressure of the condenser of the SRC, where its effect on the energy, exergy efficiencies of the combined cycle and the ammonia mass flow rate in the bottoming ORC are investigated. The results of this study are shown in Fig. 8. As shown in Fig. 9, as the operating pressure of condenser 1 increases it will result in reducing the overall energy and exergy efficiencies. However, the mass of the ammonia circulating in the bottoming Rankine cycle almost remains constant. The drop in the energy efficiency was noticed to be more dramatic than the drop in the exergy efficiency as shown in Fig. 10, where the cause is further investigated through the variation of the power produced by the two cycles as shown in Fig. 10. Fig. 11 shows the variation of the energy and exergy efficiencies with the variations of the ambient temperature. The exergy efficiency

328

Combined Energy Conversion Systems

Power production (W)

13,100 Power produced by the ORC (W) 12,600

Power produced by the SRC (W)

12,100 11,600 11,100 10,600 500

600

700

800

900

1000

Operating pressure of condenser 1 (kPa)

42

65.4

41.8

65.2

41.6

65

41.4

64.8

41.2

64.6

41 Energy en eff efficiency (%) 64.4

40.8

Exergy ex eff efficiency (%)

40.6

64.2

40.4

64

40.2

63.8

40 25

30

35 40 Ambient temperature (°C)

45

50

Exergy efficiency (%)

Energy efficiency (%)

Fig. 9 The variation of the power produced by each of the two cycles in the combined cycle with the variation of the operating pressure of condenser 1. ORC, organic Rankine cycle; SRC, steam Rankine cycle.

63.6

Fig. 10 The variation of the energy and exergy efficiencies with the variation of the ambient temperature.

changes slightly, while the energy efficiency remains constant as shown in Fig. 10 from a value of 63.7% to 65.2% when the temperature increases from 25 to 551C. 4.9.2.1.2.1 Closure It can be concluded from the study that combining the two cycles resulted in increasing the overall performance of the two cycles in terms of energy and exergy efficiencies, which can be easily noticed through the difference between the energy and exergy efficiencies of the combined cycle and the Rankine cycle presented earlier. However, from the parametric studies, it is recommended to run the condenser at relatively lower pressure than the inlet pressure of the ST in the topping SRC.

4.9.2.2

The Brayton–Kalina Cycle

In this section, various types of combined cycles are explained and future directions are discussed. Dr. Andreas Poullikkas’s review article “An overview of current and future sustainable gas turbine technologies” [5] is followed closely in those sections. The Kalina cycle is a novel bottoming cycle developed by Dr. Alexander Kalina, which uses a zeotropic mixture of two fluids with different boiling points as the working fluid (ammonia and water). Its characteristics are such that temperature of the cycle tracks the temperature of turbine exit in the waste heat boiler. Nevertheless, at the condensing processes, the thermodynamic gain of the relatively small boiler temperature difference compared to a steam cycle would be gone, supposing the cooling temperature of condenser cooling medium would be identical in both cases. In order to solve these problems the Kalina cycle was introduced [5]. A basic Kalina cycle consists of a WHR vapor generator (in this case it is the HEX where the exhaust gas of the GT is directed to supply heat to bottoming Kalina cycle), a steam-ammonia turbine and the distillation condensation system, as it is depicted in

Combined Energy Conversion Systems

Steam turbine

329

Generator

Seperator Heater Condenser HEX – Steam generator

Fuel intake

Combustion chamber

Pump

Absorber

Compressor

Gas turbine

Generator

Fig. 11 The Brayton–Kalina cycle. HEX, heat exchanger.

Fig. 11. The hot stream from the steam-ammonia turbine is cooled in the recuperator and then mixed with a lean solution of NH3 to increase the condensing temperature. Finally, basic solution is condensed in the absorber and the condensed solution is sent to the heater under pressure. A part of the flow is directed to dilute the ammonia-rich stream coming from the separator. The primary stream flows through the recuperator. Afterwards, it is flashed into the separator. Then the vapor is condensed in the condenser and pressurized by the pump and sent to the vapor generator [6]. Note that 10%–30% more energy can be generated by the Kalina cycle when it is compared with a Rankine cycle [6]. The main reason for this situation is the high exhaust pressure of the Kalina cycle, which is above the atmospheric conditions. The startup time of a Kalina cycle is much less because maintaining the vacuumed medium is not a necessity for the condenser, during the operation of the cycle. The working fluid mixture can easily be changed to obtain the optimum performance with respect to changes in load or ambient conditions. Moreover, those systems are favorable due to their compact configuration. The footprint of a Kalina plant is about 40% smaller than a Rankine plant design when they are compared [5]. In order to commercialize the Brayton–Kalina cycle plants, a special global licensing contract has been signed with an agreement between the Exergy Inc., who owned the rights to the technology, and General Electric in 1993. Afterward, GE designed a Brayton–Kalina cycle of 260 MW capacity, which was planned to be in service by 1998 but the project was suspended later [5,7].

4.9.2.3

The Brayton–Brayton Cycle

Two Brayton cycles can be combined by an air–gas HEX as shown in Fig. 12. The exhaust stream of the main cycle flows into a HEX where it supplies required heat by the working fluid of the second Brayton cycle. The working fluid is expanded in the secondary turbine to generate supplementary power. When this scheme is compared with the conventional Brayton–Rankine combined cycle, this layout does not need a large mass of steam equipment such as a condenser, boiler, and ST. Moreover, it does not require a water treating component, which allows remote and unmanned processes possible in these applications [5]. The current studies presented the viability of this configuration [6]. Depending on the design parameters it is possible to increase overall efficiency by 10%. Furthermore designing parameters such as the number of intercoolers can boost the power capacity of the systems by 18%–30%. For instance, Allison 571K topping GT can be considered as a good example for this layout [8]. By the addition of the air bottoming cycle, which consists of two intercoolers, power production of the cycle is increased from 5.9 to 7.5 MWe. Moreover, the overall thermal efficiency is inclined from 33.9% to 43.2%. Similar outcomes were found with the General Electric LM2500 topping turbine [9]. The system can also be applied for cogeneration. Instead of wasting the exhaust stream, which is discharged by the cycle at 473–523K, they can be utilized for the cycle needs that require heat of such temperatures [5].

4.9.2.4

The Brayton–Diesel Cycle

By directing the hot exhaust stream of the gas turbine to a heat exchanger for preheating of the fuel of a Diesel engine, it is possible to sufficiently improve the efficiency. A simple layout for the system is presented in Fig. 13. Another application can be a direct

330

Combined Energy Conversion Systems

Fuel intake

Combustion chamber

Compressor

Gas turbine

Generator

Discharge

Compressor

Gas turbine

Generator

Fig. 12 The Brayton–Brayton cycle.

Fuel intake

Combustion chamber

Compressor

Gas turbine

Generator

Discharge

Diesel engine

Generator Fig. 13 The Brayton–Diesel cycle.

usage of the exhaust stream of the GT in the Diesel engine. As the GT consists of an oxygen–rich mixture, there would be still enough amount O2 (14%–16%) in the combustion products that can be burned further in the Diesel engine [6].

4.9.2.5

The Brayton–Stirling Cycle

A combined Brayton–Stirling cycle can be achieved by either locating the heater of the Stirling engine in the combustion chamber of the gas turbine or next to the expander as depicted in Fig. 14. The configuration is set by the ideal operation conditions of the cycle. Another restriction for the cycle`s configuration is the type of the materials used in the head of the Stirling heater. Up to 9 MWe can be recovered by the utilization of the waste heat of a Rolls–Royce RB211 GT of 27.5 MWe in the bottoming Stirling

Combined Energy Conversion Systems

Fuel intake

331

Combustion chamber

Compressor

Gas turbine

Stirling engine

Generator

Regenerator

Generator Fig. 14 The Brayton–Stirling cycle.

cycle [8]. Such a plant can achieve an efficiency of 47.7%. Similar with the Brayton–Brayton cycle, this combination presents a compact and simple waste heat recovery configuration [5].

4.9.2.6

The Brayton–Fuel Cell Cycle

A fuel cell system, which can offer great efficiency (60%), can be run at higher pressures and can produce very high temperature exhaust gases. This situation allows integrating a GT to the scheme, therefore performance of the system is improved [10]. The diagram of the structure is illustrated in Fig. 15. The utilization of the fuel cells combined with combustion chambers allows efficiency to approach 70%. [11]. The Brayton–fuel cell cycle is appealed to have the highest efficiency compared with all the advance cycles. Thus, it can be considered as a one of the most promising power plant applications for the future [5,12].

4.9.2.6.1

Case Study 3

In this study, a combination of a high-temperature solid oxide fuel cell (SOFC) combined and a gas turbine (GT–SOFC) plant is proposed for performance investigation. A thermodynamic analysis is carried out exploiting the first and second law approaches to analyze each part of the proposed system. The overall system efficiency is investigated along with the exergy destruction occurred in the plant. The proposed system is also compared to a conventional GT cycle with the same operating conditions for further performance investigation. 4.9.2.6.1.1 System description The combined GT–SOFC is displayed in Fig. 16. The system is comprised of six main parts: (1) high-temperature SOFC, (2) GT, (3) air compressor, (4) recuperator, (5) combustor, and (6) power turbine. The ambient air is supplied to the cycle at state 1 and it is then pressurized by the compressor and exits from state 2. To enhance the system efficiency, the exhaust gases from the GT power plant are utilized to preheat the air before entering the cathode side of the SOFC stack using a recuperator as shown in Fig. 16. The methane fuel (CH4) is also supplied to the SOFC, and the electrochemical reaction takes place generating both electricity and heat. However, a portion of the methane that is supplied to the SOFC will not be oxidized, and one good solution is to provide it to the combustor to be burned along with the direct main CH4 stream that is supplied to the combustor; the generated exhaust gases from the SOFC will be sent to the combustor to be heated. The generated exhaust gases from the combustor will be directed toward the GT to convert the heat of the exhaust gases into mechanical work providing the air compressor with the necessary energy to increase the inlet air pressure as illustrated in Fig. 16. The flue gases will be expanded in the GT resulting in a reduction in its temperature, the flue gases exit the GT at state 6 and enter a power turbine to go through another expansion process until they attain the atmospheric pressure using the residual thermal energy of the flue gases to produce an electric power through the combination of the power turbine and the electric generator. The thermal energy of the flue gases will be used in the recuperator at comparatively low compression ratios, leading to an upsurge in the temperature of the ambient air supplied to the SOFC. Finally, the flue gases are emitted to the surrounding environment at state 8, respectively.

332

Combined Energy Conversion Systems

Water intake

Anode

Cathode Cathode

Combustion chamber Compressor

Generator

Gas turbine

Fig. 15 The Brayton–fuel cell cycle.

8

Recuperator

7 Inverter

3 SOFC

Air

Ignitor

Combustor

Fuel (CH4)

2

Compressor

5

6

Gas turbine

1 Air

Power turbine

Fig. 16 Schematic diagram of the proposed combined gas turbine power plant with a solid oxide fuel cell (SOFC).

Generator

Combined Energy Conversion Systems

333

4.9.2.6.1.2 Thermodynamic analysis The thermodynamic assessment of every part of the system introduced in the previous section will be provided in this section. The overall cycle is assumed to be running steadily, and all the gases in the system are treated as an ideal gas at the different state points displayed in Fig. 16. 4.9.2.6.1.3 Compressor The compressor isentropic efficiency can be expressed as follows: Wcos h2s h1 ¼ ð55Þ Wcoa h2 h1 and the ideal temperature of the working fluid leaving the compressor can be identified utilizing the subsequent equation:  ϒ ϒ 1 T2s P2 ¼ ð56Þ T1 P1 Zco ¼

The necessary work to operate the compressor to generate a compression ratio of rp can be expressed as follows: _ co ¼ m _ 1 ðh2 W

h1 Þ

ð57Þ

The entropy generation of the compressor can be determined using the following equation: _ 2 s2 þ S_ gen;c ¼ 0 m

_ 1 s1 m

ð58Þ

4.9.2.6.1.4 Recuperator The recuperator effectiveness can be described as follows: erec ¼

T3 T7

T2 T2

ð59Þ

The exit temperature of the introduced cycle can be obtained as follows: _ 2 ðh3 m

_ 7 ðh7 h2 Þ ¼ m

h8 Þ

ð60Þ

Moreover, the entropy generation of the recuperator can be obtained by the following equation: _ 7 s7 _ 2 s2 þ m m

_ 3 s3 m

_ 8 s8 þ S_ gen;rec ¼ 0 m

ð61Þ

4.9.2.6.1.5 Solid oxide fuel cell The electrochemical reactions that occur in the anode and cathode sides of the SOFC can be written as follows [13]: Anode: H2 þ O ¼ -H2 O þ 2e CO þ O ¼ -CO2 þ 2e CH4 þ 4O ¼ -2H2 O þ CO2 þ 8e Cathode: 1 O2 þ 2e -O ¼ 2 The overall cell reaction then can be determined as: CH4 þ 2O2 -CO2 þ 2H2 O The reversible cell voltage E can be determined using Nernst equation as shown below: ! 2 PCH4 PO RT 2 ln E ¼ E1 þ 2 PCO2 PH 8F 2O

ð62Þ

here E is the ideal cell voltage at standard conditions, R is the universal gas constant, T is the temperature of the stack, and F is representing the Faraday constant. The typical ideal cell potential for an intermediate-temperature SOFC running at 8001C is around 0.99 V and for high-temperature SOFC functioning at 1100 the cell potential is around 0.91 V [14]. The current density j is the rate of electron transfer per unit activation area of the fuel cell, the generated electricity from the fuel can be written as follows: _ Fc ¼ Vc jAc W

ð63Þ

here Vc denotes the actual cell potential and it can be calculated by subtracting the voltage losses in the fuel cell from the reversible voltage as follows: Vc ¼ E

Vloss

ð64Þ

334

Combined Energy Conversion Systems

Vloss is the summation of the potential losses that are occurring due to the irreversibilities in the fuel cell due to the formation of the activation, ohmic and concentration overpotentials and Vloss can be written as follows: Vloss ¼ Vact þ Vohm þ Vconc

ð65Þ

at which,   j Vact ¼ Aln jo Vohm ¼ jr  Bln 1

Vconc ¼

  j j1

The rate of heat produced by the fuel cell stack can be determined using the following equation: _ Fc ¼ IVloss ¼ jAc ðE Q

Vc Þ  10 6 ðkWÞ

ð66Þ

The air required to be supplied to the fuel cell can be calculated as follows:   _ Fc kg W Air usage ¼ 3:57  10 7  l  Vc s

ð67Þ

The mass balance equation of the fuel cell can be written as follows: _ Fc þ m _ 3 h3 þ m _ 4 h4 þ m _ fuel;Fc  ð1 _ fuel;Fc  LHV CH4 ¼ W m

Uf Þ  LHV CH4

ð68Þ

_ 4 is representing the mass flow rate of the air that is supplied but not here, LHV is representing the low heat value of CH4 and m used in the SOFC and the gases resulting from the electrochemical reaction inside the SOFC. 4.9.2.6.1.6 Combustor Mass balance of the combustor can be written as follows: _4þm _ fuel;Fc  ð1 m

_ fuel;comb ¼ m _5 Uf Þ þ m

ð69Þ

Energy balance equation of the combustor can be written as follows: _ comb ¼ m _ loss _ 5 h5 þ Q _ 4 h4 þ Q m

ð70Þ

where _ comb ¼ m _ fuel;Fc  ð1 Q _ loss ¼ m _ fuel;Fc  ð1 Q

_ fuel;comb  LHV Uf Þ  LHV þ m

_ fuel;comb  LHV  ð1 Uf Þ  LHV þ m

Zcomb Þ  LHV

ð71Þ ð72Þ

Entropy generation rate of the combustor can be calculated as follows: _ 5 s5 S_ gen;comb ¼ m

_ 4 s4 m

_ Þfuel;comb þ ðms

_ loss Q Tsurr

_ comb Q Tcomb

ð73Þ

4.9.2.6.1.7 Gas turbine The work needed to run the compressor will be supplied by the GT, therefore: _ GT ¼ W _ co W The turbine exit temperature T6 can be determined by knowing the TIT. Moreover, applying the definition of turbine isentropic efficiency will enable us to determine the ideal temperature of the working gas at the turbine exit as follows: ZGT ¼

WGTa h5 ¼ WGTs h5

h6 h6s

The downstream pressure of the GT can be expressed using the following equation:  ϒ ϒ 1 T6s P6 ¼ P5 T5

ð74Þ

ð75Þ

The entropy generation rate of the turbine can be calculated as follows: _ 5 ðs6 S_ gen;GT ¼ m

s5 Þ

ð76Þ

Combined Energy Conversion Systems

Table 2

335

Main operating parameters of the solid oxide fuel cell (SOFC) and a gas turbine (GT–SOFC) combined plant

Parameters

Value (%)

Parameters

Value

Compressor efficiency (Zco) Gas turbine efficiency (ZGT) Power turbine efficiency (ZPT) Recuperator effectiveness (erec) Combustor efficiency (Zcomb) AC generator efficiency (ZGe) Recuperator gas/air sides pressure losses (%) Fuel cell stack pressure losses Combustor pressure losses

81 84 89 80 98 95 4 4 5

Air utilization factor (Ua) Fuel utilization factor (UF) Steam-to-carbon ratio (STCR) Stack temperature (Tstack) (K) Current density (A/cm2) DC–AC inverter efficiency (Zinvert) Cell area (cm2) Ambient temperature (K) Ambient pressure (atm)

0.25 0.85 2.5 1237.15 0.3 89% 834 288 1

Source: Reproduced from Haseli Y, Dincer I, Naterer GF. Thermodynamic modeling of a gas turbine cycle combined with a solid oxide fuel cell. Int J Hydrogen Energy 2008;33.5811–22.doi:10.1016/j.ijhydene.2008.05.036.

Table 3 Modeling results of the introduced combined system at compression ratio of 4 and turbine inlet temperature of 1250K and fuel cell current density of 0.3 A/cm2 Parameters

Values

Thermal efficiency of the plant Specific power to drive compressor Specific power from generator Specific power from solid oxide fuel cell (SOFC) Total specific power produced Net power Air mass flow rate Mass flow rate of fuel to the combustor Mass flow rate of fuel to the fuel cell CO2 emissions

60.55% 175.7 kJ/kg 146.4 kJ/kg 437.5 kJ/kg 583.9 kJ/kg 2419.3 kW 4.123 kg/s 62.1 kg/h 225.3 kg/h 325.8 kg/MWh

Source: Reproduced from Haseli Y, Dincer I, Naterer GF. Thermodynamic modeling of a gas turbine cycle combined with a solid oxide fuel cell. Int J Hydrogen Energy 2008;33.5811–22. doi:10.1016/j.ijhydene.2008.05.036.

4.9.2.6.1.8 Power turbine Power turbine downstream temperature T7 can be expressed as follows: ZPT ¼

WPTa h6 ¼ WPTs h6

h7 h7s

ð77Þ

The working fluid ideal temperature at turbine exit can be determined as follows: T7s ¼ T6

 ϒ ϒ 1 P7 P6

ð78Þ

The amount of work that can be transferred to the generator is expressed as below: _ PT ¼ m _ 6 ðh6 W

h7 Þ

ð79Þ

The entropy generation of the power turbine can be determined as: _ 6 ðh6 S_ gen;PT ¼ m

h7 Þ

ð80Þ

The exergy destruction rate of any component can be determined by multiplying its entropy generation rate with the ambient temperature. 4.9.2.6.1.9 Results and discussion Operational conditions that are used in modeling the introduced combined system can be found in Table 2. The modeling results are also highlighted in Table 3. The effect of varying the compression ratio rp on the exergy destruction rate and thermal efficiency of a GT plant (conventional system) and a SOFC combined with GT (system proposed in this case study) at the same operating

336

Combined Energy Conversion Systems

1800 Conventional GT plant Exergy destruction rate (kW)

GT−SOFC plant 1600

TIT = 1250 K

1400

1200

1000

800

4

6

8 Compression ratio rp

10

12

Fig. 17 Exergy destruction rate variation with compression ratio for a gas turbine (GT) plant (conventional plant) and with combined solid oxide fuel cell (SOFC) and a gas turbine (GT–SOFC) plant under the same operating conditions.

0.7

Efficiency

0.6

0.5

0.4

Conv,GT

GT−SOFC

Conv,GT

GT−SOFC

TIT= 1250K

0.3

0.2 4

6

8 10 Compression ratio rp

12

14

Fig. 18 Thermal efficiency variation with compression ratios for a gas turbine plant (conventional plant) and a combined solid oxide fuel cell (SOFC) and a gas turbine (GT–SOFC) plant under the same operating conditions.

conditions will be represented. Moreover, the exergy destruction rate of the different component of the introduced combined system will be shown accordingly. The overall rate of irreversibility represented in exergy destruction rate of the GT–SOFC plant is found to be larger than that of the conventional plant (nearly with a uniform difference). Varying the compression ratio rp from 4 to 12 increased the exergy destruction rate of the GT standalone system from 1011 to 1312 kW; the same increase in the compression ratio led to an upsurge in the exergy destruction rate of the proposed combined SOFC–GT system from 1358 to 1699 kW as shown in Fig. 17. Regardless of the high exergy destruction rate occurred in the GT–SOFC plant, it can be observed from Fig. 17 that it provides better performance compared to the GT standalone system. This can be interpreted by the high temperature of operation of the fuel cell, which permits preheating of the working fluid before supplying it to the combustor. Thus, a significant amount of power generation. Both factors result in an improvement in the energetic efficiency of the combined system compared to the conventional plant. Furthermore, it can be observed from Fig. 18 that the maximum efficiency in the case of the GT plant takes place at a compression ratio of 6, which means that the power required to be supplied to the compressor in the GT plant is larger, considering that both cycles are functioning at their optimum point, therefore having a negative role regarding efficiency of the cycle, compared to the GT–SOFC cycle. Fig. 18 is also showing the GT–SOFC power plant has 27.8% better energy efficiency on an

Combined Energy Conversion Systems

337

440 396.22 400

370.75

Exergy destruction rate (kW)

360 320 280 240 200 160 111.32

120

87.73

80 33.96

40

13.2

0 Combustor

Compressor

Power turbine

SOFC

Recuperator

Gas turbine

Fig. 19 Exergy destruction rate of different components in the proposed combined solid oxide fuel cell (SOFC) and a gas turbine (GT–SOFC) system.

average basis than a conventional GT plant. Moreover, the exergy efficiency of the combined SOFC–GT plant declined from 57.86% to 53.64% with increasing the compression ratio from 4 to 12 and the exergy efficiency of the GT plant also witnessed a mitigation from 32.52% to 27.86% when increasing the compression ratio from 6 to 12. Additionally, from Fig. 18 we can obtain that the GT–SOFC power plant has 26.6% better exergetic performance on an average basis compared to a conventional GT plant. Calculating the exergy destruction rate of each component in the proposed combined system helps to identify the main components that are responsible for the largest irreversibilities in the plant. Fig. 19 is showing the exergy destruction rate of the main parts in the combined SOFC–GT system at which the combustor possessed the highest rank with 396.22 kW of exergy destruction rates followed by SOFC with 370.75 kW and recuperator with 111.32 kW, GT with 33.96 kW, and finally the power turbine with 13.2 kW.

4.9.2.6.1.10 Final remarks In this study, a combined SOFC with a recuperative GT cycle was assessed. A thermodynamic simulation has been carried out by utilizing the first and the second law approach for each part in the system, along with the whole cycle as a lumped control volume. The model was also validated with results available in the literature; for further reading see Ref. [13]. From the obtained results, the thermal efficiency of the GT–SOFC plant is mitigated by augmenting the TIT. This mitigation is also obvious from the compression ratio, after its maximum value when the cycle attains its optimal performance. Any upsurge in the TIT and the compression ratio results in a higher rate of exergy destruction of the plant. However, increasing the TIT enhances the specific power obtained from the cycle. It is also found that the combustor and SOFC have a high contribution to the exergy destruction rate of the plant with values of 396.22 and 370.75 kW, respectively. Furthermore, the GT–SOFC power plant is found to be more efficient energetically and exergetically compared to a traditional GT plant, with around 27.8% higher energy efficiency and around 26.6% higher exergy efficiency. Furthermore, the GT–SOFC combined system is found to be environmentally friendly since it emits carbon dioxide emissions of 325.8 kg/MWh, which is much lower than the GT system, which emits carbon dioxide emissions exceeding 600 kg/ MWh.

4.9.2.7

The Chemical Recuperation Cycle

The chemical recuperated GTC converts CH4, H2O, and occasionally CO2 into an H2 and CO2 fuel mixture that can be burned in the combustion chamber by using a reforming process. This process increases heating values of the fuel because the endothermic reaction absorbs the heat at a temperature, which is below the burning temperature. Recuperation method, which consists both thermal and chemical, has greater WHR rate than the conventional recuperation system. Furthermore, the H2 rich fuel is more flammable than methane and allows ignition at a lesser flame temperature that theoretically decreases NOX formation. Conversely, the GT flue gas temperature is not sufficient for a whole reforming reaction. At 823K, only 20% of the entire fuel is reformed. To raise the temperature, some other combustion processes can be used [5].

338

Combined Energy Conversion Systems

4.9.2.8

The Rankine–Rankine Cycle

Two Rankine cycles can be combined by a steam–water HEX as shown in Fig. 20. The exhaust stream of the main Rankine cycle passes through a HEX where it supplies the required heat to the working fluid of the second Rankine cycle. In addition, it works as a condenser for the main chamber. The working fluid is expanded in the second turbine to produce more power. In the following example, a solar-driven Rankine–Rankine cycle is analyzed thermodynamically. More about the case study can be found in “Analysis and performance assessment of a new solar-based multigeneration system integrated with ammonia fuel cell and solid oxide fuel cell–gas turbine combined cycle,” written by Siddiqui and Dincer [15].

Electricity

HPT

2

LPT 5

3 Heat exchanger 1

8 Heat exchanger 2

4

Steam turbine

1 6

Electricity

9

7 10

Con

Fig. 20 A schematic representation of a solar-driven combined cycle. HPT, high-pressure turbine; HX; LPT, low-pressure steam; ST2.

4.9.2.8.1

Case Study 4

In this case study, a combined system driven by solar energy, which consists of two SRCs (namely primary and secondary), is presented, analyzed, and discussed accordingly. It will also show how combined mode helps improve performance of the system. 4.9.2.8.1.1 System description In the present study, a solar tower based integrated system is designed and thermodynamically analyzed. The schematic of the system is shown in Fig. 20. The system comprises of a concentrated solar tower based integrated system with a primary and secondary Rankine cycle to generate electricity. The waste heat of the primary Rankine cycle is utilized by the secondary Rankine cycle to generate electricity. The central solar tower receives the solar energy, which is absorbed by the molten salt and transferred through HEX 1 to the primary Rankine cycle. Water enters HEX 1 and absorbs heat from the molten salt. After leaving HEX 1, steam enters the high-pressure turbine (HPT) at high temperature and high-pressure. The steam leaving the HPT is reheated in HEX 1 before entering the low-pressure turbine (LPT). After leaving the LPT, the steam is passed through HEX 2, where the waste heat is transferred to the secondary Rankine cycle. Water in the secondary Rankine cycle enters HEX 2 at state 7, where it absorbs heat from the exhaust of the primary Rankine cycle. Steam at high temperature and pressure leaves HEX 2 at state 8 and enters the ST of the secondary Rankine cycle. In the proposed system, the efficiency of a conventional single Rankine cycle solar-based system is shown to increase by integrating with a secondary Rankine cycle by utilizing the waste heat from the primary Rankine cycle. In order to assess the system performance, energy and exergy analyses are conducted to determine the efficiencies of the proposed combined system and compare them to a system utilizing only the primary Rankine cycle for power generation. By utilizing the secondary Rankine cycle, higher power outputs and hence higher overall system energy and exergy efficiencies can be obtained. 4.9.2.8.1.2 Analyses In order to assess the system performance, energy and exergy analyses are conducted. The overall system, as well as each system component is analyzed using thermodynamic approaches of energy and exergy. The overall system energy and exergy efficiency of the proposed integrated system are determined and compared to a system comprising of only the primary Rankine cycle. The effects of changing ambient temperature, turbine inlet pressure, beam radiation intensity, and TIT on the overall system energy and exergy efficiencies are analyzed and discussed, respectively. The solar energy received by the solar tower is expressed as _ solar ¼ ¼ I_ b Ahf Zhf Q

ð81Þ

where I_ b denotes the beam radiation, Ahf denotes the area of the heliostat field, and heliostat field efficiency is denoted by Zhf, considered in this study as 80%.

Combined Energy Conversion Systems

339

Here, both energy and exergy analyses are performed on each system component of the proposed solar tower based integrated system to determine their performance and to obtain the overall system efficiency. The reference environment temperature is taken as T0 ¼ 298K and the reference pressure P0 ¼ 100 kPa. Some quick assumptions utilized to facilitate the analysis are listed as follows:

• • • • • •

The The The The The The

pumps and turbines are adiabatic. kinetic energy changes are negligible. potential energy changes are negligible. pumps and turbines have an isentropic efficiency of 80%. steady-state conditions exist. pressure losses are negligible.

The principle of conservation of mass for a general given control volume can be applied to denote the general mass balance equation: X X dmcv _ _ ¼ ð82Þ m m i i e e dt

The first law of thermodynamics can be applied to obtain the energy balance equation for a general control volume on a rate basis:  X   X  V2 V2 dE _ W _ þ _ i hi þ i þ gZi _ e he þ e þ gZe ¼ cv Q ð83Þ m m i e 2 2 dt

The entropy is generated during a process due to irreversibilities. The rate of entropy generation for a control volume is expressed as follows: dScv X _ s S_ gen ¼ þ m e e e dt

X

i

_ i si m

X Q _k k T k

The exergy balance equation for a given control volume is expressed as: X X _ Qþ _ w þ Ex _ d _ ex ¼ _ ex þ Ex Ex m m i i i e e e

ð84Þ

ð85Þ

In this study, a steady-state steady-flow type analysis of the system components is conducted. The rate balance equations are applied on each component as follows: Pump 1: the mass balance per unit time for pump 1 can be denoted as follows: _6¼m _1 m

ð86Þ

The energy balance on a rate basis for pump 1 can be expressed as follows: _ p1 ¼ m _ 1 h1 _ 6 h6 þ W m

ð87Þ

The entropy balance on a rate basis for pump 1 is expressed as follows: _ 6 s6 þ S_ gen;p1 ¼ m _ 1 s1 m

ð88Þ

The exergy balance on a rate basis for pump 1 can be expressed as follows: _ p1 ¼ m _ d;p1 _ 1 ex 1 þ Ex _ 6 ex 6 þ W m

ð89Þ

HEX 1: the mass balance per unit time for HEX 1 is expressed as follows: _ st;h ¼ m _ st;c and m _1¼m _2 ¼m _3¼m _4 m

ð90Þ

The energy balance per unit time for HEX 1 is expressed as follows: _ 3 h3 þ m _ st;h hst;h ¼ m _ 2 h2 þ m _ 4 h4 þ m _ st;c hst;c _ 1 h1 þ m m

ð91Þ

The entropy balance per unit time for HEX 1 is expressed as follows: _ 1 s1 þ m _ 3 s3 þ m _ st;h sst;h þ S_ gen;HX1 ¼ m _ 2 s2 þ m _ 4 s4 þ m _ st;c sst;c m

ð92Þ

The exergy balance per unit time for HEX 1 is expressed as follows: _ d;HX1 _ 3 ex 3 þ m _ st;h ex st;h ¼ m _ 2 ex 2 þ m _ 4 ex 4 þ m _ st;c exst;c þ Ex _ 1 ex1 þ m m

ð93Þ

HPT: mass balance on a rate basis for the primary turbine is denoted as follows: _3 _2¼m m

ð94Þ

The energy balance on a rate basis for the HPT is denoted as follows: _ HPT þ m _ 3 h3 _ 2 h2 ¼ W m

ð95Þ

The entropy balance on a rate basis for the HPT is denoted as follows: _ 2 s2 þ S_ gen;HPT ¼ m _ 3 s3 m

ð96Þ

340

Combined Energy Conversion Systems

The exergy balance on a rate basis for the primary turbine is denoted as follows: _ HPT þ m _ d;T1 _ 3 ex 3 þ Ex _ 2 ex 2 ¼ W m

ð97Þ

LPT: The mass balance on a rate basis is written for the LPT as follows: _4¼m _5 m

ð98Þ

The energy balance on a rate basis is written for the LPT as follows: _ LPT þ m _ 5 h5 _ 4 h4 ¼ W m

ð99Þ

The entropy balance on a rate basis is written for the LPT as follows: _ 4 s4 þ S_ gen;LPT ¼ m _ 5 s5 m

ð100Þ

The exergy balance on a rate basis is written for the LPT as follows: _ LPT þ m _ d;LPT _ 5 ex 5 þ Ex _ 4 ex4 ¼ W m

ð101Þ

HEX 2: Mass balance per unit time for HEX 2 is expressed as follows: _ 6 and m _7¼m _8 _5 ¼m m

ð102Þ

The energy balance per unit time for HEX 2 is expressed as follows: _ 5 h5 þ m _ 7 h7 ¼ m _ 6 h6 þ m _ 8 h8 m

ð103Þ

The entropy balance per unit time for HEX 2 can be written as follows: _ 7 s7 þ S_ gen;HX2 ¼ m _ 6 s6 þ m _ 8 s8 _ 5 s5 þ m m

ð104Þ

The exergy balance per unit time for HEX 2 can be written as follows: _ d;HX2 _ 7 ex 7 ¼ m _ 6 ex 6 þ m _ 8 ex 8 þ Ex _ 5 ex 5 þ m m

ð105Þ

ST 2: The mass balance on a rate basis is written for the secondary turbine as follows: _8¼m _9 m

ð106Þ

The energy balance on a rate basis is written for the secondary turbine as follows: _ ST2 þ m _ 9 h9 _ 8 h8 ¼ W m

ð107Þ

The entropy balance on a rate basis is written for the secondary turbine as follows: _ 8 s8 þ S_ gen;ST2 ¼ m _ 9 s9 m

ð108Þ

The exergy balance on a rate basis is written for the secondary turbine as follows: _ ST2 þ m _ d;ST2 _ 9 ex 9 þ Ex _ 8 ex 8 ¼ W m Condenser: The mass balance for the condenser is written on a rate basis as follows: _ 10 _9 ¼m m

ð109Þ

The energy balance for the secondary Rankine cycle condenser is written on a rate basis as follows: _ l;c þ m _ 9 h9 ¼ Q _ 10 h10 m

ð110Þ

The entropy balance for the secondary Rankine cycle condenser is written on a rate basis as follows: _ 10 s10 þ _ 9 s9 þ S_ gen;c ¼ m m

_l Q Tc

The exergy balance for the secondary Rankine cycle condenser is written on a rate basis as follows:   _ l;c 1 To þ Ex _ d;c _ 10 ex 10 þ Q _ 9 ex9 ¼ m m Tc

ð111Þ

ð112Þ

Pump 2: For pump 2, the equation for the mass balance on a rate basis is expressed as follows: _7 _ 10 ¼ m m

ð113Þ

The energy balance for pump 2 on a rate basis is expressed as follows: _ P2 ¼ m _ 7 h7 _ 10 h10 þ W m

ð114Þ

The entropy balance for pump 2 on a rate basis is expressed as follows: _ 7 s7 _ 10 s10 þ S_ gen;P2 ¼ m m

ð115Þ

The exergy balance for pump 2 on a rate basis is expressed as follows: _ P2 ¼ m _ d;P2 _ 7 ex 7 þ Ex _ 10 ex 10 þ W m

ð116Þ

Combined Energy Conversion Systems

341

4.9.2.8.1.3 Energy and exergy efficiencies The energy and exergy efficiencies of the overall system are evaluated as Zen;ov ¼

Zex;ov ¼

_ HPT þ W _ LPT þ W _ ST2 W _ solar Q

ð117Þ

_ _ _ HPT þ W þW W  LPT  ST2 _ solar 1 To Q

ð118Þ

Tsun

_ LPT denotes the power output of the LPT, W _ ST2 denotes _ HPT represents the power output of the high-pressure turbine, W where W _ the power output of the secondary turbine, and Qsolar represents the solar energy input to the integrated system. In addition, the energy and exergy efficiencies of the proposed system in case of utilizing only the primary Rankine cycle can be expressed as: Zen;ov ¼

Zex;ov ¼

_ HPT þ W _ LPT W _ solar Q

ð119Þ

_ LPT _ HPT þ W W   o _ Qsolar 1 TTsun

ð120Þ

4.9.2.8.1.4 Results and discussion In conducting the thermodynamic analysis, the temperature, pressure, enthalpy, entropy, and exergy of each state are determined and tabulated in Table 4. The Engineering Equation Solver (EES) software is utilized to evaluate the thermodynamic properties. The reference ambient conditions considered in this study are a temperature of 251C and a pressure of 101 kPa. The beam radiation value is taken as 850 W/m2. The heliostat mirror dimensions are 11 m  11 m and 400 mirrors are utilized in the proposed system. To assess the system performance, energy and exergy efficiencies are evaluated. The overall integrated system energy efficiency is evaluated as 49.7% and the exergy efficiency of the overall system is obtained as 52.4%. However, the energy efficiency of the system in case of utilizing only the primary Rankine cycle is obtained as 30.4% and the exergy efficiency is obtained as 32%. Hence, a considerable increase in efficiencies is possible by utilizing a combined primary and secondary Rankine cycle system. The effect of changing ambient temperature on the overall system energy and exergy efficiencies is shown in Fig. 21. As can be seen, the exergy efficiency increases with increasing ambient temperature. However, the energy efficiency remains constant. As the ambient temperature increases from 10 to 551C, the exergy efficiency increases from 52.2% to 52.7%, respectively. The effect of the HPT inlet pressure on the overall system energy and exergy efficiency is shown in Fig. 22. As the inlet pressure increases from 500 kPa to 5 MPa, the energy efficiency increases from 48.5% to 51.2%. In addition, the exergy efficiency increases from 51.1% to 54.5%, respectively, for the same pressure increase in the turbine inlet pressure. Hence, the inlet pressure to be used should be chosen according to the required overall system efficiencies, considering the associated costs. Higher pump power would be required to reach the higher turbine inlet pressure, however, the system efficiencies will increase. The effect of beam radiation intensity on the overall system energy and exergy efficiencies is shown in Fig. 23. As can be seen from the figure, the energy efficiency decreases to 32.3% as the beam radiation intensity decreases to 300 W/m2. However, as the beam radiation intensity increases to 1500 W/m2, the energy efficiency increases to 54.8%. In addition to this, the exergy efficiency increases from 34.0% to 57.8% as the beam radiation intensity increases from 300 to 1500 W/m2. This can be attributed to the increase in power output from the system as the amount of solar radiation intensity increases. The effect of high-pressure TIT on overall system energy and exergy efficiencies is depicted in Fig. 24. As can be seen from the figure, the energy and exergy efficiencies are observed to increase with increasing TIT. In addition, Fig. 24 also shows the energy and exergy efficiencies of a solar energy driven system utilizing only a single Rankine cycle for power generation. A Table 4

Input and evaluated thermodynamic data

N

Temperature (1C)

Pressure (kPa)

Mass flow rate (kg/s)

Specific enthalpy (kJ/kg)

Specific entropy (kJ/kg K)

Specific exergy (kJ/kg)

1 2 3 4 5 6 7 8 9 10

64.9 632 545.8 1898 1197 64.9 65.1 874.4 364.5 64.9

1000 1000 550 550 25 25 1200 1200 25 25

5 5 5 5 5 5 6 6 6 6

272.5 3769 3582 7079 5142 272 273.5 4332 3206 272

0.89 8.11 8.17 10.63 10.98 0.893 0.894 8.58 9.07 0.89

11.2 1356 1152 3915 1872 10.2 11.5 1780 506 10.2

342

Combined Energy Conversion Systems

0.530

0.527

en ex

0.525

0.526

0.525

0.515 0.510

0.524

Exergy efficiency

Energy efficiency

0.520

0.505 0.523 0.500 0.495 10

15

20

25

30

35

40

45

50

0.522 55

T0 (°C) Fig. 21 Effect of ambient temperature on overall system energy and exergy efficiencies.

0.55

0.55 ex

en

0.54

Energy efficiency

0.52 0.53 0.51 0.50

Exergy efficiency

0.54

0.53

0.52

0.49 0.48

1000

2000

3000

4000

0.51 5000

High-pressure turbine inlet pressure (kPa) Fig. 22 Effect of high-pressure turbine (HPT) inlet pressure on overall system energy and exergy efficiencies.

46% energy efficiency is obtained at a TIT of 3501C for the combined system, however, an energy efficiency of 28.9% is obtained for a system utilizing only the primary Rankine cycle for power generation. The energy and exergy efficiencies of the combined system increase from 32% and 34%, respectively, at an inlet temperature of 1801C to 54.8% and 57.8%, respectively, at a temperature of 16761C. Furthermore, the energy and exergy efficiencies of the system utilizing only the primary Rankine cycle increase from 16.9% and 17.8%, respectively, at an inlet temperature of 1801C to 31.8% and 33.6%, respectively, at a temperature of 16761C. 4.9.2.8.1.5 Closing remarks In this case study, a solar-driven combined cycle is studied for power generation. In conventional solar-based power plants, single SRCs are utilized. However, the proposed system comprises of combined primary and secondary Rankine cycles. The waste heat of the primary Rankine cycle is utilized by the secondary Rankine cycle. Hence, higher power output and higher efficiencies are obtained accordingly. The energy efficiency of the combined system is evaluated as 49.7% and the exergy efficiency of the combined system is obtained as 52.4%. However, the energy efficiency of the system in case of utilizing only the primary Rankine cycle is obtained as 30.4% and the exergy efficiency is obtained as 32%. Hence, a considerable increase in efficiencies is possible by utilizing a combined primary and secondary Rankine cycle system.

Combined Energy Conversion Systems

en

0.6

ex

0.55

0.55

0.50

0.5

0.45

0.45

0.40

0.4

0.35

0.35

0.30 0.3

0.45

0.6

0.75

0.9

1.05

1.2

1.35

Exergy efficiency

Energy efficiency

0.60

343

0.3 1.5

Beam radiation (kW/m2) Fig. 23 Effect of beam radiation intensity on overall system energy and exergy efficiencies.

0.6

Energy efficiency

0.5

excomb encomb

enPRC 0.5 exPRC

0.4

0.4

0.3

0.3

0.2

0.2

0.1 200

400

600

800

1000

1200

1400

1600

Exergy efficiency

0.6

0.1 1800

Turbine inlet temperature (°C)

Fig. 24 Effect of high-pressure turbine (HPT) inlet temperature on overall system energy and exergy efficiencies.

4.9.2.9

Brayton–Organic Rankine Combined Cycle

The main difference between the traditional SRCs and the ORC is the utilization of different working fluids. In the ORC, the working fluid consists of organic substances, which include various types of refrigerants, mixtures of hydrocarbons, silicon oil, pentane, and ammonia [16]. However, the technology of the ORC is appropriate for small-scale applications. Also, energy systems based on renewable energy resources as well as for applications involving comparatively low temperature WHR, ORC is suitable. A few examples of the low temperature heat sources that can be utilized include solar irradiance, exhaust engine gases, geothermal resources of energy, biomass combustion systems, and ocean thermal energy [2]. As the main difference between the organic and SRCs is the working fluid type, the operating temperatures are also different. However, the process paths for both systems are quite similar. The operating principles of the ORC resemble that of the SRC. A pump is utilized to pump the working fluid in the liquid phase to a boiler. In the boiler, the working fluid is heated and the state changes to vapor. After leaving the boiler, mechanical work is produced from the vapor by passing it through an expansion device. As the working fluid exits the expansion device, the temperature and pressure are lowered. The fluid is then passed through a condenser, where it condenses back to the liquid state [17]. For the cases where the exhaust stream of a cycle is relatively low, utilizing it with an ORC instead of a Rankine cycle becomes more suitable. Hence combining a Brayton cycle with an ORC gives promising outcomes. In the following example, a Brayton–organic Rankine combined cycle is presented to illustrate such systems.

344

Combined Energy Conversion Systems

Q

Compressor 2

3

WGT

WC

Heat exchanger Gas turbine

1

4

Working fluid: air

5 Working fluid: isobutane

6 WEX

Heat exchanger

Expander 7 Condenser

Pump 8 WP

Q Fig. 25 A Brayton–organic Rankine cycle (ORC) combined system.

4.9.2.9.1

Case Study 5

Assume an ideal Brayton–organic Rankine combined cycle as depicted in Fig. 25, in which air goes to the compressor at ambient temperature and pressure (100 kPa and 251C) and is compressed by a compression ratio r¼15. The mass flow rate of the air _ air ¼ 30 kg=s. Compression at the compressor occurs under isentropic conditions. After heat addition at constant pressure, the m temperature of the stream reaches 8001C then hot and pressurized air expands isentropically at the GT and produces useful work. The air enters the HEX and rejects its heat to the working fluid of the ORC. Isobutane enters the HEX (boiler) at 301.5K and 31.9 bar pressure and leaves at 400K. The isobutane exits from the expander at 300K at saturation pressure. The air-cooled condenser condenses the liquid–vapor mixture and then pressurized the saturated liquid to the boiling pressure. By considering the ambient temperature is equal to the reference temperature, calculate the following:

• • •

The mass flow rate of isobutane in the ORC. The net power outputs of both cycles. The energy and exergy efficiencies of the binary cycle and the overall combined plant.

Solution: In order to complete the calculations, we need to make some assumptions as follows:

• • • • •

The system operates on steady-state and steady-flow conditions. Pump and GT work adiabatically. No pressure drop occurs through the HEXs. Kinetic and potential energy effects are negligible. Reference conditions T0 ¼ 298K, P0 ¼101.3 kPa.

Solution: 1. Reference conditions T0 ¼251C, P0 ¼100 kPa State 1 (before the compression): Air at 100 kPa and 251C s ¼ 6:862 ðkJ=kg KÞ

h ¼ 298:6 ðkJ=kgÞ:

Initial conditions equals to conditions at state 1, therefore: ex1 ¼ 0 ðkJ=kgÞ

Combined Energy Conversion Systems

345

State 2 (after the compression at compressor): r¼

P2 ¼ 15-P2 ¼ 15  100 ðkPaÞ ¼ 1500 ðkPaÞ ¼ 1:5 ðMPaÞ P1

From the equation for isentropic compression  k T2 P2 ¼ T1 P1

1=k

¼

  T2 1500 ðkPaÞ 1:4 ¼ 100 ðkPaÞ 298:15K

1=1:4

T2 ¼ 646:3K

Under ideal gas assumption at T¼ 643.3K specific enthalpy of the air h2 ¼ 656.3 kJ/kg Since the specific entropy of the stream remains constant at isentropic processes s1 ¼ s2 ¼ 6:862 ðkJ=kg KÞ Compression work for unit basis analysis wc ¼ h 2

h1 ¼ 357:7ðkJ=kgÞ

Specific exergy ex 2 ¼ ðh2 ex 2 ¼ ð656:3 ðkJ=kgÞÞ

h0

298:6 ðkJ=kgÞ

T0  ðs2

s0 ÞÞ-ex2

298:2 ðKÞ  ð6:862 ðkJ=kg KÞ

6:862 ðkJ=kg KÞÞ

State 3 (heat addition at constant pressure): Temperature of the stream reaches 8001C after heat addition T3 ¼ 1073.15K. Heat addition process occurs at constant pressure P3 ¼ P2 ¼ 1500 kPa. Under ideal gas assumption at T3 ¼1073.15K specific enthalpy of the air h3 ¼1130 (kJ/kg). Air at 1500 kPa and 1073.15K, s3 ¼ 7.439 (kJ/kg K). Heat addition at unit basis qsr ¼ cp ðT3

T2 Þ ¼ h3

h2 ¼ 474 ðkJ=kgÞ

Specific exergy ex 3 ¼ ðh3 ex3 ¼ ð1130 ðkj=kgÞ

298:6 ðkJ=kgÞ

h0

T0  ðs3

s0 ÞÞ-ex3

298:2 ðKÞ  ð7:439 ðkJ=kg KÞ

6:862 ðkj=kg KÞÞ ¼ 660 ðkJ=kgÞ

Here heat addition is assumed done by a HEX. Since no chemical reaction takes place, chemical exergies can be neglected for the system. State 4 (expansion at the turbine): Air expands to pressure level it has at state 1 P1 ¼ P4 ¼ 100 kPa From the equation which shows the temperature and pressure relation for isentropic expansion  k T3 P3 ¼ T4 P4

1=k

¼

  1073:15 ðKÞ 1500 ðkPaÞ ð1:4 ¼ T4 100 ðkPaÞ

1=1:4Þ

¼ T4 ¼ 495 ðKÞ

Under ideal gas assumption at T4 ¼495K specific enthalpy of the air h4 ¼498.3 kJ/kg. Since the specific entropy of the stream remains constant at isentropic processes

s3 ¼ s4 ¼ 7:434 ðkJ=kg KÞ

346

Combined Energy Conversion Systems

Specific exergy ex 4 ¼ ðh4 ex4 ¼ ð498:3 ðkJ=kgÞ

298:6 ðkJ=kgÞ

T0  ðs4

h0

s0 ÞÞ-ex3

298:2 ðKÞ  ð7:439 ðkJ=kg KÞ

6:862 ðkJ=kg KÞÞ ¼ 27:9 ðkJ=kgÞ

Heat inlet at unit basis qsr ¼ h3

qsr ¼ 1130 ðkJ=kgÞ

h2

656:3 ðkJ=kgÞ ¼ 474 ðkJ=kgÞ

Heat inlet to the cycle _ sr ¼ qsr m _ air ¼ 474 ðkJ=kgÞ 30 ðkg=sÞ ¼ 14221 ðkWÞ Q Heat rejection at the HEX for unit basis analysis qrej ¼ h4

h1 ¼ cp ðT4

T1 Þ ¼ 498:3 ðkJ=kgÞ

298:6 ðkJ=kgÞ ¼ 199:7 ðkJ=kgÞ

The specific work done by the GT wt ¼ h3

h4 ¼ 1130 ðkJ=kgÞ

498:3 ðkJ=kgÞ ¼ 632:1 ðkJ=kgÞ

Specific network done by the Brayton cycle wnet;B ¼ wt

wc ¼ 632:1ðkJ=kgÞ

357:7ðkJ=kgÞ ¼ 274:4ðkJ=kgÞ

Network done by the Brayton cycle _ net;B ¼ wnet m _ air ¼ 274:4 ðkJ=kgÞ 30 ðkg=sÞ ¼ 8231 ðkWÞ W Specific enthalpy and entropy of isobutane at 301.5K temperature and 31.9 bar pressure h5 ¼ 269:1 ðkJ=kgÞ

s5 ¼ 1:222 ðkJ=kg KÞ

Specific enthalpy and entropy of isobutane at 400K temperature and 31.9 bar pressure h6 ¼ 622:2 ðkJ=kgÞ s6 ¼ 2:203 ðkJ=kg KÞ From the energy balance _ rej ¼ m _ air ðh4 Q

_ isobutane ðh6 h1 Þ ¼ m

30 ðkg=sÞð498:3 ðkJ=kgÞ _ isobutane ð622:2 ðkJ=kgÞ ¼m

h5 Þ

298:6 ðkJ=kgÞÞ ¼ 5990 kW

_ isobutane ¼ 16:97 kg=s 269:1 ðkJ=kgÞÞ-m

Specific enthalpy of isobutane at 300K temperature and specific entropy s7 ¼ s6 ¼2.203 (kJ/kg K) h7 ¼ 558:3 ðkJ=kgÞ Expander work _ e ¼ m6 ðh6 W

h7 Þ ¼ 16:97 ðkg=sÞð622:2 ðkJ=kgÞ

558:3 ðkJ=kgÞÞ

_ e ¼ 1084 ðkWÞ W Specific enthalpy and entropy of isobutane at 300K temperature and in saturated liquid form x ¼ 0 h8 ¼ 264 ðkJ=kgÞ

s8 ¼ 1:222 ðkJ=kg KÞ

Combined Energy Conversion Systems

Table 5

347

Thermodynamic properties at each state point for the combined plant obtained from the Engineering Equation Solver (EES) software

State

Fluid/phase

T (K)

P (kPa)

m_ (kg/s)

h_ (kJ/kg)

s_ (kJ/kg K)

E_ x (kW)

x[

0 00 1 2 3 4 5 6 7 8

Air Isobutane Air Air Air Air Isobutene/liquid Isobutene/superheated vapor Isobutene/liquid–vapor mixture Isobutene/saturated liquid

298.2 298.2 100 1500 1500 100 3190 3190 369.7 369.7

100 100 30 30 30 30 16.97 16.97 16.97 16.97

– – 298.6 656.3 1130 498.3 269.1 622.2 558.3 264

298.6 599 6.862 6.862 7.439 7.439 1.222 2.203 2.203 1.222

6.862 2.515 0 10,731 19,799 837.1 946.2 1974 889.9 859.1

– – 100 1500 1500 100 3190 3190 369.7 369.7

– – – – – – – – 0.9 0

]

Isobutane pump work _ p ¼ m5 ðh5 W

h8 Þ ¼ 16:97 ðkg=sÞð269:1 ðkJ=kgÞ

264 ðkJ=kgÞÞ

_ p ¼ 87:05 kW W Net produced work by the ORC _e _ net;ORC ¼ W W

_ p ¼ 1083:84 ðkW Þ W

87:05 ðkW Þ ¼ 996:8 ðkWÞ

2. Overall network of the combined cycle _ net;ORC ¼ 8231 ðkW Þ þ 996:8 ðkWÞ ¼ 9228 ðkWÞ _ net;B þ W W Overall efficiency of the system can be defined as Z¼

_ net 9228 ðkWÞ W ¼ ¼ 0:6489-%64:89 _ sr 14221 ðkWÞ Q

3. Overall exergy efficiency c1 ¼ 1

1

Z   ¼1 T0 T3

1

0:6489   ¼ 0:8985-%89:85 298:2ðKÞ 1073ðKÞ

Thermodynamic properties at each state point for the combined plant are tabulated in Table 5.

4.9.2.10

Integrated Gasification Combined Cycle

The emissions of fossil fuels have led to environmental concerns to the point where coal use for power production has been threatened. Parallel to the installation of flue gas scrubbers in conventional coal-fired power plants, development of the integrated gasification combined cycle is proceeding on groundbreaking power plant models, which are not only more suitable from an environmental aspect but also feature higher efficiency. In order to achieve the combined cycle, first, the coal needs to be converted to a gaseous fuel via a gasifier [18]. The simple schematic of an integrated gasification combined cycle is illustrated in Fig. 26. In the gasifier, gasification is attained as the result of the controlled combustion of coal/biomass with oxygen and steam and syngas (synthesis gas) and solid waste is produced. Oxygen is supplied to the gasifier by the companion air separation unit. The syngas produced by the gasifier is mainly composed of CO and H2. After the syngas is formed, it is cleaned of pollutants and later on combusted in the combustion chamber and drives the GT. In integrated gasification combined cycles, contaminants as sulfur and mercury are removed before the combustion process, unlike the conventional coal power cycles. Even though integrated gasification combined cycles emit less SO2, NO, Hg, and particulate emissions than similar conventional coal plants, abundant solid waste is still needed to be managed carefully [16]. Likewise, to the Brayton–Rankine combined cycle, the exhaust heat of the cycle supplies the heat for the steam generator of the bottoming cycle consequently the ST runs and drives the generator. The amount of H2O going to gasifier from the Rankine cycle is compensated by the feed water intake, which takes place after the condenser.

4.9.2.11

Thermoelectric Generator Combined Cycles

The thermoelectric effect is a phenomenon found in the 19th century that provides that the direct conversion of the temperature difference to an electrical potential and vice versa. Peltier effect creates a temperature gradient by using electrical current where the

348

Combined Energy Conversion Systems

Particulate Sulfur Mercury removal

Gas cleaning Compressor

Combustion chamber Gas turbine

Generator

Discharge

Steam generator (HEX)

Coal

Steam turbine Pump

Generator

O2

Air separation unit

Compressor

Air intake Ash

Depleted O2 air Condenser

Feed water Fig. 26 Integrated gasification combined cycle.

Seebeck effect generates an electrical potential between cold and hot heat sources. TEGs appear to be a suitable option for WHR systems. Due to their advantages such as having no moving parts, with low maintenance cost, no operational cost, and zero emission of greenhouse gases (GHGs), thermoelectric materials (TEMs) excite the attention of researchers. Despite all the virtues of TEGs, the commercial readiness level of these devices has not reached the anticipated level yet. The key reason behind it is the relatively poor efficiencies they have. Moreover, as the thermoelectricity is still a developing technology, the prices of TEGs are higher than the conventional power generation units. Despite the stated drawbacks, they can attract people’s attention because of the virtues as mentioned above they have. Due to their simple structure, they can be easily combined with any system where a temperature gradient exists. They can be combined with latent heat storage systems [19], GTs [20], or with a gasoline engine [21,22]. In the following case study, a gasoline engine–TEG combined system thermodynamically analyzed.

4.9.2.11.1

Case Study 6

In this case study, a shell and tube HEX whose pipes are covered by TEGs is situated at the exit of an exhaust manifold of an automobile. TEG unit consists of a shell and tube HEX whose pipes are covered with TEMs. The exhaust gas of the gasoline engine supplies the heat for the TEG and incoming air from the front grill of the automobile maintains the temperature difference between the hot and cold junctions by rejecting heat from the TEG system. The main difference of the proposed system from the other systems presented in the literature systems that the suggested system has a passive cooling system, which does not require a pump for the cooling process of the TEG. A nozzle is situated just after the front grill in order to accelerate the inlet stream. Highspeed air flows through the nozzle and goes into the TEG unit. The temperature difference between exhaust and air streams in the TEG unit induces the electrical potential. A numerical study is performed by COMSOL, a multiphysics software package [23] in order to model the heat and mass transfer through the TEG unit. 4.9.2.11.1.1 System description A combined TEG–Otto cycle system is suggested. A sketch of the system is shown in Fig. 27. A shell and tube HEX is placed after the exhaust manifold of the automobile. In the suggested system TEGs cover the tubes of the shell and tube HEX. The hot and cold streams pass through the TEG unit as shown in Fig. 27. While the high-temperature exhaust gases flow inside the pipes, the relatively cool air enters the TEG unit from the contrary side and flows over the tubes. A nozzle is located at the front grill in order to accelerate and direct the air to the TEG. After the hot exhaust stream flows through the TEG, it enters to the catalytic converter in order to reduce the pollutants. After the cold stream passes the thermoelectric unit, it is discharged without further usage. Two

Combined Energy Conversion Systems

349

Exhaust manifold Front grill

Pipes covered by TEGs

To catalytic converter

Discharged

Fig. 27 Sketch of the proposed system. TEGs, thermoelectric generators. Adapted from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/j.ijheatmasstransfer.2016.12.102.

QL Heat sink

N

P

Heat source QH Fig. 28 Simple layout of the thermoelectric generator (TEG). Adapted from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/j.ijheatmasstransfer.2016.12.102.

different TEG unit configurations are selected and analyzed in this case study. In the first configuration a HEX with a length of 101 cm and the tube diameter of 3 cm is used. The TEG unit consists of 37 tubes. Similarly, in the second system a configuration the same concept is considered with different dimensions. The second configuration consists of the same amount of tubes with 61 and 1.8 cm diameter. Model of the shell and tube HEX used in this study is a modified version of the shell and tube geometry in the COMSOL application library [23]. 4.9.2.11.1.2 Analysis A TEG consists of two dissimilar types of thermoelectric semiconductor materials, i.e., negatively doped (n-type) and positively doped (p-type). A simple layout of the TEG can be seen in Fig. 28. The thermoelectric legs are connected from their terminals and an electrical current flow in the circuit as a result of the temperature difference between the top and bottom part of the TEM. The performance assessment of a TEG can be done by the definition modified dimensionless figure of merit ZT. The modified figure of merit is a function of three thermoelectric properties of the TEM and temperature. It can be determined as [24]:  2 Sp Sn T ZT ¼  1  1 2 k 2 kn 2 þ spp sn

ð121Þ

where k (W/m K) indicates the thermal conductivity, s (S/m) is the electrical conductivity, and S shows the Seebeck coefficient (V/K), which is an indicator of the magnitude of the induced thermoelectric voltage due to the temperature gradient through the TEM. T is the average temperature of the hot (Th) and cold (Tc) surfaces of the TEG. Subscriptions (p and n) show the doping type of the TEM. To attain better efficiency and thermoelectricity, maximizing S, T, s, and minimizing the k is the main goal for a TEGs.

350

Combined Energy Conversion Systems

Table 6

Design parameters

Parameters

Values

Reference temperature Reference pressure rair Ap An CD CRF ZT kTEG mcar Zmec g HVfuel AFR Diameter of tubes Length of the tubes

293 (K) 101.325 (kPa) 1.204 kg/m3 3.1 (m2) 0.83 ( ) 0.54 ( ) 0.02 ( ) 0.0001 (1/K) 2.2 W/m K 2.121 (tonnes) 15% 9.81 m/s2 43,500 (kJ/kg) 14.6 1.84 and 3.06 (cm) 1.02 and 0.61 (m)

Source: Reproduced from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/j.ijheatmasstransfer.2016.12.102.

The theoretical maximum heat-to-electricity conversion efficiency of the TEG can be defined as [25]: pffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi _ TEG Th Tc 1 þ ZT 1 W pffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi Z¼ ¼ _ TEG Th Q 1 þ ZT Tc

ð122Þ

Th

_ TEG indicates the heat energy absorbed at the hot junction of the TEG. _ TEG shows the electric work generation by TEG and Q here W As it can be seen in Eq. (122), the dimensionless modified figure of merit (ZT) is directly related to the efficiency. In order to enhance the thermoelectric performance, ZT and Th should be maximized, while Tc should be minimized. In order to determine the fluid flow characteristics and the temperature distribution throughout the TEG, a numerical study is performed by COMSOL Multiphysics software [14]. The TEG model used in this study is the modified version of the parameterized shell and tube HEX geometry in the COMSOL application library. The main design parameters are tabulated in Table 6. In order to calculate continuity, momentum, and heat transfer equations of the fluid flow, a numerical study is conducted by the COMSOL Multiphysics [23] software package. The k–e model is selected to simulate turbulent flow. Conservation of momentum and mass equations can be written under steady and incompressible conditions as follows [23]:    2 2 rðu∇Þu ¼ ∇ pI þ ðm þ mT Þ ∇u þ ð∇uÞT ðm þ mT Þð∇uÞ I rkI þ F ð123Þ 3 3

where,

∇ðruÞ ¼ 0   m rðu∇Þk ¼ ∇ m þ T ∇k þ Pk sk   m e rðu∇Þe ¼ ∇ m þ T ∇e þ Cc1 Pk se k   2 Pk ¼ mT ∇u : ∇u þ ð∇uÞT ð∇uÞ2 3 mT ¼ rCm

ð124Þ ð125Þ

re e2 k

ð126Þ

2 rk∇u 3

ð127Þ

Cc2 r

k2 Cc1 ¼ 1:44; Cc2 ¼ 1:92; Cm ¼ 0:09; sk ¼ 1; se ¼ 1:3 e

where u (m/s) represents the velocity vector, P (kPa) shows the pressure, I is the identity tensor, F (N/m3) represents the acting body forces, m (Pa s) indicates the dynamic viscosity, k shows the turbulent kinetic energy (m2/s2), and finally e shows the dissipation rate of kinetic energy (m2/s3). The subscription T indicates the turbulence in the equations. The boundary conditions can be set for the TEG model as follows: us

g

¼ vs

g

¼ ws

g

ð128Þ

Combined Energy Conversion Systems

351

here, s–g indicates the solid–gaseous interfaces. Ts; s ks

g

¼ Tg; s

 ! ∂Ts  s g ¼ ∂n 

g

kg

 ! ∂Tg  s g ∂n 

ð129Þ

Eq. (128) represents the no-slip condition at the gaseous–solid interface where the velocity vectors in all directions are zero. Eq. (129) shows the temperature of the gaseous and solid at the interface should be equal to each other. In this case study, the Otto–thermoelectric combined system is analyzed under various exhaust flow rates. The exhaust flow rate of the automobile is determined based on the overall power requirement of the vehicle. First the acting forces on the vehicle are needed to be calculated in order to determine the power need of the car (Preq). Preq ¼ FD Vcar

ð130Þ

Here, Vcar (m/s) shows the speed of the automobile and FD (N) is the overall resistance force acting on the automobile. In the equation above, no acceleration is taken into account and the vehicle is assumed to be driving at a constant speed. The overall resistance force acting on the automobile can be obtained as follows: 1 r Ap Vcar 2 CD þ mcar gCRF ð131Þ 2 air 1  where, the total resistance force acting on the car is the summation of the total air drag forces acting on the car 2 rair Ap Vcar 2 CD and the rolling resistance caused by tire–road interaction (mcar g CRF). The total amount of heat input to the Otto cycle can be obtained as follows: FD ¼

_ fuel ¼ Preq ¼ Preq Q Zmec 0:15

ð132Þ

_ fuel indicates the amount of heat released after the combustion process of the gasoline. In this case study, it is assumed that where Q 15% of the combustion heat is used for propelling the car [26] and also the combustion efficiency Zc. is assumed to be 1. _ fuel Q ð133Þ HV fuel _ fuel is the mass flow rate of the injected fuel. The mass flow rate of the air where HVfuel is the fuel’s (gasoline) heating value and m introduced in the combustion chamber can be calculated as follows: _ fuel ¼ m

AFR ¼

_ air m _ fuel m

ð134Þ

_ air is the mass flow rate of the air entering the combustion chamber and AFR is the air–fuel (AF) ratio of the where, m mixture. _ in ), the exhaust flow rate of the car is calculated In order to determine the mass flow rate entering the TEG from the hot side (m _ ex Þ. The total mass flow rate of the gasoline can be found simply by using the definition of the AFR. ðm _ in ¼ ð1 þ AFR Þm _ fuel _ ex ¼ m m

ð135Þ

The velocity of the coolant stream of the TEG unit (vg) equals the exit velocity of the nozzle Ai ¼

_ air m Vcar rair;i

Ae ¼

_ air m Vg rair;e

Ae =Ai ¼ An ¼

Vcar rair;i Vg rair;e

ð136Þ

where Ai and Ae show the inlet and exit cross-sectional area of the nozzle and An represents the ratio between them. Based on the equations stated above, for instance, the considered car in this study should burn 0.0622 kg/s gasoline to maintain its speed at 120 km/h steadily. In this case study, overall heat transfer coefficient U (W/m2 K) is also calculated to observe the influence on power generation and efficiency. Overall heat transfer coefficient is determined as follows: _ Q ð137Þ ATEG DTlm _ indicates the total heat transfer rate to the TEG unit; ATEG is the overall heat transfer surface of the TEM, and DTlm is the log here, Q mean temperature difference. For a better performance assessment, exergy analysis of the TEG unit also should be executed. The exergy balance equation of the TEG unit can be defined as follows [27]: X _ Qi Ex _ Wi þ _ di ¼ 0 _ in exin m _ out exout Ex Ex m ð138Þ U¼

_ Qi shows the exergy transfer rate by the heat and Ex _ Wi and Ex _ di are the exergy rate by work and exergy destruction rate by where, Ex the TEG unit, respectively. Also in Eq. (138), exin (kJ/kg) shows the specific exergy of the inlet stream of the TEG and exout (kJ/kg) _ (kg/s) represents the mass flow rate of the indicates the specific exergy of the air stream when it exits from the TEG unit and m working fluids. The specific exergies can be determined as follows [27]: ex i ¼ hi

h0

T0 ðsi

s0 Þ þ exch

ð139Þ

352

Combined Energy Conversion Systems

Fig. 29 Mesh structure of the thermoelectric generator (TEG) unit generator. Adapted from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/j.ijheatmasstransfer.2016.12.102.

Table 7

Inlet parameters for the thermoelectric generator (TEG) unit used in the numerical study

Case #

m_ ex (kg/s)

vg (m/s)

dpipe (cm)

lpipe (cm)

Th,i (K)

Tc,i (K)

1 2 3 4 5 6 7 8 9 10

0.062169 0.050461 0.040456 0.050461 0.040456 0.062169 0.050461 0.040456 0.050461 0.040456

40.00 36.67 33.33 36.67 33.33 40.00 36.67 33.33 36.67 33.33

3.06 3.06 3.06 3.06 3.06 1.84 1.84 1.84 1.84 1.84

102.2 102.2 102.2 102.2 102.2 61.3 61.3 61.3 61.3 61.3

800 783 767 883 867 800 783 767 883 867

293 293 293 293 293 293 293 293 293 293

Source: Reproduced from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/j.ijheatmasstransfer.2016.12.102.

where, indices “i” represent the properties at ith state point and “0” indicates the properties of the substances at the reference condition. In this case study, 101.325 kPa pressure and 293.15K is selected as the reference conditions. Specific chemical exergies are not taken into account, since no chemical reaction take place in the TEG unit. The overall exergy efficiency of the TEG unit can be defined as follows [28]: C¼

_ TEG W _ out _ in mex mex

ð140Þ

The numerical study is performed in the software package COMSOL [14]. The mesh structure of the model is shown in Fig. 29. The mesh frequency is adjusted based on the physics of the study. In order to get more precise results closer to the walls, the number of boundary layers is selected as three. The thickness adjustment factor of the boundary layer is set to be 2.5. Ten different cases are analyzed in the numerical study. The initial values for each case are tabulated in Table 7. 4.9.2.11.1.3 Results and discussion The overall heat transfer rate (U) from the hot stream to the TEG unit and the overall electric power production by the TEG unit are obtained and tabulated in Table 8. As it is given in Table 8, maximum generated power is obtained for the fourth case as 0.28 kW and it is followed by the first case with 0.24 kW. The lowest electric work production rate is obtained for the eighth case as 0.12 kW and it is followed by the third case with a power production of 0.14 kW. The decrease in the exhaust temperature and mass flow rate can be considered as the main reason for the reduction of the power production almost by half. As the temperature of the exhaust stream drops, flow energy of the stream decreases as well. Furthermore, for the third and eighth cases, the mass into the TEG unit has the minimum value, which is directly related with the total energy flow rate to the TEG system. For all the

Combined Energy Conversion Systems

Table 8

353

Results of the numerical study

Case #

m_ ex (kg/s)

vg (m/s)

T h (K)

T c (K)

T TEG (K)

DTlm (K)

U (W/m2 K)

Q_ (W)

W_ TEG (W)

1 2 3 4 5 6 7 8 9 10

0.062169 0.050461 0.040456 0.050461 0.040456 0.062169 0.050461 0.040456 0.050461 0.040456

40.00 36.67 33.33 36.67 33.33 40.00 36.67 33.33 36.67 33.33

657 636 616 705 682 697 676 654 753 730

304 303 302 305 304 316 314 313 319 317

440 432 424 461 452 474 464 454 500 490

359 339 319 406 384 389 369 349 443 421

23.8 20.8 18.1 21.5 18.7 44.7 39.7 34.8 40.8 35.8

30,436 25,092 20,444 31,038 25,454 22,272 18,738 15,538 23,126 19,300

244 192 148 282 221 195 156 123 230 183

2

2

(W/m )

(W/m )

3.46×101 ×101

174

160 3.00 140 2.50 120 2.00

100 80

1.50

60 1.00 40 0.50

(A)

(B)

23.6

3.66

Fig. 30 Power intensity of the thermoelectric generator (TEG) unit ((A) configuration 1 and (B) configuration 2). Adapted from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/ j.ijheatmasstransfer.2016.12.102.

(°C) (°C)

524

527 500 500 450 400

400

350 300

300

250 200

200

150 100

100 50

20

20 (A)

(B)

Fig. 31 Temperature distribution of the thermoelectric generator (TEG) unit ((A) configuration 1 and (B) configuration 2). Adapted from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/j.ijheatmasstransfer.2016.12.102.

354

Combined Energy Conversion Systems

300 Configuration 1 280

Configuration 2

260

∆W = 52.4 W

WTEG (W)

240 ∆W = 49.5 W

220

∆W = 37.7 W

200 180

∆W = 35.7 W

160 140

∆W = 2.52 W

120 100 0

1

2

3

4

5

6

Simulation Fig. 32 Change of generated power by the thermoelectric generator (TEG) with the size of the TEG. Adapted from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/j. ijheatmasstransfer.2016.12.102.

(m/s)

(m/s)

67.3

55.2

50 60

50

40

40 30 30 20 20 10 10

(A)

0.0215

(B)

0.0215

Fig. 33 Streamlines of the thermoelectric generator (TEG) unit ((A) configuration 1 and (B) configuration 2). Adapted from Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017. doi:10.1016/j. ijheatmasstransfer.2016.12.102.

simulations, it is observed that electricity generation by the TEG grows with both the exhaust temperature and exhaust mass flow rate. In this case study, the coolant air mass flow rates vary with the change of fuel consumption of the car and hence speed of the car. The acceleration of the vehicle induces more drag force to act on it. Therefore more fuel is required to maintain the same velocity. As more gasoline goes into the combustion chamber, higher exhaust gas temperature can be obtained in it. Configuration 2 has relatively less thermoelectricity generation compared to configuration 1. On the other hand, as it is shown in Fig. 30, configuration 2 has greater power density than configuration 1. As configuration 2 is more compact and has a smaller structure, the hot stream temperature does not decline as it does in configuration 1 when it leaves the TEG. That condition can be understood more clearly in Fig. 31. The bigger TEG unit encounters more temperature drop on the surfaces for the similar initial conditions.

Combined Energy Conversion Systems

355

As the TEMs are more efficient under greater temperature gradients, greater power density is achieved for the small-scale configuration. Fig. 32 shows the relation between the size of the TEG unit and on electricity production. Five alternative cases are analyzed for the same initial conditions for both configurations. For all simulations, the generated electric work by the TEG increases with the larger TEG unit. The velocity profile of the TEG unit is also given in Fig. 32. Since they have identical geometry, their flow characteristics are quite similar, and they have different average velocities. While average inlet velocity of the exhaust gas obtained 3.4 m/s for configuration 1, it rises to 10.1 m/s for configuration 2. For both cases, vortices can be seen at the inlet of the tubes. Vortices in the TEG system increase the heat transfer rate at the locations they occur. Higher temperature drops also can be seen in Fig. 33 for the same areas as expected. Vortex generators could be considered for further studies to see their effect on overall power generation. To increase the generated power of the TEG unit, water can be selected as a coolant. Since the cp of air is less than the cp of air, there would be less temperature rise on the cold side of the TEG, which causes high thermoelectric efficiency. However, that system would require an external power source and a pump. Moreover, adding a pump to the system will increase the maintenance cost as a result. For further reading, please see Refs. [21,22]. 4.9.2.11.1.4 Conclusions In this case study, an alternative passive system with TEG is analyzed numerically. The effects of heat transfer characteristics on the performance of the gasoline engine–TEG are analyzed and presented. Two different size TEG units are used and compared. The results are calculated for 10 different cases. It is observed that power capacity of the system is directly related to inlet temperature and mass flow rate of the exhaust gas incoming to the system. The following concluding points are listed based on the study:

• • • • • •

It is possible to increase the generated power of the system about 90.6% (from 148 to 282 W) by increasing mass flow rate of exhaust 24% (from 0.0405 to 0.05046 kg/s) and inlet temperature of exhaust by 15% (from 767 to 8831C). The exhaust flow rate and temperature mostly affect the heat transfer of the TEG unit. As the size of the TEG unit decreases by 40% (from 102 to 61 cm), power intensity of the system can be increased from 79 to 180 W/m2 (226%). Since the cross-sectional area of the pipes decreases, incoming exhaust has higher velocity leading a significant increase of conductive heat transfer coefficient. _ raises from 19.8 to Increasing the size of the system by 66.7% raises the overall heat transfer rate of the system by 33.8%. Q 26.5 kW. Increasing the size of the system by 66.7% decreases the overall heat transfer coefficient of the system by 52.5%. U declines from 39.2 to 20.5 W/m2 K. Vortices at the inlet of the TEG unit increase heat transfer rate considerably.

In closing, thermoelectric devices are environmentally friendly and easy to operate. Despite their low efficiencies, they are still useful tools for recovering heat for individual cases. The exhaust gases of vehicles are unavoidable losses. Since there is a lack of space in the car, it is hard to use the exhaust heat of the vehicle in another cycle. TEGs are a suitable choice for these conditions.

4.9.3

Cascade Refrigeration Cycle

For the refrigeration applications, where the desired temperatures are moderately low and also the temperature range of the process is too large for a conventional single vapor-compression refrigeration cycle, practical solutions are needed to be investigated. A greater temperature range also brings a large pressure variation in the refrigeration cycle and hence a reduced performance for a reciprocating compressor. One of the methods to overcome such situations is to implement the refrigeration process in multiple stages, that is, to have two or more refrigeration cycles that operate in series. Such refrigeration cycles are named as cascade refrigeration cycles. A two-stage cascade refrigeration cycle is shown in Fig. 34. A HEX in the middle, which operates as the evaporator for the topping cycle and the condenser, link the two cycles for the bottoming cycle. In the following case study, a twostage cascade refrigeration system is introduced and analyzed thermodynamically.

4.9.3.1

Case Study 7: Two-Stage Cascade Refrigeration System

Smart and purposeful integration is a promising method for resolving the shortcomings of energy systems. Combining two or multiple energy systems considering the characteristics of subsystem results in better efficiency and more energy recovery for the integrated system. Smart integration with respect to each subset: for refrigeration systems, based on the required temperature, different configurations for integration of refrigeration systems are proposed [1]. Cascade refrigeration systems are used mainly for better achievement of lower temperatures in industrial applications such as NG liquefaction. Two-stage cascade refrigeration system is used for moderately low temperature applications where a large pressure range results in an inefficient performance of a single vapor-compression refrigeration system [1].

356

Combined Energy Conversion Systems

QH

Condenser

7

6

(A)

WA

Expansion valve

Compressor

Evaporator 5

8

3

2

Condensor

Expansion valve

WB

(B)

Compressor

1

4

Evaporator

QL Fig. 34 Schematic diagram for a two-stage cascade refrigeration cycle: (A) upper cycle and (B) lower cycle.

T

QH 6

6s

7 2s

A 2

8

3

B 4

1 QL S

Fig. 35 T–s diagram of a two-stage cascade refrigeration cycle.

4.9.3.1.1

System description

As shown in Fig. 34, a two-stage cascade refrigeration system is made of two power compression refrigeration cycles: an upper cycle (A) and lower cycle (B). The two cycles are integrated through a heat exchange process in which condenser of the cycle A rejects heat to the evaporator of the cycle B via a HEX that performs as an evaporator for the upper cycle and condenser for the lower cycle.

Table 9

Balance equations for each components of two-stage cascade refrigeration system

Components

MBE

EBE

EnBE

ExBE

Compressor B Expansion valve B Evaporator B Compressor A

m_ 1 ¼ m_ 2 m_ 3 ¼ m_ 4 m_ 1 ¼ m_ 4

m_ 1 h1 þ W_ CB ¼ m_ 2 h2 m_ 3 h3 ¼ m_ 4 h4 m_ 4 h4 þ Q_ L ¼ m_ 1 h1 m_ 5 h5 þ W_ CA ¼ m_ 6 h6 m_ 7 h7 ¼ m_ 8 h8 m_ 6 h6 ¼ Q_ H þ m_ 7 h7

m_ 1 s1 þ S_ gen;C1 ¼ m_ 2 s2 m_ 3 s3 þ S_ gen;ev1 ¼ m_ 4 s4 _ m_ 4 s4 þ QTLL þ S_ gen;e ¼ m_ 1 s1 m_ 5 s5 þ S_ gen;C2 ¼ m_ 6 s6 m_ 7 s7 þ S_ gen;ev2 ¼ m_ 8 s8 _ m_ s þ þS_ gen;con ¼ Q H þ m_ 7 s7

m_ 2 h2 þ m_ 8 h8 ¼ m_ 3 h3 þ m_ 5 h5

m_ 2 s2 þ m_ 8 s8 þ S_ gen;hex ¼ m_ 3 s3 þ m_ 5 s5

_ dest m_ 1 ex1 þ W_ CB ¼ m_ 2 ex2 þ Ex _ dest m_ 3 ex3 ¼ m_ 4 ex4 þ Ex  _ dest m_ 4 ex4 þ Q_ L TT0L 1 ¼ m_ 1 ex1 þ Ex _ dest m_ 5 ex5 þ W_ CA ¼ m_ 6 ex6 þ Ex _ dest m_ 7 ex7 ¼ m_ 8 ex8 þ Ex  T _ _ dest 0 þ m_ 7 ex7 þEx m_ 6 ex6 ¼ Q H 1

Expansion valve A Condenser A Heat exchanger (HEX)

m_ 5 ¼ m_ 6 m_ 7 ¼ m_ 8 m_ 6 ¼ m_ 7 m_ 2 ¼ m_ 3 m_ 5 ¼ m_ 8

6 6

TH

TL

_ dest;HEX _ 8 ex8 ¼ m _ 3 s3 þ m_ 5 s5 þ Ex m_ 2 ex2 þ m

Combined Energy Conversion Systems 357

358

Combined Energy Conversion Systems

Normally, air cooling is used for the condenser in system A, which is in a high-pressure and can be considered as a highpressure stage, while the evaporator in system A, which cools the condenser of system B, is a low-pressure stage of the cycle. Fig. 35 demonstrates the T–s diagram of the two-stage cascade refrigeration cycle. As a result of cascading, compressor work decreases (dotted area in T–s diagram) and heat absorption increases (grid area in T–s diagram), which means an increase in refrigeration capacity and coefficient of performance (COP), consequently. 4.9.3.1.1.1 Analysis Balance equations based on the first and second law of thermodynamics for the cascade refrigeration cycle as an integrated process are provided in Table 9. State condition for analysis of each component is assumed to be steady and potential and kinetic energies are considered as negligible. Besides, for the integration section, the HEX is assumed to be well-insulated and the heat transfer to each cycle during heat exchange is considered to be equal so that the mass flow rate relation can be written using the energy balance equation for HEX as _ B ðh2 h8 Þ ¼ m

_ A ðh5 m

h3 Þ

ð141Þ

For this cascade cycle, energetic COP, as a measure for energetic performance of the system can be defined as the ratio of the heat removed from the refrigerated space to the net total work input where: _L ¼m _ B ðh1 Q

h4 Þ

ð142Þ

_ net; in ¼ W _ CA þ W _ CB W _L Q COPen ¼ _ net;in W

!

¼

ð143Þ

_ B ðh1 h4 Þ m _ B ðh2 h5 Þ þ m

_ A ðh6 m

ð144Þ

h1 Þ

Also, exergetic COP can be expressed as _ QL Ex COPex ¼ _ W net; in

!

¼

_L Q



T0 TL

1



_ B ðh2 h5 Þ þ m

_ A ðh6 m

ð145Þ

h1 Þ

4.9.3.1.1.2 Results and discussion Using an example of the cycle provided by Cengel and Boles [3], a two-stage cascade refrigeration cycle using R134a as refrigerant with mass flow rate of 0.11 kg/s for the lower cycle and operating between the pressure limits of 1.4 MPa and 160 kPa. Heat exchanges between the upper and lower cycles with the pressures of 0.4 and 0.5 MPa, respectively. Isentropic efficiency for compressors is 0.8. Each stage of the cascade refrigeration cycle operates on the ideal vapor-compression refrigeration cycle. The refrigerant enters the compressor as a saturated vapor at the evaporator pressure. Also, the refrigerant leaves the condenser as a saturated liquid at the condenser pressure and standard vapor at the compressor inlet. The T–s diagram for this case study is shown in Fig. 36. Using the properties table for R134a and the given data for state conditions, mass flow rate for the upper cycle can be obtained using Eq. (141) as 0.168 kg/s. Moreover, according to Eqs. (144) and (145), COPen and COPex of the cascade cycle are 2.12 and 0.19, respectively. As seen in Fig. 37(A), increasing of ambient temperature has no effect on COPen, while it results in an increase in

T

1.4 MPa 6

QH

WA

7

2

A 3 8

4

0.4 MPa 0.5 MPa

WB B

5

0.16 MPa

1 QL S

Fig. 36 T–s diagram for the case studied for two-stage cascade refrigeration system.

Combined Energy Conversion Systems

4

3

0.8

359

0.35 COPen

0.7

COPex

2.5

0.4

COPen

COPen

0.5

COPex

0.6

3

0.3

COPex

0.25 0.2

2 0.15

COPex

COPen

2 0.3

0.1

1.5

0.2 1 290

300

310

320

330

340

1 290

0.1 350

T0 (K)

(A)

0.05

300

310

320

330

340

0 350

T0 (K)

(B)

Fig. 37 Effect of the ambient temperature on the coefficient of performances (COPs) for the two-stage (A) and single stage (B) refrigeration systems.

2.1

0.2 COP C en COPex

2

0.046 0.044

0.19 1.9

0.17 1.8

COPen

0.18

COPex

2 COPen

COPen COPex

1.8

0.042

1.7

0.04

COPex

2.2

1.6 0.16

1.6 0.15 1200 1400 1600 1800 2000 2200 2400 2600 P6 (kPa) (A)

1.5

0.038

1.4 0.036 1200 1400 1600 1800 2000 2200 2400 2600 Pcon (kPa) (B)

Fig. 38 Effect of condenser pressure on coefficient of performances (COPs) for the two-stage (A) and single stage (B) refrigeration systems.

_ QL ), which enhances the exergetic performance according to the incremental the exergy transferred via heat in the evaporator (Ex trend of exergetic COPex in Fig. 37(A) where the corresponding value increased from 0.19 to 0.58 when the ambient temperature increased from 298 to 348K. These figures clearly show that the combination of refrigeration cycles as cascade cycle significantly improves the exergetic performance of the system. As shown, the COPen is increased from 2 to 2.12 and COPex is increased from 0.04 to 0.19. Also, as seen in Fig. 38(A) and (B) increasing the condenser pressure (upper stage in cascade cycle) decreases the energetic and exergetic COPs as a result of higher work required for compression. Here again, compared with the single stage refrigeration cycle working between the same upper and lower limit pressures, the two-stage cascade cycle shows better performance both energetically and exegetically.

4.9.3.1.1.3 Concluding remarks A two-stage cascade refrigeration system was investigated energetically and exegetically as a case study for combined thermodynamic cycles. Effect of operating conditions on energetic and exergetic COPs as parameters for a sample of cascade refrigeration system were studied. This system provides advantages over the single stage vapor-compression refrigeration systems for gaining lower temperatures with suitable performances. As shown, integration of refrigeration cycles can improve the COPx for the case studied from 0.04 to 0.19. All in all, the cascade refrigeration systems are suitable systems for achieving lower refrigeration temperatures with no significant decrease in system performance.

360

4.9.4

Combined Energy Conversion Systems

Future Directions

GT and ST technologies have been developing through the centuries. In the 18th century, the fundamentals of GTs were introduced by the pioneers of the technology. Then in the early 20th century applications of the simple cycles spread to several industries. Power plants have long used GTs and STs in a simple cycle configuration for a restricted maximum power generation. Furthermore, commercialized facilities utilize the turbine components for the ground basis power production, typically in an arrangement with the process, heat generation, such as steam generation, hot water, or air heating. Currently, the performance of industrial turbines has been improved because of the significant incentives in research and development, regarding fuel to electricity conversion efficiency, the capacity of the power plant, viability, and dependability [5]. Ongoing progress and the recent developments of innovative turbines will enhance the performance of the simple cycle facilities over 40%. However, simple cycles’ efficiencies are still restricted due to the high energy rejection they have. The integration of the GTC, such as Brayton cycle, with a moderate or low temperature bottoming cycles such as the Kalina cycle or Rankine cycle, which is known as the typical combined cycle, is the best operational way to raise the thermal efficiency of a GTC. Heavy duty NG combusted GTs together with WHR vapor generators and vapor turbines indicate the main idea of this approach [29]. Multigeneration systems provide an opportunity to fulfill multiple useful commodities. ORCs incorporated in multigeneration systems are the upcoming technologies that are being investigated for implementation in the near future. Various studies have been conducted on multigeneration systems with ORCs. Suleman et al. [30] integrated geothermal and solar energy resources with ORCs for electricity production. Furthermore, the integrated system is also comprised of a cooling system based on absorption cooling and a system for drying wet products. The energy efficiency of the system was found to be 54.7%, and the exergy efficiency was obtained as 76.4%. Furthermore, Islam and Dincer [31] proposed a new multigeneration system based on solar and geothermal sources of energy. Combined cycles were utilized for power generation, an absorption cooling system was incorporated to provide cooling, a drying process was included to dry wet products, and a heat pump was utilized to provide space heating. The overall system energy and exergy efficiencies were obtained as 51% and 62%. However, for a single generation system, the energy and exergy efficiencies were evaluated as 28% and 54%, respectively. The tendency of the energy conversion systems can be seen in Fig. 39. It starts with the simple cycles, followed by the combined systems, cascade refrigeration, cogeneration, and trigeneration systems. The main goal for the future is achieving integrating multiple systems for multigeneration. For near the future, the fuel flexibility and intermediate capacity peak power generation are considered as the trends for combined energy conversion system development, where the rise in combustion temperature and challenges to new cycles are considered as mid- and long-term trends for the energy conversion systems [32]. New technologies toward the fuel range will direct the near future developments. Claire Soares [33] listed these developments and briefly explains as follows: Coal fuel combustion related: two innovative developments regarding the coal being utilized as a fuel in GT systems appear to shape future progress: oxy-combustion and hydrogen turbines. Theoretically, oxy-fuel combustion systems can be used in both conventional and advanced power plants. Around 30% nominal efficiency can be attained by today’s systems with STs when fueled with NG and capturing the carbon dioxide. Researchers focus their studies on improving the performance of these systems by reducing the emissions to nearly zero and increasing the efficiency around 60% [33]. Hydrogen turbines: by the development of the hydrogen turbines, turbine systems, and system elements are aimed to be improved, which includes cooling and combustor technology, and material and coating research. Those concepts appear to be the crucial technologies for the next generations. A power plant system comprises an advanced turbine that could be built to operate the world’s first near-zero emissions power plant to generate electricity and H2 from coal, while capturing and storing carbon dioxide through sequestration [33].

co

nv

er

sio

n

sy

st e

m

s

Integrated multigeneration systems

ss

in

en

er

gy

Trigeneration systems

Pr

og

re

Double, binary, combined systems Single cycles (Brayton, Rankine, Stirling, etc.)

Fig. 39 Progress in energy conversion systems.

Combined Energy Conversion Systems

361

Beside the improvement on fuel flexibility, thermodynamic cycle technologies like recuperation, aftercooling, intercooling, and cycle integration can be considered as the possible ways to enhance the performance of the CCPP at feasible costs within the near future [5]. The tendency of the energy conversion systems can be seen in Fig. 39. It starts with the simple cycles and is followed by the combined systems, cascade refrigeration, cogeneration, and trigeneration systems. The primary goal for the future is achieving integrating multiple systems for multigeneration. The working fluid plays a vital role in the performance of the systems with a bottoming cycle as ORC. Due to the low operating temperature, inefficiencies accompanying the heat transfer processes are high. The fluid thermodynamic properties highly influence these inefficiencies. In order to be compatible with the ORC, the working fluid needs to have low boiling points to be able to utilize the low-grade heat. Compatible fluids include chlorofluorocarbons as well as hydrocarbons such as isobutane. New organic fluids to be utilized in these power cycles are being investigated. The working fluid needs to have suitable properties to be utilized in these power cycles. The fluid should have high critical temperatures, which allows more heat to be absorbed by the fluid without reaching the critical temperature. Furthermore, fluids with high specific enthalpies need to be investigated. In order to increase the efficiency of ORCs, it is essential to introduce new organic fluids that can absorb a large amount of heat from the heat source without reaching the critical temperature and that have high specific enthalpies. Moreover, at high temperatures, the organic fluids deteriorate chemically. Hence, the temperature of the heat source also has an upper limit, beyond which the organic fluids might become unstable. The latent heat of vaporization of the fluids is also an important fluid property that determines the required flow rates and the specific heat absorbing capacity. Beside the working fluid technology, alternative fuels, which are also known as nonconventional fuels, will play a critical role for the future combined energy conversion systems technologies. Alternative fuels include the fuels obtained from biomass (biofuels), or from fossil fuels (synthetic fossil-based fuels). Blends of fossil-derived fuels and biofuels are also considered as alternative fuels (e.g., gasoline þ bioethanol blends). For the sustainable energy development, alternative fuel technology should gain more attention as they are viewed as a cleaner means of chemical energy storage with respect to fossil fuels. The use of hydrogen as a synthetic fuel is also believed to be a key solution for the future economy. As a transition period, various paths of producing hydrogen from fossil fuel, biomass, renewable resources, and a combination of these are considered. Integration of other alternative energy conversion systems are going to have an important role for the future technologies. Energy conversion systems without any GHG emissions such as TEGs offer promising solution for clean energy systems. BMW is developing a system to recover the heat from the exhaust stream and combustion engines of its car. However, with the current technology, a TEG with a figure of merit (ZT) with 1 is considered decent, which can only convert less than 10% of the heat to electricity. The TEMs with a lower thermal conductivity and higher electrical conductivity can bring up the ZT to the level of 2 or more and so they can increase their market shares for combined systems.

4.9.5

Concluding Remarks

This chapter introduced the concept of combined energy conversion systems for practical applications. In a combined energy conversion system, two systems are coupled to obtain the same output from the same input. The combined energy conversion systems offer a wide range of virtues, such as better efficiency, better cost-effectiveness, better resource use, better environment, and hence better sustainability. Several case studies were presented and thermodynamically analyzed to highlight the importance of combined energy conversion systems for practical applications.

Acknowledgment The authors acknowledge the support provided by the Natural Sciences and Engineering Research Council of Canada.

References [1] [2] [3] [4] [5] [6] [7] [8] [9] [10] [11]

Russell CR, Woodruff EB, Landis F, et al. Energy conversion. Chatswood, NSW: Encyclopedia Britannica; 2016. Dincer I, Zamfirescu C. Advanced power generation systems. 1st ed. Oshawa, ON: Elsevier; 2014. http://dx.doi.org/10.1016/B978-0-12-383860-5.09991-0 Cengel YA, Boles MA. Thermodynamics: an engineering approach. 8th ed. New York, NY: Mcgraw-Hiill Education; 2015. Dincer I, Ratlamwala TAH. Importance of exergy for analysis, improvement, design, and assessment. WIREs Energy Env 2013;2.335–49. doi:10.1002/wene.63. Poullikkas A. An overview of current and future sustainable gas turbine technologies. Renew Sustain Energy Rev 2005;9.409–43. doi:10.1016/j.rser.2004.05.009. Korobitsyn MA. New and advanced conversion technologies: analysis of cogeneration, combined and integrated cycles [PhD thesis]. University of Twente; 1998. California Energy Commission. Available from: http://www.energy.ca.gov. Rolls-Royce. Available from: http://www.rolls-royce.com. GE Power. Available from: http://www.gepower.com; 2004 [accessed 01.01.04]. U.S. Department of Energy. Available from: http://www.fe.doe.gov. Rao AD, Yi Y, Samuelsen GS. Gas turbine based high efficiency “vision 21” natural gas and coal central plants. In: Proceedings of the first international conference on industrial gas turbine technologies; 2005. [12] Lund University. Available from: http://www.vok.lth.se.

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[13] Haseli Y, Dincer I, Naterer GF. Thermodynamic modeling of a gas turbine cycle combined with a solid oxide fuel cell. Int J Hydrogen Energy 2008;33.5811–22. doi:10.1016/j.ijhydene.2008.05.036. [14] EG&G Technical Services I. Fuel cell handbook. 7th ed. In: Morgantown WV, editors. Prepared by EG&G Technical Services and Science Applications International Corporation for the National Energy Technology Laboratory, vol. 7; 2004. doi:10.1002/zaac.200300050. [15] Siddiqui O, Dincer I. Analysis and performance assessment of a new solar-based multigeneration system integrated with ammonia fuel cell and solid oxide fuel cell–gas turbine combined cycle. J Power Sources 2017;370:1–17. [16] Moran MJ, Shapiro HN, Boettner DD, Bailey MB. Fundamentals of engineering thermodynamics. Hoboken, NJ: John Wiley & Sons; 2010. [17] Tarique MA. Experimental investigation of scroll based Organic rankine systems. Oshawa, ON: University of Ontario Institute of Technology; 2011. [18] Giampaolo T. The gas turbine handbook: principles and practice. Cambridge: Cambridge University Press; 2003. http://dx.doi.org/10.1017/CBO9781107415324.004 [19] Demir ME, Dincer I. Development of a hybrid solar thermal system with TEG and PEM for hydrogen and power production. Int J Hydrogen Energy 2017 Available from: https://doi.org/10.1016/j.ijhydene.2017.09.001. [20] Demir ME, Dincer I. Development of an integrated hybrid solar thermal power system with thermoelectric generator for desalination and power production. Desalination 2017;404.59–71. doi:10.1016/j.desal.2016.10.016. [21] Demir ME, Dincer I. Development and heat transfer analysis of a new heat recovery system with thermoelectric generator. Int J Heat Mass Transf 2017; doi:10.1016/j. ijheatmasstransfer.2016.12.102. [22] Demir ME, Dincer I. Performance assessment of a thermoelectric generator applied to exhaust waste heat recovery. Appl Therm Eng 2017;120.694–707. doi:10.1016/j. applthermaleng.2017.03.052. [23] COMSOL. COMSOL Multiphysicss Modeling Software Version 5.2. [24] Kong LB, Li T, Hng HH, Boey F, Zhang T, Li S. Waste energy harvesting 2014;24. doi:10.1007/978-3-642-54634-1. [25] Zhang X, Zhao L-D. Thermoelectric materials: energy conversion between heat and electricity. J Mater 2015;1.92–105. doi:10.1016/j.jmat.2015.01.001. [26] U.S. Department of Energy’s Office of Energy Efficiency and Renewable Energy. Available from: http://www.fueleconomy.gov/feg/atv.shtml; 2016 [accessed 20.12.16]. [27] Dincer I, Rosen MA. Chapter 2 – exergy and energy analyses. Exergy 2013;Amsterdam: Elsevier; 2013. p. 21–30. http://dx.doi.org/10.1016/B978-0-08-097089-9.00002-4 [28] Demir ME, Dincer I. Development of an integrated hybrid solar thermal power system with thermoelectric generator for desalination and power production. Desalination 2016; doi:10.1016/j.desal.2016.10.016. [29] Poullikkas A. Parametric study for the penetration of combined cycle technologies into Cyprus power system. Appl Therm Eng 2004;24.1697–707. doi:10.1016/j. applthermaleng.2003.10.033. [30] Suleman F, Dincer I, Agelin-Chaab M. Development of an integrated renewable energy system for multigeneration. Energy 2014;78.196–204. doi:10.1016/j. energy.2014.09.082. [31] Islam S, Dincer I. Development, analysis and performance assessment of a combined solar and geothermal energy-based integrated system for multigeneration. Sol Energy 2017;147.328–43. doi:10.1016/j.solener.2017.02.048. [32] Fukuizumi Y. Gas turbine technology: the future for gas turbines – Power Engineering International. Available from: http://www.powerengineeringint.com/articles/print/ volume-13/issue-5/features/gas-turbine-technology-the-future-for-gas-turbines.html; 2017 [accessed 26.02.17]. [33] Soares C. Chapter 18 – future trends in the gas turbine industry. Gas turbines, 2nd ed. Available from: https://doi.org/10.1016/B978-0-12-410461-7.00018-3; 2015. p. 887–912.

Further Reading Dincer I. Refrigeration systems and applications. 3rd ed. London: John Wiley & Sons, Ltd.; 2017. Dincer I, Rosen MA, Ahmadi P. Optimization of energy systems. London: John Wiley & Sons, Ltd.; 2016. International Energy Agency. World energy outlook 2017. Paris: IEA; 2017. International Energy Agency. World energy investments 2017. Paris: IEA; 2017. International Energy Agency. Energy efficiency indicators highlights. Paris: IEA; 2016. Kanoglu M, Cengel YA, Dincer I. Efficiency evaluation of energy systems. New York, NY: Springer Verlag; 2012. Quoilin S, van den Broek M, Declaye S, Dewallef P, Lemort V. Techno-economic survey of organic rankine cycle (ORC) systems. Renew Sustain Energy Rev 2013;22:168–86.

Relevant Websites http://www.alstom.com Alstom Power. https://www.geaviation.com/ GE Aviation. https://powergen.gepower.com/applications.html GE Power. https://www.gepower.com/steam/steam-turbines GE Steam Power. http://www.globalenergyobservatory.org/list.php?db=PowerPlants&type=Gas Global Energy Observatory (GEO). https://www.iea.org International Energy Agency. https://www.iea.org/ International Energy Agency. https://global.kawasaki.com/ Kawasaki Heavy Industries. http://turbomachinery.man.eu/ MAN Diesel and Turbo. http://web.mit.edu/16.unified/www/SPRING/propulsion/notes/node27.html MIT – Thermodynamics and Propulsion lecture notes. http://www.mhps.com/en/products/thermal_power_plant/ Mitsubishi Hitachi Power Systems.

Combined Energy Conversion Systems

https://www.netl.doe.gov National Energy Technology Laboratory. http://nptel.ac.in/courses/112105123/24 National Program on Technology Enhanced Learning (NPTEL). https://www.siemens.com/global/en/home/products/energy/power generation/ Siemens. https://energy.gov/ U.S. Department of Energy.

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4.10 Integrated Gasification Combined Cycles Murat Ozturk, Suleyman Demirel University, Isparta, Turkey Ibrahim Dincer, University of Ontario Institute of Technology, Oshawa, ON, Canada r 2018 Elsevier Inc. All rights reserved.

4.10.1 4.10.2 4.10.3 4.10.3.1 4.10.3.2 4.10.3.3 4.10.3.4 4.10.3.5 4.10.3.6 4.10.3.7 4.10.3.8 4.10.3.9 4.10.4 4.10.4.1 4.10.4.2 4.10.4.3 4.10.4.4 4.10.4.5 4.10.4.5.1 4.10.4.5.2 4.10.4.5.3 4.10.4.5.4 4.10.4.5.5 4.10.4.5.6 4.10.4.6 4.10.4.7 4.10.4.8 4.10.4.8.1 4.10.4.8.2 4.10.5 4.10.5.1 4.10.5.2 4.10.5.3 4.10.5.4 4.10.5.5 4.10.5.6 4.10.5.7 4.10.5.8 4.10.6 4.10.6.1 4.10.6.1.1 4.10.6.1.2 4.10.6.2 4.10.6.3 4.10.6.4 4.10.7 4.10.7.1 4.10.7.1.1 4.10.7.1.2 4.10.7.1.3 4.10.7.2 4.10.7.2.1

364

Introduction Gasification Process Gasification Methods Biomass Gasification Coal Gasification Stoichiometric Method of Gasification Cycles Co-Gasification Fischer–Tropsch Process Hydrogen Production via Gasification Process Methanol Production via Gasification Process Ammonia Production via Gasification Process Syngas Cleaning and Sulfur Removal Processes Thermodynamic Assessment of Integrated Gasification Combined Cycles Basic Thermodynamic Concepts Mass Balance Equation Energy Balance Equation Entropy Balance Equation Exergy Balance Equation Exergy analysis of heat transfer Exergy analysis of work transfer Exergy analysis of flow Exergy destruction rate Energy efficiency analysis Exergy efficiency analysis Specification of Coal Sources Thermodynamic Analysis of a Gasification System Environmental Impact Analysis Normalized carbon dioxide emissions Sustainability analysis Combined Cycles Rankine/Organic Rankine Cycle Combined Cycle Rankine/Kalina Combined Cycle Rankine/Stirling Combined Cycle Brayton/Rankine Combined Cycle Brayton/Kalina Combined Cycle Brayton/Brayton Combined Cycle Brayton/Stirling Engine Combined Cycle Brayton/Solid Oxide Fuel Cell Combined Cycle Integrated Gasification Combined Cycles Coal-Based Integrated Gasification Combined Cycle Coal-based integrated gasification combined cycle without CO2 capture Coal-based integrated gasification combined cycle with CO2 capture Biomass-Based Integrated Gasification Combined Cycle Waste Materials-Based Integrated Gasification Combined Cycle Heavy Oil-Based Integrated Gasification Combined Cycle Case Studies Solar Energy Combined Biomass Gasification System Balance equations for system components Coefficient of performance Parametric analysis Integrated Gasification Combined System With Biomass Gasification and Solid Oxide Fuel Cell Balance equations

Comprehensive Energy Systems, Volume 4

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doi:10.1016/B978-0-12-809597-3.00412-0

Integrated Gasification Combined Cycles 4.10.7.2.2 Solid oxide fuel cell 4.10.7.2.3 Thermal characteristics of biomass resources 4.10.7.2.4 Parametric studies 4.10.7.3 Integrated Gasification Combined System With Coal Gasification and Hydrogen Liquefaction 4.10.7.3.1 Balance equations 4.10.7.3.2 Properties of coal samples 4.10.7.3.3 Hydrogen liquefaction system 4.10.7.3.4 Thermodynamic assessment of integrated system 4.10.7.3.4.1 Gasification reaction 4.10.7.3.4.2 Combustion reaction 4.10.7.3.4.3 Gas cycle process efficiency 4.10.7.3.4.4 Steam cycle process efficiency 4.10.7.3.4.5 Combined process efficiency 4.10.7.3.5 Proton exchange membrane electrolyzer 4.10.7.3.6 Parametric studies 4.10.8 Future Directions and Potential Developments 4.10.8.1 Tar Pollution 4.10.8.2 Ash Melting 4.10.8.3 Too Low Syngas Heating Value 4.10.8.4 Feedstock 4.10.8.5 Gas Cleaning 4.10.8.6 Prime Mover 4.10.8.7 Reliability 4.10.8.8 Conditions for Commercialization in the Future 4.10.9 Closing Remarks References Relevant Websites

Nomenclature A E E_ ex _ Ex _ D Ex F g h H J J0 Jiref L m _ m P q

Area (m2) Energy (kJ) Energy rate (kW) Specific exergy (kJ/kg) Exergy rate (kW) Exergy destruction rate (kW) Faraday constant (C/mol) Gravitational acceleration (m2/s) Specific enthalpy (kJ/kg) Enthalpy (kJ) Current density (A/m2) Exchange current density (A/m2) Pre-exponential factor (A/m2) Length (m) Mass (kg) Mass flow rate (kg/s) Pressure (kPa) Specific heat transfer (kJ/kg)

Greek Letters D Change in variable Water content at anode-membrane interface la (O 1) lc Water content at cathode-membrane interface (O 1)

Q _ Q RPEM Ru s S S_ t T v V V0 Vact Vact,a Vact,c w W

_ W z

l(x) sPEM s(x) Z c

Heat (kJ) Heat rate (kW) Proton exchange membrane resistance (O) Universal gas constant (kJ/mol K) Specific entropy (kJ/kg K) Entropy (kJ/K) Entropy rate (kW/K) Time (s) Temperature (K) Velocity (m/s) Volume (m3) Reversible potential (V) Activation overpotential (V) Anode activation overpotential (V) Cathode activation overpotential (V) weight (N) Work (kJ) Work Rate (kW) Elevation (m)

Water content at location x in the membrane (O 1) Proton conductivity in PEM (s/m) Local ionic PEM conductivity (s/m) Energy efficiency Exergy efficiency

365 447 450 451 453 456 462 462 463 463 463 464 464 464 465 465 468 468 469 469 469 470 470 470 470 471 471 473

366

Integrated Gasification Combined Cycles

Subscripts a abs ch ct cooling con cv D e ej en erd eva ev ex

Air Absorber Chemical Cooling tower Cooling load Condenser Control volume Destruction Exit condition Ejector Energy Energy recovery device Evaporator Expansion valve Exergy

Superscripts

Acronyms ASU COP CSP EES ER HEX HRSG

4.10.1

Air separation unit Coefficient of performance Concentrating solar collector Engineering Equation Solver Equivalent ratio Heat exchanger Heat recovery steam generator

f fls g gen gene heating HP kn l i o ohm p ph pt tot

Fuel Flashing Gas Generation Generator Heating load High pressure Kinetic Liquid Inlet condition Dead state Ohmic Pump Physical Potential Total

:

Rate

IGCCs ORC PEM PV RE SExI SMR TE

Integrated gasification combined cycles Organic Rankine Cycle Proton exchange membrane Photovoltaic Rational efficiency Specific exergy index Steam methane reforming Task efficiency

Introduction

The energy sources have important critical roles in driving almost all practical systems and are necessary to support the quality of life. The speed of energy production is a locomotive force of the industry and development, and a significant indicator in the improving of societies. On the other hand, it is seen as one of the significant elements in sustainable growth. Considering increasing human energy consumption rate, mankind requirements have to find new forms of energy sources for the next generation for two reasons. The first reason is the environmental aspect, and the second reason is the diminishing supply of fossil fuels. These factors are encouraging researchers and societies to find new energy resources and technologies. In this chapter, overviews about integrated gasification combined cycles (IGCCs) and advantages of utilizing gasification technologies for energy, chemicals, and materials production are investigated to provide background motivations for this chapter. One of the most significant challenges of this century is keeping up with the increasing global energy demands due to growing population and increasing living qualities. In 2015, about 7 billion people in the world used 11,621 Mtoe of energy. By 2050, these values are expected to increase to 9 billion and 20,000 Mtoe, respectively. Nowadays, 81% of the world energy need is met by fossil fuels (coal, oil, and natural gas). Therewithal, as a result of their limited supply and nonhomogeneous distribution, the fossil-based fuels are not expected to provide pace with the increasing energy need. Meeting the importantly increasing world energy needs with no or minimal harmful emissions and fossil fuel dependency can only be achieved by utilizing clean energy processes. These processes can submit significant environmental, energetic, financial, and societal benefits. Reducing the dependence on coal, oil, and natural gas, and decreasing potential environmental damages should be achieved by utilizing clean and sustainable energy resources. Different alternative energy solutions, such as solar, wind, biomass, hydropower, geothermal, and nuclear energy, have been presented to supply great and sustainable energy submission and to reduce climate change by monitoring the greenhouse gas (GHG) emissions [1]. The generation and consumption rate of energy sources have an important role in deciding power conversion at each process of economic development of countries. Total energy consumption rate in the world should be obviously classified based on kinds of fuel sources, kinds of area, end-use sector, economic growth, and purchasing power as illustrated in Fig. 1, and also that has a direct impact on economic and environmental development aims.

Integrated Gasification Combined Cycles

367

Total power consumption in the world

Kinds of fuel source

OECD member countries

Residential application

Oil

Kinds of area

End-use sector

Non-OECD member countries

Commercial application

Natural gas

Economic growth

Lower level

Higher level

Industry needs

Coal

Purchasing power

Transport application

Nuclear

Higher rate

Lower rate

Different sector needs

Renewable energy sources

Fig. 1 Schematic diagram of total power consumption in the world. OECD, Organisation for Economic Co-operation and Development.

Nuclear 6%

Renewables 11%

Liquids 33%

Renewables 14% Nuclear 7%

Coal 27%

Coal 28% World primary energy 2015

Natural gas 22%

World primary energy 2035

Liquids 29%

Natural gas 23%

Fig. 2 World primary energy demand at 2015 and supposed need at 2035. Data taken from EIA. International energy outlook 2016. Washington, DC: U.S. Energy Administration Information; 2016.

The share of world primary energy for 2015 and 2035 are shown in Fig. 2, respectively. The liquid energy sources and natural gas are supposed to continue to be the primary source of fuel through the next two decades as illustrated in Fig. 2, and according to the growing rate of energy need, the liquid fuels are the slowest increasing source of energy. It is expected that, at the near future, the biomass gasification technologies will have an important role in the increasing rate of renewable energy. As seen in Fig. 3, it is clear that coal continues to be the most utilized fuel for electricity production in the world. An important decrease in the fossil-based fuels consumed per unit power can be reached by growing of renewable energy utilization efficiency, which leads to decrease in harmful gas emissions. For this reason, the coal gasification technologies are very important for clean power production. Also, its share reduces as the renewable alternatives, natural gas, and nuclear power are expected to advance during the projection area. In a close future, the power cost most likely will be added to the CO2 emissions through carbon credits and taxes, thus researchers and policy makers are more worried now with the different present alternatives with no or relatively low level of emissions to remove the existing power production processes. Based on their low or zero end-use harmful emissions and constantly reloaded resources, renewable energies (e.g., solar, wind, biomass, geothermal, waves, etc.) are considered as sustainable replacements to fossil fuels. One of the primary structures of electricity production is supplying power for rural region and for nearly one third of the world population living in isolated zones and has no obtain to a power utility grid. The suitable way for this structure is the improving of processes depending on the local alternative energy sources. There are several technologies for renewable energy resources to

Integrated Gasification Combined Cycles

Net power generation (Trillion kWh)

368

40 35 30 25 20 15 10 5 0

Liquids

2012

2020

Natural gas

2025 Coal

2030 Years Nuclear

2035

2040

Renewables

Total

2375

18.5

2275

18

2175

17.5

2075 17

1975

16.5

1875 1775

CO2 emissions

1675

Petroleum

1575

Natural gas

1475

Coal

15.5 15 14.5

1375 1275 2010

16

CO2 emissions per person (tons)

CO2 emissions (million metric tons)

Fig. 3 World net power generation by fuel type. Data taken from EIA. International energy outlook 2016. Washington, DC: U.S. Energy Administration Information; 2016.

2015

2020

2025 Years

2030

2035

14 2040

Fig. 4 Energy related CO2 emissions by source from 2010 to 2040. Data taken from DOE/EIA-0383. Annual energy outlook 2013 with projections to 2040. Washington, DC: U.S. Energy Information Administration; 2013.

be used for electricity generation. These alternative resources contain the solar energy, wind, hydropower, geothermal, biomass, and nuclear energy. Being obtainable for great-scale usage, environmentally benign, and cost effectiveness are the primary indicators that assistance the biomass gasification to play the primary role of electricity production for the near future. In addition to that, the potentials of gasification technologies in substituting fossil energy-based technics for power generation have frequently attracted growing concerns from industries, academia, and policy makers. The primary motives behind this situation are due to ecological impacts associated with the burning of fossil-based fuels and, also energy safety threats and safety conditions in countries. For environmental impacts, one of common GHG that released from fossilbased fuels combustion is CO2. The CO2, H2O vapor, and CH4 known as GHGs absorb solar energy and create the natural greenhouse cover impact around the world. Carbon dioxide molecules can stay in the atmosphere for hundreds of years. Increasing fossil fuels consumption rate has been the important contributor to GHGs emission which has been dramatically raised since preindustrial times. It is known that the world average temperatures would be 301C lower without this impact [2]. The carbon dioxide concentration has been raised by over one-third from 280 ppm in 1750 to 379 ppm in 2005 [3]. On the other hand, as the same impact in a greenhouse effect, this harmful gas traps the heat and increase the world temperature. The CO2 emission data from fossil fuels burning and CO2 emissions per person are illustrated in Fig. 4. As the population of world rises and under-developing countries develop, the energy demand continues to increase [4]. The worldwide primary energy demand from 1980 to 2035 is shown in Fig. 5.

Integrated Gasification Combined Cycles

369

4500

Energy sources (Mtoe)

4000 3500 3000 2500

Coal Petroleum

2000

Natural gas

Hydro

Nuclear

Biomass

Other renewables

1500 1000 500 2010

2015

2020

2025

2030

2035

Years Fig. 5 World primary energy demands in metric tonne of oil equivalent (Mtoe). Data taken from World Energy Council. World energy resources 2013 survey. London: World Energy Council; 2013.

Table 1

Numbers of gasifiers by primary feedstock

Feedstock

Operational plant

Construction (2017–18)

Planned plant (2019–25)

Coal Petroleum Natural gas Petcoke Biomass Waste

412 136 43 12 26 29

218 8 1 2 1 1

487 12 2 3 5 2

Source: Data taken from U.S. DoE and Gasification Technology Council. World gasification database. Washington, DC: U.S. Department of Energy; 2016.

Table 2

Primary feedstocks for gasification

Feedstock

Operational plant (synthesis gas, GWth)

Construction (2017–18) (synthesis gas, GWth)

Planned plant (2019–25) (synthesis gas, gwth)

Coal Petroleum Natural gas Petcoke Biomass Waste

123 24 23 11 2 0.5

73 5 2 17 1 0.2

95 13 5 13 2 0.7

Source: Data taken from Higman C. State of the gasification industry: worldwide gasification and syngas databases 2016 update. In: Gasification and syngas technologies conference, Vancouve, Canada; 2016.

The integrated gasification combined processes are emerging as a best exiting technic to use low quality or contaminated energy sources, such as coal, oil, or biomass. The integrated gasification cycles produce multiple outputs, such as power, hydrogen, and chemicals (methanol, higher alcohols, etc.) and by-products (sulfur, sulfuric acid, slag, etc.). Moreover, the integrated gasification cycle technologies have the potential for carbon dioxide sequestration. In the IGCCs, as coal and biomass sources are not combusted, the slightly lower volumes of synthesis gaseous are easier to clean up than the much higher volumes of exhaust gaseous of coal or biomass combustion processes. The gasification cycles can process various kinds of sources as illustrated in Table 1 at different working states, and also they can succeed various conversions, and hence may be only limited to specific productions, such as hydrogen, power, ammonia, oxy-chemicals, synthesis gaseous, and methanol. As seen in Table 1, in the year 2016, out of the total of 658 operating gasifiers around the world, about 412 of these have utilized the coal sources as the feedstock.

370

Integrated Gasification Combined Cycles

Products

CO2 removal

Sequestration Steam

Water−gas shift reactor

Cleaning cycles

Integrated gasification combined cycle

Coal

Petcok

Biomass

Steam reforming

Natural gas

Fig. 6 Integrated gasification combined cycle (IGCC) and related technologies.

Bio-fuels generation and harvesting

Transportation to gasification facility

Integrated gasification combined cycles

• • • • • • • • • • • •

Further integration for multigeneration

More useful outputs: Power Hydrogen Alternative fuels Chemicals Fresh water Heating/process heating District heating and cooling Air conditioning Cooling Evaporation Drying Hot water/steam

Fig. 7 Process parts of biofuels-based integrated gasification combined system.

Table 2 tabulates the operational, construction, and planned plants syngas capacity over time based on feedstock. The integrated gasification processes for especially liquid and gaseous fuels generation are becoming increasingly significant. Also, the biomass and coal sources are the dominant feedstock, and will continue to be so for the near future. The integrated gasification processes are becoming both larger and smaller. The large industrial coal and petroleum coke-based gasification projects (for electricity, chemicals, hydrogen production) are getting bigger. Generally, these processes are located in Asia and the Middle East. Also, the demand for smaller and modular integrated gasification systems is increasing for the biomass and waste material gasification aims. Usually, the gasification processes of biomass and waste material do not need the larger gasifiers that are used in industrial applications. The gasification capacity continues to increase on the worldwide basis. Therefore, the IGCC and its related technologies is given in Fig. 6. To produce different products, such as power, hydrogen, ammonia, nitrogen, methanol, dimethyl ether, hydrocarbons, substitute natural gas, Fischer–Tropsch hydrocarbons, etc., from integrated system, the gasification cycles begins with the suitable feedstock. Biomass, an important energy resource globally, is being examined in different countries as a potentially important renewable source. Biomass is derived from solar radiation, and forms of biomass include wood, municipal solid wastes, and industrial residues. Biomass energy can be utilized as a fuel for power production, transportation, heating and cooling application, hydrogen or other synthetic fuels production, etc. The block diagrams in Fig. 7 describe the process parts in general biofuels production, harvesting, transportation, and end-use for different useful outputs generation. Applications of

Integrated Gasification Combined Cycles

371

biomass-based gasification process for power generation are expected to show progressive substitutions of fossil energy-based technologies in the near future. In order to reach this prospect, first and foremost, biofuels-based sources need to be generated and harvested in a sustainable manner. Biofuels transportation is an important consideration for projection and operation of an integrated gasification combined system. Reducing transportation cost factors are therefore essential to achieve economic advantages since it can be the limiting indicator for financial monetary feasibility of integrated gasification combined systems. The biofuel transportation cost factors are the function of distance, density of biofuel, and transportation methods [5]. The biofuels conversion methods are usually should be divided into three different classes: such as (1) thermochemical, (2) chemical, and (3) biochemical processing ways. The gasification cycle of biomass is in the thermochemical process category. For an integrated biomass gasification cycle, if the desired latest outputs from the syngas are different useful products, such as electricity, hydrogen, other synthetic fuels, heating, cooling, hot and/or fresh water, etc., an integrated system for multigeneration as a further processing can be installed. On the other hand, if the desired latest outputs from the syngas are the different chemicals and liquid fuels, further processing like methanol generation process and Fischer–Tropsch process should be deployed, respectively. The biomass gasification cycles are the transformation of biomass sources by using partial oxidation into gaseous output, which called as synthesis gas (or syngas), containing mainly of hydrogen and carbon monoxide, with lesser amounts of carbon dioxide, water, methane, higher hydrocarbons, and nitrogen [6]. The biomass gasification cycles are regarded one of the maximum effective methods of converting the power embedded in biomass sources, and it is becoming one of the best technological options for solid wastes reuse [7]. The definition of exergetic performance is vital to calculate the amount of useful energy and is utilized in the IGCCs. The exergetic assessment of the biomass gasification process has conducted by Abuadala et al. [8], and also the impacts of temperature and the amount of steam injection on the hydrogen yield, and energetic and exergetic performances are analyzed detailed for more efficiently process design. The energetic and exergetic performance of steam gasification of biomass are investigated and compared for several kinds of raw materials [9]. It is illustrated that the exergetic performance, which is a function of feedstock, gasification temperature, and steam to biomass ratio, changes from 65% to 80%. The useful outputs from an examination of the variation in exergetic content of synthesis gas and the exergy efficiency of the IGCCs show that municipal solid wastes are a competitive fuel to biomass in these cycles [10]. The gasification combined cycles, which is characterized by partial oxidation, are a vital component of several clean energy technologies. Having analyzed the significance and advantages of utilizing integrated gasification technologies as an alternative and clean feedstock, this chapter will focus about the introduce and thermodynamic analysis of IGCCs to provide a better view of multigeneration performance of the integrated processes. The different parametric studies, based on design indicators of integrated gasification systems, are conducted to investigate the exergy destruction rate, energetic and exergetic efficiencies by considering several processes [11]. It is illustrated that the adiabatic temperature of the gasifier chamber essentially varies with the kind of gasifying environment. It is shown that the performance of air gasification cycle is higher than steam gasification cycle, because the integrated gasification process operates at higher temperature. The impact of moisture substance of input material on the heating value of the generated synthesis gas and the exergetic performance of gasifier is investigated in the gasification of municipal solid waste [12]. The higher moisture substance of raw material results in the lower heating value (LHV) of synthesis gas and exergetic performance of the IGCC. The impact of process indicators, such as the temperature and equivalent ratio (ER) on the gasification efficiency system irreversibilities, are defined based on the steam and steam–air mixture as gasifier environments [13–15]. The ER is described as the ratio of utilized oxygen for gasification of any substance divided by the required O2 for its stoichiometric combustion. In the aforementioned analysis, the ER is considered as an important indicator in the operation of integrated gasification processes. Other preferences are then proposed to develop the exergetic performance of the gasification systems. The impact of fuel composition on the thermodynamic performance of gasifiers and gasification processes is analyzed and it is found that highly oxygenated biofuels are not ideal fuels for gasifiers from the exergetic viewpoint [16]. It is also illustrated that the higher gasification temperature reduces the kinetic limitations of gasification system. The impact of ER on the performance of biomass gasification with air is directed and study outputs illustrate that rise in ER corresponds to the higher exergetic and energetic performance [17]. The IGCCs are the important integrated processes that combine advanced gasification technic with both gas and steam turbine electricity production processes for trigeneration of power, steam, and hydrogen [18]. The IGCCs present higher performance and lower harmful emissions in comparison to a conventional pulverized coal combustion electricity production process. It also supplies the opportunity of utilizing diversity of low value fuels and producing multigeneration outputs [19]. The air separation unit (ASU) should be combined to the gasification cycle to increase its performance by substituting air with oxygen which makes an oxygen-fired IGCC power facility more attractive [20]. Even though utilizing O2 instead of air increases the gasification performance and decreases NOx emissions, but it will impose additional capital investment to an IGCC power facility. It is calculated that nearly 8% of the capital investment for the common IGCC power generation facility is depend to the ASU [19]. The IGCCs can be extended beyond the integrated cycle electricity production application and has a possible to be a basis for advanced power processes. It should be utilized to generate high quality steam for heating application, or generate transportation fuels and variety of chemicals through catalytic conversion of the clean synthesis gas. It should be also integrated with improvement fuel cell technology to produce clean power with high performance. The specific outputs of this chapter can be defined as follows:

• •

to propose the novel IGCCs; to conduct energetic and exergetic assessment of the IGCCs considered to define their energetic and exergetic efficiencies;

372

• • •

Integrated Gasification Combined Cycles

to progress the exergy destruction model to predict the quantities and locations of the exergy destruction rates in the investigated IGCCs; to perform the environmental impact assessment to investigate the GHG emissions from the investigated IGCCs; and to make the comprehensive process efficiency assessment, the parametric studies are carried out to investigate the change of the IGCC efficiency. The significance of parametric studies is to predict the efficiency analysis of investigated process in order to discover the suitable design indicators for the system performance.

The current chapter aims to present the integrated coal and biomass gasification cycles for multigeneration, which is very important for countries having low-grade coal supplies and excess biomass.

4.10.2

Gasification Process

The inlet materials of gasification process are usually coal, liquid residues, natural gas, biomass, and waste. The output useful materials from gasification process are usually ammonia, methanol, carbon monoxide, hydrogen, oxo-alcohols, syngas, town gas, and reduction gas. Also, syngas can be utilized in Fischer–Tropsch process for fuels, waxes, and other useful material production or used in gas turbines for electrical power generation. These processes are illustrated in Fig. 8 as an input materials and output chemicals. Due to both worldwide and governmental concerns on sustainable power production, environmental legislations limit the environmental impact of the coal-fired combustion processes; encouraging clean coal combustion systems. The fluidized bed combustors present less harmful emissions and more performance power production with growing usage rate. Countries with an important reserve of low quality coal sources have considerable interest in clean combustion processes. In the open-literature, there are different papers focusing on how the integrated gasification combined processes can be one of the most successful resolutions playing an important role in providing better environment and sustainability. The significant influence on the accomplishment of scaled up multigeneration based on IGCCs is the synergistic incorporation of crucial technologies and additional valued commodities production and at the same time maintaining the system simplicity. For the reason that more than one products are obtained, the economic feasibility of the integrated gasification system is improved. The multigenerational systems that incorporate gasification cycles to increase the system performance might be an alternate opportunity to develop affordable integrated gasification systems. The gasification process is a way for converting solid, liquid, or gaseous materials to the combustible or synthetic gaseous (e.g., hydrogen, carbon monoxide, carbon dioxide, methane). Fig. 9 illustrates the conventional procedures for gasifying the carbonaceous materials. The schematic diagram of gasification process for power and heating generations are shown in Fig. 10. The direct steam-biomass dryer is utilized to decrease the moisture content of gasification materials before materials enter the gasification process. The steam exiting from the dryer has considerable energy. Therefore, it is used in the district heating for efficient energy utilization and utilized in steam generator for power generation. As illustrated in Fig. 11, with the IGCC we can generate the power, heating and cooling application simultaneously. Fig. 11 illustrates that the trigeneration systems have mainly four subcomponents:

•

the power generation process, which is known as the system’s prime mover, such as the gas turbine, in IGCCs;

Ammonia Methanol

Coal

Carbon monoxide

Liquid residues

Hydrogen Natural gas

Gasification process

Oxoalcohols

Biomass

Syngas

Waste

Town gas Reduction gas

Fig. 8 Input materials and output chemicals of gasification process.

Integrated Gasification Combined Cycles

Ash and char Syngas cleaning CO, H2, N2 Low temperature syngas process

Gasification process Steam

Air

Ash and char Syngas cleaning CO, H2 process Medium temperature syngas

Gasification process Steam Oxygen Gasification materials

Ash and char Syngas cleaning CO, H2, N2

Gasification process Steam

373

process

Heat

Medium temperature syngas

Ash and char Syngas cleaning CO, H2, CH4

Hydrogasification process

process

Hydrogen Heat

High temperature syngas

Ash and char Syngas cleaning CH4

Catalytic gasification process

process

Low temperature syngas

Unused chemicals Steam Fig. 9 Gasification procedures of solid, liquid, and gaseous materials.

District heating

Gasification materials

Heating applications Dryer

Superheated steam

Partially dried biomass

Electricity

Steam generator

Gasification process

Gasification products

Air/steam

Fig. 10 Schematic diagram of gasification process for power production and heating applications.

• • •

the cooling process, such as the single, double, triple, or quadruple, effect absorption chiller; the heating process, such as a boiler, of integrated system; and the electrical generator for power generation.

Nowadays, the researchers have started to investigate the trigeneration for producing more useful outputs, such as power, heating, cooling, hot water, hydrogen or other synthetic fuels, fresh water or air, etc., by using the IGCC by implementing a process called multigeneration. There are different advantages of multigeneration systems, including the higher process performance, the decreased thermal energy emissions and waste heat, the decreased operational cost, the decreased harmful emissions, the preferable utilization of sources, the short transmission lines, the fewer distribution components, the multiple generation outputs, the improved reliability, and the less grid failure. There are a lot of advantages of multigeneration processes, such as higher cycle performance, decreased thermal energy losses, decreased operating costs, decreased harmful gaseous emission, better utilize of sources, shorter transmission lines, fewer distribution components, multiple production outputs, improved reliability, and less grid failure [21]. The schematic diagram of multigeneration system based on gasification cycle is illustrated in Fig. 12. The performance of multigeneration energy processes is usually higher than trigeneration process because of the extra outputs, such as hydrogen or

374

Integrated Gasification Combined Cycles

Ash and char Solid fuels

Gasification chamber

Syngas

Syngas storage

Gas fuel

Power generator process

Mechanical power

Electrical generator

Heating unit

Power

Heating application

Waste heat Cooling unit

Cooling application

Fig. 11 Schematic diagram of gasification process-based trigeneration system.

Ash and char Solid fuels

Gasification chamber

Syngas

Syngas storage

Gas fuel

Power generator process

Waste

Mechanical power

Electrical generator

Power

Water PEM electrolyzer

heat

Thermal storage system

Hydrogen

Hot water Space heating

Cooling unit

Cooling application Air conditioning

Desalination unit Sea water

Fresh water Salt

Fig. 12 Schematic diagram of gasification process based multigeneration system. PEM, proton exchange membrane.

other synthetic fuels, fresh water, hot water, etc. To generate hydrogen, the low or high temperature electrolyzer can be utilized, which is driven by part of the electricity produced by the power generation process. Hot water exiting from power generation process goes to the proton exchange membrane (PEM) electrolyzer and is reacted electrochemically to split its molecules into H2 and O2. The heating process can be created of two components, one for hot water generation and another for residential heating. The rejected heat energy from storage process enters the single effect absorption cooling process to generate cooling effect and air conditioning. To produce fresh water from multigeneration system, the desalination process must be utilized.

Integrated Gasification Combined Cycles

4.10.3

375

Gasification Methods

Nowadays, the availability of energy resources and increasing harmful emissions are the two significant concerns for the sustainability of power generation. The energy need rate of world is growing at a very rapid rate where there is a crucial requirement for an important reduction of harmful gaseous emissions. A limited amount of fossil sources for heat and electricity provides improving of alternative resources necessary. Future of the world energy supply is based on the energy supply security, sustainable development and environmental protection, and utilization of energy resources in more efficiently and cost effectively methods. Gasification converts solid fuels, such as coal or biomass, into product gas that can be utilized in various applications. The product gas is mainly composed of CO, CO2, CH4, and H2 and can be combusted to produce work and heat. Also, the outputs from gasification process are particulate matters, tar, ammonia, and hydrogen sulfide. The impact of equivalence ratio on the quality of the syngas should be investigated for more efficiently outputs production. The chemical composition of syngas can be determined with a gas analyzer unit, which measures CO, CO2, CH4, H2, and O2 components. In addition to that, the LHV syngas can be determined by using the syngas composition computations. The IGCCs have drawn widespread interests in nowadays because it addresses each of the given above worries. The gasifier systems work at high temperatures (47001C) and can accept the large diversity of low value feedstocks, such as heavy oil, refinery residues, or refuse-derived fuels [22]. In order to offset the disadvantages of gasification-based power processes such as LHV of syngas, complexity of process, and to improve its performance and decrease its environmental impacts oxygen, can be used instead of air to increase the system performance, improve the hydrogen content of generated syngas, and decrease the CO2 and NOx harmful emissions [18]. This important case can be achieved by process integration of the gasification process. The gasification process is a thermochemical conversion of either the solid (coal, coke, biomass, and solid waste) or liquid (oil, tar, and pitch) fuel sample into the synthesis gas (or syngas) composed primarily of hydrogen (H2) and also carbon monoxide (CO) [23]. The coal samples can be selected such that their carbon and hydrogen content, as well as their higher heating values (HHVs) are different. Also, the biomass samples can be investigated to compare its impact on the syngas composition. Different from the combustion systems that only generate CO2 and H2O, the gasification cycles are the partial oxidation processes that occur in an O2-limited environment. The producing synthesis gases are more beneficial than combustion flue gaseous, and also they have the potential to produce power more effectively. In the 20th century, the gasification processes have been utilized to convert coal sources into gaseous fuels for domestic heating and lighting productions. Nowadays, the gasification processes have been utilized in the petrochemical facility for the generation of chemical outputs [24]. The investigation and assessment studies of gasification cycles should be divided up into two main parts. The main efforts for gasification cycles are focused on the techniques to increase the heating value of generated syngas where the others are concentrated on methods to improve the concentration of generated hydrogen. The different studies for integrated gasification investigate the whole conversion performance of the gasification system based on the first law of thermodynamics, where others investigate irreversibilites in the system and the quality of energy at different working cases based on the second law of thermodynamics and exergy analysis viewpoint [25]. Two techniques are proposed to model and investigate the IGCCs. The first method is based on the reduction of the Gibbs free energy that requires advanced mathematical accounts. The second method of equilibrium types is basis of the equilibrium constant which is on the based on the same concept but is used to predict the composition of the generated synthesis gas mixture [26–28].

4.10.3.1

Biomass Gasification

The novel electricity production concepts, such as co-firing or combined gas turbines plant, help to decrease the dependency on fossil fuels and decrease the carbon dioxide emissions. The gasification (partial combustion, reforming) is the cleanest, most flexible, and reliable way of utilizing of fossil-based sources. Also, gasification process convert waste into high-value products, such as substitute natural gas, transport fuels, etc. The most successful way for converting biomass sources into gaseous fuel is gasification. The gasification cycles allow the conversion of various organic feed, such as wood, agricultural residues, peat, coal, anthracite, oil residues and municipal solid waste into syngas, bio-oil and bio-char. Also, the biomass resources can range from very clean wood chips at 50% moisture, to urban wood residues that are dry but contaminated with ferrous and other materials, to agricultural residues, to animal residues, sludge, and the organic component of municipal solid waste. The gasification cycles can convert these substances into carbon and hydrogen rich fuel gaseous that can be more simply used, often with the gain in efficiency and environmental performance compared to direct combustion of biomass resources [29]. The biomass resources are the complex mixture of organic compounds and polymers, and also have low ash, nitrogen, and sulfur ingredients. On the other hand, some agricultural materials, such as straws and grasses, have substantially higher amounts. In order to evaluate yields during gasification process, the complex samples must be reduced to a simplified chemical structure, such as CHmOn. In this study, chemical elements, such as sulfur and nitrogen, are considered to be present in very small amounts and have negligible contribution to the outputs [29]. The chemistry of gasification cycles is very complex because there are different processes occurring simultaneously. The early phases of gasification process are the partial oxidation and pyrolysis, which happens in the nonexistence of O2. The partial

376

Integrated Gasification Combined Cycles

Pretreatment: • Drying • Chipping

Raw biomass

Gasifier

Gas cleaning section

Reformer for higher hydrocarbons

Steam

Syngas storage

Combined cycle/ boiler with steam turbine

Electricity

Fig. 13 Schematic diagram of an integrated gasification combined cycle (IGCC) power generation facility based on biomass resources.

oxidation begins when the carbonaceous fuel sources reacts with less than the stoichiometric amount of O2. After all of the O2 in combustion chamber is consumed, the high working temperature of the gasifier devolatilizes the volatile parts of the raw material. This system creates the two-phase process occurring of the gas phase and carbon-based solid phase known as char. The primary parts of synthesis gas, such as CO and H2, are then generated by steam reforming of the carbon-based char. If the temperature of the gasifier chamber is high enough, all the char can be converted to carbon monoxide and hydrogen, and also the water–gas reaction works to completion. The most desired output of the IGCCs is generally hydrogen. In the recent phase of gasification cycle, the steam reforming of carbon monoxide generates hydrogen according to the water-gas shift reaction (WGSR). Different from the partial oxidation and steam reforming process, the WGSR does not convert all reactants into products. The extents of gasification reactions can be determined by using the reaction equilibrium point, which is a function of gasifier temperature and steam concentration of the gasifier chamber. The partial oxidation reaction, water gas reaction, and WGSRs are illustrated in the following equalities, respectively: C þ O2 þ heat-CO2

ð1Þ

C þ H2 O þ heat-CO þ H2

ð2Þ

CO þ H2 O þ heat2 CO2 þ H2

ð3Þ

C þ 2H2 -CH4 þ heat

ð4Þ

The one-sided arrows in Eqs. (1) and (2) represent that the chemical reactions are irreversible in the forward direction. The double-sided arrow for Eq. (3) illustrates that the chemical reaction can proceed in either the forward or reverse direction. The equilibrium condition can be achieved when the rate of change of the forward and reverse chemical reactions are equivalent. The advanced gasification cycles should be integrated with electricity production processes, acting as a link between coal or heavy fuel oils and gas turbines and its subsequent steam cycle. The schematic diagram of an IGCC power generation facility based on biomass resources is shown in Fig. 13. The produced synthesis gaseous from gasification cycle can be cleaned to very low levels of pollutants, such as sulfur compounds and particulates [20]. After cleaning process, the synthesis gases can be used in gas–steam turbine combined cycle power plants to generate power more efficiently than traditional combustion-based processes [30]. The resulting system mode of IGCCs is the only for electricity generation aims, including burning coal or high sulfur residues that can approach the technical and environmental efficiency of natural gas-fired processes. The environmental damage of IGCCs can be reduced even further when integrated with carbon capture and storage methods. The oxygen-fired gasification cycles can be integrated to generate a material stream composed almost entirely of carbon dioxide, which can be captured and sequestered [31]. Inlets to a common gasification chamber include feedstock, steam, and air. The produced synthesis gases are cooled and cleaned of impurities before combustion in the gas turbine. The resulting hot exhaust gases from gas turbine are then used in a heat recovery steam generator (HRSG) to boil high pressure (HP) steam for expansion in a steam turbine. The low pressure exhaust from the steam turbine is then recycled back into the HRSG and reused.

4.10.3.2

Coal Gasification

Meeting the energy needs has become a significant indicator following the inevitable rise in the world population today. On the other hand, energy consumption has been constantly increasing with high living standards and urbanization. One of the most significant sources for power generation is coal sources, which are the largest fossil fuel energy in the world, with proven reserves that are adequate to meet the expected need, without much rise in production costs. Coal sources are playing a significant indicator in delivering energy access, because it is widely available, safe, reliable, and relatively low cost. Over 7400 Mt of coal is currently produced, and also this production rate has increased 38% during the past two decades. Annual electricity generation rate from coal-fired power plants for top 10 countries are given in Table 3. Nearly 40% of electricity worldwide is generated based on coalfired technologies. This rate is much higher in many countries, such as Poland relies on coal sources for 94% of its electricity, South

Integrated Gasification Combined Cycles

Table 3

377

Annual electricity generation from coal-fired power plants

Countries

Annual electricity production from coal (TWh)

China United States India Japan Germany South Africa Korea Australia Russia Poland Rest of the world World

2010

2011

2012

2013

2014

2015

2016

3273 1994 653 304 274 242 219 181 166 138 1254 8698

3723 1875 715 281 272 243 225 173 164 141 1332 9144

3785 1643 801 303 287 239 239 171 169 142 1387 9168

4111 1712 869 337 293 237 223 161 162 140 1388 9633

4115 1713 967 349 285 232 232 152 158 132 1372 9707

4109 1471 1042 343 284 229 237 159 159 133 1372 9538

4008 1287 1125 336 282 226 239 161 162 131 1368 9325

Source: Data taken from International Energy Agency. Key world energy statistics reports. Available from: https://www.iea.org/publications/freepublications/publication/key-worldenergy-statistics-2017.html; 2017.

Table 4

Annual coal production rate of countries

Countries

Annual coal production (Mt)

China United States India Australia Russia South Africa Germany Poland Indonesia Kazakhstan World

2000

2005

2010

2012

2013

2014

2015

2016

1355 972 336 307 242 224 205 163 79 77 4726

2317 1039 437 371 285 245 206 160 171 87 6103

3316 996 570 436 300 255 184 133 325 111 7485

3678 932 603 435 331 259 197 144 446 121 8208

3749 904 610 458 328 256 191 143 490 120 8275

3640 918 657 489 334 261 187 137 485 114 8198

3527 824 683 512 348 251 184 136 488 107 7961

3210 683 708 509 359 250 177 131 459 102 7460

Source: data taken from Enerdata. Global energy statistical year book. Available from: www.yearbook.enerdata.net; 2017.

Table 5

Annual coal production rate of regions and world

Region

Asia Pacific Europe and Eurasia North America Africa South and Central America Middle East World

Annual coal production (Mt) 2000

2005

2010

2012

2013

2014

2015

2016

2190.7 1194.7 1054.4 230.5 53.7 1.54 4725.6

3439.9 1229.7 1107.6 250 73.9 1.99 6103.2

4853.9 1220.8 1067 258.9 83.2 1.52 7485.4

5530.3 1305.2 1004.6 267.3 98.7 1.46 8207.6

5673.5 1257 976.5 267.6 98.5 1.53 8274.6

5621.4 1207.4 989.5 276.3 101.8 1.35 8197.7

5528.6 1180 887.2 266.1 97.8 1.5 7961.2

5202 1162 728.9 264.2 101.6 1.5 7460

Source: Data taken from Kaiho M, Yamada O. Stoichiometric approach to the analysis of coal gasification process. In: Innocenti A, Kamarulzaman N, editors. Stoichiometry and materials science – when numbers matter. New York, NY: In Tech; 2012 [chapter 16].

Africa for 92%, China for 77%, and Australia for 76%. Tables 4 and 5 illustrate the top coal production countries and regions in the world, respectively. The coal production rate is expected to reach over 7 billion tonnes in 2030. It has been estimated that there are over 1.1 trillion tonnes of discovered coal reserves worldwide, which shows (at current rates of production and consumption) there being sufficient supplies of coal sources to last for approximately 220 years. In this decade, coal has the fastest growing consumption rate among

378

Integrated Gasification Combined Cycles

the other fossil-based energy sources. In addition, the world currently consumes more than 4 billion tons of coal. Coal has also important role for steel and cement production, as well as other manufacturing industries. Consequently, coal sources will continue to play an important role in meeting future global energy demand. It is not possible to assess the energy sources independent from their impact on the environment. CO2, SO2, NOx, and CO emissions from fossil combustion plants play a significant role on the atmospheric and environmental pollution. The dissimilar utilizes of coal sources can be separated into four main groups, such as (1) combustion, (2) pyrolysis, (3) liquefaction, and (4) gasification. In combustion, coal is directly burned to generate heat energy. In pyrolysis technique, coal sources are decomposed through heating in the absence of oxygen. In liquefaction technique, coal sources are converted into liquid fuel. In gasification technique, coal sources are converted into synthesis gas (syngas). New coal power plant designs with lower harmful emissions are imperative in the coming decades to mitigate global temperature increase. On the other hand, clean coal technology is very important way for developing novel and innovative technologies to decrease the negative impacts from utilizing coal sources for energy production. The coal sources are solid fuel, and are less suitable for storage and transportation application than petroleum and natural gas, and generally hold unwanted chemicals, such as sulfur, nitrogen, and various others. At the coal gasification process the energy of solid fuels is converted into a combustible gaseous mixture under HP and high temperature. The gasification processes convert coal sources into H2, CO, and CH4 by the reacting with gasifying agents, such as O2 and H2O. The fossil-based fuels are still the dominant source of electricity production; particularly coal, which has the highest share among these fossil fuels. The coal energy is the most abundant fossil energy source in the world, as well as the most common primary energy source used in electricity generation cycles. The coal consumption rate for power generation also contributes remarkably to harmful gaseous emissions based on current electricity production processes. For this reason, increasing performance and performing low carbon emission technologies can decrease the related ecological effect greatly. Recently, the gasification cycles are one of the most promising techniques to make utilize of coal source with higher performance and lower environmental harmful effects. The novel and more beneficial coal gasification systems are necessary in order to provide the highest acceptable range in the variation of fuel for combustion conditions and therefore higher performance of the IGCCs. The coal gasification processes include three main steps, such as (1) conversion of coal feedstock with steam to a synthesis gas, (2) catalytic shift conversion, and (3) product purification. In the first step, coal is chemically reacted with high temperature (approximately 13301C), HP steam to produce raw syngas. In the second step, the synthesis gas passes through the shift reactor converting a portion of the CO to CO2. In the third step of coal gasification process, the produced syngas is purified. The physical absorption removes 99% of hydrogen sulfide (H2S) impurities, and about 85% of the H2 in the shifted synthesis gas is removed as 99.999% pure H2 in a pressure swing adsorption (PSA) unit. If CO2 sequestration is applied, the secondary absorption tower removes CO2 from the remaining shifted synthesis gas. The synthesis gas can be burned into the combustor with air or oxygen to generate power. There are three kinds of commercialized processes for gasification as given below: 1. The fixed bed gasifier; the lump coal is gasified in the shaft reactor process at nearly 900–10001C. 2. The fluidized bed gasifier; the crashed coal sources are gasified in the fluidized reactor at approximately 11501C. 3. The entrained bed gasifier; the pulverized coal sources are gasified by burner system at around 1350B16001C. Compared to each other the last type runs at significantly higher pressure and temperature rates, which in fact leads to a better gaseous quality, but also results in higher working costs by the extensive combustion of oxygen. The coal sources supplied to the gasification chamber are usually decomposed thermally to generate gaseous, such as H2, CO, CO2, H2O, and CH4, tar, and char. The tar and char react with O2 and H2O supplied to form H2, CO, CO2, and CH4. As the gasification factor, the oxygen and steam mixture, pure oxygen or air can be utilized in the gasification chamber. The gasification reaction is composed of various types of chemical processes, such as pyrolysis of coal sources, decomposition of tar, oxidation of char, combustion of gas, shift reaction, and formation of various organic compounds. The following chemical reactions occur in the gasification chamber: C þ H2 O2 CO þ H2

ð5Þ

C þ 1=2O2 2 CO

ð6Þ

C þ O2 2 CO2

ð7Þ

C þ 2H2 2 CH4

ð8Þ

C þ CO2 2 2CO

ð9Þ

CO þ H2 O2 H2 þ CO2

ð10Þ

CO þ 3H2 2 CH4 þ H2 O

ð11Þ

The result of these chemical reactions is the syngas mixture, which predominantly consists of hydrogen and carbon monoxide. The other components are carbon dioxide, steam, and methane. Moreover, the raw mixture includes extraneous substances and pollutants. For this reason, depending on the application state, the raw gas mixture can be cooled down at first, separated from

Integrated Gasification Combined Cycles

Electricity Hot water Space heating Cooling Air conditioning

Water

Rankine cycle

Thermal storage

379

Brayton cycle

Boiler

Electricity

Air Waste heat Hot water

Cooling unit Screw feeder

Coal gasification

Coal Air

Steam

Combustion chamber

Syngas cooling

SO2 removal

Oxygen Air separation

Sulfur recovery Sulfur

Nitrogen

Fig. 14 Schematic diagram of an integrated gasification combined cycle (IGCC) power generation facility based on coal sources.

dust and desulfurized. At the end, the gas washing process should take off to remove the undesirable accompanying substances. The coal-based IGCC process includes the complete gasification of sources under HP, and this process is illustrated in Fig. 14. The syngas, generated being fairly hot, should be cooled down in the HRSG for additional useful products generation. The cooled syngas is then desulfurized in the sulfur recovery subcomponent, and then send to the combustion chamber of Brayton cycle to generate power. The exhaust gases from the gas turbine should be utilized to heat the water for boiler subcomponent. The steam from boiler generates additional power in the Rankine cycle. The waste heat from condenser unit of Rankine cycle should be utilized for cooling, heating, or hot water production aims. This IGCC produces very low levels of harmful emissions. But, the cost effectiveness and system performances should be investigated carefully against different process designs.

4.10.3.3

Stoichiometric Method of Gasification Cycles

The equilibrium constant is usually utilized to define the composition of gasification outputs. This constant is, whereas, only available to the stable state, and hence unsuitable to the examination of transient composition of gaseous in process, which is constantly unstable. The chemical cycle in the large-scale gasification chamber cannot be exactly described by using the kinetics and equilibrium states. The combination of gases generated by gasification process is a very clear measure of the chemical state in the gasification chamber. In this study, the stoichiometric approach is accepted to investigate the coal gasification reaction structure. Therefore, the hypothetical chemical definition is improved depending on the ultimate analysis results of source feedstock. The gasification reaction can be defined as follows: CHm On þ aO2 þ bH2 O-gH2 þ δCO þ eCO2 þ κCH4 þ ζC2 H4 þ yC2 H6 þ lCHm0 On0

ð12Þ

where CHmOn and CHm On are coal sources and tar, respectively [32]. The elemental balance equations for C, H, and O in Eq. (12) can be defined as follows: 0

0

1 ¼ δ þ e þ κ þ 2ζ þ 2y þ l

ð13Þ

m þ 2b ¼ 2g þ 4κ þ 4ζ þ 6y þ m0 l

ð14Þ

n þ 2a þ b ¼ δ þ 2e þ n0 l

ð15Þ

The total moles of produced gas in Eq. (12) can be defined as follows: F¼gþδþeþκþζþy

ð16Þ

When the concentrations of H2, CO, CO2, CH4, C2H4, and C2H6 (dry and N2 free) are represented by p, q, r, s, t, and u, respectively, the mole numbers of all gases can be defined as follows: H2 : g ¼ pF

ð17Þ

CO: δ ¼ qF

ð18Þ

CO2 : e ¼ rF

ð19Þ

380

Table 6

Integrated Gasification Combined Cycles

Coefficients of all components in gasifier based on Oex value

a b g δ e κ ξ y l

OexZ0

Oexr 0

0.5(1 n) þ Oex y x z 2w 2u (1 n0 )v 0.5m þ y x 3z 4w 5u {(1 n0 ) þ 0.5m0 }v 1 2Oex y þ x z 2w 2u v 2Oex þ y x Z w u v

0.5(1 n) þ Oex 2Oex þ y z 2w 2u (1 n0 )v 0.5m 2Oex þ y 4w 5u {(1 n0 ) þ 0.5m0 }v 1 y I 2w 2u v y z w u v

CH4 : κ ¼ sF

ð20Þ

C2 H4 : ζ ¼ tF

ð21Þ

C2 H6 : y ¼ uF

ð22Þ

l ¼ vF

ð23Þ

The reference state should be essential to explain the reaction process given in Eq. (12). CHm On þ 0:5ð1

nÞO2 -0:5 mH2 þ CO

a ¼ 0:5ð1

ð24Þ ð25Þ

nÞ þ Oex

In the Oex40 case when z mole of CH4 is created, the formula can be written as follows: CHm On þ f0:5ð1 þ ð2Oex þ y

nÞ þ Oex gO2 þ ðy

x

zÞH2 O-ð0:5 m þ y

x

3zÞH2 þ ð1

2Oex

yþx

zÞCO

xÞCO2 þ zCH4

ð26Þ

In the Oexo0 case, the formula can be defined as follows: CHm On þ f0:5ð1

nÞ þ Oex gO2 þ ð 2Oex þ y

zÞH2 O-ð0:5m

2Oex þ y

3zÞH2 þ ð1

y

zÞCO þ yCO2 þ zCH4

ð27Þ

From the two formulas given above, the previously defined coefficients of each component can be defined numerically as illustrated in Table 6. The heat of gasification reaction, hr (kcal/mol-coal) can be defined as follows: hr ¼ ghH2 þ δhCO þ yhCO2 þ κhCH4

4.10.3.4

hcoal

ð28Þ

Co-Gasification

The co-gasification processes for coal with biomass sources can be advantageous for different reasons, such as high oxygen content of biomass results in less external oxygen addition to the gasification system. If biomass compared to coal, biomass sources contain less ash, sulfur, and nitrogen components. At the same time, biomass sources have the comparatively low calorific value than coal, which is not desired in the IGCCs. In actual gasification techniques, the biomass sources have advantageous position than coal in kinetic reactions concerning carbon decomposition at the uniform gasifier temperature [16]. For optimizing the cogasification of coal and biomass cycles, different indicators can be changed, such as feed particle sizes and the ratio of biomass to coal. The mixing of coal and biomass has an interactive impact concerning ash creation and harmful emissions.

4.10.3.5

Fischer–Tropsch Process

Nowadays, the conversion technologies of synthesis gaseous (CO and H2) through gas-to-liquid process have demonstrated to be a perfect option to conventional resources of liquid transportation fuels [33,34]. This by itself, along with a worldwide increasing application for clean-burning fuels, has encouraged an important attentiveness in the examining of Fischer–Tropsch Process [35]. Recently, the commercial Fischer–Tropsch processes are utilized for the coal-to-liquid and gas-to-liquid conversion purposes [36]. The aim of such plants is to convert solid or gaseous carbon-based energy sources into useful outputs that may be utilized as clean-fuels or chemicals. The kinds of supply matters that should be converted to syngas are not only limited by coal and natural gas sources. It is reasonable to utilize practically any other carbon source as supply matters. The conversion of biomass resources in the biomass-to-liquid cycle and waste in the waste-to-liquid cycle are very important models of sustainable technics, because biomass sources presents the renewable energy resource and waste conversion relates to the useful recycling of discarded substances. The Fischer–Tropsch processes are the catalytic cycles that convert synthesis gaseous into the multielement mixture of

Integrated Gasification Combined Cycles

Electricity Nitrogen

381

Combined cycle/ boiler with steam turbine

Biomass or coal Gasifier

Tar removal

Reformer

CO + H2O ↔ CO2 + H2

Compressor Ash Fresh air

Tar

FT reactor

Shift reactor

H2S and COS

Diesel fuel Wax

Product recovery and upgrading

Fig. 15 Schematic diagram of Fischer–Tropsch process.

hydrocarbons. The produced outputs from Fischer–Tropsch cycles are the great high-performance and clean diesel fuels, because of their high cetane number and lack of sulfur and aromatic complexes. The requirement to eliminate the leaving heat from Fischer–Tropsch process is a great importance in the design of reactors appropriate for the conversion of synthesis gaseous. This waste heat can be utilized by integrating the ORC to the Fischer–Tropsch process to recover the rejected heat from the exothermic reactions. The temperature level of waste heat is adequate to generate power and hot water for residential application with a suitable ORC process. The basic processes in producing Fischer–Tropsch liquids are illustrated in Fig. 15. After pretreatment process, to produce syngas mixture, the biomass resources are gasified in biomass gasifier subcomponent. The produced syngas is cleaned and its structure is modified to fit the specifications for Fischer–Tropsch synthesis in the Fischer–Tropsch reactor. The Fischer–Tropsch syngas is recycled for maximal Fischer–Tropsch liquids generation, or combusted in combustor unit to generate power from gas turbine system. After carbon dioxide separation process, the liquid Fischer–Tropsch products are treated to yield gasoline and/or diesel fuels. Different pretreatment process can be applied before gasification cycle, such as storage and size reduction. An important quantity of energy is required for drying process. The extracted heat energy from the process offgas or steam cycle can be used for this process. The performance of gasification process increases with drier biomass resources. But, the hydrogen content of generated synthesis gas decreases, which has negative effect for the Fischer–Tropsch synthesis downstream. The gaseous cleaning process may be the most important system component for biomass Fischer–Tropsch application. The gaseous cleaning process can be divided into two subcomponents, such as the tar removal and residual contaminants removal. The tar removal process should be done by cracking. The tar removal process should be achieved by using the cracking subcomponent. At temperatures above 12001C, the tars should be destroyed without the catalyst, usually by addition of steam and oxygen flows. The disadvantages of high temperature gas cleaning process are the expensive material requirement, soot production, and the loss of cold gaseous performance. The catalytic cracking avoids the mentioned problems of thermal cracking process. But, this technology is not yet fully proven, and also it is unclear whether the light aromatic compounds (BTX: benzene, toluene, xylene) are converted in gas cleaning process. Once tars are absent, further impurities, such as NH3, H2S, COS, HCI, volatile metals, dust, and soot, should be removed by using conventional wet gas cleaning techniques for improved dry gas cleaning techniques.

4.10.3.6

Hydrogen Production via Gasification Process

An expanding shortage of conventional energy sources for heat, electricity, and fuels is driving the research and improving of alternative resources [2]. Also, there is a need to decarbonize global energy systems to reduce the risks associated with climate change and GHG emissions [37]. Hydrogen is a promising clean energy carrier of the future, and a potentially best solution to climate change. Also, hydrogen is widely utilized in different industry practices. For producing hydrogen, the hydrogen bonds in water or any kind of hydrocarbon element must be broken which needs significant quantity of energy in the kind of power and heat. The steam methane reforming (SMR), gasification, and electrolysis are the most general utilizing technics for hydrogen generation [38]. Clean hydrogen generation is an important challenge for power facilities and has been the point of interest for many researchers [39–42]. Hydrogen can be produced by using different types of biomass; for instance, forestry, industrial, living organism, and municipal waste, and crops. Biochemical, biomass gasification, and thermochemical ways are usually used to produce hydrogen energy from biomass sources. The chemical reactions required for biomass sources based hydrogen production are completely comparable to those of fossil fuel-based ones. Among the available biomass gasification-based hydrogen production ways, the gasification cycles are already the valid alternative, while simulated photosynthesis is the most favorable one. The chemical reactions included in biomass gasification-based hydrogen production are quite comparable to those of fossil fuelbased production pathways. Biomass is stated as a plenty alternative energy source (participating to 12% of the worldwide energy

382

Integrated Gasification Combined Cycles

Biomass

Gas clean up process

Gasification process

Hydrogen liquefaction process

Liquid hydrogen

Compression process

Low temperature shift reaction

Catalytic steam reforming

High temperature shift reaction

Fig. 16 Biomass gasification-based hydrogen production by reforming syngas.

Biomass

Gas clean-up process

Gasification process

Liquid hydrogen

Compression process

Combined cycle power generation

Hydrogen liquefaction process

High/low temperature electrolyzer

Fig. 17 Biomass gasification-based hydrogen production by electrolysis.

supply), and could possibly reduce carbon dioxide emissions, on condition that the carbon dioxide emissions are absorbed back by the biomass itself in the course of photosynthesis [43]. The schematic diagram of two ways for hydrogen generation from biomass sources is illustrated in Figs. 16 and 17. Both processes use the low pressure indirectly heated biomass gasification cycle for the syngas generation. The first process uses synthesis gas by using the catalytic steam reforming and high/low temperature WGSRs for hydrogen generation. The other process uses synthesis gas for power generation by using the integrated system and then hydrogen generation by high or low temperature electrolysis is produced. The catalytic steam reforming system is similar to that utilized in hydrogen generation of natural gas. The reforming system is mainly consisted of two subcomponents, such as (1) primary reformer (to convert the higher hydrocarbons present in synthesis gas into hydrogen), and (2) high/low temperature WGSR. The chemical reactions of steam reforming for synthesis gas can be defined as follows: Cn Hm þ nH2 O þ heat-nCO þ ðm=2 þ nÞH2

ð29Þ

CO þ H2 O-CO2 þ H2 þ heat

ð30Þ

The methane and hydrocarbons in biomass sample are converted into carbon monoxide and hydrogen in the primary reformer subcomponent (Reaction (12)). Also, the carbon monoxide and water are converted into hydrogen in the high/low temperature WGSR subsystems (Reaction (30)).

4.10.3.7

Methanol Production via Gasification Process

In theory, the large numbers of methanol production routes based on gasification process involving conventional and advanced technologies are possible. The methanol can be produced by the reaction of hydrogen with carbon oxides as follows: CO þ 2H2 -CH3 OH

ð31Þ

CO2 þ 3H2 -CH3 OH þ H2 O

ð32Þ

and

As seen in Fig. 18, the methanol and hydrogen production system typically consist of the following basic subcomponents, such as (1) pretreatment, (2) gasification, (3) gas cleaning, (4) reforming of higher hydrocarbons, (5) shift to obtain appropriate H2/CO ratios, and (6) gas separation for hydrogen production or methanol synthesis and purification. For optional ways, the gas turbine or boiler, and the steam turbine can be used for power coproduction. The produced gas in gasifier includes the tars, dust, alkali compounds and halogens, which can block or poison the catalysts downstream, or corrode the gas turbine. The syngas should be cleaned by using conventional low temperature techniques (e.g., water scrubbing). Instead of this, the hot gas cleaning technology should be considered, using ceramic filters and reagents at 350–8001C. The hot gas cleaning

Integrated Gasification Combined Cycles

Pretreatment: • Drying • Chipping

Gasifier

Raw biomass

Reformer for higher hydrocarbons

Gas cleaning section

Methanol production

Shift to adjust CO/H2 ratio

H2 separation

383

Methanol

Hydrogen

Steam

Fig. 18 Schematic diagram of methanol production based on gasification process.

Air

Water O2 ASU

Raw materials Syngas cooling/ cleaning

Gasifie

Slag

Shift reactor

Acid gas removal H2S

O2

Claus unit Sulfur

N2

Ammonia synthesis

H2

Ammonia

PSA PSA Offgas

Fig. 19 Schematic diagram of ammonia production based on gasification process. ASU, air separation unit; PSA, pressure swing adsorption.

technology is favorable for the overall energy balance when followed by a reformer or ceramic membrane (require high inlet temperature). But, not all elements of hot gas cleaning are still approved techniques. The synthesis gas can include a substantial quantity of methane and other light hydrocarbons, representing an important part or the heating value of produced gas. The reforming processes convert these compounds to carbon monoxide and hydrogen by steam addition over the nickel catalyst. The conventional methanol production systems utilize the fixed beds and run in the gaseous phase. The conversion rate is limited by the equilibrium and by the high temperature sensitivity of catalysts. The methanol production systems under improving at nowadays focus on shifting the equilibrium to the product side to reach higher conversion ratio per pass, and effective heat removal. In slurry phase systems, the heat removal is extremely efficient, therefore allowing high conversions per pass without loss of catalyst activity. Furthermore, the conversion reaction happens at the catalyst surface. The shift reaction per pass is 15% to 40% per carbon monoxide rich gaseous and 40% to 70% for balanced and hydrogen rich gaseous, and future carbon monoxide conversion up to 97% will predicted. The capital cost for slurry systems are expected to be 5% to 23% less than for the gaseous phase system. After reforming and shifting to the H2 rich synthesis gas, and also H2 can be separated and compressed for later utilization. The PSA process separates components of a gaseous flow by selective adsorption to a solid at HP, and subsequent desorption at low pressure. The membranes are the favorable technique for gaseous separation. Also, the membranes are attractive because of their simply model, may have the potential of combining shift and separation in one reactor, and have potentially better economies than traditional separation techniques. The unshifted fuel gases that remain after the methanol and hydrogen generation process can still include an important amount of chemical exergy. These gaseous flows may be combusted in the combustion chamber for electricity generation from gas turbine.

4.10.3.8

Ammonia Production via Gasification Process

The gasification cycles can be used to convert different raw materials into the syngas for usage in making ammonia. In these cycles, the raw materials are reacted with O2 and H2O to generate synthetic gaseous. Fig. 19 is the schematic diagram, which shows the gasification cycle for generating the synthesis gaseous for ammonia production. The ASU separates air into oxygen and nitrogen. The produced oxygen enters the gasifier subcomponent and also Claus unit. The Claus process section includes two chemical reactors. The H2S rich gas enters the first reactor, where it reacts with O2 to produce SO2: 2H2 S þ 3O2 2 2SO2 þ 2H2 O

ð33Þ

The gas exiting from this reactor is mixed with the rest of H2S rich gas in the second chemical reactor to generate elemental sulfur over a bauxite or alumina catalyst.

384

Integrated Gasification Combined Cycles 2H2 S þ SO2 2 3S þ 2H2 O

ð34Þ

With this configuration, it is possible to convert 96% of H2S of the entering steam into elemental sulfur. The carbon dioxide, carbon monoxide, hydrogen, and methane are produced in the gasifier. Syngas mix exiting from gasifier is cooled and cleaned up in the syngas cooling/cleaning subcomponent. CO and H2O are converted into CO2 and H2 in the WGSR. The WGSR can contain catalysts for the low temperature or high temperature conversion. The synthesis gas goes to the acid gas removal unit which removes hydrogen sulfide and carbon dioxide. The remaining gas mixture enters the PSA unit to separate hydrogen from other gaseous. The nitrogen coming from ASU and produced hydrogen enter the ammonia synthesis unit where ammonia is produced in the well-known reaction as follows: N2 þ 3H2 -2NH3

4.10.3.9

ð35Þ

Syngas Cleaning and Sulfur Removal Processes

To removal of particulate matters, sulfur and NOx, the gas cleanup processes must be used in the gasification cycles. These aims can be achieved as given below:

• •

Particulate matter removal: integration of cyclone-filters and ceramic candle filters. SOx and NOx removal: integration of steam/water washing and removing sulfur compounds for recovery of sulfur contents.

The hot gas cleanup processes are currently under demonstration step. The wet scrubbing processes, through with a lower performance, still remain the preferred technologies for gas cleanup processes in the IGCCs. The sulfur compound from hot syngas can be captured by reducing it to H2S, COS, CS2, and so on. Nowadays, the sulfur removal processes usually utilize the zinc-based regenerative sorbents, such as zinc ferrite, zinc titanate, and so on. Also, the sulfur compound is also removed by the addition of limestone in the gasifier subcomponent. This option is commonly utilized in the air-blown fluidized bed gasifier subcomponents. In process of air-blown gasifier, the sulfur compounds are captured in the gasifier bed itself (above 90%) because of the addition of limestone. The captured sulfur compounds in the gasifier bed can be removed with ash.

4.10.4

Thermodynamic Assessment of Integrated Gasification Combined Cycles

Thermodynamic performance of the IGCCs for multigeneration aims are investigated by conducting quantitative energy and exergy analyses. In this chapter, the assumptions, basic concepts, procedure, and equations used to evaluate integrated systems performance are defined and expressed. Throughout this chapter, the following assumptions are utilized:

• • • • • •

Environmental temperature (T0) and pressure (P0) are taken as 201C and 1 atm, respectively. All streams and components are at operating temperature and pressure at all times. All processes take place in steady state and steady flow conditions. Changes in potential and kinetic energies are negligible. Change in the control volumes (gasification reactor and integrated system units) is disregarded. The heat losses to environment are neglected.

4.10.4.1

Basic Thermodynamic Concepts

In this subsection, overall mass, energy, entropy, and exergy balance equations are defined based on the assumptions and operating conditions of the IGCCs for multigeneration to develop a clear understanding of the systematic approach. The heat and work input/output rates, entropy generation rates, and exergy destruction rates, and energy and exergy efficiencies can be calculated by using these balance equations.

4.10.4.2

Mass Balance Equation

The general conservation of mass in a control volume for any process can be defined in its most general arrangement as follows: Total mass flow Net change in ¼ rate entering the mass flow rate control volume

Total mass flow rate leaving the control volume

or; X dmcv _i ¼ m dt

X

_e m

ð36Þ

_ are the mass and mass flow rate, respectively. The subscripts cv, i, and e are the control volume, inlet condition of where m and m control volume, and outlet condition of control volume, respectively.

Integrated Gasification Combined Cycles

385

In the course of steady state and steady flow conditions, the Eq. (36) becomes as follows:

4.10.4.3

X

Energy Balance Equation

_i m

X

_e m

ð37Þ

The conservation of energy equation in any control volume can be defined by using the first law of thermodynamics, which states as follows: E1 ¼ δQ

E2

ð38Þ

δE

where heat and work interactions between the control volume and its surroundings are represented by Q and W, respectively. The initial and final states are denoted by subscripts 1 and 2, respectively. The energy (E) can be any form of potential, kinetic, and/or flow that a process could have at a given state. The general transient type energy balance equation can be defined as follows [44]: Time rate of

Net rate at

change of the energy

which enenrgy is being ¼ transferred in by heat

contained within the control volume

transfer

Net rate of energy transfer

Net rate at

which enenrgy into the control is being þ volume transferred accompanying out by work mass flow

or, dEcv _ cv ¼Q dt

_ cv þ W

X

  v2 _ i hi þ i þ gzi m 2

X

  v2 _ e he þ e þ gze m 2

ð39Þ

_ and W _ are the rates of heat and electricity flows inside the control volume, respectively. And also h, v, z, and g are the where Q specific enthalpy, velocity, altitude, and gravitational acceleration, respectively. The Eq. (39) can be rewritten by the steady state steady flow process assumption, and if the potential and kinetic energy differences are neglected, the energy balance can be defined as follows [45]: X X X X X X _iþ _eþ _ i¼ _e _ e he þ _ i hi þ Q W m Q W ð40Þ m

4.10.4.4

Entropy Balance Equation

Utilizing the second law of thermodynamics, the entropy balance equation can be defined as follows: X dScv _ i si m ¼ dt

X

_ e se þ m

XQ _ cv T0

þ S_ gen

ð41Þ

where s stands for specific entropy and S_ gen is the rate of entropy generation. Unlike energy, entropy is not conserved; it is generated during a system due to process irreversibilities. Therefore, the amount of entropy leaving the control volume exceeds the input entropy due to entropy generation associated with irreversibilities. During the steady state steady flow conditions, the Eq. (41) becomes as follows: X

4.10.4.5

_ i si þ m

XQ _ cv T0

þ S_ gen ¼

X

_ e se m

ð42Þ

Exergy Balance Equation

Exergy can be described as the maximum work that can be extracted from a process interacting with its reference surroundings [46]. Performing the exergetic assessment in addition to energetic assessment is therefore expected to provide valuable additional insight, thereby contributing to the development of sustainable energy systems for the future [47]. Similar to entropy, exergy is independent from the conservation law. The exergetic assessment accepts that energy cannot be created or destroyed but it can be degraded in quality, finally reaching a complete equilibrium with the ambient condition and therefore of no further utilize for performing tasks [46]. The exergy balance is a statement of law of energy degradation as it defines the irretrievable loss of exergy due to system irreversibilities [48]. The exergy balance equation for process parts in the general style can be written as given below Rate of exergy change

¼

Rate of exergy

Rate of exergy

transfer

destruction

or, X dEx cv _ Q ¼ Ex dt

X

_ Wþ Ex

X

_ in Ex

X

_ out Ex

X

_ D Ex

ð43Þ

386

Integrated Gasification Combined Cycles

_ Q and Ex _ W indicate the exergy transfer rates associated with heat and boundary or shaft work, respectively. The exergy where Ex _ D. destruction rate defines the process irreversibility and it is illustrated in the equation as Ex The steady state and steady flow exergy balance equation can be given as follows [47]: X

4.10.4.5.1

_ i exi þ m

X

_ Qþ Ex i

X

_ W¼ Ex i

X

_ e exe þ m

X

_ Qþ Ex e

X

_ W þ Ex _ D Ex i

ð44Þ

Exergy analysis of heat transfer

_ in a control volume at an operating temperature of T, the highest When there is heat transfer included in a process with a rate of Q conversion rate from thermal energy to desired work, which states the thermal exergy flow, can be defined as follows [49]:   _ _ Q ¼ 1 T0 Q Ex ð45Þ T where (1 T0/T) is the dimensionless exergetic temperature, usually defined as Carnot efficiency working between the ambient temperature at T0 and the process temperature at T.

4.10.4.5.2

Exergy analysis of work transfer

From the basic description, the work equivalent of a given condition or energy is an evaluation of its exergy [48], it can be defined that the exergy transfer rate with shaft or boundary work equals to work and the exergy transfer rate can be defined as well by the power or the work transfer rate. The following equality illustrates the exergy transfer rate associated with work considering the change of the volume as given below: _ þ P0 dVcv _ W ¼W ð46Þ Ex dt where P0 is the reference pressure. If the control volume is assumed to be constant, then the exergy transfer rate can be defined as follows: _ _ W ¼W Ex

4.10.4.5.3

ð47Þ

Exergy analysis of flow

Exergy of a flow of a given matter can be defined as the maximum quantity of work that can be provided when the flow is taken away from its original state to the environmental state throughout a system of interactions with its surroundings [48]. The exergy transfer rate by the flow through any process part can be defined in terms of the specific flow exergy as given below: X X _ flow ¼ _ i ex i Þ ðm ð48Þ Ex Exergy related to a stream flow is consisted of four major elements: physical exergy (exph), chemical exergy (exch), potential exergy (expt), and kinetic exergy (exkn) [44]. Therefore, the specific flow exergy can be defined as follows: ex ¼ exph þ exch þ ex pt þ exkn

ð49Þ

The flow associated specific exergy of a system part (i) can be defined as follows: vi2 þ gzi ð50Þ 2 The kinetic and potential parts of exergy appear in the above equation are assumed to be negligible during the course of this chapter as the changes in velocities and within the process parts are unimportant compared to the contributions of other indicators. Therefore, the Eq. (50) can be written as follows: exi ¼ ex ph;i þ exch;i þ exke;i þ ex pe;i ¼ ðh

exi ¼ ex ph;i þ exch;i ¼ ðh

h0 Þi

h 0 Þi

T0 ðs

T0 ðs

s0 Þi þ exch;i þ

s0 Þi þ exch;i

ð51Þ

The physical exergy part of flow exergy rate is influenced by physical process including thermal interaction with the surroundings to bring the flow from its main case to the environmental case which is at a temperature and pressure T0 and P0, respectively. The specific physical exergy parts can be written as follows: exph ¼ h

h0

T0 ðs

s0 Þ

ð52Þ

where h and h0 are specific enthalpies, and s and s0 are the specific entropies at the real case and the reference environment states, respectively. The chemical exergy is the part of the flow exergy that is created by process involving heat transfer rate and exchange of materials with the surroundings to bring the material to the dead state [48]. The chemical exergy of an ideal gas mixture can be defined as follows [46]: X X ð53Þ yi ex ch yi ln yi exch mix ¼ i þ RT0 where yi is the mole fraction of the part i in the gas mixture.

Integrated Gasification Combined Cycles

The following general exergy balance equation based on the combining Eqs. (44) to (53) can be defined as      X  X  ph  ch ch _ i 1 T0 þ W _ e 1 T0 þ W _ i¼ _ e þ Ex _ D _ _ i ex i þ ex ch m þ Q m ex þ ex þ Q e e e i T i T e

4.10.4.5.4

387

ð54Þ

Exergy destruction rate

Quantity of exergy output rate from the control volume must be less than the exergy inlet rate due to exergy destruction rate within _ D in the given exergy balance equation, is directly interested to the the system. The exergy destruction rate, which is indicated as Ex entropy generation rate within the control volume, and can be defined as follows: _ D ¼ T0 S_ gen Ex

ð55Þ

The following assumptions are made for the thermodynamic analysis of IGCCs.

• • • • • •

pressure drops and heat losses in piping are negligible; the system parts are well insulated and therefore are taken as adiabatic; for the IGCCs for multigeneration aims, the steady state and steady flow operating conditions exist; kinetic and potential energy changes in all processes are negligible; dry and ash-free portion of the feedstock is completely reacted in the gasifier; and the reference temperature and pressure are taken as 251C and 101.3 kPa, respectively.

4.10.4.5.5

Energy efficiency analysis

As the measure of energy related performance, an energy efficiency of a system (Z) can be described as the ratio of beneficial production from the system boundary to the energy inlet to the system. P P useful output energy energy loss P ¼1 P ð56Þ Z¼ input energy input exergy

4.10.4.5.6

Exergy efficiency analysis

On the basis of the second law of thermodynamics, the exergy efficiency can be described based on the exergy content of the process inlets and outlets, which gives a better insight of process efficiency. Dincer and Rosen [46] describe energetic and exergetic efficiency and outline the key indicators of exergetic performance. Also, Dincer and Rosen [46] describe energetic and exergetic efficiency and outline the key indicators of exergetic performance. They define that exergetic performance usually provides the finer understanding of exergetic performance than energetic efficiency. The exergy efficiency opinion differentiates irreversibilities from losses, that supplies knowledge about the investigated process developments by reducing the losses. The exergy efficiency (c) equations for steady state conditions can be defined as follows: c¼

P

Useful output exergy P ¼1 Input exergy

P Exergy loss P Input exergy

ð57Þ

The other exergy efficiency descriptions for steady state conditions can be defined as follows [50]: RE ¼

Total exergy output ¼1 Total exergy input

TE ¼

Exergy consumption Actual exergy input

ð58Þ

Theoretical minimum exergy input required Actual exergy input

ð59Þ

where RE and TE are the rational efficiency and task efficiency, respectively.

4.10.4.6

Specification of Coal Sources

In this section, the calculations of gross calorific value (GCV), net calorific value (NCV), entropy and chemical exergy for coal samples are briefly defined in the following equations. The formula depending on the weight percentages of main coal elements, which estimates of the GCV of coal should be written as follows [51]: GCV ¼ 326  ½0:198ð%CÞ þ 0:6203ð%HÞ þ 0:0809ð%SÞ þ 0:04495ð%AÞ

5:153Š

ð60Þ

For the dry, ash-free basis GCV calculation, the equation can be defined as given below [52], where c, h, o, n, s are the molar concentrations of carbon, hydrogen, oxygen, nitrogen, and sulfur in kilograms of dry and ash-free coal basis, respectively. The GCV and GCVDAF are defined in MJ/kg basis. GCV DAF ¼ ð151:19  h þ 98:767Þ  ½c=3 þ h

ðo

sÞ=8Š

ð61Þ

It is beneficial to explain the net calorific value on original basis; this can be done with the equality given by Van Loo and Koppejan [53]:

388

Integrated Gasification Combined Cycles NCV ¼ GCV ð1

ww Þ

21:839wH ð1

2:444ww

ww Þ

ð62Þ

where ww is the moisture content by weight. The calculation of coal entropy and chemical exergy is very important for exergy analysis of energy systems involving coal combustion. The specific exergy content of dry and ash-free coal samples can be defined as follows [54]:    h o n s þ 20:1145 ð63Þ þ 54:3111 þ 44:6712 0:564682 sDAF ¼ c 37:1653 31:4767 exp cþn cþn cþn cþn where subscript DAF stands for dry and ash-free basis, and the entropy unit is kJ/kg K. The chemical exergy content of coal samples, dry and ash-free basis can be defined as follows [55]:    1 1 1 1 o sO2 c  sCO2 h  sH2 O s  sSO2 n  sN2 T0 sDAF þ c þ h þ s exch DAF ¼ GCV DAF 4 2 2 2   1 1 1 1 ch ch ch o ex ch cþ hþs þ c  ex ch CO2 þ h  ex H2 O þ s  ex SO2 þ n  ex N2 O2 2 2 4 2

ð64Þ

where the exergy value results in MJ/kg for dry coal in ash-free basis. In addition, the chemical exergy of high fixed carbon containing coals varies between 7 and 8.2 MJ/kg dry ash-free basis. The specific chemical exergy of coal with wet basis can be calculated with neglecting the ash content. Because the contribution of ash to chemical exergy is negligible [56]. ch ex ch coal ¼ ð%coalDAF Þ  ex DAF þ

ð%H2 OÞ  ex ch H2 O MWH2 O

ð65Þ

where (%coalDAF) is the weight percent of dry ash-free coal; (%H2O) is the weight percent of water; MW H2 O is molecular mass of water.

4.10.4.7

Thermodynamic Analysis of a Gasification System

The chemical analyses of gasification systems (see in Fig. 20) are very complex because there are different processes occurring simultaneously. For developing the gasification design, the chemical formula of raw materials is considered as CHaObNc and the common gasification reaction is described as below in Eq. (60), where a, b, and c demonstrate the number of atoms for H2, O2, and N2 per number of atom of carbon, respectively. c  þ 3:76w N2 ð66Þ CHa Ob Nc þ mH2 O þ wðO2 þ 3:76N2 Þ ¼ nH2 H2 þ nCO CO þ nCO2 CO2 þ nH2 O H2 O þ nCH4 CH4 þ 2 For the gasification chamber at a steady state condition as illustrated in Fig. 20, the inlet mass flow rate of gasifier chamber is equal to the outlet mass flow rate of gasifier chamber. Therefore, the mass balance can be defined as follows: N X

_i¼ m

i¼1

M X

_e m

ð67Þ

e¼1

where N and M are the total numbers of inlet and outlet flows, respectively. Also, the mass flow rate of gasification chamber can be defined based on the molar rate. N X

i¼1

n_ i MW i ¼

M X

_ e MWe m

ð68Þ

e¼1

where MW is the molecular weight. The energy balance equation of gasification chamber illustrated in Fig. 20 can be defined as follows: N X

_ þ E_ fuel ¼ _ i hi þ Q m

M X

_ e he m

ð69Þ

e¼1

i¼1

The energy flow rates of the fuel entering to the gasification chamber can be defined based on the LHV, and it can be given as follows: Q Mass in

Mass out

Gasification

Energy in Exergy in

e=N

i=N

i=1

e=1 Efuel

Fig. 20 Schematic diagram of the gasification chamber.

Energy out Exergy out

Integrated Gasification Combined Cycles

389

Table 7 Standard chemical values of different substances at T0 ¼251C and P0 ¼101.3 kPa Substance

_ 0 (kJ/kmol) Ex ch;k

CO CO2 H2O C(s) H2 O2 N2

275,430 20,141 11,710 410,260 238,500 3,900 720

Source: Data taken form Dincer I, Rosen MA. Energy, environment, and sustainable development. Oxford: Elsevier; 2012.

_ fuel LHV fuel E_ fuel ¼ m

ð70Þ

Also, the LHV can be calculated via removing the latent heat of vaporization of the water vapor created in the combustion system from the HHV of the entering fuel. The exergy balance equality can be represented in the following form utilizing the exergy values of all flows entering and leaving the gasification chamber. The primary difference between the exergy and energy balance equalities is that exergy is not conserved in any process. It means that the exergy rate leaving the gasification chamber will always be less than the entering exergy rate. The exergy balance equation of gasification chamber can be defined as follows:   N N X X _ þ Ex _ fuel ¼ _ D _ i þ 1 T0 Q _ i þ Ex ð71Þ Ex Ex T i¼1 i¼1 The exergy contents of the gasifier input and outlet flows are divided into two main components.

_ ph;k þ Ex _ ch;k _ k ¼ Ex ð72Þ Ex _ _ where Exph and Exch are the physical and chemical exergy, respectively, and can be defined for gasifier chamber as follows: _ ph;k ¼ m _ ½ðhk Ex

h0 Þ

T0 ðsk

s0 ފ

ð73Þ

The chemical exergy content of the entering fuel to the gasifier can be defined by using the composition and concentration of mixture components: _ ch;k ¼ Ex

N  X

k¼1

N  X _ 0 ðxk lnxk Þ xk Ex ch;k þ RT0

ð74Þ

k¼1

_ 0 ) is defined as the maximum work that can be provided when the The standard chemical exergy of the chemical material (Ex ch;k investigated process is brought into reaction with reference substances present in the surroundings. Describing the exergy reference ambient is one of the most important components of examining chemical exergy rate [57]. Generally, the ambient can be described as the composition of air at 251C and the pressure of 101.3 kPa, respectively. The values of standard chemical exergy of different substances are given in Table 7. Generally, the chemical exergy of solid fuels can be calculated by using the below statistical correlation for different oxygen and hydrogen to carbon ratios [16]: _ ch;sf ¼ lLHV f Ex where subscripts sf and f are the solid fuel and fuel, respectively. The statistical correlation (l) can be defined for 0:5rO C r2 as follows, respectively: l ¼ 1:0438 þ 0:0158

H O þ 0:0813 C C

   

H 1:0414 þ 0:0177 H 0:3328 O C C 1 þ 0:0537 C   l¼ 1 0:4021 O C

ð75Þ O C r0:5

and

ð76Þ

ð77Þ

The LHV of fuel is the amount of heat released during its complete combustion. The heating value of synthesis gas with a molar basis can be calculated as the sum of the heat of combustion, at 251C and 101.3 kPa, multiplied by the mole fraction for each component in the synthesis gas. There are several energy efficiency definitions for gasification system. Generally, the cold gas efficiency and carbon conversion efficiency are utilized for process performance. The cold gas efficiency can be defined as follows: Zcge ¼

LHV pg LHV fs

ð78Þ

390

Integrated Gasification Combined Cycles

where LHVpg and LHVfs are the lower heating value of product gas and feedstock, respectively. The carbon conversion efficiency can be written as  Cgr Zcc ¼ 1 ð79Þ Cfs where Cgr and Cfs are the carbon content in gasification residue and feedstock, respectively. The exergy efficiency equation of gasification chamber can be described as the ratio between useful exergy outputs to the necessary exergy input to the gasification chamber. cgasifier ¼

4.10.4.8

_ syngas Ex _ steam þ Ex _ air _Ex fuel þ Ex

ð80Þ

Environmental Impact Analysis

The environmental impact assessment is the investigation of negatively or positively possible effects on the natural surroundings including ecological system, accounting for technic, economic, and social indicators. The environmental impact assessment incorporates the recognition, estimation, and evaluation of ecological impacts and mitigation alternatives. In this chapter, the normalized carbon dioxide emissions and sustainability analysis are utilized for environmental impact assessments of IGCCs.

4.10.4.8.1

Normalized carbon dioxide emissions

To define carbon dioxide emissions for the IGCC, three conditions are investigated and the carbon dioxide emissions are defined for each condition. In the first condition, the power cycle is utilized to generate electricity. In the second condition, the electricity and heating production processes are investigated simultaneously. In the last condition, the whole integrated process for multigeneration aims is investigated. The amount of carbon dioxide produced in each condition can be defined as follows: _ CO2 m _ net W

ð81Þ

_ CO2 m _ heating _ W net þ Q

ð82Þ

epower ¼ eCHP ¼

emulti ¼

_ net þ W

P

_ heating þ Q

_ CO2 m P _ cooling þ E_ HW þ E_ H2 þ E_ FW Q

ð83Þ

where HW and FW are the hot water and fresh water, respectively.

4.10.4.8.2

Sustainability analysis

To increase environmental sustainability, it is necessary not only to utilize sustainable energy resources, but also to use nonrenewable sources like coal more efficiently, and to limit environmental damage. In this method, communities can decrease its utilize of limited sources and extend their lifetimes. The sustainability index can be utilized to relate exergy with environmental impact [46]: SI ¼

1 Dp

ð84Þ

where Dp is the depletion number and can be defined as the ratio of exergy destruction to input exergy. This relationship shows how decreasing the environmental impact of process can be achieved by decreasing its exergy destruction. Furthermore, the sustainability index is then determined as a measure of how the exergetic performance affects sustainable development as given below: SI ¼

4.10.5

1 1

c

ð85Þ

Combined Cycles

In this section possible and potential ways for combustion chamber-based combined cycles are described by using thermodynamic balance equations in order to obtain to better integrated system efficiency. The working conditions, performance, and power range are investigated for each combined process, and their benefits and disadvantages are analyzed below. The combined cycles can be divided into two sub-cycles, such as (1) topping process and (2) bottoming process, based on their temperature range. Different basic cycles have the maximum-temperature near the flame temperature range within the combustor. Therefore, the heat rejection at high temperature in these main cycles occurs, and in order to reach higher performance, the combined cycles are required with the high temperature topping-cycle and also the medium or low temperature bottoming-cycle. The investigated combined cycles can be analyzed for possible efficiency gains and feasibility. To define what integrated cycle is better suitable for topping-cycle and bottoming-cycle usages, the integrated cycles can be classified based on their operating temperature range. The integrated cycles can

Integrated Gasification Combined Cycles

Table 8

Integrated cycles based on temperature range of topping and bottoming processes

Low temperature integrated cycles

Medium temperature integrated cycles

High temperature integrated cycles

Applications:

Applications:

Applications:

• • • • • •

391

Rankine/organic Rankine cycle (ORC) Rankine/Rankine Rankine/Kalina Rankine/Stirling Kalina/Kalina Kalina/Stirling

• • • •

Brayton/Rankine Brayton/ORC Brayton/Kalina Brayton/Stirling

• •

Brayton/Brayton Brayton/solid oxide fuel cell (SOFC)

Turbine-I

7

Power 8

1 Fuel Burner Air 2 Heating 6 3 application HEX 5

9 Turbine-II Power

4 10

13

Stack 11

12

Condenser-I

Pump-I

Turbine-III Power

14

16

18 15

Heating application

Condenser-II

Pump-II 17 Fig. 21 Rankine and organic Rankine cycle (ORC)-based combined cycle. HEX, heat exchanger.

be classified based on the working temperature range of topping and bottoming processes, as shown in Table 8. These integrated cycles are detailed investigated in the subtitles.

4.10.5.1

Rankine/Organic Rankine Cycle Combined Cycle

In the Rankine–ORC combined cycle, illustrated in Fig. 21, fuel and air enter the burner to produce useful heat energy at points 1 and 2, respectively. The rejected heat energy by stack gaseous from burner is transferred to a heat exchanger (HEX) unit for heating applications. To transfer heat energy at the maximum allowable temperature, the working fluid of Rankine cycle should be selected with a high critical temperature. Also, to recover heat energy more efficiently, turbine-I and II are integrated in the Rankine cycle, where temperature range of these turbines is suitable for extra electricity generation. The ORC process is similar to the Rankine process, but utilizes the organic working fluids instead of steam. The rejected heat energy from Rankine cycle is transferred to the ORC process, which has one turbine to produce power, a condenser to cool and condense the steam coming out of turbine-III, a pump to increase pressure level of working fluid and a condenser-I to transfer waste heat. The mass, energy, entropy, and exergy balance equations are defined for Rankine and ORC-based combined cycle, which is shown in Fig. 21. Burner subcomponent: the mass, energy, entropy, and exergy balance equations of burner subcomponent under steady state and steady flow conditions are written as

392

Integrated Gasification Combined Cycles _2þm _8þm _ 12 ¼ m _3þm _7þm _9 Mass: m _1þm

ð86Þ

_ 2 h2 þ m _ 8 h8 þ m _ 12 h12 ¼ m _ 3 h3 þ m _ 7 h7 þ m _ 9 h9 Energy: m _ 1 h1 þ m

ð87Þ

_ 2 s2 þ m _ 8 s8 þ m _ 12 s12 þ S_ gen;br ¼ m _ 3 s3 þ m _ 7 s7 þ m _ 9 s9 Entropy: m _ 1 s1 þ m

ð88Þ

_ D;br _ 2 ex2 þ m _ 8 ex8 ¼ m _ 3 ex3 þ m _ 7 ex7 þ m _ 9 ex 9 þ Ex Exergy: m _ 1 ex 1 þ m

ð89Þ

The combustion reaction occurs in burner subcomponent. The exit properties of burner are the function of air mass flow rate, heating value of fuel and also burner performance. _ 2 h2 ¼ m _ 3 h3 þ ð1 _ 1 LHV fuel þ m m

_ 1 LHV fuel Zbr Þm

ð90Þ

where LHVfuel is the lower heating value of fuel and Zbr is the combustion efficiency. The species coefficients of combustion reaction are given as lCx1 Hy1 þ ðxO2 O2 þ xN2 N2 þ xH2 O H2 O þ xCO2 CO2 þ xAr Ar Þ -yCO2 CO2 þ yN2 N2 þ yO2 O2 þ yH2 O H2 O þ yNO NO þ yCO CO þ yAr Ar yCO2 ¼ ðlx1 þ xCO2 yN2 ¼ ðxN2

yCO Þ

ð92Þ

yNO Þ

ð93Þ

  ly1 yH 2 O ¼ x H 2 O þ 2  yO2 ¼ xO2

lx1

ly1 4

ð91Þ

ð94Þ

lCO 2

lNO 2



ð95Þ

yAr ¼ xAr

ð96Þ

nfuel nair

ð97Þ

l¼

HEX subcomponent: the mass, energy, entropy, and exergy balance equations for HEX subcomponent can be defined under the steady state and steady flow conditions as _ 4; m _6 _5¼m mass: m _3 ¼m

ð98Þ

_ 5 h5 ¼ m _ 4 h4 þ m _ 6 h6 Energy: m _ 3 h3 þ m

ð99Þ

_ 5 s5 þ S_ gen; HEX ¼ m _ 4 s4 þ m _ 6 s6 Entropy: m _ 3 s3 þ m

ð100Þ

_ D;HEX _ 5 ex 5 ¼ m _ 4 ex4 þ m _ 6 ex 6 þ Ex Exergy: m _ 3 ex3 þ m

ð101Þ

Turbine-I subcomponent: under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for turbine-I subcomponent are _8 Mass: m _7 ¼m

ð102Þ

_ tur _ 8 h8 þ W Energy: m _ 7 h7 ¼ m Entropy: m _ 7 s7 þ S_ gen;tur

I

_ tur _ 8 ex8 þ W Exergy: m _ 7 ex 7 ¼ m

ð103Þ

I

_ 8 s8 ¼m I

_ D;tur þ Ex

ð104Þ I

ð105Þ

Turbine-II subcomponent: the balance equations of turbine-II under steady state and steady flow conditions are written as follows: _ 10 Mass: m _9¼m _ tur _ 10 h10 þ W Energy: m _ 9 h9 ¼ m

ð106Þ II

ð107Þ

Integrated Gasification Combined Cycles Entropy: m _ 9 s9 þ S_ gen; tur

_ 10 s10 ¼m

II

_ tur _ 10 ex10 þ W Exergy: m _ 9 ex 9 ¼ m

II

_ D;tur þ Ex

393

ð108Þ ð109Þ

II

Condenser-I subcomponent: the mass, energy, entropy, and exergy balance equations are given for condenser-I under the steady state and steady flow conditions. _ 11 ; m _ 16 Mass: m _ 10 ¼ m _ 13 ¼ m

ð110Þ

_ 16 h16 ¼ m _ 11 h11 þ m _ 13 h13 Energy: m _ 10 h10 þ m

ð111Þ

_ 16 s16 þ S_ gen;con Entropy: m _ 10 s10 þ m

I

_ 11 s11 þ m _ 13 s13 ¼m

_ D; con _ 16 ex16 ¼ m _ 11 ex11 þ m _ 13 ex13 þ Ex Exergy: m _ 10 ex10 þ m

ð112Þ

I

ð113Þ

Pump-I subcomponent: for the pump-I subcomponent of Rankine and ORC-based integrated cycle, the balance equations are provided under the steady state and steady flow conditions. _ 12 Mass: m _ 11 ¼ m

ð114Þ

_ p_I ¼ m _ 12 h12 Energy: m _ 11 h11 þ W

ð115Þ

_ 12 s12 Entropy: m _ 11 s11 þ S_ gen; p_I ¼ m

ð116Þ

_ p_I ¼ m _ D; p_I _ 12 ex12 þ Ex Exergy: m _ 11 ex11 þ W

ð117Þ

Turbine-III subcomponent: the balance equations of turbine-III subcomponent under steady state and steady flow conditions are defined as follows: _ 14 Mass: m _ 13 ¼ m

ð118Þ

_ tur _ 14 h14 þ W Energy: m _ 13 h13 ¼ m Entropy: m _ 13 s13 þ S_ gen; tur

III

_ tur _ 14 ex 14 þ W Exergy: m _ 13 ex 13 ¼ m

ð119Þ

III

_ 14 s14 ¼m

III

_ D; tur þ Ex

ð120Þ ð121Þ

III

Condenser-II subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for condenser-II subcomponent can be written as _ 15 ; m _ 18 Mass: m _ 14 ¼ m _ 17 ¼ m

ð122Þ

_ 17 h17 ¼ m _ 15 h15 þ m _ 18 h18 Energy: m _ 14 h14 þ m

ð123Þ

_ 17 s17 þ S_ gen; con Entropy: m _ 14 s14 þ m

II

_ 15 s15 þ m _ 18 s18 ¼m

_ D; con _ 17 ex17 ¼ m _ 15 ex 15 þ m _ 18 ex18 þ Ex Exergy: m _ 14 ex 14 þ m

ð124Þ

II

ð125Þ

Pump-II subcomponent: for the pump-II subcomponent of Rankine and ORC-based integrated cycle, the balance equations are provided under the steady state and steady flow conditions. _ 16 Mass: m _ 15 ¼ m

ð126Þ

_ p_II ¼ m _ 16 h16 Energy: m _ 15 h15 þ W

ð127Þ

_ 16 s16 Entropy: m _ 15 s15 þ S_ gen;p_II ¼ m

ð128Þ

394

Integrated Gasification Combined Cycles

_ p_II ¼ m _ D;p_II _ 16 ex 16 þ Ex Exergy: m _ 15 ex15 þ W

4.10.5.2

ð129Þ

Rankine/Kalina Combined Cycle

The Kalina cycle could be supply a better connection between the temperature difference of waste heat and that of working fluid of cycle, because of the variable composition of working fluid. The Kalina cycle components do not operate at vacuum conditions, therefore, this cycle has low operation and maintenance cost as compared to other traditional technologies. The schematic diagram of Rankine cycle and Kalina cycle based combined cycle is given in Fig. 22. As seen from this Fig. 22, fuel and air enter the burner at point 1 and 2, respectively. Stack gaseous exiting from burner are entered through the HEX to produce residential heating at point 3 and exiting from HEX at point 4. The steam from point 7 runs through turbine-II to produce electricity. The working fluid exiting from turbine-I enters the burner to recover heat energy efficiency. The steam enters the turbine-II at point 9, and expands point 10 to produce power. The water in the condenser-I can be utilized to provide heating to the working fluid in point 18 as illustrated in Fig. 22. After leaving from condenser-I at point 11, the working fluid of Rankine cycle enters the pump-I to increase pressure level. Kalina cycle is similar to the ORC process, but utilize chemical composition of H2O and NH3 with suitable mixture as working fluid in cycle. The superheated vapor enters the Kalina turbine to generate power. Expanded fluid exiting turbine-II goes through internal-HEX to increase temperature. After going through condenser and pump, pressured fluid enters internal-HEX. The working fluid exiting from internal-HEX enters the condenser-I and the cycle completes. The balance equations are defined for Rankine and Kalina cycles based combined cycle, which is illustrated in Fig. 22. Burner subcomponent: the mass, energy, entropy, and exergy balance equations can be defined for burner subcomponent under the steady state and steady flow conditions. _2þm _8þm _ 12 ¼ m _3þm _7þm _9 Mass: m _1þm

ð130Þ

_ 2 h2 þ m _ 8 h8 þ m _ 12 h12 ¼ m _ 3 h3 þ m _ 7 h7 þ m _ 9 h9 Energy: m _ 1 h1 þ m

ð131Þ

_ 2 s2 þ m _ 8 s8 þ m _ 12 s12 þ S_ gen;br ¼ m _ 3 s3 þ m _ 7 s7 þ m _ 9 s9 Entropy: m _ 1 s1 þ m

ð132Þ

Turbine-I

7

Power 8

1 Fuel Air 2

Burner

Heating 6 application

9

3 HEX

Turbine-II Power

5

4 10 Stack 11

12

13 Turbine-III

Condenser-I

Power

Pump-I 14 18 Internal-HEX

15

17 16 Pump-II Fig. 22 Rankine and Kalina cycle based combined cycle. HEX, heat exchanger.

Heating application 20

Condenser-II 19

Integrated Gasification Combined Cycles _ D; br _ 2 ex 2 þ m _ 8 ex 8 ¼ m _ 3 ex3 þ m _ 7 ex 7 þ m _ 9 ex9 þ Ex Exergy: m _ 1 ex1 þ m

395 ð133Þ

HEX subcomponent: the mass, energy, entropy, and exergy balance equations for HEX subcomponent can be defined under the steady state and steady flow conditions as _ 4; m _6 Mass: m _3 ¼m _5¼m

ð134Þ

_ 5 h5 ¼ m _ 4 h4 þ m _ 6 h6 Energy: m _ 3 h3 þ m

ð135Þ

_ 5 s5 þ S_ gen; HEX ¼ m _ 4 s4 þ m _ 6 s6 Entropy: m _ 3 s3 þ m

ð136Þ

_ D;HEX _ 5 ex 5 ¼ m _ 4 ex4 þ m _ 6 ex 6 þ Ex Exergy: m _ 3 ex3 þ m

ð137Þ

Turbine-I subcomponent: the mass, energy, entropy, and exergy balance equations are defined for turbine-I subcomponent under steady state and steady flow conditions. _8 Mass: m _7 ¼m

ð138Þ

_ tur _ 8 h8 þ W Energy: m _ 7 h7 ¼ m Entropy:m _ 7 s7 þ S_ gen;tur

I

ð139Þ

I

_ 8 s8 ¼m

_ tur _ 8 ex 8 þ W Exergy: m _ 7 ex 7 ¼ m

I

_ D; tur þ Ex

ð140Þ ð141Þ

I

Turbine-II subcomponent: the balance equations of turbine-II under steady state and steady flow conditions are written as follows: _ 10 Mass: m _9¼m

ð142Þ

_ tur _ 10 h10 þ W Energy: m _ 9 h9 ¼ m Entropy: m _ 9 s9 þ S_ gen; tur

II

ð143Þ

II

_ 10 s10 ¼m

_ tur _ 10 ex 10 þ W Exergy: m _ 9 ex9 ¼ m

II

_ D; tur þ Ex

ð144Þ ð145Þ

II

Condenser-I subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for condenser-I component are defined as follows: _ 11 ; m _ 16 Mass: m _ 10 ¼ m _ 13 ¼ m

ð146Þ

_ 16 h16 ¼ m _ 11 h11 þ m _ 13 h13 Energy: m _ 10 h10 þ m

ð147Þ

_ 16 s16 þ S_ gen; con Entropy: m _ 10 s10 þ m

I

_ 11 s11 þ m _ 13 s13 ¼m

_ D;con _ 16 ex 16 ¼ m _ 11 ex11 þ m _ 13 ex 13 þ Ex Exergy: m _ 10 ex10 þ m

ð148Þ

I

ð149Þ

Pump-I subcomponent: for the pump-I subcomponent of Rankine- and ORC-based integrated cycle, the balance equations are provided under the steady state and steady flow conditions. _ 12 Mass: m _ 11 ¼ m

ð150Þ

_ p_I ¼ m _ 12 h12 Energy: m _ 11 h11 þ W

ð151Þ

_ 12 s12 Entropy: m _ 11 s11 þ S_ gen; p_I ¼ m

ð152Þ

_ p_I ¼ m _ D; p_I _ 12 ex12 þ Ex Exergy: m _ 11 ex11 þ W

ð153Þ

396

Integrated Gasification Combined Cycles

Turbine-III subcomponent: the mass, energy, entropy, and exergy balance equations are written for turbine-III under the steady state and steady flow conditions. _ 14 Mass: m _ 13 ¼ m

ð154Þ

_ tur _ 14 h14 þ W Energy: m _ 13 h13 ¼ m Entropy: m _ 13 s13 þ S_ gen;tur

III

ð155Þ

III

_ 14 s14 ¼m

_ tur _ 14 ex14 þ W Exergy: m _ 13 ex13 ¼ m

III

_ D;tur þ Ex

ð156Þ ð157Þ

III

Internal-HEX subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for internal-HEX subcomponent can be written as _ 15 ; m _ 18 _ 17 ¼ m Mass: m _ 14 ¼ m

ð158Þ

_ 17 h17 ¼ m _ 15 h15 þ m _ 18 h18 Energy: m _ 14 h14 þ m

ð159Þ

_ 17 s17 þ S_ gen;int Entropy: m _ 14 s14 þ m

HEX

_ 15 s15 þ m _ 18 s18 ¼m

_ D;int _ 17 ex17 ¼ m _ 15 ex 15 þ m _ 18 ex18 þ Ex Exergy: m _ 14 ex 14 þ m

ð160Þ

HEX

ð161Þ

Condenser-II subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for condenser-II component are _ 16 ; m _ 20 _ 19 ¼ m Mass: m _ 15 ¼ m

ð162Þ

_ 19 h19 ¼ m _ 16 h16 þ m _ 20 h20 Energy: m _ 15 h15 þ m

ð163Þ

_ 19 s19 þ S_ gen; con Entropy: m _ 15 s15 þ m

II

_ 16 s16 þ m _ 20 s20 ¼m

_ D;con _ 19 ex19 ¼ m _ 16 ex16 þ m _ 20 ex 20 þ Ex Exergy: m _ 15 ex15 þ m

ð164Þ

II

ð165Þ

Pump-II subcomponent: for the pump-II subcomponent, the balance equations are provided under the steady state and steady flow conditions.

4.10.5.3

_ 17 Mass: m _ 16 ¼ m

ð166Þ

_ p_II ¼ m _ 17 h17 Energy: m _ 16 h16 þ W

ð167Þ

_ 17 s17 Entropy: m _ 16 s16 þ S_ gen;p_II ¼ m

ð168Þ

_ p_II ¼ m _ D;p_II _ 17 ex 17 þ Ex Exergy: m _ 16 ex16 þ W

ð169Þ

Rankine/Stirling Combined Cycle

The utilization of Rankine cycle and Stirling engine combined cycle is theoretically interesting to produce power, but in practice this combined cycle is limited by the small power size of the Stirling engine. Stirling engine is the heat engine with external supply of heat energy that work under the closed regenerative process. The heater of Stirling engine is integrated in condenser-I subcomponent as illustrated in Fig. 23. In the Stirling engine cycle, the regenerator subcomponent is the internal-HEX and temporary heat store integrated between the hotter and colder fields such that the working fluid goes through it first in one way then the other, taking heat energy from the working fluid in one direction, and returning it in the other field. The basic effect of regeneration utilization in the Stirling engine cycle is to rise the thermal performance. The balance equations of Rankine and Stirling engine-based combined cycle components are written in subsections. Burner subcomponent: the balance equations are written for burner subcomponent under the steady state and steady flow conditions.

Integrated Gasification Combined Cycles

397

Turbine-I

7

Power 8

1 Fuel Air 2

Burner

Heating 6 application

9

3 HEX

Turbine-II Power

5

4

13

10 Stack 11

12

Regenerater

Condenser-I

Pump-I 16

14

Heating application 18

Condenser-II 17

Power

15

Stirling engine Fig. 23 Rankine and Stirling engine-based combined cycle. HEX, heat exchanger.

_2þm _8þm _ 12 ¼ m _3þm _7þm _9 Mass: m _1þm

ð170Þ

_ 2 h2 þ m _ 8 h8 þ m _ 12 h12 ¼ m _ 3 h3 þ m _ 7 h7 þ m _ 9 h9 Energy: m _ 1 h1 þ m

ð171Þ

_ 2 s2 þ m _ 8 s8 þ m _ 12 s12 þ S_ gen; br ¼ m _ 3 s3 þ m _ 7 s7 þ m _ 9 s9 Entropy: m _ 1 s1 þ m

ð172Þ

_ D; br _ 2 ex 2 þ m _ 8 ex 8 ¼ m _ 3 ex3 þ m _ 7 ex 7 þ m _ 9 ex9 þ Ex Exergy: m _ 1 ex1 þ m

ð173Þ

HEX subcomponent: the balance equations for HEX subcomponent can be defined under the steady state and steady flow conditions as _ 4; m _6 Mass: m _3 ¼m _5¼m

ð174Þ

_ 5 h5 ¼ m _ 4 h4 þ m _ 6 h6 Energy: m _ 3 h3 þ m

ð175Þ

_ 5 s5 þ S_ gen;HEX ¼ m _ 4 s4 þ m _ 6 s6 Entropy: m _ 3 s3 þ m

ð176Þ

_ D;HEX _ 5 ex 5 ¼ m _ 4 ex4 þ m _ 6 ex 6 þ Ex Exergy: m _ 3 ex3 þ m

ð177Þ

Turbine-I subcomponent: the balance equations are defined for turbine-I subcomponent under steady state and steady flow conditions. _8 Mass: m _7 ¼m

ð178Þ

_ tur _ 8 h8 þ W Energy: m _ 7 h7 ¼ m Entropy: m _ 7 s7 þ S_ gen;tur

I

_ tur _ 8 ex8 þ W Exergy: m _ 7 ex 7 ¼ m

ð179Þ

I

_ 8 s8 ¼m I

_ D;tur þ Ex

ð180Þ I

ð181Þ

398

Integrated Gasification Combined Cycles

Turbine-II subcomponent: the balance equations of turbine-II under steady state and steady flow conditions are _ 10 Mass: m _9¼m

ð182Þ

_ tur _ 10 h10 þ W Energy: m _ 9 h9 ¼ m Entropy: m _ 9 s9 þ S_ gen;tur

ð183Þ

II

_ 10 s10 ¼m

II

_ tur _ 10 ex 10 þ W Exergy: m _ 9 ex9 ¼ m

_ II þ Ex D;tur

ð184Þ ð185Þ

II

Condenser-I subcomponent: under the steady state and steady flow conditions, the balance equations for condenser-I subcomponent are defined as follows: _ 11 ; m _ 16 _ 13 ¼ m Mass: m _ 10 ¼ m

ð186Þ

_ 16 h16 ¼ m _ 11 h11 þ m _ 13 h13 Energy: m _ 10 h10 þ m

ð187Þ

_ 16 s16 þ S_ gen;con Entropy: m _ 10 s10 þ m

I

_ 11 s11 þ m _ 13 s13 ¼m

_ D;con _ 16 ex 16 ¼ m _ 11 ex11 þ m _ 13 ex 13 þ Ex Exergy: m _ 10 ex10 þ m

ð188Þ

I

ð189Þ

Pump-I subcomponent: under the steady state and steady flow conditions, the balance equations for the pump-I subcomponent are _ 12 Mass: m _ 11 ¼ m

ð190Þ

_ p_I ¼ m _ 12 h12 Energy: m _ 11 h11 þ W

ð191Þ

_ 12 s12 Entropy: m _ 11 s11 þ S_ gen;p_I ¼ m

ð192Þ

_ p_I ¼ m _ D;p_I _ 12 ex 12 þ Ex Exergy: m _ 11 ex 11 þ W

ð193Þ

Regenerator subcomponent: the mass, energy, entropy, and exergy balance equations for regenerator subcomponent can be defined under the steady state and steady flow conditions as follows: _ 14 Mass: m _ 13 ¼ m

ð194Þ

_ L;reg _ 14 h14 þ Q Energy: m _ 13 h13 ¼ m

ð195Þ

_ L;reg =Treg _ 14 s14 þ Q Entropy: m _ 13 s13 þ S_ gen; reg ¼ m

ð196Þ

_ Q þ Ex _ D;reg _ 14 ex14 þ Ex Exergy: m _ 13 ex13 ¼ m L;reg

ð197Þ

Condenser-II subcomponent: under the steady state and steady flow conditions, the balance equations for condenser-II subcomponent are _ 15 m _ 18 _ 17 ¼ m Mass: m _ 14 ¼ m

ð198Þ

_ 15 h15 ¼ m _ 17 h17 þ m _ 18 h18 Energy: m _ 14 h14 þ m

ð199Þ

_ 15 s15 þ S_ gen;con Entropy: m _ 14 s14 þ m

II

_ 17 s17 þ m _ 18 s18 ¼m

_ D;con _ 15 ex15 ¼ m _ 17 ex17 þ m _ 18 ex 18 þ Ex Exergy: m _ 14 ex14 þ m

ð200Þ

II

ð201Þ

Stirling engine subcomponent: the mass, energy, entropy, and exergy balance equations can be written for Stirling engine subcomponent under steady state and steady flow conditions. _ 16 ð202Þ Mass: m _ 15 ¼ m

Integrated Gasification Combined Cycles

399

_ Se _ 16 h16 þ W Energy: m _ 15 h15 ¼ m

ð203Þ

_ 16 s16 Entropy: m _ 15 s15 þ S_ gen;Se ¼ m

ð204Þ

_ Se þ Ex _ D;Se _ 16 ex 16 þ W Exergy: m _ 15 ex 15 ¼ m

ð205Þ

The power output rate of Stirling engine can be defined as follows:   _ in; Se Q _ Loss; Se _ Se ¼ Zpe Q W

ð206Þ _ _ where Zpe is the polytrophic efficiency, Qin;Se is the heat flow rate from hot source to Stirling engine and QLoss; Se is the heat loss rate from Stirling engine to the environment, and also it can be defined as   _ Loss; Se ¼ Q _ in; Se 1 Zmec; Se Q ð207Þ The polytrophic efficiency can be calculated as follows:  ð1 RF1 g Þ ζ ðRFg 1 Zpe ¼ ð1 RF1 g Þ þ ð1 ζÞð1

1Þ eSe Þ



ð208Þ

here, RF is the reversible factor of Stirling engine subcomponent, eSe is the efficiency of internal HEXs in Stirling engine, g is constant and equal to 1.667, ζ is the temperature of cooler gas over the heater gas, and can be calculated as follows:  Tc;g ζ¼ ð209Þ Th;g where Tc,g and Th,g are the temperature of cooler and heater gas in K. Th;g ¼ Th;w Tc;w ¼ Tw;i

DThigh

ð210Þ

0:66667ðDTw Þ

ð211Þ

Tc;g ¼ Tc;w þ DTlow

ð212Þ

where Th,w and Tc,w are the temperature of heater wall and cooler wall, respectively, and Tw,i is the inlet water temperature.

4.10.5.4

Brayton/Rankine Combined Cycle

For the different combined cycles, the Brayton and Rankine combined cycle is the most developed and widespread cycle nowadays. The principle of Brayton and Rankine combined cycle is to operate in cascade one or more Brayton turbines, followed by the steam power process whose heat resource is the waste heat of Brayton cycle. Fig. 24 shows the schematic diagram of Brayton and Rankine combine system to produce power and residential heating. As seen in this Fig. 24, Brayton subsystem consists of a compressor to provide the required pressure level for air, a combustion chamber and one gas turbine to generate electricity. The hot stack gaseous Fuel

2

4

Combustion chamber

3

Compressor

Gas turbine Power 5

1

10

Air HEX-I Heating application 9

HEX-II 8

Turbine Power

6 11

13

15

7 12 Stack

Condenser

Pump 14 Fig. 24 Brayton and Rankine cycle based combined cycle. HEX, heat exchanger.

Heating application

400

Integrated Gasification Combined Cycles

exiting from the gas turbine are transferred to the Rankine cycle to make utilize of this heat to generate electricity and hot water for residential heating applications. The balance equations of Brayton and Rankine cycle based combined cycle components are written in subsections. Compressor subcomponent: under steady state and steady flow conditions, the balance equations for compressor are defined as follows: _2 Mass: m _1 ¼m

ð213Þ

_ comp ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð214Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;comp ¼ m

ð215Þ

_ comp ¼ m _ D;comp _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð216Þ

Air at surroundings temperature and pressure enters the compressor subcomponent. The compressor exit temperature can be calculated as follows:   ga 1 1 T2 ¼ T1 1 þ r ga ZAC AC

1



ð217Þ

here ZAC is the compressor isentropic efficiency, rAC is the specific heat ratio, and ga is the air specific ratio. The compressor work rate can be calculated as follows: _ comp ¼ m _ 1 Cpa ðT2 W here Cpa is treated as a function of temperature as given below:     3:83T 9:45T 2 þ Cpa ¼ 1:048 107 104

T1 Þ 

5:49T 3 1010

ð218Þ 

þ



7:92T 4 1014



ð219Þ

Combustion chamber subcomponent: under the steady state and steady flow conditions, the balance equations for combustion chamber are defined as follows: _4¼m _3 Mass: m _2þm

ð220Þ

_ 4 h4 ¼ m _ 3 h3 Energy: m _ 2 h2 þ m

ð221Þ

_ 4 s4 þ S_ gen;cc ¼ m _ 3 s3 Entropy: m _ 2 s2 þ m

ð222Þ

_ D;cc _ 4 ex4 ¼ m _ 3 ex 3 þ Ex Exergy:m _ 2 ex2 þ m

ð223Þ

Gas turbine subcomponent: under steady state and steady flow conditions, the balance equations for gas turbine are defined as follows: _5 Mass: m _3 ¼m

ð224Þ

_ gt _ 5 h5 þ W Energy: m _ 3 h3 ¼ m

ð225Þ

_ 5 s5 Entropy: m _ 3 s3 þ S_ gen;gt ¼ m

ð226Þ

_ gt þ Ex _ D;gt _ 5 ex5 þ W Exergy: m _ 3 ex 3 ¼ m

ð227Þ

The gas turbine inlet pressure can be given by considering the pressure decrease across the burner as given below P3 ¼1 P2

DPcc

ð228Þ

here DPcc is the pressure loss across the combustion chamber. The gas turbine exit temperature can be calculated as follows:

Integrated Gasification Combined Cycles 2

T4 ¼ T3 41

0

ZGT @1

13  1 g gg P3 g A5 P4

401

ð229Þ

where ZGT is the isentropic efficiency of gas turbine. The power output rate from gas turbine can be calculated as _ gt ¼ m _ 3 Cpg ðT3 W

T4 Þ

here Cpg can be given as a function of temperature as     6:99T 2:712T 2 þ Cpg ¼ 0:991 7 10 105

ð230Þ 

1:2244T 3 1010



ð231Þ

HEX-I subcomponent: the balance equations for HEX-I subcomponent can be defined under the steady state and steady flow conditions as _ 6; m _ 13 _ 10 ¼ m Mass: m _5¼m

ð232Þ

_ 13 h13 ¼ m _ 6 h6 þ m _ 10 h10 Energy: m _ 5 h5 þ m

ð233Þ

_ 13 s13 þ S_ gen; HEX Entropy: m _ 5 s5 þ m

I

_ 6 s6 þ m _ 10 s10 ¼m

_ D; HEX _ 13 ex13 ¼ m _ 6 ex6 þ m _ 10 ex10 þ Ex Exergy: m _ 5 ex 5 þ m

ð234Þ

I

ð235Þ

HEX-II subcomponent: under the steady state and steady flow conditions, the balance equations for HEX-II can written as follows: _ 7; m _9 Mass: m _6 ¼m _8¼m

ð236Þ

_ 8 h8 ¼ m _ 7 h7 þ m _ 9 h9 Energy: m _ 6 h6 þ m

ð237Þ

_ 8 s8 þ S_ gen;HEX Entropy: m _ 6 s6 þ m

II

_ 7 s7 þ m _ 9 s9 ¼m

_ D;HEX _ 8 ex 8 ¼ m _ 7 ex7 þ m _ 9 ex 9 þ Ex Exergy: m _ 6 ex6 þ m

ð238Þ

II

ð239Þ

Turbine subcomponent: under steady state and steady flow conditions, the balance equations for Rankine turbine are defined as follows: _ 11 Mass: m _ 10 ¼ m

ð240Þ

_ st _ 11 h11 þ W Energy: m _ 10 h10 ¼ m

ð241Þ

_ 11 s11 Entropy: m _ 10 s10 þ S_ gen;st ¼ m

ð242Þ

_ st þ Ex _ D;st _ 11 ex 11 þ W Exergy: m _ 10 ex 10 ¼ m

ð243Þ

Condenser subcomponent: under the steady state and steady flow conditions, the balance equations for condenser-II subcomponent can be defined as follows: _ 12 ; m _ 15 _ 14 ¼ m Mass: m _ 11 ¼ m

ð244Þ

_ 14 h14 ¼ m _ 12 h12 þ m _ 15 h15 Energy: m _ 11 h11 þ m

ð245Þ

_ 14 s14 þ S_ gen;con ¼ m _ 12 s12 þ m _ 15 s15 Entropy: m _ 11 s11 þ m

ð246Þ

_ D;con _ 14 ex 14 ¼ m _ 12 ex12 þ m _ 15 ex 15 þ Ex Exergy: m _ 11 ex11 þ m

ð247Þ

Pump subcomponent: for the pump subcomponent, the balance equations are provided under the steady state and steady flow conditions.

402

4.10.5.5

Integrated Gasification Combined Cycles _ 13 Mass: m _ 12 ¼ m

ð248Þ

_ p¼m _ 13 h13 Energy: m _ 12 h12 þ W

ð249Þ

_ 13 s13 Entropy: m _ 12 s12 þ S_ gen;p ¼ m

ð250Þ

_ p ¼m _ D;p _ 13 ex 13 þ Ex Exergy: m _ 12 ex 12 þ W

ð251Þ

Brayton/Kalina Combined Cycle

The schematic diagram of Brayton and Kalina based combined cycle is illustrated in Fig. 25. No extra vacuum and pump are needed to be provided in the condenser during working or stand-by periods. Because the hot stack gaseous pressure level exiting from gas turbine in the Brayton cycle is above from surroundings pressure. Hence, the start-up operations could be applied in the much shorter time. The working fluid combination of Kalina cycle could easily be changed in order to get the ideal efficiency based on the changes in working or surroundings cases. Another important advantage of Brayton and Kalina process based combined system is that whole system has the smaller size. The thermodynamic balance equations of Brayton and Kalina cycle based combined cycle components are given in subsections. Compressor subcomponent: under steady state and steady flow conditions, the balance equations for compressor subsystem can be written as

Fuel

2

_2 Mass: m _1 ¼m

ð252Þ

_ comp ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð253Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen; comp ¼ m

ð254Þ

_ comp ¼ m _ D;comp _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð255Þ

4

Combustion chamber

Compressor

3 Power

Gas turbine

10 5

1 Air

Turbine

HEX-I Heating application

Power

6 9 HEX-II 8

15

11

7 Internal-HEX

Stack

12

14 13 Pump Fig. 25 Brayton and Kalina cycle based combined cycle. HEX, heat exchanger.

Heating application 17

Condenser 16

Integrated Gasification Combined Cycles

403

Combustion chamber subcomponent: under the steady state and steady flow conditions, the balance equations for combustion chamber subcomponent are _4¼m _3 Mass: m _2þm

ð256Þ

_ 4 h4 ¼ m _ 3 h3 Energy: m _ 2 h2 þ m

ð257Þ

_ 4 s4 þ S_ gen;cc ¼ m _ 3 s3 Entropy: m _ 2 s2 þ m

ð258Þ

_ D;cc _ 4 ex 4 ¼ m _ 3 ex 3 þ Ex Exergy: m _ 2 ex 2 þ m

ð259Þ

Gas turbine subcomponent: under steady state and steady flow conditions, the balance equations for gas turbine subsystem are written as _5 Mass: m _3 ¼m

ð260Þ

_ gt _ 5 h5 þ W Energy: m _ 3 h3 ¼ m

ð261Þ

_ 5 s5 Entropy: m _ 3 s3 þ S_ gen;gt ¼ m

ð262Þ

_ gt þ Ex _ D;gt _ 5 ex5 þ W Exergy: m _ 3 ex 3 ¼ m

ð263Þ

HEX-I subcomponent: the balance equations for HEX-I subcomponent can be defined under the steady state and steady flow conditions as follows: _ 6; m _ 15 _ 10 ¼ m Mass: m _5¼m

ð264Þ

_ 15 h15 ¼ m _ 6 h6 þ m _ 10 h10 Energy: m _ 5 h5 þ m

ð265Þ

_ 15 s15 þ S_ gen; HEX Entropy: m _ 5 s5 þ m

I

_ 6 s6 þ m _ 10 s10 ¼m

_ D; HEX _ 15 ex15 ¼ m _ 6 ex6 þ m _ 10 ex10 þ Ex Exergy: m _ 5 ex 5 þ m

ð266Þ

I

ð267Þ

HEX-II subcomponent: under the steady state and steady flow conditions, the balance equations for HEX-II subcomponent can written as _ 7; m _9 Mass: m _6 ¼m _8¼m

ð268Þ

_ 8 h8 ¼ m _ 7 h7 þ m _ 9 h9 Energy: m _ 6 h6 þ m

ð269Þ

_ 8 s8 þ S_ gen; HEX Entropy: m _ 6 s6 þ m

II

_ 7 s7 þ m _ 9 s9 ¼m

_ D;HEX _ 8 ex8 ¼ m _ 7 ex7 þ m _ 9 ex9 þ Ex Exergy: m _ 6 ex6 þ m

ð270Þ

II

ð271Þ

Turbine subcomponent: under steady state and steady flow conditions, the balance equations for steam turbine subcomponent can be defined as _ 11 Mass: m _ 10 ¼ m

ð272Þ

_ st _ 11 h11 þ W Energy: m _ 10 h10 ¼ m

ð273Þ

_ 11 s11 Entropy: m _ 10 s10 þ S_ gen;st ¼ m

ð274Þ

_ st þ Ex _ D;st _ 11 ex 11 þ W Exergy: m _ 10 ex 10 ¼ m

ð275Þ

Internal-HEX subcomponent: under the steady state and steady flow conditions, the balance equations for internal-HEX subcomponent can written as follows:

404

Integrated Gasification Combined Cycles _ 12 ; m _ 15 Mass: m _ 11 ¼ m _ 14 ¼ m

ð276Þ

_ 14 h14 ¼ m _ 12 h12 þ m _ 15 h15 Energy: m _ 11 h11 þ m

ð277Þ

_ 14 s14 þ S_ gen; int Entropy: m _ 11 s11 þ m

HEX

_ 12 s12 þ m _ 15 s15 ¼m

_ D;int _ 14 ex14 ¼ m _ 12 ex 12 þ m _ 15 ex15 þ Ex Exergy: m _ 11 ex 11 þ m

HEX

ð278Þ ð279Þ

Condenser subcomponent: under the steady state and steady flow conditions, the balance equations for condenser-II subcomponent can be defined as follows: _ 13 ; m _ 17 Mass: m _ 12 ¼ m _ 16 ¼ m

ð280Þ

_ 16 h16 ¼ m _ 13 h13 þ m _ 17 h17 Energy: m _ 12 h12 þ m

ð281Þ

_ 16 s16 þ S_ gen;con ¼ m _ 13 s13 þ m _ 17 s17 Entropy: m _ 12 s12 þ m

ð282Þ

_ D;con _ 16 ex 16 ¼ m _ 13 ex13 þ m _ 17 ex 17 þ Ex Exergy: m _ 12 ex12 þ m

ð283Þ

Pump subcomponent: for pump subcomponent, the balance equations are provided under the steady state and steady flow conditions. _ 14 Mass: m _ 13 ¼ m

ð284Þ

_ p¼m _ 14 h14 Energy: m _ 13 h13 þ W

ð285Þ

_ 14 s14 Entropy: m _ 13 s13 þ S_ gen;p ¼ m

ð286Þ

Fuel

4

Combustion chamber

2 Compressor-I

3 Gas turbine-I

Power

1 Air 6 Stack 10

Compressor-II

7 Air

5

HEX-I 11

Gas turbine-II

8

Power

15

12 9 HEX-II

Intercooler

14 13 Air-out

Fig. 26 Brayton/Brayton cycle based combined cycle. HEX, heat exchanger.

Heating application

Integrated Gasification Combined Cycles _ p ¼m _ D;p _ 14 ex 14 þ Ex Exergy: m _ 13 ex 13 þ W

4.10.5.6

405

ð287Þ

Brayton/Brayton Combined Cycle

As seen in Fig. 26, two Brayton cycles could be integrated by using the HEX subcomponent. The stack gases of the gas turbine-I enter the HEX-I for heating of secondary Brayton cycle working fluid. The air exiting from HEX-I is expended in the gas turbine-II to produce additional electricity. The utilization of intercooling unit in the air compressor of secondary Brayton cycle decreases the required expander work. In comparison to the Brayton and Rankine integrated process, this combined cycle does not require the bulky steam components, such as boiler, steam turbine, and condenser, or the water processing equipment. The balance equations of Brayton/Brayton cycle based combined cycle components are written in subsections. Compressor-I subcomponent: under steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for compressor-I can be written as follows: _2 Mass: m _1 ¼m _ comp Energy: m _ 1 h1 þ W

ð288Þ

I

Entropy: m _ 1 s1 þ S_ gen;comp _ comp Exergy: m _ 1 ex 1 þ W

I

_ 2 h2 ¼m

ð289Þ

_ 2 s2 ¼m

I

ð290Þ

_ D; comp _ 2 ex2 þ Ex ¼m

ð291Þ

I

Combustion chamber subcomponent: under the steady state and steady flow conditions, the balance equations for combustion chamber subcomponent are written as: _4¼m _3 Mass: m _2þm

ð292Þ

_ 4 h4 ¼ m _ 3 h3 Energy: m _ 2 h2 þ m

ð293Þ

_ 4 s4 þ S_ gen;cc ¼ m _ 3 s3 Entropy: m _ 2 s2 þ m

ð294Þ

_ D;cc _ 4 ex 4 ¼ m _ 3 ex 3 þ Ex Exergy: m _ 2 ex 2 þ m

ð295Þ

Gas turbine-I subcomponent: under steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for gas turbine-I can be defined as follows: _5 Mass: m _3 ¼m

ð296Þ

_ gt _ 5 h5 þ W Energy: m _ 3 h3 ¼ m Entropy: m _ 3 s3 þ S_ gen;gt

I

ð297Þ

I

_ 5 s5 ¼m

_ gt _ 5 ex5 þ W Exergy: m _ 3 ex3 ¼ m

I

_ D;gt þ Ex

ð298Þ ð299Þ

I

HEX-I subcomponent: under the steady state and steady flow conditions, the balance equations for HEX-I subcomponent can be defined as _ 6; m _ 11 Mass: m _5¼m _ 10 ¼ m

ð300Þ

_ 10 h10 ¼ m _ 6 h6 þ m _ 11 h11 Energy: m _ 5 h5 þ m

ð301Þ

_ 10 s10 þ S_ gen;HEX Entropy: m _ 5 s5 þ m

I

_ 6 s6 þ m _ 11 s11 ¼m

_ D;HEX _ 10 ex 10 ¼ m _ 6 ex 6 þ m _ 11 ex11 þ Ex Exergy: m _ 5 ex 5 þ m

ð302Þ

I

ð303Þ

406

Integrated Gasification Combined Cycles

Compressor-II subcomponent: under steady state and steady flow conditions, the balance equations for compressor-II can be defined as _8þm _ 10 Mass: m _7¼m _ comp Energy: m _ 7 h7 þ W

II

Entropy: m _ 7 s7 þ S_ gen;comp _ comp Exergy: m _ 7 ex7 þ W

II

ð304Þ

_ 8 h8 þ m _ 10 h10 ¼m

II

ð305Þ

_ 8 s8 þ m _ 10 s10 ¼m

_ D;comp _ 8 ex 8 þ m _ 10 ex 10 þ Ex ¼m

ð306Þ

II

ð307Þ

Intercooler subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for intercooler subcomponent are written as follows: _9 Mass: m _8 ¼m

ð308Þ

_ 9 h9 Energy: m _ 8 h8 ¼ m

ð309Þ

_ 9 s9 Entropy: m _ 8 s8 þ S_ gen;ic ¼ m

ð310Þ

_ D;ic _ 9 ex 9 þ Ex Exergy: m _ 8 ex 8 ¼ m

ð311Þ

Gas turbine-II subcomponent: under steady state and steady flow conditions, the balance equations for gas turbine-II can be defined as _ 12 Mass: m _ 11 ¼ m

ð312Þ

_ gt _ 12 h12 þ W Energy: m _ 11 h11 ¼ m Entropy: m _ 11 s11 þ S_ gen;gt

II

ð313Þ

II

_ 12 s12 ¼m

_ gt _ 12 ex 12 þ W Exergy: m _ 11 ex 11 ¼ m

II

_ D;gt þ Ex

ð314Þ ð315Þ

II

HEX-II subcomponent: under the steady state and steady flow conditions, the balance equations for HEX-II can be defined as _ 13 ; m _ 15 Mass: m _ 12 ¼ m _ 14 ¼ m

ð316Þ

_ 14 h14 ¼ m _ 13 h13 þ m _ 15 h15 Energy: m _ 12 h12 þ m

ð317Þ

_ 14 s14 þ S_ gen;HEX Entropy: m _ 12 s12 þ m

II

_ 13 s13 þ m _ 15 s15 ¼m

_ D;HEX _ 14 ex 14 ¼ m _ 13 ex 13 þ m _ 15 ex 15 þ Ex Exergy: m _ 12 ex 12 þ m

4.10.5.7

ð318Þ

II

ð319Þ

Brayton/Stirling Engine Combined Cycle

Fig. 27 shows the Brayton and Stirling engine combined cycle for power production. As seen from this Fig. 27, the heater component of Stirling engine process can be placed after the gas turbine in the stack gases flow. The integrated system composition could be determined by the optimal efficiency of Brayton and Stirling engine combined cycle and by the materials used in the head of Stirling heater. The balance equations of Brayton and Stirling engine based combined cycle components are written in subsections. Compressor subcomponent: under steady state and steady flow conditions, the balance equations for compressor subsystem can be defined as _2 Mass: m _1 ¼m

ð320Þ

_ comp ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð321Þ

Integrated Gasification Combined Cycles

Fuel

2

407

4

Combustion chamber

3

Compressor

Gas turbine

Power

1 Air

Heating application 9 7

5

10

6

HEX-II

Regenerater

HEX-I

Stack 8 13

11

Heating application 15

Condenser Power

12

14

Stirling engine Fig. 27 Brayton and Stirling engine based combined cycle. HEX, heat exchanger.

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;comp ¼ m

ð322Þ

_ comp ¼ m _ D;comp _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð323Þ

Combustion chamber subcomponent: under the steady state and steady flow conditions, the balance equations for combustion chamber are _4¼m _3 Mass: m _2þm

ð324Þ

_ 4 h4 ¼ m _ 3 h3 Energy: m _ 2 h2 þ m

ð325Þ

_ 4 s4 þ S_ gen;cc ¼ m _ 3 s3 Entropy: m _ 2 s2 þ m

ð326Þ

_ D;cc _ 4 ex 4 ¼ m _ 3 ex 3 þ Ex Exergy: m _ 2 ex 2 þ m

ð327Þ

Gas turbine subcomponent: under steady state and steady flow conditions, the balance equations for gas turbine subsystem can be written as follows: _5 Mass: m _3 ¼m

ð328Þ

_ gt _ 5 h5 þ W Energy: m _ 3 h3 ¼ m

ð329Þ

_ 5 s5 Entropy: m _ 3 s3 þ S_ gen;gt ¼ m

ð330Þ

_ gt þ Ex _ D;gt _ 5 ex5 þ W Exergy: m _ 3 ex 3 ¼ m

ð331Þ

HEX-I subcomponent: under the steady state and steady flow conditions, the balance equations for HEX-I subcomponent can be defined as _ 6; m _ 13 Mass: m _5¼m _ 10 ¼ m

ð332Þ

_ 13 h13 ¼ m _ 6 h6 þ m _ 10 h10 Energy: m _ 5 h5 þ m

ð333Þ

408

Integrated Gasification Combined Cycles

_ 13 s13 þ S_ gen; HEX Entropy: m _ 5 s5 þ m

I

_ 6 s6 þ m _ 10 s10 ¼m

_ D;HEX _ 13 ex 13 ¼ m _ 6 ex6 þ m _ 10 ex10 þ Ex Exergy: m _ 5 ex 5 þ m

ð334Þ

I

ð335Þ

HEX-II subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for HEX-II are written as follows: _ 7; m _9 _8¼m Mass: m _6 ¼m

ð336Þ

_ 8 h8 ¼ m _ 7 h7 þ m _ 9 h9 Energy: m _ 6 h6 þ m

ð337Þ

_ 8 s8 þ S_ gen; HEX Entropy: m _ 6 s6 þ m

II

_ 7 s7 þ m _ 9 s9 ¼m

_ D;HEX _ 8 ex8 ¼ m _ 7 ex7 þ m _ 9 ex9 þ Ex Exergy: m _ 6 ex6 þ m

ð338Þ

II

ð339Þ

Regenerator subcomponent: the balance equations for regenerator can be defined under the steady state and steady flow conditions as _ 11 Mass: m _ 10 ¼ m

ð340Þ

_ L;reg _ 11 h11 þ Q Energy: m _ 10 h10 ¼ m

ð341Þ

_ L; reg =Treg _ 11 s11 þ Q Entropy: m _ 10 s10 þ S_ gen; reg ¼ m

ð342Þ

_ Q _ _ 11 ex11 þ Ex Exergy: m _ 10 ex10 ¼ m L;reg þ Ex D;reg

ð343Þ

Condenser subcomponent: under the steady state and steady flow conditions, the balance equations for condenser subcomponent are defined as follows: _ 12 ; m _ 15 Mass: m _ 11 ¼ m _ 14 ¼ m

ð344Þ

_ 14 h14 ¼ m _ 12 h12 þ m _ 15 h15 Energy: m _ 11 h11 þ m

ð345Þ

_ 14 s14 þ S_ gen; con ¼ m _ 12 s12 þ m _ 15 s15 Entropy: m _ 11 s11 þ m

ð346Þ

_ D; con _ 14 ex14 ¼ m _ 12 ex 12 þ m _ 15 ex15 þ Ex Exergy: m _ 11 ex 11 þ m

ð347Þ

Stirling engine subcomponent: the balance equations can be written for Stirling engine under steady state and steady flow conditions.

4.10.5.8

_ 13 Mass: m _ 12 ¼ m

ð348Þ

_ Se _ 13 h13 þ W Energy: m _ 12 h12 ¼ m

ð349Þ

_ 13 s13 Entropy: m _ 12 s12 þ S_ gen; Se ¼ m

ð350Þ

_ Se þ Ex _ D; Se _ 13 ex13 þ W Exergy: m _ 12 ex 12 ¼ m

ð351Þ

Brayton/Solid Oxide Fuel Cell Combined Cycle

Fig. 28 illustrates the schematic diagram of a power production process with both the Brayton and solid oxide fuel cell (SOFC) to generate electricity. The working temperature range and desired output power of fuel cells are very important indicators for fuel cell-based integrated system. For instance, the SOFC operates at sufficiently high temperature ranges to allow direct internal reforming process. The anode stack gases carry enough HP steam to supply the water necessary for the reforming process. The heat energy for this endothermic reaction can be supplied by the surroundings by convection and radiation heat transfer mechanisms. One of the important benefits of SOFC is that either H2 and CO2 can be utilized as fuel. The balance equations are defined for Brayton and SOFC based combined cycle, which shown in Fig. 28.

Integrated Gasification Combined Cycles

409

Heating application 13 11

HEX-II

10 HEX-I

Stack

Fuel

Power 3

12 2

SOFC Pure 7 methane

5

4

9

Combustion chamber

8

6 Gas turbine

Compressor

Power 1 Air Fig. 28 Brayton and solid oxide fuel cell (SOFC) based combined cycle. HEX, heat exchanger.

Compressor subcomponent: under steady state and steady flow conditions, the balance equations for compressor are written as follows: _2 Mass: m _1¼m

ð352Þ

_ comp ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð353Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;comp ¼ m

ð354Þ

_ comp ¼ m _ D;comp _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð355Þ

HEX-I subcomponent: under the steady state and steady flow conditions, the balance equations for HEX-I can be defined as follows: _ 3; m _ 10 Mass: m _2 ¼m _9¼m

ð356Þ

_ 9 h9 ¼ m _ 3 h3 þ m _ 10 h10 Energy: m _ 2 h2 þ m

ð357Þ

_ 9 s9 þ S_ gen; HEX Entropy: m _ 2 s2 þ m

I

_ 3 s3 þ m _ 10 s10 ¼m

_ D;HEX _ 9 ex 9 ¼ m _ 3 ex3 þ m _ 10 ex10 þ Ex Exergy: m _ 2 ex2 þ m

ð358Þ ð359Þ

I

HEX-II subcomponent: under the steady state and steady flow conditions, the balance equations for HEX-II are written as _ 11 ; m _ 13 Mass: m _ 10 ¼ m _ 12 ¼ m

ð360Þ

_ 12 h12 ¼ m _ 11 h11 þ m _ 13 h13 Energy: m _ 10 h10 þ m

ð361Þ

_ 12 s12 þ S_ gen;HEX Entropy: m _ 10 s10 þ m

II

_ 11 s11 þ m _ 13 s13 ¼m

_ D; HEX _ 12 ex12 ¼ m _ 11 ex11 þ m _ 13 ex 13 þ Ex Exergy: m _ 10 ex10 þ m

ð362Þ

II

ð363Þ

SOFC subcomponent: the mass, energy, entropy, and exergy balance equations for SOFC can be written under steady state and steady flow conditions. _7 ¼m _4þm _8 Mass: m _3þm

ð364Þ

_ SOFC _ 7 h7 ¼ m _ 4 h4 þ m _ 8 h8 þ W Energy: m _ 3 h3 þ m

ð365Þ

410

Integrated Gasification Combined Cycles _ 7 s7 þ S_ gen; SOFC ¼ m _ 4 s4 þ m _ 8 s8 Entropy: m _ 3 s3 þ m

ð366Þ

_ SOFC þ Ex _ D; SOFC _ 7 ex 7 ¼ m _ 4 ex4 þ m _ 8 ex 8 þ W Exergy: m _ 3 ex3 þ m

ð367Þ

Combustion chamber subcomponent: under the steady state and steady flow conditions, the balance equations for combustion chamber can be written as follows: _5þm _8¼m _6 Mass: m _4þm

ð368Þ

_ 5 h5 þ m _ 8 h8 ¼ m _ 6 h6 Energy: m _ 4 h4 þ m

ð369Þ

_ 5 s5 þ m _ 8 s8 þ S_ gen;cc ¼ m _ 6 s6 Entropy: m _ 4 s4 þ m

ð370Þ

_ D;cc _ 5 ex5 þ m _ 8 ex 8 ¼ m _ 6 ex6 þ Ex Exergy: m _ 4 ex4 þ m

ð371Þ

Gas turbine subcomponent: under steady state and steady flow conditions, the balance equations for gas turbine subsystem are

4.10.6

_9 Mass: m _6 ¼m

ð372Þ

_ gt _ 9 h9 þ W Energy: m _ 6 h6 ¼ m

ð373Þ

_ 9 s9 Entropy: m _ 6 s6 þ S_ gen;gt ¼ m

ð374Þ

_ gt þ Ex _ D;gt _ 9 ex9 þ W Exergy: m _ 6 ex 6 ¼ m

ð375Þ

Integrated Gasification Combined Cycles

This chapter gives an overview of basic integrated cycles based on thermodynamic principles. The IGCC generates the power and steam from the solid or liquid sources. The commercial IGCC electricity generation processes have shown capable of exceeding the most compact emissions arrangements recently feasible to comparable combustion-based power facilities. Firstly, the fossil sources and waste materials are converted to synthesis gaseous which is a mixture of CO, H2 and often times CO2. Secondly, the synthesis gaseous are converted to power in the combined cycle consisting of the gas turbine subsystem and steam turbine subsystem which includes the HRSG. A number of IGCCs have been developed to different extents, and they may be classified based on source types, such as (1) coal, (2) biomass, (3) waste, and (4) heavy oilbased IGCCs.

4.10.6.1

Coal-Based Integrated Gasification Combined Cycle

The coal-based IGCC power systems have proven capable of exceeding the tightest emissions regulations currently applicable to comparable combustion-based power systems. The coal-based IGCC systems have reached the lower levels of NOx, SOx, and CO pollutant air emissions compared to any coal-fueled power production processes. Another important environmental indicator is the CO2 emissions reduction, by at least 10% for an equivalent net power generation, due to higher working performance comparing to existing coal-fueled combustion-based electricity production process. If more important CO2 reduction is wanted in the power production system, the coal gasification processes have significant working benefits that could be reached to CO2 capture more efficiently than currently possible with coal-fueled combustion process. Therefore, coal-based IGCC power system can be classified based on without CO2 capture and with CO2 capture.

4.10.6.1.1

Coal-based integrated gasification combined cycle without CO2 capture

Fig. 29 illustrates the schematic diagram of a coal-based IGCC plant without CO2 capture. The coal samples are supplied to the coal gasifier subcomponent where it is partially oxidized under pressure (30–80 bar). The ASU is used in the coal-based IGCC plant because its utilizes oxygen as oxidant. The temperature of coal gasifier subcomponent, which is of the entrained flow slagging type, should exceed 14001C. The high temperature of coal gasifier provides that the ash is transformed to the liquid-slag with low viscosity; therefore ash could easily exit from the coal gasifier. The coal-based IGCC plants have achieved the lower levels of NOx, SOx, and CO pollutant air emissions compared to any coal-based power facilities in the world. Fig. 29 shows the coal-based IGCC without CO2 capture containing two air compressors, an ASU, a coal gasifier, particulate removal, sulfur removal, syngas storage, nitrogen storage, combustion chamber, gas turbine, HRSG to produce superheated steam, steam turbine, condenser, pump to increase working fluid pressure level, and HEX to produce heat for domestic applications.

Integrated Gasification Combined Cycles

10 H2S

Sulfur removal

11

411

Syngas storage 12

9 8 Tar

Particulate removal

Nitrogen 13 storage

Coal gasifier

Coal

Gas turbine

17

Power

15

Steam 4

5 Waste

Combustion chamber

16

7 6

14

Air compressor-II

23

18 HRSG

Turbine

ASU

3

Power

19

22 2

HEX

Air compressor-I

21

24 Heating application

26

20

28 25

Stack

1

Condenser

Pump 27

Fig. 29 Integrated coal gasification combined cycle without CO2 capture. ASU, air separation unit; HEX, heat exchanger; HRSG, heat recovery steam generator.

The balance equations of integrated coal gasification-based combined cycle components without CO2 capture are written in subsections. Air compressor-I: under steady state and steady flow conditions, the balance equations for air compressor-I can be written as _2 Mass: m _1 ¼m

ð376Þ

_ ac_I ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð377Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;ac_I ¼ m

ð378Þ

_ ac_I ¼ m _ D;ac_I _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð379Þ

ASU: the mass, energy, entropy, and exergy balance equations for ASU can be written under steady state and steady flow conditions as follows: _3þm _4þm _5 Mass: m _2 ¼m

ð380Þ

_ 3 h3 þ m _ 4 h4 þ m _ 5 h5 Energy: m _ 2 h2 ¼ m

ð381Þ

_ 3 s3 þ m _ 4 s4 þ m _ 5 s5 Entropy: m _ 2 s2 þ S_ gen; ASU ¼ m

ð382Þ

_ D;ASU _ 3 ex 3 þ m _ 4 ex 4 þ m _ 5 ex5 þ Ex Exergy: m _ 2 ex2 ¼ m

ð383Þ

Coal gasifier: under the steady state and steady flow conditions, the balance equations for coal gasifier can be written as follows: _6þm _ 22 ¼ m _7 Mass: m _5þm

ð384Þ

_ 6 h6 þ m _ 22 h222 ¼ m _ 7 h7 Energy: m _ 5 h5 þ m

ð385Þ

_ 6 s6 þ m _ 22 s22 þ S_ gen;cg ¼ m _ 7 s7 Entropy: m _ 5 s5 þ m

ð386Þ

_ D;cg _ 6 ex6 þ m _ 22 ex22 ¼ m _ 7 ex 7 þ Ex Exergy: m _ 5 ex5 þ m

ð387Þ

412

Integrated Gasification Combined Cycles

Particulate removal: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for particulate removal subcomponent are _8þm _9 Mass: m _7¼m

ð388Þ

_ 8 h8 þ m _ 9 h9 Energy: m _ 7 h7 ¼ m

ð389Þ

_ 8 s8 þ m _ 9 s9 Entropy: m _ 7 s7 þ S_ gen;pr ¼ m

ð390Þ

_ D;pr _ 8 ex8 þ m _ 9 ex 9 þ Ex Exergy: m _ 7 ex 7 ¼ m

ð391Þ

Sulfur removal: under the steady state and steady flow conditions, the balance equations for sulfur removal subcomponent can be written as follows: _ 10 þ m _ 11 Mass: m _9¼m

ð392Þ

_ 10 h10 þ m _ 11 h11 Energy: m _ 9 h9 ¼ m

ð393Þ

_ 10 s10 þ m _ 11 s11 Entropy: m _ 9 s9 þ S_ gen;sr ¼ m

ð394Þ

_ D;sr _ 10 ex 10 þ m _ 11 ex11 þ Ex Exergy: m _ 9 ex9 ¼ m

ð395Þ

Syngas storage: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for syngas storage subcomponent are defined as _ 12 Mass: m _ 11 ¼ m

ð396Þ

_ 12 h12 Energy: m _ 11 h11 ¼ m

ð397Þ

_ 12 s12 Entropy: m _ 11 s11 þ S_ gen;ss ¼ m

ð398Þ

_ D;ss _ 12 ex12 þ Ex Exergy: m _ 11 ex 11 ¼ m

ð399Þ

Nitrogen storage: the mass, energy, entropy, and exergy balance equations for nitrogen storage subcomponent can be written under steady state and steady flow conditions as follows: _ 13 Mass: m _4¼m

ð400Þ

_ 13 h13 Energy: m _ 4 h4 ¼ m

ð401Þ

_ 13 s13 Entropy: m _ 4 s4 þ S_ gen;ns ¼ m

ð402Þ

_ D;ns _ 13 ex 13 þ Ex Exergy: m _ 4 ex 4 ¼ m

ð403Þ

Air compressor-II: under steady state and steady flow conditions, the balance equations for air compressor-II subcomponent can be defined as _ 16 Mass: m _ 15 ¼ m

ð404Þ

_ ac_II ¼ m _ 16 h16 Energy: m _ 15 h15 þ W

ð405Þ

_ 16 s16 Entropy: m _ 15 s15 þ S_ gen;ac_II ¼ m

ð406Þ

_ ac_II ¼ m _ D;ac_II _ 16 ex 16 þ Ex Exergy: m _ 15 ex15 þ W

ð407Þ

Combustion chamber: under the steady state and steady flow conditions, the balance equations for combustion chamber are written as follows: _ 16 ¼ m _ 17 Mass: m _ 14 þ m

ð408Þ

Integrated Gasification Combined Cycles

413

_ 16 h16 ¼ m _ 17 h17 Energy: m _ 14 h14 þ m

ð409Þ

_ 16 s16 þ S_ gen;cc ¼ m _ 17 s17 Entropy: m _ 14 s14 þ m

ð410Þ

_ D;cc _ 16 ex 16 ¼ m _ 17 ex17 þ Ex Exergy: m _ 14 ex14 þ m

ð411Þ

Gas turbine: under steady state and steady flow conditions, the balance equations for gas turbine can be given as _ 18 Mass: m _ 17 ¼ m

ð412Þ

_ gt _ 18 h18 þ W Energy: m _ 17 h17 ¼ m

ð413Þ

_ 18 s18 Entropy: m _ 17 s17 þ S_ gen;gt ¼ m

ð414Þ

_ gt þ Ex _ D;gt _ 18 ex18 þ W Exergy: m _ 17 ex 17 ¼ m

ð415Þ

HRSG subcomponent: under the steady state and steady flow conditions, the balance equations for HRSG subcomponent can written as _ 19 Mass: m _ 18 ¼ m

ð416Þ

_ 26 h26 ¼ m _ 19 h19 þ m _ 23 h23 Energy: m _ 18 h18 þ m

ð417Þ

_ 26 s26 þ S_ gen;HRSG ¼ m _ 19 s19 þ m _ 23 s23 Entropy: m _ 18 s18 þ m

ð418Þ

_ D;HRSG _ 26 ex26 ¼ m _ 19 ex 19 þ m _ 23 ex23 þ Ex Exergy: m _ 18 ex 18 þ m

ð419Þ

Steam turbine: under steady state and steady flow conditions, the balance equations for steam turbine can be defined as follows: _ 24 Mass: m _ 23 ¼ m

ð420Þ

_ st _ 24 h24 þ W Energy: m _ 23 h23 ¼ m

ð421Þ

_ 24 s24 Entropy: m _ 23 s23 þ S_ gen;st ¼ m

ð422Þ

_ st þ Ex _ D;st _ 24 ex 24 þ W Exergy: m _ 23 ex 23 ¼ m

ð423Þ

Condenser: under the steady state and steady flow conditions, the balance equations for condenser subcomponent can be written as _ 25 ; m _ 28 Mass: m _ 24 ¼ m _ 27 ¼ m

ð424Þ

_ 27 h27 ¼ m _ 25 h25 þ m _ 28 h28 Energy: m _ 24 h24 þ m

ð425Þ

_ 27 s27 þ S_ gen;con ¼ m _ 25 s25 þ m _ 28 s28 Entropy: m _ 24 s24 þ m

ð426Þ

_ D;con _ 27 ex 27 ¼ m _ 25 ex25 þ m _ 28 ex 28 þ Ex Exergy: m _ 24 ex24 þ m

ð427Þ

Pump: for pump subcomponent, the balance equations are provided under the steady state and steady flow conditions. _ 26 Mass: m _ 25 ¼ m

ð428Þ

_ p¼m _ 26 h26 Energy: m _ 25 h25 þ W

ð429Þ

_ 26 s26 Entropy: m _ 25 s25 þ S_ gen; p ¼ m

ð430Þ

414

Integrated Gasification Combined Cycles

14 CO2

CO2 capture

15

Syngas storage

13 12 H2S

Sulfur removal 16 11

10 Steam

WGSR 9

8 Tar

17

Nitrogen storage

Particulate removal

18

Combustion chamber

20

Gas turbine

21

7 6

Coal gasifier

Steam

Air compressor-II

Coal 5 Waste

ASU

3

Power

19

27

22 HRSG

4

Turbine Power

23

26 2

HEX

Air compressor-I

25 1

28 Heating application 32

30

24 29 Stack

Condenser

Pump 31

Fig. 30 Integrated coal gasification combined cycle with CO2 capture. ASU, air separation unit; HEX, heat exchanger; WGSR, water-gas shift reactor.

_ p ¼m _ D;p _ 26 ex 26 þ Ex Exergy: m _ 25 ex 25 þ W

ð431Þ

HEX: under the steady state and steady flow conditions, the balance equations for HEX subcomponent can written as follows:

4.10.6.1.2

_ 20 ; m _ 22 Mass: m _ 19 ¼ m _ 21 ¼ m

ð432Þ

_ 21 h21 ¼ m _ 20 h20 þ m _ 22 h22 Energy: m _ 19 h19 þ m

ð433Þ

_ 21 s21 þ S_ gen; HEX ¼ m _ 20 s20 þ m _ 22 s22 Entropy: m _ 19 s19 þ m

ð434Þ

_ D;HEX _ 21 ex 21 ¼ m _ 20 ex 20 þ m _ 22 ex 22 þ Ex Exergy: m _ 19 ex19 þ m

ð435Þ

Coal-based integrated gasification combined cycle with CO2 capture

The schematic diagram of a coal-based IGCC plant with CO2 capture is shown in Fig. 30. Air at reference conditions enters the air compressor-I at number 1 and exits after compression at number 2. The pressured air enters the ASU, and separated hot nitrogen exits at point 4 and is stored in nitrogen storage subcomponent. The hot air at point 5 and steam coming from HEX-II at point 26 enter the coal gasifier into which coal samples are injected, and hot combustion gases exit at point 7 and enter the particular removal subcomponent. The hot gas at point 9 and steam at point 10 enter the WGSR for a higher hydrogen rich syngas yield. The synthesis gaseous from gasification cycle can be cleaned to very low levels of contaminants, such as sulfur compounds and particulates. The hydrogen rich syngas enters the sulfur removal and CO2 capture subcomponents for removing H2S and CO2, respectively, and is stored in syngas storage subcomponent for continuously power generation and heat supply. The compressed air at point 20, nitrogen at point 17, and syngas at point 16 enter the combustion chamber, and hot combustion gaseous exit at point 21 and pass through the gas turbine to generate electricity. Hot stack gaseous enter the HRSG to provide HP steam at point

Integrated Gasification Combined Cycles

415

27. The HP working fluid enters the Rankine turbine to produce shaft power. The low pressure working fluid has adequate energy for utilize in heating applications. The balance equations of integrated coal gasification-based combined cycle components with CO2 capture are defined in subsections. Air compressor-I: under steady state and steady flow conditions, the balance equations for air compressor-I subcomponent are defined as follows: _2 Mass: m _1 ¼m

ð436Þ

_ ac_I ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð437Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;ac_I ¼ m

ð438Þ

_ ac_I ¼ m _ D;ac_I _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð439Þ

ASU: under steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for ASU can be written as follows: _3þm _4þm _5 Mass: m _2 ¼m

ð440Þ

_ 3 h3 þ m _ 4 h4 þ m _ 5 h5 Energy: m _ 2 h2 ¼ m

ð441Þ

_ 3 s3 þ m _ 4 s4 þ m _ 5 s5 Entropy: m _ 2 s2 þ S_ gen; ASU ¼ m

ð442Þ

_ D; ASU _ 3 ex 3 þ m _ 4 ex 4 þ m _ 5 ex5 þ Ex Exergy: m _ 2 ex 2 ¼ m

ð443Þ

Coal gasifier: under the steady state and steady flow conditions, the balance equations for coal gasifier can be written as follows: _6þm _ 26 ¼ m _7 Mass: m _5þm

ð444Þ

_ 6 h6 þ m _ 26 h26 ¼ m _ 7 h7 Energy: m _ 5 h5 þ m

ð445Þ

_ 6 s6 þ m _ 26 s26 þ S_ gen;cg ¼ m _ 7 s7 Entropy: m _ 5 s5 þ m

ð446Þ

_ D;cg _ 6 ex6 þ m _ 26 ex 26 ¼ m _ 7 ex 7 þ Ex Exergy: m _ 5 ex5 þ m

ð447Þ

Particulate removal: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for particulate removal subcomponent are _8þm _9 Mass: m _7¼m

ð448Þ

_ 8 h8 þ m _ 9 h9 Energy: m _ 7 h7 ¼ m

ð449Þ

_ 8 s8 þ m _ 9 s9 Entropy: m _ 7 s7 þ S_ gen;pr ¼ m

ð450Þ

_ D;pr _ 8 ex 8 þ m _ 9 ex 9 þ Ex Exergy: m _ 7 ex 7 ¼ m

ð451Þ

WGSR: under the steady state and steady flow conditions, the balance equations for WGSR subcomponent can be written as follows: _ 10 ¼ m _ 11 Mass: m _9þm

ð452Þ

_ 10 h10 ¼ m _ 11 h11 Energy: m _ 9 h9 þ m

ð453Þ

_ 10 s10 þ S_ gen;WGSR ¼ m _ 11 s11 Entropy: m _ 9 s9 þ m

ð454Þ

_ D;WGSR _ 10 ex10 ¼ m _ 11 ex 11 þ Ex Exergy: m _ 9 ex9 þ m

ð455Þ

416

Integrated Gasification Combined Cycles

Sulfur removal: under the steady state and steady flow conditions, the balance equations for sulfur removal are defined as _ 12 þ m _ 13 Mass: m _ 11 ¼ m

ð456Þ

_ 12 h12 þ m _ 13 h13 Energy: m _ 11 h11 ¼ m

ð457Þ

_ 12 s12 þ m _ 13 s13 Entropy: m _ 11 s11 þ S_ gen;sr ¼ m

ð458Þ

_ D;sr _ 12 ex12 þ m _ 13 ex13 þ Ex Exergy: m _ 11 ex11 ¼ m

ð459Þ

CO2 capture: under the steady state and steady flow conditions, the balance equations for CO2 capture can be defined as follows: _ 14 þ m _ 15 Mass: m _ 13 ¼ m

ð460Þ

_ 14 h14 þ m _ 15 h15 Energy: m _ 13 h13 ¼ m

ð461Þ

Entropy: m _ 13 s13 þ S_ gen;CO2

cap

_ 14 s14 þ m _ 15 s15 ¼m

_ D;CO2 _ 14 ex 14 þ m _ 15 ex 15 þ Ex Exergy: m _ 13 ex 13 ¼ m

cap

ð462Þ ð463Þ

Syngas storage: under the steady state and steady flow conditions, the balance equations for syngas storage subcomponent are written as follows: _ 16 Mass: m _ 15 ¼ m

ð464Þ

_ 16 h16 Energy: m _ 15 h15 ¼ m

ð465Þ

_ 16 s16 Entropy: m _ 15 s15 þ S_ gen;ss ¼ m

ð466Þ

_ D;ss _ 16 ex16 þ Ex Exergy: m _ 15 ex 15 ¼ m

ð467Þ

Nitrogen storage: the balance equations for nitrogen storage subcomponent can be written under steady state and steady flow conditions as follows: _ 17 Mass: m _4¼m

ð468Þ

_ 17 h17 Energy: m _ 4 h4 ¼ m

ð469Þ

_ 17 s17 Entropy: m _ 4 s4 þ S_ gen;ns ¼ m

ð470Þ

_ D;ns _ 17 ex17 þ Ex Exergy: m _ 4 ex 4 ¼ m

ð471Þ

Air compressor-II: under steady state and steady flow conditions, the balance equations for air compressor-II subcomponent can be written as follows: _ 20 Mass: m _ 19 ¼ m

ð472Þ

_ ac_II ¼ m _ 19 h19 þ W _ 20 h20 m

ð473Þ

_ 20 s20 Entropy: m _ 19 s19 þ S_ gen;ac_II ¼ m

ð474Þ

_ ac_II ¼ m _ D;ac_II _ 20 ex 20 þ Ex Exergy: m _ 19 ex19 þ W

ð475Þ

Combustion chamber: under the steady state and steady flow conditions, the balance equations for combustion chamber are written as follows: _ 20 ¼ m _ 21 Mass: m _ 18 þ m

ð476Þ

_ 20 h20 ¼ m _ 21 h21 Energy: m _ 18 h18 þ m

ð477Þ

Integrated Gasification Combined Cycles

417

_ 20 s20 þ S_ gen;cc ¼ m _ 21 s21 Entropy: m _ 18 s18 þ m

ð478Þ

_ D; cc _ 20 ex20 ¼ m _ 21 ex21 þ Ex Exergy: m _ 18 ex 18 þ m

ð479Þ

Gas turbine: under steady state and steady flow conditions, the balance equations for gas turbine can be defined as follows: _ 22 Mass: m _ 21 ¼ m

ð480Þ

_ gt _ 22 h22 þ W Energy: m _ 21 h21 ¼ m

ð481Þ

_ 22 s22 Entropy: m _ 21 s21 þ S_ gen;gt ¼ m

ð482Þ

_ gt þ Ex _ D; gt _ 22 ex 22 þ W Exergy: m _ 21 ex21 ¼ m

ð483Þ

HRSG subcomponent: under the steady state and steady flow conditions, the balance equations for HRSG can written as _ 23 ; m _ 30 Mass: m _ 22 ¼ m _ 27 ¼ m

ð484Þ

_ 30 h30 ¼ m _ 23 h23 þ m _ 27 h27 Energy: m _ 22 h22 þ m

ð485Þ

_ 30 s30 þ S_ gen;HRSG ¼ m _ 23 s23 þ m _ 27 s27 Entropy: m _ 22 s22 þ m

ð486Þ

_ D;HRSG _ 30 ex30 ¼ m _ 23 ex 23 þ m _ 27 ex27 þ Ex Exergy: m _ 22 ex 22 þ m

ð487Þ

Steam turbine: under steady state and steady flow conditions, the balance equations for steam turbine can be defined as follows: _ 28 Mass: m _ 27 ¼ m

ð488Þ

_ st _ 28 h28 þ W Energy: m _ 27 h27 ¼ m

ð489Þ

_ 28 s28 Entropy: m _ 27 s27 þ S_ gen;st ¼ m

ð490Þ

_ st þ Ex _ D; st _ 28 ex28 þ W Exergy: m _ 27 ex27 ¼ m

ð491Þ

Condenser: under the steady state and steady flow conditions, the balance equations for condenser subcomponent are written as _ 29 ; m _ 32 Mass: m _ 28 ¼ m _ 31 ¼ m

ð492Þ

_ 31 h31 ¼ m _ 29 h29 þ m _ 32 h32 Energy: m _ 28 h28 þ m

ð493Þ

_ 31 s31 þ S_ gen;con ¼ m _ 29 s29 þ m _ 32 s32 Entropy: m _ 28 s28 þ m

ð494Þ

_ D; con _ 31 ex31 ¼ m _ 29 ex 29 þ m _ 32 ex32 þ Ex Exergy: m _ 28 ex 28 þ m

ð495Þ

Pump: for pump subcomponent, the balance equations are provided under the steady state and steady flow conditions. _ 30 Mass: m _ 29 ¼ m

ð496Þ

_ p¼m _ 30 h30 Energy: m _ 29 h29 þ W

ð497Þ

_ 30 s30 _ 29 s29 þ S_ gen;p ¼ m Entropy: cm

ð498Þ

_ p ¼m _ D;p _ 30 ex 30 þ Ex Exergy: m _ 29 ex 29 þ W

ð499Þ

HEX: under the steady state and steady flow conditions, the balance equations for HEX subcomponent are defined as follows: _ 24 ; m _ 26 Mass: m _ 23 ¼ m _ 25 ¼ m

ð500Þ

418

Integrated Gasification Combined Cycles

8 Tar

Particulate removal

Syngas storage

9

10

Combustion chamber

12

7

Gas turbine

13

Power

11 Biomass gasifier

Steam

Air compressor

4 Rotary dryer

18

14 HRSG

6

Turbine Power

15

3 Wet biomass

HEX-I 2

5

19 Heating application

21

16

23 20

HEX-II 1

Condenser

Pump

17

22

Stack Fig. 31 Integrated biomass gasification combined cycle. HEX, heat exchanger; HRSG, heat recovery steam generator.

4.10.6.2

_ 25 h25 ¼ m _ 24 h24 þ m _ 26 h26 Energy: m _ 23 h23 þ m

ð501Þ

_ 25 s25 þ S_ gen;HEX ¼ m _ 24 s24 þ m _ 26 s26 Entropy:m _ 23 s23 þ m

ð502Þ

_ D; HEX _ 25 ex25 ¼ m _ 24 ex24 þ m _ 26 ex26 þ Ex Exergy: m _ 23 ex23 þ m

ð503Þ

Biomass-Based Integrated Gasification Combined Cycle

Generally, the biomass waste materials including wood waste, sugarcane bagasse, rice hull, corn cob, and cotton are stored in a storage silo or contained in a dryer hopper. A schematic diagram of biomass-based IGCC process is illustrated in Fig. 31. A predrying process is needed before the biomass materials enters the biomass gasifier. Waste heat from the stack gaseous of the HEX-II can be used to dry the biomass materials in the rotary dryer and therefore, not wasted by exhausting into the air. Air at reference conations enters the HEX-II at point 1, and heated air and wet biomass materials go into the rotary dryer at points 2 and 3, respectively. Dry biomass and steam enter the biomass gasifier at points 4 and 6, respectively. The produced syngas goes into the particulate removal subcomponent where a significant amount of particulate is removed from the syngas, and is stored in the syngas storage subcomponent for continuously electricity production generation and heat supply. Air at reference conditions enters the air compressor-II at point 11. The compressed air at point 12 and syngas at point 10 enter the combustion chamber, and hot combustion gaseous exit at point 13 and pass through the gas turbine to produce power. Hot stack gaseous enter the HRSG to provide HP steam at point 18. The HP working fluid enters the turbine to produce shaft power. The low pressure working fluid has adequate energy for utilize in heating applications. The balance equations of integrated biomass gasification-based combined cycle components are written in subsections. Rotary dryer: the mass, energy, entropy, and exergy balance equations for rotary dryer subcomponent can be written under the steady state and steady flow conditions as follows: _3þm _4 Mass: m _2¼m

ð504Þ

_ 3 h3 þ m _ 4 h4 Energy: m _ 2 h2 ¼ m

ð505Þ

_ 3 s3 þ m _ 4 s4 Entropy: m _ 2 s2 þ S_ gen;rd ¼ m

ð506Þ

_ D;rd _ 3 ex3 þ m _ 4 ex4 þ Ex Exergy: m _ 2 ex2 ¼ m

ð507Þ

Biomass gasifier: under the steady state and steady flow conditions, the balance equations for biomass gasifier subcomponent can be written as

Integrated Gasification Combined Cycles

419

_6¼m _7 Mass: m _4þm

ð508Þ

_ 6 h6 ¼ m _ 7 h7 Energy: m _ 4 h4 þ m

ð509Þ

_ 6 s6 þ S_ gen;bg ¼ m _ 7 s7 Entropy: m _ 4 s4 þ m

ð510Þ

_ D;bg _ 6 ex 6 ¼ m _ 7 ex7 þ Ex Exergy: m _ 4 ex4 þ m

ð511Þ

Particulate removal: the mass, energy, entropy, and exergy balance equations for particulate removal subcomponent can be defined under the steady state and steady flow conditions as follows: _8þm _9 Mass: m _7¼m

ð512Þ

_ 8 h8 þ m _ 9 h9 Energy: m _ 7 h7 ¼ m

ð513Þ

_ 8 s8 þ m _ 9 s9 Entropy: m _ 7 s7 þ S_ gen;pr ¼ m

ð514Þ

_ D;pr _ 8 ex8 þ m _ 9 ex 9 þ Ex Exergy: m _ 7 ex 7 ¼ m

ð515Þ

Syngas storage: under the steady state and steady flow conditions, the balance equations for syngas storage subcomponent can be written as _ 10 Mass: m _9¼m

ð516Þ

_ 10 h10 Energy: m _ 9 h9 ¼ m

ð517Þ

_ 10 s10 Entropy: m _ 9 s9 þ S_ gen;ss ¼ m

ð518Þ

_ D;ss _ 10 ex10 þ Ex Exergy: m _ 9 ex 9 ¼ m

ð519Þ

Air compressor: under steady state and steady flow conditions, the balance equations for air compressor subcomponent can be written as follows: _ 12 Mass: m _ 11 ¼ m

ð520Þ

_ ac ¼ m _ 12 h12 Energy: m _ 11 h11 þ W

ð521Þ

_ 12 s12 Entropy: m _ 11 s11 þ S_ gen;ac ¼ m

ð522Þ

_ ac ¼ m _ D;ac _ 12 ex 12 þ Ex Exergy: m _ 11 ex11 þ W

ð523Þ

Combustion chamber: under the steady state and steady flow conditions, the balance equations for combustion chamber are defined as _ 12 ¼ m _ 13 Mass: m _ 10 þ m

ð524Þ

_ 12 h12 ¼ m _ 13 h13 Energy: m _ 10 h10 þ m

ð525Þ

_ 12 s12 þ S_ gen;cc ¼ m _ 13 s13 Entropy: m _ 10 s10 þ m

ð526Þ

_ D;cc _ 12 ex 12 ¼ m _ 13 ex13 þ Ex Exergy: m _ 10 ex10 þ m

ð527Þ

Gas turbine: under steady state and steady flow conditions, the balance equations for gas turbine can be written as follows: _ 14 Mass: m _ 13 ¼ m

ð528Þ

_ gt _ 14 h14 þ W Energy: m _ 13 h13 ¼ m

ð529Þ

420

Integrated Gasification Combined Cycles

_ 14 s14 Entropy: m _ 13 s13 þ S_ gen;gt ¼ m

ð530Þ

_ gt þ Ex _ D;gt _ 14 ex14 þ W Exergy: m _ 13 ex 13 ¼ m

ð531Þ

HRSG: under the steady state and steady flow conditions, the balance equations for HRSG can be given as follows: _ 15 ; m _ 21 _ 18 ¼ m Mass: m _ 14 ¼ m

ð532Þ

_ 21 h21 ¼ m _ 15 h15 þ m _ 18 h18 Energy: m _ 14 h14 þ m

ð533Þ

_ 21 s21 þ S_ gen;HRSG ¼ m _ 15 s15 þ m _ 18 s18 Entropy: m _ 14 s14 þ m

ð534Þ

_ D;HRSG _ 21 ex21 ¼ m _ 15 ex 15 þ m _ 18 ex18 þ Ex Exergy: m _ 14 ex 14 þ m

ð535Þ

Steam turbine: under steady state and steady flow conditions, the balance equations for steam turbine are defined as _ 19 Mass: m _ 18 ¼ m

ð536Þ

_ st _ 19 h19 þ W Energy: m _ 18 h18 ¼ m

ð537Þ

_ 19 s19 Entropy: m _ 18 s18 þ S_ gen;st ¼ m

ð538Þ

_ st þ Ex _ D;st _ 19 ex 19 þ W Exergy: m _ 18 ex 18 ¼ m

ð539Þ

Condenser: under the steady state and steady flow conditions, the balance equations for condenser subcomponent can be written as follows: _ 20 ; m _ 23 _ 22 ¼ m Mass: m _ 19 ¼ m

ð540Þ

_ 22 h22 ¼ m _ 20 h20 þ m _ 23 h23 Energy: m _ 19 h19 þ m

ð541Þ

_ 22 s22 þ S_ gen;con ¼ m _ 20 s20 þ m _ 23 s23 Entropy: m _ 19 s19 þ m

ð542Þ

_ D;con _ 22 ex 22 ¼ m _ 20 ex20 þ m _ 23 ex 23 þ Ex Exergy: m _ 19 ex19 þ m

ð543Þ

Pump: for pump subcomponent, the balance equations are provided under the steady state and steady flow conditions. _ 21 Mass: m _ 20 ¼ m

ð544Þ

_ p¼m _ 21 h21 Energy: m _ 20 h20 þ W

ð545Þ

_ 21 s21 Entropy: m _ 20 s20 þ S_ gen;p ¼ m

ð546Þ

_ p ¼m _ D;p _ 21 ex 21 þ Ex Exergy: m _ 20 ex 20 þ W

ð547Þ

HEX-I: under the steady state and steady flow conditions, the balance equations for HEX-I subcomponent can be given as follows: _ 6; m _ 16 _ 15 ¼ m Mass: m _5¼m

ð548Þ

_ 15 h15 ¼ m _ 6 h6 þ m _ 16 h16 Energy: m _ 5 h5 þ m

ð549Þ

_ 15 s15 þ S_ gen;HEX Entropy: m _ 5 s5 þ m

I

_ 6 s6 þ m _ 16 s16 ¼m

_ D;HEX _ 15 ex 15 ¼ m _ 6 ex 6 þ m _ 16 ex16 þ Ex Exergy: m _ 5 ex 5 þ m

ð550Þ

I

ð551Þ

HEX-II: the mass, energy, entropy, and exergy balance equations for HEX-II subcomponent can be defined under the steady state and steady flow conditions as follows:

Integrated Gasification Combined Cycles

7 Tar

Particulate removal

Syngas storage

8

9

Combustion chamber

Gas turbine

12

11

6

Power

10 1

Waste materials gasifier Waste materials 4 5

Air compressor

16

13 HRSG

Vitrified Recycled glass metals

421

Steam

Turbine Power

14

3

HEX 2

17

19

15 18

Stack

Heating application 21

Condenser

Pump 20 Fig. 32 Integrated waste materials gasification combined cycle. HEX, heat exchanger; HRSG, heat recovery steam generator.

_ 2; m _ 17 Mass: m _1¼m _ 16 ¼ m

ð552Þ

_ 16 h16 ¼ m _ 2 h2 þ m _ 17 h17 Energy: m _ 1 h1 þ m

ð553Þ

_ 16 s16 þ S_ gen;HEX Entropy: m _ 1 s1 þ m

II

_ 2 s2 þ m _ 17 s17 ¼m

_ D;HEX _ 16 ex 16 ¼ m _ 2 ex 2 þ m _ 17 ex1 þ Ex Exergy: m _ 1 ex 1 þ m

4.10.6.3

ð554Þ

II

ð555Þ

Waste Materials-Based Integrated Gasification Combined Cycle

The Brayton cycle can be used as a prime mover for the integrated system or it can be combined with another prime mover. In this study, waste materials based IGCC process is examined. This process is combined Brayton cycle with Rankine cycle to produce power and useful heat output. The schematic diagram of integrated waste materials gasification combined cycle is shown in Fig. 32. Waste heat from the stack gaseous from the HEX can be used to produce steam for waste materials gasifier, therefore, not wasted by exhausting into the air. The produced syngas goes into the particulate removal subcomponent where a significant amount of particulate is removed from the syngas, and can be stored in the syngas storage subcomponent for continuously electricity production generation and heat supply. Air at reference conditions enters the air compressor-II at point 10. The compressed air at point 11 and syngas at point 9 enter the combustion chamber, and hot combustion gaseous exit at point 12 and pass through the gas turbine to generate electricity. Hot stack gaseous enter the HRSG to provide HP steam. The working fluid of Rankine cycle enters the HRSG in a liquid state and exits as vapor at point 16. Then, the working fluid expands through the steam turbine to generate the mechanical power. Next, the working fluid exits from the steam turbine at point 17 and supplies heat to the heating-process condenser. The heating-process condenser rejects heat to supply the heating applications. Then, the working fluid enters the condenser at point 17 as saturated vapor. The condenser absorbs heat to supply the heating load for the heating applications. Next, the working fluid exits from the condenser again as saturated liquid at point 18. After that, the pump increases the pressure of the saturated liquid at point 19. The mass, energy, entropy, and exergy balance equations for integrated waste materials gasification-based combined cycle components are defined in subsections. Waste materials gasifier: under the steady state and steady flow conditions, the balance equations for waste materials gasifier can be defined as _3¼m _4þm _5þm _6 Mass: m _1þm

ð556Þ

_ 3 h3 ¼ m _ 4 h4 þ m _ 5 h5 þ m _ 6 h6 Energy: m _ 1 h1 þ m

ð557Þ

_ 3 s3 þ S_ gen;WMG ¼ m _ 4 s4 þ m _ 5 s5 þ m _ 6 s6 Entropy: m _ 1 s1 þ m

ð558Þ

422

Integrated Gasification Combined Cycles _ D;WMG _ 3 ex3 ¼ m _ 4 ex4 þ m _ 5 ex5 þ m _ 6 ex 6 þ Ex Exergy: m _ 1 ex1 þ m

ð559Þ

Particulate removal: under the steady state and steady flow conditions, the balance equations for particulate removal subcomponent can be written as follows: _7þm _8 Mass: m _6¼m

ð560Þ

_ 7 h7 þ m _ 8 h8 Energy: m _ 6 h6 ¼ m

ð561Þ

_ 7 s7 þ m _ 8 s8 Entropy: m _ 6 s6 þ S_ gen;pr ¼ m

ð562Þ

_ D;pr _ 7 ex7 þ m _ 8 ex 8 þ Ex Exergy: m _ 6 ex 6 ¼ m

ð563Þ

Syngas storage: under the steady state and steady flow conditions, the balance equations for syngas storage subcomponent are given as _9 Mass: m _8 ¼m

ð564Þ

_ 9 h9 Energy: m _ 8 h8 ¼ m

ð565Þ

_ 9 s9 Entropy: m _ 8 s8 þ S_ gen;ss ¼ m

ð566Þ

_ D;ss _ 9 ex 9 þ Ex Exergy:m _ 8 ex 8 ¼ m

ð567Þ

Air compressor: under steady state and steady flow conditions, the balance equations for air compressor-II subcomponent can be written as follows: _ 11 Mass: m _ 10 ¼ m

ð568Þ

_ ac ¼ m _ 11 h11 Energy: m _ 10 h10 þ W

ð569Þ

_ 11 s11 Entropy: m _ 10 s10 þ S_ gen;ac ¼ m

ð570Þ

_ ac ¼ m _ D;ac _ 11 ex 11 þ Ex Exergy: m _ 10 ex10 þ W

ð571Þ

Combustion chamber: under the steady state and steady flow conditions, the balance equations for combustion chamber can be defined as follows: _ 11 ¼ m _ 12 Mass: m _9þm

ð572Þ

_ 11 h11 ¼ m _ 12 h12 Energy: m _ 9 h9 þ m

ð573Þ

_ 11 s11 þ S_ gen;cc ¼ m _ 12 s12 Entropy: m _ 9 s9 þ m

ð574Þ

_ D;cc _ 11 ex11 ¼ m _ 12 ex12 þ Ex Exergy: m _ 9 ex9 þ m

ð575Þ

Gas turbine: under steady state and steady flow conditions, the balance equations for gas turbine are given as _ 13 Mass: m _ 12 ¼ m

ð576Þ

_ gt _ 13 h13 þ W Energy: m _ 12 h12 ¼ m

ð577Þ

_ 13 s13 Entropy: m _ 12 s12 þ S_ gen;gt ¼ m

ð578Þ

_ gt þ Ex _ D;gt _ 13 ex13 þ W Exergy: m _ 12 ex 12 ¼ m

ð579Þ

Integrated Gasification Combined Cycles

423

HRSG subcomponent: under the steady state and steady flow conditions, the balance equations for HRSG can written as follows: _ 14 ; m _ 19 _ 16 ¼ m Mass: m _ 13 ¼ m

ð580Þ

_ 19 h19 ¼ m _ 14 h14 þ m _ 16 h16 Energy: m _ 13 h13 þ m

ð581Þ

_ 19 s19 þ S_ gen;HRSG ¼ m _ 14 s14 þ m _ 16 s16 Entropy: m _ 13 s13 þ m

ð582Þ

_ D;HRSG _ 19 ex19 ¼ m _ 14 ex 14 þ m _ 16 ex16 þ Ex Exergy: m _ 13 ex 13 þ m

ð583Þ

Turbine: under steady state and steady flow conditions, the balance equations for turbine can be defined as _ 17 Mass: m _ 16 ¼ m

ð584Þ

_ tur _ 17 h17 þ W Energy: m _ 16 h16 ¼ m

ð585Þ

_ 17 s17 Entropy: m _ 16 s16 þ S_ gen;tur ¼ m

ð586Þ

_ tur þ Ex _ D;tur _ 17 ex17 þ W Exergy: m _ 16 ex16 ¼ m

ð587Þ

Condenser: under the steady state and steady flow conditions, the balance equations for condenser subcomponent are given as follows: _ 18 ; m _ 21 Mass: m _ 17 ¼ m _ 20 ¼ m

ð588Þ

_ 20 h20 ¼ m _ 18 h18 þ m _ 21 h21 Energy: m _ 17 h17 þ m

ð589Þ

_ 20 s20 þ S_ gen;con ¼ m _ 18 s18 þ m _ 21 s21 Entropy: m _ 17 s17 þ m

ð590Þ

_ D;con _ 20 ex 20 ¼ m _ 18 ex18 þ m _ 21 ex 21 þ Ex Exergy: m _ 17 ex17 þ m

ð591Þ

Pump: for pump subcomponent, the balance equations are provided under the steady state and steady flow conditions. _ 19 Mass: m _ 18 ¼ m

ð592Þ

_ p¼m _ 19 h19 Energy: m _ 18 h18 þ W

ð593Þ

_ 19 s19 Entropy: m _ 18 s18 þ S_ gen;p ¼ m

ð594Þ

_ p ¼m _ D;p _ 19 ex 19 þ Ex Exergy: m _ 18 ex 18 þ W

ð595Þ

HEX: under the steady state and steady flow conditions, the balance equations for HEX subcomponent are defined as follows:

4.10.6.4

_ 3; m _ 15 Mass: m _2¼m _ 14 ¼ m

ð596Þ

_ 14 h14 ¼ m _ 3 h3 þ m _ 15 h15 Energy: m _ 2 h2 þ m

ð597Þ

_ 14 s14 þ S_ gen;HEX ¼ m _ 3 s3 þ m _ 15 s15 Entropy: m _ 2 s2 þ m

ð598Þ

_ D;HEX _ 14 ex 14 ¼ m _ 3 ex3 þ m _ 15 ex15 þ Ex Exergy: m _ 2 ex 2 þ m

ð599Þ

Heavy Oil-Based Integrated Gasification Combined Cycle

The heavy oil residues gasification processes have been shown to be a cleaner alternative way to conventional processing techniques of fossil fuels. The heavy oil gasifier should be utilized to process refinery waste, avoiding waste disposal costs and improving the yields from crude oil distillation system. The IGCC based on heavy oil residues can produce power and steam needed by a refinery. The schematic diagram of a heavy oil-based IGCC system is illustrated in Fig. 33. Air at ambient temperature

424

Integrated Gasification Combined Cycles

14 CO2

15 Syngas storage

CO2 removal 13

12 H2S

Sulfur removal 16 11

10 Steam

WGSR 9

8 Ash

Ash removal

17 18

Nitrogen storage

20

Combustion chamber

Gas turbine

21

7 6 Heavy oil

Power

19 Heavy oil gasifier

Air compressor-II

5 Waste ASU 3 2 Air compressor-I

27

22 HRSG

4

Turbine Power

23

26 Steam

HEX 25 2

28 Heating application 32

30

24 Stack

1

29 Pump

Condenser 31

Fig. 33 Integrated heavy oil gasification combined cycle. ASU, air separation unit; HEX, heat exchanger; HRSG, heat recovery steam generator; WGSR, water-gas shift reactor.

and pressure enters an air compressor at point 1 and exits after compression at point 2. The pressured air enters the ASU, and separated hot nitrogen exits at point 4 and can be stored in a nitrogen storage subcomponent. The hot air at point 5 and steam coming from HEX at point 26 enter a heavy oil gasifier into which heavy oil is injected, and hot combustion gases exit at point 7 and enter an ash removal subcomponent. The hot gas at point 9 and steam at point 10 enter a WGSR for a higher hydrogen rich syngas yield. Thereafter, the sulfur must be removed from syngas using by sulfur removal subcomponent. The hydrogen rich syngas enters a carbon dioxide capture subcomponents for removing carbon dioxide, and can be stored in syngas storage subcomponent for continuously electricity production and heat supply. The compressed air at point 20, nitrogen at point 17 and syngas at point 16 enter a combustion chamber, and hot combustion gas exits at point 21 and passes through the gas turbine to produce power. Hot stack gas enters a HRSG to generate HP steam at point 27. The HP steam enters a steam turbine to generate shaft power. The low pressure working fluid has adequate energy for utilize in heating applications. The mass, energy, entropy, and exergy balance equations of integrated heavy oil gasification-based combined cycle components are defined in subsections. Air compressor-I: under steady state and steady flow conditions, the balance equations for air compressor-I can be written as _2 Mass: m _1 ¼m

ð600Þ

_ ac_I ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð601Þ

_ 2 s2 Entropy:m _ 1 s1 þ S_ gen;ac_I ¼ m

ð602Þ

_ ac_I ¼ m _ D;ac_I _ 2 ex2 þ Ex Exergy: m _ 1 ex 1 þ W

ð603Þ

ASU: under steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for ASU subcomponent are defined as follows: _3þm _4þm _5 Mass: m _2 ¼m

ð604Þ

_ 3 h3 þ m _ 4 h4 þ m _ 5 h5 Energy: m _ 2 h2 ¼ m

ð605Þ

Integrated Gasification Combined Cycles

425

_ 3 s3 þ m _ 4 s4 þ m _ 5 s5 Entropy: m _ 2 s2 þ S_ gen;ASU ¼ m

ð606Þ

_ D;ASU _ 3 ex3 þ m _ 4 ex 4 þ m _ 5 ex 5 þ Ex Exergy: m _ 2 ex 2 ¼ m

ð607Þ

Heavy oil gasifier: under the steady state and steady flow conditions, the balance equations for heavy oil gasifier can be written as _6þm _ 26 ¼ m _7 Mass: m _5þm

ð608Þ

_ 6 h6 þ m _ 26 h26 ¼ m _ 7 h7 Energy: m _ 5 h5 þ m

ð609Þ

_ 6 s6 þ m _ 26 s26 þ S_ gen;hog ¼ m _ 7 s7 Entropy: m _ 5 s5 þ m

ð610Þ

_ D;hog _ 6 ex 6 þ m _ 26 ex 26 ¼ m _ 7 ex7 þ Ex Exergy: m _ 5 ex5 þ m

ð611Þ

Ash removal: under the steady state and steady flow conditions, the balance equations for ash removal subcomponent can be given as follows: _8þm _9 Mass: m _7¼m

ð612Þ

_ 8 h8 þ m _ 9 h9 Energy: m _ 7 h7 ¼ m

ð613Þ

_ 8 s8 þ m _ 9 s9 Entropy: m _ 7 s7 þ S_ gen;ar ¼ m

ð614Þ

_ D;ar _ 8 ex8 þ m _ 9 ex 9 þ Ex Exergy: m _ 7 ex 7 ¼ m

ð615Þ

WGSR: under the steady state and steady flow conditions, the balance equations for WGSR are written as _ 10 ¼ m _ 11 Mass: m _9þm

ð616Þ

_ 10 h10 ¼ m _ 11 h11 Energy: m _ 9 h9 þ m

ð617Þ

_ 10 s10 þ S_ gen;WGSR ¼ m _ 11 s11 Entropy: m _ 9 s9 þ m

ð618Þ

_ D;WGSR _ 10 ex10 ¼ m _ 11 ex 11 þ Ex Exergy: m _ 9 ex9 þ m

ð619Þ

Sulfur removal: under the steady state and steady flow conditions, the balance equations for sulfur removal can be defined as follows: _ 12 þ m _ 13 Mass: m _ 11 ¼ m

ð620Þ

_ 12 h12 þ m _ 13 h13 Energy: m _ 11 h11 ¼ m

ð621Þ

_ 12 s12 þ m _ 13 s13 Entropy: m _ 11 s11 þ S_ gen;sr ¼ m

ð622Þ

_ D;sr _ 12 ex12 þ m _ 13 ex13 þ Ex Exergy: m _ 11 ex11 ¼ m

ð623Þ

CO2 removal: under the steady state and steady flow conditions, the balance equations for CO2 removal can be defined as follows: _ 14 þ m _ 15 Mass: m _ 13 ¼ m

ð624Þ

_ 14 h14 þ m _ 15 h15 Energy: m _ 13 h13 ¼ m

ð625Þ

Entropy: m _ 13 s13 þ S_ gen;CO2

rmv

_ 14 s14 þ m _ 15 s15 ¼m

ð626Þ

426

Integrated Gasification Combined Cycles

_ D;CO2 _ 14 ex14 þ m _ 15 ex15 þ Ex Exergy: m _ 13 ex13 ¼ m

rmv

ð627Þ

Syngas storage: under the steady state and steady flow conditions, the balance equations for syngas storage can be written as _ 16 Mass: m _ 15 ¼ m

ð628Þ

_ 16 h16 Energy: m _ 15 h15 ¼ m

ð629Þ

_ 16 s16 Entropy: m _ 15 s15 þ S_ gen;ss ¼ m

ð630Þ

_ D;ss _ 16 ex16 þ Ex Exergy: m _ 15 ex 15 ¼ m

ð631Þ

Nitrogen storage: the balance equations for nitrogen storage can be written under steady state and steady flow conditions as _ 17 Mass: m _4¼m

ð632Þ

_ 17 h17 Energy: m _ 4 h4 ¼ m

ð633Þ

_ 17 s17 Entropy: m _ 4 s4 þ S_ gen;ns ¼ m

ð634Þ

_ D;ns _ 17 ex 17 þ Ex Exergy: m _ 4 ex 4 ¼ m

ð635Þ

Air compressor-II: under steady state and steady flow conditions, the balance equations for air compressor-II are written as _ 20 Mass: m _ 19 ¼ m

ð636Þ

_ ac_II ¼ m _ 20 h20 Energy: m _ 19 h19 þ W

ð637Þ

_ 20 s20 Entropy: m _ 19 s19 þ S_ gen;ac_II ¼ m

ð638Þ

_ ac_II ¼ m _ D;ac_II _ 20 ex 20 þ Ex Exergy: m _ 19 ex19 þ W

ð639Þ

Combustion chamber: under the steady state and steady flow conditions, the balance equations for combustion chamber can be defined as follows: _ 20 ¼ m _ 21 Mass: m _ 18 þ m

ð640Þ

_ 20 h20 ¼ m _ 21 h21 Energy: m _ 18 h18 þ m

ð641Þ

_ 20 s20 þ S_ gen;cc ¼ m _ 21 s21 Entropy: m _ 18 s18 þ m

ð642Þ

_ D;cc _ 20 ex 20 ¼ m _ 21 ex21 þ Ex Exergy: m _ 18 ex18 þ m

ð643Þ

Gas turbine: under steady state and steady flow conditions, the balance equations for gas turbine can be defined as _ 22 Mass: m _ 21 ¼ m

ð644Þ

_ gt _ 22 h22 þ W Energy: m _ 21 h21 ¼ m

ð645Þ

_ 22 s22 Entropy: m _ 21 s21 þ S_ gen;gt ¼ m

ð646Þ

_ gt þ Ex _ D;gt _ 22 ex22 þ W Exergy: m _ 21 ex 21 ¼ m

ð647Þ

Integrated Gasification Combined Cycles

427

HRSG: under the steady state and steady flow conditions, the balance equations for HRSG can be defined as follows: _ 23 ; m _ 30 Mass: m _ 22 ¼ m _ 27 ¼ m

ð648Þ

_ 30 h30 ¼ m _ 23 h23 þ m _ 27 h27 Energy: m _ 22 h22 þ m

ð649Þ

_ 30 s30 þ S_ gen;HRSG ¼ m _ 23 s23 þ m _ 27 s27 Entropy: m _ 22 s22 þ m

ð650Þ

_ D;HRSG _ 30 ex30 ¼ m _ 23 ex 23 þ m _ 27 ex27 þ Ex Exergy: m _ 22 ex 22 þ m

ð651Þ

Steam turbine: under steady state and steady flow conditions, the balance equations for steam turbine can be written as _ 28 Mass: m _ 27 ¼ m

ð652Þ

_ st _ 28 h28 þ W Energy: m _ 27 h27 ¼ m

ð653Þ

_ 28 s28 Entropy: m _ 27 s27 þ S_ gen;st ¼ m

ð654Þ

_ st þ Ex _ D;st _ 28 ex 28 þ W Exergy: m _ 27 ex 27 ¼ m

ð655Þ

Condenser: under the steady state and steady flow conditions, the balance equations for condenser subcomponent are written as follows: _ 29 ; m _ 32 Mass: m _ 28 ¼ m _ 31 ¼ m

ð656Þ

_ 31 h31 ¼ m _ 29 h29 þ m _ 32 h32 Energy: m _ 28 h28 þ m

ð657Þ

_ 31 s31 þ S_ gen;con ¼ m _ 29 s29 þ m _ 32 s32 Entropy: m _ 28 s28 þ m

ð658Þ

_ D;con _ 31 ex 31 ¼ m _ 29 ex29 þ m _ 32 ex 32 þ Ex Exergy: m _ 28 ex28 þ m

ð659Þ

Pump: for pump subcomponent, the balance equations are provided under the steady state and steady flow conditions. _ 30 Mass: m _ 29 ¼ m

ð660Þ

_ p¼m _ 30 h30 Energy: m _ 29 h29 þ W

ð661Þ

_ 30 s30 Entropy: m _ 29 s29 þ S_ gen;p ¼ m

ð662Þ

_ p ¼m _ D;p _ 30 ex 30 þ Ex Exergy: m _ 29 ex 29 þ W

ð663Þ

HEX: under the steady state and steady flow conditions, the balance equations for HEX are defined as follows:

4.10.7

_ 24 ; m _ 26 Mass: m _ 23 ¼ m _ 25 ¼ m

ð664Þ

_ 25 h25 ¼ m _ 24 h24 þ m _ 26 h26 Energy: m _ 23 h23 þ m

ð665Þ

_ 25 s25 þ S_ gen;HEX ¼ m _ 24 s24 þ m _ 26 s26 Entropy: m _ 23 s23 þ m

ð666Þ

_ D;HEX _ 25 ex 25 ¼ m _ 24 ex 24 þ m _ 26 ex 26 þ Ex Exergy: m _ 23 ex23 þ m

ð667Þ

Case Studies

In this section, to investigate thermodynamic assessment of different IGCCs, the multigeneration systems are described and their schematic representations are given. Also, the thermodynamic modeling of the IGCCs and components for multigeneration purposes are illustrated.

Integrated Gasification Combined Cycles

428 4.10.7.1

Solar Energy Combined Biomass Gasification System

Biomass fuel and solar energy are the alternative energy resources that can be integrated for multigeneration aim. Nowadays, the potential investigation that resources the applicability of these two alternative energy resources is on ongoing. For this reason, multigeneration process based on biomass gasification and solar parabolic dish collector is selected in this case study. The schematic diagram of solar energy-based IGCC for power, heating, cooling and fresh water production is illustrated in Fig. 34. This integrated system consists of mainly six subsystems, such as (1) parabolic dich collector system, (2) biomass gasifier, (3) gas turbine process, (4) Rankine cycle, (5) single effect absorption cooling system, and (6) fresh water production cycle. As illustrated in Fig. 34, the biomass source enters the gasification chamber at state 4 and hot air coming from parabolic dish collector enters the gasifier at state 3. The biomass gasification subsystem consists of a gasifier and syngas storage for continuously power, heating, cooling and fresh water generation. After that, the syngas is burned in the combustion chamber to supply heat for the Brayton cycle for power generation. Air at reference status enters the air compressor subcomponent at state 7 and exits after compression by using compressor at state 8. The hot air enters the combustion chamber into which syngas is injected, and hot exhaust gas exit at state 9 and pass through the gas turbine to generate shaft power. The hot exhaust gas expands in the gas turbine to state 10. The hot gas exiting from the gas pressure turbine is transferred to the integrated Rankine cycle to make use of this heat to produce power and hot water.

Solar radiation Ash

2

3

Air fan 1

5

Gasification chamber

6

4 Biomass

Air

Syngas storage

8

Combustion chamber

Air compressor

9 Power

Gas turbine

7 10 12

14

HEX-I

HEX-III 13

15

17

38 39

16 Pump-III 37

Saline 36 water

41

Mineralize

26

Generator 23

22

Drain

Condenser-I

Heating 18

Membrane distillation module

40

19

Pump -I 11

Exhaust gas

Power

Steam turbine

Condenser-II 27

HEX-II 42

24 Expansion valve-II 20 Pump-II 25

Fresh water tank

31 30

Expansion valve-I

21

34

Absorber

29

28 32 Evaporator

35 Fig. 34 Schematic diagram of solar energy-based integrated gasification combined cycle (IGCC). HEX, heat exchanger.

District cooling 33

Integrated Gasification Combined Cycles

429

The flow conditions in the Rankine cycle according to Fig. 34 can be defined as follows. The working fluid at point 16 exits from the condenser-I as saturated liquid. Then, the pump-I increases the pressure of the saturated liquid at point 17. Next, the working fluid goes the evaporator in a liquid state and exits as vapor at point 14. Then, the working fluid expands through the Rankine turbine to generate electricity. Next, the working fluid exits the Rankine turbine at point 15 and provides heat energy to the heating process in the condenser-I. Then, the working fluid enters the condenser-I as saturated vapor. The generator removes heat energy to supply the heating load for the residential application. Next, the working fluid exits from the condenser-I again as saturated liquid at point 16. To produce cooling effect, the exhaust gas coming from combustion chamber at point 11 enters the generator subcomponent of single effect absorption chiller. This flow transports between the system parts of this cooling process as either H2O or the mixture of LiBr–H2O. As an output of the input heat energy into the generator subcomponent, H2O evaporates from the mixture of LiBr–H2O and goes to the condenser-II at point 26. The heat energy is removed in the condenser-II subcomponent. Hence, the water rejected heat and quits the condenser-II as saturated liquid at point 27. Then, the H2O is throttled before entering the evaporator subcomponent at low temperature level at point 28. The evaporator subcomponent gives the cooling effect for residential application. Then, H2O exits from the evaporator subcomponent and goes to the absorber at point 29. Then, the H2O mixes with the mixture of LiBr–H2O. The LiBr–H2O mixture exits from the absorber at point 20 and is pumped to the HEX-I in single effect cooling system at point 21. After that, the H2O–LiBr mixture exits from the HEX-I and goes to the generator at point 22. Then, the LiBr–H2O mixture is heated in the generator and part of the H2O in the mixture evaporates and exits from this subcomponent at point 26. As an output of the H2O vaporization, the LiBr–H2O mixture exits from the generator with a higher LiBr concentration in the mixture to enter the HEX-I at point 23 to gain heat energy. Next, this mixture exits from the HEX-I at point 24 and is throttled into the absorber subcomponent at point 25. This information in this page summarizes the working fluids cycle in the single effect absorption cooling process. The point definitions of single effect absorption cooling system are given in Table 9. The schematic diagram of the biomass and solar energy-based distillation process to produce fresh water from saline water is given in the left side of Fig. 34. The desalination proses consists of the pump that pressurizes the saline water, the filter to remove the coarse-grained particles, the energy recovery subcomponent, the osmotic membrane to filter thin-grained particles and salt and the fresh water tank to store the domestic water. With the utilization of the osmotic membrane, the conversion to domestic water can be provided with better quality. As biomass sample, wheat straw is selected due to its availability and low moisture content. From the previous experience, it can be foreseen that high moisture content will result in stuck of screw-feeder. Small particles are also more feasible than large fiber-type biomass to maintain a continuous flow without wind-around. The operating parameters of gasification chamber are shown in Table 10. Also, the ultimate analysis and calorific values are illustrated in Tables 11 and 12, respectively. The composition product syngas is given in Table 13.

4.10.7.1.1

Balance equations for system components

In this subsection, the balance equations for components of solar energy-based IGCC are defined for thermodynamic analysis. Table 9 definition

Single effect absorption cooling (SEAC) system with point

Points of SEAC system

Definitions

20 21 22 23 24 25 26 27 28 29

Saturated liquid solution Subcooled liquid solution Subcooled liquid solution Saturated liquid solution Subcooled liquid solution Vapor–liquid solution Superheated steam Saturated liquid water Vapor–liquid water Saturated vapor

Table 10

Biomass gasification chamber operating conditions

Operating conditions

Value

Gasifier temperature (1C) Gasifier pressure (kPa) Steam/biomass ratio Sand/biomass ratio

826 0.17 0.45 0.03

430

Integrated Gasification Combined Cycles

Table 11 Compound

Carbon Hydrogen Nitrogen Sulfur Oxygen

Table 12

Ultimate analysis of biomass sample employed in biomass gasifier Straw Dry basis (wt%)

Dry ash-free basis (wt%)

44.86 3.82 0.73 0.11 44.59

47.66 4.06 0.78 0.12 47.37

Lower and higher heating values (HHVs) of biomass samples in the biomass gasifier subcomponent

Heating value

Straw

Lower heating value (LHV) Higher heating value (HHV)

Original basis (kJ/kg)

Dry basis (kJ/kg)

16,525 17,785

18,020 19,225

Table 13 Product syngas composition Component

% mol

Hydrogen Carbon monoxide Carbon dioxide Methane Acetylene Ethylene Ethane Tars H2 S NH3

21.28 43.16 13.45 15.83 0.36 4.62 0.62 0.40 0.08 0.37

Air fan subcomponent: the mass, energy, entropy, and exergy balance equalities can be written for air fan subcomponent under the steady state and steady flow situations as follows: _2 Mass: m _1 ¼m

ð668Þ

_ af ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð669Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;af ¼ m

ð670Þ

_ af ¼ m _ D;af _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð671Þ

Parabolic collector subcomponent: the mass, energy, entropy, and exergy balance equalities can be defined for parabolic dish collector subcomponent under the steady state and steady flow situations as follows: _3 Mass: m _2 ¼m

ð672Þ

_ solar ¼ m _ 3 h3 Energy: m _ 2 h2 þ Q

ð673Þ

_ solar =Tsc þ S_ gen;sc ¼ m _ 3 s3 Entropy: m _ 2 s2 þ Q

ð674Þ

Integrated Gasification Combined Cycles

_ solar ð1 Exergy: m _ 2 ex2 þ Q

_ D;sc _ 3 ex 3 þ Ex T0 =Tsc Þ ¼ m

431

ð675Þ

Gasification chamber subcomponent: IGCCS are the processes that usually includes chemical reactions in presence of feedstock and oxygen at high temperatures enough to break the chemical bonds of the molecules. At the end of the gasification reaction, this chemical bond energy comes out as heat. For this reason, chemical energy and exergy of the feedstock and the resultant syngas should be taken into account instead of thermo-mechanical energy and exergy. The mass, energy, entropy, and exergy balance equations of gasification chamber under steady state and steady flow conditions are written in Eqs. (676)–(679). _4¼m _5 Mass: m _3þm

ð676Þ

_ 4 h4 ¼ m _ 5 h5 Energy: m _ 3 h3 þ m

ð677Þ

_ 4 s4 þ S_ gen;gas ¼ m _ 5 s5 Entropy: m _ 3 s3 þ m

ð678Þ

_ _ 4 exch _ 5 ex ch Exergy: m _ 3 ex 3 þ m 4 ¼m 5 þ Ex D;Das

ð679Þ

In the following equalities, the chemical exergy accounts for syngas are defined accordingly: ch ch ch ex ch 5 ¼ rH2 MH2 ex H2 þ rCO MCO ex CO þ rCH4 MCH4 ex CH4

 0  0 0 ex ch CO ¼ hCO hC þ 0:5hO2

 T0 s0CO



s0C þ 0:5s0O2



ch þ ex ch C þ 0:5ex O2

ð680Þ

ð681Þ

 0  0 0 ex ch CO2 ¼ hCO2 hC þ hO2

 T0 s0CO

  ch þ ex ch s0C þ s0O2 C þ ex O2

ð682Þ

 0  0 0 ex ch CH4 ¼ hCH4 hC þ 2hH2

 T0 s0CO



ð683Þ

s0C þ 2s0H2



ch þ ex ch C þ 2ex H2

where M is the molar mass (kg/kmol) of the corresponding substance, h0 is the enthalpy (kJ/kmol), s0 is the entropy (kJ/kmol K), exch is the chemical exergy (kJ/kmol) of the corresponding substance at 251C and 101.3 kPa, the bar accent represents molar basis, r is the mass concentration (wt%) of the substance. The point 4 is considered as a biomass feedstock, moisture and ash to model gasification in a molar basis. At point 5, the composition of the syngas that consists of hydrogen, carbon monoxide, carbon dioxide, nitrogen, methane, ash and trace amounts of sulfur compounds exit from the gasification chamber. Syngas storage subcomponent: the mass, energy, entropy, and exergy balance equations can be written for syngas storage subcomponent under the steady state and steady flow situations as follows: _6 Mass: m _5 ¼m

ð684Þ

_ ss ¼ m _ 6 h6 Energy: m _ 5 h5 þ Q

ð685Þ

_ ss =Tss þ S_ gen;ss ¼ m _ 6 s6 Entropy: m _ 5 s5 þ Q

ð686Þ

_ Q ¼m _ D;ss _ 6 ex 6 þ Ex Exergy: m _ 5 ex5 þ Ex ss

ð687Þ

Air compressor subcomponent: the mass, energy, entropy, and exergy balance equalities can be defined for the air compressor subcomponent under the steady state and steady flow conditions as follows: _8 Mass: m _7 ¼m

ð688Þ

_ ac ¼ m _ 8 h8 Energy: m _ 7 h7 þ W

ð689Þ

_ 8 s8 Entropy: m _ 7 s7 þ S_ gen;ac ¼ m

ð690Þ

_ ac ¼ m _ D;ac _ 8 ex 8 þ Ex Exergy: m _ 7 ex 7 þ W

ð691Þ

432

Integrated Gasification Combined Cycles

Combustion chamber subcomponent: in the combustion chamber, the syngas coming from storage subcomponent is combusted in presence of compressed air illustrated as point 8. To simplify the accounts, point 9 is assumed to be hot air, whose temperature is increased by the heat of combustion occurring in the biomass combustion chamber. This heat is the HHV of the syngas mixture. The balance equalities of biomass combustion chamber under steady state and steady flow conditions are defined in Eqs. (692)–(695). _8¼m _9 Mass: m _6þm

ð692Þ

_ 8 h8 ¼ m _ 9 h9 Energy: m _ 6 h6 þ m

ð693Þ

_ 8 s8 þ S_ gen;cc ¼ m _ 9 s9 Entropy: m _ 6 s6 þ m

ð694Þ

_ D;cc _ 8 ex8 ¼ m _ 9 ex9 þ Ex Exergy: m _ 6 exch 6 þm

ð695Þ

Moreover, in the chemical exergy calculations part, the chemical exergy content of syngas can be defined through the Eqs. (696)–(698). ch ch ch ex ch 6 ¼ rH2 MH2 ex H2 þ rCO MCO ex CO þ rCH4 MCH4 ex CH4

ð696Þ

 0  0 0 ex ch CO2 ¼ hCO2 hC þ hO2

 T0 s0CO

  ch þ ex ch s0C þ s0O2 C þ ex O2

ð697Þ

 0  0 0 ex ch CH4 ¼ hCH4 hC þ 2hH2

 T0 s0CO



ð698Þ

s0C þ 2s0H2



ch þ ex ch C þ 2ex H2

where M is the molar mass (kg/kmol) of the corresponding substance, h0 is the enthalpy (kJ/kmol), s0 is the entropy (kJ/kmol K), exch is the chemical exergy (kJ/kmol) of the corresponding substance at 251C and 101.3 kPa, the bar accent represents molar basis, r is the mass concentration (wt%) of the substance. The syngas coming from point 10 enters the biomass combustion chamber along with air coming from the HP compressor at point 8 as illustrated in Fig. 34. The amount of oxygen is calculated by chemical reaction balance and it is assumed that 20% more air is provided for better combustion relative to the stoichiometric amount. The efficiency rate of biomass combustion subcomponent is taken as 90%. Gas turbine subcomponent: under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for gas turbine subcomponent can be written as follows: _ 10 Mass: m _9¼m

ð699Þ

_ GT;net _ 10 h10 þ W Energy: m _ 9 h9 ¼ m

ð700Þ

_ 10 s10 Entropy: m _ 9 s9 þ S_ gen;GT ¼ m

ð701Þ

_ GT;net þ Ex _ D;GT _ 10 ex 10 þ W Exergy: m _ 9 ex9 ¼ m

ð702Þ

The gas turbine exit temperature (T10) should be descried as the function of gas turbine isentropic efficiency (ZGT), gas turbine inlet temperature (T9), and inlet and outlet pressure ratio of gas turbine (P9/P10) as given below: 0 1    1 y g g g P 9 A ð703Þ T10 ¼ T9 @1 ZGT 1 P10 The power output from gas turbine can be calculated as follows:

_ GT ¼ m _ 9 Cpg ðT9 W

T10 Þ

where Cpg can be calculated based on the temperature function as given below:       6:997T 2:712T 2 1:2244T 3 Cpg ðT Þ ¼ 0:991 þ þ 107 1010 105

ð704Þ

ð705Þ

The net power generation from gas turbine can be calculated as follows: _ GT _ net ¼ W W

_ AC W

ð706Þ

HEX-I subcomponent: the mass, energy, entropy, and exergy balance equations for HEX-I subcomponent can be defined under the steady state and steady flow conditions as follows:

Integrated Gasification Combined Cycles

433

_ 11 ; m _ 17 Mass: m _ 10 ¼ m _ 14 ¼ m

ð707Þ

_ 17 h17 ¼ m _ 11 h11 þ m _ 14 h14 Energy: m _ 10 h10 þ m

ð708Þ

_ 17 s17 þ S_ gen;HEX Entropy:m _ 10 s10 þ m

I

_ 11 s11 þ m _ 14 s14 ¼m

_ D;HEX _ 17 ex 17 ¼ m _ 11 ex 11 þ m _ 14 ex 14 þ Ex Exergy: m _ 10 ex 10 þ m

ð709Þ

I

ð710Þ

Steam turbine subcomponent: under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for steam turbine subcomponent can be written as follows: _ 15 Mass: m _ 14 ¼ m

ð711Þ

_ st _ 15 h15 þ W Energy: m _ 14 h14 ¼ m

ð712Þ

_ 15 s15 Entropy: m _ 14 s14 þ S_ gen;st ¼ m

ð713Þ

_ st þ Ex _ D;st _ 15 ex 15 þ W Exergy: m _ 14 ex 14 ¼ m

ð714Þ

Condenser-I subcomponent: the mass, energy, entropy, and exergy balance equations can be written for condenser-I under the steady state and steady flow conditions as follows: _ 16 ; m _ 19 Mass: m _ 15 ¼ m _ 18 ¼ m

ð715Þ

_ con_I _ 18 h18 ¼ m _ 16 h16 þ m _ 19 h19 þ Q Energy: m _ 15 h15 þ m

ð716Þ

_ 18 s18 þ S_ gen;con_I ¼ m _ 16 s16 þ m _ 19 s19 Entropy: m _ 15 s15 þ m

ð717Þ

_ D;con_I _ 18 ex 18 ¼ m _ 16 ex16 þ m _ 19 ex 19 þ Ex Exergy: m _ 15 ex15 þ m

ð718Þ

Pump-I subcomponent: the mass, energy, entropy, and exergy balance equations for pump-I subcomponent can be written under the steady state and steady flow conditions as given below: _7 Mass: m _6 ¼m

ð719Þ

_ p¼m _ 7 h7 Energy: m _ 6 h6 þ W

ð720Þ

_ 7 s7 Entropy: m _ 6 s6 þ S_ gen;p ¼ m

ð721Þ

_ p ¼m _ D;p _ 7 ex 7 þ Ex Exergy: m _ 6 ex 6 þ W

ð722Þ

Generator subcomponent: the mass, energy, entropy, and exergy balance equalities can be written for generator subcomponent under the steady state and steady flow situations as follows: _ 12 ; m _ 23 þ m _ 26 Mass: m _ 11 ¼ m _ 22 ¼ m

ð723Þ

_ gena ¼ m _ 22 h222 þ Q _ 12 h12 þ m _ 23 h23 þ m _ 26 h26 Energy: m _ 11 h11 þ m

ð724Þ

_ gena =Tgena þ S_ gen;gena ¼ m _ 22 s22 þ Q _ 12 s12 þ m _ 23 s23 þ m _ 26 s26 Entropy: m _ 11 s11 þ m

ð725Þ

_ Q ¼m _ D;gena _ 22 ex 22 þ Ex _ 12 ex 12 þ m _ 23 ex 23 þ m _ 26 ex 26 þ Ex Exergy: m _ 11 ex 11 þ m gena

ð726Þ

Condenser-II subcomponent: the mass, energy, entropy, and exergy balance equalities can be defined for condenser-II under the steady state and steady flow conditions as follows: _ 27 ; m _ 31 Mass: m _ 26 ¼ m _ 30 ¼ m

ð727Þ

_ con_II _ 30 h30 ¼ m _ 27 h27 þ m _ 31 h31 þ Q Energy: m _ 26 h26 þ m

ð728Þ

434

Integrated Gasification Combined Cycles

_ con_II =Tcon_II _ 30 s30 þ S_ gen;con_II ¼ m _ 27 s27 þ m _ 31 s31 þ Q Entropy: m _ 26 s26 þ m

ð729Þ

_ Q _ _ 30 ex 30 ¼ m _ 27 ex27 þ m _ 31 ex 31 þ Ex Exergy: m _ 26 ex26 þ m con_II þ Ex D;con_II

ð730Þ

Expansion valve-I subcomponent: the mass, energy, entropy, and exergy balance equations for expansion valve-I subcomponent are written under the steady state and steady flow conditions as follows: _ 28 Mass: m _ 27 ¼ m

ð731Þ

_ 28 h28 Energy: m _ 27 h27 ¼ m

ð732Þ

_ 28 s28 Entropy: m _ 27 s27 þ S_ gen;ev_I ¼ m

ð733Þ

_ D;ev_I _ 28 ex28 þ Ex Exergy: m _ 27 ex27 ¼ m

ð734Þ

Evaporator subcomponent: the mass, energy, entropy, and exergy balance equalities can be defined for evaporator subcomponent under the steady state and steady flow conditions as follows: _ 29 ; m _ 33 Mass: m _ 28 ¼ m _ 32 ¼ m

ð735Þ

_ eva ¼ m _ 32 h32 þ Q _ 29 h29 þ m _ 33 h33 Energy: m _ 28 h28 þ m

ð736Þ

_ eva =Teva þ S_ gen;eva ¼ m _ 32 s32 þ Q _ 29 s29 þ m _ 33 s33 Entropy: m _ 28 s28 þ m

ð737Þ

_ Q ¼m _ D;eva _ 32 ex32 þ Ex _ 29 ex29 þ m _ 33 ex33 þ Ex Exergy: m _ 28 ex28 þ m eva

ð738Þ

Absorber subcomponent: the mass, energy, entropy, and exergy balance equalities are written for absorber subcomponent under the steady state and steady flow conditions as follows: _ 25 þ m _ 29 ; m _ 35 _ 34 ¼ m Mass: m _ 20 ¼ m

ð739Þ

_ abs _ 29 h29 þ m _ 34 h34 ¼ m _ 20 h20 þ m _ 35 h35 þ Q Energy: m _ 25 h25 þ m

ð740Þ

_ abs =Tabs _ 29 s29 þ m _ 34 s34 þ S_ gen;abs ¼ m _ 20 s20 þ m _ 35 s35 þ Q Entropy: m _ 25 s25 þ m

ð741Þ

_ Q þ Ex _ D;abs _ 29 ex 29 þ m _ 34 ex34 ¼ m _ 20 ex 20 þ m _ 35 ex35 þ Ex Exergy: m _ 25 ex 25 þ m abs

ð742Þ

Pump-II subcomponent: for the pump-II subcomponent of single effect absorption cooling system, the balance equations are provided under the steady state and steady flow conditions as follows: _ 21 Mass: m _ 20 ¼ m

ð743Þ

_ p_II ¼ m _ 21 h21 Energy: m _ 20 h20 þ W

ð744Þ

_ 21 s21 Entropy: m _ 20 s20 þ S_ gen;p_II ¼ m

ð745Þ

_ p_II ¼ m _ D;p_II _ 21 ex 21 þ Ex Exergy: m _ 20 ex20 þ W

ð746Þ

Expansion valve-II subcomponent: the mass, energy, entropy, and exergy balance equations for expansion valve-II subcomponent can be defined under the steady state and steady flow conditions as follows: _ 25 Mass: m _ 24 ¼ m

ð747Þ

_ 25 h25 Energy: m _ 24 h24 ¼ m

ð748Þ

_ 25 s25 Entropy: m _ 24 s24 þ S_ gen;ev_II ¼ m

ð749Þ

Integrated Gasification Combined Cycles

_ D;ev_II _ 25 ex25 þ Ex Exergy: m _ 24 ex24 ¼ m

435

ð750Þ

HEX-II subcomponent: the mass, energy, entropy, and exergy balance equations for HEX-II subcomponent can be defined under the steady state and steady flow conditions as follows: _ 22 ; m _ 24 Mass: m _ 21 ¼ m _ 23 ¼ m _ HEX _ 23 h23 þ Q Energy: m _ 21 h21 þ m _ HEX I =THEX _ 23 s23 þ Q Entropy: m _ 21 s21 þ m _ Q _ 23 ex 23 þ Ex Exergy: m _ 21 ex21 þ m HEX

I

I

I

ð751Þ

_ 22 h22 þ m _ 24 h24 ¼m

þ S_ gen;HEX

I

ð752Þ

_ 22 s22 þ m _ 24 s24 ¼m

_ D;HEX _ 22 ex22 þ m _ 24 ex24 þ Ex ¼m

I

ð753Þ ð754Þ

HEX-III subcomponent: the mass, energy, entropy, and exergy balance equations for HEX-III subcomponent can be defined under the steady state and steady flow conditions as follows: _ 13 ; m _ 39 _ 38 ¼ m Mass: m _ 12 ¼ m

ð755Þ

_ 38 h38 ¼ m _ 13 h13 þ m _ 39 h39 Energy: m _ 12 h12 þ m

ð756Þ

_ 38 s38 þ S_ gen;HEX Entropy: m _ 12 s12 þ m

II

_ 13 s13 þ m _ 39 s39 ¼m

_ D;HEX _ 38 ex 38 ¼ m _ 13 ex 13 þ m _ 39 ex 39 þ Ex Exergy: m _ 12 ex 12 þ m

ð757Þ II

ð758Þ

Pump-III subcomponent: the mass, energy, entropy, and exergy balance equations for pump-III subcomponent can be defined under the steady state and steady flow conditions as follows: _ 37 Mass: m _ 36 ¼ m _p Energy: m _ 36 h36 þ W

III

Entropy: m _ 36 s36 þ S_ gen;p _p Exergy: m _ 36 ex36 þ W

III

ð759Þ

_ 37 h37 ¼m

III

ð760Þ

_ 37 s37 ¼m

_ D;p _ 37 ex 37 þ Ex ¼m

ð761Þ

III

ð762Þ

Membrane distillation module: the mass, energy, entropy, and exergy balance equations for membrane distillation module can be written under the steady state and steady flow conditions as follows: _ 39 ¼ m _ 38 þ m _ 40 þ m _ 41 Mass: m _ 37 þ m

ð763Þ

_ 39 h39 ¼ m _ 38 h38 þ m _ 40 h40 þ m _ 41 h41 Energy: m _ 37 h37 þ m

ð764Þ

_ 39 s39 þ S_ gen;mdm ¼ m _ 38 s38 þ m _ 40 s40 þ m _ 41 h41 Entropy: m _ 37 s37 þ m

ð765Þ

_ D;mdm _ 39 ex39 ¼ m _ 38 ex38 þ m _ 40 ex40 þ m _ 41 ex 41 þ Ex Exergy: m _ 37 ex 37 þ m

ð766Þ

Mineralizer subcomponent: the mass, energy, entropy, and exergy balance equations for mineralizer subcomponent can be defined under the steady state and steady flow conditions as follows: _ 42 Mass: m _ 41 ¼ m

ð767Þ

_ 42 h42 Energy: m _ 41 h41 ¼ m

ð768Þ

_ 42 s42 Entropy: m _ 41 s41 þ S_ gen;mine ¼ m

ð769Þ

_ D;mine _ 42 ex42 þ Ex Exergy: m _ 41 ex41 ¼ m

ð770Þ

436

Integrated Gasification Combined Cycles

4.10.7.1.2

Coefficient of performance

The energetic coefficient of performance (COPen) and exergetic coefficient of performance (COPex) of single effect absorption process can be written as follows, respectively:

4.10.7.1.3

COPen ¼

_ eva Q _ gena þ W _p Q

ð771Þ

COPex ¼

_ Q Ex eva _ExQ _ gena þ W p

ð772Þ

Parametric analysis

In order to better understand the integrated system performance, the parametric studies are given to investigate the effects of different indicators, such as reference temperature, solar radiation intensity, gas turbine inlet temperature, and gas turbine inlet pressure on the integrated system exergy destruction rate and exergy efficiency. Exergy analysis reveals the both exergetic efficiency and exergy destruction of subunits of an integrated system. Fig. 35 shows exergy destruction rates of subunits of the integrated system proposed in this study. According to Fig. 35, the largest exergy destruction rates occur in biomass gasifier, Brayton cycle, and solar collector with 33,900, 28,200, and 22,800 kW, respectively. These results indicate that those subunits should be improved in order to achieve higher exergy efficiencies than current ones. Fig. 36 shows that exergy destruction ratios of subunits of the integrated system. Similar to Fig. 35, biomass gasifier, Brayton cycle and solar collector have the highest exergy destruction ratios with 34.51%, 28.71%, and 23.21%, respectively. Exergy efficiency is another product of exergy analysis which reveals where the irreversibilities occur and which part should be improved. According to the analysis results absorption cooling system has the lowest exergy efficiency with 14.28% and exergy efficiency of whole system is 56.54% as seen from Fig. 37.

Exergy destruction rate (kW)

120,000 98,245

100,000 80,000 60,000 33,900

40,000 22,800

28,200

20,000

9312

3015

1018

0 Solar collector

Biomass gasifier

Brayton cycle

Rankine Absorption cycle cooling

Fresh Integrated water system production

Fig. 35 Exergy destruction rates for the solar energy-based integrated biomass gasification cycle.

Exergy destruction ratio (%)

40 34.51

35

28.71

30 25

23.21

20 15 9.48

10 5

3.07

1.04

0 Solar collector

Biomass gasifier

Brayton cycle

Rankine cycle

Absorption cooling

Fig. 36 Exergy destruction ratio for the solar energy-based integrated biomass gasification cycle.

Fresh water production

Integrated Gasification Combined Cycles

70

Exergy efficiency (%)

60

437

64.72 56.54 54.28 46.24

50

45.84

44.57

40 30 20

14.28

10 0 Solar collector

Biomass gasifier

Brayton cycle

Rankine cycle

Absorption Fresh water Integrated cooling production system

Fig. 37 Exergy efficiency for the subsystems of the solar energy-based integrated biomass gasification cycle.

110,000

0.6 ExD,system

0.59

System

106,000

0.58

104,000

0.57

102,000

0.56

100,000

0.55

98,000

0.54

96,000

0.53

94,000

0.52

92,000

0

5

10

15 20 25 Reference temperature (°C)

30

35

Exergy efficiency

Exergy destruction rate (kW)

108,000

0.51 40

Fig. 38 Effects of reference temperature on the exergy destruction rate and exergy efficiency of integrated system.

The effects of the varying ambient temperature from 0 to 401C on the exergy destruction rate and exergy efficiency of the integrated gasification combined system are illustrated in Fig. 38. As seen from Fig. 38, when the ambient temperature increases, the exergy destruction rate of the integrated system decreases from 108,245 to 92,135 kW, and exergy efficiency of the integrated gasification combined system increases from 51.54% to 59.47%, respectively. Because the solar radiation intensity varies during the solar daylight, the variations of the integrated system efficiency are analyzed. Fig. 39 demonstrates the variations of exergy destruction rate and exergy efficiency of the concentrating collector subsystem for different values of solar radiation intensity. As seen from Fig. 39, the exergy destruction rate and exergy efficiency of parabolic collector for the solar mode increase with increasing solar radiation intensity. This is because increasing the solar radiation intensity increases the outlet temperature of the working fluid for the parabolic collector subsystem. In the evaluation of absorption cooling systems, COF is mostly preferred. The COF is applied to the absorption cooling system both energetic and exergetic viewpoint. As seen from Fig. 40, as reference temperature rises from 5 to 401C, energetic COP almost remains the same, however, exergetic COP increases from about 0.3175 to 0.3875. Fig. 41 illustrates the variations with turbine inlet temperature of the exergy destruction rate and exergy efficiencies for the IGCCs. When the turbine inlet temperature increases while keeping other design parameters constant, the exergy destruction rates of solar energy-based IGCC decreases as the pinch point temperature is constant. The exergy efficiencies are observed to increase with increasing in turbine inlet temperature, because of the corresponding increase in network output and relatively smaller increase in heat addition to the cycle. The impact of gas turbine inlet pressure on the exergy destruction rate and exergy efficiency of integrated process is illustrated in Fig. 42. The exergy destruction rate of integrated system decreases with increasing gas turbine inlet pressure. As expected, the higher pressure of gas has positive effect on the system exergy efficiency. As seen from Fig. 42, while gas turbine inlet pressure changes

Integrated Gasification Combined Cycles

Exergy destruction rate (kW)

105,000

0.63

104,000

ExD,system

0.62

103,000

System

0.61

102,000

0.6

101,000

0.59

100,000

0.58

99,000

0.57

98,000

0.56

97,000

0.55

96,000

0.54

95,000

0.53

94,000 500

600

700 800 Ib (W m−2)

900

Exergy efficiency

438

0.52 1000

Fig. 39 Exergy destruction rate and exergy efficiency of the parabolic dish collector system for solar mode depending on the solar radiation intensity changes.

1.5

0.39 COPen

1.4

0.38

COPex

1.3

0.37 1.1

0.36

1

0.35

0.9

COPex

COPen

1,2

0.34

0.8 0.33 0.7 0.32

0.6 0.5

5

10

15 20 25 30 Reference temperature (°C)

35

0.31 40

Fig. 40 COPen and COPex of the absorption cooling system with respect to the reference environment temperature. COP, coefficient of performance.

from 1500 to 4000 kPa, the exergy efficiency of IGCC increases from about 54.8% to 57.5%. Because of an increment in the gas turbine inlet pressure of gas mixture, Brayton cycle produces more work in the gas turbines. To supply environmental comprehensions, the environmental effects of single generation, cogeneration, trigeneration, and multigeneration of IGCC are compared in Fig. 43. It is observed that the multigeneration process has less carbon dioxide emissions than the single generation, cogeneration, and trigeneration processes, supplying an important motivation for the utilization of multigeneration processes. It is also seen that the multigeneration process has the higher exergetic performance than the other generation processes. Also, the multigeneration cycle has less carbon monoxide emissions than the other generation processes, providing another motivation for the utilization of multigeneration processes. However, the amount of carbon monoxide emission is significantly less than that of the amount of carbon dioxide emissions of the IGCC.

4.10.7.2

Integrated Gasification Combined System With Biomass Gasification and Solid Oxide Fuel Cell

A new multigeneration energy process, based on biomass gasification, for generating power with different useful outputs, such as heating, cooling, hot water and hydrogen, is proposed in this case study. One of the most interesting alternatives of multigeneration energy systems is combined biomass gasifier and SOFC technology. The exhaustive definition of the suggested integrated gasification combined process is supplied and it is expected that this process meets with sustainable development needs.

Integrated Gasification Combined Cycles

0.62

104,000

ExD,system

102,000

System

0.61 0.6

100,000

0.59

98,000 0.58 96,000 0.57

94,000

0.56

92,000 90,000

0.55

88,000

0.54

86,000 800

850

900

950 1000 1050 1100 1150 Gas turbine inlet temperature (°C)

Exergy efficiency

Exergy destruction rate (kW)

106,000

0.53 1250

1200

0.575

100,000

0.572

99,500

0.569

Exergy destruction rate (kW)

100,500

0.566

99,000 98,500 98,000

ExD,system

0.563

System

0.56 0.557

97,500

0.554

97,000

0.551

96,500

0.548

96,000 1500

2000

2500 3000 3500 Gas turbine inlet pressure (kPa)

0.545 4000

Fig. 42 Effects of gas turbine inlet pressure on the exergy destruction rate and exergy efficiency of integrated system.

Environmental impacts 500

469.5 414.6

400

386.4 327.9

300 200 100

32.83

21.72

36.57

18.18

43.64

15.57

56.54

13.14

0 Single generation

Cogeneration

Trigeneration

Exergy efficency (%) Carbon dioxide emissions (kg/kWh) Carbon monoxide emissions (kg/kWh) Fig. 43 Comparison of exergy efficiency, unit CO2 and CO emissions of generation cycles.

Multigeneration

Exergy efficiency

Fig. 41 Effects of gas turbine inlet temperature on the exergy destruction rate and exergy efficiency of integrated system.

439

440

Integrated Gasification Combined Cycles

Water Air 9 16 Pump-I

Three-way valve-I 13

Biomass dryer

10

1 Biomass

33

HRSG

14

Biomass gasifier

2

Blower

11

12

3

15 17 Flue gas 35

5

45 34

Separator Water 21

HEX-I

22 23

64

6

HT-WGS reactor

27

Electricity

HWST 28

44 38

40

Expander-II

Mixing chamber

25 3-way valve-III

43

47 41

65 HEX-II

Resorber-I

39

24

Cold water

Heating 37

Expander-I

Three-way valve-II

26

Electricity

Turbine

Syngas cleaning

Ash and char 4

LT-WGS reactor

36

Electricity

42

Hot water

Resorber-II

Pump-II

7 Compressor

Hot water

46

Cold water

18 29

Valve PSA

Burner 32

31

30

55

HEX-III

Anode SOFC

20

51

50

Cathode

Expansion valve-II 53

Pump-III 48 62 Absorber

58 Expansion valve-I

52

49 19

59

Condenser

b

a 8

54

Generator

Hydrogen

57

56 60 Evaporator

District cooling

63

61

Fig. 44 Schematic diagram of biomass resources-based integrated gasification combined cycle (IGCC). HEX, heat exchanger; HRSG, heat recovery steam generator; HWST, hot water storage tank; HT-WGS, high temperature water-gas shift; LT-WGS, low temperature water-gas shift; PSA, pressure swing adsorption; SOFC, solid oxide fuel cell.

The biomass gasification technology is utilized to provide thermal power as the prime mover of integrated gasification combined process introduced in Fig. 44. The electricity generation is provided by using the double state organic Rankine cycle (DS-ORC) where ammonia–water mixture is chosen to be as the working fluid. Also, the synthesis gaseous produced from the biomass gasifier is utilized to feed the direct reforming SOFC unit for power generation. A single effect absorption cooling process is combined with the biomass gasification-based integrated system to supply space cooling using heat of exhaust gaseous coming from HEX-I. A part of the generated synthesis gaseous is directed to be more hydrogen enriched using the high and low temperature WGSRs for hydrogen generation.

4.10.7.2.1

Balance equations

For detailed thermodynamic assessment of integrated gasification combined system, the balance equations for system components are defined in the following subsection.

Integrated Gasification Combined Cycles

441

Biomass dryer subcomponent: the mass, energy, entropy, and exergy balance equations for biomass dryer subcomponent, which is illustrated in Fig. 44, can be written under the steady state and steady flow conditions as follows: _ 13 ¼ m _2 Mass: m _1þm

ð773Þ

_ 13 h13 ¼ m _ 12 h12 Energy: m _ 1 h1 þ m

ð774Þ

_ 13 s13 þ S_ gen;bd ¼ m _ 12 s12 Entropy:m _ 1 s1 þ m

ð775Þ

_ D;bd _ 13 ex 13 ¼ m _ 12 ex12 þ Ex Exergy: m _ 1 ex 1 þ m

ð776Þ

Biomass gasifier subcomponent: the mass, energy, entropy, and exergy balance equations of biomass gasifier subcomponent under steady state and steady flow conditions are written in Eqs. (777)–(780). _ 12 þ m _ 33 ¼ m _3þm _ 14 Mass: m _2þm

ð777Þ

_ 12 h12 þ m _ 33 h33 ¼ m _ 3 h3 þ m _ 14 h14 Energy: m _ 2 h2 þ m

ð778Þ

_ 12 s12 þ m _ 33 s33 þ S_ gen;bg ¼ m _ 3 s3 þ m _ 14 s14 Entropy: m _ 2 s2 þ m

ð779Þ

_ D;bg _ 12 ex 12 þ m _ 33 ex 33 ¼ m _ 3 ex ch _ 14 ex 14 þ Ex Exergy: m _ 2 ex ch 2 þm 3 þm

ð780Þ

In the following equalities, the chemical exergy accounts for syngas are defined accordingly: ch ch ch ex ch 2 ¼ rH2 MH2 ex H2 þ rCO MCO ex CO þ rCH4 MCH4 ex CH4

 0  0 0 ex ch CO ¼ hCO hC þ 0:5hO2

 T0 s0CO





s0C þ 0:5s0O2

ch þ ex ch C þ 0:5ex O2

ð781Þ

ð782Þ

 0  0 0 ex ch CO2 ¼ hCO2 hC þ hO2

 T0 s0CO

  ch þ ex ch s0C þ s0O2 C þ ex O2

ð783Þ

 0  0 0 ex ch CH4 ¼ hCH4 hC þ 2hH2

 T0 s0CO



ð784Þ

s0C þ 2s0H2



ch þ ex ch C þ 2ex H2

where M is the molar mass (kg/kmol) of the corresponding substance, h0 is the enthalpy (kJ/kmol), s0 is the entropy (kJ/kmol K), exch is the chemical exergy (kJ/kmol) of the corresponding substance at 251C and 101.3 kPa, the bar accent represents molar basis, r is the mass concentration (wt%) of the substance. The point 2 is the dry basis biomass feedstock, moisture and ash to model gasification in a molar basis. At point 3, the composition of the synthesis gas that consists of hydrogen, carbon monoxide, carbon dioxide, nitrogen, methane, ash and trace amounts of sulfur compounds exit from the biomass gasifier. HRSG subcomponent: the mass, energy, entropy, and exergy balance equations for HRSG subcomponent can be defined under the steady state and steady flow conditions as follows: _ 11 ; m _ 15 Mass: m _ 10 ¼ m _ 14 ¼ m

ð785Þ

_ 14 h14 ¼ m _ 11 h11 þ m _ 15 h15 Energy: m _ 10 h10 þ m

ð786Þ

_ 14 s14 þ S_ gen;HRSG ¼ m _ 11 s11 þ m _ 15 s15 Entropy: m _ 10 s10 þ m

ð787Þ

_ D;HRSG _ 14 ex14 ¼ m _ 11 ex 11 þ m _ 15 ex15 þ Ex Exergy: m _ 10 ex 10 þ m

ð788Þ

Three-way valve-I subcomponent: the mass, energy, entropy, and exergy balance equations are written for three-way valve-I subcomponent under steady state and steady flow conditions. _ 12 þ m _ 13 Mass: m _ 11 ¼ m

ð789Þ

_ 12 h12 þ m _ 13 h13 Energy: m _ 11 h11 ¼ m

ð790Þ

442

Integrated Gasification Combined Cycles

_ 12 s12 þ m _ 13 s13 Entropy: m _ 11 s11 þ S_ gen;3wv_I ¼ m

ð791Þ

_ D;3wv_I _ 12 ex12 þ m _ 13 ex13 þ Ex Exergy: m _ 11 ex11 ¼ m

ð792Þ

Syngas cleaning subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for syngas cleaning subcomponent can be defined as follows: _4þm _5 Mass: m _3¼m

ð793Þ

_ 4 h4 þ m _ 5 h5 Energy: m _ 3 h3 ¼ m

ð794Þ

_ 4 s4 þ m _ 5 s5 Entropy: m _ 3 s3 þ S_ gen;sc ¼ m

ð795Þ

_ D;sc _ 4 ex 4 þ m _ 5 ex5 þ Ex Exergy: m _ 3 ex3 ¼ m

ð796Þ

Pump-I component: the mass, energy, entropy, and exergy balance equations for pump-I subcomponent can be defined under the steady state and steady flow conditions as follows: _ 10 Mass: m _9¼m

ð797Þ

_ p_I ¼ m _ 10 h10 Energy: m _ 9 h9 þ W

ð798Þ

_ 10 s10 Entropy: m _ 9 s9 þ S_ gen;p_I ¼ m

ð799Þ

_ p_I ¼ m _ D;p_I _ 10 ex 10 þ Ex Exergy: m _ 9 ex 9 þ W

ð800Þ

Blower subcomponent: the mass, energy, entropy, and exergy balance equations for blower subcomponent can be written defined under the steady state and steady flow conditions as given below: _ 17 Mass: m _ 16 ¼ m

ð801Þ

_ bl ¼ m _ 17 h17 Energy: m _ 16 h16 þ W

ð802Þ

_ 17 s17 Entropy: m _ 16 s16 þ S_ gen;bl ¼ m

ð803Þ

_ bl ¼ m _ D;bl _ 17 ex 17 þ Ex Exergy: m _ 16 ex 16 þ W

ð804Þ

HEX-I subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for HEX-I subcomponent can be written as follows: _ 6; m _ 18 ; m _ 22 ; m _ 43 Mass: m _5¼m _ 17 ¼ m _ 21 ¼ m _ 34 ¼ m

ð805Þ

_ 17 h17 þ m _ 21 h21 þ m _ 43 h43 ¼ m _ 6 h6 þ m _ 18 h18 þ m _ 22 h22 þ m _ 34 h34 Energy: m _ 5 h5 þ m

ð806Þ

_ 17 s17 þ m _ 21 s21 þ m _ 43 s43 þ S_ gen;HEX_I ¼ m _ 6 s6 þ m _ 18 s18 þ m _ 22 s22 þ m _ 34 s34 Entropy: m _ 5 s5 þ m

ð807Þ

_ D;sc _ 17 ex17 þ m _ 21 ex 21 þ m _ 43 ex 43 ¼ m _ 6 ex6 þ m _ 18 ex18 þ m _ 22 ex 22 þ m _ 34 ex 34 þ Ex Exergy: m _ 5 ex5 þ m

ð808Þ

Expander-I subcomponent: under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for expander-I subcomponent can be defined as follows: _ 35 Mass: m _ 34 ¼ m

ð809Þ

_ ep_I _ 35 h35 þ W Energy: m _ 34 h34 ¼ m

ð810Þ

_ 35 s35 Entropy: m _ 34 s34 þ S_ gen;ep_I ¼ m

ð811Þ

_ ep_I þ Ex _ D;ep_I _ 35 ex35 þ W Exergy: m _ 34 ex 34 ¼ m

ð812Þ

Integrated Gasification Combined Cycles

443

Separator subcomponent: the mass, energy, entropy, and exergy balance equalities can be defined for separator subcomponent under the steady state and steady flow conditions as follows: _ 36 þ m _ 39 Mass: m _ 35 ¼ m

ð813Þ

_ 36 h36 þ m _ 39 h39 Energy: m _ 35 h35 ¼ m

ð814Þ

_ 36 s36 þ m _ 39 s39 Entropy: m _ 35 s35 þ S_ gen;sp ¼ m

ð815Þ

_ D;sp _ 36 ex 36 þ m _ 39 ex 39 þ Ex Exergy: m _ 35 ex 35 ¼ m

ð816Þ

Turbine subcomponent: the mass, energy, entropy, and exergy balance equations are written for turbine subcomponent under the steady state and steady flows conditions. _ 37 Mass: m _ 36 ¼ m

ð817Þ

_ tur _ 37 h37 þ W Energy: m _ 36 h36 ¼ m

ð818Þ

_ 37 s37 Entropy: m _ 36 s36 þ S_ gen;tur ¼ m

ð819Þ

_ tur þ Ex _ D;tur _ 37 ex37 þ W Exergy: m _ 36 ex36 ¼ m

ð820Þ

Resorber-I subcomponent: the mass, energy, entropy, and exergy balance equations are written for resorber-I subcomponent under the steady state and steady flow conditions. _ 38 ; m _ 45 _ 44 ¼ m Mass: m _ 37 ¼ m

ð821Þ

_ 44 h44 ¼ m _ 38 h38 þ m _ 45 h45 Energy: m _ 37 h37 þ m

ð822Þ

_ 44 s44 þ S_ gen;rb_I ¼ m _ 38 h38 þ m _ 45 h45 Entropy: m _ 37 s37 þ m

ð823Þ

_ D;rb_I _ 44 ex44 ¼ m _ 38 ex38 þ m _ 45 ex45 þ Ex Exergy: m _ 37 ex 37 þ m

ð824Þ

Expander-II subcomponent: the mass, energy, entropy, and exergy balance equations are defined for expander-II subcomponent as given below: _ 35 Mass: m _ 34 ¼ m

ð825Þ

_ ep_I _ 35 h35 þ W Energy: m _ 34 h34 ¼ m

ð826Þ

_ 35 s35 Entropy: m _ 34 s34 þ S_ gen;ep_I ¼ m

ð827Þ

_ ep_I þ Ex _ D;ep_E _ 35 ex35 þ W Exergy: m _ 34 ex 34 ¼ m

ð828Þ

Mixing chamber subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for mixing chamber subcomponent can be given as _ 40 ¼ m _ 41 Mass: m _ 38 þ m

ð829Þ

_ 40 h40 ¼ m _ 41 h41 Energy: m _ 38 h38 þ m

ð830Þ

_ 40 s40 þ S_ gen;mc ¼ m _ 41 s41 Entropy: m _ 38 s38 þ m

ð831Þ

_ D;mc _ 40 ex 40 ¼ m _ 41 ex 41 þ Ex Exergy: m _ 38 ex 38 þ m

ð832Þ

Resorber-II subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations can be defined as follows:

444

Integrated Gasification Combined Cycles _ 42 ; m _ 47 Mass: m _ 41 ¼ m _ 46 ¼ m

ð833Þ

_ 46 h46 ¼ m _ 42 h42 þ m _ 47 h47 Energy: m _ 41 h41 þ m

ð834Þ

_ 46 s46 þ S_ gen;rb_II ¼ m _ 42 h42 þ m _ 47 h47 Entropy: m _ 41 s41 þ m

ð835Þ

_ D;rb_II _ 46 ex46 ¼ m _ 42 ex 42 þ m _ 47 ex47 þ Ex Exergy: m _ 41 ex 41 þ m

ð836Þ

Pump-II subcomponent: the mass, energy, entropy, and exergy balance equations for pump-II subcomponent can be written under the steady state and steady flow conditions as given below: _ 43 Mass: m _ 42 ¼ m

ð837Þ

_ p_II ¼ m _ 43 h43 Energy: m _ 42 h42 þ W

ð838Þ

_ 43 s43 Entropy: m _ 42 s42 þ S_ gen;p_II ¼ m

ð839Þ

_ p_II ¼ m _ D;p_II _ 43 ex 43 þ Ex Exergy: m _ 42 ex42 þ W

ð840Þ

Generator subcomponent: under the steady state and steady flow situation, the mass, energy, entropy, and exergy balance equalities for generator subcomponent can be defined as follows: _ 19 ; m _ 51 þ m _ 54 _ 50 ¼ m Mass: m _ 18 ¼ m

ð841Þ

_ gena ¼ m _ 50 h50 þ Q _ 19 h19 þ m _ 51 h51 þ m _ 54 h54 Energy: m _ 18 h18 þ m

ð842Þ

_ gena =Tgena þ S_ gen;gena ¼ m _ 50 s50 þ Q _ 19 s19 þ m _ 51 s51 þ m _ 54 s54 Entropy: m _ 18 s18 þ m

ð843Þ

_ Q ¼m _ D;gena _ 50 ex 50 þ Ex _ 19 ex 19 þ m _ 51 ex 51 þ m _ 54 ex 54 þ Ex Exergy: m _ 18 ex 18 þ m gena

ð844Þ

Condenser subcomponent: the mass, energy, entropy, and exergy balance equalities can be written for condenser under the steady state and steady flow conditions as follows: _ 55 ; m _ 59 Mass: m _ 54 ¼ m _ 58 ¼ m

ð845Þ

_ con _ 58 h58 ¼ m _ 55 h55 þ m _ 59 h59 þ Q Energy: m _ 54 h54 þ m

ð846Þ

_ con =Tcon _ 58 s58 þ S_ gen;con ¼ m _ 55 s55 þ m _ 59 s59 þ Q Entropy: m _ 54 s54 þ m

ð847Þ

_ Q þ Ex _ D;con _ 58 ex 58 ¼ m _ 55 ex 55 þ m _ 59 ex 59 þ Ex Exergy: m _ 54 ex54 þ m con

ð848Þ

Expansion valve-I subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for expansion valve-I subcomponent can be written as follows: _ 56 Mass: m _ 55 ¼ m

ð849Þ

_ 56 h56 Energy: m _ 55 h55 ¼ m

ð850Þ

_ 56 s56 Entropy: m _ 55 s55 þ S_ gen;ev_I ¼ m

ð851Þ

_ D;ev_I _ 56 ex56 þ Ex Exergy: m _ 55 ex55 ¼ m

ð852Þ

Evaporator subcomponent: the mass, energy, entropy, and exergy balance equations can be defined for evaporator subcomponent under the steady state and steady flow conditions as follows:

Integrated Gasification Combined Cycles

445

_ 57 ; m _ 61 Mass: m _ 56 ¼ m _ 60 ¼ m

ð853Þ

_ eva ¼ m _ 60 h60 þ Q _ 57 h57 þ m _ 61 h61 Energy: m _ 56 h56 þ m

ð854Þ

_ eva =Teva þ S_ gen;eva ¼ m _ 60 s60 þ Q _ 57 s57 þ m _ 61 s61 Entropy: m _ 56 s56 þ m

ð855Þ

_ Q ¼m _ D;eva _ 60 ex60 þ Ex _ 57 ex57 þ m _ 61 ex61 þ Ex Exergy: m _ 56 ex56 þ m eva

ð856Þ

Absorber subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equalities for absorber subcomponent are written as follows: _ 53 þ m _ 57 ; m _ 63 Mass: m _ 48 ¼ m _ 62 ¼ m

ð857Þ

_ abs _ 57 h57 þ m _ 62 h62 ¼ m _ 48 h48 þ m _ 63 h63 þ Q Energy: m _ 53 h53 þ m

ð858Þ

_ abs =Tabs _ 57 s57 þ m _ 62 s62 þ S_ gen;abs ¼ m _ 48 s48 þ m _ 63 s63 þ Q Entropy: m _ 53 s53 þ m

ð859Þ

_ Q þ Ex _ D;abs _ 57 ex 57 þ m _ 62 ex62 ¼ m _ 48 ex 48 þ m _ 63 ex63 þ Ex Exergy: m _ 53 ex 53 þ m abs

ð860Þ

Pump-III subcomponent: the mass, energy, entropy, and exergy balance equations for pump-II subcomponent can be defined under the steady state and steady flow conditions as given below: _ 49 Mass: m _ 48 ¼ m

ð861Þ

_ p_III ¼ m _ 49 h49 Energy: m _ 48 h48 þ W

ð862Þ

_ 49 s49 Entropy: m _ 48 s48 þ S_ gen;p_III ¼ m

ð863Þ

_ p_III ¼ m _ D;p_III _ 49 ex 49 þ Ex Exergy: m _ 48 ex48 þ W

ð864Þ

Expansion valve-II subcomponent: the mass, energy, entropy, and exergy balance equations for expansion valve-II subcomponent can be written under the steady state and steady flow conditions as given below: _ 53 Mass: m _ 52 ¼ m

ð865Þ

_ 53 h53 Energy: m _ 52 h52 ¼ m

ð866Þ

_ 53 s53 Entropy: m _ 52 s52 þ S_ gen;ev_II ¼ m

ð867Þ

_ D;ev_II _ 53 ex53 þ Ex Exergy: m _ 52 ex52 ¼ m

ð868Þ

HEX-III subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for HEX-III subcomponent can be written as given below: _ 50 ; m _ 52 Mass: m _ 49 ¼ m _ 51 ¼ m _ HEX _ 51 h51 þ Q Energy: m _ 49 h49 þ m

III

_ 50 h50 þ m _ 52 h52 ¼m

ð869Þ ð870Þ

_ HEX_III =THEX_III þ S_ gen;HEX_III ¼ m _ 51 s51 þ Q _ 50 s50 þ m _ 52 s52 Entropy: m _ 49 s49 þ m

ð871Þ

_ Q _ D;HEX_III _ 51 ex 51 þ Ex _ 50 ex50 þ m _ 52 ex52 þ Ex Exergy: m _ 49 ex49 þ m HEX_III ¼ m

ð872Þ

Three-way valve-II subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for three-way valve-I subcomponent are written as follows: _ 23 þ m _ 24 Mass: m _ 22 ¼ m

ð873Þ

_ 23 h23 þ m _ 24 h24 Energy: m _ 22 h22 ¼ m

ð874Þ

_ 23 s23 þ m _ 24 s24 Entropy: m _ 22 s22 þ S_ gen;3wv_II ¼ m

ð875Þ

446

Integrated Gasification Combined Cycles

Three-way valve-III subcomponent: the mass, energy, entropy, and exergy balance equations are written for three-way valve-III subcomponent under steady state and steady flow conditions. _7þm _ 25 Mass: m _6¼m

ð876Þ

_ 7 h7 þ m _ 25 h25 Energy: m _ 6 h6 ¼ m

ð877Þ

_ 7 s7 þ m _ 25 s25 Entropy:m _ 6 s6 þ S_ gen;3wv_III ¼ m

ð878Þ

_ D;3wv_III _ 7 ex7 þ m _ 25 ex 25 þ Ex Exergy: m _ 6 ex 6 ¼ m

ð879Þ

High temperature WGSR subcomponent: the balance equations of high temperature water-gas shift reactor (HT-WGSR) process under steady state and steady flow conditions are written as given below: _ 25 ¼ m _ 26 Mass: m _ 23 þ m

ð880Þ

_ steam ¼ m _ 25 h25 þ Q _ 26 h26 Energy: m _ 23 h23 þ m

ð881Þ

_ steam =THTR þ S_ gen;HTR ¼ m _ 25 s25 þ Q _ 26 s26 Entropy: m _ 23 s23 þ m

ð882Þ

_ D;HTR _ Q _ 25 ex 25 þ Ex _ 26 ex 26 þ Ex Exergy: m _ 23 ex23 þ m HTR ¼ m

ð883Þ

Low temperature WGSR subcomponent: the mass, energy, entropy, and exergy balance equations of low temperature water-gas shift reactor (LT-WGSR) process under steady state and steady flow conditions can be defined as given below: _ 26 ¼ m _ 27 Mass: m _ 24 þ m

ð884Þ

_ steam ¼ m _ 26 h26 þ Q _ 27 h27 Energy: m _ 24 h24 þ m

ð885Þ

_ steam =TLTR þ S_ gen;LTR ¼ m _ 26 s26 þ Q _ 27 s27 Entropy: m _ 24 s24 þ m

ð886Þ

_ Q ¼m _ D;LTR _ 26 ex26 þ Ex _ 27 ex 27 þ Ex Exergy: m _ 24 ex 24 þ m LTR

ð887Þ

HEX-II subcomponent: the mass, energy, entropy, and exergy balance equations are written for HEX-II subcomponent under the steady state and steady flow conditions. _ 65 ; m _ 28 Mass: m _ 64 ¼ m _ 27 ¼ m

ð888Þ

_ 27 h27 ¼ m _ 65 h65 þ m _ 28 h28 Energy: m _ 64 h64 þ m

ð889Þ

_ 27 s27 þ S_ gen;HEX_II ¼ m _ 65 s65 þ m _ 28 s28 Entropy: m _ 64 s64 þ m

ð890Þ

_ D;HEX_II _ 27 ex 27 ¼ m _ 65 ex 65 þ m _ 28 ex 28 þ Ex Exergy: m _ 64 ex64 þ m

ð891Þ

Compressor subcomponent: the mass, energy, entropy, and exergy balance equations for compressor subcomponent can be written defined under the steady state and steady flow conditions as given below: _ 29 Mass: m _ 28 ¼ m

ð892Þ

_ cp ¼ m _ 29 h29 Energy: m _ 28 h28 þ W

ð893Þ

_ 29 s29 Entropy: m _ 28 s28 þ S_ gen;cp ¼ m

ð894Þ

_ cp ¼ m _ D;cp _ 29 ex 29 þ Ex Exergy: m _ 28 ex28 þ W

ð895Þ

PSA subcomponent: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for PSA subcomponent can be defined as follows: _ 30 þ m _ 31 Mass: m _ 29 ¼ m

ð896Þ

Integrated Gasification Combined Cycles

447

_ 30 h30 þ m _ 31 h31 Energy: m _ 29 h29 ¼ m

ð897Þ

_ 30 s30 þ m _ 31 s31 Entropy: m _ 29 s29 ¼ m

ð898Þ

_ D;PSA _ 30 ex 30 þ m _ 31 ex31 þ Ex Exergy: m _ 29 ex 29 ¼ m

ð899Þ

Valve subcomponent: the mass, energy, entropy, and exergy balance equations for valve subcomponent can be written under the steady state and steady flow conditions as given below:

4.10.7.2.2

_ 32 Mass: m _ 31 ¼ m

ð900Þ

_ 32 h32 Energy: m _ 31 h31 ¼ m

ð901Þ

_ 32 s32 Entropy: m _ 31 s31 þ S_ gen;vlf ¼ m

ð902Þ

_ D;vlf _ 32 ex32 þ Ex Exergy: m _ 31 ex31 ¼ m

ð903Þ

Solid oxide fuel cell

In this section, the direct inner improving SOFC is investigated. The schematic diagram of SOFC subcomponent in integrated gasification combined process is illustrated in Fig. 44. The entering gaseous flow to the anode side is mixed with recirculated gaseous at the SOFC outlet. This supplies the required water through the high-water content gaseous flow exiting the SOFC. In this design, SOFC electricity current density is depend on the molar flow rate of utilized hydrogen rate which assists in detecting the outlet gaseous combination. Furthermore, the isolated cell is considered to define the quantity of air exiting from the cell carrying the released and unused heat energy. For the electrochemical viewpoint, the Eqs. (903)–(905) can be considered where methane steam reforming and water-gas shift reactions, along with the electrochemical reaction happens simultaneously at the SOFC [58]. CH4 þ H2 O2 CO þ 3H2

ð904Þ

CO þ H2 O2 CO2 þ H2

ð905Þ

H2 þ 1=2O2 -H2 O

ð906Þ

The stability constants for the two governing reactions of the SOFC that can be defined as function of the molar concentrations, decreased to partial pressure rates, of the chemical reaction types. The amount of equilibrium constants can be defined as given below [59]: logK ¼ A0 þ A1 T þ A2 T 2 þ A3 T 3 þ A4 T 4

ð907Þ

The values of A0 , A1 , A2 , A3 , and A4 constants are illustrated in Table 14 for reforming and also shifting processes, respectively. The equilibrium constants can be calculated as the function of the molar concentration of the gaseous flow at the point 8 of SOFC subcomponent.  2 3 yeq;H :yeq;CO P 2 KReforming ¼ ð908Þ yeq;CH4 :yeq;H2 O P0 KShift ¼

  yeq;H2 :yeq;CO2 P 2 yeq;CO :yeq;H2 O P0

ð909Þ

where yeq is the molar ratio of the types of equilibrium combination of the gaseous flow at the SOFC outlet. Table 14 Constants A0 A1 A2 A3 A4

Values of equilibrium constants Reforming 66.139488 0.195027 2.2523  10 1.2406  10 2.6312  10

Shifting

4 7 11

13.209723 3.915  10 2 4.6374  10 5 2.547  10 8 5.47  10 12

Source: Data taken from Fryda L, Panopoulos KD, Karl J, Kakaras E. Exergetic analysis of solid oxide fuel cell and biomass gasification integration with heat pipes. Energy 2008;33:292–9.

448

Integrated Gasification Combined Cycles

To define the combination of the gaseous fuel at the SOFC outlet, the equilibrium combinations should be defined. Depending on the hydrogen use ingredient, the utilized hydrogen molar rate can be defined in relationship with the chemical reactions contents as given below:  2O  þ 3cR þ cS ð910Þ cH ¼ l n_ H 1

where l is working situation coefficient, and can be given as follows: l ¼ Uf =ð1

r þ rUf Þ

ð911Þ

where r is the recirculation rate, Uf is the utilization coefficient. cH, cR, and cS are the hydrogen, methane, and carbon monoxide transformation rates during the SOFC electrochemical reaction, reforming and shifting reactions, respectively. The total molar flow rate of the gaseous flow at point 8 should be defined based on the input gaseous flow at point 7 as given below: n_ 8 ¼ n_ 7 þ 2cR

ð912Þ

The molar rates of gaseous flow combination for equilibrium state at the outlet of SOFC can be defined based on the input combinations as given below: 4 4 ¼ n_ CH n_ CH 7 8

ð913Þ

cR

_ CO n_ CO 7 þ cR 8 ¼n

ð914Þ

cs

2 2 n_ CO þ cS ¼ n_ CO 7 8

2 2 _H n_ H 8 ¼n 7 þ 3cR þ cS

2O 2O n_ H ¼ n_ H 8 7

cR

ð915Þ cH

ð916Þ

cS þ cH

ð917Þ

The molar rate equations related to equilibrium constant are used to define the combination of gaseous flow exiting from the SOFC subcomponent. The molar rate of gaseous flow at point b is related to the known input flow, 7 by adding the recirculated rate.   ð918Þ n_ ib ¼ n_ i7 þ r yeq;i n_ ia The air flow is related to transport the waste heat produced within the SOFC subcomponent. The molar flow rate of O2 based on the air flow (flow 19) can be defined associated with the transformation of H2 throughout the system as given below:   1 cH 2 ¼ n_ O ð919Þ 5 2 U0

where U0 is the air utilization indicator which is the oxidant in the SOFC reaction. The outlet molar rate of air can be defined considering the real quantity of ðcH =2Þ of O2 to be used in the chemical reaction. The molar flow rates of accompanied N2 with the inserting air can be defined based on the ratio of O2/N2 in reference air. The electricity current of SOFC subcomponent can be defined as a function of supplied flow rate of H2 as given below:  2  I ¼ 2Fl n_ H ð920Þ 7 þ 3cR þ cS This equation can be simply reduced as function of the used H2 rate as follows: I ¼ 2FcH

ð921Þ

where F is the Faraday constant. The SOFC voltage can be defined by defining the several polarizations affecting the SOFC, and subtracting them from the Nernst open circuit voltage term. The Nernst voltage term can be defined as given below [60]: ! DG0 RT P H2 O ð922Þ ln E¼ 1=2 2F 2F PH P 2

O2

where DG0 is the molar Gibbs free energy change, and can be defined based on the electrochemical reaction of SOFC subcomponent. The amounts of partial pressure (Pi) in the Eq. (919) are to be stated in form of molar ratios of equilibrium flows. The three polarizations affecting the SOFC efficiency are the ohmic, activation, and concentration polarization. The ohmic losses are caused by the flow of electrons through the anode, cathode, and interconnector and the ionic flow through the SOFC electrolyte. The ohmic losses can be calculated as the function of the cell parts thickness (L) and specific electrical resistivity (r) which is temperature dependent, as given below: X Vohm ¼ i rk Lk ð923Þ where k shows the anode, cathode, electrolyte, and interconnector components of the SOFC. The values of cell components thickness (Lk) are given in Table 15. The resistivity of parts can be defined as given below [64,65]: rk ¼ C1 expðC2 =T Þ

ð924Þ

Integrated Gasification Combined Cycles

Table 15

Input data of solid oxide fuel cell (SOFC) for system design 0.04 mm [61] 0.75 mm [61] 0.05 mm [61] 4000 A/m2 5 m2/s [62] 3.5  10 6 m2/s [62] 7.3  10 2 7 5.7  10 A/m [62] 2 9 7  10 A/m [62] 140 kJ/mol [63] 137 kJ/mol [63] 1 0.25 83% 20% 7501C 1001C 101.3 kPa

Electrolyte thickness (Le) Anode thickness (La) Cathode thickness (Lc) Operating current density a Anode effective gas diffusion factor (Deff ) c Cathode effective gas diffusion factor (Deff ) Anode pre-exponential factor (ga ) Cathode pre-exponential factor (gc ) a Anode activation energy (Eact ) c ) Cathode activation energy (Eact Anode exchange current density constant (m) Cathode exchange current density constant (n) Fuel utilization Recirculation ratio Operating temperature Temperature difference across the cell Operating pressure

Table 16

449

Specific resistivity values of anode-based solid oxide fuel cell (SOFC) subcomponent

System parts

C1

C2

Anode (Ni/yttria stabilized zirconia (YSZ) cermet) Cathode (lanthanum strontium manganite (LSM)-YSZ) Electrolyte (YSZ) Interconnector (doped LaCrO3)

2.98  10 5 8.114  10 5 2.94  10 5 3.215  10 4

1392 600 10,350 0

where C1 and C2 values are written in Table 16, also these data ensured by Wongchanapai [65], for each part of the SOFC subcomponent. The activation energies in SOFC electrochemical reaction are the energy barriers which must be overcome by using the reactant. The activation losses are accompanied with these energy barriers because of the charge transfer through the three phase regions of anode and cathode electrodes. The activation polarization value can be defined based on the Butler–Volmer equation as given below [64]:     RT 1 RT i 1 sinh 1 sinh Vact ¼ þ ð925Þ 2F 2ia0 2F 2ic0

where ia0 and ic0 are the anode and cathode exchange current density values, respectively, and can be defined based on the semiempirical correlations as follows [66]:  a     PH2 PH2 O m Eact ð926Þ exp ia0 ¼ ga RT P0 P0 ic0 ¼ gc



PO2 P0

n

 c  Eact exp RT

ð927Þ

The pre-exponential factors (ga and gc ), the values of activation energy (Eaact and Ecact ), and the empirical constant (m and n) are illustrated in Table 15. The concentration polarization is caused based on the mass transport limitations of reactants, and can be defined as a function of the limiting current density at the anode and cathode sides of SOFC: !   P ex i RT i RT Vcon;a ¼ ln 1 a þ ln 1 þ exH2 a ð928Þ PH2 O is 2F is 2F Vcon;c ¼

 RT ln 1 4F

i ics



ð929Þ

The Eqs. (928) and (929) are used to define the anode and cathode sites limiting current density outputs, which are written as the function of temperature, and also hydrogen and oxygen partial pressure at the SOFC output cases: ias ¼

ex 2F  PH  Daeff 2 R  T  La

ð930Þ

450

Integrated Gasification Combined Cycles

ics ¼  P

ex 4F  PO P  Dceff  2 ex PO  R  T  Lc 2

ð931Þ

where La and Lc are the thickness values of anode and cathode of SOFC, respectively. Daeff and Dceff are effective gas diffusion factors for the anode and cathode sites of SOFC subcomponent, respectively. The values of thickness and effective gas factors for anode and cathode are shown in Table 15. Considering the supposition of utilizing excess air to release the produced heat at the SOFC, the air use can be defined by using the energetic balance equation of the insulated cell [67]. X X X X _ SOFC þ Hair ð932Þ Hf þ Hf þ Hair ¼ W in

out

in

out

where Hf and Hair are the total enthalpy of fuel and air flow rates at input and output of the SOFC, respectively. The electricity production rate based on the air utilization can be defined as given below: _ SOFC ¼ i  V  A  Nstack  Nelement W

ð933Þ

where i is the SOFC current density, V is the working cell voltage, A is the active area of SOFC, Nstack is the number of stacks of SOFC, and Nelement is the number of elements in the SOFC subcomponent. The power production performance of SOFC can be defined as follows: ZSOFC ¼

_ SOFC W n_ 1 LHV

ð934Þ

The exergetic efficiency of SOFC can be defined based on the exergy associated with the inlet fuel flow to the SOFC: cSOFC ¼

_ SOFC W _ fuel;in Ex

ð935Þ

The cost function of SOFC subcomponent can be defined based on the specific cost of fuel cell. The specific SOFC cost can be considered as 0.1442 $/cm2 [63]. The total purchase cost of fuel cell can be defined based on the total SOFC area after identifying the number of cells required.

4.10.7.2.3

Thermal characteristics of biomass resources

The thermal characteristics of biomass resources can be defined concerning of the specific heating value, which is submitted by Jenkins [68] as the function of temperature, and it can be written for dry-biomass resources as given below: cp;db ¼ 1:16  10

3

5:0854  10

T

2

ð936Þ

where the temperature is in K. The moisture impact on specific heat definition can be written as follows: cp;b ¼ cp;db ð1

oÞ þ ocp;w

ð937Þ

where o is the moisture faction on wet basis and cp,w is the water specific heat. The enthalpy and entropy contents at several state points can be defined as follows: Z T cp dT ð938Þ h ¼ h0 þ T0

s ¼ s0 þ

Z

T

cp T0

dT T

ð939Þ

The Gibbs free energy can be written as the function of operating temperature as illustrated in the Eq. (937) as given below, which is presented as modified from Basu [67] indicated in kJ/kmol: DgTf ¼ hf0 þ z0 T þ z1 TlnT þ z2 T 2 þ z3 T 3 þ z4 T 4 þ z5 T

1

þ z6

ð940Þ

where enthalpy of formation is at the reference temperature and pressure, and temperature is in K. The variables in Eq. (938) are written in Table 17. The formation enthalpy of moist biomass resources can be defined as the function of the formation enthalpy of dry-biomass and moisture content as given below:   hfb ¼ hfdb 2:44o 21:83H ð1 oÞ ð941Þ The formation enthalpy of the dry biomass can be defined as given below [69]: hfdb ¼ 349:1C þ 1178:2H

103:4O þ 100:5S

15:1N

ð942Þ

where C, H, O, S, and N are the weight fractions of ultimate biomass combinations. The chemical exergy of biomass resources can be defined as the function of LHV. exch b ¼ bLHV b

ð943Þ

Integrated Gasification Combined Cycles

Table 17

Z0 Z1 Z2 Z3 Z4 Z5 Z6

451

Coefficients of empirical correlation for Gibbs free energy definitions CO

CO2

CH4

H2O

61.31 5.619 1.19  10 2 3.1915  10 6.153  10 10 244,550 868

120.7 19.49 3.133  10 2 1.224  10 5 2.3153  10 9 244,550 5270

223.4 46.2 1.13  10 2 6.595  10 6 2.2157  10 9 244,550 14,110

17.2 8.95 3.672  10 3 2.6045  10 4.927  10 10 0 2868

6

6

Source: Data taken from Basu P. Biomass gasification and pyrolysis: practical design and theory. Oxford: Elsevier; 2010.

Exergy destruction rate (kW)

80,000

71,415

70,000 60,000 50,000 40,000 26,570

30,000

19,270

20,000

13,850

8475

10,000

3250

0 Gasification Double stage Hydrogen system ORC production system

SOFC

Absorption cooling

Integrated system

Fig. 45 Exergy destruction rates for the biomass gasification and solid oxide fuel cell (SOFC) based integrated system. ORC, organic Rankine cycle.

The factor b can be defined depending on the composition of biomass resources as follows [70]: b¼

1:044 þ 0:016ðH=CÞ

0:3493ðO=CÞð1 þ 0:0531ðH=CÞÞ þ 0:0493ðN=CÞ ð1 0:4124ðO=CÞÞ

ð944Þ

The LHV of generated syngas can be defined as the function of syngas combinations heating values as given below: LHV gas ¼

n  X

yi LHV i

1



ð945Þ

gas

where LHV gas is illustrated as the lower heating value in molar form. The chemical exergy of produced syngas at any state is written as ex ch gas ¼

n  X 1

yi ex ch i þ RT0 yi lnyi



gas

ð946Þ

where yi is the function of chemical exergy and molar fractions of all the species of generated syngas.

4.10.7.2.4

Parametric studies

The exergy destruction rates occurring in the five subsystems and whole integrated process during operation are shown in Fig. 45. The main exergy destruction occurs in the gasification system where chemical and electrochemical reactions are considered in the biomass gasifier and the SOFC subsystem, including high temperature exhaust gaseous. Even with the usage of the heat content of exhaust gaseous for steam generation for gasification cycle, still the gasification subsystem has the most exergy destruction rate in the IGCC. The thermodynamic analysis results illustrated that the cooling subsystem does not show important exergy destruction rate, principally this subsystem does not directly use the fuel energy but instead utilizes waste heat produced by the gasification process. Fig. 46 illustrates the dimensionless exergy destruction ratio for each component of IGCC. This thermodynamic variable is beneficial for prioritizing irreversibility in the intuitive manner. Both exergy destruction rate and exergy destruction ratio are higher in the gasification system than in other system parts, suggesting that it would likely be worthwhile to focus improvement efforts on this integrated system part. Also, these outcomes illustrate that the absorption cooling process does not exhibit important exergy destruction ratio.

452

Integrated Gasification Combined Cycles

Exergy destruction ratio (%)

40

37.2

35 30

26.98

25 19.39

20 15

11.87

10 4.55

5 0 Gasification system

Double stage ORC

Hydrogen production system

SOFC

Absorption cooling

Fig. 46 Exergy destruction ratio for the biomass gasification and solid oxide fuel cell (SOFC) based integrated system. ORC, organic Rankine cycle.

60

54.18

Exergy efficiency (%)

52.73 50 37.56

40 30

36.43

26.35

20

13.93

10 0 Gasification Double stage system ORC

Hydrogen production system

SOFC

Absorption cooling

Integrated system

Fig. 47 Exergy efficiency for the subsystems of the biomass gasification and solid oxide fuel cell (SOFC) based integrated system. ORC, organic Rankine cycle.

The exergy efficiencies of each subsystem of the biomass gasification and SOFC based integrated system are shown in Fig. 47. As it is observed, the exergy efficiency of whole system is higher than all other subsystems, because the IGCC has higher useful outputs performance values than single output systems. According to the results shown in Fig. 47, the biomass gasification subsystem has the highest exergy efficiency among all subsystems. The single effect absorption cooling system has the lowest exergy efficiency, mainly due to the temperature differences of working fluid streams, and also due to the pressure drops in the cooling system components. Fig. 48 illustrated that the impact of absorption chiller evaporator temperature both on energetic COPen and exergetic COPex. According to analysis results and Fig. 48, as absorption chiller evaporator temperature increases from 5 to 201C, COPen increases from about 0.76 to 0.851C. On the other hand, with the same temperature change, COPex decreases from about 0.26 to 0.141C. Fig. 49 illustrates the effect of reference temperature both on exergy destruction rate on the left side and exergy efficiency of system at the right side of the figure. According to this Fig. 49, as reference temperature increases from 0 to 401C, the exergy destruction rate of the system decreases, and the exergy efficiency of integrated system increase with increasing reference temperature. The biomass gasifier temperature is a significant indicator that affects the efficiency of the integrated gasification combined process. Changing the biomass gasifier temperature affects a composition of generated synthesis gaseous and an accompanied exergy rate. This temperature changes affect the number of fuel cell stacks needed to cover the required output electricity. The biomass gasifier temperature is varied in reasonable operating range and also the parametric analyses are made to examine the impact of gasification temperature on the integrated process efficiency. Fig. 50 illustrates the impact of biomass gasifier temperature on the efficiency of integrated biomass process. As seen from Fig. 50, when gasifier temperature increasing from 700 to 9001C, the gasification performance creates the linear increase in the energy efficiency from 56.87% to 64.72% and the exergy efficiency from 53.19% to 61.23%, respectively.

Integrated Gasification Combined Cycles

453

1 0.9

COPen and COPex

0.8 0.7 0.6

COPen

0.5

COPex

0.4 0.3 0.2 0.1 0 5

10 15 Absorption chiller evaporator temperature (°C)

20

Fig. 48 Effect of absorption chiller evaporator temperature on COPen and COPex. COP, coefficient of performance.

0.58

77,000 ExD,system

0.57

System

75,000

0.56

74,000

0.55

73,000

0.54

72,000

0.53

71,000

0.52

70,000

0.51

69,000

0.5

68,000

0

5

10

15 20 25 Reference temperature (°C)

30

35

Exergy efficiency

Exergy destruction rate (kW)

76,000

0.49 40

Fig. 49 Effect of reference temperature on exergy destruction rate and exergy efficiency of integrated gasification combined system.

Fig. 51 shows the effect of SOFC working temperature on the power generation rate and exergy efficiency of SOFC subcomponent. It is observed that an increase in SOFC working temperature results in an increase in power generation rate and exergy efficiency of SOFC. These outputs are provided because increase in the SOFC temperature generates the decrease in heat demand. This is applied feasible by the increase in electrical energy as the SOFC temperature increases. Also, the total energy need reduces in higher SOFC subsystem temperature cases. To supply environmental comprehensions, the environmental effects of single generation, cogeneration, trigeneration, and multigeneration are investigated, and the results are illustrated in Fig. 52. It is observed that the multigeneration process has less carbon dioxide emissions than the singe generation, cogeneration and trigeneration processes, providing an important motivating force for the usage of multigeneration processes. In addition to that, the multigeneration process investigated here has the higher exergy efficiency than other generation processes.

4.10.7.3

Integrated Gasification Combined System With Coal Gasification and Hydrogen Liquefaction

The coal gasification process is the thermochemical conversion of coal sources into the synthesis gas (or syngas) composed primarily of hydrogen and carbon monoxide. Unlike combustion cycles that only generate carbon dioxide and water, gasification is a partial oxidation cycle that occurs in oxygen-limited surroundings. The producing synthesis gaseous are more beneficial than combustion exhaust gaseous and it has the potential to produce power more efficiently and cleanly. Over the last two decades, the coal gasification cycles have been used to convert coal sources into fueled-gas to produce power and domestic heating applications.

Integrated Gasification Combined Cycles

0.62

0.65 System

0.64 0.63 Energy efficiency

0.61

System

0.6

0.62

0.59

0.61

0.58

0.6

0.57

0.59

0.56

0.58

0.55

0.57

0.54

0.56 700

725

750 775 800 825 850 Biomass gasifier temperature (°C)

875

Exergy efficiency

454

0.53 900

Fig. 50 Effect of biomass gasifier temperature on energy and exergy efficiency of integrated gasification combined system.

15.75

0.4 SOFC

15.5

0.39 0.38

15.25 0.37 15 0.36 14.75

Exergy efficiency

SOFC output power (kW)

WSOFC

0.35

14.5 700

705

710

715 720 725 730 735 SOFC temperature (°C)

740

745

0.34 750

Fig. 51 Effect of solid oxide fuel cell (SOFC) temperature on output power from SOFC and exergy efficiency of SOFC.

Environmental impacts 500

451.3

402.5

400

342.6

300

232.8

200 100

26.35

19.47

32.74

16.82

45.93

13.68

54.18 9.24

0 Single generation

Cogeneration

Exergy efficiency (%)

Trigeneration

Multigeneration

Carbon dioxide emissions (kg/kWh)

Carbon monoxide emissions (kg/kWh) Fig. 52 Comparison of exergy efficiency, unit CO2 and CO emissions of generation cycles for biomass gasification and solid oxide fuel cell (SOFC) based integrated system.

Integrated Gasification Combined Cycles

Heating application

19

20 Condenser-I

16

455

17 Pump-I

Exhaust 12 gaseous

HPST

LPST

18

Electricity HRSG

13 15

34 11 30

21 Coal Syngas storage

23

24

LPC 1 Air Water

HPC 2

Pump-II 28

5 3-way valve I 4

35

32 Expansion valve-II 33

29

Absorber

42

Additional coal when required

39 38

31 HEX-I

Char Fluidized Coal bed gasification combustor

Compressed 22 air

Condenser-II

14

37

Expansion valve-I 36 40 Evaporator

District cooling

43 Electricity LPGT

HPGT

3

6

CC-I

7

8

41

CC-II

Intercooler

9 10

44

26

27

25

HEX-II

3-way valve II Electricity Mixer 45

47 PEM electrolyzer

48

Compressor 49

N2(gas)

50

HEX-IV 61 60 N2(liq)

HEX-III

46 Oxygen

51

58

59

HEX-V

57

52 N2(gas)

56 Liquid hydrogen tank

64

HEX-VI

HEX-VII

Separator 55 54 Expansion valve-III

53

63 62 N2(liq)

Fig. 53 Schematic diagram of integrated gasification combined cycle (IGCC) based on coal fuel. HEX, heat exchanger; HPC, high pressure compressor; HPGT, high pressure gas turbine; HPST, high pressure steam turbine; HRSG, heat recovery steam generator; LPC, low pressure compressor; LPGT, low pressure gas turbine; LPST, low pressure steam turbine; PEM, proton exchange membrane.

The schematic diagram of partial gasification-based IGCC is illustrated in Fig. 53. The integrated gasification combined system investigated in this study utilizes coal gasification to convert coal samples into the synthesis gas which is burned in the combustion chamber-I creating high temperature pressurized gaseous mix which expands in the HP gas turbines for power generation. The reheated low-pressure synthesis gaseous mixture enters the low-pressure gas turbine at point 7 for extra power generation. The exhaust gases leaving from low-pressure gas turbine enters the fluidized bed combustor at point 10 to increase its temperature before entering the HRSG subsystem. The producing high temperature gaseous mixture enters through the HRSG at point 11, the generating high temperature steam that enters the HP steam turbine at point 13 and low-pressure steam turbine at point 15 in the

456

Integrated Gasification Combined Cycles

bottoming process, respectively. The additional fluidized bed combustor has three aims, such as (1) reheat the exhaust gas exiting from low pressure gas turbine (from point 10 to 11), (2) preheat the compressed air exiting from HP compressor before air goes to the combustion chamber-I (from point 5 to 6), and (3) reheat the low-pressure steam in the steam process (from point 14 to 15).

4.10.7.3.1

Balance equations

Applying thermodynamic assessments to integrated gasification combined processes for multigeneration can supply a better understanding of their behaviors and enhanced steps for developing them. The general balance equations for coal gasificationbased IGCC components are defined in this subsection as given below. Coal gasifier: the mass, energy, entropy, and exergy balance equations of coal gasifier under steady state and steady flow conditions are written as follows: _ 22 ¼ m _ 23 þ m _ char Mass: m _ 21 þ m

ð947Þ

_ 22 h22 ¼ m _ 23 h23 Energy: m _ 21 h21 þ m

ð948Þ

_ 22 s22 þ S_ gen;cg ¼ m _ 23 s23 Entropy: m _ 21 s21 þ m

ð949Þ

_ _ 22 ex 22 ¼ m _ 23 ex ch Exergy: m _ 21 ex ch 21 þ m 23 þ Ex D;cg

ð950Þ

The chemical properties of coal sample can be calculated by using the equations given in Section 4.10.4.6. In the following equations, the chemical exergy calculations for synthesis gas are given as follows: ch ch ch exch 23 ¼ rH2 MH2 ex H2 þ rCO MCO ex CO þ rCH4 MCH4 ex CH4

 0  0 0 ex ch CO ¼ hCO hC þ 0:5hO2

 0  0 0 ex ch CO2 ¼ hCO2 hC þ hO2

 0  0 0 ex ch CH4 ¼ hCH4 hC þ 2hH2

 T0 s0CO

 T0 s0CO

 T0 s0CO

  ch þ ex ch s0C þ 0:5s0O2 C þ 0:5ex O2 



s0C þ s0O2



s0C þ 2s0H2

ch þ ex ch C þ ex O2



ch þ ex ch C þ 2ex H2

ð951Þ

ð952Þ ð953Þ

ð954Þ

Low pressure compressor: the mass, energy, entropy, and exergy balance equations for low pressure compressor subcomponent can be defined under the steady state and steady flow conditions. _2 Mass: m _1 ¼m

ð955Þ

_ LPC ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð956Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;LPC ¼ m

ð957Þ

_ LPC ¼ m _ D;LPC _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð958Þ

Intercooler subcomponent: the mass, energy, entropy, and exergy balance equations of intercooler subcomponent under steady state and steady flow conditions can be defined as given below: _3 Mass: m _2 ¼m

ð959Þ

_ Loss ¼ m _ 3 h3 Energy: m _ 2 h2 þ Q

ð960Þ

_ Loss =TLoss þ S_ gen;ic ¼ m _ 3 s3 Entropy: m _ 2 s2 þ Q

ð961Þ

_ Q ¼m _ D;ic _ 3 ex3 þ Ex Exergy: m _ 2 ex 2 þ Ex Loss

ð962Þ

HP compressor: under the steady state and steady flow conditions, the balance equations for HP compressor subcomponent can be written as follows:

Integrated Gasification Combined Cycles

457

_4 Mass: m _3 ¼m

ð963Þ

_ HPC ¼ m _ 4 h4 Energy: m _ 3 h3 þ W

ð964Þ

_ 4 s4 Entropy: m _ 3 s3 þ S_ gen;HPC ¼ m

ð965Þ

_ HPC ¼ m _ D;HPC _ 4 ex 4 þ Ex Exergy: m _ 3 ex3 þ W

ð966Þ

Three-way valve-I: the mass, energy, entropy, and exergy balance equations are defined for three-way valve-I under steady state and steady flow conditions. _5þm _ 21 Mass: m _4¼m

ð967Þ

_ 5 h5 þ m _ 21 h21 Energy: m _ 4 h4 ¼ m

ð968Þ

_ 5 s5 þ m _ 21 s21 Entropy: m _ 4 s4 þ S_ gen;3wv_I ¼ m

ð969Þ

_ D;3wv_I _ 5 ex 5 þ m _ 21 ex 21 þ Ex Exergy: m _ 4 ex 4 ¼ m

ð970Þ

Fluidized bed combustor: the mass, energy, entropy, and exergy balance equations are written for fluidized bed combustor under the steady state and steady flow conditions. _ 6; m _ 11 ; m _ 15 Mass: m _5¼m _ 10 ¼ m _ 14 ¼ m

ð971Þ

_ 10 h10 þ m _ 14 h14 ¼ m _ 6 h6 þ m _ 11 h11 þ m _ 15 h15 Energy: m _ 5 h5 þ m

ð972Þ

_ 10 s10 þ m _ 14 s14 þ S_ gen;fbc ¼ m _ 6 s6 þ m _ 11 s11 þ m _ 15 s15 Entropy: m _ 5 s5 þ m

ð973Þ

_ D;fbc _ 10 ex10 þ m _ 14 ex14 ¼ m _ 6 ex 6 þ m _ 11 ex 11 þ m _ 15 ex15 þ Ex Exergy: m _ 5 ex5 þ m

ð974Þ

HP steam turbine: the mass, energy, entropy, and exergy balance equations are written for HP steam turbine under the steady state and steady flows conditions. _ 14 Mass: m _ 13 ¼ m

ð975Þ

_ HPST _ 14 h14 þ W Energy: m _ 13 h13 ¼ m

ð976Þ

_ 14 s14 Entropy: m _ 13 s13 þ S_ gen;HPST ¼ m

ð977Þ

_ HPST þ Ex _ D;HPST _ 14 ex14 þ W Exergy: m _ 13 ex13 ¼ m

ð978Þ

Low pressure steam turbine: the balance equations are written for low pressure steam turbine under the steady state and steady flows conditions. _ 16 Mass: m _ 15 ¼ m

ð979Þ

_ LPST _ 16 h16 þ W Energy: m _ 15 h15 ¼ m

ð980Þ

_ 16 s16 Entropy: m _ 15 s15 þ S_ gen;LPST ¼ m

ð981Þ

_ LPST þ Ex _ D;LPST _ 16 ex 16 þ W Exergy: m _ 15 ex 15 ¼ m

ð982Þ

Condenser-I: the mass, energy, entropy, and exergy balance equations can be defined for condenser-I under the steady state and steady flow conditions as follows: _ 17 ; m _ 20 Mass: m _ 16 ¼ m _ 19 ¼ m _ con _ 19 h19 ¼ m _ 17 h17 þ m _ 20 h20 þ Q Energy: m _ 16 h16 þ m

ð983Þ I

ð984Þ

458

Integrated Gasification Combined Cycles

_ 19 s19 þ S_ gen;con Entropy: m _ 16 s16 þ m

I

_ 17 s17 þ m _ 20 s20 ¼m

_ D;con _ 19 ex 19 ¼ m _ 17 ex 17 þ m _ 20 ex 20 þ Ex Exergy: m _ 16 ex16 þ m

ð985Þ

I

ð986Þ

Pump-I: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for pump-I can be defined as _ 18 Mass: m _ 17 ¼ m _p Energy: m _ 17 h17 þ W

I

Entropy: m _ 17 s17 þ S_ gen;p _p Exergy: m _ 17 ex 17 þ W

I

ð987Þ

_ 18 h18 ¼m

I

ð988Þ

_ 18 s18 ¼m

_ D;p _ 18 ex 18 þ Ex ¼m

ð989Þ

I

ð990Þ

Combustion chamber-I: under steady state and steady flow condition, the mass, energy, entropy, and exergy balance equations of combustion chamber-I are written as follows: _ 26 ¼ m _7 Mass: m _6þm

ð991Þ

_ 26 h26 ¼ m _ 7 h7 Energy: m _ 6 h6 þ m

ð992Þ

_ 26 s26 þ S_ gen;cc Entropy: m _ 6 s6 þ m

I

_ 7 s7 ¼m

_ _ 26 exch _ 7 ex ch Exergy: m _ 6 ex6 þ m 26 ¼ m 7 þ Ex D;cc

ð993Þ

I

ð994Þ

HP gas turbine: the mass, energy, entropy, and exergy balance equations are written for HP gas turbine under the steady state and steady flows conditions. _8 Mass: m _7 ¼m ð995Þ _ HPGT _ 8 h8 þ W Energy: m _ 7 h7 ¼ m

ð996Þ

_ 8 s8 Entropy: m _ 7 s7 þ S_ gen;HPGT ¼ m

ð997Þ

_ HPGT þ Ex _ D;HPGT _ 8 ex 8 þ W Exergy: m _ 7 ex 7 ¼ m

ð998Þ

Combustion chamber-II: under steady state and steady flow condition, the balance equations of combustion chamber-II can be written as _ 27 ¼ m _9 Mass: m _8þm

ð999Þ

_ 27 h27 ¼ m _ 9 h9 Energy: m _ 8 h8 þ m

ð1000Þ

_ 27 s27 þ S_ gen;cc Entropy: m _ 8 s8 þ m

II

_ 9 s9 ¼m

_ _ 27 ex ch _ 9 ex ch Exergy: m _ 8 ex 8 þ m 27 ¼ m 9 þ Ex D;cc

ð1001Þ

II

ð1002Þ

Low pressure gas turbine: the balance equations are written for low pressure gas turbine under the steady state and steady flows conditions. _ 10 Mass: m _9¼m

ð1003Þ

_ LPGT _ 10 h10 þ W Energy: m _ 9 h9 ¼ m

ð1004Þ

_ 10 s10 Entropy: m _ 9 s9 þ S_ gen;LPGT ¼ m

ð1005Þ

Integrated Gasification Combined Cycles

_ LPGT þ Ex _ D;LPGT _ 10 ex 10 þ W Exergy: m _ 9 ex 9 ¼ m

459

ð1006Þ

HRSG: the mass, energy, entropy, and exergy balance equations are written for HRSG under the steady state and steady flow conditions as follows: _ 12 ; m _ 18 ; m _ 31 þ m _ 34 Mass: m _ 11 ¼ m _ 13 ¼ m _ 30 ¼ m

ð1007Þ

_ HRSG ¼ m _ 18 h18 þ m _ 30 h30 þ Q _ 12 h12 þ m _ 13 h13 þ m _ 31 h31 þ m _ 34 h34 Energy: m _ 11 h11 þ m

ð1008Þ

_ HRSG =THRSG þ S_ gen;HRSG ¼ m _ 18 s18 þ m _ 30 s30 þ Q _ 12 s12 þ m _ 13 s13 þ m _ 31 s31 Entropy: m _ 11 s11 þ m

ð1009Þ

_ Q _ D;HRSG _ 18 ex 18 þ m _ 30 ex 30 þ Ex _ 12 ex12 þ m _ 13 ex13 þ m _ 31 ex 31 þ Ex Exergy: m _ 11 ex11 þ m HRSG ¼ m

ð1010Þ

Condenser-II: the balance equations can be defined for condenser-II under the steady state and steady flow conditions as _ 35 ; m _ 39 Mass: m _ 34 ¼ m _ 38 ¼ m

ð1011Þ

_ con _ 38 h38 ¼ m _ 35 h35 þ m _ 39 h39 þ Q Energy: m _ 34 h34 þ m _ 38 s38 þ S_ gen;con Entropy: m _ 34 s34 þ m

II

_ con _ 35 s35 þ m _ 39 s39 þ Q ¼m

_ Q _ 38 ex 38 ¼ m _ 35 ex 35 þ m _ 39 ex 39 þ Ex Exergy: m _ 34 ex 34 þ m con

II

ð1012Þ

II

II =Tcon II

_ D;con þ Ex

II

ð1013Þ ð1014Þ

Expansion valve-I: the mass, energy, entropy, and exergy balance equations for expansion valve-I can be written under the steady state and steady flow conditions as follows: _ 36 Mass: m _ 35 ¼ m

ð1015Þ

_ 36 h36 Energy: m _ 35 h35 ¼ m

ð1016Þ

_ 36 s36 Entropy: m _ 35 s35 þ S_ gen;ev_I ¼ m

ð1017Þ

_ D;ev_I _ 36 ex36 þ Ex Exergy: m _ 35 ex35 ¼ m

ð1018Þ

Evaporator: the mass, energy, entropy, and exergy balance equations can be defined for evaporator under the steady state and steady flow conditions as follows: _ 37 ; m _ 41 Mass: m _ 36 ¼ m _ 40 ¼ m

ð1019Þ

_ eva ¼ m _ 40 h40 þ Q _ 37 h37 þ m _ 41 h41 Energy: m _ 36 h36 þ m

ð1020Þ

_ eva =Teva þ S_ gen;eva ¼ m _ 40 s40 þ Q _ 37 s37 þ m _ 41 s41 Entropy: m _ 36 s36 þ m

ð1021Þ

_ Q ¼m _ D;eva _ 40 ex40 þ Ex _ 37 ex37 þ m _ 41 ex41 þ Ex Exergy: m _ 36 ex36 þ m eva

ð1022Þ

Absorber: the mass, energy, entropy, and exergy balance equations can be given for absorber under the steady state and steady flow conditions as follows: _ 33 þ m _ 37 ; m _ 43 Mass: m _ 28 ¼ m _ 42 ¼ m

ð1023Þ

_ abs _ 37 h37 þ m _ 42 h42 ¼ m _ 28 h28 þ m _ 43 h43 þ Q Energy: m _ 33 h33 þ m

ð1024Þ

_ abs =Tabs _ 37 s37 þ m _ 42 s42 þ S_ gen;abs ¼ m _ 28 s28 þ m _ 43 s43 þ Q Entropy: m _ 33 s33 þ m

ð1025Þ

_ D;abs þ Ex _ Q _ 37 ex 37 þ m _ 42 ex42 ¼ m _ 28 ex 28 þ m _ 43 ex43 þ Ex Exergy: m _ 33 ex 33 þ m abs

ð1026Þ

460

Integrated Gasification Combined Cycles

Pump-II: for the pump-II of single effect absorption cooling system, the balance equations are provided under the steady state and steady flow conditions as follows: _ 29 Mass: m _ 28 ¼ m _p Energy: m _ 28 h28 þ W

_ 29 h29 ¼m

II

Entropy: m _ 28 s28 þ S_ gen;p _p Exergy: m _ 28 ex 28 þ W

II

ð1027Þ

II

ð1028Þ

_ 29 s29 ¼m

_ D;p _ 29 ex 29 þ Ex ¼m

ð1029Þ ð1030Þ

II

Expansion valve-II: under the steady state and steady flow conditions, the balance equations for expansion valve-II can be written as _ 33 Mass: m _ 32 ¼ m

ð1031Þ

_ 33 h33 Energy: m _ 32 h32 ¼ m

ð1032Þ

_ 33 s33 Entropy: m _ 32 s32 þ S_ gen;ev_II ¼ m

ð1033Þ

_ D;ev_II _ 33 ex33 þ Ex Exergy: m _ 32 ex32 ¼ m

ð1034Þ

HEX-I: the mass, energy, entropy, and exergy balance equations for HEX-I can be expressed under the steady state and steady flow conditions as follows: _ 30 ; m _ 32 Mass: m _ 29 ¼ m _ 31 ¼ m _ HEX _ 31 h31 þ Q Energy: m _ 29 h29 þ m

I

_ 31 s31 þ S_ gen;HEX Entropy: m _ 29 s29 þ m

ð1035Þ

_ 30 h30 þ m _ 32 h32 ¼m

I

ð1036Þ

_ 30 s30 þ m _ 32 s32 ¼m

_ D;HEX _ 31 ex31 ¼ m _ 30 ex 30 þ m _ 32 ex32 þ Ex Exergy: m _ 29 ex 29 þ m

ð1037Þ I

ð1038Þ

HEX-II: the balance equations for HEX-II can be expressed under the steady state and steady flow conditions as _ 25 ; m _ 45 Mass: m _ 24 ¼ m _ 44 ¼ m

ð1039Þ

_ 44 h44 ¼ m _ 25 h25 þ m _ 45 h45 Energy: m _ 24 h24 þ m

ð1040Þ

_ 44 s44 þ S_ gen;HEX Entropy: m _ 24 s24 þ m

II

_ 25 s25 þ m _ 45 s45 ¼m

_ D;HEX _ 44 ex 44 ¼ m _ 25 ex 25 þ m _ 45 ex 45 þ Ex Exergy: m _ 24 ex 24 þ m

ð1041Þ

II

ð1042Þ

PEM electrolyzer: under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for PEM electrolyzer is written as _ 46 þ m _ 47 Mass: m _ 45 ¼ m

ð1043Þ

_ PEM ¼ m _ 46 h46 þ m _ 47 h47 Energy: m _ 45 h45 þ W

ð1044Þ

_ 46 s46 þ m _ 47 s47 Entropy: m _ 45 s45 þ S_ gen;PEM ¼ m

ð1045Þ

_ PEM ¼ m _ D;PEM _ 46 ex46 þ m _ 47 ex47 þ Ex Exergy: m _ 45 ex45 þ W

ð1046Þ

Mixer: the balance equations for mixer can be written under the steady state and steady flow conditions as follows: _ 59 ¼ m _ 48 Mass: m _ 47 þ m

ð1047Þ

_ 59 h59 ¼ m _ 48 h48 Energy: m _ 47 h47 þ m

ð1048Þ

Integrated Gasification Combined Cycles

461

_ 59 s59 þ S_ gen;mixer ¼ m _ 48 s48 Entropy: m _ 47 s47 þ m

ð1049Þ

_ D;mixer _ 59 ex 59 ¼ m _ 48 ex 48 þ Ex Exergy: m _ 47 ex 47 þ m

ð1050Þ

Compressor: the mass, energy, entropy, and exergy balance equations for compressor subcomponent can be defined written under the steady state and steady flow conditions as follows: _ 49 Mass: m _ 48 ¼ m

ð1051Þ

_ cmp ¼ m _ 49 h49 Energy: m _ 48 h48 þ W

ð1052Þ

_ 49 s49 Entropy: m _ 48 s48 þ S_ gen;cmp ¼ m

ð1053Þ

_ cmp ¼ m _ D;cmp _ 49 ex 49 þ Ex Exergy: m _ 48 ex 48 þ W

ð1054Þ

HEX-III: under the steady state and steady flow conditions, the balance equations for HEX-III can be expressed as _ 50 ; m _ 59 _ 58 ¼ m Mass: m _ 49 ¼ m

ð1055Þ

_ 58 h58 ¼ m _ 50 h50 þ m _ 59 h59 Energy: m _ 49 h49 þ m

ð1056Þ

_ 58 s58 þ S_ gen;HEX Entropy: m _ 49 s49 þ m

III

_ 50 s50 þ m _ 59 s59 ¼m

_ D;HEX _ 58 ex58 ¼ m _ 50 ex50 þ m _ 59 ex 59 þ Ex Exergy: m _ 49 ex49 þ m

ð1057Þ

III

ð1058Þ

HEX-IV: under the steady state and steady flow conditions, the balance equations for HEX-IV are _ 51 ; m _ 61 Mass: m _ 50 ¼ m _ 60 ¼ m

ð1059Þ

_ 60 h60 ¼ m _ 51 h51 þ m _ 61 h61 Energy: m _ 50 h50 þ m

ð1060Þ

_ 60 s60 þ S_ gen;HEX Entropy: m _ 50 s50 þ m

IV

_ 51 s51 þ m _ 61 s61 ¼m

_ D;HEX _ 60 ex 60 ¼ m _ 51 ex 51 þ m _ 61 ex 61 þ Ex Exergy: m _ 50 ex50 þ m

ð1061Þ

IV

ð1062Þ

HEX-V: under the steady state and steady flow conditions, the balance equations for HEX-V are written as _ 52 ; m _ 58 Mass: m _ 51 ¼ m _ 57 ¼ m

ð1063Þ

_ 57 h57 ¼ m _ 52 h52 þ m _ 58 h58 Energy: m _ 51 h51 þ m

ð1064Þ

_ 57 s57 þ S_ gen;HEX Entropy: m _ 51 s51 þ m

V

_ 52 s52 þ m _ 58 s58 ¼m

_ D;HEX _ 57 ex 57 ¼ m _ 52 ex 52 þ m _ 58 ex 58 þ Ex Exergy: m _ 51 ex 51 þ m

ð1065Þ

V

ð1066Þ

HEX-VI: under the steady state and steady flow conditions, the balance equations for HEX-VI can be defined as _ 53 ; m _ 63 Mass: m _ 52 ¼ m _ 62 ¼ m

ð1067Þ

_ 62 h62 ¼ m _ 53 h53 þ m _ 63 h63 Energy: m _ 52 h52 þ m

ð1068Þ

_ 62 s62 þ S_ gen;HEX Entropy: m _ 52 s52 þ m

VI

_ 53 s53 þ m _ 63 s63 ¼m

_ D;HEX _ 62 ex 62 ¼ m _ 53 ex53 þ m _ 63 ex 63 þ Ex Exergy: m _ 52 ex52 þ m

ð1069Þ

VI

ð1070Þ

HEX-VII: under the steady state and steady flow conditions, the balance equations for HEX-IV are _ 54 ; m _ 57 _ 56 ¼ m Mass: m _ 53 ¼ m

ð1071Þ

462

Integrated Gasification Combined Cycles

_ 56 h56 ¼ m _ 54 h54 þ m _ 57 h57 Energy: m _ 53 h53 þ m _ 56 s56 þ S_ gen;HEX Entropy: m _ 53 s53 þ m

VII

ð1072Þ

_ 54 s54 þ m _ 57 s57 ¼m

_ D;HEX _ 56 ex56 ¼ m _ 54 ex54 þ m _ 57 ex57 þ Ex Exergy: m _ 53 ex53 þ m

ð1073Þ

VII

ð1074Þ

Expansion valve-III: the balance equations for expansion valve-III are defined under the steady state and steady flow conditions as follows: _ 55 Mass: m _ 54 ¼ m

ð1075Þ

_ 55 h55 Energy: m _ 54 h54 ¼ m

ð1076Þ

Entropy: m _ 54 s54 þ S_ gen;ev

III

_ 55 s55 ¼m

_ D;ev _ 55 ex55 þ Ex Exergy: m _ 54 ex54 ¼ m

III

ð1077Þ

ð1078Þ

Separator: the mass, energy, entropy, and exergy balance equations for separator can be defined under the steady state and steady flow conditions as follows:

4.10.7.3.2

_ 56 þ m _ 64 Mass: m _ 55 ¼ m

ð1079Þ

_ 56 h56 þ m _ 64 h64 Energy: m _ 55 h55 ¼ m

ð1080Þ

_ 56 s56 þ m _ 64 s64 Entropy: m _ 55 s55 þ S_ gen;sep ¼ m

ð1081Þ

_ D;sep _ 56 ex 56 þ m _ 64 ex 64 þ Ex Exergy: m _ 55 ex 55 ¼ m

ð1082Þ

Properties of coal samples

One of the most significant indicators for IGCC investigation is the equivalence ratio, which is the dimensionless variable that illustrates the real fuel/oxidant ratio normalized by the stoichiometric fuel/oxidant ratio as given below: eER ¼

F=Oa F=Os

ð1083Þ

where eER is the equivalence ratio, F is the fuel, O is the oxidant, subscripts a and s represent the actual and stoichiometric case, respectively. If ER is equal to 1, this case represents the stoichiometric combustion condition. The mixture of syngas can be categorized as lean, where excess oxidant exists and hence ER is lower than 1. Moreover, if excess fuel is present, the mixtures are categorized as rich, and hence ER is higher than 1. The LHV of syngas can be calculated as follows, which results in units of kJ/m3 [71]. LHV sg ¼

282:99 VCO þ 802:34 VCH4 þ 241:83 VH2 0:1  22:4

ð1084Þ

The chemical characteristics and chemical exergy of coal samples can be defined by using the approaches illustrated in Section 4.10.4.6. The ultimate analysis and proximate analysis of coal samples employed in coal gasifier are illustrated in Tables 18 and 19, respectively, where their calorific values are given in Table 20.

4.10.7.3.3

Hydrogen liquefaction system

The Linde Hampson hydrogen liquefaction system is integrated to liquefy of the produced hydrogen from integrated gasification combined process. In this chapter, the mass flow rate of liquefied hydrogen is kept constant value. Based on this state, the yield of the liquid phase hydrogen can be calculated as follows [72]: y ¼ ðh57 h50 ¼ h49

h53 Þ=ðh57 ð1

h64 Þ

ð1085Þ

yÞðh59

h58 Þ

ð1086Þ

h59 ¼ h58 þ A HEX ðh590

h58 Þ

ð1087Þ

h52 ¼ h51

h57 Þ

ð1088Þ

ð1

yÞðh58

Integrated Gasification Combined Cycles

Table 18

Ultimate analysis of coal sample employed in coal gasifier

Compound

Dry basis (wt%)

Dry ash-free basis (wt%)

Carbon Hydrogen Nitrogen Sulfur Oxygen

73.85 3.65 2.76 1.42 1.86

88.40 4.37 3.30 1.70 2.23

Table 19 gasifier

463

Proximate analysis of coal sample employed in coal

Compound

Original basis (wt%)

Dry basis (wt%)

Moisture Volatile matter Fixed carbon Ash

5.65 32.86 45.96 15.53

– 34.83 48.71 16.46

Table 20 gasifier

Calorific values of coal sample employed in coal

Heating value

Original basis (kJ/kg) Dry basis (kJ/kg)

Lower heating value (LHV) 24,974 Higher heating value (HHV) 26,004

26,589 27,543

h58 ¼ h57 þ A HEX ðh580  h57 ¼ hg þ A HEX h570

h54 ¼ h53

ð1

h57 Þ

ð1089Þ



ð1090Þ

hg

 yÞ h56

hg



ð1091Þ

here A HEX is the HEX effectiveness indicator, and it changes between 0.85 and 1. For an assumed efficiency, it is possible to determine the entalpy values of stream points which leads to calculation of compression power requirement for hydrogen and nitrogen mass flow rates. The work requirement of nitrogen is assumed as 7760 kJ/kg N2.

4.10.7.3.4

Thermodynamic assessment of integrated system

In this subsection, thermodynamic assessment of integrated gasification combined process is analyzed through energetic and exergetic analysis viewpoint. 4.10.7.3.4.1 Gasification reaction The chemical reaction of coal gasification process can be defined as follows [73]: Air

zfflfflfflfflfflfflfflfflfflfflfflffl}|fflfflfflfflfflfflfflfflfflfflfflffl{ Coal zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl}|fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{ þ FaG ðO2 þ 3:76N2 Þ ðMC C þ MH2 H2 þ MN2 N2 þ MO2 O2 þ MS S þ Mash Þ Steam

zfflfflffl}|fflfflffl{   þ FstRG ðH2 OÞ -FFg FCO2 CO2 þ FCO CO þ FpH2 H2 þ FCH4 CH4 þFstPG H2 O þ FN2 PG N2 þ MCch C þðMS S þ Mash Þ |{z} |fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl} Fuel gas

ð1092Þ

Char

4.10.7.3.4.2 Combustion reaction The combustion reaction for combustor chamber-I can be written as   FFg1 FCO2 CO2 þ FCO CO þ FpH2 H2 þ FCH4 CH4 þ FN2PG N2 þ FstPG H2 O þ MaGC ðO2 þ 3:76N2 Þ -FCO2 PGC CO2 þ FstPGC H2 O þ FN2PGC N2 þ FO2PGC O2

ð1093Þ

464

Integrated Gasification Combined Cycles

The combustion reaction for combustor chamber-II for all cases can be defined as follows:   FFg2 FCO2 CO2 þ FCO CO þ FpH2 H2 þ FCH4 CH4 þ FN2PG N2 þ FstPG H2 O þ ðFCO2 PGC CO2 þ FstPGC H2 O þ FN2PGC N2 þ FO2 PGC O2 Þ -FCO2 PRH CO2 þ FstPRH H2 O þ FN2PRH N2 þ FO2PRH O2 ð1094Þ The combustion reaction for char combustor for all cases is 0 1 Additional coal zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl}|fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{ @MCCC C þ MH2 CC H2 þ MN2 CC N2 þ MO2 CC O2 A þ MCch ðCÞ

ð1095Þ

þ ðFCO2 PRH CO2 þ FstPRH H2 O þ FN2 PRH N2 þ FO2 PRH O2 Þ þ MaCC ðO2 þ 3:76N2 Þ -ðFCO2 PCC CO2 þ FstPCC H2 O þ FN2 PCC N2 þ FO2 PCC O2 Þ

4.10.7.3.4.3 Gas cycle process efficiency The power production rate from gas cycle process is given as follows: _ GC ¼ m _ 1 ðh2 W

_ 3 ðh4 h1 Þ þ m

h3 Þ

_ 7 ðh7 m

h8 Þ

_ 9 ðh9 m

h10 Þ

ð1096Þ

The heat supplied rate of gas cycle process can be defined as follows: _ GC ¼ QL ½m _ 7 h7 Q

_ 6 h6 þ m _ 9 h9 m

_ 8 h8 Š m

ð1097Þ

The energetic performance of gas cycle process can be defined as follows: ZGC ¼

_ GC W _ GC Q

ð1098Þ

cGC ¼

_ GC W _ Q Ex GC

ð1099Þ

The exergetic performance is

4.10.7.3.4.4 Steam cycle process efficiency The power production rate from steam cycle process can be defined as _ SC ¼ m _ 13 ðh13 W

_ 15 ðh15 h14 Þ þ m

h16 Þ

_ 17 ðh18 m

h17 Þ

_ 1 h1 m

ð1100Þ

_ 1 h1 should be added in the Eq. (1099) because of the excess oxygen balance. The heat supplied rate of steam cycle Here m process is _ SC ¼ QL ½m _ 11 h11 Q

_ 12 h12 þ m _ 15 h15 m

_ 14 h14 Š m

ð1101Þ

The energetic performance of steam cycle process is ZSC ¼

_ SC W _ SC Q

ð1102Þ

cSC ¼

_ SC W _ Q Ex

ð1103Þ

The exergetic performance is

SC

4.10.7.3.4.5 Combined process efficiency The produced total power from combined process is given as follows: _ GC þ W _ SC _ Total ¼ W W

ð1104Þ

The energetic and exergetic performance can be defined as follows: Zcomb ¼

_ Total W _ GC þ Q _ SC Q

ð1105Þ

ccomb ¼

_ Total W _ Q _ Q þ Ex Ex GC SC

ð1106Þ

and

Integrated Gasification Combined Cycles 4.10.7.3.5

465

Proton exchange membrane electrolyzer

The net chemical reaction in the PEM electrolyzer can be written as follows: H2 OðlÞ -H2ðgÞ þ 12O2ðgÞ

ð1107Þ

where subscript l and g refer to liquid and gas phases, respectively. The following reactions take place in the anode and cathode sites of PEM electrolyzer, respectively. H2 OðlÞ -12O2ðgÞ þ Hþ ðaqÞ þ 2e

ð1108Þ

Hþ ðaqÞ þ 2e -2H2

ð1109Þ

The produced hydrogen and oxygen output flow rates are given as follows, respectively: _ H2;out ¼ J=2F ¼ N_ H2 O N

ð1110Þ

N_ O2;out ¼ J=4F

ð1111Þ

and _ H2 O is the water consumed rate in the electrolyzer. where J and F are the current density and Faraday constant, respectively, and N To generate hydrogen from the electrolyzer, electrical power must be input to the PEM electrolyzer and this can be written as follows: _ elec ¼ JV ð1112Þ E_ elec ¼ Ex _ _ where Eelec and Exelec are the rate of electrical power and electrical exergy input, respectively. V is the cell potential, and V can be given as follows: V ¼ V0 þ Zact;a þ Zact;c þ Zohm

ð1113Þ

where V0 is the reversible potential, and should be determined using by the Nernst equation. Zact,a is the activation potential of anode side, Zact,c is the activation potential of cathode side and Zohm is the ohmic potential of PEM-electrolyte, and can be written as follows: Zohm ¼ JR where R is overall ohmic resistance. This resistance can be expressed as Z D dx R¼ ½lðxފ s PEM 0

ð1114Þ

ð1115Þ

where D is PEM thickness, sPEM[l(x)] is the local ionic PEM conductivity of membrane and can be calculated as follows:    1 1 sPEM ½lðxފ ¼ ½0:5139lðxÞ 0:326Šexp 1268 ð1116Þ 303 T

where x is the interval in PEM evaluated based on a cathode side interface. l(x) is water amount at the position x in PEM, and can be calculated as follows [74]: lc x þ lc ð1117Þ D where la is the amount of water at anode PEM interface, and lc is amount of water at the cathode-membrane interface. The activation potential of the electrolyzer can be calculated as [75,76]   RT J Zact;i ¼ sinh 1 ; i ¼ at; ct ð1118Þ F 2J0;i      azFZact;i ð1 aÞzFZact;i exp ; i ¼ at; ct ð1119Þ J ¼ J0;i exp RT RT   Eact;i ; i ¼ at; ct ð1120Þ J0;i ¼ Jiref exp RT lðxÞ ¼

la

where subscripts at and ct are anode and cathode sites of PEM electrolyzer, J0 is exchange current density, a is charge transfer factor for anode and cathode side reactions, and generally equal to ½. Z is the count of electrons included per water splitting reactions. For PEM electrolyzer Z must be two. Jiref is pre-exponential indicator and Eact,i is activation power for the anode and cathode side of electrolyzer.

4.10.7.3.6

Parametric studies

The exergy destruction rates of subsystems for integrated process, such as (1) coal gasification system, (2) steam power cycle, (3) gas power cycle, (4) hydrogen production and liquefaction process, and (5) single effect absorption cooling system, are shown in Fig. 54. As seen from Fig. 54, the coal gasification system has the maximum exergy destruction rate among the other subsystems. This case is associated with the fact that the high heat transfer rates across high temperature difference between coal gasification

Integrated Gasification Combined Cycles

Exergy destruction rate (kW)

466

100,000 90,000 80,000 70,000 60,000 50,000 40,000 30,000 20,000 10,000 0

88,411

34,875

27,880 12,711

9385

Steam power Gas power Coal cycle cycle gasification system

3560

Hydrogen production and liquefaction

Absorption cooling

Integrated system

Fig. 54 Exergy destruction rates for the coal gasification-based integrated system.

Exergy destruction ratio (%)

45 40

39.45

35

31.53

30 25 20

14.38

15

10.61

10 4.03

5 0 Coal gasification system

Steam power cycle

Gas power cycle

Hydrogen production and liquefaction

Absorption cooling

Fig. 55 Exergy destruction ratio for the coal gasification-based integrated system.

process components and reference ambient. The gas power subsystem has the next highest irreversibility rate in the integrated gasification combined system. The steam power cycle has the third irreversibility rate. The dimensionless exergy destruction ratio of subsystems of coal gasification-based integrated process are analyzed by using the given above procedure, and analysis outputs are shown in Fig. 55. This exergetic indicator is beneficial step for prioritizing exergy destruction in an intuitive behavior. As seen from Fig. 55, the exergy destruction rate is higher in the coal gasification system than in other subsystems in the integrated gasification combined process. Also, the single effect absorption cooling system does not exhibit important exergy destruction ratio. The exergy efficiency of coal gasification-based integrated cycle subsystems and whole system are calculated, and also shown in Fig. 56. It is seen that, the exergy efficiency of hydrogen production and liquefaction process and coal gasification system are higher than other subsystems of integrated system. Different parametric studies are applied for coal gasification-based integrated system to submit the impacts of variations in reference temperature, coal gasifier temperature, and gas turbine inlet temperature on the exergy destruction rate, energy efficiency and exergy efficiency of integrated process. In this subsection, the outputs of these aforementioned parametric studies are plotted along with their corresponding statements. In Fig. 57, the variation of energy efficiency of integrated system and its subsystems are shown with respect to reference temperature. In this case study, the base case reference temperature is taken as 251C and the dead state temperature is varied between 5 and 401C. It is observed that the reference temperature does not have an important effect on energy efficiencies of steam power cycle, gas power cycle, and hydrogen production and liquefaction process. The coal gasification system and single effect absorption cooling system are the most affected subsystems. Overall system efficiency is increased approximately 6% with 351C change in the ambient temperature. Fig. 58 illustrates that the exergy efficiencies are more affected by the reference temperature, as expected. There is around 1%–3% of change in the coal gasification system, steam power cycle, gas power cycle, hydrogen production and liquefaction system and whole system, where coal gasification system shows the largest increase as a response to increasing reference temperature. The exergy efficiency of single effect absorption cooling system is increased with increasing reference temperature from 5

Integrated Gasification Combined Cycles

467

Exergy efficiency (%)

70 58.15

56.38

60 50.82

50

48.37 39.83

40 30 20

13.93

10 0 Steam power Coal cycle gasification system

Gas power cycle

Hydrogen production and liquefaction

Absorption cooling

Integrated system

Fig. 56 Exergy efficiency for the subsystems of coal gasification-based integrated system.

0.65 0.6

Energy efficiency

0.55 0.5 0.45 0.4

CGS

0.35

SPC GPC

0.3

HPL

0.25

ACS System

0.2 0.15

5

10

15 20 25 30 Reference temperature (°C)

35

40

Fig. 57 Variation of energy efficiencies of integrated system and its subsystems with respect to ambient temperature. ACS, absorption cooling system; CGS, coal gasification system; GPC, gas power cycle; HPL, hydrogen production and liquefaction; SPC, steam power cycle.

to 401C. The exergy efficiency of coal gasification-based integrated system increases approximately 13% by an increase in the reference temperature. The variations of energetic and exergetic performance with coal gasifier temperature are shown in Fig. 59. The energetic and exergetic performance of coal gasification subsystem and integrated system increases linearly, when the coal gasifier temperature increases from 1000 to 14001C. The exergy contained in coal sample is transferred by the gasification process into the chemical and the physical exergy of syngas. Thus, the exergy content in synthesis gaseous are increased with increasing the coal gasification temperature. Fig. 60 illustrates the variation of energetic and exergetic performance for the gas power cycle and coal gasification-based IGCC with varying gas turbine inlet temperature. The energetic performance of gas power cycle and integrated process nearly remain constant where exergy efficiency increases with increasing the gas turbine inlet temperature since the thermal energy transfer rate by absorbing greater thermal energy from the exhaust gas in the boiler and generator increases in the given cycle. To supply environmental analysis, the environmental impacts of single generation, cogeneration, trigeneration, and multigeneration of coal gasification-based integrated process are given in Fig. 61. It is observed that, the multigeneration system with power, heating, cooling, hot water, and liquid hydrogen production has minimum carbon dioxide emissions than the other generation options, supplying a significant motivation for the usage of multigeneration systems. In addition to that, the multigeneration system has the higher exergy efficiency than the other generation systems. Also, the multigeneration cycle has lower carbon monoxide emissions than the other generation system, providing another motivation for the usage of multigeneration system. However, the amount of carbon monoxide emission is significantly less that the amount of carbon dioxide emissions of the IGCC.

468

Integrated Gasification Combined Cycles

0.65 0.6 0.55 Exergy efficiency

0.5 0.45 0.4

CGS

0.35

SPC

0.3

GPC

0.25

HPL

0.2

ACS System

0.15 0.1

5

10

15 20 25 30 Reference temperature (°C)

35

40

Fig. 58 Variation of energy efficiencies of integrated system and its subsystems with respect to ambient temperature. ACS, absorption cooling system; CGS, coal gasification system; GPC, gas power cycle; HPL, hydrogen production and liquefaction; SPC, steam power cycle.

0.64

Energy efficiency

0.62

0.64

CGS

0.62

System

0.6 0.58

0.6

0.56 0.58 0.54 0.56

0.52

CGS

0.54

System

0.52 0.5 1000

1050

1100 1150 1200 1250 1300 Coal gasifier temperature (°C)

1350

Exergy efficiency

0.66

0.5 0.48 0.46 1400

Fig. temperature on energy and exergy efficiency for coal gasification subsystem and integrated system  59 Effects of coal gasification  Pcoal gafication ¼ 5:5 MPa .

4.10.8

Future Directions and Potential Developments

The future directions and potential development techniques associated with the improvement of performance for IGCCs have investigated in this section. The development of IGCCs has traditionally been conducted based on the examination of the multigeneration system properties. In addition to gas cleanup and conditioning other barrier areas that could reduce the cost of fuel products from thermochemical conversion of biomass includes feed handling and drying, gasification, production of different products and coproducts, and process integration.

4.10.8.1

Tar Pollution

Because biomass resources have the higher H/C ratio and O/C ratio than fossil fuels, the higher product of both gases and hydrocarbons, such as tars, is the result. The combinations of tars are dependent on the biomass resources, gasification conditions and secondary gas phase chemical reactions. Tars are the significant problem if they polymerize or condense when the synthesis gaseous are cooled down (condensation temperature of tar from woody biomass is 200 to 5001C). Also, the significant problems with pollution of downstream gaseous cooling and gas cleaning equipment can be formed.

Integrated Gasification Combined Cycles

469

0.66

0.64 0.62

0.62

0.58

0.58

0.56 GPC

0.54

0.54

System

0.52

GPC

Exergy efficiency

Energy efficiency

0.6

0.5

System

0.5 0.48 1100

1150

1200 1250 1300 1350 1400 Gas turbine inlet temperature (°C)

0.46 1500

1450

Fig. 60 Impacts of gas turbine inlet temperature on energetic and exergetic performance for gas power cycle and integrated system.

Environmental impacts

600

513.8

500

425.3

382.5

400

297.6

300 200 100

38.17

27.38

42.82

50.86

22.67

18.54

58.15

12.45

0 Single generation Exergy efficiency (%)

Cogeneration

Trigeneration

Multigeneration

Carbon dioxide emissions (kg/kWh)

Carbon monoxide emissions (kg/kWh) Fig. 61 Comparison of exergy efficiency, unit CO2 and CO emissions of generation cycles for coal gasification-based integrated system.

Therefore, the tar removal technologies (like scrubbers) and tars conversion processes (thermal and catalytic) must be detailed investigated and also the novel technologies and processes of these approaches must be developed for more efficiently and environmentally system design aims.

4.10.8.2

Ash Melting

The quantity of ash between several kinds of biomass feedstocks differs widely (0.1% for wood up to 15% for some agriculture products) and affects the design of the gasification chamber, particularly the ash removal process. Also, the chemical composition of ash is very significant for gasification process, because it affects the melting behavior of ash. The ash melting can produce the slagging and channel formation in gasification chamber. The slags in gasifier can finally block the whole gasification chamber.

4.10.8.3

Too Low Syngas Heating Value

The biomass gasification with air causes in the syngas heating value of about 4–5 MJ/Nm3 with the given below problems:

• • • •

Peak electricity production decreasing in internal combustion engines. Flame instabilities in gas turbines for electricity production. Gas pipeline transport becomes less economical. Too diluted chemical properties for syngas.

4.10.8.4

Feedstock

The diversity in biomass morphology, combinations and moisture ingredient cause different problems with gasification feed and operation. For most gasification chambers, the specific degree of calibration is needed for which pretreatment steps, such as

470

Integrated Gasification Combined Cycles

sieving, drying, densifying, and chipping. The contaminants, such as sand, stone, plastic, paper, etc. can produce operational and emission problems. The biomass and waste feedstock availability, logistics, costs are other important indicators for sustainable gasification cycles. To decrease the high initial investment costs of gasification chamber, the long-term contracts of low-priced, quality, controlled feedstock must be supplied (sustainable fuel stock). In practical, this seems to be difficult; feeding of low-dense ingredients and mixed residues with variation characteristics is still difficult. In addition to that, the standardized fuel definitions are required for sawdust, wood chips, rice husk, etc. The feeding technology of feedstock is very significant issue and particularly the feeding technologies of pressurized processes have some problems to solve (mechanical problems, gas tightness problems, etc.). Different solution technologies and approaches are required in this area.

4.10.8.5

Gas Cleaning

The several gas cleaning technologies have improved with varying successfulness, such as wet/dry cleaning, tar removal/cracking processes. There is a requirement for both technical and economical evaluation factors of several gasification systems. Also, The CO2 separation, storage and utilization ways need to be investigated. A novel air separation process should be development to greatly decrease cost and cycle complexity.

4.10.8.6

Prime Mover

Concerning prime movers, the gas engines are most frequently utilized for power generation. The research and development studies on the fuel cells, Stirling engines and gas turbines must be performed for more efficiently system design. In practical, there is inadequate operational experience to investigate the specifications of produced syngas during gasification process. Hence, numerous strict demands are requested for gas analyses. In other respects, numerous strict limitations of allowable emissions are required for adequate clean gas. There are hardly any research and development studies on adaptations of prime movers to producer syngas.

4.10.8.7

Reliability

Generally, the gasification systems are still considered as being unreliable. Also, in most cases guarantees are lacking or very limited. These cases primarily happen due to the absence of operational experience, unfamiliarity with the process and the trend to closedown designs.

4.10.8.8

Conditions for Commercialization in the Future

The different conditions for commercialization of IGCCs are given as follows:

•

• • • • • • • • • • • •

The integrated gasification technology should be matured for multigeneration aims and proven in demonstration models. This demonstration design should be contained the long-term test efforts of prototype installations. Different development investigations are required for the high-performance gasification chamber design. The installed gasification plants should be reliable having an availability of 90% with low maintenance. Because of the limited operational knowledge, the maintenance requirements of gasifier unit are not well-known. Also, the improvement options should also come from learning of operational experience. The quality of biomass product should meet the client's specification. The tar removal or cracking technologies is an important point of concern in integrated gasification cycles, and still needs focus and solution efforts. The local capacity for manufacturing the gasifier process must exit. The low-pressure integrated gasification technologies must be developed to produce power and heat for local community. The gasifier manufacturers must provide after-sales service, guarantees and training of operators. The reactor design of gasifier should be development for more difficult type of fuels (high alkali and high ash content). Successful models are not existing to demonstrate what is going on with particles during updraft and downdraft gasification, such as shape changes, and flow through the channels between particles. Therefore, to improve the efficiency of integrated gasification process, the reactor design of gasifiers should be optimized by using the physical and mathematical modeling. The hot syngas cleaning solutions must be cost effective and sustainable in the long term. The corrosion and fouling effects in hot gas HEX are another important problem. Reasons for these changes need to be defined and knowledge shared. The personnel, with the skills and motivation needed for operation of gasification cycles, should be present. The incentives to operators will stimulate optimal operation of the gasification system. The correct and objective knowledge on the gasification cycles, the gasifier chamber and its capabilities, its limitations, and how it compares with competing systems should be available to potential customers. The credit finance should be present for customers of confirmed processes and for the decline of risk with first-of-its-kind installations. The administration policies and incentives will decrease the risks for private investments.

Integrated Gasification Combined Cycles

• • • • •

471

The definition of limitations and demonstration projects of proven concepts followed by reproduction and optimization is the best alternative to commercialize gasification process. The arrangements covering of safety, generated gaseous quality, permitted level of noxious emissions and discharges, and other features of gasification chamber operation should be in force. The certification of gasification facility is very significant behavior in promoting and gaining confidence in the technology. When the products meet certain and defined quality standards, the manufacturers can market their product with full operational guarantees. The standardization studies of the biomass gasification components as well as for acceptance tests and guarantee measurements are very important for commercialization. In many cases, with novel techniques for gasifier, ultimately one dominant design will remains becoming the market leader. Any legal barriers to the sale of surplus produced electricity from gasification plant should be removed. The long-term fixed prices for gasification plant are required for feed-in power, sale of heat and purchase of quality controlled gasification fuels.

4.10.9

Closing Remarks

The objective of this chapter is to thermodynamically examine the IGCCs, that are as cheap as existing conventional supplies, as convenient in use, less environmentally damaging, and reasonably safe in use, to convert the biomass and coal source into different useful outputs. In this study, three different IGCCs are comparatively evaluated for case studies. The first integrated system consists of mainly six subsystems, such as (1) biomass gasifier, (2) parabolic dish collectors, (3) Brayton cycle, (4) Rankine cycle, (5) single effect absorption cooling system, and (6) fresh water production cycle to meet the demands of power, heating, cooling, hot water and fresh water. The second system is integrated with the biomass gasifier, double state ORC, single effect absorption cooling process, hydrogen production system and SOFC to meet the demands of power, heating, cooling, hot water and hydrogen. The third integrated system consist of mainly six subsystems, such as (1) coal gasifier, (2) Brayton cycle, (3) Rankine cycle, (4) single effect absorption cooling system, (5) hydrogen production system, and (6) hydrogen liquefaction system to meet the demands of power, heating, cooling, hot water and liquid hydrogen. Also, the thermodynamic model calculations for IGCCs are performed by utilizing the Engineering Equation Solver (EES) software program to examine the impact of the operating temperature, inlet mass flow rate, and environmental temperature on useful outputs production, energy and exergy efficiencies, and exergy destruction. The relationship between exergy destruction, exergy efficiency and environmental impact are investigated. The sustainable development aims need the sustainable supply of energy sources that, in the long term, is sustainably available at reasonable cost and can be used for all required tasks without causing negative societal effects. The improved understanding of sustainability and environmental problems concerning with the energy source utilization is required both to allow the problems to be better addressed and to assist improve solutions that are useful for the economy and society. For these reasons, the IGCCs for multigeneration aims are very essential for environmentally benign integrated system designs. The several concluding outputs should be drawn from this work:

• • • • • • • • • • • •

Thermodynamic analysis results show that the exergetic analysis represents a better sensitivity analysis of the system performance than energy. Capacities and performances of the integrated gasification combined process depend on the design parameters of components. The advanced gasification process is more efficient than the other conventional systems. Integrated energy system for multigeneration is a promising technology for power, hydrogen, chemicals, fresh and/or hot water production, heating and cooling application for rural areas. They are undergoing research and experimentation and further research is needed to improve understanding of integrated systems. The efficiently next generation coal and biomass gasifiers must be development. Multigeneration ways from IGCCs must be investigated for better system design aims. The multigeneration design model can also be utilized to help decision makers in selecting the optimum size and/or efficiencies of the components when designing IGCCs. The small-scale gasification technologies for distributed production market, including hydrogen generation need to be development. At HP and temperature ammonia produced from IGCC must be investigated in industrial usage area. The gas cleaning, tar cracking or removal processes remain the largest problem in gasification cycles. The new processes for hot gas cleaning, ammonia removal and desulfurization from synthesis gaseous mixture must be investigated for more efficiently system design. Hydrogen storage is still an obstacle for hydrogen system due to the necessity of the HP tanks. The modeling and optimization can improve design of gasifier.

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Relevant Websites https://www.netl.doe.gov/publications/proceedings/02/turbines/Izzo.pdf Calpine Fuels Diversity Initiative. http://www.global-greenhouse-warming.com/integrated-gasification-combined-cycle-IGCC.html Global Greenhouse Warming. https://sequestration.mit.edu/pdf/LFEE_2005-002_WP.pdf Laboratory for Energy and the Environment. https://www.mhi.com/discover/earth/technology/gtcc_igcc.html Mitsubishi Heavy Industries. https://www.nrel.gov/docs/legosti/old/6080.pdf National Renewable Energy Laboratory. https://www.netl.doe.gov/research/coal/energy-systems/gasification/gasifipedia/igcc National Energy Technology Laboratory. http://www.climatetechwiki.org/technology/igcc National Energy Technology Laboratory.

4.11 Geothermal Energy Conversion Murat Ozturk, Suleyman Demirel University, Isparta, Turkey Ibrahim Dincer, University of Ontario Institute of Technology, Oshawa, ON, Canada r 2018 Elsevier Inc. All rights reserved.

4.11.1 Introduction 4.11.2 Geothermal Energy and Classification 4.11.2.1 By Potential 4.11.2.2 By Content 4.11.2.3 By Temperature or Enthalpy 4.11.2.4 By Exergy 4.11.3 Geothermal Energy Sources and Applications 4.11.4 General Analysis and Performance Assessment 4.11.4.1 Mass Balance Equation 4.11.4.2 Energy Balance Equation 4.11.4.3 Entropy Balance Equation 4.11.4.4 Exergy Balance Equation 4.11.4.5 Energy and Exergy Efficiencies 4.11.5 Geothermal Energy Conversion 4.11.5.1 Cooling and Heating Application 4.11.5.2 Power Generation 4.11.5.3 Hydrogen Energy Production 4.11.5.4 Alternative Fuels 4.11.5.5 Fresh Water Production 4.11.5.6 Other Useful Commodities 4.11.6 Case Studies 4.11.6.1 Direct Steam Power Generation 4.11.6.2 Single Flash Steam Power Generation 4.11.6.3 Double Flash Steam Power Generation 4.11.6.4 Triple Flash Steam Power Generation 4.11.6.5 Binary Cycle Power Generation 4.11.6.6 Combined Power Generation 4.11.6.7 Geothermal Energy Based Cooling System 4.11.6.8 Geothermal Energy Based Hydrogen Production and Liquefaction System 4.11.7 Future Directions 4.11.8 Concluding Remarks References Further Reading Relevant Websites

Nomenclature A Area (m2) E Energy (kJ) E_ Energy rate (kW) ex Specific exergy (kJ/kg) _ Ex Exergy rate (kW) _ D Exergy destruction rate (kW) Ex F Faraday constant (C/mol) g Gravitational acceleration (m2/s) h Specific enthalpy (kJ/kg) H Enthalpy (kJ) J Current density (A/m2) Jo Exchange current density (A/m2) Jiref Preexponential factor (A/m2) L Length (m)

474

475 477 478 478 479 479 479 480 481 481 482 482 482 483 483 493 496 502 502 505 506 506 509 512 517 522 526 530 535 541 542 543 544 544

m Mass (kg) _ Mass flow rate (kg/s) m P Presuure (kPa) q Specific heat transfer (kJ/kg) Q Heat (kJ) _ Heat rate (kW) Q RPEM Proton exchange membrane resistance (O) Ru Universal gas constant (kJ/mol K) s Specific entropy [kJ/kgK) S Entropy (kJ/K) S_ Entropy rate (kW/K) t Time [s] T Temperature [K] v Volume (m3) V Velocity (m/s)

Comprehensive Energy Systems, Volume 4

doi:10.1016/B978-0-12-809597-3.00415-6

Geothermal Energy Conversion

V0 Reversible potential (V) Vact Activation overpotential (V) Vact,a Anode activation overpotential (V) Vact,c Cathode activation overpotential (V)

w weight (N) W Work (kJ) _ Work rate (kW) W z Elevation (m)

Greek Letters D Change in variable la Water content at anode–membrane interface (O 1) lc Water content at cathode–membrane interface (O 1) l(x) Water content at location x in the membrane (O 1)

sPEM Proton conductivity in PEM (s/m) s(x) Local ionic PEM conductivity (s/m) Z Energy efficiency c Exergy efficiency

Subscript a Air abs Absorber ct Cooling tower cooling Cooling load con Condenser cv Control volume D Destruction e Exit condition ej Ejector en Energy erd Energy recovery device eva Evaporator ev Expansion valve ex Exergy f Fuel fls Flashing g Gas

gen Generation gene Generator heating Heating load HP High pressure l Liquid LP Low pressure i Inlet condition mdm Membrane distillation module mine Mineralizer MP Medium pressure ohm Ohmic p Pump pr Particulate remover pst Posttreatment pt Pretreatment pur Purifier rhs Radiator heating system tot Total

Superscripts
Rate

Ch

Acronyms COP Coefficient of performance CSP Concentrating solar collector EES Engineering Equation Solver HEX Heat exchanger ORC Organic Rankine cycle

PEM Proton exchange membrane PV Photovoltaic PV/T Photovoltaic/thermal collector RE Rational efficiency SExI Specific exergy index TE Task efficiency

4.11.1

475

Chemical

Introduction

Geothermal resources are vastly available worldwide and are widely used for power generation or space heating applications. They are clean (effectively no harmful gas emissions, including CO, CO2, NOx, SOx, etc.), safe, and reliable (renewable and sustainable), and they can play an important role in meeting world energy requirements [1]. Geothermal fluids, as they come from the underground, contain gaseous impurities, such as hydrogen sulfite and radon gaseous, that usually are not permitted to be released to the ambient. Reinjection of used geofluids back into the injection well may, however, solve any harmful emission problems. The different benefit is that geothermal power processes are easy, safe, and adaptable. The heat energy is created by the natural decay over millions of years of radiogenic materials including uranium, thorium, and potassium. The geothermal energy production systems have important advantages [2]. They are environmentally benign, renewable energy sources, and also they can be used for providing baseload electric and heat energy for industrial applications. The geothermal energy sources that have been utilized, or that are under investigation for improving, range from shallow ground to hot water and rock several kilometers below the Earth’s surface.

476

Geothermal Energy Conversion

Throughout the world, geothermal energy systems use geothermal heat to provide electricity and to use directly for nearly 90 years. While there were only 11 countries having installed capacity of more than 100 MWT in 1985, this number had increased to 36 countries by the end of 2014. Five countries appear to be dominant in the direct use of geothermal energy (without heat pumps in MWt) accounting for 68.3% of the world usage, respectively. As seen from Table 1, these countries also have the largest annual energy use. Nowadays, 26 countries are engaged in generating electricity energy from geothermal energy resources. The total installed power generation capacity based on geothermal energy is about 13,300 MW from 582 geothermal power generation plants in 2016. As seen in Fig. 1, the power generation rate based on geothermal energy resources has witnessed a steady, albeit modest, growth over the past four decades. After the oil crisis in 1970, Iceland revolutionized its energy policies, decreasing its oil usage and returning to its own energy sources: geothermal and hydropower. Fig. 2 shows the percentage of energy types employed effectively through various applications for space heating in Iceland. The status of electricity generation rates worldwide from 1990 to 2040 based on fossil and renewable energy sources are illustrated in Table 2. The total energy supply of the world has increased by nearly 100% in 14 years. As seen in this table, world geothermal energy production rate has grown from 36 TWh in 1990 to 77 TWh in 2014. The fossil fuels (coal, oil, and natural gas) will continue to generate nearly 55–60% of the world’s primary energy right through to 2040. Also, it is expected that the electricity generation rate based on installed geothermal power generation plants will reach 361 TWh in 2040. On the other hand, the electricity generation share of geothermal power systems is 0.323% in 2014, and it is expected to reach about 1% in 2040. Hot water comes up through interrelated faults and fractures and appears on the surface of the ground in the form of hot springs or hot steam. In the last decade, the notion of geothermal power systems has significantly developed in its improvement, abilities, and implementation through the reforming of traditional opinion and approaches [3]. The most important property of a geothermal power system is that it generates zero harmful gaseous emissions potentially making it one of the cleanest resources for power generation. Another important feature is that, unlike other renewable energy resources such as solar and wind energy, geothermal energy can produce a constant 24 h of baseload electricity [4,5]. The high temperature geothermal power sources are generally classified as having a geofluid temperature of greater than 150ºC [6]. The medium temperature geothermal power sources are usually classified as having temperatures between 90 and 150ºC [7,8]. The low temperature geothermal power sources refer to those that have a temperature below 90ºC [9]. Furthermore, the efficient use of lower geothermal resources of about 75 to 1001C is under investigation presently. The geothermal resources with temperature scale exist in many regions of the world [7]. Therefore, effective progress for their use will ensure wide potential for geofluid based organic Rankine cycle (ORC) to generate

Table 1

Distribution of geothermal energy among the top five countries

Countries

MWt

Countries

TJ/year

China Turkey Japan Iceland India

6089 2894 2086 2035 986

China Turkey Iceland Japan Hungary

74,041 44,932 26,700 25,630 9,573

Source: John WL, Tonya LB. Direct utilization of geothermal energy 2015 worldwide review Geothermics 2016;60:66–93.

Installed power capacity (MW)

Geothermal power generation 1975−2015 12,995

14,000 10,780

12,000 8912

10,000

7974 6798

8000

5832

6000

3887

4764

4000 2000

1300

0 1975

1980

1985

1990

1995

2000

2005

Year Fig. 1 World cumulative installed geothermal power generation capacity during the period 1975–2015.

2010

2015

Geothermal Energy Conversion

477

100% 90% 80% 70%

Geothermal

60%

89%

50% 40% 30% 20%

Oil

10%

Electricity

0% 1970

1980

1990

10% 1% 2000

Fig. 2 Relative share of energy resources in the heating of houses in Iceland. Data taken from Bardadottir H. Geothermal Development and Research in Iceland, National Energy Authority and Ministries of Industry and Commerce; 2006.

Table 2

Status of electricity generation rate from 1990 up to 2040 based on the fossil and renewable energy sources

Energy sources

Electricity generation (TWh) 1990

2014

Share (%) 2020

2025

2030

2035

2040

Coal Oil Natural gas Nuclear energy Hydropower Bioenergy Wind turbine Geothermal Solar PV CSP Marine

4,425 1,358 1,753 2,013 2,143 131 4 36 0 1 1

9,707 1,035 5,148 2,535 3,894 495 717 77 190 9 1

9,741 822 5,804 3,053 4,387 642 1,508 111 599 30 3

9,934 727 6,513 3,405 4,887 785 2,118 150 953 61 6

10,245 633 7,305 3,847 5,382 954 2,706 207 1,329 109 15

10,547 585 8,155 4,205 5,834 1,147 3,296 283 1,731 175 30

10,786 547 8,909 4,532 6,230 1,353 3,881 361 2,137 254 54

Total generation

11,863

23,809

26,698

29,540

32,732

35,989

39,045

2014 40.77 4.347 21.62 10.65 16.36 2.079 3.012 0.323 0.798 0.038 0.004 100

2040 27.63 1.401 22.82 11.61 15.96 3.465 9.94 0.925 5.473 0.651 0.138 100

Note: CSP, Concentrating solar collector. Source: International Energy Agency. World Energy Outlook. Paris: International Energy Agency; 2016.

electricity [7]. Green and Gerald [10] have given that novel low temperature power production processes may greatly expand the geothermal sources that should be improved economically today. The high and medium temperature geothermal power sources are generally the product of thermal streams that are generated by the molten core of the earth. The heat energy flows from deep within the world and collects in areas of water or rock. The low temperature geothermal sources are generally created through the collection of solar radiation within the ground [11]. The general processes of geothermal power systems are direct steam; single, double, and triple flash, and binary and combined/hybrid cycles. The primary disadvantages of generating power by using geothermal resources are having higher investment cost and lower performance than conventional power production processes. The reason for having low performance is that geofluid resource temperature is much lower. The performance of geothermal process can be increased by using some different process based on the geothermal fluid types, for instance steam, wet steam, and hot water. The overall objective of this chapter is, therefore, to investigate the geothermal energy, geothermal energy sources, geothermal energy conversion techniques, and geothermal energy based integrated system for some useful outputs. In the case study subsection, thermodynamic assessments investigate how various operating conditions and reference conditions within these geothermal power generation arrangements impact the exergy efficiencies and exergy destruction rates. This chapter further aims to discuss the geothermal energy based integrated system for multigeneration, such as power generation, cooling, space or greenhouse heating applications, hydrogen and alternative fuels generation, water heating, fresh water, industrial process heating, and other commodities (drying air, food drying, cooking, etc.).

4.11.2

Geothermal Energy and Classification

The growing needs for energy in many sectors and the harmful environmental effects caused by fossil fuels (coal, natural gas, and oil) have motived researchers, scientists, engineers, technologists, and policy makers to give more attention to switching toward

478

Table 3

Geothermal Energy Conversion

Status of geothermal power plants in two particular years as 2005 and 2012

Country

2005

2012

No of units

MWe

No of units

Australia Austria China Costa Rica El Salvador Ethiopia France Germany Guatemala Iceland Indonesia Italy Japan Kenya Mexico New Zealand Nicaragua Papua New Guinea Philippines Portugal Russia Turkey USA

1 2 13 6 7 0 2 1 9 24 15 33 22 9 37 39 7 6 58 5 12 2 193

0.15 1.125 27.6 163 204.3 0 14.7 0.2 44.6 422.4 807 811.2 537.74 130.2 953.3 572.1 108.9 56 1979.91 16 79 27.8 2555.5

1 3 8 8 7 1 2 4 9 31 23 35 21 13 39 43 5 6 48 6 12 8 253

Total

503

9512.725

586

MWe 0.15 1.145 24 205 204.3 8.5 14.7 6.75 44.6 715.4 1,134 882.3 535.26 166.2 983.3 783.3 87.5 56 1,840.9 26 79 94.98 2,774.43 10,667.72

alternative and sustainable energy sources (solar, wind, geothermal, biomass, etc.). Nowadays, 24 different countries are generating electricity by using geothermal energy resources. The status of geothermal plant units and power generation rates for these countries between 2005 and 2012 are illustrated in Table 3. Among the alternative energy technologies, geothermal energy is found in abundance, is a completely free source of energy, and also is mainly used for electricity production, residential or greenhouse heating and cooling processes, and industrial drying, distillation, and desalination, depending on the geofluid source conditions. For this reason, the classification of geothermal energy has long been necessary since geothermal energy is abundant and is an important for source for humans. Therefore, there have been many classification types for geothermal energy such as by accessibility, potential, content, temperature, exergy, etc. In this part of this study, geothermal energy is classified by potential, content, temperature, and exergy.

4.11.2.1

By Potential

Geothermal energy is abundant; however, its direct use or electricity production is expensive. The reason for this situation is that geothermal energy is theoretically abundant, but there are some technical, economical, and sustainability-related limitations. Fig. 3 shows economic and realizable parts of renewable energy sources, and this is also technically valid for geothermal sources [12]. The theoretical potential defines the physically utilizable energy supply. Nowadays, only small fractions of the theoretical potential of geothermal energy can actually be utilized because of the technical, structural, and administrative limitations. Also, the technical potential describes the fraction of theoretical potential that can be used with current technology. The economic potential defines the time and location dependent fraction of technical potential that can be economically utilized. Then sustainability takes place, that is, the use of geothermal energy in a sustainable manner [13]. The top left part of Fig. 3 shows the reserves of geothermal energy and rest of them is called as resource. The developable potential defines the fraction of economic potential that can be improved under practical conditions (regulations, environmental and social restrictions). Thus, as seen in Fig. 3, it is generally smaller than the economic potential.

4.11.2.2

By Content

Geothermal energy reservoirs have been classified also in terms of their content or phase of fluid. Geothermal reservoirs can be either water dominated fields or vapor dominated fields [14]. Water dominated fields can be also divided into two groups: (1) hot water fields and (2) wet steam fields. Hot water fields contain water with temperature up to 1001C and these fields are the lowest temperature fields. Wet steam fields have both hot water and partly vapor. The temperatures of these fields are higher than 1001C.

Geothermal Energy Conversion

479

Realizable

Developable Sustainable Economic

Economic

Technical

Theoretical Fig. 3 Classification of geothermal sources by potential. Modified from DiPippo R, Marcille D. Exergy analysis of geothermal power plants. Geotherm Resources Council Trans 1984;8:47–52.

Table 4

Thermodynamic limits of geothermal power conversion

Geofluid

Temperature (1C)

Sensible heat exchange (%)

Latent heat exchange (%)

Low temperature Medium temperature High temperature

150 220 500

17 24 44

29 39 61

Source: Dincer I, Zamfirescu C. Sustainable energy systems and applications. New York, NY: Springer; 2011.

On the other hand, vapor dominant fields contain superheated or dry saturated vapor having pressure higher than atmospheric pressure.

4.11.2.3

By Temperature or Enthalpy

Temperatures of geothermal sources range between about 501C and approximately 3501C. Geothermal sources are divided into three groups according to their temperatures; however there is no exact boundary between these groups. According to Muffer and Cataldi [15], reservoirs having lower than 901C are low temperature reservoirs, reservoirs with temperature between 90 and 1501C are medium temperature reservoirs, and finally reservoirs having higher than 1501C are high temperature reservoirs. In order to indicate maximum conversion performance from low temperature, medium temperature, and high temperature geofluid resources, thermodynamic limits of geothermal power conversion based on the resource temperature, sensible heat exchange, and latent heat exchange rate are given in Table 4.

4.11.2.4

By Exergy

Classifying geothermal resources according to only temperature or enthalpy is not convenient for deciding the feasibility of the geothermal resources. Lee has developed a new idea about classifying geothermal resources in terms of exergy content. According to Lee [6], specific exergy of fluid can be normalized by the maximum saturated steam exergy in order to obtain specific exergy index (SExI), and can be written as follows: SExI ¼ hbrine

273:16sbrine =1192

ð1Þ

which is a straight line on a specific enthalpy (h) and specific entropy (s) plot of the Mollier diagram. Finally, the geothermal resources with SExIo0.05 are defined as low exergy resources, 0.05rSExIo0.5 as medium exergy sources, and SExIZ0.5 as high exergy resources.

4.11.3

Geothermal Energy Sources and Applications

One other possible resource of renewable energy is the Earth’s heat, which is called geothermal energy. Geofluid sources vary widely from one place to another, based on the geofluid temperature and depth from ground level, rock structure, and richness of geofluid. The core temperature of the earth is calculated as nearly 43001C, and because of the rock conductivity, the ground temperature at almost 4 km below the earth’s surface can reach nearly 901C. But, the geothermal heat intensity is low compared to solar radiation intensity, namely B0.1 W/m2 versus B240 W/m2 for geothermal solar, respectively [16]. The geothermal energy resources can be utilized for power generation or any different suitable industrial, agricultural, or domestic applications.

480

Geothermal Energy Conversion

Turbine Particulate remover

Particulate matter Production well

Moisture remover

Power

Condenser

Reinjection well

Reinjection well

Recharge region Hot spring or steam vent

Geothermal well Impermeable cap rock region

Rainwater

Cold feed water

Hot geofluids

Reservoir region Heat fow from magma to surface Impermeable rock region Magma reservoir

Fig. 4 The schematic presentation of an optimal geothermal recharge field, impermeable cover, reservoir and heat resource of geothermal energy. Adapted from Barbier E. Geothermal energy technology and current status: an overview. Renew Sustainable Energy Rev 2002;6:3–65.

Geothermal energy sources are the thermal power that could reasonably be competed with other renewable energy based electricity generation prices in the near future. It is expected that the electrical energy generation rate from geothermal energy worldwide can grow by nearly tenfold the current technological limit. The developed and developing countries are intensely researching and evaluating their geothermal resources to supply their power requirements. Actually, the near future utilization of geothermal technology generally depends on overcoming design parameters in generation and utilization applications, and also its installation and maintenance cost indicators compared to other energy technologies. The geothermal action in a space is absolutely the first important indicator that subsurface rocks in the field are warmer than the ground level. The regional heat resource could be a magma core between 600 and 10001C, intruded within several kilometers of the ground level. Generally, the geothermal resources are covered with impermeable rocks that prevent the warm geofluids from easily reaching the surface level and keep them under pressure. As seen in Fig. 4, the superheated steam, steam mixed with water or only hot water for industrial applications, should be obtained based on the hydrogeological case and rock temperature. Actually, the thermal working fluid is generally rain water that infiltrates into the recharge fields. Therewithal, the temperature of thermal fluid is increased while penetrating the hot rocks of the reservoir. Determination of geofluid resources are made on the basis of geological or geophysical indicators, such as (1) depth, thickness, and extent of geofluid aquifer; (2) characteristics of geothermal field formation; (3) salinity rate and geochemistry of geofluid presented in aquifer area; and (4) temperature, porosity, and permeability rate of rock structure. Further details are available elsewhere [17]. The geothermal energy sources should be divided into three groups based on the geofluid temperature range such as low temperature (until 901C), moderate temperature (901C to 1501C), and high temperature (above 1501C). These temperature ranges are suitable in many industrial applications. Direct utilization of geothermal power sources supply a wide variety of applications based on the temperature range. The broad classification of different direct utilizations of geothermal energy on the basis of their temperature requirements is given in Table 5.

4.11.4

General Analysis and Performance Assessment

In this subsection, thermodynamic assessment relevant to energy and exergy analyses is described. Thermodynamic analysis is generally based on four balance equations: (1) mass balance equation, (2) energy balance equation, (3) entropy balance equation,

Geothermal Energy Conversion

Table 5

Direct use applications of geothermal energy resources based on temperature range

20–501C

50–801C

Applications

• • • • • • • •

481

80–1201C

Applications

Fish farming Swimming pool Thermal bath Fermentation Aquaculture Soil warming Mushroom growing Heat pumps

• • • • • •

120–1601C

Applications

Space heating Dry air Greenhouse heating Grains drying Fruits drying Vegetable drying

• • • • • • •

Applications

Fresh water Food drying Drying of stock fish Leather and fur treatment Washing and dying of textiles Pulp and paper processing Process heating

• • • • • •

Flash cycles Space cooling Direct steam Fresh water by distillation Evaporation in sugar refining Industrial space air conditioning

160–2201C Applications

• • • • • • •

Binary cycle Kalina cycle Drying farm products at high rates Canning of food Refrigeration by ammonia absorption Digestion in paper pulp Chemical production

42201C Applications

• • • •

Hydrogen production Alternative fuel production Conventional power production Cogeneration

and (4) exergy balance equations [18–20]. The general models of balance equations should be explained to develop a clear understanding of the systematic approach adopted in the geothermal energy based processes. In the most general sense, any balance equation for a quantity in a process can be defined as the following equation: Input þ Generation

Output

Consumption ¼ Accumulation

ð2Þ

This balance equation is defined as the quantity balance and can be stated as quantity accumulated in the process is equal to the difference between the net quantity transfer through the process boundary plus the quantity generated and the quantity consumed within the process boundaries. Based on this procedure, the general mass, energy, entropy, and exergy balance equation can be defined as follows: _ input m

_ output ¼ m _ accumulation m

ð3Þ

E_ input

E_ output ¼ E_ accumulation

ð4Þ

S_ input þ S_ generation _ input Ex

_ output Ex

S_ output ¼ S_ accumulation

ð5Þ

_ consumption ¼ Ex _ accumulation Ex

ð6Þ

To investigate the thermodynamic analysis of geothermal power system components, the detailed mass, energy, entropy, and exergy balance equations and also energy and exergy efficiency equations are defined in the next subsections.

4.11.4.1

Mass Balance Equation

The conservation of mass in any process is the fundamental indicator in thermodynamic analysis. The mass balance equation can be defined as follows: X X dmcv _e¼ _i m m ð7Þ dt _ are the mass and mass flow rate, respectively, subscripts i and e are the inlet and outlet flow conditions, where m and m respectively, and subscript cv is the control volume. In the steady state and steady flow conditions, the mass balance equation can be written as follows: X X _i¼ _e m m ð8Þ

4.11.4.2

Energy Balance Equation

The net energy of the control volume of any process is always conserved within the system based on the first law of thermodynamics. Also, the energy in an isolated process is always constant. The energy balance equation can be given as follows:  X   X  v2 v2 dE _ W _ net þ _ i hi þ i þ gzi _ e he þ e þ gze ¼ cv Q ð9Þ m m 2 2 dt _ is the heat transfer rate, W _ is the power, h is the specific enthalpy, v is the velocity, g is the gravitational acceleration, z is where Q the elevation, E is the energy, and t is the time.

Geothermal Energy Conversion

482

In the steady state conditions, the energy balance equation can be given as follows: X X X X X X _iþ _eþ _ i¼ _e _ e he þ _ i hi þ Q m Q m W W

4.11.4.3

ð10Þ

Entropy Balance Equation

The entropy generation rate (Sgen) is associated with the losses in a process, and can be defined as entropy balance equation for a control volume as follows: X X Q X X Q _ _  dScv _ e se _ i si þ m m S_ gen ¼ ð11Þ þ T e T i dt where s is the specific entropy, S_ gen is the entropy generation rate or entropy flow, S is the entropy, and T is the temperature at which heat fluxes cross the process boundary. In the steady state conditions, the entropy balance equation can be written as follows: X X Q X Q X _ _ _ e se þ _ i si þ þ S_ gen ¼ ð12Þ m m T i T e Note that, in actual life, the exiting entropy flow rate is always higher than that of entering the process, where the difference due to internal irreversibility is named as entropy generation.

4.11.4.4

Exergy Balance Equation

According to the second law of thermodynamics, the exergy balance equation can be defined as follows: X Q X X X X X _ þ _ W¼ _ Qþ _ W þ Ex _ D _ e exe þ _ i exi þ m m Ex Ex Ex Ex i e e i

ð13Þ

_ W is the exergy rate associated with shaft work, and _ Q is the exergy rate of heat energy transfer, Ex where ex is the specific exergy, Ex _ D is the exergy destruction rate. The specific exergy can be defined as follows: Ex ex ¼ exph þ exch þ ex pt þ exkn

ð14Þ

where exph, exch, expt, and exkn are the physical, chemical, potential, and kinetic exergy, respectively. For this paper, expt and exkn are accepted as negligible, because the exergy exchange rates of these physical quantities are very small. In geothermal power plants, the physical (or flow exergy) and chemical exergy rates must be written, because the working fluid coming from underground has different dissolved solid salts or minerals. These exergy values can be written as follows: ex ph ¼ h

ho To ðs so Þ X   ex ch ¼ ni u0i u00 i

u0i

ð15Þ ð16Þ u00 i

is chemical potential of ith component in thermomechanical equilibrium and is the chemical potential of ith where component in chemical equilibrium [21]. The exergy rates associated with heat transfer and work across the boundary of a control volume can be defined as follows, respectively:   _ _ Q ¼ 1 To Q ð17Þ Ex T _ W ¼W _ Ex

ð18Þ

_ D ¼ T0 S_ gen Ex

ð19Þ

Also, the exergy destruction rate can be calculated as

4.11.4.5

Energy and Exergy Efficiencies

The definitions of energy and exergy efficiencies are a significant factor in decision-making regarding energy source usage. For any energy production or utilization process, the nondimensional ratio of quantities is frequently utilized to define the energy and exergy efficiencies. The energy and exergy efficiencies of geothermal power plants and their components are evaluated in order to investigate the performances of investigated geothermal power processes. The energy efficiency (Z) and exergy efficiency (c) equations for steady state conditions can be defined as follows, respectively: Z¼

c¼

energy in product outputs ¼1 total energy inputs

exergy in product outputs ¼1 total exergy inputs

energy loss energy in inputs

ð20Þ

exergy loss þ exergy consumption exergy in inputs

ð21Þ

Geothermal Energy Conversion

483

The other exergy efficiency definitions for steady state conditions can be given as follows [22]: RE ¼

total exergy output ¼1 total exergy input

TE ¼

exergy consumption actual exergy input

theoretical minimum exergy input required actual exergy input

ð22Þ ð23Þ

where RE and TE are the rational efficiency and task efficiency, respectively. It is noted that the exergy efficiency equation often presents more illuminating insights into system behavior than the energy efficiency equation because the exergy efficiency equation supplies a dimension of potential for improvement.

4.11.5

Geothermal Energy Conversion

The power of geothermal energy has been utilized by many cultures for centuries. In the world, the first working prototype of geothermal process for power generation was constructed in Italy by Prince Gionori Conti in 1905. Also, the first commercial geothermal energy based power process with 250 kW electric generation was designed in 1913 at Larderello, Italy. After these geothermal energy installation applications, different geothermal energy based plants were installed, such as in New Zealand at Wairakei in 1958, an application process at Pathe, Mexico in 1959, and the first commercial process at The Geysers in the United States in 1960. Also, Japan followed with 23 MW at Matsukawa city in 1966. All of these geothermal energy systems utilized steam directly from the earth, except for Wairakei, which was the first to utilize flashed steam for power generation. Nowadays, geothermal energy can be used for power generation, cooling or heating, hydrogen and alternative fuels generation, water heating, fresh water, industrial process heating, other commodities (drying air, food drying, cooking, etc.), and health purposes. The geothermal energy based energy conversion options are given in Table 6. The installed geothermal power plants based on each type of plant in the world in 2015 are given in Table 7. As seen in this table, the single flash geothermal power plant for power generation application is the most used technology in the world. The distributions of geothermal power generation capacity and under development capacity of countries are illustrated in Fig. 5.

4.11.5.1

Cooling and Heating Application

As shown in Fig. 6, the geothermal energy resources can be used for heating and cooling application separately or simultaneously. The heating and cooling applications are treated as significant power processes in most developed and developing countries and are often responsible for an important part of their energy consumption rate. The absorption cooling systems use low grade or waste thermal energy, instead of mechanical energy, to provide cooling, and do not require a compressor subunit. Popular renewable energy based low grade thermal energy sources are concentrating solar energy technologies and low temperature geothermal resources. Geothermal energy resources can also be used to provide heat energy to absorption refrigeration systems in places where the geothermal source temperature is above 901C. The geothermal energy based single effect cooling system is illustrated in Fig. 7. The mechanical vapor compressor is replaced by a waste heat thermal compressor that consists of the generator, condenser, expansion valve, evaporator, absorber, pump, and heat exchanger (HEX). The absorption cooling process uses the chemical mixture as the cooling fluid. The two most common refrigerant/ absorbent mixtures used in absorption chillers are (1) lithium bromide and water (LiBr/H2O) pair (H2O behaviors as refrigerant), and (2) ammonia and water (NH3/H2O) pair (NH3 behaviors as refrigerant). The refrigerant vapor in the evaporator subcomponent of a single effect absorption cooling system is absorbed by a solution mixture in the absorber. This solution is then pumped to the generator where the refrigerant is revaporized using a low-grade steam source coming from geothermal brine. The refrigerant-depleted solution is then returned to the absorber via a throttling device. Applying thermodynamic assessments to processes for heating and cooling can supply a better understanding of their behaviors and enhanced steps for developing them. The general balance equations, i.e., mass, energy, entropy, and exergy, for geothermal energy based single effect absorption cooling system components are defined in this subsection as given below. Generator: The mass, energy, entropy, and exergy balance equations are written for a generator under the steady state and steady flow conditions as follows: _ 2; m _6þm _9 Mass: m _1 ¼m _5 ¼m

ð24Þ

_ gena ¼ m _ 5 h5 þ Q _ 2 h2 þ m _ 6 h6 þ m _ 9 h9 Energy: m _ 1 h1 þ m

ð25Þ

_ gena =Tgena þ S_ gen;gen ¼ m _ 5 s5 þ Q _ 2 s2 þ m _ 6 s6 þ m _ 9 s9 Entropy: m _ 1 s1 þ m

ð26Þ

_ Q ¼m _ D;gena _ 5 ex5 þ Ex _ 2 ex2 þ m _ 6 ex6 þ m _ 9 ex 9 þ Ex Exergy: m _ 1 ex1 þ m gena

ð27Þ

Condenser: The mass, energy, entropy, and exergy balance equations can be defined for a condenser under the steady state and steady flow conditions as follows: _ 10 ; m _ 14 Mass: m _9 ¼m _ 13 ¼ m

ð28Þ

_ con _ 13 h13 ¼ m _ 10 h10 þ m _ 14 h14 þ Q Energy: m _ 9 h9 þ m

ð29Þ

484 Geothermal Energy Conversion

Table 6

Geothermal energy conversion options

Thermal

Power

Applications

• •

Heating Cooling

•

Absorption cooling systems (single, double or triple) District heating systems Swimming pool Heat pumps

Technologies

• • •

Application

•

Electricity

• • • • • •

Direct steam Single flash Double flash Triple flash Binary cycle Combined and hybrid cycle Kalina cycle

Technologies

•

Hydrogen Application

•

Hydrogen

• •

Electrolyzer High temperature electrolyzer Low temperature thermochemical cycles Direct production from steam

Technologies

• •

Alternative fuels Application

• • • • •

Ethanol Methanol Butanol Propone Nonfossil methane

•

Suitable integrated technologies

Technologies

Fresh water Application

•

Fresh water

•

membrane distillation system desalination process

Technologies

•

Other commodities Application

• • • •

Dry air Food drying Food cooling Food cooking

•

Suitable technologies

Technologies

Health option Application

• • •

Thermal bath Swimming pool Thermal water therapy

Technologies

•

Suitable technologies

Geothermal Energy Conversion

Table 7

485

Geothermal power plants for each technology per country (installed capacity MW)

Country

Back pressure

Dry steam

1-flash

2-flash

3-flash

Binary

Hybrid

Total

Australia Austria China Costa Rica El Salvador Ethiopia France Germany Guatemala Iceland Indonesia Italy Japan Kenya Mexico New Zealand Nicaragua Papua New Guinea Philippines Portugal Russia Turkey USA Totals Percent of total

– – – 5 – – – – – – – – – 48 75 44 10 – – – – – – 182 1.44

– – – – – – – – – – 460 795 24 – – – – – – – – – 1584 2863 22.64

– – 1 140 160 – 10 – – 564 873 120 355 543 466 209 142 50 1286 – 82 20 60 5081 40.19

– – 24 – 35 – 5 – – 90 – – 135 – 475 356 – – 365 – – 178 881 2544 20.12

– – – – – – – – – – – – – – – 132 – – – – – – 50 182 1.44

1 1 3 63 9 7 2 27 52 10 8 1 7 4 3 265 8 – 219 29 – 198 873 1790 14.16

– – – – – – – – – – – – – – – – – – – – – – 2 2 0.016

1 1 28 208 204 7 17 27 52 664 1,341 916 521 595 1,019 1,006 160 50 1,870 29 82 396 3,450 12,644 100.00

Source: Bertani R. Geothermal power generation in the world 2010–2014 update report. Geothermics 2016;60:31–43.

4500 4000 3500

MW

3000 2500 2000 1500 1000 500

ly Ita

pa n Ke ny a M e N ew xic Ze o a Pa N lan pu ica d ra a g N ew ua G Ph uin ilip ea pi n Po es rtu ga R l us si a U Tur ni k ey te d St at es

Et

Ja

ad or hi op ia G er m G ua an de y l G oup ua e te m al Ic a el an In do d ne si a

ic a

lv

R

Sa

ta

EI

C

os

C

hi

na

0

Operating capacity (MW)

Planned capacity additions (MW)

Fig. 5 Geothermal power operating and planned capacity by country. Adapted from Bertani R. Geothermal power generation in the world 2010–2014 update report. Geothermics 2016;60:31–43.

_ con =Tcon _ 13 s13 þ S_ gen;con ¼ m _ 10 s10 þ m _ 14 s14 þ Q Entropy: m _ 9 s9 þ m

ð30Þ

_ Q þ Ex _ D;con _ 13 ex13 ¼ m _ 10 ex10 þ m _ 14 ex14 þ Ex Exergy: m _ 9 ex9 þ m con

ð31Þ

Expansion valve I: The mass, energy, entropy, and exergy balance equations for expansion valve I can be written under the steady state and steady flow conditions as follows: _ 11 Mass: m _ 10 ¼ m

ð32Þ

486

Geothermal Energy Conversion

Geothermal energy source

Heating unit

Cooling unit

Heating

Cooling

Fig. 6 A block diagram of heating and cooling applications based on geothermal energy.

9

Generator

1

6

5

Geothermal brine

2 3

8

13 Expansion valve I

7 Pump

17 Production well

10

HEX 4

14

Condenser

Expansion valve II

Absorber

12

11 15 Evaporator

District cooling

18

16

Injection well Fig. 7 Schematic diagram of a common geothermal energy based cooling system.

_ 11 h11 Energy: m _ 10 h10 ¼ m

ð33Þ

_ 11 s11 Entropy: m _ 10 s10 þ S_ gen;ev_I ¼ m

ð34Þ

_ D;ev_I _ 11 ex11 þ Ex Exergy: m _ 10 ex10 ¼ m

ð35Þ

Evaporator: The mass, energy, entropy, and exergy balance equations can be defined for an evaporator under the steady state and steady flow conditions as follows: _ 12 ; m _ 16 _ 15 ¼ m Mass: m _ 11 ¼ m

ð36Þ

_ eva ¼ m _ 15 h15 þ Q _ 12 h12 þ m _ 16 h16 Energy: m _ 11 h11 þ m

ð37Þ

_ eva =Teva þ S_ gen;eva ¼ m _ 15 s15 þ Q _ 12 s12 þ m _ 16 s16 Entropy: m _ 11 s11 þ m

ð38Þ

_ Q ¼m _ D;eva _ 15 ex15 þ Ex _ 12 ex12 þ m _ 16 ex16 þ Ex Exergy: m _ 11 ex 11 þ m eva

ð39Þ

Absorber: The mass, energy, entropy, and exergy balance equations can be given for an absorber under the steady state and steady flow conditions as follows: _8þm _ 12 ; m _ 18 Mass: m _3¼m _ 17 ¼ m

ð40Þ

_ abs _ 12 h12 þ m _ 17 h17 ¼ m _ 3 h3 þ m _ 18 h18 þ Q Energy: m _ 8 h8 þ m

ð41Þ

_ abs =Tabs _ 12 s12 þ m _ 17 s17 þ S_ gen;abs ¼ m _ 3 s3 þ m _ 18 s18 þ Q Entropy: m _ 8 s8 þ m

ð42Þ

_ D;abs þ Ex _ Q _ 12 ex 12 þ m _ 17 ex17 ¼ m _ 3 ex 3 þ m _ 18 ex 18 þ Ex Exergy: m _ 8 ex 8 þ m abs

ð43Þ

Pump: For the pump of single effect absorption cooling system, the balance equations are provided under the steady state and steady flow conditions as follows: _4 Mass: m _3 ¼m

ð44Þ

_ p¼m _ 4 h4 Energy: m _ 3 h3 þ W

ð45Þ

Geothermal Energy Conversion

487

_ 4 s4 Entropy: m _ 3 s3 þ S_ gen;p ¼ m

ð46Þ

_ p ¼m _ D;p _ 4 ex 4 þ Ex Exergy: m _ 3 ex 3 þ W

ð47Þ

Expansion valve II: The mass, energy, entropy, and exergy balance equations for expansion valve II can be written under the steady state and steady flow conditions as follows: _8 Mass: m _7 ¼m

ð48Þ

_ 8 h8 Energy: m _ 7 h7 ¼ m

ð49Þ

_ 8 s8 Entropy: m _ 7 s7 þ S_ gen;ev_II ¼ m

ð50Þ

_ D;ev_II _ 8 ex 8 þ Ex Exergy: m _ 7 ex 7 ¼ m

ð51Þ

Heat exchanger: The mass, energy, entropy, and exergy balance equations for HEX can be expressed under the steady state and steady flow conditions as follows: _ 5; m _7 _6¼m Mass: m _4¼m

ð52Þ

_ HEX ¼ m _ 6 h6 þ Q _ 5 h5 þ m _ 7 h7 Energy: m _ 4 h4 þ m

ð53Þ

_ HEX =THEX þ S_ gen;HEX ¼ m _ 6 s6 þ Q _ 5 s5 þ m _ 7 s7 Entropy: m _ 4 s4 þ m

ð54Þ

_ D;HEX _ Q ¼m _ 5 ex 5 þ m _ 7 ex 7 þ Ex _ 6 ex6 þ Ex Exergy: m _ 4 ex 4 þ m HEX

ð55Þ

The energetic coefficients of performances (COPen) and exergetic coefficients of performances (COPex) of single effect absorption process can be written as follows, respectively: COPen ¼

_ eva Q _ _p Qgena þ W

ð56Þ

COPex ¼

_ Q Ex eva _ _ExQ gena þ W p

ð57Þ

The thermomechanical cooling processes convert thermal energy to mechanical energy to generate the required amount of cooling applications. The steam ejector process is generally used in the thermomechanical cooling process. The schematic diagram of integrated steam ejector process with geothermal energy resource is illustrated in Fig. 8. In this integrated process, the geofluid enters the HEX, which then supplies high temperature working fluid to the ejector subcomponent where the pressure rate of the high temperature working fluid drops. The exiting low pressure (LP) working fluid from the ejector passes through the evaporator where a water–vapor mixture extracts heat from the working fluid flowing through the evaporator. The mass, energy, entropy, and exergy balance equations for geothermal energy based thermomechanical cooling system components are defined in the next subsections. Heat exchanger: The mass, energy, entropy, and exergy balance equations for the HEX can be expressed under the steady state and steady flow conditions as follows: _ 2; m _9 Mass: m _1¼m _3¼m

ð58Þ

_ 9 h9 ¼ m _ 2 h2 þ m _ 3 h3 Energy: m _ 1 h1 þ m

ð59Þ

3

HEX

1 Geothermal brine 2

Ejector 9

7

Pump

Evaporator

12

4 6

Expansion valve

13 8

Production well

5

Injection well Fig. 8 Schematic diagram of a geothermal energy based thermomechanical cooling system.

10 Condenser 11

488

Geothermal Energy Conversion _ 9 s9 þ S_ gen;HEX ¼ m _ 2 s2 þ m _ 3 s3 Entropy: m _ 1 s1 þ m

ð60Þ

_ D;HEX _ 9 ex 9 ¼ m _ 2 ex2 þ m _ 3 ex 3 þ Ex Exergy: m _ 1 ex1 þ m

ð61Þ

Ejector: The mass, energy, entropy, and exergy balance equations for an ejector can be given under the steady state and steady flow conditions as follows: _7¼m _4 Mass: m _3þm

ð62Þ

_ 7 h7 ¼ m _ 4 h4 Energy: m _ 3 h3 þ m

ð63Þ

_ 7 s7 þ S_ gen;ej ¼ m _ 4 s4 Entropy: m _ 3 s3 þ m

ð64Þ

_ D;ej _ 7 E7 ¼ m _ 4 ex 4 þ Ex Exergy: m _ 3 ex 3 þ m

ð65Þ

Condenser: The mass, energy, entropy, and exergy balance equations for a condenser can be written under the steady state and steady flow conditions as follows: _5; m _ 11 _ 10 ¼ m Mass: m _4¼m

ð66Þ

_ 10 h10 ¼ m _ 5 h5 þ m _ 11 h11 Energy: m _ 4 h4 þ m

ð67Þ

_ 10 s10 þ S_ gen;con ¼ m _ 5 s5 þ m _ 11 s11 Entropy: m _ 4 s4 þ m

ð68Þ

_ D;con _ 10 ex 10 ¼ m _ 5 ex5 þ m _ 11 ex11 þ Ex Exergy: m _ 4 ex 4 þ m

ð69Þ

Expansion valve: The mass, energy, entropy, and exergy balance equations for the expansion valve can be defined under the steady state and steady flow conditions as follows: _6 Mass: m _5 ¼m

ð70Þ

_ 6 h6 Energy: m _ 5 h5 ¼ m

ð71Þ

_ 6 s6 Entropy: m _ 5 s5 þ S_ gen;ev ¼ m

ð72Þ

_ D;ev _ 6 ex 6 þ Ex Exergy: m _ 5 ex 5 ¼ m

ð73Þ

Evaporator: The mass, energy, entropy, and exergy balance equations for the evaporator can be given under the steady state and steady flow conditions as follows: _7; m _ 13 _ 12 ¼ m Mass: m _6¼m

ð74Þ

_ 12 h12 ¼ m _ 7 h7 þ m _ 13 h13 Energy: m _ 6 h6 þ m

ð75Þ

_ 12 s12 þ S_ gen;eva ¼ m _ 7 s7 þ m _ 13 s13 Entropy: m _ 6 s6 þ m

ð76Þ

_ D;eva _ 12 ex12 ¼ m _ 7 ex 7 þ m _ 13 ex 13 þ Ex Exergy: m _ 6 ex6 þ m

ð77Þ

Pump: For the pump of thermomechanical cooling system, the balance equations are provided under the steady state and steady flow conditions as follows: _9 Mass: m _8 ¼m

ð78Þ

_ p¼m _ 9 h9 Energy: m _ 8 h8 þ W

ð79Þ

_ 9 s9 Entropy: m _ 8 s8 þ S_ gen;p ¼ m

ð80Þ

_ p ¼m _ D;p _ 9 ex 9 þ Ex Exergy: m _ 8 ex 8 þ W

ð81Þ

The residential heating application by using geothermal resource is one of the most common and widespread direct uses of geothermal energy. In addition to that, the space or greenhouse heating is one of the oldest direct uses of geothermal resources. Recently, the district heating application is designed to supply space heating to multiple consumers from a single geofluid production well or from multiple wells or fields. In the world, the first space heating by using geothermal energy was in Chaude Aigues in France in the 14th century. The first municipal district heating process by geothermal energy was installed in Reykjavik in Iceland in 1930. Nowadays, the geothermal energy based district heating process has been successfully installed in both developed and developing countries, for example, USA, France, Romania, Canada, Italy, Iceland, and more recently Japan, New Zealand, China, and Turkey. The geothermal energy based district heating application process is illustrated in Fig. 9. The geothermal district

Geothermal Energy Conversion

3 1 Geothermal brine

6

HEX I

HEX II

Thermal energy production

Thermal energy distribution

5 2

4

Pump I Production well

489

b HEX III

Thermal energy 7 consumption

8

a

Pump II

Injection well

Fig. 9 Schematic diagram of geothermal energy based heating system.

heating system consists generally of three cycles, i.e., the (1) thermal energy production cycle, (2) thermal energy distribution cycle, and (3) thermal energy consumption cycle. The mass, energy, entropy, and exergy balance equations for geothermal energy based heating system components are written in the next subsections. Heat exchanger I: The mass, energy, entropy, and exergy balance equations for HEX I can be expressed under the steady state and steady flow conditions as follows: _2; m _5 Mass: m _1 ¼m _3¼m

ð82Þ

_ 5 h5 ¼ m _ 2 h2 þ m _ 3 h3 Energy: m _ 1 h1 þ m

ð83Þ

_ 5 s5 þ S_ gen;HEX_I ¼ m _ 2 s2 þ m _ 3 s3 Entropy: m _ 1 s1 þ m

ð84Þ

_ D;HEX_I _ 5 ex 5 ¼ m _ 2 ex2 þ m _ 3 ex 3 þ Ex Exergy: m _ 1 ex1 þ m

ð85Þ

Heat exchanger II: The mass, energy, entropy, and exergy balance equations for HEX II under the steady state and steady flow conditions can be written as follows: _4; m _8 Mass: m _3 ¼m _6¼m

ð86Þ

_ 8 h8 ¼ m _ 3 h3 þ m _ 6 h6 Energy: m _ 5 h5 þ m

ð87Þ

_ 8 s8 þ S_ gen;HEX_II ¼ m _ 3 s3 þ m _ 6 s6 Entropy: m _ 5 s5 þ m

ð88Þ

_ D;HEX_II _ 8 ex 8 ¼ m _ 3 ex3 þ m _ 6 ex 6 þ Ex Exergy: m _ 5 ex5 þ m

ð89Þ

Heat exchanger III: The mass, energy, entropy, and exergy balance equations for HEX III can be defined under the steady state and steady flow conditions as follows: _7; m _b Mass: m _6¼m _a ¼m

ð90Þ

_ b hb ¼ m _ a ha þ m _ 7 h7 Energy: m _ 6 h6 þ m

ð91Þ

_ a sa þ S_ gen;HEX_III ¼ m _ a sa þ m _ 7 s7 Entropy: m _ 6 s6 þ m

ð92Þ

_ D;HEX_III _ a ex a ¼ m _ 7 ex 7 þ m _ b exb þ Ex Exergy: m _ 6 ex 6 þ m

ð93Þ

Pump I: The mass, energy, entropy, and exergy balance equations for pump I can be written under the steady state and steady flow conditions as follows: _5 Mass: m _4 ¼m

ð94Þ

_ p_I ¼ m _ 5 h5 Energy: m _ 4 h4 þ W

ð95Þ

_ 5 s5 Entropy: m _ 4 s4 þ S_ gen;p_I ¼ m

ð96Þ

_ p_I ¼ m _ D;p_I _ 5 ex 5 þ Ex Exergy: m _ 4 ex 4 þ W

ð97Þ

Pump II: The mass, energy, entropy, and exergy balance equations for pump II can be defined under the steady state and steady flow conditions as follows: _8 Mass: m _7 ¼m

ð98Þ

490

Geothermal Energy Conversion _ p_II ¼ m _ 8 h8 Energy: m _ 7 h7 þ W

ð99Þ

_ 8 s8 Entropy: m _ 7 s7 þ S_ gen;p_II ¼ m

ð100Þ

_ p_II ¼ m _ D;p_II _ 8 ex 8 þ Ex Exergy: m _ 7 ex 7 þ W

ð101Þ

The geothermal resources for heat pump processes have recently been used in different countries, such as United States, Canada, and France. Unfortunately, the geothermal energy resources are usually localized and do not generally coincide with fields of high population density. Also, the geofluid often has a high salt content that leads to some difficulties with the HEXes. Due to the high and constant temperatures of these geofluids, the efficiency rates are usually high. The energy demand of buildings has one of the highest energy supplies in the world, accounting for one-quarter to one-third of total energy consumption rate and the parallel amount of harmful gaseous emissions. The ground source heat pump process, also called the geothermal heat pump process, utilizes the heat thermal energy stored underground, providing different useful outputs, such as heating, cooling, and hot water for residential applications. The schematic diagrams of ground source heat pump systems, such as (1) ground to water, and (2) ground to air are illustrated in Fig. 10(A) and (B), respectively. The geothermal heat pump process has become increasingly common in different residential and commercial buildings because of its higher coefficient of performance (COP) and easy installation. Based on the American Society of Heating, Refrigerating, and Air-Conditioning Engineers (ASHRAE) [23], the geothermal heat pump process is the most energy-efficient system among heating and cooling options. The schematic diagram of the geothermal heat pump process for heating application is illustrated in Fig. 11. As seen from Fig. 11, to produce electricity for pumps and to collect thermal energy for improving the working potential of evaporator subcomponent, the Photovoltaic/thermal collector (PV/T) evaporator should be utilized in geothermal energy based heat pump design. This geothermal heat pump system can be divided into four subcycles: (1) radiator

Geothermal brine

Water HEX I

HEX II

Heat pump

Ground

(A)

Geothermal brine

Air HEX I

(B)

HEX II

Heat pump

Ground

Fig. 10 Some types of ground source heat pumps, namely (A) ground to water and (B) ground to air.

Expansion valve 9 Radiator heating system 5 6

12

4

3

HEX

Condenser

Cooling tower

7

Pump I

2

PV-T evaporator

14

1

13

10 8

Compressor

Pump III

Pump II

11 Ground level Ground heat exchanger

Fig. 11 Schematic diagram of geothermal energy based heat pump process.

Geothermal Energy Conversion

491

heating, (2) PV/T evaporator, (3) ground HEX, and (4) cooling tower. The cooling tower subcomponent is utilized from the piping loops of the geothermal heat pump process with the plate HEXs for increasing thermal energy transfer rate from the cooling tower to the ground (see Ref. [24] for details). The mass, energy, entropy, and exergy balance equations for geothermal energy based heat pump system components are written in the following subsections: Radiator heating system: The mass, energy, entropy, and exergy balance equations for radiator heating system can be expressed under the steady state and steady flow conditions as follows: _6 Mass: m _5 ¼m

ð102Þ

_ rhs _ 6 h6 þ Q Energy: m _ 5 h5 ¼ m

ð103Þ

_ rhs =Trhs _ 6 s6 þ Q Entropy: m _ 5 s5 þ S_ gen;rhs ¼ m

ð104Þ

_ Q þ Ex _ D;rhs _ 6 ex6 þ Ex Exergy: m _ 5 ex 5 ¼ m rhs

ð105Þ

Pump I: The mass, energy, entropy, and exergy balance equations for pump I can be written under the steady state and steady flow conditions as follows: _7 Mass: m _6 ¼m

ð106Þ

_ p_I ¼ m _ 7 h7 Energy: m _ 6 h6 þ W

ð107Þ

_ 7 s7 Entropy: m _ 6 s6 þ S_ gen;p_I ¼ m

ð108Þ

_ p_I ¼ m _ D;p_I _ 7 ex 7 þ Ex Exergy: m _ 6 ex 6 þ W

ð109Þ

Condenser: The mass, energy, entropy, and exergy balance equations for condenser can be defined under the steady state and steady flow conditions as follows: _ 7; m _3 Mass: m _5¼m _2 ¼m

ð110Þ

_ 7 h7 ¼ m _ 3 h3 þ m _ 5 h5 Energy: m _ 2 h2 þ m

ð111Þ

_ 7 s7 þ S_ gen;con ¼ m _ 3 s3 þ m _ 5 s5 Entropy: m _ 2 s2 þ m

ð112Þ

_ D;con _ 7 ex 7 ¼ m _ 3 ex3 þ m _ 5 ex 5 þ Ex Exergy: m _ 2 ex2 þ m

ð113Þ

Expansion valve: The mass, energy, entropy, and exergy balance equations for expansion valve can be written under the steady state and steady flow conditions as follows: _4 Mass: m _3 ¼m

ð114Þ

_ 4 h4 Energy: m _ 3 h3 ¼ m

ð115Þ

_ 4 s4 Entropy: m _ 3 s3 þ S_ gen;ev ¼ m

ð116Þ

_ D;ev _ 4 ex 4 þ Ex Exergy: m _ 3 ex 3 ¼ m

ð117Þ

Compressor: The mass, energy, entropy, and exergy balance equations for compressor can be given under the steady state and steady flow conditions as follows: _2 Mass: m _1 ¼m

ð118Þ

_ cmp ¼ m _ 2 h2 Energy: m _ 1 h1 þ W

ð119Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;cmp ¼ m

ð120Þ

_ cmp ¼ m _ D;cmp _ 2 ex 2 þ Ex Exergy: m _ 1 ex 1 þ W

ð121Þ

PV/T evaporator: The mass, energy, entropy, and exergy balance equations for the PV/T evaporator can be defined under the steady state and steady flow conditions as follows: _4; m _9 Mass: m _1 ¼m _8¼m

ð122Þ

_ PVT ¼ m _ PVT _ 8 h8 þ Q _ 1 h1 þ m _ 9 h9 þ W Energy: m _ 4 h4 þ m

ð123Þ

_ PVT =TPVT þ S_ gen;PVT_eva ¼ m _ 8 s8 þ Q _ 1 s1 þ m _ 9 s9 Entropy: m _ 4 s4 þ m

ð124Þ

492

Geothermal Energy Conversion _ Q ¼m _ PVT þ Ex _ D;PVT_eva _ 8 ex 8 þ Ex _ 1 ex1 þ m _ 9 ex9 þ W Exergy: m _ 4 ex4 þ m PVT

ð125Þ

Heat exchanger: The mass, energy, entropy, and exergy balance equations for HEX can be written under the steady state and steady flow conditions as follows: _ 10 ; m _ 14 Mass: m _9¼m _ 12 ¼ m

ð126Þ

_ 14 h14 ¼ m _ 10 h10 þ m _ 12 h12 Energy: m _ 9 h9 þ m

ð127Þ

_ 14 s14 þ S_ gen;HEX ¼ m _ 10 s10 þ m _ 12 s12 Entropy: m _ 9 s9 þ m

ð128Þ

_ D;HEX _ 14 ex14 ¼ m _ 10 ex10 þ m _ 12 ex 12 þ Ex Exergy: m _ 9 ex9 þ m

ð129Þ

Cooling tower: The mass, energy, entropy, and exergy balance equations for the cooling tower can be given under the steady state and steady flow conditions as follows: _ 13 Mass: m _ 12 ¼ m

ð130Þ

_ ct _ 13 h13 þ Q Energy: m _ 12 h12 ¼ m

ð131Þ

_ ct =Tct _ 13 s13 þ Q Entropy: m _ 12 s12 þ S_ gen;ct ¼ m

ð132Þ

_ Q þ Ex _ D;ct _ 13 ex 13 þ Ex Exergy: m _ 12 ex 12 ¼ m ct

ð133Þ

Pump II: The mass, energy, entropy, and exergy balance equations for pump II can be written under the steady state and steady flow conditions as follows: _ 14 Mass: m _ 13 ¼ m

ð134Þ

_ p_II ¼ m _ 14 h14 Energy: m _ 13 h13 þ W

ð135Þ

_ 14 s14 Entropy: m _ 13 s13 þ S_ gen;p_II ¼ m

ð136Þ

_ p_II ¼ m _ D;p_II _ 14 ex 14 þ Ex Exergy: m _ 13 ex13 þ W

ð137Þ

Pump III: The mass, energy, entropy, and exergy balance equations for pump III can be defined under the steady state and steady flow conditions as follows: _ 11 Mass: m _ 10 ¼ m

ð138Þ

_ p_III ¼ m _ 11 h11 Energy: m _ 10 h10 þ W

ð139Þ

_ 11 s11 Entropy: m _ 10 s10 þ S_ gen;p_III ¼ m

ð140Þ

_ p_III ¼ m _ D;p_III _ 11 ex 11 þ Ex Exergy: m _ 10 ex10 þ W

ð141Þ

Ground heat exchanger: The mass, energy, entropy, and exergy balance equations for the ground HEX can be given under the steady state and steady flow conditions as follows: _ 11 Mass: m _8¼m

ð142Þ

_ g_HEX ¼ m _ 8 h8 Energy: m _ 11 h11 þ Q

ð143Þ

_ g_HEX =Tg_HEX þ S_ gen;g_HEX ¼ m _ 8 s8 Entropy: m _ 11 s11 þ Q

ð144Þ

_ Q _ 8 ex 8 þ E Exergy: m _ 11 ex11 þ Ex g_HEX ¼ m

ð145Þ

The COP is the most general measure of heat pump performance and is defined as the ratio of the product heat output of the heat pump system to its power energy input. The equation form of the heat pump COP can be written as follows: COPHP ¼ Product heat output=Electrical energy input

ð146Þ

The ground source heat pumps generally have COPs ranging from 3 to 5, representing that they deliver 3–5 times more thermal energy than they consume in terms of power energy.

Geothermal Energy Conversion 4.11.5.2

493

Power Generation

The low and high temperature geofluid resources have a very high potential as alternative energy resources for power generation. A schematic diagram of the power generation cycles based on geothermal energy and some potential options of energy storage solution, such as (1) thermal energy (storage heaters and molten salts), (2) electrostatic (capacitors and supercapacitors), (3) potential (pumped hydro and compressed air), (4) kinetic (flywheels), (5) chemicals (batteries, methanol, regenerative fuel cell, and hydrogen), and (6) electromagnetics (superconducting coils), is illustrated in Fig. 12. The amount of geothermal energy is huge, but due to the lower temperature levels, the power conversion performance is lower than other common power generation processes. Therefore, different types of geothermal power generation process are proposed in the literature, such as direct steam power generation, single flash steam power generation, double flash power generation, triple flash power generation, binary cycle power generation, combined/hybrid power generation, ORC and Kalina cycle. These can also be classified as open cycles (Fig. 13(A) and (B)), closed cycles (Fig. 13(C)), and combined cycles (Fig. 13(D)). The open cycles avoid the evaporator. But, the produced low-pressured steam needs very large diameter steam turbines. This negative effect can be accomplished by using the closed or combined cycles. Except for the ORC and Kalina cycle, the other geothermal power processes are given in the case studies section. The ORC processes are similar to the Rankine process, but utilize the organic working fluids instead of steam. Also, the Kalina cycles are similar to the ORC processes, but utilize the chemical composition of H2O and NH3 with the correct mixture as working fluid. The schematic diagrams of the geothermal energy based ORC process and Kalina cycle are shown in Figs. 14 and 15, respectively. The mass, energy, entropy, and exergy balance equations for geothermal energy based ORC system components are written in the next subsections. Heat exchanger: The mass, energy, entropy, and exergy balance equations for HEX can be written under the steady state and steady flow conditions as follows: _ 2; m _6 _3 ¼m Mass: m _1¼m

ð147Þ

_ 6 h6 ¼ m _ 2 h2 þ m _ 3 h3 Energy: m _ 1 h1 þ m

ð148Þ

_ 6 s6 þ S_ gen;HEX ¼ m _ 2 s2 þ m _ 3 s3 Entropy: m _ 1 s1 þ m

ð149Þ

_ D;HEX _ 6 ex 6 ¼ m _ 2 ex2 þ m _ 3 ex 3 þ Ex Exergy: m _ 1 ex1 þ m

ð150Þ

Turbine: Under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for turbine can be defined as follows: _4 Mass: m _3 ¼m

ð151Þ

Geothermal energy

Heat/steam

Power generation cycle

Electrical energy

Storage options

Thermal; • Storage heaters • Molten salts

Potential; • Pumped hydro • Compressed air

Kinetic; • Flywheels

Electrostatic ; • Capacitors • Supercapacitors

Fig. 12 Electricity generation from geothermal energy and storage options.

Electromagnetic; • Super conducting coils

Chemicals; • Batteries • Methanol • Regenerative fuel cell • Hydrogen

494

Geothermal Energy Conversion

Particulate matter

Electricity

Particulate remover

Injection well

Condenser

Turbine

Geofluid (A)

Cold water

Production well Particulate matter

Particulate remover

Hot water

Electricity

Flash separator

Condenser

Turbine

Injection well

Geofluid (B)

Production well

Cold water

Hot water

Hot water

Electricity Cold water

Production well Geofluid

Turbine

Evaporator

Evaporator

Injection well

Pump Hot water Electricity

(C)

Particulate matter Particulate remover

Electricity Flash separator

Condenser

Turbine

Injection well

Geofluid Production well

Hot water

Cold water

Hot water

Electricity Cold water Turbine

Evaporator

Evaporator

Pump Injection well (D)

Hot water Electricity

Fig. 13 Geothermal power generation systems, including: (A) direct steam cycle, (B) flash steam cycle, (C) binary cycle, and (D) combined cycle.

Geothermal Energy Conversion

495

3

1 HEX

Power Turbine

Geothermal brine

6 Pump

2

4

5 Condenser Production well

a

b

Injection well Fig. 14 Schematic diagram of geothermal energy based ORC process.

1

3 Turbine Power HEX 4 8 Internal-HEX

Production well

2 5

7 Injection well

6

10 Condenser

Pump

9

Fig. 15 Schematic diagram of geothermal energy based Kalina cycle.

_ tur _ 4 h4 þ W Energy: m _ 3 h3 ¼ m

ð152Þ

_ 4 s4 Entropy: m _ 3 s3 þ S_ gen;tur ¼ m

ð153Þ

_ tur þ Ex _ D;tur _ 4 ex 4 þ W Exergy: m _ 3 ex 3 ¼ m

ð154Þ

Condenser: The mass, energy, entropy, and exergy balance equations are given for the condenser under the steady state and steady flow conditions. _ 5; m _b Mass: m _4 ¼m _a ¼m

ð155Þ

_ a ha ¼ m _ 5 h5 þ m _ b hb Energy: m _ 4 h4 þ m

ð156Þ

_ ¼m _ a sa þ SE _ 5 s5 þ m _ b sb Entropy: m _ 4 s4 þ m

ð157Þ

_ D;con _ a ex a ¼ m _ 5 ex 5 þ m _ b exb þ Ex Exergy: m _ 4 ex 4 þ m

ð158Þ

Pump: The mass, energy, entropy, and exergy balance equations for pump can be defined under the steady state and steady flow conditions as follows: _6 Mass: m _5 ¼m

ð159Þ

496

Geothermal Energy Conversion _ p¼m _ 6 h6 Energy: m _ 5 h5 þ W

ð160Þ

_ 6 s6 Entropy: m _ 5 s5 þ S_ gen;p ¼ m

ð161Þ

_ p ¼m _ D;p _ 6 ex 6 þ Ex Exergy: m _ 5 ex 5 þ W

ð162Þ

Heat exchanger: The mass, energy, entropy and exergy balance equations for HEX can be written under the steady state and steady flow conditions as follows: _ 2; m _8 Mass: m _1¼m _3 ¼m

ð163Þ

_ 8 h8 ¼ m _ 2 h2 þ m _ 3 h3 Energy: m _ 1 h1 þ m

ð164Þ

_ 8 s8 þ S_ gen;HEX ¼ m _ 2 s2 þ m _ 3 s3 Entropy: m _ 1 s1 þ m

ð165Þ

_ D;HEX _ 8 ex 8 ¼ m _ 2 ex2 þ m _ 3 ex 3 þ Ex Exergy: m _ 1 ex1 þ m

ð166Þ

Turbine: Under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for the turbine can be defined as follows: _4 Mass: m _3 ¼m

ð167Þ

_ tur _ 4 h4 þ W Energy: m _ 3 h3 ¼ m

ð168Þ

_ 4 s4 Entropy: m _ 3 s3 þ S_ gen;tur ¼ m

ð169Þ

_ tur þ Ex _ D;tur _ 4 ex 4 þ W Exergy: m _ 3 ex 3 ¼ m

ð170Þ

The mass, energy, entropy, and exergy balance equations for geothermal energy based Kalina cycle components are defined in the next subsections. Internal-HEX: The mass, energy, entropy, and exergy balance equations for internal-HEX can be written under the steady state and steady flow conditions as follows: _ 8; m _5 Mass: m _7¼m _4 ¼m

ð171Þ

_ 7 h7 ¼ m _ 5 h5 þ m _ 8 h8 Energy: m _ 4 h4 þ m

ð172Þ

_ 7 s7 þ S_ gen;I_HEX ¼ m _ 5 s5 þ m _ 8 s8 Entropy: m _ 4 s4 þ m

ð173Þ

_ D;I_HEX _ 7 ex 7 ¼ m _ 5 ex5 þ m _ 8 ex 8 þ Ex Exergy: m _ 4 ex4 þ m

ð174Þ

Condenser: The mass, energy, entropy, and exergy balance equations are given for the condenser under the steady state and steady flow conditions. _ 6; m _b Mass: m _5 ¼m _a ¼m

ð175Þ

_ a ha ¼ m _ 6 h6 þ m _ b hb Energy: m _ 5 h5 þ m

ð176Þ

_ a sa þ S_ gen;con ¼ m _ 6 s6 þ m _ b sb Entropy: m _ 5 s5 þ m

ð177Þ

_ D;con _ a ex a ¼ m _ 6 ex 6 þ m _ b exb þ Ex Exergy: m _ 5 ex 5 þ m

ð178Þ

Pump: The mass, energy, entropy, and exergy balance equations for the pump can be defined under the steady state and steady flow conditions as follows:

4.11.5.3

_7 Mass: m _6 ¼m

ð179Þ

_ p¼m _ 7 h7 Energy: m _ 6 h6 þ W

ð180Þ

_ 7 s7 Entropy: m _ 6 s6 þ S_ gen;p ¼ m

ð181Þ

_ p ¼m _ D;p _ 7 ex 7 þ Ex Exergy: m _ 6 ex 6 þ W

ð182Þ

Hydrogen Energy Production

The production pathway of hydrogen with carbon based sources may be the main process in the near future. But, the fossil fuel limitations and ecological harm of steam conversion of methane are stimulating the improvement of renewable energy based

Geothermal Energy Conversion

497

hydrogen. Nowadays, between 200 and 2501C of thermal energy source input in the hydrogen generation system integrated with the geothermal power resource is feasible, but geothermal energy based hydrogen production processes will change within the next few decades. A brief overview of geothermal energy based hydrogen production potential options is illustrated in Fig. 16. In many developed and developing countries, the geothermal processes are being considered as a primary energy source for producing hydrogen energy, because geothermal technology provides a reliable energy supply and is relatively benign environmentally. On the other hand, the hydrogen production and utilization technologies can be integrated with geothermal energy resources and stand-alone power processes. The thermochemical hydrogen production process was first studied at the end of the 1960s as an alternative and potentially more efficient way to generate hydrogen from water. Thermochemical hydrogen production cycles consist of a sequence of chemical reactions in which water is thermally decomposed into hydrogen and oxygen (see Eq. (183)) and also all other chemicals entering the chemical reactions are recycled. Only the heat/electricity energy and water are consumed in the thermochemical reaction. 2H2 O þ Energy-2H2 þ O2

ð183Þ

As seen in Figs. 17 and 18, the thermochemical hydrogen production cycles use only thermal energy or the combination of heat and electricity energy. While the thermochemical reaction temperature range remains as a constraint for the geothermal resource utilization for hydrogen production, there are some different options to upgrade the heat energy and allow such processes in a more suitable way to operate chemical cycles for hydrogen generation applications. The thermochemical cycle temperatures are the most important indicators for thermochemical based hydrogen generation purposes. Therefore, the optimization application of heat flows is a significant key factor for thermal to hydrogen conversion performance. The thermochemical reaction temperatures in the hydrogen production range actually from 1001C to 30001C based on the number of reaction cycles. The schematic diagram of hydrogen production and liquefaction system driven by geothermal energy is presented in Fig. 19. The geothermal energy based integrated system investigated in this chapter consists of mainly four subsystems: (1) geothermal

Geothermal energy

Thermal energy

Thermochemical processes

Electrical energy

Hybrid process

Direct sources

Electrolysis

Hydrogen energy

Fig. 16 Potential options of geothermal energy based hydrogen production.

Water Geofluid thermal energy

Hydrogen Thermochemical hydrogen production cycle Oxygen

Fig. 17 Thermochemical water decomposition system for hydrogen production from geothermal energy resources.

Geofluid thermal energy

Geofluid thermal energy

Water Hydrogen

Electricity generation

Electricity

Hybrid thermochemical hydrogen production cycle Oxygen

Fig. 18 Hybrid thermochemical water decomposition system for hydrogen production from geothermal energy resources.

498

Geothermal Energy Conversion

4 HEX I

1

Power Turbine

Geothermal brine

7

2

Pump

5

6 Condenser Production well

8 Electrolysis water

b

a Mixer PEM electrolyzer preheating 3

9

PEM electrolyzer

11

Compressor 13

12

N2(gas)

14

25

HEX III

HEX II

24

N2(liq) 10

15 22

23 Oxygen

HEX IV Injection well

16

21 Liquid hydrogen tank

20

28

HEX V

HEX VI

Separator 19

18

Expansion valve

17

N2(gas)

27 26 N2(liq)

Fig. 19 Schematic diagram of geothermal energy based hydrogen production and liquefaction system.

cycle, (2) ORC, (3) proton exchange membrane (PEM) electrolyzer, and (4) hydrogen liquefaction process. The geothermal heat energy is used in the integrated system to produce heat and power for the PEM electrolyzer. The geofluid transfers its heat energy to the ORC subsystem before entering the PEM electrolyzer preheater. The mass, energy, entropy, and exergy balance equations for geothermal energy based hydrogen production and liquefaction system components are written in the next subsections. Heat exchanger I: The mass, energy, entropy, and exergy balance equations for HEX I can be written under the steady state and steady flow conditions as follows: _ 2; m _7 Mass: m _1¼m _4 ¼m

ð184Þ

_ 7 h7 ¼ m _ 2 h2 þ m _ 4 h4 Energy: m _ 1 h1 þ m

ð185Þ

_ 7 s7 þ S_ gen;HEX_I ¼ m _ 2 s2 þ m _ 4 s4 Entropy: m _ 1 s1 þ m

ð186Þ

_ D;HEX_I _ 7 ex 7 ¼ m _ 2 ex2 þ m _ 4 ex 4 þ Ex Exergy: m _ 1 ex1 þ m

ð187Þ

Turbine: Under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for turbine can be defined as follows: _5 Mass: m _4 ¼m

ð188Þ

_ tur _ 5 h5 þ W Energy: m _ 4 h4 ¼ m

ð189Þ

_ 5 s5 Entropy: m _ 4 s4 þ S_ gen;tur ¼ m

ð190Þ

_ tur þ Ex _ D;tur _ 5 ex 5 þ W Exergy: m _ 4 ex 4 ¼ m

ð191Þ

Geothermal Energy Conversion

499

Condenser: The mass, energy, entropy, and exergy balance equations are given for the condenser under the steady state and steady flow conditions. _ 6; m _b _a ¼m Mass: m _5 ¼m

ð192Þ

_ a ha ¼ m _ 6 h6 þ m _ b hb Energy: m _ 5 h5 þ m

ð193Þ

_ a sa þ S_ gen;con ¼ m _ 6 s6 þ m _ b sb Entropy: m _ 5 s5 þ m

ð194Þ

_ D;con _ a ex a ¼ m _ 6 ex 6 þ m _ b exb þ Ex Exergy: m _ 5 ex 5 þ m

ð195Þ

Pump: The mass, energy, entropy, and exergy balance equations for the pump can be defined under the steady state and steady flow conditions as follows: _7 Mass: m _6 ¼m

ð196Þ

_ p¼m _ 7 h7 Energy: m _ 6 h6 þ W

ð197Þ

_ 7 s7 Entropy: m _ 6 s6 þ S_ gen;p ¼ m

ð198Þ

_ p ¼m _ D;p _ 7 ex 7 þ Ex Exergy: m _ 6 ex 6 þ W

ð199Þ

Proton exchange membrane electrolyzer preheating: The mass, energy, entropy, and exergy balance equations for PEM electrolyzer preheating can be written under the steady state and steady flow conditions as follows: _ 3; m _9 _8 ¼m Mass: m _2¼m

ð200Þ

_ 8 h8 ¼ m _ 3 h3 þ m _ 9 h9 Energy: m _ 2 h2 þ m

ð201Þ

_ 8 s8 þ S_ gen;peph ¼ m _ 3 s3 þ m _ 9 s9 Entropy: m _ 2 s2 þ m

ð202Þ

_ D;peph _ 8 ex 8 ¼ m _ 3 ex3 þ m _ 9 ex 9 þ Ex Exergy: m _ 2 ex2 þ m

ð203Þ

Proton exchange membrane electrolyzer: The mass, energy, entropy, and exergy balance equations for the PEM electrolyzer can be written under the steady state and steady flow conditions as follows: _ 10 þ m _ 11 Mass: m _9¼m

ð204Þ

_ T ¼m _ 10 h10 þ m _ 11 h11 Energy: m _ 9 h9 þ W

ð205Þ

_ 10 s10 þ m _ 11 s11 Entropy: m _ 9 s9 þ S_ gen;PEM_el ¼ m

ð206Þ

_ T ¼m _ D;PEM_el _ 10 ex10 þ m _ 11 ex11 þ Ex Exergy: m _ 9 ex 9 þ W

ð207Þ

Mixer: The mass, energy, entropy, and exergy balance equations for the mixer can be defined under the steady state and steady flow conditions as follows: _ 23 ¼ m _ 12 Mass: m _ 11 þ m

ð208Þ

_ 23 h23 ¼ m _ 12 h12 Energy: m _ 11 h11 þ m

ð209Þ

_ 23 s23 þ S_ gen;mixer ¼ m _ 12 s12 Entropy: m _ 11 s11 þ m

ð210Þ

_ D;mixer _ 23 ex 23 ¼ m _ 12 ex 12 þ Ex Exergy: m _ 11 ex 11 þ m

ð211Þ

Compressor: The mass, energy, entropy, and exergy balance equations for the compressor can be given under the steady state and steady flow conditions as follows: _ 13 Mass: m _ 12 ¼ m

ð212Þ

_ cmp ¼ m _ 13 h13 Energy: m _ 12 h12 þ W

ð213Þ

_ 13 s13 Entropy: m _ 12 s12 þ S_ gen;cmp ¼ m

ð214Þ

_ cmp ¼ m _ D;cmp _ 13 ex 13 þ Ex Exergy: m _ 12 ex 12 þ W

ð215Þ

Heat exchanger II: The mass, energy, entropy, and exergy balance equations for HEX II can be written under the steady state and steady flow conditions as follows: _ 14 ; m _ 23 Mass: m _ 13 ¼ m _ 22 ¼ m

ð216Þ

500

Geothermal Energy Conversion _ 22 h22 ¼ m _ 14 h14 þ m _ 23 h23 Energy: m _ 13 h13 þ m

ð217Þ

_ 22 s22 þ S_ gen;HEX_II ¼ m _ 14 s14 þ m _ 23 s23 Entropy: m _ 13 s13 þ m

ð218Þ

_ D;HEX_II _ 22 ex 22 ¼ m _ 14 ex 14 þ m _ 23 ex 23 þ Ex Exergy: m _ 13 ex13 þ m

ð219Þ

Heat exchanger III: The mass, energy, entropy, and exergy balance equations for HEX III can be defined under the steady state and steady flow conditions as follows: _ 15 ; m _ 25 Mass: m _ 14 ¼ m _ 24 ¼ m

ð220Þ

_ 24 h24 ¼ m _ 15 h15 þ m _ 25 h25 Energy: m _ 14 h14 þ m

ð221Þ

_ 24 s24 þ S_ gen;HEX_III ¼ m _ 15 s15 þ m _ 25 s25 Entropy: m _ 14 s14 þ m

ð222Þ

_ D;HEX_III _ 24 ex 24 ¼ m _ 15 ex15 þ m _ 25 ex 25 þ Ex Exergy: m _ 14 ex14 þ m

ð223Þ

Heat exchanger IV: The mass, energy, entropy, and exergy balance equations for HEX IV can be written under the steady state and steady flow conditions as follows: _ 16 ; m _ 22 Mass: m _ 15 ¼ m _ 21 ¼ m

ð224Þ

_ 21 h21 ¼ m _ 16 h16 þ m _ 22 h22 Energy: m _ 15 h15 þ m

ð225Þ

_ 21 s21 þ S_ gen;HEX_IV ¼ m _ 16 s16 þ m _ 22 s22 Entropy: m _ 15 s15 þ m

ð226Þ

_ D;HEX_IV _ 21 ex21 ¼ m _ 16 ex16 þ m _ 22 ex 22 þ Ex Exergy: m _ 15 ex15 þ m

ð227Þ

Heat exchanger V: The mass, energy, entropy, and exergy balance equations for HEX V can be given under the steady state and steady flow conditions as follows: _ 17 ; m _ 27 Mass: m _ 16 ¼ m _ 26 ¼ m

ð228Þ

_ 26 h26 ¼ m _ 17 h17 þ m _ 27 h27 Energy: m _ 16 h16 þ m

ð229Þ

_ 26 s26 þ S_ gen;HEX_V ¼ m _ 17 s17 þ m _ 27 s27 Entropy: m _ 16 s16 þ m

ð230Þ

_ D;HEX_V _ 26 ex 26 ¼ m _ 17 ex17 þ m _ 27 ex 27 þ Ex Exergy: m _ 16 ex16 þ m

ð231Þ

Heat exchanger VI: The mass, energy, entropy, and exergy balance equations for HEX VI can be defined under the steady state and steady flow conditions as follows: _ 21 ; m _ 18 Mass: m _ 20 ¼ m _ 17 ¼ m

ð232Þ

_ 20 h20 ¼ m _ 18 h18 þ m _ 21 h21 Energy: m _ 17 h17 þ m

ð233Þ

_ 20 s20 þ S_ gen;HEX_VI ¼ m _ 18 s18 þ m _ 21 s21 Entropy: m _ 17 s17 þ m

ð234Þ

_ D;HEX_VI _ 20 ex20 ¼ m _ 18 ex18 þ m _ 21 ex21 þ Ex Exergy: m _ 17 ex17 þ m

ð235Þ

Expansion valve: The mass, energy, entropy, and exergy balance equations for the expansion valve can be written under the steady state and steady flow conditions as follows: _ 19 Mass: m _ 18 ¼ m

ð236Þ

_ 19 h19 Energy: m _ 18 h18 ¼ m

ð237Þ

_ 19 s19 Entropy: m _ 18 s18 þ S_ gen;ev ¼ m

ð238Þ

_ D;ev _ 19 ex19 þ Ex Exergy: m _ 18 ex18 ¼ m

ð239Þ

Separator: The mass, energy, entropy, and exergy balance equations for the separator can be defined under the steady state and steady flow conditions as follows: _ 20 þ m _ 28 Mass: m _ 19 ¼ m

ð240Þ

_ 20 h20 þ m _ 28 h28 Energy: m _ 19 h19 ¼ m

ð241Þ

_ 20 s20 þ m _ 28 s28 Entropy: m _ 19 s19 þ S_ gen;sep ¼ m

ð242Þ

_ D;sep _ 20 ex 20 þ m _ 28 ex 28 þ Ex Exergy: m _ 19 ex 19 ¼ m

ð243Þ

Geothermal Energy Conversion

501

Also, the power generated by using the ORC system is used in the PEM electrolyzer to generate hydrogen. The produced hydrogen is in gaseous form at reference conditions. The hydrogen liquefaction subsystem is used for more efficient hydrogen storage. The hydrogen liquefaction process is relatively more energy intensive than compression of hydrogen, whereas, the density of liquid hydrogen is nearly 1120 kg/m3 and, also liquid hydrogen is 29 times better than compressed hydrogen at 700 bar, in terms of volume work. Therefore, the Linde–Hampson hydrogen liquefaction process with a secondary nitrogen cooling is investigated for hydrogen storage [25,26]. The thermodynamic analysis of a PEM electrolyzer subsystem, which is used in hydrogen production and liquefaction process, is given in the next subsection. Proton exchange membrane electrolyzer: The overall chemical reaction equation of water decomposition in the PEM electrolyzer in Fig. 19 can be written as follows: 1 ð244Þ H2 OðlÞ -H2ðgÞ þ O2ðgÞ 2 where subscript l and g are the liquid and gas phases, respectively. The following reactions take place in the anode and cathode parts of PEM electrolyzer, respectively. 1 H2 OðlÞ - O2ðgÞ þ Hþ ðaqÞ þ 2e 2

ð245Þ

Hþ ðaqÞ þ 2e -2H2

ð246Þ

and

The produced hydrogen and oxygen output flow rates are given as follows, respectively: _ H2;out ¼ J=2F ¼ N_ H2 O N

ð247Þ

N_ O2;out ¼ J=4F

ð248Þ

and

_ H2 O is the water consumed rate in the PEM where J and F are the current density and Faraday constant, respectively, and N electrolyzer. To produce hydrogen from the electrolyzer, the electrical power must be input to the PEM electrolyzer, and this can be written as follows: _ elec ¼ JV E_ elec ¼ Ex

ð249Þ _ elec are the rate of electrical power and electrical exergy input, respectively. V is the cell potential, and can be where E_ elec and Ex given as follows: V ¼ Vo þ Zact;a þ Zact;c þ Zohm

ð250Þ

where Vo is the reversible potential, which is related to the difference in free energy between reactants and products, and can be determined using by the Nernst equation. Zact;a is the activation overpotential of the anode, Zact;c is the activation overpotential of the cathode, and Zohm is the ohmic overpotential of the electrolyte. The ohmic overpotential of the PEM is attached to the resistance of the membrane to hydrogen ions crossing over PEM, and the ohmic overpotential can be defined as follows: Zohm ¼ JR where R is the overall ohmic resistance, and can be expressed as follows: Z D dx R¼ 0 sPEM ½lðxފ

ð251Þ

ð252Þ

where D is the PEM thickness, sPEM ½lðxފ is the local ionic PEM conductivity of the membrane and can be calculated as follows:    1 1 sPEM ½lðxފ ¼ ½0:5139lðxÞ 0:326Šexp 1268 ð253Þ 303 T where x is the distance in the PEM evaluated from the cathode–membrane interface, l(x) is the water content at a location x in the PEM, and can be calculated as follows [27]: lc x þ lc ð254Þ D where la and lc are the water contents at the anode and cathode–membrane interface, respectively. The activation overpotential (Zact), given in the right-hand side of Eq. (240), caused by a deviation from current from its equilibrium, and also e- transfer reaction, should be diversified from the concentration of the oxidized and reduced species. The activation overpotential of the PEM can be given as follows [28,29]:   RT J sinh 1 Zact;i ¼ ; i ¼ a; c ð255Þ F 2Jo;i      azFZact;i ð1 aÞzFZact;i exp ; i ¼ a; c ð256Þ J ¼ Jo;i exp RT RT lðxÞ ¼

la

502

Geothermal Energy Conversion   Eact;i ; i ¼ a; c Jo;i ¼ Jiref exp RT

ð257Þ

where subscripts a and c are the anode and cathode parts of the PEM electrolyzer, respectively; Jo is the exchange current density; a is the charge transfer coefficient for anode and cathode part reactions, and generally equal to ½. z is the number of electrons involved per reaction. For the PEM electrolyzer, z must be equal to 2. Jiref is the preexponential factor and Eact,i is the activation energy for the anode and cathode parts of the PEM electrolyzer.

4.11.5.4

Alternative Fuels

Thermal energy gained from geothermal sources can be used in different ways and one of them is to use thermal energy to produce electrical energy and later in thermochemical cycle in order to produce alternative fuels such as ethanol, methanol, butanol, propane, ammonia, and nonfossil methane, as seen from Fig. 20. These alternative fuels can be stored in liquid phase; then they can be transported to the final usage. These fuels can be used for transportation, space heating, electricity generation, in fuel cells, and as chemicals.

4.11.5.5

Fresh Water Production

In the world, 97.5% of the total water resources are saline water and not suitable for human demands, and only 2.5% are fresh water. Almost 70% of these global freshwater sources are in the polar icecaps, and the major part of the remaining 30% lies in remote underground aquifers. Actually, only a minor fraction of fresh water (less than 1% of total freshwater sources) that is present in rivers, lakes, and reservoirs is easily reachable for direct human utilization. Desalination processes are very energyintensive technology. Therefore, the renewable energy based technologies urgently need to be improved. The low temperature geothermal resources in the world can effectively run the seawater or brackish-water desalination process in order to generate fresh water for residential and irrigation applications. Nowadays, two processes by using the geothermal sources are generally used in freshwater production from saline water. The schematic diagram of the membrane distillation system driven by geothermal energy to generate desalinated water is shown in

Geothermal energy

Thermal energy

Electrical energy

Thermochemical cycle

Alternative fuels • Ethanol • Methanol • Butanol • Propone • Ammonia • Non-fossil methane

Liquid storage

Liquid transport

End usage

Transportation fuels

Space heating

Electricity generation

Fig. 20 Alternative fuels productions based on geothermal energy and usage options.

Fuel cells

Chemical production

Geothermal Energy Conversion

503

Fig. 21. With the use of a membrane in distillation subsystem, the conversion of saline water to drinking water should be ensured with better quality. The schematic diagram of the geothermal energy based distillation process to generate fresh water is illustrated in Fig. 22. The desalination process consists of the pump that pressurizes the saline water, the filter to remove the coarse-grained particles, the energy recovery subcomponent, the osmotic membrane to filter thin-grained particles, and salt and the freshwater tank to store the domestic water. With the utilization of the osmotic membrane, the conversion to domestic water can be provided with better quality. The mass, energy, entropy, and exergy balance equations for geothermal energy based membrane distillation unit components are defined in the next subsection. Heat exchanger: The mass, energy, entropy, and exergy balance equations for HEX can be given under the steady state and steady flow conditions as follows: _ 2; m _6 Mass: m _1¼m _5 ¼m

ð258Þ

HEX

1

5

6

Geothermal brine

Membrane distillation module

2

Pump

Saline water 3

4 Production well

7

8

Drain

Mineralize

9

Fresh water tank

Injection well Fig. 21 Schematic diagram of geothermal source based membrane distillation unit.

1 Geothermal brine

HEX

2 7 Production well

8

Injection well Minerals and chlorine 5

6

9

Pre-treatment 4

RO train I

11

21 20

Pump I 17

Post-treatment RO train II

Pump II

3

12

10

Saline water

14

16

18

Energy recovery device

15 19 Brine discharge

Fig. 22 Schematic diagram of geothermal source based distillation process.

13

22 Fresh water tank

504

Geothermal Energy Conversion _ 5 h5 ¼ m _ 2 h2 þ m _ 6 h6 Energy: m _ 1 h1 þ m

ð259Þ

_ 5 s5 þ S_ gen;HEX ¼ m _ 2 s2 þ m _ 6 s6 Entropy: m _ 1 s1 þ m

ð260Þ

_ D;HEX _ 5 ex 5 ¼ m _ 2 ex2 þ m _ 6 ex 6 þ Ex Exergy: m _ 1 ex1 þ m

ð261Þ

Membrane distillation module: The mass, energy, entropy, and exergy balance equations for the membrane distillation module can be defined under the steady state and steady flow conditions as follows: _6¼m _5þm _7þm _8 Mass: m _4þm

ð262Þ

_ 6 h6 ¼ m _ 5 h5 þ m _ 7 h7 þ m _ 8 h8 Energy: m _ 4 h4 þ m

ð263Þ

_ 6 s6 þ S_ gen;mdm ¼ m _ 5 s5 þ m _ 7 s7 þ m _ 8 h8 Entropy: m _ 4 s4 þ m

ð264Þ

_ D;HEX_V _ 6 ex 6 ¼ m _ 5 ex 5 þ m _ 7 ex 7 þ m _ 8 ex8 þ Ex Exergy: m _ 4 ex 4 þ m

ð265Þ

Pump: The mass, energy, entropy, and exergy balance equations for the pump can be defined under the steady state and steady flow conditions as follows: _4 Mass: m _3 ¼m

ð266Þ

_ p¼m _ 4 h4 Energy: m _ 3 h3 þ W

ð267Þ

_ 4 s4 Entropy: m _ 3 s3 þ S_ gen;p ¼ m

ð268Þ

_ p ¼m _ D;p _ 4 ex 4 þ Ex Exergy: m _ 3 ex 3 þ W

ð269Þ

Mineralizer: The mass, energy, entropy, and exergy balance equations for the mineralizer can be given under the steady state and steady flow conditions as follows: _9 Mass: m _8 ¼m

ð270Þ

_ 9 h9 Energy: m _ 8 h8 ¼ m

ð271Þ

_ 9 s9 Entropy: m _ 8 s8 þ S_ gen;mine ¼ m

ð272Þ

_ D;mine _ 9 ex 9 þ Ex Exergy: m _ 8 ex 8 ¼ m

ð273Þ

The mass, energy, entropy, and exergy balance equations for geothermal energy based distillation process components are given in the next subsections. Heat exchanger: The mass, energy, entropy, and exergy balance equations for HEX can be given under the steady state and steady flow conditions as follows: _ 2; m _8 Mass: m _1¼m _7 ¼m

ð274Þ

_ 7 h7 ¼ m _ 2 h2 þ m _ 8 h8 Energy: m _ 1 h1 þ m

ð275Þ

_ 7 s7 þ S_ gen;HEX ¼ m _ 2 s2 þ m _ 8 s8 Entropy: m _ 1 s1 þ m

ð276Þ

_ D;HEX _ 7 ex 7 ¼ m _ 2 ex2 þ m _ 8 ex 8 þ Ex Exergy: m _ 1 ex1 þ m

ð277Þ

Pretreatment: The mass, energy, entropy, and exergy balance equations for pretreatment can be written under the steady state and steady flow conditions as follows: _7 Mass: m _6 ¼m

ð278Þ

_ 7 h7 Energy: m _ 6 h6 ¼ m

ð279Þ

_ 7 s7 Entropy: m _ 6 s6 þ S_ gen;pt ¼ m

ð280Þ

_ D;pt _ 7 ex7 þ Ex Exergy: m _ 6 ex 6 ¼ m

ð281Þ

Pump I: The mass, energy, entropy, and exergy balance equations for pump I can be defined under the steady state and steady flow conditions as follows: _5 Mass: m _4 ¼m

ð282Þ

_ p_I ¼ m _ 5 h5 Energy: m _ 4 h4 þ W

ð283Þ

Geothermal Energy Conversion

505

_ 5 s5 Entropy: m _ 4 s4 þ S_ gen;p_I ¼ m

ð284Þ

_ p_I ¼ m _ 5 ex 5 þ E Exergy:m _ 4 ex4 þ W

ð285Þ

Pump II: The mass, energy, entropy, and exergy balance equations for pump II can be written under the steady state and steady flow conditions as follows: _ 17 Mass: m _ 16 ¼ m

ð286Þ

_ p_II ¼ m _ 17 h17 Energy: m _ 16 h16 þ W

ð287Þ

_ 17 s17 Entropy: m _ 16 s16 þ S_ gen;p_II ¼ m

ð288Þ

_ p_II ¼ m _ D;p_II _ 17 ex 17 þ Ex Exergy: m _ 16 ex16 þ W

ð289Þ

Energy recovery device: The mass, energy, entropy, and exergy balance equations for the energy recovery device can be given under the steady state and steady flow conditions as follows: _ 16 ; m _ 19 Mass: m _ 15 ¼ m _ 18 ¼ m

ð290Þ

_ 18 h18 ¼ m _ 16 h16 þ m _ 19 h19 Energy: m _ 15 h15 þ m

ð291Þ

_ 18 s18 þ S_ gen;erd ¼ m _ 16 s16 þ m _ 19 s19 Entropy: m _ 15 s15 þ m

ð292Þ

_ D;erd _ 18 ex18 ¼ m _ 16 ex 16 þ m _ 19 ex19 þ Ex Exergy: m _ 15 ex 15 þ m

ð293Þ

RO train I: The mass, energy, entropy, and exergy balance equations for RO train I can be defined under the steady state and steady flow conditions as follows: _ 11 þ m _ 13 Mass: m _9¼m

ð294Þ

_ 11 h11 þ m _ 13 h13 Energy: m _ 9 h9 ¼ m

ð295Þ

_ 11 s11 þ m _ 13 s13 Entropy: m _ 9 s9 þ S_ gen;RO_I ¼ m

ð296Þ

_ D;RO_I _ 11 ex11 þ m _ 13 ex 13 þ Ex Exergy: m _ 9 ex 9 ¼ m

ð297Þ

RO train II: The mass, energy, entropy, and exergy balance equations for RO train II can be written under the steady state and steady flow conditions as follows: _ 12 þ m _ 14 Mass: m _ 10 ¼ m

ð298Þ

_ 12 h12 þ m _ 14 h14 Energy: m _ 10 h10 ¼ m

ð299Þ

_ 12 s12 þ m _ 14 s14 Entropy: m _ 10 s10 þ S_ gen;RO_II ¼ m

ð300Þ

_ D;RO_II _ 12 ex 12 þ m _ 14 ex 14 þ Ex Exergy: m _ 10 ex 10 ¼ m

ð301Þ

Posttreatment: The mass, energy, entropy, and exergy balance equations for posttreatment can be defined under the steady state and steady flow conditions as follows:

4.11.5.6

_ 21 ¼ m _ 22 Mass: m _ 20 þ m

ð302Þ

_ 21 h21 ¼ m _ 22 h22 Energy: m _ 20 h20 þ m

ð303Þ

_ 21 s21 þ S_ gen;pst ¼ m _ 22 s22 Entropy: m _ 20 s20 þ m

ð304Þ

_ D;pst _ 21 ex21 ¼ m _ 22 ex22 þ Ex Exergy: m _ 20 ex20 þ m

ð305Þ

Other Useful Commodities

The primary idea of drying applications of agricultural products is to provide some properties, such as long-term storage without degradation, early harvesting to decrease field losses, higher prices of agricultural products, and better quality. The industrial drying systems are performed with consumption of heat and power for drying the auxiliary equipment. Generally, the drying applications need relatively temperature (401C to 901C) heat energy. Therefore, the low temperature geothermal energy resources can be used to heat the air for drying of agricultural products.

506

Geothermal Energy Conversion

Raw material

Drying process

Pulverizing

Palletizing

Weighing checking

Geothermal heat energy (60−80°C)

Product storage

Fig. 23 Process line for drying grains by using geothermal energy.

Raw material

Washing

Selection

Cutting

Drying process

Weighing checking

Geothermal heat energy (60−80°C)

Product storage

Fig. 24 Process line for drying fruits and vegetables by using geothermal energy.

Power Turbine

5

1

Particulate remover

2

Moisture remover

a

Condenser

3 Particulate matter

Production well

6

4 b

Reinjection well

7

Reinjection well

Fig. 25 Schematic diagram of direct steam geothermal power generation.

The technology of predrying and post production drying processes is different based on the species of dried product, such as grains, vegetables and fruits, and also desired results, such as moisture, shape, and further processing. The common drying lines of grains and fruits/vegetables are illustrated in Figs. 23 and 24, respectively.

4.11.6

Case Studies

Nowadays, the geothermal resource based power generation systems provide an environmentally benign alternative to conventional source based power generation systems. In this chapter, the comprehensive case studies are presented to cover energy and exergy analyses for geothermal energy based power generation systems, such as (1) direct steam power generation, (2) single flash steam power generation, (3) double flash power generation, (4) triple flash power generation, (5) binary cycle power generation, and (6) combined/hybrid power generation, and also (7) geothermal energy based double effect cooling system and (8) geothermal energy based hydrogen production and liquefaction system. In addition, optimization studies are applied. Furthermore, the parametric studies are conducted accordingly to analyze how changing operating parameters, environmental conditions, and state properties impacts the efficiency of geothermal energy based power generation systems.

4.11.6.1

Direct Steam Power Generation

In this case study, the direct steam power generation based on geothermal energy resources is assessed by using thermodynamic assessment. The simplified flow diagram of direct steam power generation in the first case study is illustrated in Fig. 25. As seen in

Geothermal Energy Conversion

507

this figure, steam is separated from the water at the borehole and enters directly through the turbine and the exhaust geofluid going into the reinjection well. This geothermal power process is the first type of geothermal power system, and also comprises nearly a quarter of power production from geothermal energy in the world today [30]. Direct steam geothermal power process has the maximum efficiency rate, between 50% and 70%, and also has low construction cost among all geothermal energy processes. This geothermal power process is very simple, and is commercially preferred, simple to work with, and needs comparatively lower capital cost than other geothermal power processes. But, to obtain maximum power generation from geothermal energy, the flashing pressure should be adjusted for optimum value. As seen in Fig. 25, the pressure level of geothermal fluid from the production well at point 1 must be reduced at fixed enthalpy using the flashing process. The geothermal water from the production well of the geothermal power system enters the particulate remover at point 1, where most of the particulate matter is removed. After that, the working fluid enters the moisture remover (or separator) at point 2, where any moisture that is presented is removed from geothermal fluid. To produce power, the working fluid is sent directly to the turbine at point 5. The outlet stream from the turbine at point 6 is condensed in the condenser unit. Then, the pump increases the pressure level of geothermal fluid at point 7. After leaving the geothermal pump, high pressured geothermal fluid is sent to the reinjection well connecting with water droplets coming from the moisture remover. The thermodynamic balance equations of direct steam geothermal power components are given in the next subsections. Particulate remover: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the particulate remover can be given as follows: _2þm _3 Mass: m _1¼m

ð306Þ

_ 2 h2 þ m _ 3 h3 Energy: m _ 1 h1 ¼ m

ð307Þ

_ 2 s2 þ m _ 3 s3 Entropy: m _ 1 s1 þ S_ gen;pr ¼ m

ð308Þ

_ D;pr _ 2 ex2 þ m _ 3 ex 3 þ Ex Exergy: m _ 1 ex 1 ¼ m

ð309Þ

Moisture remover: The balance equations of the moisture remover under steady state and steady flow conditions are written as follows: _4þm _5 Mass: m _2¼m

ð310Þ

_ 4 h2 þ m _ 5 h5 Energy: m _ 2 h2 ¼ m

ð311Þ

_ 4 s4 þ m _ 5 s5 Entropy: m _ 2 s2 þ S_ gen;mr ¼ m

ð312Þ

_ D;mr _ 4 ex4 þ m _ 5 ex5 þ Ex Exergy: m _ 2 ex2 ¼ m

ð313Þ

Turbine: Under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for the turbine can be defined as follows: _6 Mass: m _5 ¼m

ð314Þ

_ tur _ 6 h6 þ W Energy: m _ 5 h5 ¼ m

ð315Þ

_ 6 s6 Entropy: m _ 5 s5 þ S_ gen;tur ¼ m

ð316Þ

_ tur þ Ex _ D;tur _ 6 ex 6 þ W Exergy: m _ 5 ex 5 ¼ m

ð317Þ

Condenser: The mass, energy, entropy, and exergy balance equations are given for the condenser under the steady state and steady flow conditions. _7; m _b _a ¼m Mass: m _6¼m

ð318Þ

_ a ha ¼ m _ 7 h7 þ m _ b hb Energy: m _ 6 h6 þ m

ð319Þ

_ a sa þ S_ gen;con ¼ m _ 7 s7 þ m _ b sb Entropy: m _ 6 s6 þ m

ð320Þ

_ D;con _ a ex a ¼ m _ 7 ex 7 þ m _ b exb þ Ex Exergy: m _ 6 ex 6 þ m

ð321Þ

As reference conditions, ambient temperature and pressure are taken as 251C and 101.3 kPa, respectively. The assumptions used in the operating conditions of direct steam geothermal power generation system are given in Table 8. The heat and work input/output rate, entropy generation rate, and exergy destruction rates, and energy efficiency and exergy efficiency are calculated from these balance equations and given assumptions.

508

Geothermal Energy Conversion

Table 8 Assumptions for the direct steam geothermal power system Variables

Values

Geofluid source temperature (T1) Geofluid source pressure (P1) _ 1Þ Geofluid mass flow rate ðm Moisture inlet pressure (P2) Turbine output pressure (P6) Geofluid reinjection temperature (T7)

150–2301C 1500 kPa 75 to 225 kg/s 1485 kPa 95 kPa 47.551C

Exergy destruction rate (kW) 1% 1% 36%

62%

Separator

Purifier

Turbine

Condenser

Fig. 26 Exergy destruction rates for the direct steam geothermal power generation system.

9000 8250

0.49

ψD

0.48 0.47

7500 Wturbine (kW)

0.46 6750

0.45 0.44

6000

0.43 0.42

5250

0.41 4500

0.4 0.39

3750 3000 75

System exergy efficiency

0.5 Wturbine

0.38 100

125

150

175

200

0.37 225

mgeothermal (kg/s) Fig. 27 Effect of mass flow rate of geofluid on power generation and exergy efficiency.

The exergy destruction rates of direct steam geothermal power generation system components are analyzed by using the abovegiven procedure, and analysis results are given in Fig. 26. As seen in this figure, the exergy destruction rate is higher in the turbine than in other system parts. In order to better understand the process efficiency, the parametric studies are given to investigate the impacts of different indicators, such as temperature and mass flow rate of geothermal fluid, and ambient temperature, on the process exergy destruction rate and power generation rate. Power generation and exergy efficiency are directly proportional to the geothermal mass flow rate as seen from Fig. 27. While mass flow rate increases from 75 to 225 kg/s, net electricity generation increases from 3000 to 9000 kW. Similarly, the exergy efficiency increases from 37 to 50%, too.

Geothermal Energy Conversion

0.43

5000 Wturbine

0.41 0.4

4000

0.39 0.38

3500

0.37 3000

0.36

System exergy efficiency

0.42

ψD

4500 Wturbine (kW)

509

0.35

2500 150

160

170

180 190 200 Tgeothermal (°C)

210

220

0.34 230

Fig. 28 Effect of geofluid temperature on power generation and exergy efficiency.

0.48

4800 Wturbine ψD

0.46 0.44 0.42

4000 0.4 0.38

3600

0.36 3200

System exergy efficiency

Wturbine (kW)

4400

0.34 0.32

2800 0

5

10

15

20

25

30

35

40

Tambient (°C) Fig. 29 Effect of ambient temperature on power generation and exergy efficiency.

Fig. 28 demonstrates the relation between geothermal fluid temperature and electricity generation at the left side and exergy efficiency at the right. As seen from that figure, increase of geothermal fluid temperature has a positive effect on the system efficiency and the production rate. Electricity production nearly doubles and exergy efficiency increases about 40% with the increment of geothermal fluid temperature. Fig. 29 shows how ambient temperature affects the turbine work or electricity generation and exergy efficiency of the system. As ambient temperature increases from 0 to 401C, electricity generation increases from 2800 kW to almost 4700 kW and exergy efficiency increases from 32% to 47%, respectively. It is apparent that increasing ambient temperature helps decrease the irreversibilities.

4.11.6.2

Single Flash Steam Power Generation

A schematic diagram of a single flash steam geothermal power generation system in the second case study is presented in Fig. 30. This geothermal power system is suitable for two phase resources of low noncondensable gas content. The geothermal working fluid is produced from a production well, and enters a flashing at point 1. After that geothermal fluid is separated into steam and brine in a separator subcomponent. The brine is sent to the reinjection well at point 3. The separated steam at point 4 enters a purifier to improve the quality of working fluid and to avoid different fouling materials. The steam from state 6 runs through the turbine to generate power. The water in the condenser can be utilized to provide heating to the water in state a, as shown in Fig. 30. After leaving the condenser, the geofluid at point 7 is sent to the reinjection well. The geofluid is reinjected without utilizing its thermal energy. The simple design of a single flash geothermal power system allows a wide range of applicable configuration conditions. The balance equations of a single flash steam geothermal power components are written in the next subsections.

510

Geothermal Energy Conversion

6 Purifier Power

5 Turbine 4

Fouling material 7

2

1

b Flashing

Condenser

Separator a 3

8

Production well

Reinjection well

Reinjection well

Fig. 30 Schematic diagram of single flash steam geothermal power generation.

Flashing: The balance equations of flashing subcomponent under steady state and steady flow conditions are written as follows: _2 Mass: m _1 ¼m

ð322Þ

_ 2 h2 Energy: m _ 1 h1 ¼ m

ð323Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;fls ¼ m

ð324Þ

_ D;fls _ 2 ex 2 þ Ex Exergy: m _ 1 ex1 ¼ m

ð325Þ

Separator: The balance equations of the separator under steady state and steady flow conditions can be given as follows: _3þm _4 Mass: m _2¼m

ð326Þ

_ 3 h3 þ m _ 4 h4 Energy: m _ 2 h2 ¼ m

ð327Þ

_ 3 s3 þ m _ 4 s4 Entropy: m _ 2 s2 þ S_ gen;sep ¼ m

ð328Þ

_ D;sep _ 3 ex 3 þ m _ 4 ex4 þ Ex Exergy: m _ 2 ex2 ¼ m

ð329Þ

Purifier: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the purifier subcomponent can be defined as follows: _5þm _6 Mass: m _4¼m

ð330Þ

_ 5 h5 þ m _ 6 h6 Energy: m _ 4 h4 ¼ m

ð331Þ

_ 5 s5 þ m _ 6 s6 Entropy: m _ 4 s4 þ S_ gen;pur ¼ m

ð332Þ

_ D;pur _ 5 ex 5 þ m _ 6 ex 6 þ Ex Exergy: m _ 4 ex 4 ¼ m

ð333Þ

Turbine: Under the steady state and steady flows conditions, the mass, energy, entropy, and exergy balance equations for the turbine can be defined as follows: _7 Mass: m _6 ¼m

ð334Þ

_ tur _ 7 h7 þ W Energy: m _ 6 h6 ¼ m

ð335Þ

_ 7 s7 Entropy: m _ 6 s6 þ S_ gen;tur ¼ m

ð336Þ

_ tur þ Ex _ D;tur _ 7 ex 7 þ W Exergy: m _ 6 ex 6 ¼ m

ð337Þ

Geothermal Energy Conversion

511

Condenser: The mass, energy, entropy, and exergy balance equations are given for the condenser under the steady state and steady flow conditions. _8; m _b _a ¼m Mass: m _7¼m

ð338Þ

_ a ha ¼ m _ 8 h8 þ m _ b hb Energy: m _ 7 h7 þ m

ð339Þ

_ a sa þ S_ gen;con ¼ m _ 8 s8 þ m _ b sb Entropy: m _ 7 s7 þ m

ð340Þ

_ D;con _ a ex a ¼ m _ 8 ex 8 þ m _ b exb þ Ex Exergy: m _ 7 ex 7 þ m

ð341Þ

The ambient conditions To and Po are assumed to be 251C and 101.3 kPa, respectively. The assumptions used in the operating conditions of a single flash geothermal power generation system are given in Table 9. The heat and work input/output rate, entropy generation rate, and exergy destruction rates, and both energetic and exergetic effectiveness are evaluated from these balance equations and assumptions of variables. The exergy destruction rates of single flash steam geothermal power generation system parts are calculated by using the abovegiven balance equations. The analysis results are illustrated in Fig. 31. As seen in this figure, the exergy destruction rate is higher in the turbine than in other system components. In order to better investigate the system performance, the parametric studies are given below to analyze the effects of different indicators on the exergy destruction rates of system components and power generation rate. Geothermal fluid mass flow rate is an important factor affecting net electricity generation and exergy efficiency of the system. When geothermal fluid mass flow rate increases, the turbine of the system produces higher work. According to Fig. 32, as geothermal mass flow rate increases from 75 to 225 kg/s, net electricity generation increases sharply from 2500 to 8000 kW. Due to increments in power generation, exergy efficiency of the system increases too, from 34% to about 46%. As seen from Fig. 33, as temperature of geothermal fluid increases from 150 to 2301C, both net electricity generation and exergy efficiency of the system increases. Electricity generation increases nearly 2.5 times and similarly exergy efficiency increases about 20%. The reason for these increments is that higher temperature fluid transfers more energy to the turbine of the system. Fig. 34 shows how ambient temperature affects the electricity production rate and exergy efficiency of the system. As seen from that figure, there is a direct proportion between ambient temperature and both electricity generation and exergy efficiency. As ambient temperature increases, difference between ambient temperature and geothermal fluid temperature decreases. While ambient temperature varies from 0 to 401C, electricity generation increases from about 2600 kW to nearly 4000 kW and exergy efficiency increases from 30% to almost 41%.

Table 9

Assumptions for the single flash geothermal power system

Variables

Values

Geofluid source temperature (T1) Geofluid source pressure (P1) Geofluid mass flow rate ðm_ 1 Þ Separator inlet pressure (P2) Turbine output pressure (P7) Geofluid reinjection temperature (T8)

150–2301C 1500 kPa 75 to 225 kg/s 530 kPa 95 kPa 47.551C

Exergy destruction rate (kW) 1% 17% 30%

1%

51%

Flashing

Separator

Purifier

Turbine

Fig. 31 Exergy destruction rates for the single flash geothermal power generation system.

Condenser

Geothermal Energy Conversion

8000

0.46

7500

Wturbine (kW)

7000

ψSF

0.44

Wturbine (kW)

6500 0.42

6000 5500

0.4

5000 4500

0.38

4000 3500

0.36

System exergy efficiency

512

3000 2500 75

100

125

150

175

200

0.34 225

mgeothermal (kg/s) Fig. 32 Effect of mass flow rate of geofluid on net electricity generation and exergy efficiency.

4500

0.4 Wturbine ψSF

0.39 0.38

Wturbine (kW)

0.37 3500

0.36 0.35

3000

0.34 0.33

2500

System exergy efficiency

4000

0.32 2000 150

160

170

180

190

200

210

220

0.31 230

Tgeothermal (°C) Fig. 33 Effect of geofluid temperature on electricity generation and exergy efficiency.

4.11.6.3

Double Flash Steam Power Generation

Fig. 35 illustrates the simplified schematic diagram of double flash steam geothermal power generation system in the third case study. To reduce the pressure level of steam, the geothermal fluid enters the flashing I at point 1. After that, the geofluid enters the separator I at point 2. In the separator I, the steam is removed from separator at point 3, and enters the purifier to eject fouling materials from working fluid. The steam enters the high pressure (HP) turbine at point 5, and expands to point 6 to generate electricity. On the other hand, the stream from point 10 goes into flashing II to reach the separator II at a decreased pressure level at point 11. In the separator II, the working fluid is removed at point 13, which after that mixes with working fluid coming from the HP turbine in the mixing room. Then the mixing working fluid enters the LP turbine at point 7 to produce electricity. Also, the geofluid exiting from separator II is sent to the reinjection well at point 12. The water in the condenser should be utilized to provide heating to the water in state a, as illustrated in Fig. 35. After leaving the condenser, the geothermal fluid at point 9 is reinjected to the reinjection well. The performance of this system is higher than a single flash system. But, the construction cost is nearly 6–8% higher than the single flash system. The balance equations of double flash steam geothermal power components are written in the next subsections. Flashing I: The balance equations of flashing I under steady state and steady flow conditions can be defined as _2 Mass: m _1 ¼m

ð342Þ

_ 2 h2 Energy: m _ 1 h1 ¼ m

ð343Þ

Geothermal Energy Conversion

0.41

4200 Wturbine ψSF

0.4 0.39

3800

0.38

3600

0.37 0.36

3400

0.35

3200

0.34 0.33

3000

0.32 2800

System exergy efficiency

4000

Wturbine (kW)

513

0.31

2600 0

5

10

15

20

25

30

35

0.3 40

Tambient (°C) Fig. 34 Effect of ambient temperature on electricity generation and exergy efficiency.

5

Purifier 4

HP turbine

3 Fouling material 2

1 Flashing I

Mixing room

7 8 b Condenser

10

Flashing II

Production well

6

Separator I

Power

LP turbine

13 11

Separator II

12 Reinjection well

a 9

Reinjection well

Fig. 35 Schematic diagram of double flash steam geothermal power generation.

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;fls_I ¼ m

ð344Þ

_ D;fls_I _ 2 ex 2 þ Ex Exergy: m _ 1 ex1 ¼ m

ð345Þ

Separator I: The balance equations of separator I under steady state and steady flow conditions are given as follows: _3þm _ 10 Mass: m _2¼m

ð346Þ

_ 3 h3 þ m _ 10 h10 Energy: m _ 2 h2 ¼ m

ð347Þ

_ 3 s3 þ m _ 10 s10 Entropy: m _ 2 s2 þ S_ gen;sep_I ¼ m

ð348Þ

_ D;sep_I _ 3 ex3 þ m _ 10 ex10 þ Ex Exergy: m _ 2 ex 2 ¼ m

ð349Þ

Flashing II: The balance equations of flashing II under steady state and steady flow conditions can be defined as _ 11 Mass: m _ 10 ¼ m

ð350Þ

_ 11 h11 Energy: m _ 10 h10 ¼ m

ð351Þ

514

Geothermal Energy Conversion _ 11 s11 Entropy: m _ 10 s10 þ S_ gen;fls_II ¼ m

ð352Þ

_ D;fls_II _ 11 ex 11 þ Ex Exergy: m _ 10 ex 10 ¼ m

ð353Þ

Separator II: The balance equations of separator II under steady state and steady flow conditions are given as follows; _ 12 þ m _ 13 Mass: m _ 11 ¼ m

ð354Þ

_ 12 h12 þ m _ 13 h13 Energy: m _ 11 h11 ¼ m

ð355Þ

_ 12 s12 þ m _ 13 s13 Entropy: m _ 11 s11 þ S_ gen;sep_II ¼ m

ð356Þ

_ D;sep_II _ 12 ex 12 þ m _ 13 ex 13 þ Ex Exergy: m _ 11 ex 11 ¼ m

ð357Þ

Purifier: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the purifier can be given as _4þm _5 Mass: m _3¼m

ð358Þ

_ 4 h4 þ m _ 5 h5 Energy: m _ 3 h3 ¼ m

ð359Þ

_ 4 s4 þ m _ 5 s5 Entropy: m _ 3 s3 þ S_ gen;pur ¼ m

ð360Þ

_ D;pur _ 4 ex 4 þ m _ 5 ex 5 þ Ex Exergy: m _ 3 ex 3 ¼ m

ð361Þ

High pressure turbine: The mass, energy, entropy, and exergy balance equations are written for the high pressure (HP) turbine under the steady state and steady flow conditions. _6 Mass: m _5 ¼m

ð362Þ

_ HP_tur _ 6 h6 þ W Energy: m _ 5 h5 ¼ m

ð363Þ

_ 6 s6 Entropy: m _ 5 s5 þ S_ gen;HP_tur ¼ m

ð364Þ

_ HP_tur þ Ex _ D;HP_tur _ 6 ex 6 þ W Exergy: m _ 5 ex 5 ¼ m

ð365Þ

Low pressure turbine: The mass, energy, entropy, and exergy balance equations are written for the LP turbine under the steady state and steady flow conditions. _8 Mass: m _7 ¼m

ð366Þ

_ LP_tur _ 8 h8 þ W Energy: m _ 7 h7 ¼ m

ð367Þ

_ 8 s8 Entropy: m _ 7 s7 þ S_ gen;LP_tur ¼ m

ð368Þ

_ LP_tur þ Ex _ D;LP_tur _ 8 ex 8 þ W Exergy: m _ 7 ex7 ¼ m

ð369Þ

Mixing room: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the mixing room can be given as _ 13 ¼ m _7 Mass: m _6þm

ð370Þ

_ 13 h13 ¼ m _ 7 h7 Energy: m _ 6 h6 þ m

ð371Þ

_ 13 s13 þ S_ gen;mr ¼ m _ 7 s7 Entropy: m _ 6 s6 þ m

ð372Þ

_ D;mr _ 13 ex 13 ¼ m _ 7 ex7 þ Ex Exergy: m _ 6 ex 6 þ m

ð373Þ

Condenser: The mass, energy, entropy, and exergy balance equations are given for the condenser under the steady state and steady flow conditions. _ 9; m _b Mass: m _8 ¼m _a ¼m

ð374Þ

_ a ha ¼ m _ 9 h9 þ m _ b hb Energy: m _ 8 h8 þ m

ð375Þ

_ a sa þ S_ gen;con ¼ m _ 9 s9 þ m _ b sb Entropy: m _ 8 s8 þ m

ð376Þ

_ D;con _ a ex a ¼ m _ 9 ex 9 þ m _ b exb þ Ex Exergy: m _ 8 ex 8 þ m

ð377Þ

Geothermal Energy Conversion

515

The reference conditions’ ambient temperature and pressure are taken as 251C and 101.3 kPa, respectively. The assumptions used in the operating conditions of double flash geothermal power generation system are given in Table 10. The heat and work input/output rate, entropy generation rate, and exergy destruction rates, and energy and exergy effectiveness are evaluated from these balance equations and assumptions of variables. The exergy destruction rates of double flash geothermal power generation components are illustrated in Fig. 36. The turbine and flashing subcomponents exhibit higher exergy destruction rate than in other system components. Also, the condenser has the next largest exergy destruction rate, mainly due to the temperature difference between geothermal fluid and heating application working fluid passing through the condenser, but also due to the pressure drop across the subcomponent. The dimensionless exergy destruction rates of process components are shown in Fig. 37. This exergetic indicator is a beneficial step for prioritizing Table 10 Assumptions for the double flash geothermal power system Variables

Values

Geofluid source temperature (T1) Geofluid source pressure (P1) Geofluid mass flow rate (m_ 1 ) Separator I inlet pressure (P2) Separator II inlet pressure (P11) HP turbine output pressure (P6) LP turbine output pressure (P8) Geofluid reinjection temperature (T9)

150–2301C 1500 kPa 75 to 225 kg/s 530 kPa 95 kPa 95 kPa 10 kPa 47.551C

Exergy destruction rate (kW) 1904 2000 1800 1600 1400 1200 1000 800 600 400 200 0

1681.4 1194

1095

755.7

er

e

ns

in

de

rb C

LP

on

tu

ro M

H

ix

P

in

g

tu

Pu

rb

rif

in

om

e

74.74

ie

I to

Se

Fl

pa

as

ra

hi

ra pa Se

rI

II

to

ng

rI

I ng hi as Fl

70.3

r

81.6

45.38

Fig. 36 Exergy destruction rates for the double flash geothermal power generation system.

Dimensionless exergy destruction ratio (%) 27.59 30

24.36

25 17.30

15.86

20

10.95

15 10 1.18

0.66

5

1.08

1.02

er

e

ns

in

de

rb

Fig. 37 Dimensionless exergy destruction ratios for the double flash geothermal power generation system.

on C

g in M

ix

tu

rb tu P H

LP

in

ro om

e

r ie rif

to Se

pa

ra

in sh Fl a

Pu

rI

I

II g

rI to

Se pa ra

Fl

as

hi

ng

I

0

516

Geothermal Energy Conversion

exergy destruction in an intuitive behavior. Both exergy destruction rate and dimensionless exergy destruction rate in the double flash geothermal power system are higher in the turbine than in other system parts. Also, the separator, purifier, and mixing room do not exhibit an important exergy destruction ratio. The exergy efficiency of double flash geothermal steam power generation process components and whole system are calculated, as illustrated in Fig. 38. It is seen that the exergy efficiency of purifier, and mixing room separators I and II are higher than other process components. In order to better understand analysis of process efficiency, the parametric study results are shown below to investigate the impacts of different indicators on the exergy destruction rates and power generation rate. The impact of mass flow rate of geothermal working fluid on the power production from double flash geothermal steam power process and exergy efficiency is illustrated in Fig. 39. According to this figure, while mas flow rate of geothermal fluid varies from 75 kg/s to 225 kg/s, power generation increases nearly three times and exergy efficiency increases from 44% to 57%. The higher mass flow rate causes to produce higher work in the turbine of the system. The impact of geothermal working fluid temperature on the power production from the double flash geothermal steam power process and exergy efficiency is illustrated in Fig. 40. As expected, the higher temperature of geothermal working fluid has positive effect on the net power generation and exergy efficiency. As seen from Fig. 40, while temperature of geothermal fluid changes from 150 to 2301C, produced power varies from 1500 to 10,500 kW, and exergy efficiency increases from about 41% to 49.5%. Because of an increment in the temperature of geothermal fluid, working fluid produces more work in the turbines. The impact of ambient temperature on the power production from double flash geothermal steam power process and exergy efficiency is illustrated in Fig. 41. Ambient temperature is an important parameter for determining the exergy efficiency due to irreversibilities of system components dependent on the ambient temperature. As ambient temperature increases from 0 to 401C, total power production increases from 5200 to 8400 kW, too. Similarly, exergy efficiency increases from about 38% to 53% with increasing ambient temperature. The reason for this increment is that as ambient temperature rises, the difference between the

88.54

89.53

88.51

78.85

77.87

45.98

42.27

em st

er ns on

C

Sy

e de

rb tu

LP

ro M

ix

in

g

tu P H

in

om

e rb

rif

in

ie

I Pu

to

Se

pa

ra

hi as Fl

ra pa Se

rI

II ng

rI to

ng hi as Fl

r

32.88 32.84

I

Exergy efficiency (%)

98.24 100 90 80 70 60 50 40 30 20 10 0

System components Fig. 38 Exergy efficiencies for the double flash steam power generation components.

0.58

17,000 Wtotal ψDF

15,000

0.56

14,000

0.54

Wtotal (kW)

13,000 12,000

0.52

11,000 0.5

10,000 9000

0.48

8000 7000

0.46

6000 5000 75

100

125

150

175

200

mgeothermal (kg/s) Fig. 39 Effect of mass flow rate of geothermal fluid on net power generation and exergy efficiency.

0.44 225

System exergy efficiency

16,000

Geothermal Energy Conversion

0.5

10,500 Wtotal ψDF

0.49

8500

0.48

7500

0.47

6500

0.46

5500

0.45

4500

0.44

3500

0.43

2500

0.42 160

170

180

190

200

210

220

System exergy efficiency

Wtotal (kW)

9500

1500 150

517

0.41 230

Tgeothermal (°C) Fig. 40 Effect of geothermal fluid temperature on net power generation and exergy efficiency.

0.54

8400 Wtotal ψDF

0.52

7600

0.5

7200

0.48

6800

0.46

6400

0.44

6000

0.42

5600

0.4

5200 0

5

10

15

20

25

30

35

System exergy efficiency

Wtotal (kW)

8000

0.38 40

Tambient (°C) Fig. 41 Effect of ambient temperature on net power generation and exergy efficiency.

temperature of working fluid and the environment temperature decreases. From the definition of exergy, the decrease of temperature difference causes exergy efficiency to go up.

4.11.6.4

Triple Flash Steam Power Generation

A schematic flow diagram of a triple flash steam geothermal power generation system in the fourth case study is illustrated in Fig. 42. The geothermal power generation process runs when the geofluid enters the flashing I at point 1 [31]. The destruction in pressure level of working fluid occurs in flashing I, and geofluid enters separator I at point 2 in the form of saturated liquid. In separator I, the steam is removed from the separator at point 3, and enters the purifier to eject fouling materials from geothermal fluid. The steam enters the HP turbine at point 5, and expands to point 6 to produce power. On the other hand, the stream from point 12 goes into flashing II to reach the separator II at a decreased pressure level at point 13. Similarly, in separator II, the working fluid is removed at point 14, which after that mixes with working fluid coming from the HP turbine in the mixing room I. Then the mixing working fluid enters the medium pressure (MP) turbine at point 7 to generate power. Also, the geofluid stream from point 15 goes into flashing III to enter separator III at a decreased working fluid pressure level at point 16. Similarly, in separator III, the geofluid is removed at point 17, and after that mixes with working fluid coming from the MP turbine in the mixing room II. Then the mixing working fluid enters the LP turbine at point 9 to generate power. Also, the geofluid exiting from separator III is sent to the reinjection well at point 18. The water in the condenser can be utilized to generate heating application to the water in state a, as shown in Fig. 42. After leaving the condenser, the geothermal fluid at point 11 is reinjected to the reinjection well. The balance equations are defined for the double flash steam geothermal power system, which is shown in Fig. 42.

518

Geothermal Energy Conversion

5

Purifier

Power

4

HP turbine

3 Fouling material 1

2

Flashing I

6

MP turbine

LP turbine

7 8

9 10

Separator I

Mixing room I

Mixing room II

b

14

12

Condenser 17

Flashing II

a 11

13

Separator II 15

Production well

16

Separator III Reinjection well

Flashing III 18 Reinjection well

Fig. 42 Schematic diagram of triple flash steam geothermal power generation.

Flashing I: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for flashing I can be given as follows: _2 Mass: m _1 ¼m

ð378Þ

_ 2 h2 Energy: m _ 1 h1 ¼ m

ð379Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;fls_I ¼ m

ð380Þ

_ D;fls_I _ 2 ex 2 þ Ex Exergy: m _ 1 ex1 ¼ m

ð381Þ

Separator I: The mass, energy, entropy, and exergy balance equations are defined for separator I under steady state and steady flow conditions. _3þm _ 12 Mass: m _2¼m

ð382Þ

_ 3 h3 þ m _ 12 h12 Energy: m _ 2 h2 ¼ m

ð383Þ

_ 3 s3 þ m _ 12 s12 Entropy: m _ 2 s2 þ S_ gen;sep_I ¼ m

ð384Þ

_ D;sep_I _ 3 ex3 þ m _ 12 ex12 þ Ex Exergy: m _ 2 ex 2 ¼ m

ð385Þ

Flashing II: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for flashing II are written as follows: _ 13 Mass: m _ 12 ¼ m

ð386Þ

_ 13 h13 Energy: m _ 12 h12 ¼ m

ð387Þ

_ 13 s13 Entropy: m _ 12 s12 þ S_ gen;fls_II ¼ m

ð388Þ

_ D;fls_II _ 13 ex 13 þ Ex Exergy: m _ 12 ex 12 ¼ m

ð389Þ

Separator II: The mass, energy, entropy, and exergy balance equations are defined for separator II under steady state and steady flow conditions. _ 14 þ m _ 15 Mass: m _ 13 ¼ m

ð390Þ

_ 14 h14 þ m _ 15 h15 Energy: m _ 13 h13 ¼ m

ð391Þ

_ 14 s14 þ m _ 15 s15 Entropy: m _ 13 s13 þ S_ gen;sep_II ¼ m

ð392Þ

_ D;sep_II _ 14 ex 14 þ m _ 15 ex 15 þ Ex Exergy: m _ 13 ex 13 ¼ m

ð393Þ

Geothermal Energy Conversion

519

Flashing III: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for flashing III are written as follows: _ 16 Mass: m _ 15 ¼ m

ð394Þ

_ 16 h16 Energy: m _ 15 h15 ¼ m

ð395Þ

_ 16 s16 Entropy: m _ 15 s15 þ S_ gen;fls_III ¼ m

ð396Þ

_ D;fls_III _ 16 ex 16 þ Ex Exergy: m _ 15 ex 15 ¼ m

ð397Þ

Separator III: The mass, energy, entropy, and exergy balance equations are defined for separator III under steady state and steady flow conditions. _ 17 þ m _ 18 Mass: m _ 16 ¼ m

ð398Þ

_ 17 h17 þ m _ 18 h18 Energy: m _ 16 h16 ¼ m

ð399Þ

_ 17 s17 þ m _ 18 s18 Entropy: m _ 16 s16 þ S_ gen;sep_III ¼ m

ð400Þ

_ D;sep_III _ 17 ex17 þ m _ 18 ex 18 þ Ex Exergy: m _ 16 ex 16 ¼ m

ð401Þ

Purifier: For the purifier of the triple flash steam geothermal power system, the balance equations are provided under the steady state and steady flow conditions. _4þm _5 Mass: m _3¼m

ð402Þ

_ 4 h4 þ m _ 5 h5 Energy: m _ 3 h3 ¼ m

ð403Þ

_ 4 s4 þ m _ 5 s5 Entropy: m _ 3 s3 þ S_ gen;pur ¼ m

ð404Þ

_ D;pur _ 4 ex 4 þ m _ 5 ex 5 þ Ex Exergy: m _ 3 ex 3 ¼ m

ð405Þ

High pressure turbine: The mass, energy, entropy, and exergy balance equations are written for the high pressure (HP) turbine under the steady state and steady flow conditions. _6 Mass: m _5 ¼m

ð406Þ

_ HP_tur _ 6 h6 þ W Energy: m _ 5 h5 ¼ m

ð407Þ

_ 6 s6 Entropy: m _ 5 s5 þ S_ gen;HP_tur ¼ m

ð408Þ

_ HP_tur þ Ex _ D;HP_tur _ 6 ex 6 þ W Exergy: m _ 5 ex 5 ¼ m

ð409Þ

Mixing room I: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for mixing room I are defined as follows: _ 14 ¼ m _7 Mass: m _6þm

ð410Þ

_ 14 h14 ¼ m _ 7 h7 Energy: m _ 6 h6 þ m

ð411Þ

_ 14 s14 þ S_ gen;mr_I ¼ m _ 7 s7 Entropy: m _ 6 s6 þ m

ð412Þ

_ D;mr_I _ 14 ex 14 ¼ m _ 7 ex7 þ Ex Exergy: m _ 6 ex 6 þ m

ð413Þ

Middle pressure turbine: The mass, energy, entropy, and exergy balance equations are written for the middle pressure (MP) turbine under the steady state and steady flow conditions. _8 Mass: m _7 ¼m

ð414Þ

_ MP_tur _ 8 h8 þ W Energy: m _ 7 h7 ¼ m

ð415Þ

_ 8 s8 Entropy: m _ 7 s7 þ S_ gen;MP_tur ¼ m

ð416Þ

_ MP_tur þ Ex _ D;MP_tur _ 8 ex 8 þ W Exergy: m _ 7 ex7 ¼ m

ð417Þ

Mixing room II: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for mixing room II are defined as follows: _ 17 ¼ m _9 Mass: m _8þm

ð418Þ

520

Geothermal Energy Conversion _ 17 h17 ¼ m _ 9 h9 Energy: m _ 8 h8 þ m

ð419Þ

_ 17 s17 þ S_ gen;mr_II ¼ m _ 9 s9 Entropy: m _ 8 s8 þ m

ð420Þ

_ D;mr_II _ 17 ex 17 ¼ m _ 9 ex9 þ Ex Exergy: m _ 8 ex 8 þ m

ð421Þ

Low pressure turbine: The mass, energy, entropy, and exergy balance equations are written for the LP turbine under the steady state and steady flow conditions. _ 10 Mass: m _9¼m

ð422Þ

_ LP_tur _ 10 h10 þ W Energy: m _ 9 h9 ¼ m

ð423Þ

_ 10 s10 Entropy: m _ 9 s9 þ S_ gen;LP_tur ¼ m

ð424Þ

_ LP_tur þ Ex _ D;LP_tur _ 10 ex 10 þ W Exergy: m _ 9 ex 9 ¼ m

ð425Þ

Condenser: The mass, energy, entropy, and exergy balance equations are written for the condenser under the steady state and steady flow conditions. _ 11 ; m _b _a ¼m Mass: m _ 10 ¼ m

ð426Þ

_ a ha ¼ m _ 11 h11 þ m _ b hb Energy: m _ 10 h10 þ m

ð427Þ

_ a sa þ S_ gen;con ¼ m _ 11 s11 þ m _ b sb Entropy: m _ 10 s10 þ m

ð428Þ

_ D;con _ a ex a ¼ m _ 11 ex11 þ m _ b ex b þ Ex Exergy: m _ 10 ex10 þ m

ð429Þ

The reference conditions To and Po are assumed to be 251C and 101.3 kPa, respectively. The assumptions used in the operating conditions of triple flash geothermal power generation system are written in Table 11. The heat and work input/output rate, entropy generation rate, and exergy destruction rates, and energy and exergy efficiencies are calculated from the mass, energy, entropy, and exergy balance equations and assumptions of variables. The exergy destruction rate, dimensionless exergy destruction ratio, and exergy efficiency of triple flash steam geothermal power generation components are given in Table 12. The exergy analysis results showed that the HP, MP, and LP turbines, and also flashing subcomponents, are the main sources of irreversibility. The purifier subcomponent has the maximum exergy efficiency rate. Therewithal, the exergy efficiencies of separators and mixing rooms in the geothermal process vary between 81.95% and 84.76%, and 87.25% and 88.24%, respectively. These exergy efficiencies can be observed to be higher than other system components. To investigate the performance of triple flash steam geothermal power system more comprehensively, the parametric studies are given below to examine the impacts of different indicator variables on the exergy destruction rate and exergy efficiency. According to the findings, the predominant parameter affecting net electricity production is mass flow rate. As seen from Fig. 43, as mass flow rate triples from 75 to 225 kg/s, electricity production and exergy efficiency increase from 7000 kW to about 20,500 kW and 52 to 64%, respectively. The increment in geothermal mass flow rate makes turbine producing more work. Because higher temperature fluid transfers more energy to the turbine of the system, increment in geothermal fluid temperature has positive effect on both electricity generation and exergy efficiency. As seen from Fig. 44, the electricity production increases from 7300 to 10,500 kW and exergy efficiency increases from 49% to 57%, respectively. According to the findings of this study, ambient temperature is the most important factor affecting exergy efficiency. Definition of exergy clarifies this increment. As seen from Fig. 45, as ambient temperature changes from 0 to 401C, exergy efficiency increases from about 43% to 63%. Proportionally, produced electricity increases about 2000 kW with the same temperature change. Table 11

Assumptions for the triple flash geothermal power system

Variables

Values

Geofluid source temperature (T1) Geofluid source pressure (P1) _ 1) Geofluid mass flow rate (m Separator I inlet pressure (P2) Separator II inlet pressure (P13) Separator III inlet pressure (P16) HP turbine output pressure (P6) MP turbine output pressure (P8) LP turbine output pressure (P10) Geofluid reinjection temperature (T11)

150–2301C 1500 kPa 75 to 225 kg/s 530 kPa 95 kPa 50 kPa 95 kPa 50 kPa 10 kPa 42.51C

Geothermal Energy Conversion

Table 12

521

Thermodynamic assessment results for the triple flash steam geothermal power generation process components Exergy destruction rate (kW)

Exergy destruction ratio (%)

Exergy efficiency (%)

Flashing I Separator I Flashing II Separator II Flashing III Separator III Purifier HP Turbine Mixing Room I MP Turbine Mixing Room II LP Turbine Condenser

1095 45.38 1194 81.6 438.3 94.28 70.3 1681.4 76.54 1520 74.74 1373.2 879.9

12.7 0.53 13.8 0.95 5.08 1.09 0.82 19.5 0.89 17.6 0.87 15.9 10.2

74.52 84.76 73.28 83.26 72.52 81.95 97.25 44.62 88.24 41.86 87.25 40.28 31.28

22,000 20,500

Wtotal

19,000

ψTF

Wtotal (kW)

17,500 16,000 14,500 13,000 11,500 10,000 8500 7000 75

100

125

150

175

200

0.64 0.63 0.62 0.61 0.6 0.59 0.58 0.57 0.56 0.55 0.54 0.53 0.52 0.51 225

System exergy efficiency

System components

mgeothermal (kg/s) Fig. 43 Effect of mass flow rate of geothermal working fluid on net power generation and exergy efficiency.

11,000

0.57 Wtotal ψTF

0.56 0.55

Wtotal (kW)

10,000

0.54

9500

0.53 9000 0.52 8500

0.51

8000

0.5

7500 7000 150

0.49 160

170

180

190

200

210

220

Tgeothermal (°C) Fig. 44 Effect of geothermal working fluid temperature on net power generation and exergy efficiency.

0.48 230

System exergy efficiency

10,500

Geothermal Energy Conversion

0.63

10,250 Wtotal ψTF

10,000

0.61 0.59

9750 Wtotal (kW)

0.57 9500

0.55

9250

0.53 0.51

9000

0.49 8750 0.47 8500

System exergy efficiency

522

0.45 0.43 40

8250 0

5

10

15

20

25

30

35

Tambient (°C) Fig. 45 Effect of ambient temperature on net power generation and exergy efficiency.

9

Evaporator I

1

Turbine I Power 13

2

Preheater I 12

5 Production well

10 a

Pump I 11 Condenser I

b 4

14

Evaporator II

Turbine II Power

3 18 3-way valve I

15 6

Preheater II 17

Pump II

c

16 Condenser II

d 3-way valve II

7 Reinjection well

8

Fig. 46 Schematic diagram of binary cycle geothermal power generation.

4.11.6.5

Binary Cycle Power Generation

The binary cycle geothermal power generation process in the fifth case study is shown in Fig. 46. The binary cycle process runs when the geofluid enters evaporator I at point 1. The thermal energy of geothermal working fluid is transferred by using the evaporator to another working fluid, such as isobutane, isopentane, R-113, R-123, ethanol, etc., for use in a fairly conventional ORC process [32]. The ORC working fluid enters ORC turbine I at point 9 to produce electricity, and leaves at point 10. The expansion step causes temperature and pressure level to decrease. The ORC working fluid exiting from the ORC turbine passes through condenser II to supply heat energy for heating application. After exiting from condenser I, the ORC working fluid passes

Geothermal Energy Conversion

523

through ORC pump I at point 11 to increase the pressure level and exits at point 12, which goes into preheater I to transfer useful heat by using the geofluid coming from 3-way valve I at point 4. The similar binary cycle occurs in process 2. The geothermal working fluid coming from preheater I at point 5 and preheater II at point 7 mixes in 3-way valve II. Finally, geofluid is reinjected to the reinjection well at point 8. The mass, energy, entropy, and exergy balance equations are defined for a double flash steam geothermal power system, which is shown in Fig. 46. Evaporator I: The mass, energy, entropy, and exergy balance equations are written for evaporator I under the steady state and steady flow conditions. _ 2; m _ 13 Mass: m _1¼m _9 ¼m

ð430Þ

_ 13 h13 ¼ m _ 2 h2 þ m _ 9 h9 Energy: m _ 1 h1 þ m

ð431Þ

_ 13 s13 þ S_ gen;eva_I ¼ m _ 2 s2 þ m _ 9 s9 Entropy: m _ 1 s1 þ m

ð432Þ

_ D;eva_I _ 13 ex 13 ¼ m _ 2 ex2 þ m _ 9 ex 9 þ Ex Exergy: m _ 1 ex 1 þ m

ð433Þ

Preheater I: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for preheater I are defined as follows: _ 5; m _ 13 Mass: m _2¼m _ 12 ¼ m

ð434Þ

_ 12 h12 ¼ m _ 5 h5 þ m _ 13 h13 Energy: m _ 2 h2 þ m

ð435Þ

_ 12 s12 þ S_ gen;pht_I ¼ m _ 5 s5 þ m _ 13 s13 Entropy: m _ 2 s2 þ m

ð436Þ

_ D;pht_I _ 12 ex12 ¼ m _ 5 ex 5 þ m _ 13 ex 13 þ Ex Exergy: m _ 2 ex2 þ m

ð437Þ

Turbine I: The mass, energy, entropy, and exergy balance equations are written for turbine I under the steady state and steady flow conditions. _ 10 Mass: m _9¼m

ð438Þ

_ tur_I _ 10 h10 þ W Energy: m _ 9 h9 ¼ m

ð439Þ

_ 10 s10 Entropy: m _ 9 s9 þ S_ gen;tur_I ¼ m

ð440Þ

_ tur_I þ Ex _ D;tur_I _ 10 ex 10 þ W Exergy: m _ 9 ex9 ¼ m

ð441Þ

Condenser I: The mass, energy, entropy, and exergy balance equations are written for condenser I under the steady state and steady flow conditions. _ 11 ; m _b Mass: m _ 10 ¼ m _a¼m

ð442Þ

_ a ha ¼ m _ 11 h11 þ m _ b hb Energy: m _ 10 h10 þ m

ð443Þ

_ a sa þ S_ gen;con_I ¼ m _ 11 h11 þ m _ b hb Entropy: m _ 10 s10 þ m

ð444Þ

_ D;con_II _ a ex a ¼ m _ 11 ex11 þ m _ b ex b þ Ex Exergy: m _ 10 ex10 þ m

ð445Þ

Pump I: For pump I of the binary cycle geothermal power system, the balance equations are provided under the steady state and steady flow conditions. _ 12 Mass: m _ 11 ¼ m

ð446Þ

_ p_I ¼ m _ 12 h12 Energy: m _ 11 h11 þ W

ð447Þ

_ 12 s12 Entropy: m _ 11 s11 þ S_ gen;p_I ¼ m

ð448Þ

_ p_I ¼ m _ D;p_I _ 12 ex 12 þ Ex Exergy: m _ 11 ex 11 þ W

ð449Þ

Evaporator II: The mass, energy, entropy, and exergy balance equations can be defined for evaporator II under the steady state and steady flow conditions. _ 3; m _ 18 _ 14 ¼ m Mass: m _2¼m

ð450Þ

524

Geothermal Energy Conversion _ 18 h18 ¼ m _ 3 h3 þ m _ 14 h14 Energy: m _ 2 h2 þ m

ð451Þ

_ 18 s18 þ S_ gen;eva_II ¼ m _ 3 s3 þ m _ 14 s14 Entropy: m _ 2 s2 þ m

ð452Þ

_ D;eva_II _ 18 ex18 ¼ m _ 3 ex 3 þ m _ 14 ex 14 þ Ex Exergy: m _ 2 ex2 þ m

ð453Þ

3-Way valve I: The mass, energy, entropy, and exergy balance equations are defined for 3-way valve I under steady state and steady flow conditions. _4þm _6 Mass: m _3¼m

ð454Þ

_ 4 h4 þ m _ 6 h6 Energy: m _ 3 h3 ¼ m

ð455Þ

_ 4 s4 þ m _ 6 s6 Entropy: m _ 3 s3 þ S_ gen;3wv_I ¼ m

ð456Þ

_ D;3wv_I _ 4 ex4 þ m _ 6 ex 6 þ Ex Exergy: m _ 3 ex 3 ¼ m

ð457Þ

Preheater II: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for preheater II is defined as follows: _7; m _ 18 Mass: m _6¼m _ 17 ¼ m

ð458Þ

_ 17 h17 ¼ m _ 7 h7 þ m _ 18 h18 Energy: m _ 6 h6 þ m

ð459Þ

_ 17 s17 þ S_ gen;pht_II ¼ m _ 7 s7 þ m _ 18 s18 Entropy: m _ 6 s6 þ m

ð460Þ

_ D;pht_II _ 17 ex 17 ¼ m _ 7 ex 7 þ m _ 18 ex 18 þ Ex Exergy: m _ 6 ex6 þ m

ð461Þ

Turbine II: The mass, energy, entropy, and exergy balance equations are written for turbine II under the steady state and steady flow conditions. _ 15 Mass: m _ 14 ¼ m

ð462Þ

_ tur_II _ 15 h15 þ W Energy: m _ 14 h14 ¼ m

ð463Þ

_ 15 s15 Entropy: m _ 14 s14 þ S_ gen;tur_II ¼ m

ð464Þ

_ tur_II þ Ex _ D;tur_II _ 15 ex15 þ W Exergy: m _ 14 ex14 ¼ m

ð465Þ

Condenser II: The mass, energy, entropy, and exergy balance equations are written for condenser II under the steady state and steady flow conditions. _ 16 ; m _d _c ¼m Mass: m _ 15 ¼ m

ð466Þ

_ c hc ¼ m _ 16 h16 þ m _ d hd Energy: m _ 15 h15 þ m

ð467Þ

_ c sc þ S_ gen;con_II ¼ m _ 16 h16 þ m _ d hd Entropy: m _ 15 s15 þ m

ð468Þ

_ D;con_II _ c ex c ¼ m _ 16 ex16 þ m _ d exd þ Ex Exergy: m _ 15 ex 15 þ m

ð469Þ

Pump II: For pump II of the binary cycle geothermal power system, the balance equations are provided under the steady state and steady flow conditions. _ 17 Mass: m _ 16 ¼ m

ð470Þ

_ p_II ¼ m _ 17 h17 Energy: m _ 16 h16 þ W

ð471Þ

_ 17 s17 Entropy: m _ 16 s16 þ S_ gen;p_II ¼ m

ð472Þ

_ p_II ¼ m _ D;p_II _ 17 ex 17 þ Ex Exergy: m _ 16 ex16 þ W

ð473Þ

3-Way valve II: The mass, energy, entropy, and exergy balance equations are defined for 3-way valve II under steady state and steady flow conditions. _7¼m _8 Mass: m _5þm

ð474Þ

Geothermal Energy Conversion

525

_ 7 h7 ¼ m _ 8 h8 Energy: m _ 5 h5 þ m

ð475Þ

_ 7 s7 þ S_ gen;3wv_II ¼ m _ 8 s8 Entropy: m _ 5 s5 þ m

ð476Þ

_ D;3wv_II _ 7 ex7 ¼ m _ 8 ex 8 þ Ex Exergy: m _ 5 ex 5 þ m

ð477Þ

The reference conditions’ ambient temperature and pressure are taken as 251C and 101.3 kPa, respectively. The similar balance equations can be expressed for pump II using by the above procedure. The assumptions used in the operating conditions of the binary cycle geothermal power generation system are shown in Table 13. The exergy destruction rate, dimensionless exergy destruction ratio, and exergy efficiency of the binary cycle geothermal power generation components are illustrated in Table 14. The exergy analysis results show that condensers I and II, and also turbines I and II are the main sources of irreversibility. The 3-way valves I and II have the maximum exergy efficiency rate. Therewithal, the exergy efficiencies of evaporators and preheaters in the geothermal process vary between 86.24 and 88.5%, and 72.5 and 74.8%, respectively. In order to analyze the performance of binary cycle geothermal power generation process more effectively, the parametric studies are investigated below to examine the effects of some different indicator variables on the exergy destruction rate and exergy efficiency. Fig. 47 shows the relation between geothermal fluid mass flow rate and the net energy production and exergy efficiency of the system. According to the calculations and Fig. 47, increase in mass flow rate increases net energy generation. As mass flow rate triples from 75 to 225 kg/s, the amount of generated power increases from 3750 to 9000 kW. This result is logical because any increase in mass flow rate increases turbine work in the system. However, this increase in mass flow rate causes a decrease in exergy efficiency from 51 to 39%. The reason for this decrease is that the binary cycle is increasing generated power because of the second cycle, however, a limited part of the energy can be transferred from the first cycle to the second one. Losses occurring during this transfer decrease the exergy efficiency of whole system. As seen from Fig. 48, the temperature of the geothermal fluid has a positive effect on generated power but a negative effect on the exergy efficiency. As geothermal water temperature varies from 150 to 2301C, produced electricity increases from 3200 to 5400 kW. This is because fluid having higher temperature produces more work in the turbine. Although increase in geothermal fluid temperature increases generated power, it makes exergy efficiency decrease from 60 to 52%. Fig. 49 shows a direct proportion between ambient temperature and both generated power and exergy efficiency. As ambient temperature increases, losses occurred by irreversibilities decrease. While generated power is about 3700 kW at 01C, it goes up to 5200 kW at 401C. With this change of ambient temperature, exergy efficiency of the system increases from 44 to 52%. Table 13 Assumptions for the binary cycle geothermal power system

Table 14

Variables

Values

Geofluid source temperature (T1) Geofluid source pressure (P1) Geofluid mass flow rate (m_ 1 ) Evaporator II inlet temperature (T2) Turbine I output pressure (P10) Turbine II output pressure (P15) Geofluid reinjection temperature (T8)

150–2301C 1500 kPa 75 to 225 kg/s 130.71C 130 kPa 114 kPa 66.151C

Thermodynamic analysis outputs for the binary cycle geothermal power generation system components

System components

Exergy destruction rate (kW)

Exergy destruction ratio (%)

Exergy efficiency (%)

Evaporator I Preheater I Evaporator II Preheater II 3-Way valve I Turbine I Condenser I Pump I Turbine II Condenser II Pump II 3-Way valve II

260.5 180.9 248.6 220.5 84.2 458.2 795 101.3 447.9 641 94.4 76.2

7.22 5.01 6.89 6.11 2.33 12.7 22 2.81 12.4 17.8 2.62 2.11

88.5 74.8 86.24 72.5 90.24 44.8 30.5 76.4 42.6 26.8 80.7 89.1

Geothermal Energy Conversion

9000

0.52 0.51

Wnet ψBC

8500 8000

0.5 0.49

Wnet (kW)

7500

0.48

7000

0.47

6500

0.46

6000

0.45

5500

0.44 0.43

5000

0.42

4500

System exergy efficiency

526

0.41

4000

0.4

3500 75

100

125

150

175

200

0.39 225

mgeothermal (kg/s) Fig. 47 Effect of mass flow rate of geothermal working fluid on net power generation and exergy efficiency.

0.61

5600 Wnet

5200

ψBC

0.6 0.59

5000

0.58

Wnet (kW)

4800 4600

0.57

4400 0.56

4200

0.55

4000 3800

System exergy efficiency

5400

0.54

3600 0.53

3400 3200 150

160

170

180

190

200

210

220

0.52 230

Tgeothermal (°C) Fig. 48 Effect of geothermal working fluid temperature on power generation and exergy efficiency.

4.11.6.6

Combined Power Generation

The combined geothermal power generation system proposed in this case study includes the following design parameters given in Fig. 50. To decrease the pressure level [33], the geothermal working fluid enters the flashing subcomponent at point 1. The exiting geofluid from flashing enters the separator at point 2. In the separator unit, the liquid and steam fluid are separated into steam and liquid phases based on their varied specific volumes. The geothermal fluid enters the purifier to eject fouling materials at point 4. The saturated working fluid at point 5 goes into turbine I where it is expanded to generate electricity. The expanded working fluid leaves from the turbine at point 6, and passes through condenser I to supply heating application. The binary cycle process runs when the geothermal working fluid goes into the evaporator at point 9. In this process, the thermal energy of geothermal working fluid is transferred by using the evaporator to the different working fluid for use in the conventional ORC process. The ORC working fluid goes into turbine II at point 11 to generate power, and leaves at point 12. The expansion step creates a decrease in the temperature and pressure. The working fluid that is exiting from turbine II enters condenser II to supply energy for heating application. After exiting from condenser II, the working fluid enters the pump at point 13 to increase the pressure level and exit at point 14, which goes into the evaporator. The geothermal working fluid coming from condenser I at point 7 and evaporator at point 10 mixes in the 3-way valve. Finally, the geothermal fluid is reinjected to the reinjection well at point 8. The mass, energy, entropy, and exergy balance equations are defined for combined geothermal power generation, which is illustrated in Fig. 50.

Geothermal Energy Conversion

0.53

5200 Wnet ψBC

5000

0.52 System exergy efficiency

0.51

4800 Wnet (kW)

527

0.5

4600

0.49 4400 0.48 4200

0.47

4000

0.46

3800

0.45 0.44

3600 0

5

10

15

20

25

30

35

40

Tambient (°C) Fig. 49 Effect of ambient temperature on power generation and exergy efficiency.

Purifier 5 Power

4 Turbine I 3

Fouling material

c Condenser II

2

1

12 Power

Flashing Separator

6

a

d

Turbine II

Condenser I 13

9

7 11

b

14 Evaporator Pump

Production well

3-way valve

10 Reinjection well

8

Fig. 50 Schematic diagram of combined geothermal power generation.

Flashing: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for flashing can be defined as _2 Mass: m _1 ¼m

ð478Þ

_ 2 h2 Energy: m _ 1 h1 ¼ m

ð479Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;fls ¼ m

ð480Þ

_ D;fls _ 2 ex 2 þ Ex Exergy: m _ 1 ex1 ¼ m

ð481Þ

Separator: The mass, energy, entropy, and exergy balance equations are defined for the separator under steady state and steady flow conditions. _3þm _9 Mass: m _2¼m

ð482Þ

_ 3 h3 þ m _ 9 h9 Energy: m _ 2 h2 ¼ m

ð483Þ

_ 3 s3 þ m _ 9 s9 Entropy: m _ 2 s2 þ S_ gen;sep ¼ m

ð484Þ

_ D;sep _ 3 ex 3 þ m _ 9 ex9 þ Ex Exergy: m _ 2 ex2 ¼ m

ð485Þ

528

Geothermal Energy Conversion

Purifier: For the purifier of the combined geothermal power generation system, the balance equations are provided under the steady state and steady flow conditions. _4þm _5 Mass: m _3¼m

ð486Þ

_ 4 h4 þ m _ 5 h5 Energy: m _ 3 h3 ¼ m

ð487Þ

_ 4 s4 þ m _ 5 s5 Entropy: m _ 3 s3 þ S_ gen;pur ¼ m

ð488Þ

_ D;pur _ 4 ex 4 þ m _ 5 ex 5 þ Ex Exergy: m _ 3 ex 3 ¼ m

ð489Þ

Turbine I: The mass, energy, entropy, and exergy balance equations are written for turbine I under the steady state and steady flow conditions. _6 Mass: m _5 ¼m

ð490Þ

_ tur_I _ 6 h6 þ W Energy: m _ 5 h5 ¼ m

ð491Þ

_ 6 s6 Entropy: m _ 5 s5 þ S_ gen;tur_I ¼ m

ð492Þ

_ tur_I þ Ex _ D;tur_I _ 6 ex 6 þ W Exergy: m _ 5 ex 5 ¼ m

ð493Þ

Condenser I: The mass, energy, entropy, and exergy balance equations are written for condenser I under the steady state and steady flow conditions. _ 7; m _b Mass: m _6 ¼m _a ¼m

ð494Þ

_ a ha ¼ m _ 7 h7 þ m _ b hb Energy: m _ 6 h6 þ m

ð495Þ

_ a sa þ S_ gen;con_I ¼ m _ 7 s7 þ m _ b sb Entropy: m _ 6 s6 þ m

ð496Þ

_ D;con_I _ a ex a ¼ m _ 7 ex 7 þ m _ b exb þ Ex Exergy: m _ 6 ex 6 þ m

ð497Þ

3-Way valve: The mass, energy, entropy, and exergy balance equations are defined for the 3-way valve under steady state and steady flow conditions. _ 10 ¼ m _8 Mass: m _7þm

ð498Þ

_ 10 h10 ¼ m _ 8 h8 Energy: m _ 7 h7 þ m

ð499Þ

_ 10 s10 þ S_ gen;v ¼ m _ 8 s8 Entropy: m _ 7 s7 þ m

ð500Þ

_ D;v _ 10 ex10 ¼ m _ 8 E8 þ Ex Exergy: m _ 7 ex7 þ m

ð501Þ

Evaporator: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the evaporator can be defined as _ 10 ; m _ 11 Mass: m _9¼m _ 14 ¼ m

ð502Þ

_ 14 h14 ¼ m _ 10 h10 þ m _ 11 h11 Energy: m _ 9 h9 þ m

ð503Þ

_ 14 s14 þ S_ gen;eva ¼ m _ 10 s10 þ m _ 11 s11 Entropy: m _ 9 s9 þ m

ð504Þ

_ D;eva _ 14 ex14 ¼ m _ 10 ex 10 þ m _ 11 ex11 þ Ex Exergy: m _ 9 ex9 þ m

ð505Þ

Turbine II: The mass, energy, entropy, and exergy balance equations are written for turbine II under the steady state and steady flow conditions. _ 12 Mass: m _ 11 ¼ m

ð506Þ

_ tur_II _ 12 h12 þ W Energy: m _ 11 h11 ¼ m

ð507Þ

_ 12 s12 Entropy: m _ 11 s11 þ S_ gen;tur_II ¼ m

ð508Þ

_ tur_II þ Ex _ D;tur_II _ 12 ex12 þ W Exergy: m _ 11 ex11 ¼ m

ð509Þ

Condenser II: The mass, energy, entropy, and exergy balance equations are written for condenser II under the steady state and steady flow conditions. _ 13 ; m _d Mass: m _ 12 ¼ m _c ¼m

ð510Þ

Geothermal Energy Conversion

529

_ c hc ¼ m _ 13 h13 þ m _ d hd Energy: m _ 12 h12 þ m

ð511Þ

_ c sc þ S_ gen;con_II ¼ m _ 13 s13 þ m _ d sd Entropy: m _ 12 s12 þ m

ð512Þ

_ D;con_II _ c ex c ¼ m _ 13 ex13 þ m _ d exd þ Ex Exergy: m _ 12 ex 12 þ m

ð513Þ

Pump: For the pump of the combined geothermal power system, the balance equations are provided under the steady state and steady flow conditions. _ 14 Mass: m _ 13 ¼ m

ð514Þ

_ p¼m _ 14 h14 Energy: m _ 13 h13 þ W

ð515Þ

_ 14 s14 Entropy: m _ 13 s13 þ S_ gen;p ¼ m

ð516Þ

_ p ¼m _ D;p _ 14 ex 14 þ Ex Exergy: m _ 13 ex 13 þ W

ð517Þ

The reference conditions To and Po are assumed to be 251C and 101.3 kPa, respectively. The assumptions used in the operating conditions of the combined geothermal power generation system are given in Table 15. The values for exergy destruction rates (kW), exergy destruction ratio (%), and exergy efficiency (%) of combined geothermal power generation system, corresponding to the detailed thermodynamic analysis, are illustrated in Table 16. The exergy destruction rate represents the decrease in useful energy availability; however, this thermodynamic term cannot be used to analyze the energy and exergy utilization efficiency of the process components. The exergy efficiency rates of the process components are more useful for investigating exergy losses. It is seen in Table 16 that the exergy destruction rate and exergy destruction ratio of the evaporator are higher than the other system components. According to the findings, mass flow rate of geothermal fluid has positive effect on generated power from geothermal energy systems. As seen from Fig. 51, as geothermal mass flow rate alters from 75 to 225 kg/s, the amount of produced power increases from 11,500 to 21,000 kW; in other words it almost doubles. However, losses occurring in the system decrease the amount of transferred energy. This means that produced power increases with increasing mass flow rate whereas exergy efficiency of the system decreases. These results show the importance of exergy efficiency of second law efficiency because energy efficiency analyses do not point out these losses. The effect of geothermal fluid temperature on produced electricity and exergy efficiency is shown in Fig. 52. While geothermal fluid temperature increases from 150 to 2301C, electricity generation increases from 12,350 to 13,500 kW whereas with the same temperature change exergy efficiency of the system decreases from about 78% to about 40%.

Table 15 Assumptions for the combined geothermal power system

Table 16

Variables

Values

Geofluid source temperature (T1) Geofluid source pressure (P1) Geofluid mass flow rate (m_ 1 ) Separator inlet pressure (P2) Turbine I output pressure (P6) Turbine II output pressure (P12) Geofluid reinjection temperature (T8)

150–2301C 1500 kPa 75 to 225 kg/s 600 kPa 10 kPa 400 kPa 45.81C

Thermodynamic assessment results for the combined geothermal power generation system components

System components

Exergy destruction rate (kW)

Exergy destruction ratio (%)

Exergy efficiency (%)

Flashing Separator Purifier Turbine I Condenser I 3-Way valve Evaporator Turbine II Condenser II Pump

636 54 62 775 875.2 85 1586 611.83 896.24 73.68

11.2 0.95 1.1 13.7 15.5 1.5 28 10.8 15.8 1.3

95.42 98.82 96.27 42.85 34.82 88.75 74.28 46.27 32.85 84.28

Geothermal Energy Conversion

21,000

0.58

20,000 19,000

Wnet

0.56

ψC

0.54 0.52

Wnet (kW)

18,000

0.5

17,000

0.48

16,000

0.46

15,000

0.44 0.42

14,000

0.4 13,000

0.38

12,000 11,000 75

System exergy efficiency

530

0.36 100

125

150

175

200

0.34 225

mgeothermal (kg/s) Fig. 51 Effect of mass flow rate of geothermal fluid on net electricity generation and exergy efficiency.

0.8

13,600 Wnet ψC

13,400

0.76 0.72

Wnet (kW)

13,300 13,200

0.68

13,100

0.64

13,000

0.6

12,900

0.56

12,800

0.52

12,700 12,600

0.48

12,500

0.44

12,400 12,300 150

System exergy efficiency

13,500

0.4 160

170

180

190

200

210

220

230

Tgeothermal (°C) Fig. 52 Effect of geothermal fluid temperature on net electricity generation and exergy efficiency.

Increase in ambient temperature increases both generated electricity and exergy efficiently significantly. As seen in Fig. 53, as ambient temperature rises from 0 to 401C, the amount of generated electricity increases from 11,500 kW to about 14,000 kW and the exergy efficiency increases from 37 to 58%.

4.11.6.7

Geothermal Energy Based Cooling System

In order to use geothermal heat, the double effect lithium bromide–water absorption system is chosen instead of a conventional refrigeration system. The schematic diagram of a geothermal energy based double effect absorption cooling system is illustrated in Fig. 54. As seen from this figure, the cooling system consists of two generators, a condenser, an evaporator, an absorber, a pump, four expansion valves, and two HEXes. In this paper, required energy for the double effect absorption system is supplied from geothermal resources. For this reason, at point 1, the geothermal working fluid goes through generator I for cooling generation. It should be noted that the geothermal energy based cooling system is modeled according to the optimum operating parameters for the double effect absorption subsystem. To analyze the inlet and outlet conditions of the double effect absorption cooling system components, the mass, energy, entropy, and exergy balance equations are written in the next subsections. Generator I: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for generator I are defined as follows: _ 2; m _ 10 þ m _ 13 Mass: m _1 ¼m _9¼m

ð518Þ

Geothermal Energy Conversion

14,500

0.56

ψC

0.54 0.52

13,500

0.5 0.48

13,000

0.46 0.44

12,500

0.42

System exergy efficiency

0.58 Wnet

14,000

Wnet (kW)

531

0.4

12,000

0.38 0.36 40

11,500 0

5

10

15

20

25

30

35

Tambient (°C) Fig. 53 Effect of ambient temperature on net electricity generation and exergy efficiency.

16

1

15

Geothermal brine

Condenser

Generator I

13 10

9

EV II

HEX I 2 Production well

EV I

11

14

Q

Generator II

8

Injection well

3

12

17 EV III HEX II 4 7

18

Pump 6

Q (cooling)

EV IV 19 5

Q

Absorber

Evaporator

Fig. 54 Schematic diagram of geothermal energy based cooling system.

_ gen_I ¼ m _ 9 h9 þ Q _ 2 h2 þ m _ 10 h10 þ m _ 13 h13 Energy: m _ 1 h1 þ m

ð519Þ

_ gen_I =Tgen_I þ S_ gen;gen_I ¼ m _ 9 s9 þ Q _ 2 s2 þ m _ 10 s10 þ m _ 13 s13 Entropy: m _ 1 s1 þ m

ð520Þ

_ Q ¼m _ D;gen_I _ 9 ex 9 þ Ex _ 2 ex 2 þ m _ 10 ex 10 þ m _ 13 ex13 þ Ex Exergy: m _ 1 ex 1 þ m gen_I

ð521Þ

Heat exchanger I: The mass, energy, entropy, and exergy balance equations are written for HEX I under the steady state and steady flow conditions. _9; m _ 11 Mass: m _8¼m _ 10 ¼ m

ð522Þ

_ 10 h10 ¼ m _ 9 h9 þ m _ 11 h11 Energy: m _ 8 h8 þ m

ð523Þ

_ 10 s10 þ S_ gen;HEX_I ¼ m _ 9 s9 þ m _ 11 s11 Entropy: m _ 8 s8 þ m

ð524Þ

_ D;HEX_I _ 10 ex 10 ¼ m _ 9 ex9 þ m _ 11 ex11 þ Ex Exergy: m _ 8 ex 8 þ m

ð525Þ

Expansion valve I: The mass, energy, entropy, and exergy balance equations for expansion valve I can be written under the steady state and steady flow conditions as follows: _ 12 Mass: m _ 11 ¼ m

ð526Þ

532

Geothermal Energy Conversion _ 12 h12 Energy: m _ 11 h11 ¼ m

ð527Þ

_ 12 s12 Entropy: m _ 11 s11 þ S_ gen;ev_I ¼ m

ð528Þ

_ D;ev_I _ 12 ex12 þ Ex Exergy: m _ 11 ex11 ¼ m

ð529Þ

Generator II: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for generator II can be defined as follows: _ 14 ; m _ 16 þ m _ 17 Mass: m _ 13 ¼ m _ 12 ¼ m

ð530Þ

_ gen_II ¼ m _ 13 h13 þ Q _ 14 h14 þ m _ 16 h16 þ m _ 17 h17 Energy: m _ 12 h12 þ m

ð531Þ

_ gen_II =Tgen_II þ S_ gen;gen_II ¼ m _ 13 s13 þ Q _ 14 s14 þ m _ 16 s16 þ m _ 17 s17 Entropy: m _ 12 s12 þ m

ð532Þ

_ Q _ D;gen_I _ 14 ex 14 þ m _ 16 ex 16 þ m _ 17 ex17 þ Ex _ 13 ex 13 þ Ex Exergy: m _ 12 ex12 þ m gen_II ¼ m

ð533Þ

Expansion valve II: The mass, energy, entropy, and exergy balance equations for expansion valve II can be given under the steady state and steady flow conditions as follows: _ 15 Mass: m _ 14 ¼ m

ð534Þ

_ 15 h15 Energy: m _ 14 h14 ¼ m

ð535Þ

_ 15 s15 Entropy: m _ 14 s14 þ S_ gen;ev_II ¼ m

ð536Þ

_ D;ev_II _ 15 ex15 þ Ex Exergy: m _ 14 ex14 ¼ m

ð537Þ

Condenser: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the condenser are defined as follows: _ 16 ¼ m _3 Mass: m _ 15 þ m

ð538Þ

_ con _ 16 h16 ¼ m _ 3 h3 þ Q Energy: m _ 15 h15 þ m

ð539Þ

_ con =Tcon _ 16 s16 þ S_ gen;con ¼ m _ 3 s3 þ Q Entropy: m _ 15 s15 þ m

ð540Þ

_ Q þ Ex _ D;con _ 16 ex 16 ¼ m _ 3 ex3 þ Ex Exergy: m _ 15 ex15 þ m con

ð541Þ

Expansion valve III: The mass, energy, entropy, and exergy balance equations for expansion valve III can be defined under the steady state and steady flow conditions as follows: _4 Mass: m _3 ¼m

ð542Þ

_ 4 h4 Energy: m _ 3 h3 ¼ m

ð543Þ

_ 4 s4 Entropy: m _ 3 s3 þ S_ gen;ev_III ¼ m

ð544Þ

_ D;ev_III _ 4 ex 4 þ Ex Exergy: m _ 3 ex3 ¼ m

ð545Þ

Evaporator: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the evaporator are defined as follows: _5 Mass: m _4 ¼m

ð546Þ

_ eva ¼ m _ 5 h5 Energy: m _ 4 h4 þ Q

ð547Þ

_ eva =Teva þ S_ gen;eva ¼ m _ 5 s5 Entropy: m _ 4 s4 þ Q

ð548Þ

_ Q ¼m _ D;eva _ 5 ex 5 þ Ex Exergy: m _ 4 ex 4 þ Ex eva

ð549Þ

Absorber: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the absorber can be written as follows: _ 19 ¼ m _6 Mass: m _5þm

ð550Þ

_ abs _ 19 h19 ¼ m _ 6 h6 þ Q Energy: m _ 5 h5 þ m

ð551Þ

Geothermal Energy Conversion

533

_ abs =Tabs _ 19 s19 þ S_ gen;abs ¼ m _ 6 s6 þ Q Entropy: m _ 5 s5 þ m

ð552Þ

_ Q þ Ex _ D;abs _ 19 ex19 ¼ m _ 6 ex 6 þ Ex Exergy: m _ 5 ex5 þ m abs

ð553Þ

Pump: For the pump of the combined geothermal cooling system, the balance equations are provided under the steady state and steady flow conditions. _7 Mass: m _6 ¼m

ð554Þ

_ p¼m _ 7 h7 Energy: m _ 6 h6 þ W

ð555Þ

_ 7 s7 Entropy: m _ 6 s6 þ S_ gen;p ¼ m

ð556Þ

_ p ¼m _ D;p _ 7 ex 7 þ Ex Exergy: m _ 6 ex 6 þ W

ð557Þ

Expansion valve III: The mass, energy, entropy, and exergy balance equations for expansion valve III can be written under the steady state and steady flow conditions as follows: _ 19 Mass: m _ 18 ¼ m

ð558Þ

_ 19 h19 Energy: m _ 18 h18 ¼ m

ð559Þ

_ 19 s19 Entropy: m _ 18 s18 þ S_ gen;ev_III ¼ m

ð560Þ

_ D;ev_III _ 19 ex 19 þ Ex Exergy: m _ 18 ex18 ¼ m

ð561Þ

Heat exchanger II: The mass, energy, entropy, and exergy balance equations are written for HEX II under the steady state and steady flow conditions. _8; m _ 18 _ 17 ¼ m Mass: m _7¼m

ð562Þ

_ 17 h17 ¼ m _ 8 h8 þ m _ 18 h18 Energy: m _ 7 h7 þ m

ð563Þ

_ 17 s17 þ S_ gen;HEX_II ¼ m _ 8 s8 þ m _ 18 s18 Entropy: m _ 7 s7 þ m

ð564Þ

_ D;HEX_II _ 17 ex 17 ¼ m _ 8 ex8 þ m _ 18 ex18 þ Ex Exergy: m _ 7 ex 7 þ m

ð565Þ

The evaporator provides the cooling applications, and the energetic coefficient of performance (COPen) of the double effect absorption system can be defined as follows: COPen ¼

_ cooling Q _ gen I þ W _p Q

ð566Þ

The exergetic coefficient of performance (COPex) of the double effect absorption cooling system can be given as follows: COPex ¼

_ Q Ex cooling _ Q Ex gen

I

_p þW

ð567Þ

The exergy destruction rate, exergy destruction ratio, exergy efficiency, and heat transfer rate for the geothermal energy based double effect absorption cooling system devices are calculated by using balance and exergy efficiency equations, and given in Table 17. As shown in this table, the highest exergy destruction rate occurs in generator I and the evaporator with 248.12 and 201.13 kW, respectively, and the exergy efficiencies of these components are 56.27 and 45.46%, respectively. The component having the highest exergy efficiencies are expansion valve VI and expansion valve I with 99.77% and 99.34%, respectively. According to the thermodynamic assessment results, it is necessary to improve the development aims on this double effect absorption cooling process for the more efficient geothermal energy based cooling system design. As seen from Fig. 55, the ambient temperature increases from 5 to 401C, while energetic coefficient performance remains the same, but the COPex increases from about 0.43 to nearly 0.6. The reason for energetic COP remaining the same is that energy analysis is independent from the ambient temperature. However, the exergetic COP increases because the definition of exergy says that the exergy is related to the environment conditions. The impacts of varying ambient temperature on the geothermal energy based double effect absorption cooling system exergy destruction rate and exergy efficiency are shown in Fig. 56. It can be observed that the exergy destruction rate of the double effect absorption cooling system increases by increasing ambient temperature and decreasing exergetic efficiency. According to the findings, mass flow rate of geothermal fluid has positive effect on generated cooling effect from geothermal energy resources. As seen from Fig. 57, as geothermal mass flow rate increases from 7.5 to 22.5 kg/s, the amount of produced cooling effect increases from 1450 to 2050 kW. However, losses occurring in the system decrease the amount of transferred energy. This means that produced power increases with increasing mass flow rate whereas exergy efficiency of the system decreases.

534

Table 17

Geothermal Energy Conversion

Thermodynamic analysis results for geothermal energy based double effect absorption cooling system devices

Devices

Exergy destruction rate (kW)

Exergy destruction ratio (%)

Exergy efficiency (%)

Heat transfer rate (kW)

Generator I Generator II HEX I HEX II Pump Condenser Expansion valve Expansion valve Expansion valve Expansion valve Evaporator Absorber

248.12 156.57 56.26 34.14 88.55 62.47 4.053 4.6 1.57 1.72 201.13 60.57

0.85 0.53 0.19 0.11 0.30 0.21 0.013 0.015 0.005 0.005 0.68 0.20

56.27 21.09 78.05 59.68 60.12 72.68 99.34 84.77 82.68 99.77 45.46 22.64

253.5 158.4 161.1 95.75 55.36 161.1 3.366 1.932 1.509 3.039 301.9 49.57

I II III IV

0.65

2.5 COPen

2.4

COPex

2.3

0.6

2.1

0.55

2 1.9

COPex

COPen

2.2

0.5

1.8 1.7

0.45

1.6 1.5

0.4 5

10

15

20

25

30

35

40

T0 (°C) Fig. 55 Effect of ambient temperature on COPen and COPex of geothermal energy based double effect absorption cooling system.

0.173

950 ExD,DEACS ψDEACS

0.172 0.171

ExD,DEACS (kW)

850

0.17

800

0.169 750 0.168 700

0.167

650

0.166

600

0.165

System exergy efficiency

900

0.164

550 5

10

15

20

25

30

35

40

T0 (°C) Fig. 56 Exergy destruction rate and exergy efficiency of the geothermal energy based double effect absorption cooling system depending on the reference temperature changes.

Geothermal Energy Conversion

2100 Qcooling

0.178

ψDEACS

0.176

1900

0.174 0.172

1800

0.17 1700

0.168 0.166

1600

0.164 1500

System exergy efficiency

0.18

2000

Qcooling (kW)

535

0.162

1400 7.5

10

12.5

15

17.5

20

0.16 22.5

mgeothermal (kg/s) Fig. 57 Effect of mass flow rate of geothermal fluid on net cooling generation and exergy efficiency.

1699

Qcooling

0.169

ψDEACS

0.168

Qcooling (kW)

1659

0.167

1619

0.166

1579

0.165 0.164

1539

0.163 1499

0.162

1459 1419 150

System exergy efficiency

0.17

1739

0.161 160

170

180

190

200

210

220

0.16 230

Tgeothermal (°C) Fig. 58 Effect of geothermal fluid temperature on net cooling generation and exergy efficiency.

These results show the importance of exergy efficiency of second law efficiency because energy efficiency analyses do not point out these losses. The effect of geothermal fluid temperature on produced cooling effect and exergy efficiency is shown in Fig. 58. While geothermal fluid temperature increases from 150 to 2301C, cooling effect generation increases from 1418 to 1739 kW, and also with the same temperature change exergy efficiency of the system increases from about 16.12% to about 16.91%.

4.11.6.8

Geothermal Energy Based Hydrogen Production and Liquefaction System

The schematic diagram of integrated hydrogen production system driven by geothermal energy is presented in Fig. 59. The integrated system investigated in this case study consists of mainly four subsystems: (1) double flash geothermal process, (2) ORC, (3) PEM electrolyzer, and (4) hydrogen liquefaction process. The double flash geothermal process and ORC are used in the integrated system to produce heat and power for the PEM electrolyzer. The ORC process runs when the geothermal working fluid goes into the vaporizer at points 7 and 12. In this process, the thermal energy of geothermal working fluid is transferred by using the vaporizer to the different working fluid for use in the ORC process. The ORC working fluid goes into the turbine at point 16 to generate power, and leaves at point 17. Exiting from HEX I, the geofluid transfers its heat energy to the water before entering the PEM electrolyzer. Also, the power generated using by the ORC process is used in the PEM electrolyzer to generate hydrogen. The produced hydrogen is in gaseous form at reference conditions. The hydrogen liquefaction subsystem is used for more efficient hydrogen storage. The hydrogen liquefaction process is relatively more energy intensive than compression of hydrogen, whereas, the density of liquid hydrogen is nearly 1120 kg/m3 and liquid hydrogen is 29 times better than compressed hydrogen at 700 bar,

536

Geothermal Energy Conversion

Flash chamber I 1 2

16

7 3

Vaporizer

Flash separator I

12 13

22

9

8

Turbine HEX I

Production well Flash chamber II 4 5

Pump I 21 Flash separator II

17

6 HEX II 14

10

20

11

18

Pump-III Pump II

19 Condenser

Production well

a Compressor

Mixer 23 Electrolysis water

PEM electrolyzer preheating 15

24

PEM electrolyzer

26

b

28

27

N2(gas)

29

40

HEX IV

HEX III

39

N2(liq) 25

30 37

38 Oxygen

HEX V

Injection

31 36

Well

N2(gas)

35 Liquid hydrogen tank

43

HEX VI

HEX VII

Separator 34 33 Expansion valve

32

42 41 N2(liq)

Fig. 59 Schematic diagram of geothermal energy based hydrogen production and liquefaction system.

in terms of volume work. Therefore, the Linde–Hampson hydrogen liquefaction process with a secondary nitrogen cooling is defined for hydrogen storage. The balance equations for double flash geothermal process based hydrogen production and liquefaction system components are defined in the next subsections. Flash chamber I: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for flash chamber I can be given as follows: _2 Mass: m _1 ¼m

ð568Þ

_ 2 h2 Energy: m _ 1 h1 ¼ m

ð569Þ

_ 2 s2 Entropy: m _ 1 s1 þ S_ gen;fc_I ¼ m

ð570Þ

_ D;fc_I _ 2 ex2 þ Ex Exergy: m _ 1 ex 1 ¼ m

ð571Þ

Flash separator I: The mass, energy, entropy, and exergy balance equations are defined for flash separator I under steady state and steady flow conditions. _3þm _8 Mass: m _2¼m

ð572Þ

Geothermal Energy Conversion

537

_ 3 h3 þ m _ 8 h8 Energy: m _ 2 h2 ¼ m

ð573Þ

_ 3 s3 þ m _ 8 s8 Entropy: m _ 2 s2 þ S_ gen;fs_I ¼ m

ð574Þ

_ D;fs_I _ 3 ex 3 þ m _ 8 ex 8 þ Ex Exergy: m _ 2 ex 2 ¼ m

ð575Þ

Pump I: The mass, energy, entropy, and exergy balance equations for pump I can be written under the steady state and steady flow conditions as follows: _9 Mass: m _8 ¼m

ð576Þ

_ p_I ¼ m _ 9 h9 Energy: m _ 8 h8 þ W

ð577Þ

_ p_I ¼ m _ 9 h9 Energy: m _ 8 h8 þ W

ð578Þ

_ p_I ¼ m _ D;p_I _ 9 ex 9 þ Ex Exergy: m _ 8 ex 8 þ W

ð579Þ

Flash chamber II: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for flash chamber II can be written given as follows: _5 Mass: m _4 ¼m

ð580Þ

_ 5 h5 Energy: m _ 4 h4 ¼ m

ð581Þ

_ 5 s5 Entropy: m _ 4 s4 þ S_ gen;fc_II ¼ m

ð582Þ

_ D;fc_II _ 5 ex 5 þ Ex Exergy: m _ 4 ex 4 ¼ m

ð583Þ

Flash separator II: The mass, energy, entropy, and exergy balance equations are defined for flash separator II under steady state and steady flow conditions. _6þm _ 10 Mass: m _5¼m

ð584Þ

_ 6 h6 þ m _ 10 h10 Energy: m _ 5 h5 ¼ m

ð585Þ

_ 6 s6 þ m _ 10 s10 Entropy: m _ 5 s5 þ S_ gen;fs_II ¼ m

ð586Þ

_ D;fs_II _ 6 ex 6 þ m _ 10 ex 10 þ Ex Exergy: m _ 5 ex5 ¼ m

ð587Þ

Pump II: The mass, energy, entropy, and exergy balance equations for pump II can be given under the steady state and steady flow conditions as follows: _ 11 Mass: m _ 10 ¼ m

ð588Þ

_ p_II ¼ m _ 11 h11 Energy: m _ 10 h10 þ W

ð589Þ

_ 11 s11 Entropy: m _ 10 s10 þ S_ gen;p_II ¼ m

ð590Þ

_ p_II ¼ m _ D;p_II _ 11 ex 11 þ Ex Exergy: m _ 10 ex10 þ W

ð591Þ

Vaporizer: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the vaporizer can be defined as follows: _ 12 ¼ m _ 13 ; m _ 22 _ 16 ¼ m Mass: m _7þm

ð592Þ

_ 12 h12 þ m _ 22 h22 ¼ m _ 13 h13 þ m _ 16 h16 Energy: m _ 7 h7 þ m

ð593Þ

_ 12 s12 þ m _ 22 s22 þ S_ gen;vap ¼ m _ 13 s13 þ m _ 16 s16 Entropy: m _ 7 s7 þ m

ð594Þ

_ D;vap _ 12 ex12 þ m _ 22 ex 22 ¼ m _ 13 ex 13 þ m _ 16 ex 16 þ Ex Exergy: m _ 7 ex7 þ m

ð595Þ

Turbine: The mass, energy, entropy, and exergy balance equations are written for the turbine under the steady state and steady flow conditions. _ 17 Mass: m _ 16 ¼ m

ð596Þ

_ tur _ 17 h17 þ W Energy: m _ 16 h16 ¼ m

ð597Þ

_ 17 s17 Entropy: m _ 16 s16 þ S_ gen;tur ¼ m

ð598Þ

_ tur þ Ex _ D;tur _ 17 ex17 þ W Exergy: m _ 16 ex16 ¼ m

ð599Þ

538

Geothermal Energy Conversion

Heat exchanger II: The mass, energy, entropy, and exergy balance equations are written for HEX II under the steady state and steady flow conditions. _ 18 ; m _ 21 _ 20 ¼ m Mass: m _ 17 ¼ m

ð600Þ

_ 20 h20 ¼ m _ 18 h18 þ m _ 21 h21 Energy: m _ 17 h17 þ m

ð601Þ

_ 20 s20 þ S_ gen;HEX_II ¼ m _ 18 s18 þ m _ 21 s21 Entropy: m _ 17 s17 þ m

ð602Þ

_ D;HEX_II _ 20 ex 20 ¼ m _ 18 ex 18 þ m _ 21 ex 21 þ Ex Exergy: m _ 17 ex17 þ m

ð603Þ

Condenser: Under the steady state and steady flow conditions, the mass, energy, entropy, and exergy balance equations for the condenser can be defined as follows: _ 19 ; m _b _a¼m Mass: m _ 18 ¼ m

ð604Þ

_ 18 h18 ¼ m _ b hb þ m _ 19 h19 Energy: m _ a ha þ m

ð605Þ

_ 18 s18 þ S_ gen;Con ¼ m _ b sb þ m _ 19 s19 Entropy: m _ a sa þ m

ð606Þ

_ D;Con _ 18 ex 18 ¼ m _ b exb þ m _ 19 ex19 þ Ex Exergy: m _ a exa þ m

ð607Þ

Pump III: The mass, energy, entropy, and exergy balance equations for pump III can be written under the steady state and steady flow conditions as follows: _ 20 Mass: m _ 19 ¼ m

ð608Þ

_ p_III ¼ m _ 20 h20 Energy: m _ 19 h19 þ W

ð609Þ

_ 20 s20 Entropy: m _ 19 s19 þ S_ gen;p_III ¼ m

ð610Þ

_ p_III ¼ m _ D;p_III _ 20 ex 20 þ Ex Exergy: m _ 19 ex19 þ W

ð611Þ

Proton exchange membrane electrolyzer preheating: The mass, energy, entropy, and exergy balance equations for PEM electrolyzer preheating can be given under the steady state and steady flow conditions as follows: _ 15 ; m _ 24 _ 23 ¼ m Mass: m _ 14 ¼ m

ð612Þ

_ 23 h23 ¼ m _ 15 h15 þ m _ 24 h24 Energy: m _ 14 h14 þ m

ð613Þ

_ 23 s23 þ S_ gen;peph ¼ m _ 15 s15 þ m _ 24 s24 Entropy: m _ 14 s14 þ m

ð614Þ

_ D;peph _ 23 ex 23 ¼ m _ 15 ex 15 þ m _ 24 ex 24 þ Ex Exergy: m _ 14 ex14 þ m

ð615Þ

Proton exchange membrane electrolyzer: The mass, energy, entropy, and exergy balance equations for PEM electrolyzer can be defined under the steady state and steady flow conditions as follows: _ 25 þ m _ 26 Mass: m _ 24 ¼ m

ð616Þ

_ T ¼m _ 25 h25 þ m _ 26 h26 Energy: m _ 24 h24 þ W

ð617Þ

_ 25 s25 þ m _ 26 s26 Entropy: m _ 24 s24 þ S_ gen;PEM_el ¼ m

ð618Þ

_ T ¼m _ D;PEM_el _ 25 ex 25 þ m _ 26 ex 26 þ Ex Exergy: m _ 24 ex 24 þ W

ð619Þ

Mixer: The mass, energy, entropy, and exergy balance equations for the mixer can be written under the steady state and steady flow conditions as follows: _ 38 ¼ m _ 27 Mass: m _ 26 þ m

ð620Þ

_ 38 h38 ¼ m _ 27 h27 Energy: m _ 26 h26 þ m

ð621Þ

_ 38 s38 þ S_ gen;mixer ¼ m _ 27 s27 Entropy: m _ 26 s26 þ m

ð622Þ

_ D;mixer _ 38 ex 38 ¼ m _ 27 ex 27 þ Ex Exergy: m _ 26 ex 26 þ m

ð623Þ

Compressor: The mass, energy, entropy, and exergy balance equations for the compressor can be given under the steady state and steady flow conditions as follows: _ 28 Mass: m _ 27 ¼ m

ð624Þ

Geothermal Energy Conversion

539

_ cmp ¼ m _ 28 h28 Energy: m _ 27 h27 þ W

ð625Þ

_ 28 s28 Entropy: m _ 27 s27 þ S_ gen;cmp ¼ m

ð626Þ

_ cmp ¼ m _ D;cmp _ 28 ex 28 þ Ex Exergy: m _ 27 ex 27 þ W

ð627Þ

Heat exchanger III: The mass, energy, entropy, and exergy balance equations for HEX III can be written under the steady state and steady flow conditions as follows: _ 29 ; m _ 38 Mass: m _ 28 ¼ m _ 37 ¼ m

ð628Þ

_ 37 h37 ¼ m _ 29 h29 þ m _ 38 h38 Energy: m _ 28 h28 þ m

ð629Þ

_ 37 s37 þ S_ gen;HEX_III ¼ m _ 29 s29 þ m _ 38 s38 Entropy: m _ 28 s28 þ m

ð630Þ

_ D;HEX_III _ 37 ex 37 ¼ m _ 29 ex29 þ m _ 38 ex 38 þ Ex Exergy: m _ 28 ex28 þ m

ð631Þ

Heat exchanger IV: The mass, energy, entropy, and exergy balance equations for HEX IV can be defined under the steady state and steady flow conditions as follows: _ 30 ; m _ 40 _ 39 ¼ m Mass: m _ 29 ¼ m

ð632Þ

_ 39 h39 ¼ m _ 30 h30 þ m _ 40 h40 Energy: m _ 29 h29 þ m

ð633Þ

_ 39 s39 þ S_ gen;HEX_IV ¼ m _ 30 s30 þ m _ 40 s40 Entropy: m _ 29 s29 þ m

ð634Þ

_ D;HEX_IV _ 39 ex39 ¼ m _ 30 ex30 þ m _ 40 ex 40 þ Ex Exergy: m _ 29 ex29 þ m

ð635Þ

Heat exchanger V: The mass, energy, entropy, and exergy balance equations for HEX V can be written under the steady state and steady flow conditions as follows: _ 31 ; m _ 37 Mass: m _ 30 ¼ m _ 36 ¼ m

ð636Þ

_ 36 h36 ¼ m _ 31 h31 þ m _ 37 h37 Energy: m _ 30 h30 þ m

ð637Þ

_ 36 s36 þ S_ gen;HEX_V ¼ m _ 31 s31 þ m _ 37 s37 Entropy: m _ 30 s30 þ m

ð638Þ

_ D;HEX_V _ 36 ex 36 ¼ m _ 31 ex31 þ m _ 37 ex 37 þ Ex Exergy: m _ 30 ex30 þ m

ð639Þ

Heat exchanger VI: The mass, energy, entropy, and exergy balance equations for HEX VI can be given under the steady state and steady flow conditions as follows: _ 32 ; m _ 42 Mass: m _ 31 ¼ m _ 41 ¼ m

ð640Þ

_ 41 h41 ¼ m _ 32 h32 þ m _ 42 h42 Energy: m _ 31 h31 þ m

ð641Þ

_ 41 s41 þ S_ gen;HEX_VI ¼ m _ 32 s32 þ m _ 42 s42 Entropy: m _ 31 s31 þ m

ð642Þ

_ D;HEX_VI _ 41 ex41 ¼ m _ 32 ex32 þ m _ 42 ex42 þ Ex Exergy: m _ 31 ex31 þ m

ð643Þ

Heat exchanger VII: The mass, energy, entropy, and exergy balance equations for HEX VII can be defined under the steady state and steady flow conditions as follows: _ 33 ; m _ 36 Mass: m _ 32 ¼ m _ 35 ¼ m

ð644Þ

_ 35 h35 ¼ m _ 33 h33 þ m _ 36 h36 Energy: m _ 32 h32 þ m

ð645Þ

_ 35 s35 þ S_ gen;HEX_VII ¼ m _ 33 s33 þ m _ 36 s36 Entropy: m _ 32 s32 þ m

ð646Þ

_ D;HEX_VII _ 35 E35 ¼ m _ 33 ex 33 þ m _ 36 ex 36 þ Ex Exergy: m _ 32 ex32 þ m

ð647Þ

Expansion valve: The mass, energy, entropy, and exergy balance equations for the expansion valve can be written under the steady state and steady flow conditions as follows: _ 34 Mass: m _ 33 ¼ m

ð648Þ

_ 34 h34 Energy: m _ 33 h33 ¼ m

ð649Þ

540

Geothermal Energy Conversion _ 34 s34 Entropy: m _ 33 s33 þ S_ gen;ev ¼ m

ð650Þ

_ D;ev _ 34 ex34 þ Ex Exergy: m _ 33 ex33 ¼ m

ð651Þ

Separator: The mass, energy, entropy, and exergy balance equations for the separator can be defined under the steady state and steady flow conditions as follows: _ 35 þ m _ 43 Mass: m _ 34 ¼ m

ð652Þ

_ 35 h35 þ m _ 43 h43 Energy: m _ 34 h34 ¼ m

ð653Þ

_ 35 s35 þ m _ 43 s43 Entropy: m _ 34 s34 þ S_ gen;sep ¼ m

ð654Þ

_ D;sep _ 35 ex 35 þ m _ 43 ex 43 þ Ex Exergy: m _ 34 ex 34 ¼ m

ð655Þ

The effects of varying reference temperature on the geothermal energy based power production rate and hydrogen production rate are illustrated in Fig. 60. It can be observed that the power and hydrogen production rate of the double flash geothermal power system based integrated system increases by increasing the reference temperature from 0 to 401C. The effects of geothermal fluid temperature on power production rate and hydrogen production rate are illustrated in Fig. 61. While the geothermal fluid temperature increases from 150 to 2001C, the power production rate increases from 1041 to 7937 kW, 0.03

4300

Power production rate (kW)

mH2

4100

0.029

4000 0.028

3900 3800

0.027

3700 3600

0.026

Hydrogen production rate (kg/s)

Wturbine

4200

3500 3400 0

5

10

15

20

25

30

35

0.025 40

Reference temperature (°C) Fig. 60 Effect of reference temperature on net power and hydrogen production.

8000

0.056

Power production rate (kW)

7000

m H2

0.046

6000 5000

0.036

4000 0.026 3000 0.016

2000 1000 130

140

150

160

170

180

Geothermal source temperature (°C) Fig. 61 Effect of geothermal source temperature on net power and hydrogen production.

190

0.006 200

Hydrogen production rate (kg/s)

Wturbine

Geothermal Energy Conversion

541

4800

Wturbine

4500

m H2

4200

0.036 0.033 0.03

3900 0.027 3600 0.024

3300 3000

0.021

2700

0.018

2400 1500

2000

2500

3000

3500

Hydrogen production rate (kg/s)

Power production rate (kW)

5100

0.015 4000

Turbine inlet pressure (kPa) Fig. 62 Effect of turbine inlet pressure on net power and hydrogen production.

and also with the same temperature change hydrogen production rate of the integrated system increases from 0.0062 kg/s to about 0.0558 kg/s, respectively. Fig. 62 shows the impact of ORC turbine inlet pressure on power production rate and hydrogen production rate of the integrated system. As shown in this figure, an increase in this pressure reduces the power and hydrogen production rate. The energy balance equation for control volume around the vaporizer illustrates that when the energy input from the geothermal resource is constant, the reduction in turbine inlet enthalpy increases the ORC mass flow rate.

4.11.7

Future Directions

In this section, it is important to discuss the primary research areas and priorities of geothermal energy and the future directions on conversion of geothermal heat into numerous useful outputs in order to provide some abstract guidance. The possible main research areas and application fields of geothermal energy resources in the future based on the five subsections, namely (1) advances in exploration and drilling technologies, (2) power generation, (3) heating and cooling applications, (4) desalination, and (5) environmental impact mitigation public acceptance, are given in Table 18. Based on the report of researchers at the Massachusetts Institute of Technology, the geothermal power systems can provide a strong, long-lasting alternative with attributes to complement other significant energy production processes from clean coal gasification, nuclear, solar, wind, hydropower, and biomass [34]. Geothermal power sources show numerous benefits over the alternative energy resources. As an example, unlike biomass, fossil fuel, or nuclear energy sources, geothermal energy is locationspecific and requires no transportation of raw material from the source of extraction to the power plant. The primary utilizations of geothermal energy cover a wide range of applications, such as residential heating and domestic hot water supply, aquaculture, greenhouse heating, swimming pools and balneology, industrial heating processes, heat pumps and power production. But for the higher geothermal system efficiency, decreased thermal losses and wastes, decreased operating costs, decreased harmful gaseous emissions, better use of geofluid sources, multiple generation options, and increased reliability, the geothermal energy based integrated systems depending on the local availability of resources should be designed and operated. The outputs from this study can assist designers in developing more energy-efficient geothermal power systems in an integrated form. The geothermal energy should be incorporated with different alternative energy sources, such as solar, biomass, wind, etc., in the actual processes, depending on the local availability of resources for some useful output generations. The suggestions and thermodynamic analyses of geothermal energy based integrated systems offer several main areas of future research as summarized below:

• •

In the proposed integrated geothermal power process model for multigeneration aims, such as power, hydrogen, heating, cooling, and freshwater production, the required capital investment and operating cost for the desired production capacities per day can be further estimated. The geothermal energy based ORC process, triple flash power generation, binary cycle power generation, combined/hybrid power generation, and also Kalina cycle are the efficient options to generate electricity from low grade temperature heat sources. But, the required capital investment and operating cost and also life cycle assessment for the desired production capacities per day should be further estimated based on the design parameters of the installation area. The conclusions should be evaluated and compared, so that a suitable integrated system can be found for the specific operating conditions.

Geothermal Energy Conversion

542

Table 18

Possible main research areas and application fields of geothermal energy resources

•

Advances in exploration and drilling technologies:

•

Power generation:

•

Heating and cooling applications:

•

Desalination:

•

Environmental impact mitigation public acceptance:

To To To To To To To To To To To To To To To To To To To To To To To To To To To To To

decrease drilling cost reduce boreholes amounts decrease drilling times investigate new drilling technologies develop novel 3D models of geothermal resources better estimate geothermal power potential. build large-scale demonstration of geothermal power plant for cost reduction investigate new materials and operational methods for decreasing the corrosion effect and increasing efficiency of geothermal power plant develop novel technologies for decreasing the O&M costs of geothermal power plants analyze the new low temperature based power and other synthetic fuel production processes with high performance improve the novel designs of both water-cooled and air-cooled condenser equipment investigate the ORC working fluids with high enthalpy content for low temperature geothermal power generation systems build off-shore geothermal power systems based on production of deep marine environments investigate the innovative mechanism for power generation. investigate more profitable options for residential and commercial heating and cooling applications with geothermal resources build large-scale research projects for heating and cooling sectors with geothermal resources develop novel district heating and cooling network solve problems of local procedures for the district heating and cooling networks promote the development of residential geothermal heating with the closed-loop system (or without water exchange) decrease components costs, such as heat exchanger (HEX), compressor, expansion valve, pump and fan investigate the innovative mechanism for geothermal heating and cooling systems. build the seawater demonstration project with low enthalpy geothermal resources analyze the multistage distillation process for more efficiently system design aims investigate more efficient membrane for desalination applications investigate and decrease the commercial scale problems, technical design problems, and high investment cost for desalination applications. investigate suitable technologies for removing of some potentially toxic elements, condensable gaseous and solid residues develop new technologies for analysis of the seismic activity induced by geothermal energy resources support the reduction in greenhouse gaseous emissions by using the geothermal energy for building of public acceptance investigate smart controlled heating and cooling systems and other innovation solutions for higher living standards.

Source: International Energy Agency (IEA). Technology Roadmap, Geothermal Heat and Power. http://www.iea.org/publications/freepublications/publication/Geothermal_roadmap.pdf; 2011 [accessed March 2017].

• • • •

• •

Using geothermal energy sources to produce hydrogen can reduce the costs even further. The more detailed cost accounting and exergoeconomic assessment for different designs of geothermal energy based hydrogen production process should be analyzed for comparison purposes. The low temperature thermochemical cycles for hydrogen or alternative fuel production can be integrated with geothermal energy resources and investigated in terms of thermodynamic feasibilities. The effects of different alternative fuels instead of hydrogen energy in the geothermal energy based integrated system on the total operating cost are suggested to be evaluated. Hydrogen can be stored in gas, liquid, or solid form. Because of its low-density rate, the gas hydrogen needs large volumes for storage aims, thereby necessitating compression, extremely low temperatures to convert hydrogen to a cryogenic liquid form or combinations with other materials for solid storage. For liquid storage, the geothermal energy based hydrogen liquefaction processes should be designed and analyzed based on exergy analysis viewpoint. The detailed exergy and exergoeconomic analysis should be utilized to investigate the impact of the avoidable and unavoidable portions of exergy destruction in each subcomponent and optimize for determining the avoidable exergy destruction in each subcomponent. Finally, improving the geothermal energy based multigeneration performance would reduce the greenhouse gas emissions and harmful environmental impact, and enhance sustainability.

4.11.8

Concluding Remarks

In the present study, a review of geothermal energy sources and geothermal energy conversion models is provided for power, hydrogen, heating, cooling, and freshwater generation, and the main balance equations for thermodynamic analysis of different

Geothermal Energy Conversion

543

geothermal power systems are described and compared, and recent model developments are discussed. The energetic and exergetic analyses are described and contrasted for direct, single flash, double flash, triple flash, binary cycle, and combined geothermal power generation processes. The exergy based thermodynamic assessment is demonstrated to ensure an important tool for design and improving of geothermal power systems. The parametric analyses define the impacts on exergy destruction rate and exergy efficiency of varying different operating indicators, such as geofluid temperature and mass flow rate. Also, the effects of increasing reference temperature on exergy destruction rate and exergy efficiency are investigated. Different concluding outputs should be drawn from this study:

• • • • • • • •

The new potential fields for the construction of geothermal energy based power generation plants need to be investigated. Identifying design procedures, construction techniques, systematic operation, and maintenance practices of geothermal resource integrated systems for multigeneration aims is of immense value to the designers. Therefore, it is recommended to start the design application on the geothermal energy based integrated systems based on resource temperature. One of the biggest drawbacks that limit the utilization of geothermal power systems is the lack of a reliable long-term energy storage system. Therefore, it is recommended that a research work should be carried out for the application of geothermal power systems with other means of storage, such as the liquid hydrogen storage system. The exergy efficiency of direct, single flash, double flash, triple flash, binary cycle, and combined geothermal power generation systems are found to be 39.98, 36.21, 39.5, 43.99, 48.35, and 49.15%, respectively, at a reference state temperature of 251C and pressure of 101.3 kPa. The advanced combined geothermal power generation system is more efficient than the other investigated systems. The largest irreversibility in the geothermal power generation processes is associated with flashing, condenser, and turbine. It is demonstrated that increasing the geofluid temperature, mass flow rate, and reference temperature increase the power generation rate. It is shown that the reference temperature increases the exergy efficiency. The geofluid temperature and mass flow rate increase the exergy efficiency of direct, single flash, double flash, and triple flash geothermal power generation systems whereas they decrease the exergy efficiency of binary cycle and combined geothermal power generation systems.

References [1] [2] [3] [4] [5] [6] [7] [8] [9] [10] [11] [12] [13] [14] [15] [16] [17] [18] [19] [20] [21] [22] [23] [24] [25] [26] [27] [28] [29] [30] [31] [32] [33]

Ozturk M, Yuksel YE. Energy structure of turkey for sustainable development. Renew Sustainable Energy Rev 2016;53:1259–72. Coskun C, Oktay Z, Dincer I. Performance evaluations of a geothermal power plant. Appl Therm Eng 2011;31:4074–82. John WL, Tonya LB. Direct utilization of geothermal energy 2015 worldwide review. Geothermics 2016;60:66–93. Fallah M, Mohammad S, Mahmoudi S, Yari M, Ghiasi R. Advanced exergy analysis of the Kalina cycle applied for low temperature enhanced geothermal system. Energy Convers Manag 2016;108:190–201. Balta M, Dincer I, Hepbasli A. Thermodynamic assessment of geothermal energy use in hydrogen production. Int J Hydrogen Energy 2009;34(7):2925–39. Lee K. Classification of geothermal resources by exergy. Geothermics 2001;30(4):431–42. DiPippo R. Geothermal power plants: principles, applications, case studies and environmental impact. 2nd ed. Burlington: Elsevier; 2008. Erdlac R, Gross P, McDonald E. A proposed new geothermal power classification system. Trans Geotherm Resources Council 2008;32:379–84. Williams C, DeAngelo J. Mapping geothermal potential in the western United States. Trans Geother Resources Council 2008;32:181–8. Green B, Gerald R. Geothermal the energy under our feet NREL/TP-840-40665, 2006. DiPippo R, Marcille D. Exergy analysis of geothermal power plants. Geotherm Resources Council Trans 1984;8:47–52. Rybach L. The future of geothermal energy and its challenges. In: Proc. World geothermal congress, Bali, Indonesia, April 25–29; 2010. Falcone G, Gnoni A, Harrison B, Alimonti C. Classification and reporting requirements for geothermal resources. In: European Geothermal Congress, Pisa, Italy, June 3–7 June; 2013. Barbier E. Geothermal energy technology and current status: an overview. Renew Sustainable Energy Rev 2002;6:3–65. Muffler P, Cataldi R. Methods for regional assessment of geothermal resources. Geothermics 1978;7:53–89. Blackwell DB, Steele JL, Carter LS. Heat-flow patterns of the North American continent; a discussion of the geothermal map of North America. In: Selmons DB, Engdahl ER, Zoback MD, Blackwell DB, editors. Neotectonics of North America. vol. 1. Boulder, CO: Geological Society of America; 1991. Rummel F, Kappelmeyer O, editors. Geothermal energy-future energy source? Karlsruhe: Verlag C.F. Müller; 1993. Dincer I, Rosen MA. Exergy: energy, environment and sustainable development. 2nd ed. Oxford: Newnes; 2012. Moran M. Availability analysis: a guide to efficient energy usage. Englewood Cliffs, NJ: Prentice-Hall; 1982. Bejan A, Tsatsaronis G, Moran M. Thermal design and optimization. New York, NY: Wiley Interscience; 1996. Kotas T. The exergy method of thermal plant analysis. Malabar, FL: Krieger Publishing Company; 1985. Dincer I, Rosen MA. Exergy analysis of heating, refrigerating, and air: conditioning methods and applications. New York, NY: Elsevier; 2015. American Society of Heating, Refrigerating and Air-Conditioning Engineers; Caneta Research, Inc. Commercial/Institution Ground-Source Heat Pumps Engineering Manual. Atlanta, GA, USA; 1995. Ozturk M. Energy and exergy analysis of a combined ground source heat pump system. Appl Therm Eng 2014;73(1):360–8. Nandi T, Sarangi S. Performance and optimization of hydrogen liquefaction cycles. Int J Hydrogen Energy 1993;18(2):131–9. Bracha M, Lorenz A, Patzelt A, Wanner M. Large-scale hydrogen liquefaction in Germany. Int J Hydrogen Energy 1994;19(1):53–9. Gurau V, Barbir F, Liu H. An analytical solution of a half-cell model for PEM fuel cells. J Electrochem Soc 2000;147(7):2468–77. Hamann CH, Hamnett A, Vielstich W. Electrochemistry. Weinheim: Wiley-VCH; 2007. Thampan T, Malhotra S, Zhang J, Datta R. PEM fuel cell as a membrane reactor. Catalysis Today 2001;67(1-3):15–32. Sigfusson B, Uihlein A. Technology, market and economic aspects of geothermal energy in Europe, JRC geothermal energy status report, 2015. Ratlamwala T, Dincer I. Energetic and exergetic investigation of novel multi-flash geothermal systems integrated with electrolyzers. J Power Sources 2014;254:306–15. Kanoglu M. Exergy Analysis of a dual-level binary geothermal power plant. Geothermics 2002;31:709–24. Yari M. Exergetic analysis of various types of geothermal power plants. Renew Energy 2010;35:112–21.

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[34] Massachusetts Institute of Technology. The future of geothermal energy: impact of enhanced geothermal systems on the United States in the 21st century, An assessment by Massachusetts Institute of Technology-led Interdisciplinary Panel. http://geothermal.inel.gov/publications/future_of_geothermal_energy.pdf; 2006 [accessed March 2017].

Further Reading Arnold W. 2013. Geothermal Engineering: Fundamentals and Applications. New York, NY: Springer; 2013. Boden DR. 2016. Geologic Fundamentals of Geothermal Energy. New York, NY: Taylor and Francis, CRC Press; 2016. Dincer I. 2017. Refrigeration Systems and Applications. third ed. London: John Wiley & Sons, Ltd.; 2017. p. 727. Dincer I, Ratlamwala T. 2016. Integrated Absorption Refrigeration Systems: Comparative Energy and Exergy Analyses. New York, NY: Springer Verlag; 2016. p. 270. Dincer I, Rosen MA. 2015. Exergy Analysis of Heating, Refrigerating and Air Conditioning. Oxford: Elsevier Science, Ltd; 2015. p. 388. Dincer I, Rosen MA, Ahmadi P. 2017. Optimization of Energy Systems. London: John Wiley & Sons, Ltd.; 2017. p. 453. Dincer I, Zamfirescu C. 2014. Advanced Power Generation Systems. Oxford: Elsevier Science, Ltd; 2014. p. 644. Dincer I, Zamfirescu C. 2016. Sustainable Hydrogen Production. Oxford: Elsevier Science, Ltd.; 2016. p. 479. DiPippo R. 2012. Geothermal Power Plants: Principles, Applications, Case Studies and Environmental Impact. third ed. Boston, MA: Elsevier; 2012. European Communities. 1999. Blue Book on Geothermal Resources. Luxembourg: Office for Official Publications of the European Communities; 1999. Grant M, Bixley P. 2011. Geothermal Reservoir Engineering. second ed. New York, NY: Academic Press; 2011. Gupta H, Roy S. 2006. Geothermal Energy: An Alternative Resource for the 21st Century. Oxford: Elsevier; 2006. Hance C. 2005. Factors Affecting Costs of Geothermal Power Development. Geothermal Energy Association, Department of Energy; 2005. Hance CN. 2005. Factors Affecting Costs of Geothermal Power Development. Washington, DC: Geothermal Energy Association; 2005. Stober I, Bucher K. 2013. Geothermal Energy: From Theoretical Models to Exploration and Development. New York, NY: Springer; 2013. Suleman F, Dincer I, Agelin-Chaab M. 2014. Development of an integrated renewable energy system for multigeneration. Energy 2014;78:196–204. Tester JW, Anderson BJ, Batchelor AS, et al. 2006. The Future of Geothermal Energy. Cambridge, MA: MIT Press; 2006.

Relevant Websites http://www.ferc.gov/ Federal Energy Regulatory Commission. http://www.geo-energy.org/ Geothermal Energy Association. http://www.geothermal.org Geothermal Resources Council. http://www.geothermal-energy.org International Geothermal Association. http://www.nationalgeographic.com/environment/global-warming/geothermal-energy/ National Geographic Partners, LLC. http://www.eere.energy.gov/geothermal Office of Energy Efficiency & Renewable Energy. http://www.renewableenergyworld.com/geothermal-energy/tech.html Renewable Energy World. http://www.smu.edu/geothermal SMU Geothermal Laboratory. http://www.unr.edu/geothermal/links.html University of Nevada. https://www.eia.gov/energyexplained/index.cfm?page=geothermal_power_plants US Department of Energy.

4.12 Hydropower Conversion Zekâi S-en, Turkish Water Foundation, Istanbul, Turkey and Istanbul Technical University, Istanbul, Turkey r 2018 Elsevier Inc. All rights reserved.

4.12.1 4.12.2 4.12.3 4.12.3.1 4.12.3.2 4.12.3.3 4.12.4 4.12.4.1 4.12.4.2 4.12.4.3 4.12.4.4 4.12.4.4.1 4.12.4.4.2 4.12.4.4.2.1 4.12.4.4.2.2 4.12.4.4.2.3 4.12.4.4.2.4 4.12.4.4.2.5 4.12.4.4.2.6 4.12.4.4.2.7 4.12.4.4.2.8 4.12.4.4.2.9 4.12.4.4.2.10 4.12.4.5 4.12.4.5.1 4.12.4.5.2 4.12.4.5.3 4.12.5 4.12.5.1 4.12.5.1.1 4.12.5.1.2 4.12.5.1.3 4.12.5.1.4 4.12.5.1.4.1 4.12.5.1.4.1.1 4.12.5.1.4.1.2 4.12.5.1.4.1.3 4.12.6 4.12.7 4.12.8 4.12.9 4.12.10 4.12.11 References

4.12.1

Introduction Focus on Hydropower Hydropower Generation Fundamentals Hydrological Cycle and Hydropower Hydropower and Climate Change Hydropower Sustainability Hydropower Systems Reservoir Impoundments Run-of-River Pumped Storage Other Alternatives Tidal power Wave power Oscillating float system Oscillating paddle system Oscillating snake system Oscillating water column Pressure transducer system Wave capture systems Overtopping wave systems Tapchan model Lever systems Technical challenges Small Hydropower Plants Micro hydropower plants Pico hydropower plants Underground hydropower plants Analysis and Assessment Determination of the Required Capacity Graphical method using mass curve Analytical method Energy conversion Water turbines Turbine power output Pelton turbine Francis reaction turbine Propeller and Kaplan turbines Power From Dams (Potential Energy) Social and Political Water Power Case Study Future Directions Advantages and Disadvantages of Hydroelectric Power Closing Remarks

545 546 548 549 550 552 552 553 554 555 556 556 556 556 557 557 557 558 558 559 560 560 560 560 560 560 561 561 561 561 562 562 563 563 564 564 564 565 565 567 568 569 570 570

Introduction

Energy is a major development measure in any society with two different categories as nonrenewable, fossil (conventional) sources, such as coal, fuel, natural gas; on the other hand, renewable (unconventional) sources as solar, wind, hydro, geothermal, wave, and tide. After the industrial revolution as primary energy sources coal and oil are used for various human activities. They are available in limited quantities as deposits in the subsurface, but their use contributes to greenhouse gases emission giving rise to the global warming and climate change impacts. Carbon dioxide (CO2) accumulation in the troposphere causes to climate change and to the intensive rainfall, flood, and drought occurrences. Reduction in such dangerous effects, countries must try and to

Comprehensive Energy Systems, Volume 4

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Hydropower Conversion

improve energy resource qualities by replacing the fossil fuels as much as possible with renewable alternatives of hydropower, wind, solar, and other energy sources. Fossil fuel exhaustion is expected within the next 50–150 years, but although alternative energy resources are intermittent but clean, sustainable, and friendly sources with the environment. In the 21st century, water is expected to have additional stresses in supporting essential commodity for running industries, growing energy demand, and food demands. Recently a number of governmental organizations and especially private sectors have direct and indirect interests in the development and conservation of water resources for hydroelectric power (HEP) generation augmentation projects by means of different size dam constructions and small run-of-water turbines that are of vital importance for any country from renewable energy prospect points of view. Although completion of HEP projects cannot cover total energy demand of a country, but they support the existing energy demand with additional national alternatives. Additionally, operation, management and maintenance of these alternatives collectively need an integrated automatic program for the optimum distribution of energy sources to consumption centers. The social and economic development of any country gear the energy and electricity steady growths due to industrial and urban developments. Revitalization of the regional economies, speeding up the land developments, agricultural expansions, and improvement of the irrigation and energy efficiency are all dependent on the water resources development for the purpose of HEP generation, water supply, and irrigation. Natural hydrological cycle is the sole source of hydropower (HP) as a renewable energy alternative in different parts of the world. It is the most mature, reliable, and cost-effective renewable power generation technology available [1]. HP plant designs have significant flexibility for peak demand supply with high and low capacity factors. For the time being, HP is the largest renewable energy source with about 20% share in the world’s electricity generation and over four-fifths of the world’s renewable electricity. More than 25 countries depend on HP for more than 90% electricity supply (99.3% in Norway), and 12 countries are 100% reliant on hydropower, which produces the bulk of electricity in 65 countries and plays some role in more than 150 countries. Canada, China, and the United States are the countries which have the largest HP generation capacity [2–4]. Flexibility of HP to generate power comes from the accumulation of surface water behind a dam in the reservoir ready for use at any minute of time. None of the other energy generation plants is capable to respond to extra energy demand as quickly as the HP plants, especially during peak times. Reservoir and pumped storage HP can be used to reduce the frequency of start-ups and shutdowns of conventional thermal plants and maintain a balance between supply and demand, thereby reducing the loadfollowing burden of thermal plants [5]. HP system integration capabilities are useful particularly for allowing the large-scale large penetration of wind and other variable power sources [6]. Systems with significant shares of large-scale hydro with significant reservoir storage are able to integrate higher levels of variable renewables at low cost than systems without the benefit of HP [6]. HP is a historical renewable source and its present exploitation is enhanced through the use of new technologies [7] and minimization of environmental impact on rivers [8]. New HP plant constructions require detailed feasibility studies and economic evaluations [9,10]. As Priscoli [11] stated that “just as water is life, and we are predominantly water, water infrastructure is the precondition for civilization.” Abu-l Iz Al-Jazari, who lived during the 12th century, is the father of robotics that worked by water power hydraulically [12]. He has reviewed previous technology developers, such as Vitruvius and Heron who have lived during the first few centuries after the prophet Christ in the Roman domain. They have not left proper designs or procedures for the few technological ideas of their origin, but Abou-l Iz Al-Jazari has drawn many mechanically proper designs in his hand written book, which is printed in its original form by the Cultural Ministry of Turkey in 1990. His book has been translated into English with the modern drawings corresponding to the original ones [13]. A glance through his book brings to one’s mind the question of who was the first man in the human history in visualization and drawing mechanically drivable water power devices closest to today’s technological level. It is important to emphasize at this state that medieval period was an enlightenment time for Islamic countries, whereas Europa was living in dark ages as a riddle. There were many full-scale machines that were described by Al-Jazari in order to raise water and all of them incorporated features that are of great significance in the history of machine technology and even today some of these parts are in common use in any machine design. The first of this is presented in Fig. 1 where water power is used to raise the left and right hands of a robotic man on an elephant. Water injection from the right nozzle hits the left of the scoop at the back of the elephant and hence, the left hand of the robotic man is raised. This is the simplest first design toward robotic studies in the history of technology. Fig. 1 seems to be of little importance, being simply a chain of pots driven by a concealed scoop wheel to raise water and then fall on the present day Pelton type turbine planned at around AD 1200. As shown in Fig. 2, Al-Jazari provided the first example in the history of a crank hand, cylinder, and piston combination to raise water. This machine had a significant place in the development of the steam engine and pumping machinery [13]. It is the main purpose of this chapter to present the historical evolution of water power and also different technological devices to generate HP energy. Detailed information is reflected from available literature concerning various HP generation mechanisms, methodology, and technology.

4.12.2

Focus on Hydropower

The main energy policy is to obtain cheap and reliable electrical energy for consumers. It is essential to balance demand and supply sides without any problem by taking into consideration effective development and management aspects under the light of strategic

Hydropower Conversion

547

Fig. 1 Chain of pots for water rising.

Turbine Fig. 2 Wind or water powered flume-beam sweep.

analysis. Electricity generators can make independent decisions instead of a centralized one. The profit maximization should be cared for according to the prediction trends. The following points are important for energy plans: 1. 2. 3. 4. 5.

For sustainable socio-economic development cheap, clean, and good quality of energy. Security guarantee in energy supply. Privatization encouragements in energy sector. Integration of renewable energy alternatives into the overall energy supply system. It is not possible that one country cannot satisfy all of the energy demand from renewable alternatives, but addition of private HEP development support the overall energy system in a positive manner. 6. In case of any near future energy deficit possibility, HEP generation augmentation must be considered in addition to the renewable energy alternatives development; HEP has top priority in many countries. The following points are important for

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Hydropower Conversion

HEP favorite [14], Hydropower is environmentally friendly and renewable source of energy with insignificant gas emissions. Energy security of a country can be supported by construction of additional hydropower plants, if possible. Hence, dependency of a country on foreign oil import will decrease significantly. • Adaptive and simultaneous HEP supply provides flexibility in the energy demand fluctuations, which is not possible by thermal plants. • Dams for HEP production also help to regulate river flows, and hence, reduce the dangers of flood and drought in the region. • Especially at the times of peak energy demand HEP plants provide the cheapest way of supply. • HEP plant initial investment is rather high, but operation, management, and maintenance costs are comparatively very low. Almost 80% of the investment costs are for construction and the remaining 20% are for the electro-mechanical costs.

• •

HEP plants are preferable due to their environment-friendly behaviors with low risk. They are helpful for sudden respond to unexpected demand increases. “HEP is environment friendly, clean, renewable, able to meet peak demands, highly efficient (over 90%), involves no fuel cost, is a balancer of energy prices, has a long life-span (200 years), its cost recovery is short-run (5–10 years), its operational costs are low, (approximately 0.2 US cent/kWh), and it is an indigenous source of energy which is national and natural” [15].

4.12.3

Hydropower Generation Fundamentals

The hydroelectricity extraction from water depends not only on the water volume (discharge, which is the volume of water per time duration), but also on the difference in height between the water head elevations (the height difference) at high and the water outflow, which is generally the elevation of the turbines. The amount of water potential energy is directly proportional to the head and also the discharge of water led to the turbines either by a penstock or pipes. It is advantageous to build water power generation dams as high as possible to benefit from the maximum energy yield at a dam location. The energy derivation for power by the force of water moving from a higher elevation to a lower elevation is possible through a large “tube,” which is known in technical terms as a “penstock.” At the end of the penstock water power turns a “wheel” or “turbine” at enormous speeds. The turbine rotates by means of a connected shaft to an electric generator, and this generator generates electricity. It is the turbine and generator working in combination that converts “mechanical energy” into “electric energy.” The water that makes this possible is a renewable energy resource, just like the wind that turns the turbine attached to a generator. HEP is relatively clean and safe in comparison to fossil fuels born energy types (coal or oil or natural gas). Most of the large HEP plants have significant impacts on nearby environmental habitats, because they significantly impede the flow of water in rivers and reservoirs behind the dams. This causes to significant increases in water levels in the dam upstream water systems, but comparatively much lower water levels downstream. HEP is due to flowing water from mountain streams and clear lakes during winter and spring. Water that falls by the force of gravity is used to turn turbines and generators for electricity production. Populations increase, economic development, and modern technologies require vast amounts electricity generation for societal activities. In the 1920s, HEP plants supplied as much as 40% of the electric energy but the amount of energy produced in this way has increased steadily. Hydropower is an essential contributor in national power grid, because of its quick response to meet the rapidly varying demands or system disturbances. Hydropower does not pollute the water or the air, but its facilities can have large environmental impacts by changing the local environment affecting land use, homes, and natural habitats in the dam area. It is well known that most HEP plants are at the downstream of a dam and a reservoir, which may obstruct fish migration and affect their population growth. Operation of a HEP plant may also change the water temperature and the river discharge rate. Consequently, these may harm native plants and animals in the nearby environment. Reservoirs may cause people to leave their settlements and homes by relocation, important agricultural land, and archeological sites. Methane (CH4) is one of the strong greenhouse gases, which may form in some reservoirs and be emitted into the atmosphere (EPA energy kids). HEP is form of renewable energy apart from other renewable resources (geothermal, wave power, tidal power, wind power, and solar power). These power plants generate electricity depending on the hydrological cycle activities after each rainfall and snowmelt process, but they do not pollute the air, land, or water as other power plants. Small and large HEP developments were instrumental in the early expansion of the electric power industry. HEP is a result of flowing water from mountain streams and clear lakes during winter and spring. Water falls by the force of gravity is used to turn turbines and generators for electricity production. Hydropower has different forms as potential energy from high heads of water stored in dams, kinetic energy in rivers and tidal barrages, and kinetic energy also from the movement of waves and also from currents, such as Bosporus of Istanbul, in Turkey on relatively static water masses [15]. Hydropower can be harnessed by many indigenous ways, but directly from the water that flows through a turbine to generate electricity. Another possible conflict between adaptation and mitigation might arise over water resources. One obvious mitigation option is to shift to energy sources with low greenhouse gas emissions such as small HEP plants. In regions, where HEP potentials are still available, and also depending on the current and future water balance, this would increase the competition for water, especially if irrigation might be a feasible strategy to cope with climate change impacts in agriculture and the demand for cooling water by the

Hydropower Conversion

549

power sector. This reconfirms the importance of integrated land and water management strategies to ensure the optimal allocation of scarce natural resources (land, water) and economic investments in climate change adaptation and mitigation and in fostering sustainable development. Hydropower leads to the key area of mitigation, energy sources and supply, and energy use in various economic sectors beyond land use, agriculture, and forestry.

4.12.3.1

Hydrological Cycle and Hydropower

Water is in a continuous movement depending on the local hydrological cycle supported by the large global cycles. The initial source of the running water starts as evaporation from lakes and oceans, which forms clouds and they generate rainfall or snow depending on the elevation difference (orographic type of precipitation), temperature difference (convective type of precipitation), or pressure difference (frontal type of precipitation). The precipitation is converted to surface runoff, which runs from high elevations of a drainage basin toward the lakes and oceans. During the surface channel confined flow of the runoff, the way is cut by damming so as to convert the kinetic energy of water to potential energy and then by means of technological arrangements finally electricity is generated. Fig. 3 indicates hydrological cycle and also its energy cycle correspondence. The sole cause of the cycles is the sun solar irradiation. The running water energy can be tapped to produce electricity or mechanical energy that is useful for grain grinding. In a drainage basin depending on the hydrological cycle component of runoff fluctuation run-of-river type of HP to generate electricity without water storage. The physical power of water arises along the hydrological cycle through which potential, kinetic, and heat energy forms are converted from one to each other. Fig. 1 shows energy powered hydrologic cycle with the inexhaustible energy source as sun [16]. Solar radiation from the sun activates the free surface water for evaporation and in this manner the kinetic energy within the upper surface layer of the water leads to entrance of water molecules into the lower atmosphere as latent heat energy embedded particles. The rainfall has maximum power as it reaches earth’s surface, and consequently, it might cause soil erosion and consequent sediment transport. On the other hand, water at high elevations has comparatively higher potential energy than lower locations, and subsequently, this potential energy is converted to kinetic energy in the rivers. It is this kinetic energy as water power turns water mills and haulage of water at a point from low to high elevations by water wheel through the centuries. Construction of big dams at any suitable cross-section along the river conserves this kinetic energy into potential energy in the reservoir accumulation. Hence, in any dam reservoir water power is ready for electric energy generation. The energy components in a hydrological cycle are shown in Fig. 4. Since the vertical temperature distribution has a decreasing trend upwards and according to thermo-dynamical laws, heat flows from warm regions to cooler areas, the particles that escape from the free water surface raises upward into the atmosphere. Such a rise cools down the water vapor and after some height the water vapor appears in the form of cloud, where water power is in the form of potential energy although within the cloud there are partial kinetic energy movements from the bottom of the cloud toward the ceiling. The clouds are transported toward inlands by horizontal air movements in the form of winds, and due to further rise along the mountain sides, or cyclonic occurrences further cooling in the cloud renders the water drops to form, and hence, they become heavier than the gravitational force and start to fall downwards in the form of precipitation. The potential impact of climate change on the hydrologic regime is a crucial question for water resources management. Potential change in hydrologic regime resulting from climate change is an important topic in contemporary hydrology and water resources management. Increasing trend in the CO2 concentration may lead to unwanted climatic disruptions and local imbalances in the hydrological as well as atmospheric cycles will be the consequences, which may lead to excessive rainfall or drought, in addition to excessive heat and cold. The environmental and social problems can be alleviated by effective energy policies with consideration of renewable energy sources, such as hydropower, solar, wind, biomass, wave, and geothermal energies as well as the solar hydrogen energy possibilities (Section 4.12.4). Wind

Rainfall

Cloud

no

ff

Cooling

lar

So

Interception

Ru

Sun

irra

Evapotranspiration

on

ti dia

Evaportation

Depression

oir eserv

Spring

G

r water round

Water body

Fig. 3 Hydrological cycle [16].

Impervious layer

550

Hydropower Conversion

Sun Cloud Wind energy Potential energy Precipitation

Kinetic energy

Solar energy

Kinetic energy runoff

Evaporation

Potential energy Dam

Hydroelectric energy

Sea

Fig. 4 Natural energy cycle [17].

HEP is one of the renewable energy than is well established technologically for clean energy generation. In the surface water (runoff) rich countries, the full-scale developments of HEP generation by turbines at large-scale dams are active in addition to smaller hydropower systems such as run-of-the-river plants. The world’s total annual rainfall is, on the average, 108.4  1012 l/year of which 12  1012 t recharges the groundwater resources in the aquifers, 25.13  1012 t appears as surface runoff and 71.27  1012 t evaporates into atmosphere. If the above rainfall amount falls from a height of 1000 m above the earth surface, then the kinetic energy of 1.062  1015 kJ/year is imparted to the earth every year. Some of this huge energy is stored in dams, which confine the potential energy so that one can utilize it to generate HEP [15]. Hydropower generation is likely to be impacted because it is sensitive to the amount, timing, and geographical pattern of precipitation as well as temperature (rain or snow, timing of melting) [18]. Reduced stream flows are expected to jeopardize hydropower production in some areas, whereas greater stream flows, depending on their timing, might be beneficial [19,20]. Climate variability and long-term climate change should be considered in siting wind power facilities [21,22]. As a result of climate change by the 2070s, hydropower potential for the whole of Europe is expected to decline by 6%, translated into a 20–50% decrease around the Mediterranean, a 15–30% increase in northern and eastern Europe and a stable HP pattern for western and central Europe [23].

4.12.3.2

Hydropower and Climate Change

Among the renewable energy sources HP does not emit the CO2 and other greenhouse gases (SO2, NOx, and NH4) that affect the atmospheric environment. Additionally, HP has the most energy payback ratio when compared among all electricity generation systems [6]. For instance, 1 GWh of electricity produced by small HP corresponds to a reduction of CO2 emissions by 480 t [24]. Climate is defined as the average of weather events over long time spans and large regions. It is caused by changes in the wind direction and speed, ocean currents, atmospheric circulation, sunshine radiation, precipitation, pressure, and temperature. In local climatic signatures especially temperature and the precipitation records and variations provide basic initial information for hydropower generation potential. Future water resources under different modern stresses (population growth, land use, urbanization, and consumptive life style) are expected to play more effective and dominant role in any societal activity, and therefore, their planning, design, operation, and maintenance should be based on future scenarios intermixed with factual information and knowledges. One of the definite nuisance in the future is the climate change due to anthropogenic effects [25]. The same panel has provided several scenarios for future direction of the world climate based on environmental, agricultural, energy, land use, economy, vegetation, and consumption rates all of which affect atmospheric composition with consequent reflection in the climate and as a chain interaction in the hydrological cycle, and hence, water resources. Accordingly, any country for the water safety should assess available resources under climate change impacts on different future time horizons up to 2100. It is a fact that the longer the future scenario horizon the less will be reliability of the results. Engineering structures help to manage the water resources utilization according to demand and supply side requirements in the best possible (optimum) manner. Different alternatives are developed and applied in water sector over many years but they do not take into account the climate change affects explicitly. However, some countries water resources managers started to care for the

Hydropower Conversion

551

climate change effects, which are rather significant in the coming decades especially in the mid-latitudes and some subtropical climate belt of the world. If reservoirs are full after a wet period then a short-lived summer flood may not end a water resources drought caused by prolonged lack of reservoir inflows. Hence, droughts are not dependent on possible climate changes only but critically on the water resources system characteristics and especially on their management [26]. In-stream and off-stream consumptive and nonconsumptive exploitations of water resources are expected to be affected in the long-run. Anthropogenically induced gas emissions such as CO2 that trap heat in the atmosphere have contributed to an increase in global mean surface air temperatures of about 0.3–0.61C [27]. Moreover, based on the IPCC’s mid-range scenario of future greenhouse gas emissions and aerosols and their best estimate of climate sensitivity, a further increase of 21C is expected by the year 2100. Uncertainties are enormous as to how the climate and hydrology of a region will change in response to a global greenhouse warming. However, one of the more likely impacts involves areas where precipitation currently comes largely in the form of winter snowfall, and where streamflow comes largely from spring and summer snowmelt. A warming would likely result in a distinct shift in the relative amounts of snow and rainfall and in the timing of snowmelt and runoff. A shift from snow to rainfall could increase the likelihood of flooding early in the year and reduce the availability of water during periods of peak demand, especially for irrigation and HEP generation. Meteorological events and climate phenomena are related to energy sources (especially HEP) and the atmospheric composition drives them. The use of fossil fuel led to atmospheric chemical composition change with additional greenhouse gases especially CO2 and CH4 causing increase in global average temperature [28,29]. The CO2 increase is due to fossil fuels usage, which in turn related to human activity, population rise, and consumption increase. Increase in the greenhouse gases emission rates especially CO2 is a treat to the world climate. Atmospheric research and development studies have led to the conclusion that an increasing trend appeared in the average temperature tendency up to 0.51C. Due to this increase, some areas of the world are expected to have extreme (low and high) rainfall events leading to floods, droughts and local imbalance in the natural climatic behaviors all of which reflects in the agricultural and HEP generation activities. It is predicted by IPCC that by the end of this century the global temperature may rise up to 1–3.51C, which may lead further global climate change [25]. However, the use of renewable energies including hydropower may help to manage the CO2 emissions and it may possible to limit further climate change effects to adaptable levels. From demand-side impact point of view, energy production is likely to be affected by climate change and this is especially valid for HEP generation. Apart from extreme weather event impacts energy consumption is very significant to manage. However, the climate change could affect HEP production and supply as a result of the following points [18]: 1. 2. 3. 4.

Extreme weather events become more intense. Regions dependent on water supplies for HEP plant and/or thermal power plant cooling face reductions in water supplies. Climate change conditions affect HEP plant siting decisions. Conditions change (positively or negatively) for hydropower, biomass, wind power, or solar energy productions.

Climate change is likely to affect both energy consumption and energy production in many parts of the world and where the climate warms; less heating is needed for different human activities. For cooling the main energy source is electricity, while coal, oil, gas, biomass, and hydro energy are used for space heating. Regions with substantial requirements for both cooling and heating could find that net annual electricity demands increase while demands for other heating energy sources decline [30]. Due to infrastructure limitations, peak energy demand could go beyond the maximum capacity of the transmission systems. Energy and climate are related concerning cooling during hot weather. The energy sector can adapt to climate change vulnerabilities and impacts by taking steps to increase the resilience, for example, by diversifying energy supply sources, expanding its linkages with other regions, and investing in technological change to further expand its portfolio of options [22]. Climate change could have a negative impact on the HEP generation on one hand, and on the other, on the thermal power production since the availability of water, and especially, cooling water may be reduced at some locations because of climaterelated decreases [31] or seasonal shifts in river runoff [32]. The distribution of HEP is also vulnerable to climate change. It is well known by now that the economic growth and the population increase are the major factors that lead to increase in the energy demand. Table 1 presents possible future energy demand up to 2020 [33]. In this table, Moet means million oil equivalent ton as energy unit. The population has increased in the last century by a factor of 6 and against the energy consumption factor of 80. The average power consumption by one capita is equal to 2 kW, but in the United States it is about 10 kW and in Europe about 5 kW. Almost two billion people do not consume any fossil fuels at all. The world’s population growth rate is at 1.3–2%, level, but it is expected to double within the next 60 years. Presently population is about 6.5  109 and growing toward 12  109 in 2060 [34]. Table 1

Future energy demand predictions

1000 million oil equivalent ton

1990

2020

Increase (%)

Industrialized countries Central and eastern Europe Developing countries World

4.1 1.7 2.9 8.7

4.6 1.8 6.9 13.3

12 5 137 52

552

Hydropower Conversion

In the past, prior to the discovery of the fossil fuels, hydropower played the most significant role in a society as for the transportation as the river and ocean navigations; early trains ran on the steam power, which was the combination of coal and water vapor. The society is affected by climate, and hence, energy in one of the three major factors [35]. 1. Economic sectors that support the settlement are affected, because of changes in productive capacity or in market demand for goods and services produced there (energy demand). The importance of this impact depends in part on whether the settlement is rural (which generally means that it is dependent on one or two resource-based industries with very less energy consumption) or urban, in which case there usually is a broader array of alternative resources including energy resources consumption centers, 2. Some aspects of physical infrastructure (including energy transmission and distribution systems), buildings, urban services (including transportation systems), and specific industries (such as agro-industry and construction) may be directly affected. For example, buildings and infrastructure in deltaic areas may be affected by coastal and river flooding; urban energy demand may increase or decrease as a result of changed balances in space heating and cooling (additional energy consumption). Coastal and mountain tourism may be affected by changed seasonal temperature and precipitation patterns and sea level rise. Concentration of population and infrastructure in urban areas can mean higher numbers of persons and higher value of physical capital at risk, although there also are many economies of scale and proximity in ensuring well-managed infrastructure and service provision. 3. As a result of climate change society may be affected directly through extreme weather conditions leading to changes in health status and migration. Population movements caused by climate changes may affect the size and characteristics of settlement populations, which in turn changes the demand for urban services (including energy demand). The problems are somewhat different in the largest population centers (e.g., those of more than 1 million population) and mid-sized to small-sized regional centers. The former are more likely to be destinations for migrants from rural areas and smaller settlements and cross-border areas, but larger settlements generally have much greater command over national resource. The largest amount of construction work to counter balance climate change impacts will be in water management and in coastal zones. The former involves hard measures in flood protection (dykes, dams, flood control reservoirs) and in coping with seasonal variations (storage reservoirs and interbasin diversions), while the latter comprises coastal defense systems (embankment, dams, storm surge barriers). Adaptation to changing hydrological regimes and water availability will also require continuous additional energy input. In water-scarce regions, the increasing reuse of wastewater and the associated treatment, deep-well pumping, and especially largescale desalination, would increase energy use in the water sector [35].

4.12.3.3

Hydropower Sustainability

Generally, in any human settlement activity sustainability is thought to last for long-range. Hydropower projects may have beneficial and to a certain extent negative impacts on the society or nearby environment, because there may happen a series of physical alterations in the natural environment. For these reasons, the sustainability procedure must include hazard reduction policies as much as possible. It is within this context that prior to any hydropower construction in a region, Environmental and Social Impact Assessment reports must be prepared by a group of specialists. This is very essential for understanding pre-project conditions, potential impact estimations and for preparation public awareness reports for vulnerability, combat and mitigation activities, if necessary. Additionally, sustainability assessments must be completed prior to final decision on whether to start construction. After these steps, it is possible to guarantee promotion and guidance for sustainability. In the preparation reports not only engineering and environmental aspects, but other regional social, economic, public health, and physiological conditions must also be cared for. Dam effected individuals and areas must be compensated at equal share bases such that construction of a hydropower benefits must be many folds more than undesirable impacts. Social and environmental risks must be avoided for the safety of the region. For instance, the consequences of, say, a dam break must be taken into consideration. The prime importance is to safeguard policies that must be developed continuously such that after the identification of unwanted cases their avoidance or minimization of harmful cases as much as possible.

4.12.4

Hydropower Systems

A methodology is proposed to investigate potential sites of small HP projects in India by using remote sensing data which is used for extraction and mapping of water resources and its associates, such as inhabitation and settlement pattern, forest and vegetation coverage, snow coverage and selection of probable sites for small HP projects [36]. On the other hand, Kaldellis [37] investigated the existing and the proposed small HP plants in Greece by considering technical and economic factors. They found that the main obstacle to construction of a project was decision-making problems, which include the administrative bureaucracy, the absence of a rational national water resources management plan and the over-sizing of the proposed installations. The procedure of spatial planning is presented by integrating social, economic, and environmental factors in countries with a complicated administrative and legislative system [38]. A Geographic Information System (GIS) application is given to assess small HP resources in a region of India by collecting the existing data of distribution and capacity of small HP [39]. The GIS use for mapping of potential source of

Hydropower Conversion

553

small HP is proposed by calculating of topographic drop and the annual mean flow [40]. They did not consider economic and social aspects in their study. Site selection with the aid of GIS technology is a widely used procedure in a variety of fields. The role of GIS in spatial decision-making is to aid the decision-maker in designating priority weights to the criteria, to evaluate the feasible alternatives and to visualize the results of the choice. Multi-criteria analysis and integrated spatial information for decision-making are employed in the assessment and selection process of areas suitable for landfill implantation [41]. The GIS mapping is suggested to select the optimal sites for small dam installation [42]. They used both qualitative and quantitative criteria which were based on satellite, hydrological, and climatological data. Generation of HP needs a system to convert the water power to directly useable energy through a sequence of conversion processes. Generally, four different types of HP systems depend on the impoundment, diversion, pumped storage, and other water power technologies, such as tide, wave, heat, and current flows. Some HP plant systems have dams and some others without storage facility. The most frequently used HP plant systems have impoundment facility mostly behind a dam to store drainage basin river flow in a reservoir. Water is led through a pipe system to a turbine to spin it, and finally it activates a generator for electricity production. A diversion is another system, which does not have any impoundment facility and it is referred to as the run-of-river plant. It diverts a portion of river through a canal in which turbines are located for electricity production. Pumped storage is another type of HP system and it works like a battery by storing the electricity generated by other power sources like solar, wind, and nuclear for later use. It converts these power sources by pumping water uphill to a reservoir at higher elevation from a second reservoir at a lower elevation. At times of low energy demand this system stores energy by pumping water from a lower reservoir to an upper reservoir. Hence, during high electrical demand periods, the water is released back to the lower reservoir and turns a turbine for electricity generation. The fourth type of HP generation systems include tidal, wave, heat, and current flows for electricity generation through various technological system designs. New HP plant constructions require detailed feasibility studies and economic evaluations [9,10]. It is necessary to assess carefully the drainage basin river network potential as a function of the hydrological and morphological features [43,44]. There is particular interest in small HP plants with installed capacities from medium to low [45,46]. The key information about the construction of a HP plant is the flow duration curve (FDC), which facilitates evaluation of the HP potential of the site and the environmental constraints for plant management [47]. In the meantime it is necessary to identify the best water management practice and to evaluate water quality [48,49], for low flow estimation [50,51], flow data quality control [52], catchment response to afforestation [53] and HP generation [54–57]. Several solutions and approaches are proposed for the multipurpose water utilization [58] for search to identify the best strategy for the exploitation of water resources [59].

4.12.4.1

Reservoir Impoundments

Reservoir is generally defined as a hollow space where water can be stored for use at the times of need at a later stage. Most often, it refers to artificial lake (behind dams) for water impoundment, which can be used for different purposes and especially for HEP generation, water supply, irrigation, and agriculture. In general, reservoirs are next to a dam construction on a stream, which allows the stream to flow thereby filling it to capacity (Fig. 5). The reservoirs may be single-purpose type, such as for HEP generation, flood control, groundwater recharge, water supply, and agriculture. On the other hand, in a multiple-purpose reservoir the whole volume has different zones (sub-volumes) and each one is planned for a certain task (Fig. 6). Each reservoir sub-volume zone is explained below starting from the upper element toward the bottom. 1. Spill zone is the storage space above the flood control zone between the full reservoir level, which corresponds to the spillway level and it is occupied mostly during high floods. 2. Flood control zone is the storage space earmarked as temporary storage for absorbing high flows to avoid the downstream damages.

Evaporation Inflow

Runoff Outflow

Groundwater inflow Groundwater outflow Fig. 5 Reservoir after damming streams.

554

Hydropower Conversion

Maximum water level Spill zone Full reservoir level Flood zone control

Conservation zone DAM Minimum drawdown level

Buffer zone Dead storage level Dead storage zone

Fig. 6 Reservoir zones.

3. Conservation zone is the storage zone between the full reservoir level and the dead storage level and serves various site and downstream water uses, including HEP generation, irrigation, municipal and industrial water supply, navigation, water quality, fish and wildlife, and recreation. 4. Buffer zone is the storage space above the dead storage level and it is used to satisfy very essential water needs during extreme drought situations. 5. Dead storage zone is the lowest zone and it is kept full at all times to provide minimum head for HEP generation, sedimentation storage space, and other uses. These storage zones are applicable only in reservoirs, which have storage round the year. For dams in arid regions where rainfall is very scanty and may not occur every year, the reservoirs may remain dry for most of the year. There are always some losses in the quantity of water stored in a reservoir due to evaporation, percolation, and seepage. The reservoir basin should be carefully checked for water-tightness and percolation in order to limit the possible losses to a small or negligible quantity. Evaporation losses depend directly on the reservoir surface area and usually expressed in millimetre of depth per unit time. The three major factors that are effective on the evaporation losses are the temperature, wind velocity, and relative humidity. It is necessary to account for such losses in any study regarding the reservoir operations.

4.12.4.2

Run-of-River

“Run-of-river” plants are convenient for small schemes of generation less than 10 MW (megawatt) output. A fast flowing runoff is diverted through a turbine drives the electrical generator. There is no water head in these systems, but the turbine converts the kinetic energy of the runoff into the rotational energy of the turbine and the generator. In this case, the available energy depends on the discharge that passes through the turbine and the square of water velocity (Fig. 7). Impulse turbines that have partial submergence only are frequently used in fast flowing run-of-river installations. On the contrary, in deeper but slower runoff cases completely submerged Kaplan turbines are more helpful to extract the energy from the runoff. Run-of-river plants are cheaper than dams, because of the simpler civil works, requirements, but they are susceptible to rainfall and consequent runoff variations as a result of which there are occasional reductions or even cut off potential power output during dry spells. However, in flood conditions the installation may not be able to accommodate the higher flow rates and water must be diverted around the turbine losing the potential generating capacity of the increased water flow. If the construction of a dam is not possible, run-of-river installations may need to incorporate some form of supply back-up, such as battery storage, emergency generators, or even a grid connection. These are run-of-the-river hydroelectric stations with no or small reservoir storage capacity, which uses the running water at the point instantaneously for power generation and then water, runs at the downstream without any loss. Sites suitable for this type of energy generation have usually water supply from a lake or upstream reservoir. In the site selection of the HP project, a conventional method using manual operation is quite time consuming. Also, in some cases, a field survey must be conducted, and it requires significant manpower due to the undulating topography, dense forest cover and bad climatic conditions at the site. The conventional method of site selection focuses on engineering and economic criteria, disregarding environmental criteria, and social impact. In order to reduce cost and time consumption for the site selection, and to integrate all criteria (engineering, economic, environmental, social impact) into the decision-making process, this study proposed a new method using GIS application with the consideration of engineering, economic, and environmental criteria, and social impact. This proposed method comprises the following steps [43]. 1. Engineering analysis by discharge analysis and GIS application to locate project sites. 2. Economic analysis by GIS application to evaluate the economic potential of each project.

Hydropower Conversion

Generator

Headwater level

555

Tailwater level Gross head

Turbine

Kaplan turbine with vertical shaft Fig. 7 Run-of-river plants [43].

Upper reservoir

Transmission connection Upper water conductor Lower water conductor

Powerhouse Lower reservoir Fig. 8 Pumped storage system [64].

3. Ranking of the selected sites by total weighted scores of environmental parameters. 4. Social impact study through a public participation process, including interview questionnaire survey and focus group discussion. It is stated by Rojanamon et al. [43] that a small-scale run-of-river project is a kind of HP project that generates electricity according to the available hydrological fluctuations of the site. Although, there is still no internationally agreed definition of small hydro, but a maximum of 10 MW is the most widely accepted value worldwide [60]. HP projects are ease of smaller investments, shorter period for planning and construction, use of a smaller area, use of local labor and material, and cheaper generation cost as compared to other power projects [61,62]. Their power generation cost is higher than larger sized plants due to economy of scale of some instrumentation, control, and monitoring systems [63].

4.12.4.3

Pumped Storage

It is one of the most established technologies that it has been in use since 1890 [64,65]. As shown in Fig. 8 it consists of two interconnected reservoirs, through tunnels that convey water from one reservoir to another, a powerhouse with a pump and turbine, a generator, and a transmission line connection. Depending on the meteorological, hydrologic, and geological set up of the location, there are various types of pumped storage systems. It is also possible to benefit from the existence of natural lakes, large rivers, or reservoirs of conventional hydro facilities as their reservoirs.

556

Hydropower Conversion

Few methods are proposed for the optimization of large-scale HP system operations [66]. An economic analysis of the inclusion of pumped storage is presented in a small island system that has abundant renewable energy available, but that at times cannot accept all of this power, because of limits imposed by security criteria [65]. The question of whether or how much pumped storage to include is addressed by formulating a linear programming optimization problem. The stochastic nature of load and renewable production is addressed using scenarios developed through fuzzy clustering [67]. Both the unit capacity in megawatt and the reservoir storage capacity in megawatt hour are optimized, and optimal operating strategies for the scenarios are produced.

4.12.4.4

Other Alternatives

Small-scale HP generation is also possible from different natural water storages, especially oceans, through a variety of technological devices. They are useful for local electricity demand satisfactions.

4.12.4.4.1

Tidal power

Tides also have power and a tidal power station is based on the daily fluctuations of water level in an estuary and predictability of energy generation is very high and it is also possible to generate power during high demand periods provided that there are convenient reservoirs. The most common types are without reservoir and generate power from water kinetic energy or undammed sources such as undershot waterwheels. These types of power stations are viable in a relatively small number of locations around the world [68]. It is also possible to harness hydropower from the tides by placing bidirectional turbines in the path of the tidal water flow in bays, and especially, in river estuaries. It is necessary to have a large tidal range with a barrier across the bay or estuary to funnel the water through the turbines as the tide comes in and out (Fig. 9). The tidal energy capture in tidal ponds has been in use since Roman times to power mills, but there are few modern installations. In modern times, the first tidal power plant on a large-scale for electricity generation was built at Rance in France in 1966. Tidal power provides unlimited, continuous and predictable power output, but unfortunately there are few suitable sites in the world with environmental constraints, where tidal power generation plants can be constructed. Deep located turbines work through tidal currents that show better potential for exploitation. In tidal power plants, the generation is possible only for 6–12 h/day depending on the ebb and flow of the tides. Scientists and engineers still continue to search for different energy generation alternatives from wave power, which is a carbon and emissions free method.

4.12.4.4.2

Wave power

Various methods have been employed to turn the wave motion into electrical energy. Surface wave motion is limited and very difficult to capture, many ingenious systems have been proposed for power generation. Very small installations generate electricity commercially and most have been hindered by practical problems. There are various versions of the wave power and some of them are outlined below. Most of them are still in experimental phases and many are not scalable into high capacity systems [69–71]. 4.12.4.4.2.1 Oscillating float system This is one of the simplest and most common energy conversion mechanisms depending on the oscillating float system with a float housed inside a cylinder shaped buoy, which is open at the bottom and moored to the seabed. Inside the cylinder the float moves up and down on the surface of the waves as they pass through the buoy (Fig. 10) [72,73]. Among the hydraulic systems comes the air compression in a pneumatic reservoir above the float during its upward movement on the crests of the waves. After the crests passage, the air expands and forces the float downwards into the following troughs of the waves. Additionally, a hydraulic system uses the reciprocal float movement to pump water through a water turbine, which drives a rotary for electrical generation. Pneumatic systems displace the air in a cylinder, which is used to power an air turbine for the generator function. There are also linear generators that turn the reciprocating motion of the float directly into electrical power. Another alternative is that Electric power from tidal flows Sea level Generator Tidal flows

Turbine

Concrete base

Fig. 9 Tidal flow plant mechanisms [68].

Sea water

Hydropower Conversion

Floating object

Elliptical trajectory

557

Wave propagation

Ripples Fig. 10 Oscillating wave power systems [72].

Axis horizontal offset, Xa

Paddle length, L

Axis vertical offset, Ya

Clearance, Cl

Angle of backplane, 

Fig. 11 Schematic diagram of an oscillating wave surge converter (OWSC) [74].

instead of electricity generation on board of the buoy, some systems pump the hydraulic fluid ashore to power shore based generators. 4.12.4.4.2.2 Oscillating paddle system They have large paddles moored to the ocean floor to mimic the swaying motion of sea plants in the presence of ocean waves, where the paddles are fixed to special hinged joints at the base, which use the swaying motion of the paddles to pump water through a turbine generator [74]. In its most primitive form the oscillating wave surge converter (OWSC) consists of a paddle suspended from a hinge located above the water surface so that it can rotate about an axis approximately parallel to the wave crests, together with an angled backwall behind this paddle (Fig. 11). Enhancements to this basic concept of the OWS could include the paddle oscillating about a non-vertical position and more complex profiles of the back-wall. 4.12.4.4.2.3 Oscillating snake system It uses a series of floating cylindrical sections linked by hinged joints and floating snake is tethered to the seabed maintaining a position head on into the waves. The wave-induced motion at the hinges is used to pump high-pressure oil through hydraulic motors via smoothing accumulators. The hydraulic motors in turn drive electrical generators to produce the electrical power [75]. In general, it is made of movable, coupled segments. In normal sea conditions, with moderate vertical oscillation, cause segments of the plant to perform a horizontal evasive motion, like a sea snake swimming. Finally, hydraulic assemblies convert this motion into usable energy (Fig. 12). 4.12.4.4.2.4 Oscillating water column As shown in Fig. 13 water columns are formed within large concrete structures built on the shore line or on rafts and the structure is open at both the top and the bottom [76,77].

558

Hydropower Conversion

Fig. 12 Oscillating snake system [75].

Air turbine Oscillating water column Generator

Air

W ate

r

Waves

Sea

Cliff

Fig. 13 Water column oscillations [76].

The lower end of the cylinder is submerged in the sea and an air turbine fills the aperture at the top. The rising and falling of the water column inside the cylinder moves the air column above driving the air through the turbine generator. The turbine has movable vanes, which rotate to maintain unidirectional rotation when the movement of the air column reverses. 4.12.4.4.2.5 Pressure transducer system The hydraulic pump system has a submerged gas-filled tank with rigid sides and base in addition to a flexible bellows-like, top (Fig. 14). The gas in the tank compresses and expands in response to pressure changes from the waves passing overhead causing the top to rise and fall. A lever attached to center of the top drives pistons, which pump pressurized water ashore for driving hydraulic generators. It is often called a pressure transmitter, which is a transducer that converts pressure into an analog electrical signal. The conversion of pressure into an electrical signal is achieved by the physical deformation of strain gages, which are bonded into the diaphragm of the pressure transducer and wired into a wheat stone bridge configuration. Pressure applied to the pressure transducer produces a deflection of the diaphragm, which introduces strain to the gages. The strain will produce an electrical resistance change proportional to the pressure. 4.12.4.4.2.6 Wave capture systems They use a narrowing ramp to direct waves into an elevated reservoir (Fig. 15). Waves enter the funnel over a wide front and they are concentrated into a narrowing channel, which causes the amplitude of the wave to increase [78]. The increased wave height

Hydropower Conversion

27.12 (1.066)

75.57 (2.03)

1/2−1/4 NPT

559

24.58 (1.125)

Zero adjustment span on outer side

Conductor PVC cable

23.6 (1.125)

Fig. 14 Pressure transducer systems. PVC, poly vinyl chloride.

Wave capture system

Re

ser

voi r

Clif

f

Ra

mp

Tur b

ine

Fig. 15 Wave capture plant [78].

Overtopping

Reservoir

Turbine

Outlet

Fig. 16 Overtopping system [79].

coupled with the momentum of the water is sufficient to raise a quantity of water up a ramp and into a reservoir situated above the sea level. Water reservoir can then be released through a hydroelectric turbine located below the reservoir to generate electricity. 4.12.4.4.2.7 Overtopping wave systems They are floating plants as explained before focus waves onto a tapered ramp, which causes their amplitude to increase [79,80]. The wave crests overtop the ramp and spill into low lying dam then water from the low dam flows through hydroelectric turbines back into the sea beneath the floating structure (Fig. 16).

560

Hydropower Conversion

Tapered channel

Reservoir

Cliff face

Turbine house

Fig. 17 Tapchan wave power systems.

4.12.4.4.2.8 Tapchan model This system is located at the convenient places along the sea shore, where wave has significant heights for energy generation. It has a reservoir on the land for wave pushed water storage, and a turbine on the side for energy generation. As can be seen from Fig. 17, after the entrance of water into the reservoir, it flows through a turbine back to the sea. 4.12.4.4.2.9 Lever systems These energy capture systems have long levers that are mounted on steel piles or on floating platforms. Large floats or buoys are attached to the extremities of the levers, which move up and down with the waves. Movement of the lever arms forces water into a central hydraulic accumulator and through to a generator turbine. Alternatively high-pressure water can be pumped ashore to power shore based generators. 4.12.4.4.2.10 Technical challenges Wave energy capture has a set of formidable technical challenges and practical design systems. Variability of the sea conditions is very high and the system may be able to cope with a wide range of wave amplitudes and frequencies as well as changes in the directions of currents. Matching the generating equipment to the wave characteristics is required to convert the power of the irregular oscillating mechanical forces induced by the waves into electrical power. Typical rotating machines for power generation operate at a synchronous speed of 1200 rpm (20 revolutions/s) whereas the frequency of waves that drive the generator is likely to be between 5 and 10 s/cycle. A mechanical gearing system is needed to match this 200:1 ratio in operating speeds, possibly combined with special purpose, slow speed generators, incorporating a large number of pole pairs. One way around all of these problems is to use hydraulic accumulators either in situ or on shore to smooth out the energy delivery to the generator.

4.12.4.5

Small Hydropower Plants

These serve as hydroelectricity generation units to small community or individual industrial plants. Their definition is variable but in general 10 MW generation capacity is accepted as the upper limit of these hydroelectricity generation stations. However, in some countries this upper limit is regarded as 25 MW in Canada and 30 MW in the United States. The small HP stations may be connected to conventional electrical distribution networks as a source of low-cost renewable energy. On the other hand, their projects may be in isolated areas away from uneconomic to serve from a network, or in areas where there is no national electrical distribution network. They usually have minimal reservoirs and they have relatively low environmental impact in comparison to large HP plants. Hence, environmental impact is low and depends strongly on the balance between streamflow and power production.

4.12.4.5.1

Micro hydropower plants

Their capacity for HEP generation is up to 100 kW of power and installations can provide power to an isolated home or small community, or are sometimes connected to electric power networks. Around the world there are many of such type of power plants, which is mostly preferred in developing countries, because they can provide energy economically. These systems complement photovoltaic solar energy systems, because in many areas, water flow, and thus available hydropower, is highest in the winter when solar energy is at a minimum.

4.12.4.5.2

Pico hydropower plants

HP generation capacities less than 5 kW are referred to as Pico HP plants, which may help to light small and remote communities with only a small amount of electricity requirement [81]. Even smaller turbines of 200–300 W may power a single home in a

Hydropower Conversion

561

developing country with a drop of only 1 m. A typically setup of run-of-the-river type is without dams, but rather pipes divert some of the flow, drop water down a gradient.

4.12.4.5.3

Underground hydropower plants

These are used generally make use of a large natural height difference between two waterways, such as a waterfall or mountain lake. An underground tunnel takes water from the high reservoir to the power generation plant built in an underground cavern near the lowest point of the water tunnel [82,83].

4.12.5

Analysis and Assessment

Provide the necessary analysis method(s) for the system(s) and/or application(s) considered and their methods of assessment particularly for performance assessment through, for example, energy and exergy efficiencies. It is possible to view hydropower plant analysis in four interrelated branches. 1. Hydrological analysis: such plants are feasible provided that there are enough rainfall or snow occurrences and hence, the runoff in a river in a sustainable manner. This analysis includes also the extreme events, such as flood and drought [26]. The main tool is the discharge continuity curve, which reflects the discharge values versus percentage of time. 2. Topographic analysis: the most significant point in deciding about dam location is the possible water head fall availability for power generation, which can be deduced from a topographic map or better after detailed field trips in an optimum manner. Alignment for structures and also distances for transmission can be obtained from a topographic map. However, in these days Digital Elevation Model (DEM) data are very useful for preliminary field investigation on the computer screen electronically. 3. Geological analysis: dam stability is closely related to surface and also subsurface geological features, such as porous, fractured and karstic medium, faults, cleavages, dykes, etc. [16] especially, for the stability seepage below the dam must be reduced to ignorable levels. Additionally, dam embedment must be studied in detailed geological investigations through borehole loggings. 4. Socio-economic analysis: the main item is the demand level consideration for uninterrupted support. For this purpose, the demand level should be identified by consideration of possible risks that may arise during the hydropower plant operation and management. The feasibility to fulfill obligations of legislation, donors, and lenders are also very important issues. On the other hand the execution of any hydropower plant plan the following steps are necessary in sequence. 1. Office and desk studies: concerning the aforementioned analyses procedures especially engineers must look for any document and previous reports, studies, literature. 2. Reconnaissance field trips: these help to familiarize the personnel with the study area based on the documents collected from different sources. Validation of the available information can be carried out is such trips. 3. Feasibility studies: with all office and field trips a preliminary feasibility study should be carried out so as to assess the socioeconomic value of the project. 4. Decision and tender: after lengthy discussions by contribution of various shareholders the specifications are confirmed or altered and then the tendering, contracting, construction stages take place. 5. Implementation: operation and management works are started for hydropower generation and distribution.

4.12.5.1

Determination of the Required Capacity

The complexity of a reservoir capacity design depends upon the type of flow regulation. If regulation is of the over the year type, the analysis is based on the annual streamflow and a given degree of development. If the within-year storage fluctuations are considered, the analysis is usually made with monthly, weekly, or daily streamflow and the reservoir capacity can be determined by the following methods [84–87].

4.12.5.1.1

Graphical method using mass curve

1. Prepare a mass inflow curve from the flow hydrograph of the site for a number of consecutive years including the driest years when the discharge is low, as illustrated in Figs. 18 and 19 show the mass inflow curve for 4 years. 2. Assuming uniform demand, prepare a mass demand curve which is considered as a straight line as shown in Fig. 19. 3. Draw the lines AB, FG, etc. such that: a. They are parallel to the mass demand curve. b. They are tangential to the crests A, F, etc. c. The points A, F, etc., in Fig. 19 indicate the beginning of the dry period. 4. Determine the vertical intercepts CD, HJ, etc. between the tangential lines and the mass inflow curve. These intercepts indicate the volumes by which the inflow volumes fall short of demand. 5. Determine the largest of the vertical intercepts which represents the net storage capacity required, (R). If the tangential lines do not intersect the mass curve, the reservoir will not be filled again. 6. The gross capacity of the reservoir will be more than the net storage capacity. It is obtained by adding the storage losses (evaporation and seepage).

562

Hydropower Conversion

Mass inflow curve 10×105

Inflow hydrograph 300 Discharge

Accumulated flow

20×105

A1 1960

1961

1962

1963

200 A1

100 0 1960

1964

1961

1962

1963

1964

Accumulated inflow (×106 m3)

Fig. 18 Mass inflow curves.

FULL

G

12×103 FULL

F

EMPTY K

SPILL

8×103 B C

FULL

A

Demand curve

Mass inflow curve

R

4×103

M

L

EMPTY

D

N 1-year

E FULL 0 1950

1951

1952

1953

1954

Years Fig. 19 Determination of the storage capacity (R).

4.12.5.1.2

Analytical method

The following procedure is used for the determination of storage capacity for a long-term reservoir. 1. 2. 3. 4. 5.

Collect the annual inflow rates at the reservoir site for a long period of records (N) as possible (N450 years). For the period of records, calculate the mean value of the annual inflow, Imean. Calculate the departure of each annual inflow from the mean inflow (d). P Calculate the accumulated sums of the departures ( d). The storage capacity (R) can be calculated as the difference between the maximum and minimum of the accumulated sums.

4.12.5.1.3

Energy conversion

Moving water kinetic energy conversion to electric energy needs that water should move with sufficient speed with water volume to spin a turbine, which in turn rotates a generator to generate electricity. Approximately, one gallon of water per second falling one hundred feet can generate one kilowatt of electricity. In order to increase the volume of moving water impoundments or dams are used to store water behind these structures. Additionally, an opening in the dam or movement of water through pipes down to the turbines uses gravity (potential energy) to drop water down a pipe called a penstock. In HP generation there are a variety of turbines and their use depends on the amount of hydraulic head (vertical distance between the water level behind dam and the turbine). The most common ones are Kaplan, Francis, and Pelton types. Some of these designs use not just the kinetic force of the moving water but also the water pressure due to the water head.

Hydropower Conversion

563

The Kaplan turbine is similar to a boat propeller, with a runner (the turning part of a turbine) and three to six blades and it can provide power up to 400 MW. This turbine is different from other in the sense of HP turbines, because its performance can be improved by changing the pitch of the blades. The Francis turbine has a runner with nine or more fixed vanes and they are designed for up to 800 MW in size, where the runner blades direct the water so that it moves in an axial flow [6]. Another type of turbines is the Pelton turbine, which consists of a set of specially shaped buckets that are mounted on the outside of a circular disk and makes it look similar to a water wheel. These turbines are used typically in high hydraulic head sites and can be as large as 200 MW. On the other hand, run-of-the-river processes can also be generated HP without a dam and the volume and speed of water is not augmented by a dam. In a run-of-river project the turbine blades are sinned by capturing the kinetic energy of the moving water in the river. The run-of-river projects do not store water as behind dams and they have much less ability to control the amount and timing of when electricity is generated. Pumped storage is another type of HP technology, where water is pumped from a lower reservoir to a higher one during off-peak times when electricity is relatively cheap, using electricity generated from other types of energy sources. Haulage of water to higher elevations generates potential energy that can be used later at the time of energy need to generate HP. For this purpose, stored water at higher elevations is released back into the lower reservoir through turbines. Although, some power is lost during this process, but pumped storage systems can be up to 80% efficient, which corresponds to more than 90 GW of pumped storage capacity worldwide, with about 20% of that in the United States. The conversion of potential energy behind the dam is very easy through an intake gate to the pipe that leads water to turbines. The pipe that leads water to the turbines is referred to as the penstock. The falling water rushes toward the turbines and hit its blades with force and causes the turbine to spin. This is equivalent to the conversion of the kinetic energy to mechanical energy. The water that gives a great part of its kinetic energy to the mechanical energy leaves the turbines as wary water and runs to the downstream channel after the dam. The generator is connected to the turbine by means of a shaft, and hence, the spin of generator is satisfied by the turbine spin. In the generator there is an electromagnetic field that converts the mechanical energy to electrical energy.

4.12.5.1.4

Water turbines

These depend on the impulse of the working fluid on the turbine blades or the reaction between the working fluid and the blades to turn the turbine shaft, which in turn drives the generator. Several different turbines have been developed to optimize performance depending on particular water supply conditions. As mentioned before there are different turbine types that are convenient depending on the rate of water flow and the head or pressure of water [88–92]. 4.12.5.1.4.1 Turbine power output Turbines are means of converting moving water kinetic energy into rotational motion of the turbine shaft. In 1754 Leonhard Euler indicated the equivalence of the torque on the shaft to the change in angular momentum of the water flow with deflection by the turbine blades and power generation is equal to the torque on the shaft multiplied by the rotational speed of the shaft as in Fig. 20. The energy conversion is the change in angular momentum of the fluid between the turbine input and output.

Euler's turbine equation Q = Fluid flow rate

qin in Vin

 = Fluid density rin

q = Fluid velocity  = Incidence angle rout

out Vout

V = Tangential fluid velocity V = q cos 

qout Turbine

r = Turbine radius  = Turbine rotational speed T = Torque P = Power output

Torque T= Q (rinVin − routVout) Power P = T = Q (rinqin cosin − routqout cosout ) Fig. 20 Conversion mechanism of water kinetic energy into mechanical energy [92].

564

Hydropower Conversion

4.12.5.1.4.1.1 Pelton turbine It is an impulse turbine, which requires tangential water flow on one side of the wheel, and therefore, must operate when partly submerged (Fig. 21). Its use is the most appropriate in cases of high heads with low discharges, which is possible in a wide range of situations with heads that vary from 15 m up to almost 2000 m. High-pressure heads provide fast water jets impingements in the blades leading to very high turbine rotational speeds. These turbines are ideal for low power installations with outputs of 10 kW or less, but their output capacity may reach up to 200 MW with 95% efficiency [93]. 4.12.5.1.4.1.2 Francis reaction turbine It is a reaction turbine and operates in fully submerged position with water flow entrance in a radial direction toward the axis and exit in the axis direction [94] (Fig. 22). These turbines are suitable for rather low water heads of 500 m or less, and they are the most commonly used high power turbines. Large-scale turbines are capable of over 500 MW power delivery from a water head of 100 m with efficiencies of up to 95%. 4.12.5.1.4.1.3 Propeller and Kaplan turbines The propeller turbine is also a reaction turbine that works in full submergence (Fig. 23). It is, simply, similar in its geometrical form to a ship propeller, which is the most suitable design for low head water sources with a high flow rate such as those in slow running rivers. Various designs are optimized for a particular discharge (flow rate), but if the flow rate falls below the design rating accordingly efficiencies also drop of rapidly [95]. On the other hand, the Kaplan version has variable pitch vanes for its work efficiently over a range of flow rates.

Pelton turbine

Fig. 21 A typical Pelton turbine [93].

Fig. 22 A typical Francis turbine [94].

Hydropower Conversion

565

Fig. 23 A typical Kaplan turbine [95].

Generator

Stator

Rotor

Turbine generator shaft

Turbine

Water flow

Wicket gate

Turbine blades Fig. 24 Turbine and generator combination [96].

4.12.6

Power From Dams (Potential Energy)

Dam installations are for hydroelectric energy generation from the water potential energy stored to activate a water turbine, which in turn drives an electric generator. The energy output depends on the water head above the turbine elevation and also on the discharge (flow rate). Generally, as aforementioned turbines are of reaction types, whose blades are fully submerged in the water flow. Fig. 24 is a typical turbine and generator configuration and Fig. 25 is a cross-section of a dam, which shows all the components in any HEP generation [96–98]. As for the costs, the civil works of hydropower unit with dam construction are usually equal to many times the cost of the turbines and the associated electricity generators. However, dams provide a large water reservoir ready for the controllable water flow that generates desired level of hydroelectricity. For this purpose, the reservoirs serve as a supply buffer for excess water during rainy periods and for its release during dry spells.

4.12.7

Social and Political Water Power

The sanctity of water is recognized by societies. It has immersed social power to move people. The societies are deeply concerned with this power. Since, it has been the key driving force for sustenance and transportation of civilization. A very good example for such a case has appeared with the advent of steam engine leading to industrialized societies by using water power in the form of steam. Water constitutes socio-economic foundations of a society and environmental issues.

566

Hydropower Conversion

Hydroelectric power Reservoir Long distance power lines

Powerhouse Intake Generator

Penstock Turbine

River

Fig. 25 Hydroelectric energy generation dam cross-sections [96].

Fresh water is basic requirement of human being. It is the basic ingredient of food production, a moderator of economic growth and an essential social agent. As a source of incommensurable benefits to the planet and its people, it is also a vehicle of looming threads to human health and environment. In the next 30 years, more than 60% of the world’s population is expected to face water related problems in large cities. Availability of limited fresh water resources places extra stress on the water related issues due to population growth and rising living standards which are the driving forces of the industry at large. The magnitude of stress on water varies widely across the world. This brings the concerned people to ponder upon the rational and more efficient use of available resources without extravagant wastage and according to suitable management and operation rules. Otherwise, the existing small-scale issues arising from the water limitations might lead to political and economic crises. Unfortunately, today in the global world local political issues are becoming swiftly international conflicts which pave ways to some other unconcerned nations to put their nose into such water related sensitive questions. According to World Meteorological Organization about one-third of all low-income people live in areas of moderate to severe water stress. In a river basin there are many water resources activities both on the surface and in the subsurface in the form of groundwater reservoirs. Increasing demands bring additional stresses on these resources. The local conflicts emerge between city municipalities. For example, Istanbul city in Turkey as one of the world’s mage poles with more than 10 million populations presently try to meet its water demand not from the water resources within the municipal boundaries but transportation of water from nearby cities. In a way this presents a small-scale water rights agreement between the cities, the largescale of which might occur between neighbor countries as trans-boundary water resources rights. There are different conflicts between the cities of the same country let along the possibilities of trans-boundary water rights. In order to offset the water demand, today in many cities of the world and especially in developing or oil rich countries or in Israel where the water scarcity prevail so to say “supersurface” water resources are being exploited by weather modification studies, i.e., by artificial rainfall generation procedures. Unfortunately, this approach is not yet well documented but at the research stages. Hence, there are competitions for water resources with huge ramifications for water policy, particularly in countries with high water stresses especially in the Middle East, southwest United States and South Africa. As explained above, many such competitions will arise in the future not only concerning the quantity of water but also perhaps more significantly the quality of water with pollution frees preservation. Not only the global population growth but social, environmental, and industrial activities will escalate the risks of political conflicts over the limited and unevenly distributed water resources. Such tension and conflict possibilities make fresh water hence more social and international power than its physical power. In fact, physical power creates social power and therefore prior to any physical intervention on water resources social and political means must be debated with rightful understanding. A substantial portion of fresh water resources is contained in international drainage basins, basins shared by two or more nations. These basins make up nearly 50% of the earth’s land area, and 60% of the area of both Africa and Latin America [99]. Social and economic powers of water became more interesting especially that the international agreements are often either inadequate or lacking entirely in some parts of the world. In the future, it is most likely that political tensions over social and economic power of water might lead to tensions over shared rivers, lakes, aquifers and even clouds, if the artificial rainfall generation procedures become more efficient. Beijing Declaration adopted by the Conference at its concluding session on 21 March 1996 brought more than 150 international experts from some 50 countries. The conference based its considerations on the following internationally accepted principles and recommended their consistent applications. The following points are considered for collaboration: 1. Fresh water is a finite, vulnerable resource, essential to sustain life, development and environment. 2. Water development and management should be based on a participatory approach, involving users, planners, and policymakers at all levels.

Hydropower Conversion

567

3. Women play a central part in the provision, management, and safeguarding of water. 4. Water has an economic value in all its competing uses and should be recognized as an economical good. 5. Mobilizing financial resources is critical to effective water resources management. Rojanamon et al. [43] stated that public participation techniques are used for data collection to help explain social impact and they also presented a detailed literature review, which is summarized in the following sequel. Social impacts are changes that occur in people’s everyday life, livelihood, culture or heritage and community from the implementation of a project, program, policy, or plan. It is often not possible to predict exactly what will happen to people and their community as a result of a development project. It is possible to provide an estimate and understanding of what might happen, why and what should be done to prevent harm, and to respond to the needs and concerns of those people who might be affected.

4.12.8

Case Study

Niksar is located in Tokat province of Turkey, where Kelkit Stream exists with its 11,445 km2 surface area and main channel length of 245.5 km as one of the main branches of Yes- ilırmak River (Fig. 26). Kelkit Stream is formed by small creeks at an elevation of 1500 m.

Sea

Aegean

Sea

Black

Turkey

Sea

Mediterranean

Fig. 26 Yes¸ilırmak and Kelkit drainage basins.

500 450

Discharge (m3/s)

400 350 300 250 200 150 100 50 0 0

20

40

60 % of time

Fig. 27 Kelkit flow duration curve (FDC) [101].

80

100

568

Hydropower Conversion

Table 2 Energy production flow duration curve (FDC) Q (m3/s)

Hydro energy (MWh)

50 60 70 80 90 100

204,366 233,626 250,874 262,945 271,871 279,144

This stream drainage basin constitutes 55% of Yes- ilırmak total flow [101]. The climate of the region has transitional characteristics between Black Sea and Central Anatolian climates, where rainfalls mostly appear in spring and in December–March period snow falls. With snowmelt in spring Kelkit Stream discharge increases significantly. As explained before the FDC gives the relationship between the percentages of time of flow has being equal or more for the period of record. The FDC for Kelkit stream is given in Fig. 27. This curve provides information about the future regime of the river and is used to estimate the electricity generation potential of a power scheme at a given location. FDC can be computed from mean monthly flows or daily flows. It is preferable to use daily flows since there may be significant variation in flow within a month. These variations cannot be seen in monthly FDCs. In order the curve to represent the flow with sufficient accuracy, it is necessary to have a record length of at least 30 years. The FDC is derived from the past flow data, and it is accepted to be time-invariant. Therefore, this procedure includes some uncertainties, but it is acceptable for feasibility studies [102]. In most cases the firm energy is adapted as for discharge value that corresponds to 95% of the time, and hence, from Fig. 27 one can see that discharge equals to about 40 m3/s. Turkish State Hydraulic Works calculated the design discharge for Niksar HEP as 60 m3/s, exceeding 40% of the time. In Turkey as a rule of thumb discharges corresponding to 20–30% of time are often identified as design discharges for small HEPPs in Turkey [101]. Fig. 2 is used for energy calculations, where the data covers the monthly average discharges between 1966 and 2001 (Table 2).

4.12.9

Future Directions

Recent advances in turbines and data collection techniques increase fish passage effectiveness, which provides new opportunities for the HP industry. Improvements in construction and operation so as to minimize environmental and cultural impacts, HP projects provide low cost and clean sources of electricity to urban and rural areas throughout the world. In the future dependence increment on HP will result in reduction on fossil fuel, and hence, reduction of anthropogenic gases. Coal and natural gas dependence can be reduced, because of the HP flexibility, which is operated by up and down water releases to turbines, and hence, help to integrate larger amounts of variable renewable energy resources, like wind and solar power. In some countries there are HP expansion possibilities, and therefore, depending on the population growth and economic development the number of HEP plants is bound to increase. According to the International HP Association, more than 30 GW of new HP capacity was commissioned in 2012, with significant investment occurring in South America, Asia, and Africa [100]. In Brazil, three large projects are under construction in the Amazon region totaling more than 22 GW of generation capacity. In general, according to a set of scenarios the world energy is expected to increase steadily until 2030 and the global primary energy demand is projected to increase by 1.7% per year from 2000 to 2030, reaching at annual level of 15.3  109 t of oil equivalent (toe). On the other hand, the global oil demand is expected to increase about 1.6% per year from 75  106 to 120  106 barrels/day. The major energy consumer is the transportation sector at almost three quarters of the stated amounts. Oil will remain the fuel of choice in transportation. The long-term energy sources are hoped to have the following important points for a safer and pleasant environment in the future [35]. 1. Diversity of various alternative energy resources both nonrenewable and renewables with increasing trend coupled with decreasing trend in the nonrenewables. 2. Quantities must be abundant and sustainable for the long future. 3. Acceptable cost limits and compatible prices with strong economic growth. 4. Energy supply options must be politically reliable. 5. Friendly energy resources for the environment and climate changes. 6. Renewable are domestic resources that help to reduce the important energy alternatives. 7. They can support small to medium scale local industries.

Hydropower Conversion

569

The renewable energies are expected to play an active role in the future energy share because they satisfy the following prerequisites. 1. They are environmentally clean, friendly and do not produce greenhouse gases. 2. They should have sufficient resources for larger scale utilization. For instance, the solar energy resources are almost evenly distributed all over the world with maximum possible generatable amounts increasing toward the equator. 3. Intermittent nature of solar and wind energy should be alleviated by improving the storage possibilities. 4. Cost-effectiveness of the renewable is one of the most important issues that must be tackled in a reduction direction. However, new renewable energies are now, by and large, becoming cost competitive with conventional forms of energy. Although throughout of this chapter energy generation is explained for HEP in detail, but another very important issue is efficient energy use and savage. On this point the following points is a list of some guidance 1. Conservation and more efficient use of energy. Since the first energy crisis, this has been the most cost-effective mode of operation. It is much cheaper to save a barrel of oil than to discover new oil. 2. Reduce demand to zero growth rate and begin a steady state society, and 3. Redefine the size of the system and colonize the planets and space. For instance, the resources of the solar system are infinite and our galaxy contains over 100 billion stars. Because the earth resources are finite for population a change to a sustainable society depends primarily on renewable energy that becomes imperative on a long time scale. The following adaptation and mitigation policies must be enhanced in every society. 1. 2. 3. 4.

Practice conservation and efficiency. Increase the use of renewable energy. Continue dependence on natural gas. Use of coal, but include all social costs (externalities).

Regional and local polices must be the same. Efficiency can be improved in all major sectors including residential, commercial, industrial, transportation, and even the primary electrical utility industry. The most gains can be accomplished in the transportation, residential, and commercial sectors. National, state, and even local building codes will improve energy efficiency in buildings. Finally, there are a number of things that each individual can do in conservation and energy efficiency.

4.12.10

Advantages and Disadvantages of Hydroelectric Power

Any HP generation system with all its components has the following advantages, which can also be improved with new scientific procedures and technology. 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11.

After the dam construction, the energy generation is virtually free. There is no waste or pollution production. Hydroelectricity generation is much more reliable than other renewable energy sources. Water storage provided the ability to cope with peaks in demand. Hydroelectric power stations can respond sudden energy loads very quickly, unlike other power stations. It is possible to generate electricity energy at a fixed level constantly, Fossil fuel is not burned so there is very little pollution, and hence, HEP plays a major role in reducing greenhouse gas emissions. There is no cost of water to run the power plant. Operations and maintenance costs are comparatively low. HEP plant technology has been proven over time and it is reliable. The dam reservoir is renewed after each storm rainfall over the drainage basin.

As mentioned before the disadvantages of HEP generation are comparatively rather of minor significance compared with other renewable energy alternatives. 1. Although the dam construction is very expensive, but in order to reduce the cost many dams are also used for flood control or irrigation. 2. The inundation area behind a dam covers very large area, which may cause problems to the human settlers and to animal habitant. 3. Selection on a suitable dam construction cross-section is very difficult decision-making, because of the impact on residents and the environment may be unacceptable. 4. Water quality and quantity at the downstream of dam can be affected, which in turn causes an impact on plant life. 5. Valuable and limited natural resources are exploited in ant HEP plant construction. 6. Loss or modification of fish habitat and fish entrainment or passage restriction. 7. Local population displacement.

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Unfortunately, limited fresh water resources are bound to be more limited and scarce with economic value in the future if present rate of water quality deterioration mechanism, such as environmental pollution, global climate change, CO2 emission rate, greenhouse effects, and ozone layer depletion problems are not combated jointly all over the world. However, the real problem is to try and solve these problems without creating any other problems of economic, political, military or other types. In other words, water problems must be confined within the water problems only with solutions. For this purpose the following guidance must be considered: 1. Data collection, dissemination, and treatment procedures must be rendered to more efficient and internationally recognized standards by collaborations. Temporal and spatial distribution of water resources of any type must be assessed both on national and international levels. 2. The socio-economic and cultural weights of water must be emphasized because the importance of one or the other varies temporally and spatially depending on each nation. 3. Every society should have commonsense and should search for rightful solutions to water related problems especially within the next 50 years during which water conflict scenarios are expected to be fabricated even without any sound basis. 4. Water power of any type should serve societies on equitable rights with mutual understanding and scientific management programs.

4.12.11

Closing Remarks

The world is in need of any energy source for future prosperous, reliable and sustainable development even though present energy sources are sufficient at some parts, whereas other parts lack sufficiency. Populations increase, economic development and modern technologies require vast amounts electricity generation for societal activities. Presently, the global energy challenge is to tackle the threat of climate change, meet the rising demand for energy and to safeguard security of energy supplies. Renewable energy sources (hydropower, solar irradiation, wind, etc.) are effective energy technologies that are ready for global deployment today on a scale that can help tackle climate change problems. Among the sources HP energy generation is significant in the climate convenient regions of the world, where there are frequent storm rainfalls that give rise to surface flow (runoff). This paper presents fundamentals and systems of HP generation devices by taking into consideration the historical, climatic, hydrologic, engineering, and technological points. Hydrological cycle plays the major role for HP generation possibility as a renewable energy alternative in different parts of the world. It is the most mature, reliable and cost-effective renewable power generation technology. HP is at the service of human since time unmemorable for various social activities and initially for grinding of grains such as wheat. Abu-l Iz Al-Jazari [12], who lived during the 12th century, is the father of robotics that worked by water power hydraulically. The electrical turbines and generators are developed by the second half of the 19th century, and accordingly, electricity generation from water power has also started to serve society. HP energy is environment friendly, clean, renewable, and able to meet peak demands, highly efficient (more than 9%), involves no fuel cost, is a balancer of energy prices, and has a long life-span about 200-years. The HP electricity extraction depends not only on the water volume per time (discharge), but also on the elevation difference between the water inflow head and the water outflow, where turbines are located. The water power turns turbine at enormous speeds by means of a connected shaft to an electric generator, which generates electricity. In order to increase the volume of moving water impoundments or dams are used to store water behind these structures. Additionally, an opening in the dam or movement of water through pipes down to the turbines uses gravity (potential energy) to drop water down a pipe called a penstock. The potential impact of climate change on the hydrologic regime is a crucial question for water resources management. Potential change in hydrologic regime resulting from climate change is an important topic in contemporary hydrology and water resources management. Increasing trend in the CO2 concentration may lead to unwanted climatic disruptions and local imbalances in the hydrological as well as atmospheric cycles will be the consequences, which may lead to excessive rainfall or drought, in addition to excessive heat and cold. The environmental and social problems can be alleviated by effective energy policies with consideration of renewable energy sources, such as hydropower, solar, wind, biomass, wave and geothermal energies as well as the solar hydrogen energy possibilities. HEP production methods are functional by means of technological plants that depend on water movement and convert its potential and kinetic energy laden volumes into hydroelectric energy. The most common plants are dams at convenient locations and hydrological cycle function to produce rainfall and snow. In this paper, various technological aspects of HP energy generation are presented in a sequential manner including reservoir impoundments, run-of-river, pumped storage and other alternatives, such as tidal and wave powers, oscillating float, paddle, and snake, overtopping wave systems, Tapchan model and small HP plants. Several different turbines have been developed to optimize performance depending on particular water supply conditions. As mentioned before there are different turbine types that are convenient depending on the rate of water flow and the head or pressure of water. Three types of turbines that are used in different HP electricity generation are Pelton, Francis, and Kaplan turbine designs.

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4.13 Magnetic Energy Conversion Ercüment Yüzüak, Recep Tayyip Erdog˘an University, Rize, Turkey Gizem Durak Yüzüak, ˙Ilker Dinçer, and Yalçın Elerman, Ankara University, Ankara, Turkey r 2018 Elsevier Inc. All rights reserved.

4.13.1 Introduction 4.13.2 Background/Fundamentals 4.13.2.1 Isothermal Magnetic Entropy Change 4.13.2.2 Magnetic Entropy Change Under Adiabatic Conditions 4.13.2.3 Thermodynamic Equilibrium 4.13.2.4 The First-Order Phase Transitions and Giant Magnetocaloric Effect 4.13.2.5 The Effect of Magnetic Field on the First-Order Phase Transition 4.13.2.6 Hysteresis Effect 4.13.2.7 Determination of Magnetocaloric Effect 4.13.2.7.1 Direct measurement technique 4.13.2.7.2 Indirect measurement technique 4.13.2.8 Calculation of Magnetocaloric Effect from the Heat Capacity Measurement 4.13.2.9 Calculation of the Magnetic Entropy Change by Using the Clausius–Clapeyron Equation 4.13.3 Analysis and Assessment 4.13.3.1 Magnetocaloric Materials 4.13.3.1.1 Gadolinium (Gd) 4.13.3.1.2 La(Fe–Si)13 based alloys 4.13.3.1.3 (Mn–Fe)2–(P–Z) based alloys 4.13.3.1.4 Manganites 4.13.3.1.5 XMnGe (X: Co or Ni) based alloys 4.13.4 Case Studies 4.13.4.1 Magnetic Energy Conversion 4.13.4.2 Magnetic Cooling 4.13.5 Results and Discussion 4.13.6 Future Directions 4.13.7 Closing Remarks Acknowledgment References Further Reading Relevant Website

Abbreviations AMR

Active magnetic regenerator

Nomenclature f F G H Hd M m P S SE SL SM

Frequency (Hz) Helmholtz free energy Gibbs free energy Enthalpy Demagnetization field (T) Magnetization (emu g 1) Magnetic moment (mB) Pressure (mTorr) Entropy (J kg 1 K 1) Electronic entropy (J kg 1 K 1) Lattice entropy (J kg 1 K 1) Magnetic entropy (J kg 1 K 1)

Comprehensive Energy Systems, Volume 4

CFC HCFC

Chlorofluorocarbon Hydrochlorofluorocarbon

T TC U V q Q a g ∆H DSiso ∆SM ∆Tad

Temperature (K) Curie temperature (K) Internal energy Volume (cm3) Cooling power (J kg−1) Latent heat (J g 1) Alpha phase Gamma phase Magnetic field change (T) Isotropic entropy change (J kg 1 K 1) Magnetic entropy change (J kg 1 K 1) Adiabatic temperature change (K)

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4.13.1

Introduction

Today, conventional vapor compressed refrigeration systems are widely used for home appliances and large commercial applications. Nevertheless, the research on methods is ongoing because of the amount of energy consumption and to avoid damage to the environment by refrigerant cooling materials (CFC, HFCF, etc.). Accordingly, research and development (R&D) studies continue on magnetic cooling systems, which are being developed. The magnetic cooling system is explained as ferromagnetic or paramagnetic material that exhibits a magnetocaloric effect (MCE) using a heat exchanger and a heat transfer fluid, to make the heat exchange from the ambient. A material that bears MCE is warmed up due to the alignment of magnetic moments in parallel position in the direction of the magnetic field when the magnetic field is applied. When the magnetic field is removed, the moments of the array are restored and the material tends to cool down. Gadolinium shows the most efficient MCE among the materials according to thermodynamic calculations and experiments. The MCE was discovered by German physicist Emil Gabriel Warburg in 1881 [1]. It was first observed in an iron sample in the range of mK (mili-Kelvin). It warms up under the magnetic field, and after removing the magnetic field, it cools down to the beginning temperature. Independently of each other, in 1926 Debye [2] and in 1927 Giauque [3] created a reversible process and reached the lower temperatures to observe the MCE in the paramagnetic salt under the magnetic field. The first experiment on magnetic cooling was conducted by Giauque and Macdougall in 1933 [4]. With the technology being used, they reached lower temperature values below 1K by a cool-down process. Nowadays, by using simple technologies, magnetic cooling can reach extremely low-temperature values. The terms of magnetic energy and electromagnetic energy are used as an alternative source of magnetic cooling technology where the magnetic field is applied. Magnetic cooling applications near room temperature are important alternatives to the conventional vapor compression cooling systems. That situation is mainly originated from the cooling cycle, which is environmentally friendly due to not using harmful gas, has a simple design, has low operating cost, has low efficiency because of not using components (such as conventional cooling systems compressors, etc.) in the system, has high energy productivity, is noiseless, and reaches lower pressure, as well. In addition to these advantages, it is possible to reach high yield values by making various modifications to the system during the heat transfer process. Investigation of magnetocaloric materials (MCMs) gained importance in the mid-1970s. In 1976, experiments were done and the magnetic cooling system was first created by Brown near room temperature (temperature distribution is 47K) [5]. In 1982, Barcalay and Steyert [6] developed the active magnetic regenerator (AMR) system and were granted the first patent on refrigeration. In 1998, Ames Laboratories and Aviation-Astronautics Corporation collaboration created another active magnetic cooling system by using 5 T magnetic field magnet. They reached the 12K temperature span at room temperature, with 6.6 as the performance coefficient (COP) and 550 W cooling capacity resulting with this system. In 2001, second and third generation magnetic cooling systems were developed by the same scientists. At the second generation, 1.5 T permanent magnet was used and reached 251C temperature span with a cooling power of 50 W. At the third generation, two 1.5 T magnets were used and 101C temperature span with 840 W cooling power was obtained. Within the scope of these studies, the use of difficult-to-control and expensive superconducting electromagnets is proved to be unnecessary. This situation is evidence of magnetic refrigeration suitable for commercial applications. Recently, some works have been carried out on magnetic cooling regarding the magnetic cooler and fluid, the use of regenerator, and design and morphology of the system. In this study, the effect of magnetocaloric concept, MCMs, magnetic cooling principles, basics of thermodynamic approach, comparison with a conventional vapor compression cooling system, and evaluation of advantages and disadvantages of some up-to-date investigations are examined.

4.13.2

Background/Fundamentals

All magnetic materials show a MCE, but the severity of MCE depends on the properties of the materials. MCE depends on the arrangement of magnetic moments in the magnetic materials under the applied magnetic field. This order brings about a change in the entropy of the material [7,8]. The magnitude of MCE is set as the isothermal magnetic entropy change ∆SM;iso or the adiabatic temperature change ∆Tad derived from an applied external magnetic field. The total entropy of magnetic material S, depending on the local magnetic moments where magnetization is at constant pressure, is expressed as the sum of the electronic, the lattice, and the magnetic entropy. These three functions are represented as the functions of the temperature and the applied external magnetic field [7,9]. SðT; HÞP ¼ ½SE ðT; HÞ þ SL ðT; HÞ þ SM ðT; HފP

ð1Þ

When the value of magnetic field is altered from H1 to H2, there is some variation observed from SE and SL while magnetic entropy change SM can be increased. In Fig. 1, a magnetic material of Gd5Ge2.025Si1.925In0.05 of temperature-dependent entropy diagrams is given under two different magnetic fields (zero and under 5 T magnetic field). MCE thermodynamic is explained by two different transformations in the graph.

Magnetic Energy Conversion

160

GdSiGeIn ΔTad(T) = T2 −T1

140 S(T )H,P (J mol K)

575

S1, T1 S1, T2

120

100

H 80

=0

H

=5

240

S2, T1 ΔS (T ) = S −S M 2 1

T T

260

280 T (K)

300

320

Fig. 1 Total entropy function of GdSiGeIn compound under magnetic field. Reproduced from Yüzüak E, Dincer I, Elerman E. Giant magnetocaloric effect in the Gd5Ge2.025Si1. 925In0. 05 compound. Chin Phys B 2010;19:037502.

4.13.2.1

Isothermal Magnetic Entropy Change

At a constant temperature, the magnetic field is changed from H1 to H2 value and MCE is defined as isothermal entropy change. The resulting heat in a reversible process is detracted from the system, i h i h ð2Þ ¼ ½SðTÞH2 SðTÞH1 DSM ðTÞT;DH;P ¼ SM ðTÞH2 SM ðTÞH1 T;P

T;P

The magnitude and size of the ∆SM is dependent on the relation between SðTÞH1 and SðTÞH2 at constant pressure.

4.13.2.2

Magnetic Entropy Change Under Adiabatic Conditions

If the value of magnetic field changes from H1 to H2 under adiabatic conditions, magnetic entropy changes, but the total entropy change remains constant as defined well in advance ðDS ¼ DSL þ DSE þ DSM ¼ 0Þ. Also, lattice and electronic entropy can be changed as DSL þ DSE ¼

ð3Þ

DSM

Adiabatic temperature change DTad is defined as temperature change under adiabatic conditions between S(T)H curves. It is also used to calculate the MCE of a material. DTad(T)DH is a function of temperature. The measured adiabatic temperature change DTad includes also a lattice, electronic, and magnetic entropy changes. For a constant DH and for any T is defined as h i ð4Þ DTad ðTÞT;DH;P ¼ TðSÞH2 TðSÞH1 S;P

As a result, according to Eqs. (2) and (4) the MCE is characterized by the total entropy of behavior of the magnetic material where both temperature and magnetic field of the function are known. With increasing magnetic field, magnetic moments in the material are arranged. Thus, the magnetic entropy change ∆SM is observed to be negative. Under adiabatic conditions, material is heated and ∆Tad is measured as a positive value. Similarly, if the applied magnetic field is reduced, the magnetic ordering decreases and ∆SM becomes a positive value. The cooling process is observed in the material under adiabatic conditions. In Fig. 1, total entropy function of Gd5Ge2.025Si1.925In0.05 compound is given under the magnetic field of 0 and 5 T. MCE is represented by both isothermal entropy change and adiabatic temperature change with the horizontal and vertical directions of temperature [7].

4.13.2.3

Thermodynamic Equilibrium

MCE shows larger values usually where the second-order phase transition occurs in the materials. In a typical regular irregular magnetic transition, the total entropy is a continuous function of temperature and this transition is completely reversible. All changes in the magnetic system are assumed to be under applied thermodynamic equilibrium and equilibrium process. For defining the MCE in the magnetic systems, the overall thermodynamic terms are used: the internal energy (U), the Gibbs free energy (G), and the free energy (F) [9]. The U of the ordinary system can be presented as a function of the magnetization (M), the volume (V), the entropy (S); U ¼ UðS; V; MÞ

ð5Þ

576

Magnetic Energy Conversion

The Helmholtz free energy is expressed as a function of volume (V) and magnetic field (H) at constant T, in order to use in constant volume systems is described as F¼U

ð6Þ

TS

The Gibbs free energy is used with constant pressure systems. The temperature (T), magnetic field (H), and the pressure is expressed as follows: G¼U

TS þ pV

m0 MH

ð7Þ

If the first derivative of the Gibbs free energy is discontinuous during phase transition, the transition is called first-order phase transition. If magnetic material exhibits the first-order phase transitions, generally the volume, magnetization, and entropy show discontinuity. If the first derivative of Gibbs free energy is continuous during the transition but the second-order derivative is discontinuous, this transition is called second-order phase transition [9]. Total differential according to the U, S, and G is given dU ¼ TdS dF ¼

pdV

SdT

dG ¼ Vdp

m0 HdM m0 HdM

pdV SdT

m0 MdH

ð8Þ ð9Þ ð10Þ

The external parameters entropy (S), pressure (P), and magnetic field (H), which contains the free energy F, are expressed by the following equations: SðT; H; VÞ ¼

ð∂F=∂T ÞH;V

ð11aÞ

HðT; M; VÞ ¼

ð∂F=∂MÞV;T

ð11bÞ

pðT; H; VÞ ¼

ð∂F=∂V ÞH;T

ð11cÞ

According to the Gibbs free energy of the above equation it is expressed as follows: SðT; H; pÞ ¼

ð∂G=∂T ÞH;p

ð12aÞ

MðT; H; pÞ ¼

ð∂G=∂HÞT;p

ð12bÞ

VðT; H; pÞ ¼

ð∂G=∂pÞT;H

ð12cÞ

If the magnetic moment M is chosen in G as an external variable instead of the magnetic field H, then ð12dÞ

H ¼ ð∂G=∂MÞT;p

It is possible to derive entropy from external magnetic field, pressure, and magnetization by using above the Maxwell equations related to (12a,b) to (12a,c) and (12a,d).     ∂S ∂M ¼ ð13aÞ ∂H T;p ∂T H;p   ∂S ¼ ∂p T;p





  ∂H ∂T M;p

 ∂S ¼ ∂M T;p

∂V ∂T



ð13bÞ H;p

The entropy change for a reversible process according to the second law of thermodynamics is given as follows:   δQ dS ¼ T While x is the fixed parameter of the system and the system’s heat capacity Cx is given by the following:   δQ Cx ¼ T x

ð13cÞ

ð14Þ

ð15Þ

where δQ is the heat quantity changing the system temperature on dT. Using the second law of thermodynamics δQ ¼ TdS, the heat capacity can be represented as   ∂S Cx ¼ T ð16aÞ ∂T x

Magnetic Energy Conversion

577

and total entropy change is calculated: CðTÞx dT ð16bÞ T Total entropy S as a function of the temperature, pressure, and magnetic field in the magnetic system with the form of total differential is expressed as follows:       ∂S ∂S ∂S dT þ dH þ dp ð17Þ dS ¼ ∂T H;p ∂H T;p ∂p T;H dSðTÞx ¼

Under an adiabatic–isobaric process (dS¼ 0 and dp¼ 0) the temperature change due to the change of the magnetic field and magnetization (the MCE) is obtained from Eqs. (13a,c), (17), and (16b) as   T ∂M dH ð18Þ dT ¼ CH;p ∂T H;p dT ¼

  T ∂H dM CM;p ∂T M;p

ð19Þ

CH,p and CM,p are heat capacity constants of the system at constant pressure-magnetic field and constant pressure-magnetization. Similarly, the total entropy of the system as a function of temperature, pressure, and magnetization of the material is written as follows:       ∂S ∂S ∂S dT þ dM þ dp ð20Þ dS ¼ ∂T M;p ∂M T;p ∂p T;M by integrating the Eqs. (19) and (20), ∆Tad (adiabatic temperature change) for a finite magnetization is acquired and MCE is defined as follows:    Z H2  T ∂MðT; HÞ DTad ðT; DHÞ ¼ dH ð21Þ CðT; MÞ M ∂T H1 H    Z M2  T ∂HðT; MÞ dM ð22Þ DTad ðT; DMÞ ¼ CðT; MÞ M ∂T M1 H Finite entropy change under isotemperature and isopressure according to the Maxwell equations is observed in the following:  Z H2  ∂MðT; HÞ DSM ðT; DHÞ ¼ dH ð23Þ ∂T H1 H As a result, Eqs. (21)–(23) will lead to the following conclusions: 1. The magnetic entropy change depends on both the temperature change at constant magnetic field and derivative of magnetization. As a result, large MCE is expected at the large value of derivative of temperature-dependent magnetization. In ferromagnetic materials the maximum value of ð∂M=∂T ÞH is closed to the Curie temperature (TC) of the system. At T¼ TC, ∆Tad and ∆Tiso should have maximum value. 2. If this effect is usually observed in ferromagnetic and paramagnetic material, in a constant magnetic field, magnetization decrease might cause temperature increase ð∂M=∂T ÞHo0. Therefore, adiabatic temperature and isotemperature entropy change are calculated positive and magnetic field change is calculated negative. This is defined as a direct MCE. The opposite is defined as the inverse MCE.

4.13.2.4

The First-Order Phase Transitions and Giant Magnetocaloric Effect

A large MCE is observed in the materials that exhibit the first-order structural and magnetic phase transition. In these materials, the first-order structural phase transition is due to the change of the interaction force between the magnetic moments. As a result of structural parameter (unit cell volume, the lattice parameters) discontinuities, the discontinuity is observed in the temperature dependence magnetization measurements (M(T)). The value of the MCE will increase with the applied/removed magnetic field under first-order phase transition. This effect is defined as a “giant MCE.” The word “giant” was first used in 1973 [8]. The unexpected value of MCE in Gd5(Si2Ge2) and other Gd5(Si4 xGex), x Z2 then this effect is taken into the important case from Pecharsky and Gschneidner [10]. The giant MCE (GMCE) may also be observed with these materials in many different studies. Some of these are MnFe(P1 xAsx), La(Fe13 xSix), MnAs compounds, and Ni–Mn–Ga, Ni–Mn–Sn, Ni–Mn–In, Ni–Mn–Sb as Heusler compounds [8,11]. The Clausius–Clapeyron equation is a valid equation for first-order phase transitions. Clausius–Clapeyron equation is to establish a relationship among transition temperature (Tt) at the specific magnetic field (Ht), magnetization change (DM), and entropy change (DS). DS ¼

DM

∂Ht ∂Tt

ð24Þ

578

Magnetic Energy Conversion

Entropy - S(T)H,P

ΔEH1 + Tpt,H1

Tpt,H1

C (T )H1,P dT T

0

ΔEH2 + Tpt,H2

S (T )H1

Tpt,H 2

Tpt,H1

0

0

C (T )H 2,P dT T

C (T )H1,P dT T Tpt,H2 S (T )H 2 0

C (T )H 2,P dT T

Temperature - T (K) Fig. 2 The diagram of magnetic system having a first-order transition. Adapted from Pecharsky VK, Gschneidner KA Jr., Pecharsky AO, Tishin AM. Thermodynamics of the magnetocaloric effect. Phys. Rev. B. 2001;64:144406.

DE

T is expressed as “transition enthalpy,” which occurs during the field-induced magnetostructural phase transitions. During this transition, the MCE is expected to increase because the transition enthalpy change occurs. Furthermore, Tt , the terms resulting t from the magnetic field changes, and the ∂H ∂Tt rate should be attempted as big as possible. In the presence of a large entropy change, with additional contributions brought about by the structural phase transition occurring, this is known as the GMCE [10]. Paired entropy of a system under magnetostructural phase transition is shown in Fig. 2. The first-order transition is manifested in the presence of a temperature and magnetic field hysteresis. Critical points are marked on the graph of the energy. If the magnetic field is small enough to be neglected, the contribution to the heat capacity at the under-phase transition temperature is Tpt ;H1 and the over-phase transition temperature is Tpt ;H2 , the impact of the heat capacity is suppressed by a high magnetic field [8]. At a constant temperature and pressure, the observed phase transition is calculated from Eq. (17) and written as Z T h Z Tt;H l C ðTÞH;P C ðTÞH;P DEH dT þ dT ð25Þ þ SðTÞH; P ¼ Tt;H T T Tt;H T1 -0

wherein H is the applied magnetic field, Tt;H is the phase transition temperature, and DEH is the transition enthalpy. Cl and Ch show heat capacity values at the low- and high-temperature phase where two phases are different from each other [7]. The materials that show first-order phase transition materials generally affect the transition temperature, but below and above phase transition temperature, the impact of the heat capacity is relatively small [7]. Cl ðTÞH DCh ðTÞH ¼ CðTÞ

ð26Þ

While the applied magnetic field change is constant, the first order of magnetic entropy change during the phase transition causes a large increase. This is because of the first-order phase transition, which increases the MCE value because of the enthalpy values [7]. DSM ðTÞDH;p D

4.13.2.5

DEH Tt;H

ð27Þ

The Effect of Magnetic Field on the First-Order Phase Transition

The most important effect of the magnetic field on the magnetic and structural phase transitions is to force to phase transition and made a shift to the phase transition temperature. In a crystal, the crystal structure affects directly structural phase change and magnetic properties significantly. The effect of magnetic field on the first-order phase transition temperature is calculated using the Clausius–Clapeyron thermodynamic approach [12]. g and a are different phases from each other and they are expressed with the Gibbs free energy as follows: Gg ¼ Ug

TSg þ pVg

HMg ¼ Fg ðT; pÞ

HMg

ð28aÞ

Ga ¼ Ua

TSa þ pVa

HMa ¼ Fa ðT; pÞ

HMa

ð28bÞ

In the case where the system is in equilibrium, the thermodynamic potential (free energy),Gg and Ga phases are equal to each other. Gg ðT; H; pÞ ¼ Ga ðT; H; pÞ

ð29Þ

Magnetic Energy Conversion By calculating differential on both terms of Eq. (29), the first equation is obtained.       ∂Ga ∂Gg ∂Ga ∂Gg ∂Ga ∂Gg dT þ dH þ dp ¼ 0 ∂T ∂T ∂H ∂H ∂p ∂p

579

ð30Þ

Gibbs free energy is given in Eqs. (12a)–(12c), and by utilizing Eq. (33) the following can be written. ∂Ga ∂T

∂Gg ¼ Sa ∂T

Sg ¼

Q Tt

ð31Þ

∂Ga ∂H

∂Gg ¼ Ma ∂H

Mg

ð32Þ

∂Ga ∂p

∂Gg ¼ Va ∂p

Vg

ð33Þ

By combining the equity of Eqs. (30) and (31) and ignoring the volume change during phase transitions, structural phase transition temperature change is calculated under magnetic field at constant pressure.   Ma Mg Tt dT DM ¼ ¼ ð34Þ Q dH DS   Z Z T Ma Mg Tt H dT ¼ dH ð35Þ Q Tt H0 ¼ 0     Ma Mg Tt H Ma Mg H ¼ ð36Þ DTt ¼ Q DS DS represents the entropy change during the transition. For structural phase transition in cases a and g with the high temperature phase in both phases, the magnetization is written as follows for each phase: Mg;a ¼ χg;a H

ð37Þ

wherein χ is the magnetic susceptibility, Tt variation is given as follows:     χa χg Tt H2 χa χg H2 ¼ DTt ¼ 2Q 2DS

4.13.2.6

ð38Þ

Hysteresis Effect

The observed temperature and magnetic field hysteresis during the first-order structural and magnetic phase transition can cause problems while defining systems constitute. If Gibbs free energy of the two structural phases is written from Eq. (28a,b), the effect on the phase transition of the temperature and magnetic field can be found. Energy potential governing the structural phase transitions is shown in Fig. 3. One of the energy potentials of the dual phase shows the high temperature and the other shows the low-temperature phase. According to Eq. (28a,b), the total system forms a potential well while moving from one structural phase (alpha phase) to another structural phase (gamma phase). The resulting moves toward the phase system from one phase to another can cause an experience a force outside the potential wells system and arrange the energy to overcome this difficulty. The internal energies of the alpha and gamma phases are different, so for moving the gamma phase to the alpha phase (vice versa) the necessary energy that is taken from outside is different. Thus, the hysteresis effect is caused by the phenomenon described above.

Energy potential

Alpha phase

Process A

Gamma phase

Ordering parameter (A)

(B)

Process B

Heating without magnetic field +

Heating without magnetic field +

Apply magnetic field

Remove magnetic field

(C)

Fig. 3 Schematic representation of dual energy potential well for (A) equilibrium, (B) and (C) non-equilibrium states. Adapted from Sasso CP, Küperling M, Giudici L, Basso V, Pasquale M. Direct measurements of the entropy change and its history dependence in Ni–Mn–Ga alloys. J Appl Phys 2008;103:07B306.

580

Magnetic Energy Conversion

4.13.2.7

Determination of Magnetocaloric Effect

MCE can be calculated with two methods: direct measurement technique and indirect measurement technique.

4.13.2.7.1

Direct measurement technique

Providing direct measurements to characterize in detail the structure of the MCM is used for materials under first-order phase transition. Direct measurements are measurements that require a special experimental setup. The adiabatic and isothermal calorimeters that will be used in the measurements are produced from the specific geometry and appropriate materials. Then the system can be placed in the electromagnet and measurements are made under not very high field intensity. The value of each field is applied and the changing temperature of the alloy is measured by the calorimeter. DTad ¼ TH2

4.13.2.7.2

ð39Þ

TH1

Indirect measurement technique

Normal MCE is observed during the reversible second-order paramagnetic–ferromagnetic phase transition. The system total entropy is a continuous function of the temperature. The equations written in thermodynamic equilibrium condition are used to assess the MCE in this system. The entropy change is calculated experimentally, so indirectly. During calculations heat capacity or magnetization curves are used. One of the experimental measurements of heat capacity, as a function of temperature, when external magnetic field is applied at different values, allows the calculation of both DTad(T)DH and DSM(T)DH. The magnetization measurements allow only the calculation of the DSM(T)DH. Yet, during magnetization measurements, if the heat capacity of the material value of the constant magnetic field is known, DTad(T)DH can be calculated using Eqs. (21)–(23); [13]. The Maxwell equations are used to characterize the MCM. From the cycles of isothermal magnetization M(T, H), the entropy change is calculated from Maxwell’s integral Eq. (40):     ∂S ∂M ¼ ð40Þ ∂H T;p ∂T H;p Isothermal magnetic entropy changes are based on numerical integration of experimental data. Magnetization of finite growth, put in place at the same time as the differential temperature and magnetic field value, entropy change is calculated as follows:     X MðTiþ1 ; Hj Þ MðTi ; Hj Þ DSTi þTiþ1   ¼ DHj ð41Þ   T T 2 iþ1

j

i

Ti and Ti þ 1 temperatures in a magnetic field of H and the magnetization values are M(Ti þ 1,H) and M(Ti,H). By using Eqs. (21), (22), and (23), the temperature change can be calculated from magnetization curves.    Z H1  T ∂MðT; HÞ dH DTad ðT; DHÞ ¼ Cp ðT; HÞ H ∂T H2 H

ð42Þ

As can be seen from Eq. (42), by only using the magnetization data the temperature change DTad cannot be calculated. To calculate the temperature change, heat capacity value of the alloy must be known. Analytic integration of Eq. (42) is difficult and complex, because the external magnetic field and temperature values during measurement are dependent on the heat capacity [7]. By rearranging Eq. (42), the adiabatic temperature change is calculated as follows: DTad ¼

T DSM ðT; HÞ Cp;H ðT; HÞ

ð43Þ

DSM is calculated from the Maxwell equations from (41).

4.13.2.8

Calculation of Magnetocaloric Effect from the Heat Capacity Measurement

The entropy change on the MCE, the heat capacity, is calculated according to the second law of thermodynamics as follows: Cp ðH; TÞ dT ð44Þ T The whole characterization of MCM is ensured by measuring the function of the heat capacity in the constant magnetic field and pressure as a function of temperature. If two fixed magnetic fields H1 and H2 are (H2 4 H1 and H1 is usually equal to zero) between T1 and T2 temperature (usually T2 4 T1 and T1 close to zero), the measured value Cp(H,T), the system entropy change in the continuous function of temperature, any T temperature (the temperature T between T1 and T2 temperature) and constant H1 and H2, the total entropy of the system is calculated as follows: Z T CðTÞH1 ;p SðTÞH1 ;p ¼ dT ð45Þ T T1 -0 dSðTÞH;p ¼

SðTÞH2 ;p ¼

Z

T

T1 -0

CðTÞH2 ;p T

dT

ð46Þ

Magnetic Energy Conversion

2.0

581

Gd

−ΔSw (J/mol K)

1.5

1.0

0.5

0.0

180

200

220

240 260 280 300 Temperature (K)

320

340

Fig. 4 The magnetic entropy change calculated from the heat capacity and magnetization curves of Gd single crystal. Adapted from Dan’kov SY, Tishin AM, Pecharsky VK, Gschneidner KA Jr. Magnetic phase transitions and the magnetothermal properties of gadolinium Phys Rev B 1998;57:3478–90.

According to the third law of thermodynamics, as the temperature approaches absolute zero, magnetic entropy change in value is considered to be zero and independent of the magnetic field. Here the magnetic entropy change DSM(T)DH, T¼ 0K and T is written as follows: h i Z T CðTÞH CðTÞ h i h i H 2 1 p dT ð47Þ ¼ ¼ SðTÞH2 SðTÞH1 DSM ðTÞT;DH;P ¼ SM ðTÞH2 SM ðTÞH1 T;p T;p T 0 Adiabatic temperature change, the temperature value, and the entropy value observed between the two different magnetic fields are obtained as follows [13,14]: DTad ðTÞDH ¼ TðS0 ; H2 Þ

TðS0 ; H1 Þ

ð48Þ

Gd single-crystal samples [15] resulting in the measurements made on the magnetic entropy change are shown in Fig. 4. The accuracy of the determination of the MCE of heat capacity depends on the accuracy of the heat capacity data. Small differences between DSM and DTad come from measurements made at a higher temperature value than 0K.

4.13.2.9

Calculation of the Magnetic Entropy Change by Using the Clausius–Clapeyron Equation

The Clausius–Clapeyron equation is used in the first-order phase transitions. The following equation is used instead of the derivative of the Maxwell equations. dH ¼ dTt

DS DM

ð49Þ

DM and DS are known as the change of the magnetization and the entropy between the two phases. dH/dTt is the amount of change at Tt at the transition temperature with magnetic field. The entropy change value found here is known to be very consistent with the values found in the Maxwell equations when magnetic materials are under the external magnetic field. From the Clausius–Clapeyron equation, the equivalent thermal magnetic entropy change during the phase transition is calculated. According to the field-induced entropy changes, the magnetization change associated with the first-order phase transition and ∂Tt/∂H calculation is carried out. The Clausius–Clapeyron equation in calculating magnetic entropy change will be discussed for the case of MCMs undergoing a first-order magnetic phase transition with large magnetic thermal hysteresis. In calculating the magnetic entropy change of MCMs, the Clausius–Clapeyron is related to the first-order phase shift with hysteresis. It is understood that the measurements of isothermal magnetization must be carried out in some way. They say that a single-phase transition from the ferromagnetic state to the paramagnetic state explores temperature and field dependence. In this way, the change of magnetic entropy can be reliably determined from the magnetization data through the Maxwell’s relationships.

4.13.3

Analysis and Assessment

4.13.3.1

Magnetocaloric Materials

MCMs are one of the most important parts of magnetic refrigerators. Because of this importance in magnetic refrigeration technology, MCMs should have superior magnetocaloric properties. The best candidate MCMs for room temperature application should have the following properties:

582

• • • • • •

Magnetic Energy Conversion

GMCE around room temperature, showing the magnetocaloric properties in a wide temperature range, tunable to the phase transition temperature, low thermal and magnetic hysteresis near phase transition temperature, having high electrical resistivity and thermal conductivity properties, having easily producibility properties. In this section, an overview for major classes of MCMs is given.

4.13.3.1.1

Gadolinium (Gd)

Gd is the most well-known MCM and used widely for many room temperature magnetic refrigerators. Gd crystallizes in hexagonal structure (P63/mmc). Its unit cell parameters are a ¼ b¼ 3.63 Å and c¼ 5.78 Å at room temperature. Gd has a second-order phase transition from paramagnetism to ferromagnetism with decreasing temperature at TC ¼ 294K [15]. Its magnetocaloric properties are summarized in Table 1 and are quite good at room temperature. However, the adiabatic temperature change values for polycrystalline Gd under different magnetic field changes are exhibited in Fig. 5. It is possible to write an equation to estimate the DTad value under different magnetic field change for Gd. Compiling many experimental data from very pure Gd samples yields the equation DTad (K) ¼ 3.675(m0H[T])0.7. This equation is in good accordance with the mean field theory (MFT) calculation result [20]. On the other hand, Bahl et al. find this relation as DTad (K) ¼ 2.85(m0H[T])0.78 for medium purity (1N) Gd sheets [17]. Hence, the magnetocaloric properties of Gd depends on its purity and homogeneity. The thermal conductivity (κ) of Gd (purity 99.7%) is 8.0 W/K m at 300K [21]. Consequently, Gd shows second-order phase transition at near room temperature and exhibits fairly good magnetocaloric properties because it has no thermal and magnetic hysteresis. It shows good thermal transport properties. Gd can conventionally Table 1

Magnetic and magnetocaloric properties of Gd

Material

TC (K)

m0DH (T)

Polycrystalline Gd (Purity 99.7%)

294

Sheet Gd (Purity 99.4%) Polycrystalline Gd Polycrystalline Gd (Commercial Purity)

295 295 294

1 2 1.1 1 1

DSMax (J/kg K) 3.2 6.1 – 2.8 –

Max ðKÞ DTad

Reference

2.9 5.2 3.1 2.6 2.7

[16] [17] [18] [19]

Note: TC, curie temperature; m0DH, magnetic field change value; DSMax.: entropy change value; DTadMax: adiabatic temperature change value.

20 oΔH=2 T, ΔT max. ad =5.8K max.=11.5K oΔH=5 T, ΔT ad

16

max.=15.3K oΔH=7.5 T, ΔT ad

oΔH=10 T, ΔT max. =18.7K ad

ΔTad (K)

12

8

4

0 200

230

260

290

320

350

Temperature (K) Fig. 5 The DTad (T) curves for Gd determined from C(T) curves. Adapted from Bahl CRH, Nielsen KK. The effect of demagnetization on the magnetocaloric properties of gadolinium. J Appl Phys 2009;105:013916.

Magnetic Energy Conversion

583

fabricate in different shapes such as sphere and plate forms. Despite its high price, its excellent mechanical, magnetic, magnetocaloric, thermal, and fabrication properties make it a suitable reference MCM to compare the performance of other MCMs and magnetic refrigerators with each other. As observed clearly in Fig. 5, the highest MCE in the element Gd is near the temperature at which this element passes through the magnetic phase. In the vicinity of this transition temperature, Gd element transitions from an irregular magnetic structure to a regular magnetic structure. This temperature is given as Curie temperature for Gd element. Moreover, according to the adiabatic temperature change measurements made near room temperature, a temperature difference of 18.7K can be created in the element of Gd approximately at the external magnetic field change of 10 T. This material has a magnetic phase transition especially near room temperature, and besides, this temperature change under the magnetic field makes this material and its derivatives very popular.

4.13.3.1.2

La(Fe–Si)13 based alloys

La(Fe1 xSix)13 based alloys are commonly used as MCMs in magnetic refrigerators instead of expensive Gd or Gd based alloys (such as Gd–R (R: Y, Dy, Tb, Er) alloys). These alloys exhibit the first-order phase transition from the ferromagnetism to paramagnetism at their TC with increasing temperature. While the LaFe13alloy does not crystallize in NaZn13-type cubic structure, the La(Fe1 xSix)13 based alloys can crystallize in this cubic structure (space group Fm3̅c ) by substitution of Fe for Si with the concentration ranging from x ¼0.81 to x ¼ 0.89 [22]. In the cubic NaZn13-type structure, La and FeI atoms occupy in 8a (1/4,1/ 4,1/4) and 8b (0,0,0) sites, respectively [23]. 12 FeII(Si and Fe) atoms are randomly distributed among 96i (0,y,z) sites [24] (Fig. 6). In the paramagnetic state near TC, the La(Fe1 xSix)13 alloys show an itinerant-electron metamagnetic phase transition (that is field-induced first-order transition). This phase transition leads to a large MCE around 200K [25]. In order to go up their TC to higher temperature, the researchers substitute the 3d transition metal (Co and Mn) for Fe or rare-earth element (Ce,Pr and Nd) in various La(FexSi1 x)13 based alloys [26–28]. The substitution of Co for Fe causes an increase in the Tc up to above room temperature from 200K, but their phase transition type changes to second order. Therefore, the entropy change and adiabatic temperature change values are decreased to lower values as seen in Fig. 7 [29]. Adding hydrogen to La(Fe–Si)13 alloys is a well-known way to shift their Curie temperature from 190 to 340K. With hydrogen insertion to these alloys, the unit cell volume of 1:13 phase is expanded and the magnetic interaction between Fe atoms is modified. Thus, the Curie temperature of La(Fe–Si)13Hn (0ono2.4) alloys increases up to 350K while the first-order transition remains and they exhibit giant MCE [30–33]. For powders of La(Fe–Si)13 alloys, the hydrogenation process is conducted around 500K in an atmosphere of helium and hydrogen for several hours [33]. Nevertheless, La(Fe–Si)13Hn alloys exhibit some decomposition problems at their Curie temperature. Because of these, by substitution of Co and Mn for Fe or Ce, Pr and Nd for La, the La1 yRy(Fe1 xTx Si)13(R:Ce, Pr or Nd; T:Mn or Co) alloys shift to high temperature and then the hydrogenation process is applied to these alloys [33]. Having shown good stability and giant magnetocaloric properties of La(Fe–Mn–Si)13Hn alloys, these alloys are the best candidates for MCM for application. The magnetic entropy change values of LaFe11.74MnySi1.76H1.53 (y¼ 0.322, 0.338, 0.356, 0.373, and 0.390) are shown in Fig. 8 for moDH¼ 1.6 T. These alloys are known as CALORIVAC-H and are commercially sold by Vacuumschmelze GmbH & Co. The La(Fe–Si)13 based alloys exhibit very good magnetocaloric properties (large MCE with small thermal and magnetic hysteresis during the phase transition process). These alloys show large MCE when compared with Gd (Table 2). The cost of raw materials for these alloys is cheap and these alloys contain no critical elements. The production technology of these alloys is available for large-scale production. The good magnetocaloric properties and reproducible properties of these alloys make them ideal candidates for magnetic refrigeration technology.

Fe′

La Fe″

Fig. 6 The crystal structure of La(Fe1 xSix)13 based alloys. Adapted from Wang W, Huang R, Li W, et al. Zero thermal expansion in NaZn13-type La(Fe,Si)13 compounds, Phys Chem Chem Phys 2015;17: 2352.

584

Magnetic Energy Conversion

25 LaFe11.5Si1.5

Δ0H = 1.9 T

LaFe11.40Co0.52Si1.09

6

20

LaFe11.05Co0.91Si1.04 ΔTad (K)

2

LaFe10.71Co1.30Si1.00 10

ΔS (J/kg K)

15

2 5

0 160

200

240

280

0 360

320

T (K) Fig. 7 Temperature dependence of entropy change and adiabatic temperature change values of La(Fe–Co–Si)13 alloys for moDH ¼1.9 T. Adapted from Skokov KP, Yu. Karpenkov A, Yu. Karpenkov D, Gutfleisch O, The maximal cooling power of magnetic and thermoelectric refrigerators with La(FeCoSi)13 alloys. J Appl Phys 2013;113:17A945.

14 y=0.356 12

y=0.338

y=0.373

10 ΔS (J/kg K)

y=0.390

y=0.322

8

6

4

2

0 270

280

290

300

310

T (K) Fig. 8 Temperature dependence of entropy change values of LaFe11.74MnySi1.76H1.53 (y ¼0.322, 0.338, 0.356, 0.373, and 0.390) alloys for a field change of 1.6 T. Adapted from Barcza A, Katter M, Zellmann V, et al. Stability and magnetocaloric properties of sintered La(Fe, Mn, Si)13Hz Alloys. IEEE Trans Magn 2011;47:3391

4.13.3.1.3

(Mn–Fe)2–(P–Z) based alloys

The Fe2P alloy exhibits a first-order magnetoelastic phase transition around 215K with a small MCE. On the other hand, the (Mn–Fe)2-(P–Z) alloys exhibit a first-order magnetoelastic phase transition at room temperature with a large MCE. The (Mn–Fe)2–(P–Z) based alloys (Z: p-element such as B, Si, Ge, As) crystallize the Fe2P type hexagonal structure (space group P6̅2 m) (as seen in Fig. 9) [39]. There are sudden changes in the parameters a and c (∆a/a and ∆c/c, respectively) during first-order magnetoelastic transition for the (Mn–Fe)2–(P–Z) based alloys. Due to the large changes of ∆a/a and ∆c/c, their unit cell volumes hardly change [40]. These alloys display a giant MCE, because the changes in crystal lattice enhance the magnetocaloric properties.

Magnetic Energy Conversion Table 2

585

The magnetic and magnetocaloric properties of some La(Fe–Si)13 based alloys

Material

TC (K)

m0DH (T)

LaFe11.7Si1.3 LaFe11.6Si1.4 LaFe11.18Si1.82 LaFe10.96Co0.97Si1.07 LaFe10.71Co1.30Si1.00 LaFe11.7Si1.3H1.53 LaFe11.6Si1.4H1.6 LaFe11.048 Mn0.322Si1.26H1.53 La0.7Pr0.3Fe11.5Si1.5C0.2H1.2

184 196 215 289 325 287 333 296 325

2 2 2 1 1.9 2 2 1.6 2

DSMax. (J/kg K) 28 24 9.5 5.3 28 29 15 11.6 17.1

Max ðKÞ DTad

Reference

8.2 7.8 3.5 2.2 11.5 7.5 3.8 4 –

[30] [34,35] [30] [36] [17] [30] [35,37] [33] [38]

Mn 3g (x,0,1/2) Fe 3f (x,0,0) P/As 1b (0,0,1/2) P/As 2c (1/3,2/3,0)

Fig. 9 The crystal structure of the (Mn–Fe)2–(P–Z) alloys.

y=0.78 y=0.76

120

y=0.76

y=0.80 y=0.78 2T

y=0.70 2T

15

2T

y=0.70 60

10 2T

1T

30

5 1T

1T

1T 0 120 (A)

20

ΔS (J kg K)

M (emu/g)

90

y=0.80

0 160

200

240 T (K)

280

320 (B)

220

240

260

280

300

T (K)

Fig. 10 (A) The M(T) curves of Mn2 yFeyP0.75Ge0.25 melt-spun ribbons for moH ¼0.5 T. (B) The DS(T) curves for moDH¼1 T and 2 T. Adapted from Trung NT, Klaasse JCP, Tegus O, et al. Determination of adiabatic temperature change in MnFe(P,Ge) compounds with pulse-field method. J Phys D: Appl Phys 2010;43:015002.

The Curie temperature of MnFeP1 xAsx alloys increases with increasing As content from 150K to above 300K, while their total magnetic moment values stay constant. The MnFeP0.45As0.55 alloy prepared by solid-state reactions shows magnetic phase transition from paramagnetism to ferromagnetism at TC ¼293K with decreasing temperature. Its magnetic entropy change and adiabatic temperature change values are 15 J/kg K (moDH¼2 T) and 4K (moDH¼ 1.45 T), respectively [41]. The MnFeP1 xAsx alloys are promising candidates for magnetic refrigeration, exhibiting large MCE with small thermal hysteresis (less than 2K). Unfortunately, the toxicity of As hinders the use of these alloys for applications. To dispose of toxic As from these giant MCE materials, various attempts have been tried by researchers. Brück and his research group members have developed (Mn–Fe)2(P–Z) with Z: Si and Ge [42–44]. In Fig. 10, the M(T) and DS(T) curves of Mn2 yFeyP0.75Ge0.25 melt-spun ribbons are shown. Their phase transition temperatures can also be changed by doping of B, C, and N atoms. The average thermal hysteresis of these alloys decreases by boron doping [45]. The substitution of Z for P in (Mn–Fe)2(P–Z) alloys is caused by changing in the electronic configuration and the mixed magnetism that is responsible for the good magnetocaloric properties in these compounds. Table 3 shows the magnetic and magnetocaloric properties of some (Mn–Fe)2(P–Z) alloys. For practical application, the tunability of Curie temperature

586

Magnetic Energy Conversion

Table 3

The magnetic and magnetocaloric properties of some (Mn–Fe)2(P–Z) alloys

Material

TC (K)

m0DH (T)

MnFeP0.45As0.55 Mn1.1Fe0.9P0.47As0.53 Mn1.1Fe0.9P0.75Ge0.25 Mn1.25Fe0.7P0.51Si0.49 Mn1.3Fe0.675Co0.025P0.46Si0.54 Mn1.2Fe0.7Ru0.1P0.5Si0.5

306 294 285 281 280 289

MnFe0.95P0.595B0.075Si0.33

282

1 1 1 1 1.5 1 2 1

DSMax.(J/kg K) 13 12 12 10 7.5 6.5 13.5 10

Max ðKÞ DTad

Reference

4(m0DH¼1.45 T) 4.2(m0DH¼ 1.45 T) 3 1.9 2 2.1 4.3 2.7

[41] [41] [46] [47] [48] [49] [50]

10 THot = 281K, U = 0.5 THot = 284K, U = 0.5

Temperature span (K)

8

THot = 284K, U = 0.4

6

4

2

0 0

10 20 Cooling power (W/kg)

30

Fig. 11 The dependence of temperature span on cooling power for 1 T magnetic refrigerator for parallel plate LCMS and U ¼0.4 and 0.5. U is the utilization factor. THot shows temperature of the hot end. Adapted from Bahl CRH, Velaź quez D, Nielsen KK, et al. High performance magnetocaloric perovskites for magnetic refrigeration. Appl Phys Lett 2012;100:121905.

by changing of Mn and Fe or P and Si or B composition, no toxic and critical elements, large MCE (high DS and high DTad), small thermal hysteresis, and good mechanical stability make these alloys highly promising for magnetic refrigeration.

4.13.3.1.4

Manganites

Although the manganites have been known since 1960s, their interesting magnetocaloric properties were only discovered in 1996 [50]. The general formula of manganite is R1 xMxMn1 yTyO3. In this formula, R is trivalent rare-earth elements (for example, La, Pr, Nd, Sm, Eu, Gd, Ho, Tb, Y), M is divalent alkaline earth ions (Sr, Ca, Ba, Pb, Na1 þ , K1 þ ,Ag1 þ ), and T is 3d metals (Fe or Co) [51]. In manganites, the colossal magnetoresistive (CMR) effect is observed around phase transition temperature. Since there exists a definite relation between the resistivity and magnetic entropy [52], these alloys show large MCE at Curie temperature. The most promising alloys among manganites for magnetic refrigeration are the La0.67(Ca,Sr)0.33MnO3 (called as LCSM) alloys. LCMS alloys showed no hysteresis and excellent chemical stability [53]. Bahl et al. used the LCMS alloys (La0.67Ca0.2925Sr0.0375MnO3 and La0.67Ca0.2850Sr0.0450MnO3) to find the magnetic refrigeration performance of these alloys in their magnetic refrigerator. They found that the maximum cooling power for zero-span is 35 W/kg (as seen in Fig. 11) and this is greater than that of Gd plates (16 W/kg) under similar conditions [53]. Magnetic and magnetocaloric properties of some manganites are shown in Table 4. The adiabatic temperature change values of manganites are relatively low when compared to other MCMs, for example Gd, La (Fe–Si)13 based and (Mn–Fe)2(P–Z) based alloys, but their entropy change values are comparable with that of Gd. The main advantages of manganites over other MCMs are low cost, no oxidation, high electric resistance, and minimum Eddy current loss. The Curie temperatures of these alloys can be changed by substitution of other elements. However, they can be produced in fine structures that lead to high performance in magnetic refrigeration. These advantages make manganites promising MCMs for magnetic refrigeration technology.

4.13.3.1.5

XMnGe (X: Co or Ni) based alloys

In the first-order phase transition materials, there is a strong coupling between their structure and magnetism. Therefore, a magnetic field can induce a change of their lattice and magnetic entropies. For that reason, the first-order phase transition materials

Magnetic Energy Conversion Table 4

587

The magnetocaloric properties of some manganites

Material

TC (K)

m0DH (T)

La0.67Ca0.33MnO3 La0.67Ca0.2925Sr0.0375MnO3 La0.67Ca0.2850Sr0.0450MnO3 La0.67Sr0.33MnO3 Pr0.67Sr0.33MnO3 La0.67Ba0.33MnO3 (La0.7Sm0.3)0.67Pb0.33MnO3

267 275 282 369 300 337 280

1.2 1 1 1.2 1 1 3

DSMax.(J/kg K) 5.9 3.7 3.5 1.8 8.5(m0DH ¼5 T) 2.70 2.6

Ge

X X

Max DTad ðKÞ

Reference

2.0 1.3 1.2 0.93 1.5 – 1.3

[54] [53] [53] [54] [55] [56] [57]

Ge

Cooling Mn

Mn

Heating TCr.

Orthorhombic (Pnma)

Hexagonal (P63/mmm)

Fig. 12 The crystal structures in the XMnGe (X: Co and Ni) based alloys. TCr represents the structural phase transition temperature.

exhibit giant magnetocaloric properties. The researchers have proposed the XMnGe (X: Co and Ni) based alloys that show first order phase transition around room temperature for a new family of MCMs. The stoichiometric CoMnGe and NiMnGe alloys undergo a structural phase transition at 610 and 470K with decreasing temperature, respectively [58–60]. The Curie temperature of CoMnGe alloy is TC ¼ 345K, while for NiMnGe alloy the antiferromagnetic phase transition occurs at TN ¼ 345K [59,60]. In the XMnGe (X: Co and Ni) based alloys, the structural phase transition occurs from Ni2In-type hexagonal structure to TiNiSi-type orthorhombic structure with decreasing temperature (as seen in Fig. 12). Since the structural and magnetic properties of these alloys are very sensitive to interatomic distance, there are many methods to decrease the structural phase transition temperature to lower temperatures. These methods include (1) doping interstitial atoms (such as XMnGeBy or XMnGeCy); (2) partial substitution such as XMn1 yTyGe, X1 yTyMnGe and XMnGe1 yZy (Z: Al, Si, Ga, In, Sn), and (XMnGe)1 y(Fe2Ge)y; (3) vacancy on X or Mn sublattices such as X1 yMnGe and XMn1 yGe; and (4) applying hydrostatic pressure [61–70]. The decrease in structural phase transition temperature in these alloys and the overlap of structural and magnetic phase transitions could occur. Thus it is possible to observe a first-order magnetostructural phase transition from paramagnetic hexagonal to ferromagnetic orthorhombic states with a large magnetization difference and accordingly MCE. According to Gschneidner et al. [71], the large volume change (DV/V) plays an important role for observing the large MCE. The CoMnGe0.95Ga0.05 alloy coexists with the structural and magnetic phase transformations around 310K and this alloy shows the DV/V value of 3.9% (as seen in Fig. 13) [62]. This DV/V value is comparable to other MCMs such as MnAs (DV/VB2.2%), LaFe11.2Co0.7Si1.1(DV/VB1.3%), FeRh(DV/VB0.9%), and Gd5Si1.8Ge2.2(DV/VB0.4%) [70]. The magnetocaloric properties of some of the XMnGe (X: Co and Ni) based alloys are summarized in Table 5. Generally, large magnetic entropy changes have been observed in the XMnGe (X: Co and Ni) based alloys. The Ge-free XMn(Al, Si) alloys show large MCE near room temperature [78]. Since these alloys are produced from cheap elements, they are economically attractive for applications. In conclusion, the magnetocaloric properties of MCMs are observed in narrow temperature range. For the application, they should have giant MCE in wide temperature range. The layered regenerators could improve the large temperature span by using various MCMs with different TCs to enhance the MCE along the active magnetic regenerator (AMR) bed (as seen in Fig. 14). Astronautics has built a rotary magnetic refrigerator with 2 kW of cooling power for DTSpan ¼12K by using seven layered magnetocaloric (LaFeSiH) AMR [79]. Its COP is higher than 2. In this section we gave a brief summary of well-known and commonly used MCMs. The magnetocaloric properties of wellknown MCMs are summarized in Table 6. The maximum specific cooling power (q_ Max ) of the MCMs are estimated by the following equation [81,82]: q_ Max ¼ TC
DS
f þ D

ð2
TC þ DTad Þ
DS
f 2

ð50Þ

here, TC is Curie temperature of the MCM, DS is magnetic entropy change at TC, DTad is adiabatic temperature change at TC, and f is operation frequency of the refrigeration. The average magnetic field value of the permanent magnets is approximately 1 T and the

Magnetic Energy Conversion

588

100 (A)

(B) Yobs 75

10

Ycalc

Hex. for cooling

Bragg pos

Hex. for heating

Yobs−Ycalc

Ort. for cooling

50

Ort. for heating 25

0

16

18

20 2θ (°)

22

24

0 162

40 (D)

(C)

160

Hex. for cooling

20

5T 4T

Ort. for cooling

10

3T 2T 1T

158

Hex. for heating 78

Volume (Å3)

7T 6T

30 −ΔS (J/kg.K)

Phase fraction (%)

Intensity (×103 counts)

20

Ort. for heating 77

76

0 280

300

320

360

T (K)

0

100

200 T (K)

300

400

Fig. 13 (A) The Rietveld refinement results of CoMnGe0.95Ga0.05 at 315K. (B) The phase fractions of the hexagonal and orthorhombic structures for CoMnGe0.95Ga0.05. (C) The volumes of the hexagonal and orthorhombic structures for CoMnGe0.95Ga0.05. (D) The entropy change values of CoMnGe0.95Ga0.05. Adapted from Dincer I, Yuzuak E, Durak G, et al. J Alloys Comp, 2012;540:236. Table 5

The magnetocaloric properties of some XMnGe (X: Co and Ni) based alloys

Material

TC (K)

m0DH (T)

Co0.95MnGe0.97 CoMnGeB0.02 CoMn0.97Cr0.03Ge CoMn0.9Fe0.1Ge CoMnGe0.945Ga0.55 CoMnGe0.99In0.01 (NiMnGe)0.9(NiCoGe)0.1 (CoMnGe)0.89(NiCoGe)0.11 (NiMnSi)0.65(Fe2Ge)0.35

282 287 322 299 310 305 236 292 262

1.9 2 2 2 2 DP ¼3kPa 2 2 2

DSMax.(J/kg K) – 20 12 15 13 52 15 10 15.2

Max DTad ðKÞ

Reference

1.5 – – 8.5 6.5 18.5 – – –

[72] [73] [74] [75] [76] [70] [71] [76] [77]

average operation frequency is about 1 Hz. According to the results in Table 6, the La(Fe–Si)13 based alloys give the best specific cooling power for 1 T and 1 Hz.

4.13.4 4.13.4.1

Case Studies Magnetic Energy Conversion

As a broad application prospective green renewable new energy technology, the magnetic and the electromagnetic energy conversion technology has been studied by more and more countries all over the world. These studies are mainly focused on the design and the application of magnetic and electromagnetic energy conversion devices. This issue briefly introduces the current

Temperature

Magnetic Energy Conversion

589

When magnetic field applied

ΔT1 ΔT2 ΔT3

ΔTspan

ΔT4 ΔT5 When removing magnetic field

ΔT6 ΔT7

Magnetocaloric materials

Hot end

TC1 TC2 TC3 TC4 TC5 TC6 TC7

Location

Cold end

AMR BED Fig. 14 Schematic diagram of a layered AMR bed and the temperature span between the ends of the AMR bed.

Table 6 f¼1 Hz

The magnetocaloric properties and maximum specific cooling powers of some magnetocaloric materials (MCMs), for m0DH¼1 T and

Material

TC (K)

Gd CALORIVAC-H (La(Fe–Mn–Si)13Hz) CALORIVAC-C (La(Fe–Co–Si)13) Mn1.25Fe0.7P0.51Si0.49 CoMn0.9Fe0.1Ge La0.67Ca0.2850Sr0.0450MnO3

294 300 300 281 299 282

DSMax. (J/kg K) 3.2 12.4 4.6 10 7 3.5

Max DTad ðKÞ

q_ Max (W/kg)

Reference

2.9 2.8 1.6 1.9 4.1 1.2

945 3737 1387 2820 2107 989

[16] [80] [80] [47] [75] [53]

research status of magnetic and electromagnetic energy conversion technology, and deeply analyzes the magnetic refrigeration devices. Through referencing the achievement of experimental research, theoretical analysis, and calculations, this issue is specific on the following aspects of research work. Processes of energy conversion are omnipresent in our lives, and electricity is the most important form of energy as we use it every day. In addition to macroscopic energy sources such as sun, wind, and water, more and more research activities have been engaged in discovering techniques for converting ambient energy sources available around us to realize energy conversion. The development of energy conversion techniques for ambient energy sources indeed led to fundamental insights and applications for materials and devices. Energy conversion is a process of converting one form of energy into another; in particular, the conversion of magnetic, electromagnetic, and mechanical energy into thermal energy, and to reveal how this performance is influenced by microstructural features tailored using carefully selected areas such as magnetic refrigeration in green technology branches, and heat treatment in medical branches. As an example, one significance of energy conversion is that sustainable and ambient energy sources around us can be used to power consumer and medical electronics. For example, pacemakers are crucial to patients with cardiac arrhythmia. Even though the battery life of pacemakers has improved, the battery still needs to be replaced every 5–10 years. However, if portable energy conversion can be applied to power implantable biomedical devices such as pacemakers, the battery life can be further extended. Therefore, research related to energy conversion from omnipresent ambient energy sources has gained broad research interest in the hope of exploring feasible energy sources for portable energy conversion. Ambient sources such as heat dissipated from passengers, and vibration generated due to daily motions or sound waves, have been applied as energy sources for energy conversion. In addition, it has been proposed that the movements from our daily activities show great potential to power personal electronics. For example, arm motions, footfalls, and blood pressure are estimated to generate power of 0.33 W, 67 W, and 0.37 W, respectively. Various ambient energy sources are indeed around us, but the question is how to utilize these sources to realize energy conversion.

Magnetic Energy Conversion

Entropy, S

0ΔH=0 T0

0ΔH>0

Heat load 4

ΔS

N

1

2

Heat rejection

ΔTad

T0−ΔTad

T0

N

3

T0−ΔTad

S

T0+ΔTad

T0

S

590

T0+ΔTad

Temperature, T

Fig. 15 The schematic representation of S–T diagram for the Brayton cycle. Steps 1 and 3 are the adiabatic processes. Steps 2 and 4 are isofield processes.

There are two main references to determine the performance of the energy conversion systems. In order to identify how magnetic refrigerators and heat treatments impact the selected alloys energy conversion performance, an established thermodynamic structure refrigerant cooling (RC) was utilized with the thermodynamic relations defining first-/second-order martensitic phase transitions. Significant research efforts have been dedicated to energy conversion mechanisms for exploring magnetic materials. Mechanisms based on magnetic materials have been widely used for energy conversion from ambient energy sources. The efficiency of a heat exchanger is called its the coefficient of performance or COP. It is defined in the same way as the efficiency of an energy conversion device: COP ¼

Energy output Energy input

ð51Þ

There is a significant difference between energy transfer devices and energy conversion devices. Only a part of the energy input is acquired as utility energy output, and an efficiency is a number between zero and one in a conversion device. The utility energy output is the amount of heat removed from the TL (temperature low), or the amount of heat removed from the TH (temperature high) in a transfer device. It is not exactly the part of the energy input or output. In fact, the output of useful energy can exceed energy. Whence, heat pumps can be quite attractive for 3D space heating/cooling purposes. Consequently, the performance coefficient (energy efficiency ratio) can be in multiple, and it means that does not disrupt the first law. In a more scientific way of writing, the term “work” will be more productive as the height increases. If we do not need to reenter the thermodynamic analysis, it is feasible to define the maximum the coefficient of performance of any system (COPmax). The definition relies on the use of the heat exchanger as a heater or a cooler. If it is a heater or cooler, the definition is as follows: COP ¼

TH TL or COP ¼ TH TL TL TH

ð52Þ

In the case of ideal efficiency (maximum performance coefficient), these defined temperature units must be expressed in absolute units (K). Over all for an equated refrigerator without freezer sight, the mean cooling power is required in the range of 50–150 W. The COP of standard vapor compression systems is in the range of 1–2.5. The lowest temperature in the cooling sight is around 51C, as the room temperature varies from 20 to 351C.

4.13.4.2

Magnetic Cooling

The working principle of the magnetic refrigeration is analogous to that of commercial vapor compression refrigeration. Unlike commercial vapor compression refrigeration, the magnetic field is used instead of pressure. The magnetic refrigerator is composed of an AMR. The parts of an AMR are a MCM, a permanent magnet, hot/cold heat exchangers, and a heat transfer fluid. The magnetic refrigeration cycle is shown in Fig. 15. When the MCM at a temperature (T0 and just above TC) is adiabatically magnetized, the temperature of the MCM increases up to T0 þ DTad. In the second step, the heat transfer fluid is transferred from magnetized magnetocaloric to expel the heat from the MCE material. Its entropy and temperature goes down to lower temperature (T0). In the third step, the magnetic field is removed and the

Magnetic Energy Conversion Table 7

591

The properties of some reciprocating and rotary magnetic refrigerators

Magnetocaloric material

Magnet assembly

Type

f (Hz)

DTSpan(K)

q_ (W/kg)

Reference

Gd plates, 1 mm thick Gd spheres, 0.15–0.3 mm, 500 g Gd plates, 1 mm thick, 223 g Gd packed bed, 1167.4 g Gd particles, 0.6 mm, 62 g Gd plates, 0.9 mm thick Layered La–Fe–Co–Si plates, 1 mm thick, 150 g 7 Layered La–Fe–Co–Si plates, 0.5 mm thick Gd plates, 560 g Gd particles, 0.5 mm Gd plates, 516 g

SMA, SMA, PMA, PMA, PMA, PMA, PMA, PMA, PMA, PMA, PMA,

7T 1.5 T 0.8 T 1.5 T 1T 1.03 T 0.8 T 1.15 T 1.17 T 1.5 T 1.5 T

REC REC REC REC REC REC REC REC REC ROT ROT

– – 0.42 1 0.35 – 1 0.15–0.45 0.55 5 2

PMA, PMA, PMA, PMA, PMA,

1T 0.85 0.98 1.17 1.47

T T T T

ROT ROT ROT ROT ROT

1.3 3 4 6 1.4

Gd spheres, 0.25–0.8 mm, 2800 g

PMA, 1.24 T

ROT

Layered La–Fe–Si–H particles, 1520 g

PMA, 1.5 T

ROT

1.8 1.5 4

No-load 70 39.5 43.9 No-load No-load No-load 15 No-load No-load No-load 240 150 W 157.1 W 120 300 No-load 454.5 360.7 142.8 1375 2001

[83] [84] [85] [86] [87] [88] [89] [90] [91] [92] [93]

Gd or LaFeCoSi packed bed Gd or Gd alloy packed bed, 1273 g Gd and Gd-Tb Gd, GdEr, LaFeSi Gd spheres, 0.3 mm, 110 g

47 7 4 18.2 12 10.2 10.5 8 33 25 11.5 0 7.6 13.8 42 38 29 2.5 0.3 12.9 12 0

[82] [82] [94] [82] [95] [96] [79]

q_ , the specific cooling power; f, frequency of device; DTSpan, the temperature span between hot and cold ends. Abbreviations: SMA, superconducting magnet assembly; PMA, permanent magnet assembly; REC, reciprocating magnetic refrigerator; ROT, rotary magnetic refrigerator.

MCMs cool down to lower temperature below the starting temperature. At the last step, by moving the heat transfer fluid in the opposite direction, its temperature increases up to higher temperature. This thermodynamic (Brayton) cycle is shown in Fig. 15 and it is commonly used in many magnetic energy converted refrigerator prototypes. In 1976, the first room temperature magnetic refrigerator had been built by Brown [5]. In 1997, Pecharsky and Gschneidner discovered the Gd–Si–Ge based alloys as a new giant MCM. This discovery was a milestone and accelerated the development of room temperature magnetic refrigerators. After that, many room temperature magnetic refrigerators have been developed to date. The published magnetic refrigerators are classified in two groups: (1) reciprocating magnetic refrigerators and (2) rotating magnetic refrigerators. In reciprocating magnetic refrigerators, the magnet assembly moves linearly over the AMR assembly or vice versa. Their operating frequency is limited to 1 or 2 Hz and consequently their cooling powers are below 60 W/kg. Due to their low cooling power, reciprocating magnetic refrigerators are commonly used for testing devices. On the other hand, rotary magnetic refrigerators are more complex and advanced devices when compared with reciprocating magnetic refrigerators. The rotating magnetic refrigerators can operate in higher frequency as continuous mode. These magnetic refrigerators and some properties are summarized in Table 7. In magnetic refrigerators, the heat is periodically stored and transferred from/to MCM by using a heat transfer fluid. It is pumped through MCMs in an oscillatory. Thus, the heat is transferred to the hot end from the cold end by using MCE and thermodynamic cycle. After some completed thermodynamic cycles, the system reaches the steady state and the maximum temperature span between hot and cold ends can be established in the magnetic refrigerator. The heat transfer fluid is moving through the AMR beds by two different methods. One of them is the pump-valve system. In this method, the heat transfer fluid is transferred from hot/cold to cold/hot ends by a pump-valve as seen in Fig. 15 during different thermodynamic cycle steps. The other method is via a piston system to move heat transfer fluid during different thermodynamic cycle steps. Such a piston system is shown in Figs. 16 and 17. In particular, the linear displacements used in such systems consume very little energy. In this case, the total amount of energy consumed causes me to keep it at least. In addition, the maintenance and repair costs are very small due to the small amount of moving parts used in these systems. The magnetic refrigerators can be operated with different thermodynamic cycles to obtain high performance. There are four different processes for each thermodynamic cycle of a magnetic refrigerator. These are respectively: (1) magnetization, (2) fluid flow from cold end to hot end, (3) demagnetization, and (4) fluid flow from hot end to cold end. The duration of these four processes is different for the different thermodynamic cycles. The thermodynamic cycles are the Brayton, the Ericson, the Carnot, and the hybrid [97]. The Brayton cycle is shown in Fig. 18. In its S–T diagram, (1) is adiabatic magnetization, (2) is heat rejection from system to outside, (3) is adiabatic demagnetization, and (4) is heat absorption from the cooled part of the system. Here tMag. is the time of magnetization process, tCold is the time of fluid flow from cold end to hot end, tDemag. is the time of demagnetization process, and tHot is the time of fluid flow from hot end to cold end.

Magnetic Energy Conversion

592

Pump

Pump Valve-1

Valve-1

Valve-2

Valve-2 S

Q

Q

Q

Q

CHEX

CHEX

N

HHEX

Q

Q HHEX Q

Q N

(A)

S

(B)

Fig. 16 Schematic showing a magnetic refrigerator with two AMR beds and pump-valve methods for (A) cycle-1 and (B) cycle-2.

Q

Q

CHEX

S

AMR-1

AMR-2

N

HHEX-1 Q

CHEX

AMR-1

HHEX-2

HHEX-1

Q

(A)

N

HHEX-2

Q Linear actuator

AMR-2

S

Q Linear actuator

(B)

Fig. 17 Schematic representing a magnetic refrigerator with two AMR beds and piston methods for (A) cycle-1 and (B) cycle-2.

Magnetic field

Entropy

H 0H=0 H>0 1 4 2 3

0 Fluid flow

Time

+V 0

Time

−V (A)

Temperature

tMag

tCold

tDemag

tHot

(B)

Fig. 18 (A) S–T diagram of Brayton cycle. (B) The magnetic field and fluid flow periods of Brayton cycle.

Magnetic Energy Conversion

593

Magnetic field

Entropy

H 0H=0 H >0

0 Fluid flow

41 3 2

Time

+V 0

tMag

tDemag

Time

−V Temperature

(A)

tCold

tHot

(B)

Fig. 19 (A) S–T diagram of Ericson cycle. (B) The magnetic field and fluid flow periods of Ericson cycle.

Magnetic field

Entropy

H 0H=0 H>0

0 Fluid flow

12 6 3 5 4

Time

+V 0

tMag

tDemag

Time

−V (A)

Temperature

tCold

tHot

(B)

Fig. 20 (A) S–T diagram of hybrid cycle. (B) The magnetic field and fluid flow periods of hybrid cycle.

The Ericson cycle is shown in Fig. 19. In its S–T diagram, (1) is isothermal magnetization, (2) is isofield fluid flow at magnetic field from cold to hot end, (3) is isothermal demagnetization, and (4) is isofield fluid flow at zero magnetic field from hot to cold end. An adiabatic magnetization occurs in the process in the Carnot cycle. It goes with a further magnetization in the stage, which is an isothermal magnetization. Heat is extracted from the system while this process created. The next step is an adiabatic demagnetization process. An isothermal demagnetization is induced by connecting the system to a heat source. It's clear that the Carnot cycle can just be run if a minimum of four different magnetic fields happens, through which the MCM is moved. The combination of the Brayton and the Ericson cycles is called a hybrid cycle and it is shown in Fig. 20. The hybrid cycle has six different steps: (1) adiabatic magnetization, (2) isothermal magnetization, (3) isofield fluid flow at higher magnetic field from cold to hot end, (4) adiabatic demagnetization, (5) isothermal magnetization, and (6) isofield fluid flow at zero magnetic field from hot to cold end. In conclusion, the cooling capacity of a magnetic refrigerator strongly depends on many operation parameters (such as type of heat transfer fluid, shape of MCMs, designing of magnet, type of thermodynamic cycle, heat transfer fluid flow rate, etc.). The most important operation parameters are utilization factor (U) and operation frequency (f). In order to increase the performance of a magnetic refrigerator, these two different parameters have to be carefully selected during operation. The U parameter is determined by U¼

_ f
cf
tf m mM
cM

ð53Þ

_ f is the mass flow rate of the heat transfer fluid, cf is the specific heat value of the heat transfer fluid, tf is the flow period of Here, m heat transfer fluid, mM is the mass of MCM, and cM is average specific heat value of the MCM. The utilization factor should be determined and carefully optimized for each magnetic refrigerator by controlling the heat transfer fluid flow rate. Another important operation parameter is frequency and it depends on the thermodynamic cycle. The frequency (completed thermodynamic cycles per second) is given in the following equation: f¼

1 tMag: þ tCold þ tDemag: þ tHot

ð54Þ

594

Magnetic Energy Conversion

These parameters (U and f) influence the temperature span and the cooling power of the magnetic refrigerators. The geometry of the MCMs strongly affects their cooling performance. Fine MCM geometry (such as thin plates, small particle, spheres with small diameter, and wire with small diameter) speeds up the heat transfer mechanism from/to MCM. This increases the optimum operating frequency to higher values and its cooling power.

4.13.5

Results and Discussion

As the world population is rapidly increasing, the importance of having highly efficient, environmentally friendly, and energysaving technology is increasing every day. Rapid economic growth has been seen in the last 15 years in the developing world, and as economic growth is widely seen in parallel with cooling and heating technology, this brings a variety of environmental problems such as global warming. From business centers to homes, hospitals, and cars, there is a very wide range of applications for magnetic refrigerators that can resolve these types of problems caused by conventional cooling systems. When compared with a higher yield gas compression systems, the magnetic field controlled cooling/heating technology systems are especially saved energy due to environment-friendly properties. In particular, the damage to the environment from the origin and the limited use of energy resources in the world to increase the minimum level of problems in downloading, working with scientists to develop environmentally friendly high efficiency and it is an innovative cooling technology.

4.13.6

Future Directions

This high technology is in its middle R&D stage with some stakeholders attempting to commercialize in refrigeration for magnetic electromagnetic energy conversion applications within the next few years. With the new discoveries of the properties of MCMs, we would see this cooling technology in a few years and it could become part of everyday home appliances. Hopefully, we would see magnetocaloric air conditioning in the market in a few years and many prototypes of various capacities and for various applications, including high vacuum have been developed over the years. The final impact is that within magnetic and electromagnetic energy conversion of these systems, it’s going to be eventually 45% more efficient than the current technology, which is based on vapor compression systems. Some of the innovative ideas that scientists in this study believe will drive the MCE forward are listed below.

•

•

•

•

•

We anticipate that spring-shaped magnetic cooler materials that can be produced with 3D printers will cause a very important problem to be left behind. New spring-like material results obtained from non-rare-earth magnetic materials are very important to deal with including rare earth element magnets and have opened an era of developing permanent magnet technologies, suggesting that magnetic cooling materials in this geometry suggested in this section of this work will especially increase the heat exchange rate. Increased heat exchange will result in more heat exchange from the MCM. This will allow the magnetic cooling device to operate at higher frequencies. At the end of all this, we will achieve a more efficient system by acquiring the magnetic refrigerant operating at higher frequencies. With this study, magnetic materials that can be used in magnetic coolers near room temperature were revisited again. It is possible to make an effective magnetic cooling with the materials mentioned via this study. However, the increase in the magnitude of the MCE of these materials will cause a direct boost of the magnetic refrigerant. We believe that this increase is the fastest and quickest way of passing through nanotechnology. In recent years, the rapid increase in nanotechnology and the efficient adaptation of nanotechnology to many applications today have led to many requirements of modern society. Future works might obtain better permanent magnet features through doping third small elements like B, C, etc., and nanostructuring processes. These materials are thought by the authors to have significant effects in terms of cancer therapy applications, especially for those with transition temperatures above and below room temperatures, and some of the theoretical studies on the application have been made by the authors. From the perspective of cancer therapy applications of such materials, it is based on the destruction of cancerous cells by placing high-temperature differences using a pulsed magnetic field by placing these materials in a cancerous structure. However, such experiments of thought are still in the beginning stages. Work on this should continue and empirical findings should be presented. According to the scientists who carried out this study, the phenomenon that is thought to be another product related to the future use of the MCE is the magnetocaloric heat pump. As is known, the heat pumps used today are generally very fragile and provide mechanical flow of heat from one place to another. Because they are multi-parts, they cannot provide the heat transfer with the required and optimized high efficiency. Materials exhibiting MCE can transfer the heat from one heat region to another by keeping the internal heat at certain levels during the phase transition. Thus, MCMs can carry heat and cooling, as well as heat from one region to another. When recent scientific publications are examined, it is observed that some theoretical and experimental studies about magnetocaloric heat pumps have participated in the literature [98,99]. Microprocessors of today’s high-tech products are required to have a temperature around room temperature in order to be able to work with desired performance under load. Nowadays, it is almost impossible to keep processors at these temperatures under load, especially with conventional air-cooled, liquid-cooled systems. In addition, such systems can make quite a lot of noise and do not work at high performance. We believe that the introduction of such methods in the future of microprocessor

Magnetic Energy Conversion

595

coolers based on MCE will be a new breaking point for microprocessors. The MCMs that will be enlarged as a thin film upon production of microprocessors and the permanent magnet thin films, which will be grown as a magnetic field source, will enable microprocessors to perform self-cooling without any external cooling. This will allow the processor and electronic cooling to be performed without spending any extra power. Separate pieces of this imaginative approach (magnetocaloric thin film, permanent magnet thin film, etc.) are studied in the literature in a quick and detailed way [100–104]. These materials show superior properties separately. A superior cooler prototype can be achieved by enlarging these individual fine thin films in a suitable manner on a microprocessor.

4.13.7

Closing Remarks

The magnetic cooling system, which is an alternative to the other cooling technologies, is important in terms of being an environmentally friendly and energy-efficient system. From the standpoint of the system, the MCMs can be used like solid Gd, LaFeSi, CoMnGe, MnAs as a system. Some liquid materials and water are used as heat transfer fluid. The magnetocaloric loop is largely reversible, with a cycle of high environmental sensitivity. The efficiency is high because the system is reversible and has no inefficient parts such as a compressor. Another advantage of the magnetic cooling system is that high energy density can be achieved in compact devices. These novel results obtained from MCMs are very important to deal with problems about energy conversions and have opened an era of developing cooling technologies with magnetic and electromagnetic energy conversion. Future works might obtain better features through developing new elements and designs. Current refrigerators are among the biggest consumers of energy in home applications and they can leak fluorocarbons into the atmosphere, which can cause global warming. By using these magnetic and electromagnetic energy converted new technologies, we will try to help save the world from global warming.

Acknowledgment This work was done at the Magnetic Materials Research Laboratory in Ankara University and Functional Materials Research Laboratory in Recep Tayyip Erdoğan University.

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J Alloys Comp 2015;641:69. [76] Zhang CL, Shi HF, Ye EJ, et al. Magnetostructural transition and magnetocaloric effect in MnNiSi–Fe2Ge system. J Alloys Comp 2015;639:36. [77] Zhang CL, Shi HF, Ye EJ, et al. Magnetostructural transition and magnetocaloric effect in MnNiSi–Fe2Ge system. Appl Phys Lett 2015;107:212403. [78] Zhang CL, Shi HF, Nie YG, et al. Thermal-cycling-dependent magnetostructural transitions in a Ge-free system Mn0.5Fe0.5Ni(Si,Al). Appl Phys Lett 2014;105:242403. [79] Jacobs S, Auringer J, Boeder A, et al. The performance of a large-scale rotary magnetic refrigerator. Int J Refrigeration 2014;37:84. [80] Barcza A, Katter M, Private Commun. In: Proc.magnetic materials for the 21st century seminars, Hanau, 22–23 October 2012; 2015. [81] Kitanovski A, Egolf PW. Application of magnetic refrigeration and its assessment. J Magn Magn Mater 2009;321:777. [82] Kitanovski A, Tušek J, Tomc U, et al., Magnetocaloric energy conversion: from theory to applications. Green Energy and Technology: Springer; 2009. ISBN 978-3-31908740-5. [83] Brown GV. Magnetic heat pumping near room temperature. J Appl Phys 1976;47:3673. [84] Zimm C, Jastrab A, Sternberg A, et al. Description and performance of a near-room temperature magnetic refrigerator. Adv Cryog Eng 1998;43:1759.

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[85] Clot P, Viallet D, Allab F, et al. A magnet-based device for active magnetic regenerative refrigeration. IEEE Trans Magn 2003;39:3349. [86] Yao GH, Gong MQ, Wu JF. Experimental study on the performance of a room temperature magnetic refrigerator using permanent magnets. Int J Refrigeration 2006;29:1267. [87] Kawanami T. Heat transfer characteristics and cooling performance of an active magnetic regenerator. In: Second IIF-IIR international conference on magnetic refrigeration atroom temperature, Thermag II, Portoroz, Slovenia; 2007. p. 23–34. [88] Engelbrecht K, Bahl CRH, Nielsen KK. Experimental results for a magnetic refrigerator using three different types of magnetocaloric material regenerators. Int J Refrigeration 2011;34:1132. [89] Legait U, Guillou F, Kedous-Lebouc A, Hardyb V, Almanzae M. An experimental comparison of four magnetocaloric regenerators using three different materials. Int J Refrigeration 2014;37:147. [90] Tušek J, Kitanovski A, Tomc U, Favero C, Poredos A. Experimental comparison of multi-layered La–Fe–Co–Si and single-layered Gd active magnetic regenerators for use in a room-temperature magnetic refrigerator. Int J Refrigeration 2014;37:117. [91] Gatti JM, Muller C, Vasile C, et al. Magnetic heat pumps – configurable hydraulic distribution for a magnetic cooling system. Int J Refrigeration 2014;37:165. [92] Zimm C, Boeder A, Chell J, et al. Near-room-temperature refrigeration through voltage-controlled entropy change in multiferroics. Int J Refrigeration 2006;29:1302. [93] Zimm C, Auringer J, Boeder A, et al., Design and initial performance of a magnetic refrigerator with a rotating permanent magnet. In: Second IIF-IIR international conference onmagnetic refrigeration at room temperature, Thermag II, Portoroz, Slovenia; 2007. p. 341–7. [94] Vasile C, Muller C. Innovative design of a magnetocaloric system. Int J Refrigeration 2006;29:1318. [95] Tura A, Rowe A. Permanent magnet magnetic refrigerator design and experimental characterization. Int J Refrigeration 2011;34:628. [96] Engelbrecht K, Eriksen D, Bahl CRH, et al. Experimental results for a novel rotary active magnetic regenerator. Int J Refrigeration 2012;35:1498. [97] Plaznik U, Tusek J, Kitanovski A, Poredos A. Numerical and experimental analyses of different magnetic thermodynamic cycles with an active magnetic regenerator. Appl Therm Eng 2013;59:52. [98] Benedict AM, Sherif SA, Beers DG, Schroeder MG. Design and performance of a novel magnetocaloric heat pump. Sci Technol Built Environ 2016;22:520. [99] de Vries W, van der Meer TH. Application of Peltier thermal diodes in a magnetocaloric heat pump. Appl Therm Eng 2017;111:377. [100] Millera CW, Belyea DD. Magnetocaloric effect in nanoscale thin films and heterostructures. J Vac Sci Technol A: Vac Surf Films 2014;32:040802. [101] Millera CW, Williams DV, Bingham NS, Srikanth H. Magnetocaloric effect in Gd/W thin film heterostructures. J Appl Phys 2010;107:09A903. [102] Yüzüak GD, Yüzüak E, Elerman Y. Hf2Co11 thin films: rare-earth-free permanent nanomagnets. Thin Solid Films 2017;625:115–21. [103] Jiang JS, Bader SD. New thin film permanent magnets. Mater Sci Technol 2013;17:1491–9. [104] Uehara M. Microstructure and permanent magnet properties of a perpendicular anisotropic NdFeB/Ta multilayered thin film prepared by magnetron sputtering. J Magn Magn Mater 2004;284:281.

Further Reading Coey JMD, Magnetism and magnetic materials, ISBN-13: 978-0521816144. Kitanovski A, Tušek J, Tomc U et al., Magnetocaloric energy conversion, from theory to applications, ISBN 978-3-319-08741-2. Löwe K, Liu J, Skokov K, et al., The effect of the thermal decomposition reaction on the mechanical and magnetocaloric properties of La(Fe,Si,Co)13, Acta Mater, 60 (2012) 4268. Pecharsky Vitalij K, Gschneidner Karl A Jr., Magnetocaloric materials, DOI: 10.1002/9780470022184.hmm417. Tishin AM, Spichkin YI, The magnetocaloric effect and its applications, ISBN 9780750309226.

Relevant Website http://www.malzemebilimi.ankara.edu.tr/index-en.php Ankara University – MMAG.

4.14 Electromechanical Energy Conversion Ahmet Cansiz, Istanbul Technical University, Maslak, Istanbul r 2018 Elsevier Inc. All rights reserved.

4.14.1 4.14.2 4.14.2.1 4.14.2.2 4.14.2.3 4.14.2.3.1 4.14.2.4 4.14.2.5 4.14.2.6 4.14.3 4.14.3.1 4.14.3.2 4.14.3.3 4.14.3.4 4.14.3.5 4.14.3.6 4.14.3.7 4.14.3.8 4.14.3.9 4.14.4 4.14.4.1 4.14.4.2 4.14.5 4.14.5.1 4.14.5.2 4.14.5.3 4.14.5.3.1 4.14.5.3.2 4.14.5.3.2.1 4.14.5.3.2.2 4.14.5.3.2.3 4.14.5.3.2.4 4.14.6 4.14.6.1 4.14.6.1.1 4.14.6.1.2 4.14.6.1.3 4.14.6.2 4.14.6.3 4.14.6.4 4.14.7 4.14.7.1 4.14.7.2 4.14.7.3 4.14.7.4 4.14.8 4.14.8.1 4.14.8.2 4.14.8.3 4.14.9 4.14.10 References

598

Introduction Renewable Energy Consideration Biomass Energy Geothermal Energy Hydropower Energy Hydropotential and hydrokinetic energy conversion systems Ocean Energy Solar Energy Wind Energy Conversion Process for Energy Storage Chemical Energy Storage and Batteries Fuel Cells and Hydrogen Energy Storage Thermal Energy Storage Pumped Hydro Energy Storage Compressed Air Energy Storage Flywheel Energy Storage Superconducting Magnetic Energy Storage Energy Storage in Supercapacitors Electricity and Energy Storage Fundamentals of Electromechanical Energy Conversion Coupling Field Reaction Conservation of Energy Principle and Energy Balance Equation Electromechanical Energy Conversion Devices Transducers Transformers Rotating Electrical Machines: Motor-Generators DC machines AC machines Induction machine Operation modes Power flow and efficiency Synchronous machine Energy Losses in Electromechanical Energy Conversion Systems Engineering Materials Permanent magnets Powdered cores Ferrite cores Saturation, Hysteresis in the Core Materials and Associated Losses Hysteresis and Associated Losses Eddy Current Losses Design of Efficient Electromechanical Energy Conversion System Thermodynamic Analysis of Energy Conversion Energy Transport by Heat Energy Transfer by Work Energy Conversion Efficiencies in Thermodynamic Systems Case Studies Nanostructured Energy Conversion Devices Superconducting Magnetic Energy Storage Superconducting Flywheels as Electromechanical Energy Conversion Future Directions Closing Remarks

Comprehensive Energy Systems, Volume 4

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doi:10.1016/B978-0-12-809597-3.00425-9

Electromechanical Energy Conversion Further Reading Relevant Website

Nomenclature

635 635

NdFeB PE PHES PM PV RE RMS RPM SFES SHS SMES SmCo TCS VA

Neodymium iron boron Potential energy Pumped hydro energy storage Permanent magnet Photovoltaic Renewable energy Root mean square Round per minute Superconducting flywheel energy storage Sensible heat storage Superconducting magnetic energy storage Samarium cobalt Thermochemical storage Volt ampere

n ns P p Pw Pin Pout Ph Pe Qcond Qconv Qrad q R r s T U V v WE WM Wfc Wfg

Rotor speed Synchronous speed Power Power density Wind power Input power Output power Hysteresis loss Eddy current loss Heat conduction Heat convection Heat radiation Electric charge Resistance Radius Slip Temperature Internal energy Voltage Speed (m/s) Electrostatic energy density Magneto static energy density Field energy in the core Field energy in the gap

Greek Letters r Density (kg/m3) Z Energy efficiency (%) Permittivity of free space e0 m0 Permeability of free space f Magnetic flux

l y o ℜ t tT

Flux linkage Angular position Angular speed Reluctance Response time Torque

Subscripts c Core

Cu f

Copper Field

Abbreviations AC Alternative current CAES Compressed air energy storage COF Coefficient of friction DC Direct current ES Energy storage EMF Electromotive force EMEC Electromechanical energy conversion EMECD Electromechanical energy conversion device FES Flywheel energy storage HTS High temperature superconductor KE Kinetic energy LHS Latent heat storage MMF Magneto motive force

Symbols A a B Br C E Etot e ed F G g H Hc h I i kt L l lc lg m N

Cross-sectional area Turn ratio Magnetic flux density Remnant field Cycle life Electric field Total energy Electromotive force Energy density Force Giga Gravitational acceleration Magnetic field strength Coercive field Convective heat transfer Current (RMS) Current (instantaneous) Thermal conductivity Inductance Length Core length Gap length Mass Number of turn

599

600

g in m max out

4.14.1

Electromechanical Energy Conversion

Gap Input Mechanical Maximum Output

p r s sr

Pole Rotor Stator Stator-rotor

Introduction

The history of mankind is closely linked with the history of energy. The use of energy may have begun with the use of fire. Survival and daily needs, such as cooking and protection, would have been the first use of fire. The first inspiration for the use of fire may have been due to a lightning strike. Until humans learned how to start a fire they might have experienced many natural events that would have helped them. The first use of fire might have been started accidentally with the observation of a nearby strike of lightning resulting in a fire, which would have also accidentally helped humans understand the benefits of fire. Based on these experiences, the history of the use of energy might have started with controlling the use of fire. It may also be assumed that people first enjoyed the comfort of heat energy at this time. This in turn accelerated the development of the first industry, via cooking, warming against the cold, lighting against the dark, and protection from the wild. Fire would be the first serious arm used by human kind. The use of fire throughout history has increased in time resulting in a vast number of examples up to the present time. It is clear that the use of energy started with fire has made a substantial impact on the history of mankind. It has taken a long time to reach maturity of today and is still in a continuous development process. In addition to the naturally provided energy from the sun, the history of energy use by humans evolved by harnessing energy from domesticated animals. Animals are accustomed to carrying out heavy duties, such as plowing land, turning mills, and carrying loads. This would have been followed by wind power, which made possible the discoveries in navigation, which led humans to explore new lands and continents. Wind power was important in the transformation of primary products by means of windmills, and ignited the first industrial processes developed by man. Industrial energy use by mankind started in the 18th century with the invention of steam power. It was a milestone in the history of energy development, and initiated the first industrial revolution of the modern age. Heat energy is converted into a mechanical energy in the steam engine, which constitutes a thermomechanical energy conversion device. Thermomechanical energy conversion was the inspiration for all kinds of energy conversion processes. With the use of coal and oil, thermomechanical energy conversion technology reached its peak and is inevitable even today, despite serious environmental issues. Long after the invention of electricity, the electromechanical energy conversion (EMEC) age started. Indeed, electricity has existed since ancient times, but its use in industrial means only commenced during the mid-19th century and was responsible for the great developments in industry. As in the scenario mentioned above, mankind would have first experienced external energy from lightning (a form of electrical energy). However, its complete use was developed in the last two centuries, leading to rapid industrial development. The use of electricity has forced the EMEC in a way that has encircled our lives in every way. Energy is formally defined as the ability or the capacity to do work and has always existed in one form or another. According to this definition, which makes the plants and machines work, energy can be divided into two different types: energy that is stored is called potential energy (PE) and energy that is moving is called kinetic energy (KE). Energy can be found in a number of different forms. It can be in the form of chemical, electrical, thermal, light, mechanical, and nuclear, and can be transformed from one form into another. This transformation is necessary to provide continuous machine work where the energy conversion process occurs. According to the second law of the thermodynamics energy can neither be created nor destroyed. This is known as the principle of energy conservation and it must be realized when an energy conversion device is being developed. All forms of energy are stored in certain ways based on the convenience of the energy sources. The energy sources are divided into two main groups: renewable, where the energy source can be used over and over again; and conventional, where the energy source cannot be reused or created. Renewable energy (RE) sources include solar energy (which comes from the sun and can be transformed into electricity and heat), wind energy from temperature differences of the earth’s surface, geothermal energy (from inside the earth), biomass from plants, and hydropower from water are also RE sources. Most of the energy consumed in the industry comes from nonrenewable energy sources, which include fossil fuels, such as oil, natural gas, and coal. There is a controversy to classify a unique energy source obtained from the splitting of the element uranium through the process of nuclear fission. By definition, in order to count any specific kind of energy as renewable, it should not pollute the Earth and it should exist naturally, without the necessity of any process. For example, sunlight is available all the time and the wind blows whether or not it is converted into any form to put into use. From this perspective nuclear energy might be considered as renewable, since it does not pollute the environment, except for the possibility of nuclear accidents. This judgment generally is left to the reader. According to recent reports from energy storage (ES) databases, less than 2% of the total electric energy capacity of the world is stored, while the remaining 98% of the world’s available energy is not being used effectively [1]. This report points out the fact that energy conversion technology is still in need of development. In this context, EMEC devices are the driving force for maintaining a stable and continuous development for the industry. The first industrial revolution started with the use of steam power to mechanize production, the second used the power of electricity to create mass production, and the third one used electronics and information technology to automate production. The fourth industrial revolution, also known as the digital revolution, which the

Electromechanical Energy Conversion

601

world is experiencing at the present time, is based on the previous one, providing a digitalized industry. Considering the history of these revolutions, invention of steam power in the 1780s initiated mechanical production equipment, the introduction of electricity in the 1870s initiated mass production, the introduction of electronics and information technology in the 1970s initiated automated production, and eventually introducing digital technology during the 1990s initiated cyber-physical systems. When compared with previous industrial revolutions, the fourth is evolving at an exponential rather than a linear pace, initiating changes and transformation of entire systems of production, transportation, distribution, storage, and management of energy. Industrial development is closely linked to the availability of energy in useful forms, requiring vast amounts of energy in all forms, such as light, heat, electrical, mechanical, chemical, and nuclear. The demand for energy has increased steadily, not only because of the growing industry but also because of the technological improvements to meet the demands of mass production. The main reason for the increased demand rate is the needs of the growing population of the world. As a result of the increase in the consumption of the energy, concerns have arisen about the exploration of natural resources, because most of the existing energy sources result in a contaminated environment. Most of the energy is generated by the combustion of fossil fuels, such as coal, oil, and natural gas. Beside the world has only a finite supply of these fuels, the combustion of these fuels releases various pollutants, such as carbon monoxide and other poisonous gases. These create health risks and may contribute to environmental disasters, such as acid rain and global warming. For reasons the main focus in this analysis is aimed toward RE sources and related storage and conversion processes. EMEC is a tool to interconnect certain forms of energy sources. The energy conversion process closely links with energy generation, transmission, distribution, and storage. Because the focus of modern technologies is on RE, the structure of the present analysis will be based on the introduction of RE sources first. Throughout the text, energy sources and storage will be discussed in the general form, including all kinds of energy sources (renewable and nonrenewable) and ES technologies. Once these topics are well understood in their contexts, the energy conversion processes associated with these energy sources will be introduced. The device applications and associated engineering materials are also introduced together with the principles of operation, design, and efficiency evaluation. The operation and design of EMEC systems and devices are discussed in terms of thermodynamic principles to provide a systematics of the evolution of any conversion device to be built. Covering the energy conversion process with various emerging popular technologies will be focused on, with case studies including nanostructured systems and superconducting systems. The case studies about these topics will be particularly focused on the EMEC associated with the process of the conversion itself, operation principles, and energy losses.

4.14.2

Renewable Energy Consideration

RE sources are essential in order to provide sustainable energy generation for developing industry. Due to environmental issues, the focus on the RE sources related to conversion devices has highly increased in recent decades. RE sources are considered to be hydropower, wind power, solar power (photovoltaic (PV)), biomass, and geothermal energies. Working with RE sources will also require dealing with the conversion devices for wind, solar, and hydropower turbine systems, which are the most important types of device applications of RE. There are advantages and disadvantages of renewable energies for the society we live in. Since extracting energy for sustainable development is a necessity, some of the disadvantages of the RE are disregarded via taking into account the required precautions. Based on these facts, when we evaluate biomass, for example, the release of methane during the production of biofuels is attracting the attention of environmentalists because methane is a global warming gas. Another RE is geothermal energy. Geothermal installations pollute waterways. Hydropower changes the ecosystem and impacts social and cultural life. Ocean energy kills fishes and prevents water from circulating. Solar energy generates hazardous waste, and wind turbines kill birds, create noise, and cause mandatory changes to the landscape. These are minor disadvantages compared to their major benefits for the society. In addition to the reduction of CO2 emissions, governments tend to regulate energy policies in terms of RE to meet the objectives of environmental and health policies. Renewable energies are energy sources that are continually restored by nature and they are not directly or indirectly derived from fossil fuels. The main source of RE is the sun. The sun provides thermal, chemical, and electric energies and it is the driver for other energy sources, such as wind, hydro, and biomass. Moreover, the sun also indirectly affects the formative environmental mechanisms for potential RE, such as geothermal and tidal energy. Fig. 1 shows the main RE sources. Renewable energy sources

Wind

Solar

Wave

Onshore Offshore

PV

Heating

Fig. 1 Classification of renewable energy (RE) sources. PV, photovoltaic.

Hydro

Geothermal

Electricity

Bio

Fuel

602

Electromechanical Energy Conversion

4.14.2.1

Biomass Energy

Biomass is an organic material originating from plants, trees, and crops. Biomass energy is classified as one of the most important renewable energies. This type of energy is obtained during the conversion of biomass into useful forms of energy, such as heat, electricity, and biofuels. While biomass can be directly burned to obtain energy, it can also be converted to biofuels. Biofuels can be used for power generation. The utilization of biofuels help to enhance the energy alternatives for economic development. In addition to these advantages, the potential benefits and technical limitations of biomass energy are widely discussed in the literature [2–4]. According to the studies about biomass energy, one of the major benefits is that it reduces the use of fossil fuels. However, the high demand of fertilizers, which creates potential for the environmental pollution and a wide use of biomass energy, is a concern due to carbon emissions from wood burning. In recent decades, biomass energy has received much attention as a RE source and research on the conversion of biomass to energy is being boosted in many countries [3]. The conversion technologies regarding biomass have received much attention from researchers and industry [4].

4.14.2.2

Geothermal Energy

Geothermal energy is a thermal energy and is classified as RE. Geothermal energy is a stored energy of steam or liquid water trapped in the earth’s interior and it can be transformed into useful energy by using a geothermal power plant. Geothermal power is considered as a cost-effective, reliable, and environmentally friendly energy source. In addition, the technology for electricity generation from geothermal reservoirs has been well developed and matured for more than a century. It is known that the temperature gradient of the earth along the center is 25–301C/km. For the case of geothermal systems, however, this temperature gradient may vary as much as 50–701C/km. This temperature gradient can constitute a potential for energy generation. In this sense, the geothermal fields can be used for electrical power generation, for heat generation, or for combined heat and power in cogeneration applications. The conversion process of the geothermal energy is also a developing technology and it is highly studied in terms of present and future perspectives in the literature [5,6].

4.14.2.3

Hydropower Energy

Hydropower is a power that is derived from the energy of moving water. Flowing water creates energy that can be captured and converted into electricity by using turbines. The most general form of hydropower is obtained by constructing water dams. Hydropower is generated from water moving from high to low PE levels. Hydropower plants may have power levels of a few watts to several gigawatts. Hydropower plants serve the industry in various ways, such as electricity generation and irrigation purposes. Pumped hydro is also used for energy generation purposes for later use. Since pumped hydro is mostly used for ES purposes it will be discussed in the storage context later in Section 4.14.3.4. Hydropower is a well-developed technology based on more than a century of experience and it is among the best conversion efficiency systems of all energy sources due to its direct transformation of hydraulic energy to electricity. There is still further improvement by refining operation, reducing environmental impacts, and developing more robust and cost-effective technological solutions [7].

4.14.2.3.1

Hydropotential and hydrokinetic energy conversion systems

Water is an industrial material. The importance of water within the context of energy comes from its use in energy production. As mentioned for the hydropower, it can be stored in water dams fed by the river and used for hydropotential and hydrokinetic energy purposes. RE sources are the best alternative to the power plants using fossil fuels to answer growing demand. Hydropotential energy as a RE source is as old as early human civilization. Ancient dams were mostly built to provide water for irrigation purposes. The energy production by using water dams effectively started after the 1950s. Although hydropower plants are considered as RE sources, there are some public reservations against building new water dams. Energy production using water dams provides almost 20% of the world’s energy demand. A hydroelectric power plant produces electricity during the conversion of PE to KE by means of a turbine and generator. A typical hydroelectric power plant, shown in Fig. 2, basically consists of a water reservoir to accumulate PE, inclined penstock to guide water, a turbine to obtain rotational power from water impinging on the blades, a generator to convert mechanical (rotational) power to electrical power, and most of all a river to feed the reservoir of the water dam. The electricity generated by the power plant is transformed into a high voltage with step-up transformer to reduce power losses. The electricity is transmitted to long distances via transmission lines. The electricity at high voltage is transformed to low voltage again by using a distribution transformer for public and industrial use. As mentioned above, public perceptions of water dams are still controversial due to their pros and cons. The operation cost of a hydropower plant can be considered as minimal since it does not require fossil fuel. From this point of view, these plants can be classified as beneficial for the environment. Hydroelectric power plants produce fewer methane emissions compared to a thermal power plant. However, an effect on the ecosystem is inevitable since the construction of a water dam changes the surrounding area. Pumped hydropower plants are not energy sources, but they can be used as ES devices. In such a system, water is pumped from a lower reservoir to an upper reservoir, usually during off-peak hours. The water is released to generate electricity during the daily

Electromechanical Energy Conversion

603

Transmission line

Turbine Generator Reservoir

Transformer

Dam

Control gate Penstock

Outflow

Fig. 2 A typical water dam and its components.

peak load period. Since the hydropower plants are also an energy consumer during the pumping process, in order to be efficient the plants need to be installed on a large scale. In fact, pumped storage is the largest-capacity form of grid ES now readily available worldwide. One of the emerging types of RE technology is hydrokinetic conversion, which is based on the capture of energy from flowing water [8]. Hydrokinetic energy conversion systems are designed and generally constructed in natural rivers and ocean wave currents. Hydrokinetic conversion devices are arranged in such a way to extract the energy from the PE accumulated in a water reservoir or from ocean waves. The energy conversion system established on a river current can be described as an energy converter that harnesses the KE of river streams. On smaller scale grids, hydrokinetic energy can be used to deliver electrical power directly for commercial, residential, agricultural, or public facilities [9].

4.14.2.4

Ocean Energy

Ocean energy is one of the mainstream RE sources, and comes from various sources, such as waves, tidal currents, and ocean currents. Waves are generated by wind passing over the ocean surface due to pressure and temperature variations. Therefore the height of sea waves and duration of the wind determine the amount of energy transferred to the energy conversion system. Wave energy devices extract energy from the surface motion of the waves or from pressure fluctuations below the surface. This energy is converted to mechanical energy that drives a generator to produce electricity. These types of energy sources are newly emerging compared to the other mainstream sources shown in Fig. 1. Extracting energy from these sources is improving with different energy conversion technologies [8,10]. For example, harnessing energy from ocean waves started in the early 18th century and is still improving today.

4.14.2.5

Solar Energy

Solar energy generation systems involve the use of the energy from the sun. There are two main methods to extract energy from these systems. In most of these applications the purpose is to provide hot water. The other main use is to generate electricity. These technologies are technically well proven with numerous systems installed around the world. PV systems directly convert solar energy into electricity. PV systems or modules are formed by semiconducting components, which convert solar energy into direct current (DC) electricity, typically up to 50–200 W [11]. PV modules are combined with a set of additional system components, such as inverters, batteries, electrical components, and mounting systems. The most established solar PV technologies are siliconbased systems. More recently, so-called thin film modules, which can also consist of nonsilicon semiconductor materials, have become increasingly important. Other technologies, such as organic PV cells, are still in the research phase [11]. PV systems are classified as off-grid and grid-connected applications. Off-grid PV modules are usually suitable in areas where there is no electricity. Centralized systems, such as PV for local areas, have technical advantages concerning electrical performance, reduction of storage needs, and availability of energy, and provide cost-efficient services [12]. Grid-connected PV systems use an inverter to convert electricity from DC to alternative current (AC) electricity, and then supply the generated electricity to the electric grid [7].

604 4.14.2.6

Electromechanical Energy Conversion Wind Energy

Wind power is defined as the conversion of wind energy into a useful form by using wind turbines. Wind power can be used to generate electricity, generate mechanical power for the windmills, and for the wind pumps to pump water. Wind turbines for electricity generation were first developed at the beginning of the 20th century. The technology has gradually improved since the early 1970s. By the end of the 1990s, wind energy has reemerged as one of the most important sustainable energy resources because of the technological advances of conversion efficiency in wind turbines [13]. Generating electricity from the wind requires that the translational KE of moving air first needs to be converted to rotational mechanical energy. The mechanical energy rotates the mill of the generator to produce electricity. The development of wind turbines accelerated during the last few decades, providing cost-effective wind turbines and power plants to perform efficient energy conversion. The amount of KE in the wind that is theoretically available for extraction increases with the cube of the wind speed. The power Pw available in the wind is, Pw ¼

1 rAv3 2

ð1Þ

where r is the air density (kg/m3), A is the cross-sectional area through which wind passes (m2), and v is wind speed (m/s) [13]. A turbine only captures a fraction of the available wind energy. Thus the wind turbine design must be focused on maximizing energy capture over the range of wind speeds experienced by wind turbines, while at the same time minimizing the cost of wind energy taking into account such factors as material use and turbine size. Wind energy is a clean energy system since it does not pollute the environment, and is the cheapest energy source of all the renewable sources. However, the capital cost of the wind energy systems is high because the generated energy is intermittent due to changing wind speed. Wind turbines also make more noise and cause the death of birds. From the perspective of electric system reliability, an important part of the wind turbine is the EMEC system. For large gridconnected turbines, electrical conversion systems come in different forms. For modern turbines, the designs have variable-speed machines, such as doubly fed induction generators and synchronous generators. These turbines can provide the real and reactive power required by electric networks [13].

4.14.3

Conversion Process for Energy Storage

ES can be defined as a medium that provides interconnection between variable energy sources and loads. The ES devices function as energy buffers or backups to prevent unregulated electric power flow between the supply and the demand sides. Due to its dynamic nature, electricity is not easy to store directly; it can instead be stored in other forms and converted back to electricity when it is needed. Storage technologies for electricity can also be classified in terms of the storage types, such as those given in Table 1. As shown in this table the major ESs are electrical, mechanical, chemical, and thermal. From this table the various forms of energy can lead to the devices or systems, such as supercapacitors, batteries, flywheel energy storage (FES), superconducting flywheel energy storage (SFES), superconducting magnetic energy storage (SMES), pumped hydro energy storage (PHES), and compressed air energy storage (CAES) [14–16]. Supercapacitors are electricity-based storage devices. Batteries store the energy in the form of electrochemical energy and are utilized in various sizes [15,17,18]. CAES is also one of the methods for bulk ES [19]. Another mechanical type of storage is the pumped hydro, which operates by the principle that when there is an energy surplus the electricity can be used to pump the water from a lower reservoir to a higher one when cheap energy is available. The most effective mechanical ES devices are flywheels, which store electrical energy in the form of rotational KE [16,20,21]. SMES devices store the electric energy in the form of magnetic fields. For grid applications, the major ES devices are batteries, flywheels, supercapacitors, SMES, pumped hydro, and compressed air units [15–21]. The specifications and the comparison of some of the novel ES devices are given in Table 2 [18]. Thermal, fuel cells, pumped hydro, and compressed air are not indicated in the table since they constitute different energy density specifications. This table suggests that there is no particular ES device that Table 1

Energy storage (ES) methods

ES type

Form of energy

Examples

Electrical energy

Electrostatic Magnetic/current Kinetic Potential Electrochemical Chemical Thermochemical Low temperature High temperature

Capacitors and supercapacitors Superconducting magnetic energy storage (SMES) Flywheels, superconducting flywheel energy storage (SFES) Pumped hydro and compressed air Batteries Fuel cells Solar hydrogen Cryogenic ES Steam, latent heat systems

Mechanical energy Chemical energy

Thermal energy

Electromechanical Energy Conversion

Table 2

605

Energy storage (ES) devices and their specifications

Type

Z (%)

ed (Wh/kg)

p (W/kg)

t (ms)

C (times)

Cost ($/kWh)

Lithium-ion SMES Flywheel Supercapacitor

70–85 95–98 95 95

100–200 30–100 5–50 o50

25–1000 104–105 1000–5000 4000

30 5 5 5

500–2000 106 420,000 450,000

150–1300 High 380–2500 250–350

Notes: Z, energy efficiency; ed, energy density; p, power density; t, response time; C, cycle life. Source: Reproduced from Vazquez S, Lukic SM, Galvan E, Franquelo LG, Carrasco JM. Energy storage systems for transport and grid applications. IEEE Trans Ind Electron 2010;57 (12):3881–95

meets both the technical and economical requirements of the growing grid. In order to solve the ES issue of the industry, the combination of various types of ES systems to form a hybrid system is encouraged.

4.14.3.1

Chemical Energy Storage and Batteries

Batteries have been around for more than a hundred years. There is a wide range of technologies used in the fabrication of batteries, such as lead acid, nickel–cadmium, nickel–metal hydride, nickel–iron, zinc–air, iron–air, sodium–sulfur, lithium-ion, and lithium polymer [20,22,23]. Battery systems are modular, quiet, and nonpolluting [22,24,25]. The response time of the batteries is about 20 ms and their round trip efficiency is in the range of 60–80% [25]. For example, as shown in Table 2 Li-ion batteries are leading in efficiency. Batteries store energy as an electrochemical process. During an electrical charge and discharge cycle the temperature change in the battery must be controlled [22]. Another major concern is the battery’s life cycle. The battery/cycle application may require a mechanism to charge and discharge multiple times a day. The maximum discharge rate of the battery is also of concern because the battery can also be damaged by a high discharge rate. For power quality issues, such as voltage regulation, frequency control, short term interrupts, and spinning reserve, these systems can be quite effective [24]. Batteries are often used in portable systems. As opposed to capacitors, voltage remains stable as a function of charge level in the batteries [14].

4.14.3.2

Fuel Cells and Hydrogen Energy Storage

Fuel cells generate electricity from the fuel on the anode and the oxidant on the cathode and react in the electrolyte. Basically, a fuel cell combines hydrogen and oxygen to produce electricity through water electrolysis. The storage system proposed includes three key components: electrolysis, which consumes off-peak electricity to produce hydrogen; the fuel cell, which uses hydrogen and oxygen from air to generate peak-hour electricity; and a hydrogen buffer tank to ensure appropriate resources in periods of demand [14,16]. There are many types of fuel cells, such as alkaline, polymer exchange membrane, direct methanol, phosphoric acid, molten carbonate, and solid oxide. Fuel cells can offer a solution for isolated areas where the installation of power lines is too difficult or expensive. There are several hydrogen storage modes, including compressed, liquefied, and metal hydride [19].

4.14.3.3

Thermal Energy Storage

Thermal ES is a method that stores thermal energy by heating or cooling a storage medium so that the stored energy can be used at a later time for power generation. The generated power is generally used to produce electricity, which can be used to provide heating and cooling. There are three types of thermal ES systems, classified in terms of whether they use sensible or latent heat. The first one is sensible heat storage (SHS), which is based on storing thermal energy by heating or cooling a storage medium, such as water, sand, molten salts, and rocks. The second is latent heat storage (LHS) using phase change materials and the third one is thermochemical storage (TCS) using chemical reactions. During accumulation, the material will shift from the solid state to liquid state and during the retrieval of the heat, will transfer back to solid state. The heat transfer between the thermal accumulator and the environment is carried out through a fluid material. The energy can be stored at high concentration depending on the temperature [14]. Once the storing is completed the heat is recovered to produce water vapor, which can be used to drive a generator for electricity. During off-peak hours, the hot water for storage can be obtained from a thermal plant. Generating extra electricity during peak hours can be achieved when retrieving stored energy in the thermal plant.

4.14.3.4

Pumped Hydro Energy Storage

The pumped hydro uses the power that is obtained from accumulated potential/KE in the water. The main advantage of pumped hydro is that it is readily available. Pumped hydro storage systems are essential for the storage of electrical energy. The principle is that during the periods when demand is low, these stations use electricity to pump the water from the lower reservoir to the upper reservoir. When demand is very high, the water flows out of the upper reservoir and activates the turbines to generate high-value electricity for peak hours. Pumped hydroelectric systems have a conversion efficiency of about 65–80% [14].

606 4.14.3.5

Electromechanical Energy Conversion Compressed Air Energy Storage

CAES relies on the mechanical energy formed by pressurized air in a container. A power plant with a standard gas turbine uses its available power to compress the air. Based on this process, the air is compressed by using the electrical power during off-peak hours (storage hours), to then produce electricity during peak hours (retrieval hours) by expanding the air in a combustion chamber. CAES is achieved at high pressures (40–70 bars). Large caverns made of high-quality rock deep in the ground, and underground natural gas storage caves, are the best options for compressed air storage, as they benefit from geostatic pressure, which facilitates the containment of the air mass [14].

4.14.3.6

Flywheel Energy Storage

FES systems are comprised of a massive or composite flywheel coupled with a motor–generator set inside a housing at very low pressure to reduce self-discharge losses. They have a great cycling capacity as indicated in Table 2. In order to store energy in an electrical power system, high-capacity flywheels are needed. Friction losses of a 200-ton flywheel are estimated at about 200 kW [14]. As indicated in the literature given in [14], assuming the instantaneous efficiency of 85%, the overall efficiency would drop to 78% after 5 h, and 45% after one day. Long-term storage with this type of apparatus is therefore not feasible. The flywheel energy will be discussed later in the text in detail with the consideration of superconductor use.

4.14.3.7

Superconducting Magnetic Energy Storage

SMES is achieved by inducing DC current into a coil made of superconducting cables of nearly zero resistance, generally made of niobium titane (NbTi) filaments that operate at very low temperature. The conventional superconductors usually operate in liquid helium (4K) and high temperature superconductors (HTSs) in liquid nitrogen (77K) temperatures. The current increases when charging and decreases during discharge and has to be converted for AC or DC voltage applications. One advantage of this storage system is its great instantaneous efficiency, nearly 95% for a charge–discharge cycle, as shown in Table 2. Moreover, these systems are capable of discharging nearly the total of the stored energy instantly, as opposed to batteries. They are very useful for applications requiring continuous operation with a great number of complete charge–discharge cycles. The fast response time (under 100 ms) of these systems makes them ideal for regulating network stability (load leveling). Their major shortcoming is the refrigeration system, which is quite costly and makes operation more complicated [14]. SMES will be discussed in more detail in the case studies section.

4.14.3.8

Energy Storage in Supercapacitors

ES in supercapacitors is achieved in the form of an electric field between two electrodes. The principle is the same as capacitors except that the dielectric material is replaced by electrolyte ionic conductor in which ion movement is made along a conducting electrode. The energy density is superior to that of capacitors. Serial connection, as opposed to capacitors, is required to reach normal voltages in power applications. As can be inferred from Table 2, supercapacitors generally are durable with 95% efficiency and 5% per day self-discharge. This means that the stored energy must be consumed. The technology and associated concepts of ES, RE sources, and electromechanical (or any kind of) energy conversion systems are all interrelated. This is due to the double-conversion chain, which can be described as “electricity – storable intermediary energy – electricity.” Based on this chain, the storable intermediary energy is established by sophisticated conversion systems between the energy sources and the electricity. For a complete understanding of energy conversion all of the energy sources and their conversion technologies are reviewed in this analysis. As a matter of fact, when a particular energy source is examined the analysis is carried out in terms of the double-conversion chain [22,23,19,26].

4.14.3.9

Electricity and Energy Storage

Electricity is more an energy carrier than an energy source. Electricity is produced from various sources of energy, such as the PE of water in hydroelectric power plants, chemical energy of petroleum, thermal energy of natural gas, and nuclear energy of fission in the nuclear power plants. In these examples, generally the available energy is converted into mechanical energy and then this mechanical energy is converted into electrical energy. Electricity is not very often used directly as a final usage of energy, but is instead converted in other forms. There are, however, important cases in which electricity is the final usage of electric energy, such as in electronic appliances that operate using electricity. Since electricity is a dynamic form of energy and it can only have meaning during its consumption, it is rational to focus on its storage and conversion in various forms and convert it back to electricity when needed. Electricity can be stored in electrical and magnetic field forms. The energy-storing capacity of the magnetic field is much greater than that of the electric field. Applying voltage to capacitor terminals causes a separation of charges, which results in an electric field inside the capacitor. The field persists once the voltage source is removed and the field diminishes when an external resistor allows them to flow to the other terminal. Similarly, in an inductor, according to Ampere’s law a current flowing through a conductor generates a proportional magnetic field around it. This magnetic field opposes changes in the current through the inductor by inducing back voltages across it. In both cases the analogy is much like storing energy in the gravitational field.

Electromechanical Energy Conversion

607

Capacitors and inductors are also known as the ES elements in the circuit theory, where the energy is stored in electric and magnetic field forms, respectively. It has to be noted that a capacitor acts similar to a temporary battery when charged and is capable of storing electrical energy. While on the other hand, an inductor has no storage capacity. It only creates a magnetic field as current passes through it. A capacitor is capable of storing electrical energy in electrostatic form. While the inductor stores electrical energy in the form of current-carrying conductor or magnetostatic form. The use of the magnetic field as a coupling field is more common in machine applications since the energy-storing capacity of the magnetic field is much higher than that of the electric field. Energy density, associated with the energy-storing capacity, is obtained from electromagnetic field theory in connection with the circuit theory and it is an important criterion for designing an efficient EMEC device. EMEC devices use electricity to convert any energy form into mechanical energy form and vice versa. The energy is storable in electric and magnetic field forms and can be used to operate EMEC devices. In most cases the energy conversion procedure requires interaction of the various energy forms with the mechanical parts through the air via coupling field, especially in rotating machines where the air gap is unavoidable. In this respect, besides the energy density, the coupling field is another important criterion to design a device for a particular purpose. Both electric and magnetic fields store energy and mechanical forces can be obtained via derivation of energy with respect to displacement. The coupling field capacity of electric and magnetic fields are different from each other. Under the normal atmospheric pressure the dielectric strength of air restricts the working electric field intensity (E) to about E¼ 3  106 V/m. From the field theory, the electrostatic energy density is given as, WE ¼

1 e0 E2 2

ð2Þ

where e0 is the permittivity of free space. With a typical dielectric having a strength of E ¼3  106 V/m and finite permittivity of the material the energy density can be estimated. Since the force is derived from the energy equation given in Eq. (2) we can obtain approximately 40 N/m2 force density. Similarly, magnetostatic energy density is given as, WM ¼

1 2 B 2m0

ð3Þ

where m0 is the permeability of free space and B is the magnetic flux density. Assuming that an attainable amount of magnetic flux density in a ferromagnetic material can go up to 1.6 T, the stored energy density in a free space can provide force density as much as 1  106 N/m2, which is at least four orders of magnitude higher than that of the electric case. This is the main reason for using magnetic materials in the EMEC devices.

4.14.4

Fundamentals of Electromechanical Energy Conversion

The energy is converted from mechanical to electrical or vice versa by the use of electromechanical devices. Electrical energy can be transmitted, utilized, and controlled more easily, reliably, and efficiently compared to other forms of energy. From the customer’s point of view, energy conversion devices are expected to convert various forms of energy into electrical energy. This conversion provides useful forms, such as light, sound, and mechanical energy. The energy conversion process is a reversible process, except for energy losses. All of the energy conversion devices operate on the same principles, but their constructional features are different. The coupling between the electrical and mechanical systems of these devices is through the magnetic or electric field. As mentioned before, according to the energy conservation the energy can neither be created nor destroyed; it can only be converted from one form to another. In an energy conversion device, out of the total input energy, some energy is converted into the required form, and the rest is dissipated. In the following, the conversion of the energy with the device applications will be further explored.

4.14.4.1

Coupling Field Reaction

The interaction between electrical and mechanical parts in an EMEC system is achieved through the medium of energy stored in the coupling field. In other words, the coupling field is the link between the electrical and mechanical systems. A conceptual map of an electromechanical system modeling is shown in the diagram in Fig. 3, where the circuit equations are obtained from Kirchhoff voltage or current laws, while on the other hand the force and torque equations are obtained from Newton’s law. As this diagram indicates, the electromechanical system may go along with the path that can either end up with motoring or generating action. In order for a moving part be able to rotate or move with respect to the stationary part of a machine, an air gap is inevitable in between the parts. The energy stored in the coupling field must produce action and reaction on the electromechanical systems for the conversion of energy from electrical to mechanical and vice versa. If the output is mechanical as in motor, the coupling flied must react with the electrical system in order to receive electrical energy from it. Assume that a conductor of length l is placed in a uniform magnetic field of flux density B. When a conductor moves at a speed v, the induced electromotive force (EMF) e in the conductor can be determined from the following expression, e ¼ lvB

ð4Þ

608

Electromechanical Energy Conversion

Electromechanical system

Electrical system

Magnetic system

Mechanical system

Voltage and current

Magnetic flux

Position, speed and acceleration

Circuit equations

Force /torque

EMF

Motoring

Generating

Force and torque equations

Fig. 3 Conceptual map of electromechanical system modeling. EMF, electromotive force.

The force on a moving particle of electric charge q in a magnetic field is given by Lorentz's force law as F ¼ qðv  BÞ

ð5Þ

From the definition of current (I¼ dq/dt), the force acting on a current-carrying conductor can be directly derived from Eq. (5) as, F¼I

Z

dl  B

ð6Þ

C

For a homogeneous conductor of length l carrying current I in a uniform magnetic field B, the above expression can be reduced to F ¼ Iðl  BÞ

ð7Þ

In a rotating system, the torque about an axis is given as tT ¼ r  F

ð8Þ

where r is the radius vector from the axis towards the conductor.

4.14.4.2

Conservation of Energy Principle and Energy Balance Equation

The above-given force and torque expressions are fundamental equations of the energy conversion. By using these formulations energy conversion can be realized in various forms. Energy conversion takes place between pairs of the energy forms, such as electrical–electrical, electrical–chemical, electrical–thermal, electrical–optical, electrical–sound, and electrical–mechanical. Electrical–mechanical conversion is the main focus of the discussion here. Electromechanical devices convert electrical energy into mechanical energy and vice versa. Energy conversion takes place through the medium of electric field or magnetic field coupling in terms of the force and torques given above. The use of the magnetic field as a coupling medium between electrical and mechanical systems is common in almost all of the EMEC devices. There are three basic principles associated with all electromagnetic devices: induction, interaction, and alignment. The induction in an electromagnetic device is governed by Faraday’s law: the voltage (e) induced between the terminals of a device circuit equals the time derivation of the magnetic flux f, given with the following expression, e¼

df dt

ð9Þ

where the minus sign indicates that the polarity of the induced voltage is established in opposition to the magnetic flux change, which is also known as Lenz’s law. The interaction principle in an electromechanical device can be visualized in terms of the following configuration. A flux density created by an arbitrary source interacts with another magnetic flux density that is produced from a current-carrying conductor. This configuration develops the resultant flux density as shown in Fig. 4(A). As seen in Fig. 4(B), the flux density is not homogeneously distributed in the neighborhood of the conductor, which creates the resultant flux density greater on one side

Electromechanical Energy Conversion

B

609

F

IF

I F

(A)

(B)

(C)

Fig. 4 Induction, interaction, and alignment of current-carrying conductor in magnetic field: (A) conductor is free to move, (B) conductor is fixed, (C) loop of conductor.

Ic

Stator Poles

 i

+

 N Rotor

(A)

(B)

lg

Ni –

Rg

Rg

(C)

Fig. 5 Composite structure: (A) rotating machine. (B) Magnetic core with air gap. (C) Magnetic equivalent circuit. Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004

Electrical input

Electrical system

Field coupling

Electrical losses Field losses

Mechanical system

Mechanical output

Mechanical losses

Fig. 6 Energy conversion flow diagram.

compared to the other. As a result of this configuration, the direction of the developed mechanical force tends to restore the field to its original undisturbed configuration. This interaction creates mechanical force (governed by Eq. (7)) to be used in an electromechanical conversion device to initiate translational motion, such as in actuators and relays. Consider now the flux density B of an undisturbed uniform field shown in Fig. 4(C), in which the introduction of a currentcarrying conductor loop imposes a corresponding field component, developing the resultant torque (governed by Eq. (8)) to initiate the rotational motion, such as in motors or generators. In the neighborhood of the conductor, as seen in Fig. 4(C), the resultant flux density is greater than B on one side and less than B on the other. This configuration of the flux distribution stores PE, meaning that if disturbed we get back the energy we put into the system. For example, transformer operation is based on induction as given in Eq. (9). Most of the rotating electrical machines use this principle of induction. In structures, such as rotating electrical machines or transducers, the air gap is inevitably required and associated magnetic circuit is modified by taking the reluctances into account. A representation of rotational case with the air gap is given in Fig. 5(A), the air gap introduced into a magnetic core is given in Fig. 5(B), and the circuit representation is given in Fig. 5(C). In a rotating machine the poles are formed by either permanent magnets (PM) or current-carrying conductors with the number of turns N, which is also the case in the magnetic circuit given in Fig. 5(B). In both cases of Fig. 5(A) and (B) field coupling occurs through the gap. In the circuit representation the source is represented by magneto motive force (MMF) and the reluctances of the core and gap are connected in series with each other. The EMEC system, as shown in Fig. 5(A), contains three fundamental components called a mechanical system, a coupling field, and electrical system. The mechanical components in the EMECs can be connected to the electrical system through field coupling and associated losses occur between the conversion processes. Considering the flow diagram given in Fig. 3, the energy flow from electrical to mechanical terminals of the system can be represented with the configuration given in Fig. 6. For now, only the construction mechanism of the conversion system is discussed and the losses associated with the electrical, field, and mechanical will be discussed in the next section.

610

Electromechanical Energy Conversion

EMEC devices are categorized in terms of their operating principle. The first category of devices, involving small motions, processes only low-energy signals from electrical to mechanical or vice versa. These are the conversion devices, such as microphones, loudspeakers, and transducers. The second category consists of force- or torque-producing devices with limited mechanical motion. These are electromagnets, relays, and actuators. The third category includes continuous energy conversion devices like motors and generators. These are used for bulk energy conversion and utilization. All of the EMEC devices operate according to the energy conservation stated by the second law of thermodynamics. Electric generators convert mechanical energy into electrical energy, while electric motors convert electrical energy into mechanical energy. An EMEC system is basically divided into three parts: mechanical, field coupling, and electrical. The principle of energy conversion is formalized based on the energy conservation principle. The energy transfer equation for generator action can be written as, ! !  !  Mechanical Electrical Total Losses ¼ þ þ ð10Þ energy input energy output energy losses in field Eq. (10) indicates that a prime mover forms the mechanical energy and this energy is converted into electrical energy through the field coupling. The energy transfer equation for motoring action can be written as, ! ! ! ! Electrical Mechanical Stored energy Total ¼ þ þ ð11Þ energy input energy output by field energy losses Eq. (11) indicates that the electrical energy input provides a mechanical energy through the field coupling. During the energy conversion a certain amount of energy losses occur, which are core losses (or iron losses), electrical losses (or copper losses), and mechanical losses. In this case the energy balance equation given in Eq. (11) can be rewritten as 1 0 0 1 1 0 Electrical energy Increase in stored Mechanical energy B input from source C B C C B ð12Þ A ¼ @ output þ mechanical A þ @ field energy @ A copper losses þcore loss losses

Based on the energy balance equation, assume that an increment of electrical energy dWe (excluding electrical loss) flows to the system in differential time, there will be a differential energy supplied to the field dWf (in stored form or loss), and a differential amount of energy dWm will be converted to mechanical form (in useful form or a loss). This can be formulated in the following equation as, dWe ¼ dWm þ dWf

ð13Þ

Since the core, windage, and friction losses are small, they can be neglected. Thus dWf will represent the change in the stored field energy and similarly, all of dWm will be available as useful mechanical energy output. Eq. (13) is the starting formulation for any EMEC device design that inherits the force and torque expressions given in Eqs. (7) and (8), respectively, and Faraday’s law of induction given in Eq. (9). Following is an overview of the various devices focused on conversion.

4.14.5

Electromechanical Energy Conversion Devices

In order to provide a complete classification of EMEC devices given in the previous section the formal definition of the conversion process needs to be explored further. As previously given, the electromechanical conversion devices are classified as transducer, force producing, and continuous conversion devices. In addition to the above classifications, even though there is actually no energy conversion in the transformers, these devices play an important role in providing proper working conditions in the conversion process. Besides, the analysis of these devices is almost identical with the induction machines. For this reason it is customary to investigate transformers together with general conversion devices. As mentioned earlier in the text, the energy conservation principle imposes that energy cannot be created or destroyed; it can only be converted from one form of the energy to another. Consider now the electromechanical system with a turn number N, current i, gap g, and magnetic core length l in Fig. 7(A). The movable part can be held in static equilibrium by the spring. If the core loss is neglected, all the incremental energy is stored as incremental field energy. Assume now that the movable part is held stationary at some air gap and the current is increased from zero to a value i, flux linkage will increase from zero to l, and the energy stored in the field will be [25], Z l idl ð14Þ Wf ¼ 0

This integral represents the area between the l axis and the curve representing l i characteristic, the entire area shown shaded in Fig. 7(B). Another expression of the field energy can be derived from Ampere’s law in terms of volume of magnetic material and the gaps considering, Ni ¼ Hc lc þ 2Hg lg

ð15Þ

Electromechanical Energy Conversion

611

g Movable part

 i e

λ

dWf

Fm Spring Core

i (A)

(B)

Fig. 7 (A) Electromechanical system. (B) Corresponding l–i characteristic. Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004.

where Hc and Hg are magnetic field strength in the core and gap, respectively. lc and lg are the core length and gap, respectively. Since the flux linkage is the product of number of turns and magnetic flux f, with f¼ BA, where A is the cross section of core and B is the magnetic flux density, Eq. (16) can be modified as, Wf ¼ Wfc þ Wfg

ð16Þ

where, Wfc and Wfg are the field energy in the magnetic material and the field energy in the air gap, respectively and they are given as, Z Wfc ¼ Vc Hc dBc ð17Þ Wfg ¼ Vg

1 2 B 2m0

ð18Þ

where Vc and Vg are the volume of magnetic material and volume of the air gap, respectively. The energy stored in the air gap is much larger than that of the magnetic material. For a linear magnetic system with B ¼ mH, energy in the magnetic material is obtained as, Wfc ¼ Vc

1 2 B 2mc c

ð19Þ

where mc is the permeability of the core and the field energy of the system given in Fig. 7 can be obtained from Eq. (18). Assuming that the movement has occurred very quickly, the flux linkage has remained essentially constant, during the motion the mechanical work done is in fact the decrease in the field energy. Now, a mechanical force Fm causing a differential displacement dx is expressed as, ∂Wf ðl; xÞ Fm ¼ ð20Þ l ¼ constant ∂x

Above force expressions given in Eqs. (19) and (20) can be implemented for the linear EMEC systems, such as relays and actuators. Particularly, considering the electromagnetic system given in Fig. 7(A), if the reluctance of the magnetic core path is negligible compared to that of the air gap, the flux linkage will vary with the current linearly. In this case, flux linkage will be the product of inductance L and current i, and Eq. (16) will provide the following simplified formula for the field energy, 1 LðxÞi2 2

ð21Þ

1 2 ∂LðxÞ i 2 ∂x

ð22Þ

Wf ¼ and of course force is obtained as, Fm ¼

For the case of rotating systems the expressions given in Eqs. (21) and (22) need to be implemented in terms of rotational motion, such as motors and generators. Most of the electromechanical energy converters produce rotational motion. The essential part of a rotating electromechanical system is shown in Fig. 8, where the fixed part of the system is called the stator, and the moving part is called the rotor. Let us consider a general case in which both stator and rotor have windings carrying currents, in which the current can be fed into the rotor circuit through brushes and slip rings. The stored field energy Wf of the system can be evaluated by establishing the similar analogy provided by Eq. (16). For a linear magnetic system the flux linkages of the stator and rotor windings can be expressed in terms of self-inductance of stator winding Ls, self-inductance of rotor winding Lr, and mutual inductance between stator and rotor Lsr, whose values depend on the angular position (y) of the rotor.

Wf ¼

1 2 1 2 1 Ls i þ Lr i þ Lsr is ir 2 s 2 r 2

ð23Þ

612

Electromechanical Energy Conversion

Stator is

ir  m

Rotor

Fig. 8 Basic configuration of a rotating electromagnetic system. Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004.

Flexible diagram S Magnet Sound wave

Moving coil

N Former S

Electrical output Fig. 9 Electromechanical energy conversion (EMEC) device: microphone as sound transducer.

Following the procedure used to determine an expression for force developed in a translational actuator, it may be shown that the torque developed in a rotational electromagnetic system is ∂Wf ði; yÞ tT ¼ ð24Þ i ¼ constant ∂y tT ¼

1 2 dLs 1 2 dLr 1 dLsr þ i þ is ir i dy 2 s dy 2 r dy 2

ð25Þ

The first two terms on the right-hand side of Eq. (25) represent torques produced in the machine due to variation of selfinductance with rotor position. This component of torque is called the reluctance torque. The third term represents torque produced by the variation of the mutual inductance between the stator and rotor windings.

4.14.5.1

Transducers

Transducers are devices that are used for obtaining signals for measurement systems. A microphone, shown in Fig. 9, is a typical transducer that produces an electrical output signal proportional to the sound wave acting upon its flexible diaphragm as input. The output signal from a microphone is an analogue signal of a voltage or current that is proportional to the amplitude of a sound wave. A sound wave hits the flexible diagram of the microphone and this mechanical movement is converted into an electrical signal through the inductance variation in the magnetic system of the microphone. The construction of a loudspeaker resembles that of a microphone, but in a reverse analogy. Unlike the microphone, which is a pressure-sensing device, the loudspeaker is classified as a pressure-generating device.

4.14.5.2

Transformers

Transformers are static electromagnetic devices that transfer the electrical energy from one voltage level to another. The rated power may vary from a few to 109 VA and voltage may vary from a few volts to 1000 kV. The operation frequency may also vary up to a few hundred hertz (Hz). The main applications of transformers are to increase voltage in power transmission and to decrease

Electromechanical Energy Conversion

613

 i2

i1 v1

N1

v2

N2

Load

Fig. 10 Ideal transformer and load.

voltage in distribution lines. Transformers are also used in electronic circuits of current- and voltage-measuring devices where isolation is needed. A transformer consists of two or more electric circuits and a common magnetic circuit. Transformers that use ferromagnetic materials as magnetic coupling to provide high flux densities are called iron core transformers. In order to simplify the transformer analysis an ideal transformer approach is generally applied. In this approach the winding resistance is neglected, and the leakage flux and core losses are not taken into account. In addition, the permeability of the core material is assumed to be infinite. In reality all of the parameters are taken into account for a practical transformer. As indicated in Fig. 10, if a time-varying voltage v1 is applied to the primary winding of the transformer, a voltage e1 is induced in the primary winding and is equal to applied voltage if the winding resistance is neglected. v 1 ¼ e1 ¼ N

df dt

ð26Þ

The flux also induces a voltage e2 in the other winding (secondary winding), which is the same as output voltage. v 2 ¼ e2 ¼ N

df dt

ð27Þ

From the ratio of Eqs. (26) and (27) we have a useful relation that leads to a turn ratio a, v1 N1 ¼ ¼a v2 N2

ð28Þ

Thus an ideal transformer transforms the voltage in terms of the turn ratio of its windings. If a load is connected to the secondary of the transformer, a current i2 sets MMF in the secondary. Since the core permeability is assumed, a very large net exciting MMF acting on the core is negligible. In other words, a compensating primary MMF must result to cancel that of the primary providing that, N1 i1 ¼ N2 i2

ð29Þ

For sinusoidal supply voltage, the above equations are generally written for root mean square (RMS) or phasor values. In the text the RMS values of the quantities will be indicated with capital letters. From Eqs. (28) and (29) following power balance expression is obtained. v1 i1 ¼ v2 i2

ð30Þ

The above expressions indicate that the voltages are transformed in the direct turn ratio, the currents in the inverse turn ratio, and instantaneous power input to the primary equals the instantaneous power output from the secondary. In a practical transformer, the windings have resistance and not all the windings link the same flux. The permeability of the core material is not infinite and the core loss occurs when the core material is subject to time-varying flux. For real analysis of transformers, the above mentioned factors and some of the assumptions must be considered. In a practical core material having finite permeability, a magnetizing current Im is required to establish a flux in the core material, which also creates the effect of magnetizing reactance Xm. Another imperfection in the transformer is the core loss in the magnetic core material, which can be represented by a resistance Rc. As indicated in the equivalent circuit in Fig. 11, a practical transformer is equivalent to an ideal transformer plus external impedances that represent imperfections of an actual transformer. In this circuit representation, R1 and R2 are respectively the resistances of primary and secondary winding. If the effect of leakage reactance Xl is taken into account the transformer windings are coupled by a mutual flux. Leakage inductances due to leakage flux for the primary and secondary side are, Ll1 ¼ N1

fl1 i1

ð31Þ

Ll2 ¼ N2

fl2 i2

ð32Þ

The rating value of a transformer is generally written on its nameplate and indicates what the transformer is designed for. The turn ratio of the transformer is calculated as a ¼ N1/N2 from the voltage ratio.

614

Electromechanical Energy Conversion

I1

R1

I1

X1

I2 Ic

R2

Im N1

v1 Rc1

X2

N2

v2

Xm1

Fig. 11 Equivalent circuit for a transformer. Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004.

The model parameters of R1, X1, Rcl, Xm1, R2, X2, and turn ratio of N1/N2 shown in Fig. 11 must be known so that the equivalent circuit model can be established. These model parameters can be calculated if the dimension and magnetic properties of the materials used in the transformer are known. The winding resistances R1 and R2 can be calculated from the resistivity and the dimensions of the copper wire. The magnetizing inductances Lm can be calculated from the number of turns of the windings and reactance of the magnetic path. These parameters can also be determined easily by performing two main tests on a transformer with a little power consumption, such as open circuit and short circuit tests. The primary current is the exciting current and the losses measured by the wattmeter are essentially the core losses, which depend on the maximum value of flux in the core. Any equipment in the conversion process should operate at high efficiency. The efficiency is closely related to the input and output power of the device, which also depends on the losses during the operation. In this respect, the losses in transformers are small. In a well-designed transformer the efficiency can be as high as 99%. The efficiency in a transformer device is defined as the ratio of the output power (Pout) to input power (Pin). Since the input power is the sum of output power plus the losses due to core losses (Pc) and copper losses (PCu), the general form of efficiency can be stated as, Z¼

Pout Pout þ Pc þ PCu

ð33Þ

The copper loss PCu can be written in terms of winding currents and their resistances, while the copper loss is a function of the load current. In other words, the core loss depends on the peak flux density in the core, which in turn depends on the voltage applied to the transformer. Since a transformer remains connected to an essentially constant voltage, the core loss is almost constant and can be obtained from the no-load test of a transformer. Therefore, if the parameters of the equivalent circuit of a transformer are known, the efficiency of the transformer under any operating condition can be determined.

4.14.5.3

Rotating Electrical Machines: Motor-Generators

Most AC power used today is generated, transmitted, and distributed as three-phase power. The three-phase power system has many advantages, such as efficiency, power capacity, and constant power. Three-phase systems are more efficient as the weight of conductors and other elements are much less than that in a single-phase system. With a three-phase system, more energy can be produced, transmitted, and consumed with higher power capacity. A three-phase balanced system can deliver constant instantaneous power. This in turn makes the three-phase motors have nonzero starting torque unlike their single-phase counterparts. Electric machines are used almost everywhere in industry, households, traction, vehicles, ships, aircrafts, military equipment, medical equipment, and agriculture. Electrical machines that are used extensively for EMEC are classified as DC and AC. Energy conversion in electric machines is based on two electromagnetic phenomena: 1. Generator action: when a conductor moves in a magnetic field, voltage is induced in the conductor. 2. Motoring action: when a current-carrying conductor is placed in a magnetic field, it experiences a mechanical force. Any electric machine can operate as a motor or generator. The “voltage induction” or “mechanical force” of two effects occur simultaneously whenever energy conversion takes place from electrical to mechanical or vice versa. In a motoring action, the electrical system makes current flow through conductors that are placed in the magnetic field. The armature of the machine is connected to an AC or DC supply depending on the type of the machine. The output of the machine in motoring action is the mechanical energy. A force is produced on the conductors immediately as a result of this action. If the conductors are placed on a structure that is free to rotate, an electromagnetic torque will be produced, tending to make the rotating structure rotate at certain speed. In generating action, on the other hand, this process is reversed. If the conductors rotate in a magnetic field, a voltage will be induced in each conductor. In this case, the rotating structure, the rotor, is driven by a prime mover (such as gas or steam turbine, hydro turbine, or diesel engine). The output of the machine is electrical energy. A voltage will be induced in the conductors that are rotating with the rotor. If an electrical load is connected to the winding formed by these conductors, a current will flow, delivering electrical power to the load. Moreover, the current flowing through the conductor will interact with the magnetic field to produce a reaction torque, which will tend to oppose the torque applied by the prime mover. Note that in both motoring and generating actions, the coupling magnetic field is involved in producing a torque and an induced voltage.

Electromechanical Energy Conversion

615

As previously introduced in Eq. (4), if a conductor with a length of l moves at a linear speed v in a magnetic field B, the voltage e is induced in the conductor. Note that B, l, and v are mutually perpendicular with each other. The polarity of the voltage can be defined by the right-hand screw rule. If a right-hand screw is turned in the same way the motion of the screw will indicate the direction of positive polarity of the induced voltage. For the current-carrying conductor i with a length l placed in the magnetic field, the force experienced by the conductor is given by Eq. (6). Note again that B, l, and i are mutually perpendicular. The direction of the force can be defined by right-hand rule as well. Both Eqs. (4) and (6) represent the fundamental formulation of the generating and motoring actions, respectively. There are two major components of an electric machine: stator and rotor, which are separated by air gap. The stator is the part of the machine that does not move and generally is the outer part. The rotor is the part of the machine that is free to move and generally is the inner part. Both parts are made of ferromagnetic materials, such as certain alloys of iron. If the flux is time varying in the stator or rotor (or both), the iron core is laminated to reduce eddy current losses. The main reason for the use of the iron core is to maximize the coupling between the coils placed on the stator and rotor, which in turn increase the flux density in the machine. The conductors placed in the slots of the stator or rotor are interconnected to form windings. The winding in which voltage is induced is called the armature winding. The winding through which a current is passed to produce the primary source of flux in the machine is called the field winding. The three basic and most common rotating machines are DC machines, induction machines (asynchronous), and synchronous machines. There are also other kinds of machines, such as PM machines, hysteresis machines, and stepper machines.

4.14.5.3.1

DC machines

DC machines can operate as either a generator or a motor. In DC machines the field windings are placed on the stator and armature windings in the rotor. The use of DC machines as generators is limited because of widespread use of AC power in generation. DC machines are mostly used as motors in industry, because of their motor characteristics and easy control. Large DC motors are used as conveyors, fans, pumps, paper mills, textile mills, cranes, and many more examples. Small DC machines are used primarily as control devices, such as tachogenerators for speed sensing and servomotors for positioning and tracking. In a DC machine, the field winding is placed on the stator and the armature winding on the rotor. A schematic cross-sectional view for a two-pole DC machine (field and armature windings are not displaced) is shown in Fig. 12. The DC current is passed through the field winding to produce flux in the machine. Voltage induced in the armature winding is alternating. A mechanical commutator and a brush assembly function is a rectifier or inverter, making the armature terminal voltage unidirectional. A more detailed schematic configuration of a DC machine is given in Fig. 13. The brushes are placed so that when the sides of an armature turn (or coil) passes through the middle of the region between the field poles, the current through it changes direction. This makes all the conductors under one pole carry current in one direction. As a consequence, the MMF due to the armature current is along the axis midway between the two adjacent poles, called the quadrature (or q) axis. The brushes are placed on the q-axis. Until now, we have only mentioned various parts related to the electrical machines. In order to be concise the definition of certain parts will help to understand the electrical machines. In a general electrical machine a turn consists of two conductors

N

Armature

S

Fig. 12 Schematic cross-sectional view for a two-pole DC machine (windings are not displaced). Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004.

616

Electromechanical Energy Conversion

q-axis

Field winding

d-axis

S

N

Compensating winding Armature winding

Fig. 13 Schematic configuration of a direct current (DC) machine. Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004.

End connections Conductors

S

F S

Turn

S1

F

Coil

F1

SN

FN

Winding

Fig. 14 The turn, coil, and winding in an electrical machine. Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004.

connected to one end by an end connector. A coil is formed by connecting several turns in series. A winding is formed by connecting several coils in series. The turn, coil, and winding are shown schematically in Fig. 14. The beginning of the turn, or coil, is identified by the symbol S, and the end of the turn or coil by the symbol F. The distance between the centers of two adjacent poles is known as pole pitch or pole span. The two sides of a coil are placed in two slots on the rotor surface. The distance between the two sides of a coil is called the coil pitch. If the coil pitch is a one-pole pitch, it is called a full-pitch coil. If the coil pitch is less than a one-pole pitch, the coil is known as a short-pitch (or fractional pitch) coil. Short-pitch coils are desirable in AC machines for various reasons. The DC armature winding is mostly made of fullpitch coils. Most DC machines, particularly large ones, have more than two poles. For the four-pole machine, as the rotor completes one rotation (cycle), two cycles of flux and induced voltage variations are encountered. Mechanical degree and electrical degree are not equal if the machine has more than two poles. For a p-pole machine, ye ¼

P ym 2

ð34Þ

The angle between centers of adjacent poles is 180 degree (electrical). In other words, if coil sides are placed 180 degree electrical apart, the coil is said to be full‐pitch. When the coil pitch differs from 180 degree, the winding is called a fractional pitch.

4.14.5.3.2

AC machines

AC Machines are often constructed in three phases and are usually classified as induction and synchronous machines. Windings of electric machines are usually made of copper. Winding configurations are depending on the machine type and number of phases.

Electromechanical Energy Conversion

617

Stator winding

Squirrel cage rotor Fig. 15 Single-phase squirrel cage induction motor.

4.14.5.3.2.1 Induction machine The most widely used machines are the induction machines (asynchronous). The rotor winding current is obtained from the stator current by induction, thus, the machine is called an induction machine. They are mostly used in washing machines, refrigerators, dryers, and fans. The stator windings serve as both armature windings and field windings in the induction machine. When the stator windings are connected to an AC supply, flux is produced in the air gap and revolves at a fixed speed known as synchronous speed. This revolving flux induces voltage in the stator windings as well as in the rotor windings. If the rotor circuit is closed the current flows in the rotor winding and reacts with the revolving flux to produce torque. The steady-state speed of the rotor is very close to the synchronous speed. There are two types of rotor windings: squirrel cage type and wound rotor type. In the squirrel cage type, shown in Fig. 15, rotor winding is made from aluminum or copper bars of which both sides are short circuited by end rings. In the wound rotor type, rotor winding (same form with the stator winding) terminals are connected to slip rings and brushes so that rotor winding can be connected to an external circuit. If the stator windings are connected to a three-phase supply and the rotor circuit is closed, the induced voltages in the rotor windings produce rotor currents that interact with the air gap field to produce torque. If the stator winding is connected to the three-phase supply, induced voltage produces currents in the closed-rotor windings. When the motor reaches the steady-state operation, rotating speed n reaches a value that is less or greater than synchronous speed ns. The synchronous speed is the speed of the rotating field and it does not change with the operating conditions. The difference between the rotor speed n and synchronous speed of the rotating field ns is called slip (s) and is defined as, s¼

ns

n ns

ð35Þ

If you were sitting on the rotor, you would find that the rotor was slipping behind the rotating field by the slip round per minute (RPM) ¼ ns–n¼ sns. The frequency f2 of the induced voltage and current in the rotor circuit will correspond to this slip rpm, because this is the relative speed between the rotating field and the rotor winding. In a p-pole machine, synchronous speed (the speed of rotating field) can be written as ns ¼

120f p

ð36Þ

where p is number of poles, f is the supply frequency. The frequency (also called slip frequency) of the voltage and current induced in the rotor in terms of the stator frequency f1 is f2 ¼ sf1

ð37Þ

4.14.5.3.2.2 Operation modes The induction machine can be operated in three modes: motoring, generating, and plugging. If the stator terminals are connected to a three-phase supply, the rotor will rotate in the direction of the stator rotating magnetic field. This is the motoring mode of operation of the induction machine. The steady-state speed n is less than the synchronous speed ns. In the generating mode the DC motor can be adjusted so that the speed of the system is higher than the synchronous speed and the system rotates in the same direction as the stator rotating field. If the DC motor is adjusted so that the system rotates in a direction opposite to the stator rotating magnetic field, the torque will be in the direction of the rotating field but will oppose the motion of the rotor. This is known as plugging mode. This mode of operation is sometimes utilized in drive applications where the drive system is required to stop very quickly.

618

Electromechanical Energy Conversion

Stator core loss Rotor core Rotational loss loss Pin Electric power (A)

Pout Mechanical power Stator Cu loss

Rotor Cu loss

Rotational Rotor core Stator loss loss core loss Pin Mechanical power

(B)

Pout Electric power Rotor Cu loss

Stator Cu loss

Fig. 16 Power flow in an induction machine: (A) motoring action and (B) generating action.

4.14.5.3.2.3 Power flow and efficiency In order to determine the efficiency of an induction machine as an EMEC device the losses must be taken into account. A schematic representation of the power flow in an induction machine is illustrated in Fig. 16(A) for motoring action and Fig. 16(B) for generating action. In motoring action the input is the electrical power and output is the mechanical power, while on the other hand, the input is the mechanical power and the output is the electrical power. In both cases there is core loss, rotor core loss, rotational loss, stator copper loss, and rotor copper loss. For a three-phase induction machine, the power input to the stator is Pin ¼ 3V1 I1 Cosy

ð38Þ

where V1 is the induced voltage in the stator and y is the phase angle of the stator current I1. The power loss in the stator windings is P1 ¼ 3I12 R1

ð39Þ

where R1 is the AC resistance of each phase winding at the operation frequency. The power is also lost as hysteresis and eddy current loss in the magnetic materials of the stator core, which is discussed in more detail in Section 4.14.6. Other part of the loss occurs in the resistance of the rotor circuit as, P2 ¼ 3I22 R1

ð40Þ

where R2 is the AC resistance of the rotor winding. Power is also lost in the rotor core. Because the core losses are dependent on the frequency f2 of the rotor, these may be negligible at normal operating speeds, where f2 is very low. The remaining power is converted into mechanical form. Part of this is lost as windage and friction losses, which are dependent on speed. The rest is the mechanical output power Pout, which is the useful power output from the machine. The efficiency of the induction motor is the ratio of the output power to input power, which can be expressed in terms of slip s [27], Z¼1

s

ð41Þ

4.14.5.3.2.4 Synchronous machine A synchronous machine is an AC machine; its rotor rotates with the same speed of rotating field (synchronous speed). Synchronous machines are predominantly used in power generation. They are called synchronous generators (or alternators). Synchronous generators are the primary energy conversion devices and their power ratings can reach from fractional horsepower to several hundred megavolt-amperes (MVAs). They are generally used in pumps, servomotors, and electric clocks, where constant speed is desired. In the synchronous machines the rotor has DC excited winding or PMs for establishing excitation field, slip rings, and brushes for external connection. The stator carries the armature winding. The field winding is excited by DC to produce flux in the air gap. When the rotor rotates, voltage is induced in the armature winding placed on the stator. The armature current produces a revolving flux in the air gap, whose speed is the same as the speed of the rotor, hence the name synchronous machine. Synchronous machines are often classified into two groups according to the rotor configuration: cylindrical rotor and salient pole. Cylindrical or round rotor synchronous machines have uniform air gap, 2–4 rotor poles with high speed, a rotor that usually has a small diameter but is axially long, and are usually driven by steam turbines. Salient pole synchronous machines have

Electromechanical Energy Conversion

619

Fig. 17 Synchronous machine. Schematic cross-sectional view for a two-pole machine. Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004.

nonuniform air gap, high number of rotor poles with low speed, a rotor that usually has large diameter but is axially short, and are usually driven by water turbines. The operation principle of synchronous generators relies on the poles, which are magnetized either by PMs or by a DC current. The armature, normally containing a three-phase winding, is mounted on the shaft. The armature winding is fed through three slip rings and a set of brushes sliding on them. The rotor is rotated by a prime mover. Schematic cross-sectional view for a two-pole synchronous machine is shown in Fig. 17. The synchronous machine has the field-winding wound on the rotor and the armature wound on the stator. A DC current, creating a magnetic field that must be rotated at synchronous speed, energizes the rotating field winding. The rotating field winding can be energized through a set of slip rings and brushes. In externally fed fields, the source can be a shaft-driven DC generator; several variations to these arrangements exist. The stator core is made of insulated steel laminations to minimize eddy current and hysteresis losses. The number of poles in an electric machine is always an even number. While increasing the number of poles, the rotating speed decreases. Synchronous speed is the rotating speed of the excitation field. The rotor speed is decided by the excitation frequency and the number of poles.

4.14.6

Energy Losses in Electromechanical Energy Conversion Systems

An ideal EMEC is considered as a reversible process having minimum losses in the system. Energy losses occur in EMEC devices due to various reasons and these losses take place in various forms. In the first group the energy is dissipated as heat due to electrical or copper losses, which are also known as Ohmic losses. In the second group, losses occur in the field, which are called as core losses or iron losses, which consist of hysteresis and eddy current losses. Mechanical losses occur in the last group, which consist of friction and windage losses. Core losses are important in determining heating, temperature rise and rating, and the efficiency. Since the electrical and mechanical losses are related to the material properties the core losses require more focus than from the design of efficient conversion system. Following is a review of the magnetic core materials and associated core losses in the converters. Before examining energy losses in magnetic materials the engineering materials are reviewed in general.

4.14.6.1

Engineering Materials

Technological developments in the last two centuries led newer types of conversion systems and associated magnetic material. PMs are among these materials, and are capable of establishing a magnetic field without a current flow. The fundamental purpose of a magnetic core in an EMEC device is to provide an easy path for flux to facilitate flux linkage so that the magnetic energy can be stored in a nonmagnetic low permeability region. In an inductor, the core provides the flux linkage path between the circuit winding and a nonmagnetic gap. Virtually all of the energy is stored in the gap. High permeability ferrites, or magnetic metal alloys, such as permalloy, are incapable of storing significant energy.

620

Electromechanical Energy Conversion

Engineering materials

Ferrous

Non ferrous

Plastic/rubber

Ceramics

Composites

Steel Stainless Cast iron

Aluminium Copper Zinc Titanium Tungsten

Nylon ABS Polyethylene PVC Polyester Rubber Silicone

Carbides Nitrates Oxides Graphite Diamond Glass

Metal-matrix Ceramic-matrix Laminates

Fig. 18 General classifications of engineering materials. ABS, acrylonitrile butadiene styrene; PVC, polyvinyl chloride.

Magnetic materials are used in a wide range of applications in EMEC devices, such as distribution transformers, motors, generators, magnetic switches, high frequency inductors, and more. Components made from these materials reduce operating costs, and strengthen energy conservation and application efficiency due to their extremely low core loss and high permeability. For high frequency applications, soft ferrite, amorphous, and nanocrystalline laminations are used, while for low frequency applications ferrosilicon laminations and iron–nickel alloy laminations are used. In order to obtain optimum efficiency, correct selection of engineering is necessary. A general classification of engineering materials is given in Fig. 18. The engineering materials may be grouped in terms of metallurgical and physical properties. Metallurgical properties of the materials provide tensile, ductility, hardness, toughness, fatigue, and creep. Physical (also chemical) properties of the materials include thermal and electrical conductivities, magnetic properties, corrosion, density, and melting point. In recent decades some new materials, the so-called nanomaterials, have started to emerge.

4.14.6.1.1

Permanent magnets

PMs are capable of establishing a magnetic field without a current flow. They are normally alloys of iron, nickel, and cobalt. They are characterized by high remnant field (Br) and high coercive field (Hc) and called hard materials. PMs are widely used in electromechanical devices, such as loudspeakers and PM motors. Development of PMs started in Japan at the beginning of the 19th century. A family of alloys of aluminum nickel and cobalt (called alnico) that has a high residual flux density Br (remnant field) was discovered in the 1930s. They are good in high temperature applications. In the 1950s, the ferrite PMs were placed in the market. During the 1960s a new class of magnetic materials known as rare-earth PMs was developed. The development of the rare earth magnet, called neodymium iron boron (NdFeB), came about in the 1980s. NdFeB PM has a high energy product with the most cost-effective material used in PM applications in electrical machines. Some of the major drawbacks of the NdFeB magnets are the limited temperature range and the need to protect from corrosion. Ferrite PMs are ceramic materials and are made from iron oxide and barium or strontium carbonate powders. They have less residual flux density compared to alnico and therefore they are less subject to demagnetization than alnico. Fig. 19 shows the magnetization characteristics for common PM materials [28]. The grades of PMs are generally defined in terms of their energy density, which is formally expressed with the energy product. The maximum energy product of a PM is defined as the measure of the maximum amount of magnetic energy stored. It is related to the product maximally attainable with a material made out of flux density B and field strength H. The standard unit of measurement is kJ/m³ in SI units. The maximum energy products of NdFeB PMs have the highest value, followed by alnico, samarium cobalt (SmCo), and ferrites; these can also be obtained from Fig. 19.

4.14.6.1.2

Powdered cores

Composite powdered cores, such as powdered iron, Kool Mu, and permalloy, are able to store considerable amounts of energy. Due to their ES capacity, they are used in inductor and flyback transformer applications. Since the resistivity of these metal alloys is low to minimize losses due to induced eddy currents, they are built in very thin laminated sheets.

4.14.6.1.3

Ferrite cores

Ferrites are the most popular core materials used in transformer and power supply applications. Ferrites are ceramic materials made by sintering a mixture of iron, manganese, nickel, and zinc. They are low cost and low loss materials, which are available in a wide variety of core shapes. At low frequencies, core loss is almost entirely hysteresis in ferrites, while on the other hand eddy current loss overtakes hysteresis loss in high frequency applications. For example, in metal alloy cores, eddy current loss dominates above a few hundred hertz. Some ferrites are ceramic-like compounds with very low conductivities. This is actually a required property for most applications especially at high frequencies since low conductivity limits the eddy current losses, such as in high frequency transformers.

Electromechanical Energy Conversion

621

1.5 NdFeB

1.4

Alnico SmCo

1.2

Ferrite

1.1 1.0

B (T)

1.3

0.9 0.8 0.7 0.6 –1000 –900

–800

–700

–600

–500

–400

–300

–200

–100

H (kA/m) Fig. 19 Magnetization curves for common permanent magnets (PMs). Reproduced from Fitzgerald AE, Kingsley C, Umans SD. Electric machinery. 6th ed. Boston, MA: Mc Graw Hill; 2003.

4.14.6.2

Saturation, Hysteresis in the Core Materials and Associated Losses

In EMEC devices, the magnetic circuits may be formed by ferromagnetic materials only (as in transformers) or by ferromagnetic materials in conjunction with an air medium (as in rotating machines). In most electrical machines, except PM machines, the magnetic field (or flux) is produced by passing an electrical current through coils wound on ferromagnetic materials. According to Ampere’s circuit law for magnetostatic case, the line integral of the magnetic field intensity H around a closed path l is equal to total current I surrounded by the counter. Z X Hdl ¼ I ð42Þ In order to increase the magnetic effect of the current, a coil is often wound on a magnetic core material. Considering the number of turn N, Eq. (42) states Z Hdl ¼ NI ð43Þ where NI is called as MMF. MMF is actually the external force required to set up the magnetic flux lines within the magnetic core material. Since the magnetic field intensity H produces magnetic flux density B the relationship can be analytically written as, B ¼ mH

ð44Þ

where m represents the characteristics of the magnetic material, called magnetic permeability. Permeability is analogous to the conductivity in an electrical circuit. For ferromagnetic materials the relative permeability is much higher than 100. For example, inserting a ferromagnetic material into a magnetic circuit, as shown in Fig. 20, may increase the magnetic flux density several hundred times. The configuration shown in Fig. 20 forms the basis of the magnetic circuit, which is a closed path made from ferromagnetic material, to generate flux inside by using a current-carrying coil as a source of MMF. A PM also can be used in order to drive the magnetic circuit. If the cross-section of the core is A, the flux f in Weber is defined as the number of flux lines crossing the surface area A, Z f¼ BdA ð45Þ For a smooth and homogeneous core with a length l, Eq. (45) can be written as, Hl ¼ NI

ð46Þ

f ¼ BA

ð47Þ

and also flux in Eq. (45) can be written as,

622

Electromechanical Energy Conversion

B

B

Iron core

Air core

I

I

(A)

(B)

Fig. 20 Magnetic flux lines of a coil driven by a current source. (A) Air core. (B) Iron core.

Magnetic flux

Electric current

Reluctance

Resistance emf

mmf

Fig. 21 Analogy between electric and magnetic circuits. EMF, electromotive force; MMF, magneto motive force.

Then, the flux density equation is established as f¼

NI l=ðmAÞ

ð48Þ

where the term in the denominator of Eq. (48) is called the reluctance of the magnetic core material, which is given by, ℜ¼

l mA

ð49Þ

The reluctance is analogous to the resistance in the conductor. The analogy between the magnetic circuit and electric circuit is given in Fig. 21. As indicated in Fig. 21, the driving force, current, and resistance in the electric circuit corresponds to the MMF, flux, and reluctance in the magnetic circuit, respectively. The core materials used in EMEC devices exhibit high nonlinearity. This nonlinearity can be easily seen when the mathematical relationship between magnetic field intensity and the magnetic flux density is established in a magnetic material with permeability. Magnetic field intensity can be thought of as the field force (MMF) divided by the length of the material. As shown in Fig. 22, the magnetic intensity in the core is increased in nonlinear fashion with increasing the field intensity, where the hysteresis is neglected. The curve representing the relationship in the core is also known as the magnetization curve (or B–H curve). At high values of magnetic field intensity the magnetic flux density cannot be increased as it increases at the beginning values. This effect is called saturation in the core material. In addition to the phenomenon of saturation there is also the phenomenon of hysteresis in ferromagnetic materials. The defining difference is that if saturation existed alone, the flux would be a unique function of the field intensity. When hysteresis is present, flux density for a given value of field intensity depends also on the history of its magnetic flux density, which will be discussed in the following.

4.14.6.3

Hysteresis and Associated Losses

The magnetic cores of EMEC devices are generally very good ferromagnetic materials, which behave like magnets whenever a magnetic flux passes through. Ferromagnetic materials have domain structure as shown in Fig. 23. Domains are small regions in the material, where all the dipoles are aligned in the same direction [29]. While on the other hand, the domains are distributed

Electromechanical Energy Conversion

623

B(T)

High

Saturation

Low

H(A/m) Fig. 22 Magnetic flux density as a function of magnetic field strength for a ferromagnetic core material. Reproduced from Sen PC. Principles of electrical machines and power electronics. New York, NY: John Wiley & Sons; 2004.

Magnetized domain Domain wall

Fig. 23 Magnetic domain structures. Reproduced from Cheng DK. Fundamentals of applied electromagnetics. Reading, MA: Addison-Wesley Publishing Co.; 2004.

inside the material randomly, which means that the net magnetic field of the material is zero. Whenever external magnetic field is applied to the material, the randomly directed domains are aligning themselves in parallel to the applied field. As indicated in Fig. 23 when the applied field is small (small force), only a few atoms are in alignment. However, as more flux accumulates into the same cross-sectional area of a ferromagnetic material, fewer atoms are available within that material to align their electrons with additional force, and so it takes more and more force (H) to get less and less “help” from the material in creating more flux density. When the applied field is removed, the domains come to random positions once again. However, the domains do not return to their original positions completely, leaving a remnant magnetic flux in the material. In other words, the material tends to stay magnetized after the applied field force has been removed if the force has reversed its direction, which is called hysteresis, as shown in Fig. 24. In order to understand how the domain structure is related to the hysteresis phenomenon let’s examine in detail the hysteresis curve given in Fig. 24. As shown in this figure, when the applied external magnetic field strength is increased in the magnetic material the walls of the domains (shown in Fig. 23) move in such a way that the magnetic flux density increases in the material. For relatively small applied fields up to point P1 on the B–H magnetization curve in Fig. 24, the domain-wall movements are almost reversible. However, when the applied field is increased up to point P2 the domain-wall movements are no longer reversible and the domains start to align themselves with the direction of the applied field. If the applied field keeps decreasing from point P2 the curve will go down to P20 along the broken line creating a minor hysteresis curve and ending back at P2 again. As the applied field gets even stronger, the domain motion will cause essentially a total alignment of the microscopic moments with the applied field in the whole structure. This is represented with point P3, where the material reaches saturation. The curve starting from O and tracing P2, P3, P30 and ending up at P3 again is called a magnetization curve. If the applied field is reduced to zero from the value at P3, the magnetic flux density in the material is not reduced to zero but a finite value at Br. This value is called the remnant flux density and is dependent on the maximum applied field intensity and the type of material. The existence of a remnant flux density in a ferromagnetic material makes PM possible. In order to make the magnetic flux density of the ferromagnetic material zero, it is necessary to apply magnetic field intensity Hc in the opposite direction. This required Hc is called coercive force or is known as coercive field intensity. Hc also depends on the maximum value of the applied magnetic field intensity and the type of material. The magnetic field response of ferromagnetic material all depends on the temperature of the environment. Since most of the electrical machines (except the ones used in fire hazard conditions) work in a close range of temperatures, it is not much of a concern during the EMEC. In an actual situation when the temperature of a ferromagnetic material is raised to such an extent

624

Electromechanical Energy Conversion

B

P3

Br

P2 P1

Hc

H O

P ′2 P ′3

Fig. 24 Hysteresis loop in the B–H curve.

where the thermal energy exceeds the coupling energy of the magnetic dipoles moments, the magnetized domains become disorganized. Above this critical temperature, known as the Curie temperature, a ferromagnetic material behaves like a paramagnetic substance. The Curie temperature of most ferromagnetic materials lies between a few hundred to a thousand degrees Celsius, with that of iron being 7701C [29]. As shown in the magnetization curve in Fig. 24, in one cycle the tip of the curve does not follow the same line but instead follows different paths due to the hysteresis in the magnetic material. During the process in which the material is exposed under varying magnetic field (continuous cycles) the electrical energy is lost in the material. Two types of losses appear in the magnetic core material: hysteresis loss (Ph) and eddy current loss (Pe). Hysteresis loss occurs due to the reversal of magnetization of the core whenever it is subjected to alternation of magnetizing force. In other words, whenever the core is subjected to a varying magnetic field, the domains of the material will change every half cycle. The power consumed by the magnetic domains for every half cycle is called a hysteresis loss. Energy is lost in each hysteresis cycle within the magnetic core and it is dependent on the properties of core material and is proportional to the area of the hysteresis loop. As shown in Fig. 24, throughout the whole cycle response of flux density Bmax is slower than field intensity H. The power loss due to hysteresis effect in an electrical machine is given with the following formula, Ph ¼ Kh Bnmax f

ð50Þ

where, Bmax is the maximum flux density, Kh is a constant depending on the ferromagnetic material and volume of the core. n is the number that varies in the range from 1.5 to 2.5 and f is frequency of the current [27]. Ferromagnetic materials for use in EMEC devices, such as in electric generators, motors, and transformers, should have a large magnetization for a very small applied field; they should have tall, narrow hysteresis loops. As the applied field intensity varies periodically the hysteresis loop is traced once per cycle. The hysteresis loop shown in Fig. 24 corresponds to the energy loss (hysteresis loss) per unit volume per cycle. The hysteresis loss is the energy lost in the form of heat in overcoming the friction encountered during the domain-wall motion. Ferromagnetic materials, which have tall and narrow hysteresis loops with small areas are referred as “soft” materials; they are usually well-annealed materials with very few dislocations and impurities so that the domain walls can move easily [29]. Good PMs, on the other hand, should show a high resistance to demagnetization. This requires that they be made with material that has large coercive field intensities and hence fat hysteresis loops. These materials are referred to as hard ferromagnetic materials as discussed in Section 4.14.6.

4.14.6.4

Eddy Current Losses

As mentioned in the above-given description, core losses occur when a magnetic material experiences a variable magnetic field. Eddy currents are electrical currents induced in the conducting material due to Faraday’s law. Eddy current losses can be reduced by increasing the resistivity of the core material and using laminated core when laminations are isolated from each other with thin layers. The power loss due to eddy currents depends on the frequency f and maximum flux density Bmax, which is given with the following formula [27], Pe ¼ Ke B2max f 2 where Ke is a constant and its value depends on type of material and lamination thickness.

ð51Þ

Electromechanical Energy Conversion

625

As mentioned earlier, the core losses consist of hysteresis and eddy current losses. Using a wattmeter, these losses can be measured. However, it is not easy to determine how much of the loss is due to hysteresis and how much is due to eddy currents. Since it is not necessary to know the losses separately, the losses are considered under the same procedure, such as lamination. In electrical machines that have a magnetic core and a time-varying flux, core loss occurs and the loss appears as heat in the core. These losses must be taken into account, while designing the electrical machines. As indicated in Eqs. (50) and (51) the losses are also the function of frequency. For this reason, not only the geometrical shape and type of the material but also the frequency of the application is also important for proper device designs.

4.14.7 4.14.7.1

Design of Efficient Electromechanical Energy Conversion System Thermodynamic Analysis of Energy Conversion

The interrelation between the energy source and conversion of energy procedure relies on the thermodynamic principles and laws. Based on these principles, it is necessary to establish how to apply the first law of thermodynamics (conservation of energy). The thermodynamics principles are correlated with the energy conversion process in such a way that the energy is transported across the boundary of a general thermodynamic system. According to the definition, for the closed systems (fixed mass systems) the energy can cross the boundaries only in the forms of the heat transfer and work. While on the other hand, for the open systems (control volumes) the energy can cross the boundaries not only in the form of heat transfer and work but also by the mass flow. For these forms of energy transport across the boundaries the energy interaction is established as the heat transfer if the driving force is the temperature difference between the boundaries, otherwise it is the work utilized by the conversion devices. In this case, the energy transfer across a system boundary due solely to the temperature difference between a system and its surroundings is defined as heat. Energy transferred across a system boundary that can be thought of as the energy expended to lift a weight is called work. The total energy E of a system is the sum of all forms of energy that can exist within the system, such as thermal, mechanical, kinetic, potential, electric, magnetic, chemical, and nuclear. The total energy of the system is normally thought of as the sum of the internal energy (U), KE, and PE. The internal energy is that energy associated with the molecular structure of a system. The KE exists as a result of the system's motion relative to an external reference frame. When the system having a mass m moves with velocity v the KE is expressed as, KE ¼ m

ðnÞ2 2

ð52Þ

The energy that a system possesses as a result of its elevation in a gravitational field relative to the external reference frame is called PE and is expressed as, PE ¼ mgz

ð53Þ

where g is the gravitational acceleration and z is the elevation of the center of gravity of a system relative to the reference frame. The total energy of the system is then expressed as Etot ¼ U þ KE þ PE ðkJÞ

ð54Þ

Most closed systems remain stationary during a process and, thus, experience no change in their kinetic and potential energies. The change in the stored energy is identical to the change in internal energy for stationary systems.

4.14.7.2

Energy Transport by Heat

Since the heat is an energy transition across the system boundary solely due to the temperature difference between the system and its surroundings the net heat transferred to a system is defined as [30], X X Qin Qout ð55Þ Qnet ¼

where Qin and Qout are the magnitudes of the heat transfer values. There are basically three modes of heat transfer at the boundary that depend on the temperature difference between the boundary surface and the surroundings. These are conduction, convection, and radiation. The representative picture of the heat transfer is shown in Fig. 25. When solving problems in thermodynamics involving heat transfer to/from a system, the heat transfer is usually calculated by applying the first law (the conservation of energy) to the system. Conduction type of the heat transfer is a progressive exchange of energy between the molecules of a substance. The heat flow per unit time (the rate), which is formulated by Fourier's law of heat conduction, is, Qcond ¼

Akt

dT dx

ð56Þ

where kt is the thermal conductivity, A is the area normal to heat flow, and dT/dx is temperature gradient in the direction of heat flow.

626

Electromechanical Energy Conversion

Conduction Convection

Radiation Fig. 25 Heat transfer by conduction, convection, and radiation.

Convection type of the heat transfer is the mode of energy transfer between a solid surface and the adjacent liquid or gas that is in motion and it involves the combined effects of conduction and fluid motion. The rate of heat transfer by convection is determined from Newton's law of cooling, expressed as Qconv ¼ hAðTs

Tf Þ

ð57Þ 2

where A is the heat transfer area, h is the convective heat transfer coefficient (W/m
K), Ts is surface temperature, and Tf is bulk fluid temperature away from the surface. The convective heat transfer coefficient depends on the surface geometry, the nature of the fluid motion, the properties of the fluid, and the bulk fluid velocity. Radiative heat transfer is energy in transition from the surface of one body to the surface of another due to electromagnetic radiation. The radiative energy transferred is proportional to the difference in the fourth power of the absolute temperatures of the bodies exchanging energy. Heat transfer per unit time is,
4 Qrad ¼ esA Ts4 Tsurr ð58Þ where A is surface area for heat transfer, s is Stefan–Boltzmann constant,  is emissivity, Ts is absolute temperature of surface, and Tsurr is absolute temperature of surroundings.

4.14.7.3

Energy Transfer by Work

The energy transferred by work is achieved by electrical and mechanical forms, which is the fundamental principle of the EMEC as stated before. The rate of electrical work done by electrons crossing a system boundary is called electrical power (Pe) and is given by the product of the voltage drop (V) in volts and the current (I) in amperes. Pe ¼ VI

ð59Þ

The amount of electrical work done in a time period is found by integrating the rate of electrical work over the time period. Z 2 VIdt ð60Þ We ¼ 1

Mechanical work is energy expended by force acting across a distance. Thermodynamic work is defined as energy in transition across the system boundary and is done by a system if the sole effect external to the boundaries could have been the raising of a weight. Z 2 Wm ¼ Fdx ð61Þ 1

4.14.7.4

Energy Conversion Efficiencies in Thermodynamic Systems

Efficiency is the measure of performance of a thermodynamic cycle. The efficiency of a device is defined as the measure of its performance. The efficiency is actually the ratio of the desired result to the required input. Thermodynamic cycle is defined as a process in which a working fluid undergoes a series of state changes and finally returns to its initial state. In this sense a heat engine is a work-producing device based on a thermodynamic process operating in a thermodynamic cycle. The thermal efficiency of a heat engine is defined by the ratio of the network output (the desired result) to the heat input (the required input to obtain the desired result). Heat exchangers are normally well-insulated cyclic devices that allow energy exchange between hot and cold fluids without mixing them [31]. The pumps, fans, and blowers causing the fluids to flow across the control surface are normally located outside the control surface. A heat pump is a thermodynamic system operating in a thermodynamic cycle that removes heat from a lowtemperature body and delivers heat to a high temperature body. To accomplish this energy transfer, the heat pump receives external energy in the form of work or heat from the surroundings. A refrigerator is a device that operates on a thermodynamic

Electromechanical Energy Conversion

627

cycle and extracts heat from a low-temperature medium. The heat pump also operates on a thermodynamic cycle but rejects heat to the high temperature medium.

4.14.8

Case Studies

Even though the world has already encountered four industrial revolutions, the developments and current knowledge regarding the EMEC technologies are still not at a satisfactory level. This is what drives most of the research groups and companies working on effective energy conversion technologies, especially to explore new energy sources and improve the ES methods associated with the device technologies. EMEC involves various energy sources to convert energy from one form to another. As introduced in this study, there are a vast number of energy sources utilized in the industry. All these kinds of energy forms from the sources must be converted into usable forms of energy. It is important to note that all forms of energies are mostly converted into electricity. This is a convenience for all kinds of utility applications, which makes the EMEC so important. Any kind of energy conversion requires special device applications from power electronics to control engineering. Up to now, the conversion related devices and systems have been examined and some of their fundamentals were brought to light. In addition, EMEC was examined and explored in terms of conventional conversion devices, such as transducers, transformers and rotating electrical machines. In this chapter we will focus on conversion processes of specifically renewable energies, such as solar power and its associated systems. The new aspects regarding energy conversion devices are considered in a popular fashion regarding the future of the energy. Some of the popular device applications considered in this chapter are expected to be on the market in the near future. Nanotechnology and superconductivity’s energy-related applications are of that kind. For example, apart from the medical imaging applications of superconductors, there are limited device applications that are used in public service except in laboratory purpose applications. This is due to the fact that the superconductors are still expensive to fabricate in large amounts and the cryogenics is still a problem considered by firms. However, for the future, nanotechnology and superconductivity have the potential to provide high efficiency device applications for the EMEC process.

4.14.8.1

Nanostructured Energy Conversion Devices

Nanostructures are considered as fundamental building blocks of future electronic, electromechanical, and optoelectronic nanodevices and sensors [32]. Considering the great progresses in the last two decades, researchers have been able to demonstrate unique and novel nanosystems with unprecedented functionality, such as high-mobility single nanowire transistors, strain controlled logic gates, single nanowire lasers, and strain sensors [33–35]. Due to the semiconducting and piezoelectric properties, some of the nanomaterials exhibit great advantages in the energy conversion devices and detectors. Many efforts have been made to study the piezotronic effect and various novel devices have been fabricated by utilizing piezotronic properties, such as piezoelectric field effect transistors, piezoelectric diodes, actuators, and flexible piezotronic strain sensors [36]. The piezo resistance effect is based on the change in electric conductance due to the strain induced in the material. The enhancement of piezoelectric effect is important for pushing the potential applications of the nanomaterial-based devices related to the piezoelectric effect, such as piezoelectric nanogenerators. Nanogenerators are nanodevices that harvest mechanical energies in the form of vibration into electricity. These nanogenerators can effectively convert the mechanical energy into electricity. The future of this field will be in the development of functional units or integrated nanosystems for practical application in self-powered monitoring systems, such as heart beating.

4.14.8.2

Superconducting Magnetic Energy Storage

In this section, the application of superconductors on the EMEC devices is introduced. The application of superconductors depends on their extraordinary properties, such as their conductivity, with zero resistance and the levitation because of the Meissner effect. Zero resistance provides more current in the coil, which is an advantage for SMES. On the other hand, the levitation property provides magnetic bearings and maglev train applications. Superconductors play an important role in EMEC devices. Besides the use of superconductors in magnetic imaging devices, at the present time, the only realistic application of superconductivity is the SMES. Both are based on the superconductor cable. In the last few decades the RE source penetration into the power grid market has increased. The grid structures have now started to be considered as small and smart. SMES is important in various device applications of the industry and the stable energy needs of the grid. A typical microgrid is a part of the electric distribution system, having various distributed generation units, such as wind turbines and PVs being able to operate connected to ES units, control units, and loads, as shown in Fig. 26. The scenario shown in this figure indicates that the battery and SMES are working together to support the microgrid as the ES system [37]. Superconducting magnetic ES has been studied well enough, providing a large amount of the literature [38–41]. Due to its high energy and power density, SMES is suitable to work with the battery to have an optimum storage system. While the main power is supplied from the grid to the load, the battery and SMES, as ES devices, supply transient power and peak load demand with an appropriate control strategy. The SMES can provide peak power with a faster response than the battery, but it lasts shorter than the battery. So the SMES is able to recoup some amount of peak power, and if necessary, trigger the battery to supply excess peak power. This way the life cycle of the battery can be kept high.

628

Electromechanical Energy Conversion

DC/DC converter

DC/AC converter

PV

Battery WT

Switching unit

SMES

Storage units AC/DC converter

Fig. 26 Schematic structure of a microgrid. AC, alternative current; DC, direct current; PV, photovoltaic; SMES, superconducting magnetic energy storage. Reproduced from Cansiz A, Faydaci C, Qureshi MT, Usta O, McGuiness DT. Integration of a SMES-battery based hybrid energy storage system into micro-grids. In: ICSM; 2016.

Distribution network Power electronics Liquid nitrogen

HTS coil

Liquid nitrogen tank

Cryostat Fig. 27 Superconducting magnetic energy storage (SMES) general structure. HTS, high temperature superconductor. Reproduced from Cansiz A, Faydaci C, Qureshi MT, Usta O, McGuiness DT. Integration of a SMES-battery based hybrid energy storage system into micro-grids. In: ICSM; 2016.

SMES systems store energy in the form of a magnetic field. The stored energy is equal to one half of the inductance (L) of the superconducting coil and the square of the circulating current (I), which is [18]. E¼

1 2 LI 2

ð62Þ

SMES has a fast response to changing energy flow, which provides advantages with load leveling, stability, and frequency shifting during outages [17,38]. A typical SMES structure is shown in Fig. 27 [37]. The power conditioning system provides a power electronics interface between the distributing network and the SMES, which aims to convert DC to AC and to efficiently charge/ discharge the coil. When a battery and SMES are combined to supply load demand, the SMES can provide a sudden increase in peak power with a faster response than the battery, whereas the battery lasts longer than the SMES. The configuration from the transformer to a SMES unit in Fig. 28 represents the individual circuit configuration for control of the SMES. The circuit configuration is simpler in the battery than the SMES, having a multiport DC–DC converter connected to the battery. Fig. 29 shows the system’s response to a three-phase fault when ES devices of both the SMES and battery are in use. This figure elaborates on the system would work. As soon as a three-phase fault occurs at the local bus, the SMES takes over and starts

Electromechanical Energy Conversion

629

DC-to-DC chopper IGBT Transformer C

SMES

3AC

VSC

DC link

Fig. 28 Basic configuration of voltage source converter (VSC)-based superconducting magnetic energy storage (SMES) system. AC, alternative current; C, capacitor; DC, direct current; IGBT, insulated gate bipolar transistor. Reproduced from Cansiz A, Faydaci C, Qureshi MT, Usta O, McGuiness DT. Integration of a SMES-battery based hybrid energy storage system into micro-grids. In: ICSM; 2016.

300

Voltage (V)

200 100 0 –100 –200 –300 0.96

0.98

1

1.02

1.04

1.06

Time (s) Fig. 29 System voltage following the fault while superconducting magnetic energy storage (SMES) and battery are functioning. Reproduced from Cansiz A, Faydaci C, Qureshi MT, Usta O, McGuiness DT. Integration of a SMES-battery based hybrid energy storage system into micro-grids. In: ICSM; 2016.

providing power. After the SMES voltage is dropped below the required amount, the SMES is switched off and the battery is switched on. This configuration provides an uninterruptible power flow when connected with solar PVs or wind turbine systems in Fig. 26. The results shown in Fig. 29 will create a system that deals with quality management and control algorithms to produce an energy management system to handle all kinds of voltage fluctuations that occur for various reasons in the power and conversion systems.

4.14.8.3

Superconducting Flywheels as Electromechanical Energy Conversion

As mentioned in the previous section, most of the realistic application of superconductors is in SMES devices. As the superconducting material quality improves, potential applications regarding EMEC will continue to arise; superconducting bearings and flywheels are some of them. FES is a device with a long history. It has a rotating part, usually called a rotor and stator, to support the whole system. Flywheels are energy conversion devices and store any kind of intake energy as a mechanical energy for later use. In the early times the flywheels were mechanical type. With the developments in power electronics and control methods the magnetic bearing associated with the flywheels came on the scene. Magnetic bearings are constructed in various sizes and configurations based on their particular applications. The stability and the rotation for the magnetic bearing systems are generally provided by active control. The active control not only requires extra energy but also sensors and associated complicated electronics. These extra complexities are not desired especially for device applications. From this point of view the running conditions of the bearings require focusing on the stability and driving mechanism of the rotor. The improvements on the driving mechanism of the magnetic bearings would provide more effective bearing systems, which can be constructed in a smaller size and lighter weight than that of conventional ones. From the earlier studies of levitation to the present time there have been material and design improvements in the literature.

630

Electromechanical Energy Conversion

Since the discovery of HTSs, many engineering applications have emerged relating to levitation [21,42–48]. These applications generally focus on the interactions among the PMs, superconductors, and conducting materials, such as aluminum, copper, and brass. The interaction between the superconductor and PMs is highly complex and require sophisticated analysis. However, in most cases the interaction is modeled in terms of conceptual approaches, such as the frozen image model [49,50], which is beyond the scope of this analysis. One of the most important applications is to develop an effective rotor design and driving mechanism for the superconducting magnetic bearing devices. In this respect, the bearing can serve as a preliminary investigation for the medium scale SFES systems. Flywheels are the main components of the energy conversion systems. The addition of flywheels to ES systems enables the peaks in demand to be supplied by the flywheel, thus smoothing the load on the grid. Also, they can provide a permanently receptive supply for the brake energy that is subsequently transferred back into the steady-state load on the system. This particular application is not only appropriate for ES but also for the power quality management systems where energy is stored in anticipation of short timescale dropouts [51]. FES is one of the most important applications of the HTSs. For many purposes FES has an advantage over batteries in the amount of stored energy and the speed with which it can be delivered. The use of HTS passive bearings in a flywheel also has an advantage over mechanical or active bearings. The parasitic power loss associated with the bearing is less compared to the mechanical bearings, even taking into account the cooling power. They are also reliable, require no maintenance, and they can operate at very high speeds [52]. A flywheel is a rotating mass that stores energy mechanically in the form of KE. In order to rotate the flywheel for maintaining the continuous rotating motion, usually an electric motor is used. During the rotation of the flywheel it can be thought of as a mechanical battery that has a certain amount of energy stored in its system depending on its rotational velocity and its moment of inertia. The stored energy can be retrieved by slowing down the flywheel via a decelerating torque and returning the KE to the electrical motor, which is used as a generator. The main components of a flywheel storage system are the rotor, bearing, and the power interface. A schematic description of the SFES system is shown in Fig. 30. The electricity storage system comprises two major subcomponents: storage and the power conversion electronics. In flywheel storage systems, the power conversion system is a bidirectional process that allows the DC to flow to the load after it is converted to AC and vice versa to charge the battery or flywheel. One of the drawbacks of SFES systems is that they have higher standby losses compared to batteries. The losses on the rotor are of the rotational kind, which consists of hysteresis on superconducting parts, eddy currents in any conducting parts, and some other windage losses. These are caused by inhomogeneity in the rotating magnets. Hysteresis loss would not normally be frequency dependent. However, hysteresis loss is dependent on the amplitude of the variation in the magnetic field at the superconductor, which is also dependent on the amplitude of the vibrations. Hence, if the amplitude of the vibration on the rotor was sufficiently frequency dependent then the hysteresis loss would also be frequency dependent. Generally, if a rotor is assembled from many parts it is susceptible to several potential vibrations. There can be two types of rotation due to the unbalanced rotor, which in turn creates whirl amplitude around the rotating axis. One is the cylindrical whirl, which is when the bottom and top of the rotor diverge from the center of the axis of rotation with relatively equal amplitude. The other is the conical one, which is when the bottom and top of the rotor diverge again with relatively equal amplitude but in opposite directions. These types of rotation, which are due to the unbalanced rotor, increase in amplitude when the rotor reaches the resonance frequency. For this reason adjustments must be made while designing the rotor. These involve addition or subtraction of small quantities of weight to or from the rotor mass along the direction of the rotation axis of the rotor.

Magnetic bearing

Motorgenerator power coupling

Converter AC/DC

Control unit

Spinning rotor

Vaccum housing

Speed and position Superconducting bearing

Fig. 30 Superconducting flywheel energy storage (SFES) system. AC, alternative current; DC, direct current.

Electromechanical Energy Conversion

631

Rotational frequency (Hz)

12 Rotational frequency (Hz)

10 8 6 4 2 0 0

2000

4000

6000

8000

10,000

Time (s) Fig. 31 Typical time history of spin up and free spin down of rotor. Reproduced from Cansiz A. Correlation between free oscillation frequency and stiffness in high temperature superconducting bearings. Physica C 2003;390:356–62.

Fig. 31 shows the typical time history of spin up and free spin down of frequency as a function of time. There are three regions. The first region represents driving the rotor up to speed, and in the second region the driving mechanism is separated from the rotor, which is left to spin down freely. A marked resonance at about 3 Hz separates this region from the final low speed region. The rotational loss is higher above the resonance than below the resonance due to the change from rotation about the rotor center of mass to rotation about the magnetic center. This decreases the field fluctuations at the superconductor. The frequency decay is approximately linear, indicating a hysteresis loss. During one spin down test the gap to the magnet was increased from 4 to 5 mm by raising the magnets on the central shaft. The loss at the lower height is nearly four times that at the higher. This is due to fact that the closer the rotor magnet is to the superconductor the higher the field variation experienced by the superconductor, indicating that hysteresis loss in the superconductor is the main loss mechanism. The rotational loss is also a function of the temperature of the superconductor elements. For example, when the temperature was kept at 65K and the decay rate of the rotational frequency was observed to be 0.0021 Hz/s while it was 0.0012 Hz/s where the temperature was around 45K. Again this is consistent with the hypothesis that the loss is due to hysteresis in the superconductor since the loss per cycle is inversely proportional to the critical current density. We first see whether the decay rate can be explained by the hysteresis loss in the superconductor. The distribution of magnetic field across the superconductor is not uniform and the waveform is complex, so only an estimate of the hysteresis loss can be made. However, since the hysteresis loss varies as the cube of the field amplitude it is dominated by the main fluctuation with the period of the rotor and minor loops can be ignored. We make an estimate based on the field fluctuations, which are about 45 mT and correspond to DC measurements of the field from the magnet. Assuming planar geometry the loss per cycle per unit area is WH ¼

4 ðDBÞ3 3 m20 Jc

ð63Þ

where B is the amplitude of the field variation. If the mean radius of the superconducting ring is R the total loss per cycle is WH ¼

4 ðDBÞ3 2pR 3 m20 Jc

ð64Þ

The total energy is U ¼ 2p2 mk2 f 2

ð65Þ

where k is the radius of gyration and f is the frequency. The loss in energy per cycle is WH ¼

1 dU df ¼ mk2 f dt dt

ð66Þ

Substituting for WH the decay rate is df 2R ðDBÞ3 ¼ dt 3mk2 m20 J02

ð67Þ

The effective drag force on the rotor is F¼

2pmk2 df R dt

ð68Þ

632

Electromechanical Energy Conversion

Free rotational speed of rotor (Hz)

70 Free rotational speed of rotor (Hz)

60 50 40 30 20 10 0 0

5

10

15

20

25

30

35

Time (min) Fig. 32 Free spin down frequency of the rotor as a function of time. Reproduced from Cansiz A, Yildizer I. The design considerations for a superconducting magnetic bearing system. Cryogenics 2014;63:180–5.

From the decay rate we can define a coefficient of friction (COF) as COF ¼

F 2pk2 df ¼ gR dt mg

ð69Þ

The main weight of the rotor is assumed at the circumference so we take R and k to be the same at 10 cm. The effective width of the superconducting ring is approximately 3 cm. With the given value of critical current density as 104 A/cm2 and estimated value of the magnetic field variation as 0.045 T the decay rate of the rotor, df/dt, is predicted to be 1.05  10 3 s 2. The stored energy in the flywheel is proportional with the operational speed. In recent studies [53,54], the mass of the rotor produced is rather small compared to previous projects [51,52]. However, the speed is increased up to 70 Hz to explore the dynamics of the bearing mechanism at high speed. In Fig. 32, the free spin down frequency of the rotor as a function of time is measured. In high speeds, in the absence of a perfect balance among the components of the bearing the rotor tends to run away. From this respect, the design and fabrication criteria of an Evershed type [53,54] of superconducting magnetic bearing with a stable rotor can be operated at high speeds by an efficient driving mechanism. The rotor of the bearing is accelerated with the use of a noncontact driving system up to a frequency of 70 Hz. Once the ultimate speed is obtained, the driver system is successfully removed through the control unit so that the free spin down behavior of the rotor can be obtained. The result indicates that even though there exist some mechanical and magnetic imperfections, the bearing system is able to sustain the smooth rotation and promises future applications for efficient conversion devices.

4.14.9

Future Directions

EMEC is the bridge between all kinds of energy sources and devices associated with energy generation, transmission, distribution, and storage. Considering the existing energy sources, the future is in RE. The use of RE resources for power generation is an increasing trend. The benefits of RE sources are significantly enhanced when they are integrated into electric utility systems. With greater use of smart grid enabling technologies, higher degrees and rates of penetration will be accommodated to suppress the intermittency of the RE. Several factors need to be addressed during integration, such as power quality, reliability, energy conversion cost, and power system efficiency. As industrial development spreads throughout the world, the electric power systems and associated conversion systems are faced with different challenges, such as aging infrastructure, continuous demand growth, and the necessity of integration of more RE resources. Smart systems are required to cope with environmental issues. Smart grid technologies may be able to develop a cleaner energy supply that is more energy efficient, more affordable, and more sustainable. Smart grid tools and technologies implemented in the electrical grid infrastructure enable bidirectional flows of energy and communication. As previously mentioned, renewable resources are intermittent. Due to the climate conditions they fluctuate independently from demand, such as in the solar and wind power. However, since RE sources are plentiful and not restricted to only solar and wind, it is up to the effectiveness of the conversion systems to be developed. For this reason, the future directions should be focus on more development on the conversion technologies related to RE. Their contribution to sustainable energy will depend on the development of generation, conversion, and storage methods. ES is as important as the exploration of new RE sources since none of the particular energy sources are sufficiently available to meet demand. Improvements in conversion technologies will also increase the use of more RE sources for electric ES purposes.

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Electric ES can provide power for network operation and load balancing, such as helping in meeting peak electrical load demands, providing time-varying energy management, regulating the intermittence of renewable source power generation, improving power quality/reliability, and supporting the realization of smart grids. For the storage system to be really competitive, it must have good overall efficiency. This means that, for optimum operation, the power transfer chain must have rather fewer losses in terms of energy transfer and self-discharge. The development of ES techniques requires the improvement and optimization of power electronics, often used in the transformation of electricity into storable energy, and vice versa. The rate of penetration of RE will require studies on the influence of the different storage options. The study of complete systems including ES, conversion, and generation together with the power electronics and control systems will lead to the optimization of the techniques in terms of cost, efficiency, reliability, maintenance, and social and environmental impacts. Investment in research and development on the possibility of combining several storage methods with RE sources will lead to hybrid systems in an optimization of the overall efficiency of the system. Assessing the thermal storage systems also has potential to create advantage in terms of power distribution. The development of supercapacitors and fuel cells and their integration into the different types of ES devices also have potential to increase the penetration of RE to main grids. According to the present analysis the EMEC itself and associated devices are the driving force for maintaining a stable and continuous development for the industry. Technological advances have pushed the limits of these devices for the benefit of society. According to today’s observations regarding these benefits, it is a fact that they are all the result of industrial revolutions. Based on the expectation of future industrial revolutions, the device technologies associated with energy conversion will have to exceed the limits in terms of connecting people. The technological innovations related to artificial intelligence, smart vehicles, nanotechnology, robotics, 3D printing, biotechnology, ES, and quantum computing are all the progress that the world has come to see in terms of conversion technology. Transportation and logistics will especially gain from this effective and productive technological innovation. In respect to supply, various industries are becoming aware of the new technologies, which are creating new methods of meeting current needs of research, marketing, development, sales, and lastly distribution. On the other hand, the demand is also changing with consumer engagement, new consumer behavior patterns, and transparency, which in return are forcing companies to widen their delivery of products and services. As a result of the physical and digital worlds coming together, new technologies, such as surveillance systems, via digital infrastructure will enable people to engage with bodies, such as the government. Governments have come to change their current policies regarding the distribution of new technologies for good. The fourth industrial revolution changes consumption patterns of people, who develop careers and cultivate skills in the society they live. The last revolution, however, can shape and direct a future that reflects our common objectives and values, and may have the potential to “robotize” humankind! Technology stemming from the fourth industrial revolution will have a huge effect on industries and will be the main inspiration of next coming industrial revolution.

4.14.10

Closing Remarks

EMEC systems and associated devices are common in industrial development since these devices provide the most energy conversion. The electromechanical conversion deals with not only ES systems and devices but also other types of conversion systems. In the analysis presented here the energy conversion was examined on a fundamental level, including the conversion process itself and device technologies. In the context of device applications, certain case studies regarding superconductivity were introduced. All of the conversion systems require conventional device applications, such as transducers, transformers, motors, and generators, as force- and electricity-producing devices. To be able to develop efficient and environmentally friendly devices for energy sector researchers try to establish a bridge between the fundamentals of energy conversion and engineering tools. All of these applications are improved through the development of new engineering materials, advanced design techniques, advance power electronics, converters, and inverters. Power electronics and design technologies related to energy conversion have gained significant development after several decades of progress on semiconductor devices, converters, inverters, electrical machines, motor drives, and advanced control techniques. With the lowered cost, the reduced size, and the improved performance, the development in industry has made significant impact on RE systems, specifically in energy generation, conversion, and storage. Power electronics are needed in almost all kinds of RE systems. There is a wide range of studies in the literature concerning the impact of power electronics on industrial development [55,56]. From the energy conversion perspective, one of the most important applications of power electronics on the RE systems is the operation of the converters in wind power generation systems. The converters are applied to regulate the fluctuating input power and maximize the electrical energy harvested from the wind energy. In PV systems, on the other hand, PV inverters are used for efficiently converting the DC voltage for AC applications or integration of the output energy into the electrical grid. Since the solar and wind energy sources are intermittent in nature, they strictly depend on climatic conditions during their operation of electrical energy production. For example, wind energy depends on the density and speed of the wind, and solar energy, on the other hand, depends on the availability of solar radiation. These two energy sources can be merged to form more effective electricity generation systems called hybrid systems. These kinds of merging can also be applied to ES systems. As indicated in Table 2, the ES systems are classified in terms of energy–power density and the charge–discharge time and life cycle. Some of the ES systems have high energy densities, while others may have high power densities. In order to have an efficient ES system both high energy and power density are desired in device applications. Since there is not one storage system alone that

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meets the requirements of the industry to obtain an efficient storage system the combination of high energy and high power storage elements are proposed. The advantage of such hybrid systems is an overall increase in specific power and/or specific energy. For example, battery–supercapacitor systems are important ES systems. One of the hybrid systems was discussed in Section 4.14.8.2, where the battery and SMES are integrated for the ES and power conditioning purposes. Battery and flywheel is also a good candidate for hybrid ES [57]. In Section 4.14.8.3 SFES was introduced. In principle, all of the storage systems can be combined to form hybrid ES systems as long as high power and high energy density coupling is satisfactory. The possible coupled ES systems were also discussed in detail in the literature [57]. Besides ES integration with energy conversion, the hybrid system may also combine various energy sources, such as wind turbine and PV solar panels for the purpose of energy generation, with their outputs optimized by power controllers. The extracted energy is used to charge a battery bank or supply energy to an inverter, which is connected to the consumer loads and, when it is present, to the electrical power grid. Developments in research on modeling of hybrid energy resources (PV systems), backup energy systems (fuel cell, battery, ultra-capacitor, diesel generator), and power conditioning units (converters, battery chargers) have been reviewed in the literature [58]. Finally, despite the fact that we have not described all the characteristics of the different storage techniques in a detailed fashion, we have shown the possibility of storing electrical energy whenever and wherever it is needed in any quantity. It must be noted that ES is the weakest link of the energy market and it should be considered as the key element for the growth of renewable energies. ES is inevitable since RE sources are intermittent and may be located in isolated areas, which may not be able to connect efficiently with the distribution network. Based on these facts, while developing the EMEC technology the energy sources (especially renewable) and ES methods must be considered all together.

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Further Reading Smith SC, Sen PK, Kroposki B. Advancement of energy storage devices and applications in electrical power system. In: IEEE power and energy society general meeting – conversion and delivery of electrical energy in the 21st century; 2008.

Relevant Website www.irena.org International Renewable Energy Agency.