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Design and thermodynamic and thermoeconomic analysis of an organic Rankine cycle for naval surface ship applications
Energy Conversion and Management 148 (2017) 623–634
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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Design and thermodynamic and thermoeconomic analysis of an organic Rankine cycle for naval surface ship applications Ibrahim Girgin ⇑, Cüneyt Ezgi Mechanical Engineering Department, Beykent University, Istanbul 34396, Turkey
a r t i c l e
i n f o
Article history: Received 28 February 2017 Received in revised form 23 May 2017 Accepted 11 June 2017
Keywords: Organic Rankine cycle Waste heat recovery Ship Cogeneration
a b s t r a c t This paper presents the thermodynamic modeling of an organic Rankine cycle (ORC) that uses waste exhaust energy of a 1000 kW diesel generator on a naval ship. Seven different working fluids have been selected as the ORC fluids. The commercial software (EES) has been used to predict the thermodynamic properties of the selected fluids. The efficiency of the ORC goes up to 32% with toluene. For the needed generator power of 500 kW on cruising, the ideal ORC can produce 118 kW power with the working fluid toluene. Assuming an isentropic efficiency for the turbine and the pump of the case ORC to be 0.75 and 0.20, respectively, and neglecting the losses at the ORC electric generator, the electric power output of the ORC cycle becomes 92 kW. The power of the diesel-ORC system becomes 592 kW while the combined efficiency is calculated as 0.349. The ORC saves 25,500 L of diesel fuel (US$24,870) and reduces 67.2 tons of CO2 emissions at the end of 1000 operating hours. ORC working fluids may result different efficiencies at different temperatures. Therefore, a combined ORC system is proposed to get higher efficiencies at different thermal loads. The exergy efficiencies and irreversibilities were calculated. Ó 2017 Elsevier Ltd. All rights reserved.
1. Introduction Energy is ability of a system to do work. High cost of energy and the environmental effects of energy consumption cause enormous stresses on the world. Fossil fuel has commonly been used to produce energy. But, the consumption of fossil fuels has caused many environmental problems including air pollution, acid rain, and global warming. Rankine Cycle using fossil fuel is still a dominant power supply method. Low-grade waste heat accounts for 50% or more of the total heat generated in industry, and it has generally been discarded due to lack of efficient recovery methods. Therefore, recovery of lowtemperature waste heat and renewable energy is attracting much research attentions [1]. Rankine cycle using water as working fluid does not allow efficient recovery of waste heat below 370 °C. Many problems are encountered in steam Rankine cycle when water is used as the working fluid: Superheating is needed to prevent condensation during expansion, risk of erosion of turbine blades, high pressure in the evaporator, and complex and expensive turbines [2,3]. Various thermodynamic cycles can be used for conversion of low-grade heat to work, such as supercritical Rankine cycle,
⇑ Corresponding author. E-mail addresses: [email protected] (I. Girgin), cuneytezgi@beykent. edu.tr (C. Ezgi). http://dx.doi.org/10.1016/j.enconman.2017.06.033 0196-8904/Ó 2017 Elsevier Ltd. All rights reserved.
organic Rankine cycle, and Kalina cycle. But organic Rankine cycle has the characteristic of simple structure, easy maintenance and high reliability [4]. Therefore, it has many uses including industrial waste heat [5], biomass energy [6], geothermal energy [7], and solar energy [8]. Guo et al. studied thermal efficiency, influence of recuperator and exergy destruction for a 240 MW pulverized coal-fired power plant. The analytical results show that the mixture that matches with heat sink has the greatest efficiency and the mixture that matches with heat source has the lowest superheat degree. There exists no optimal working fluid for all indicators (thermal efficiency, heat exchanger area, mass flow and volumetric flow etc.) [5]. Liu et al. presented the results of thermodynamic modeling studies of a 2 kW biomass-fired system with organic Rankine cycle. They used three different environmentally friendly refrigerants to predict the efficiency of the system [6]. Desai and Bandyopadhyay reported thermo-economic comparisons of organic Rankine and steam Rankine cycles powered by linefocusing concentrating solar collectors. They also proposed a simple selection methodology, based on thermo-economic analysis, and a comparison diagram for working fluids of power generating cycles [8]. Hung et al. investigated the effects of turbine inlet temperature and condenser outlet temperature at different pressures on the efficiency of the Rankine Cycle for different working fluids [9]. Bao and Zhao have reviewed the influence of working fluid
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Nomenclature A h I _ m
t
P R s S T U V_ w
heat transfer area (m2) enthalpy (kJ kg1) irreversibility mass flow rate (kg s1) specific volume (m3 kg1) pressure (Pa) resistance (m2 KW1) specific entropy (J kg1 K1) entropy (J K1) temperature (K) overall heat transfer coefficient (Wm2 K1) volumetric flow rate (m3 s1) specific work (J kg1)
g
efficiency (%)
Subscripts ex exergy f fouling i internal lm log mean o external ORC organic Rankine cycle P pump PP pinch point rev reversible s isentropic t turbine
Greek letters q density (kg m3) 2 effectiveness of heat exchanger
properties on organic Rankine cycle, the screening of working fluid, and the comparison of different types of expansion turbines [4]. Different types of waste heat recovery technologies can be used onboard ships, which can be turbo charging of air into the engine, absorption refrigeration, thermoelectric generation, or combined power cycles [10]. The International Maritime Organization has developed a global CO2 reduction index known as the Energy Efficiency Design Index for new ships and the Ship Energy Efficiency Management Plan for all ships. The new chapter added to MARPOL ANNEX VI Regulations for the prevention of air pollution from ships, which was implemented on January 1, 2013, aims to reduce the emission of greenhouse gases, specifically CO2 emissions [11]. Implementing CO2 reduction measures will result a significant reduction in fuel consumption. Therefore, searching for new energy conservation methods that can be applied onboard ships is necessary [12]. The potential of waste heat recovery is among the most important technologies to lower fuel consumption. Larsen et al. compared ORC, the Kalina cycle, and the steam Rankine Cycle in a combined cycle application with a large marine two-stroke diesel engine. It is concluded that the ORC has the greatest potential for increasing the combined thermal efficiency [13]. Carcasci et al. illustrated the results of the simulations of an ORC combined with a gas turbine in order to convert waste heat into electric power. Four different ORC working fluids were compared to identify best choice. Rankine cycle was optimized by varying the main pressure of the fluid at different temperatures. The possible use of a superheater was also investigated in order to increase electrical power [14]. Khaljani et al. optimized a cogeneration system consisting a gas turbine and an ORC with selected decision parameters. It is reported that the gas turbine inlet temperature has important role on the trade-off between exergy efficiency and cost criteria [15]. In this study, seven different ORC fluids have been compared on an ORC design which uses the thermal energy of a diesel engine of a naval ship. A combined ORC cycle using two different ORC fluids has been proposed to reach higher efficiencies. 2. Organic Rankine cycle (ORC) Organic Rankine cycle (ORC) applies the principles of Rankine cycle using an organic fluid that has a low boiling point to recover heat from low temperature heat sources. A simple ORC converting waste heat from exhaust gases into useful work, and a typical T-s diagram is seen in Fig. 1.
There are four main components and ideal processes of this system: An evaporator for recovering waste heat from exhaust gases, a turbine for expansion of the working fluid and producing work, a condenser for transferring heat to the environment, and a pump for increasing the pressure of the working fluid. Evaporator is the component to recover waste heat which may be in various types, such as solar heat, waste heat of flue gas, and geothermal heat. The fluid is heated in the evaporator and a phase change occurs from compressed liquid to saturated vapor. After the evaporator, the working fluid at high pressure expands in the turbine and produces electrical work at the generator connected to the turbine. According to the 2nd law of Thermodynamics, a heat engine should run between at least two heat sources to produce net work. Therefore, some amount of heat should be transferred to outside environment in condenser to change phase of the working fluid to saturated liquid at an ideal Rankine cycle. Then, the liquid is pumped into the evaporator to absorb waste heat. Even though the points 1 and 2 seem to be the same points in T-s diagrams as seen in Fig. 1, actually they are two different points. The point 2 is after the pump where the point 1 is before the pump. A higher evaporator pressure gives better efficiencies in ORC power systems. But one restriction for evaporator pressure is pinch point (PP) temperature which shall be kept above zero. Pinch point is defined as the point at which the temperature difference between the hot and the cold fluid is minimum. This point is a fundamental parameter when designing a practical ORC. Pinch value must always be positive, in order to make the heat exchange possible. A small value of pinch corresponds to a very difficult heat transfer and therefore requires more heat exchange area. The minimum temperature difference occurs in evaporator at the point of saturated liquid as seen in Fig. 2. 3. Working fluids The saturation curve is the most crucial characteristic of an ORC working fluid. This characteristic affects the cycle efficiency, fluid applicability, and the arrangement of the associated equipment in an ORC power-generation system [9]. The slope of the saturation vapor curve of a fluid in T-s diagram can be negative (e.g. water), vertical (e.g. R11), or positive (e.g. n-hexane), and the fluids are called ‘‘wet”, ‘‘isentropic”, and ‘‘dry”, respectively. Typical T-s diagrams of wet, isentropic, and dry fluids are seen in Fig. 3. Wet fluids like water need to be superheated, because as they enter the turbine in saturated vapor phase, the percentage of satu-
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625
Fig. 1. The simple ideal organic Rankine cycle.
boiling point to reach a higher efficiency [9]. Recommended fluids for different applications, working conditions, performance indicators, and heat source temperature levels have been presented in [4]. Based on these features, R-123, R-141b, isopentane, npentane, n-hexane, benzene, and toluene have been selected as working fluids for thermodynamic analysis. Since the shape of the saturation curve is so important for ORC fluids, T-s diagrams and some thermodynamic properties of the selected fluids are presented in Fig. 4 and Table 1, respectively. 4. System selection for naval surface ships
Fig. 2. Pinch points in an ORC.
rated liquid increases at the exit and it may be harmful for turbine blades, as shown in Fig. 3a. But many organic fluids, which are ‘‘isentropic (Fig. 3b) or dry (Fig. 3c), do not need superheating. The saturated vapor phase of a dry fluid becomes superheated after the expansion (Fig. 3c). The exit temperature of the turbine is higher than the saturation temperature of the condenser. So, a regenerator can be used to recover some heat. But an isentropic fluid which is saturated vapor at the turbine inlet remains saturated throughout the expansion without condensation. Therefore, a regenerator is not needed after the expansion. The features of persistent saturation during the expansion and the fact that a regenerator after the turbine is not needed make the isentropic fluid ideal for ORC’s. Another advantage of the organic fluids is that the enthalpy drop during expansion for isentropic and dry fluids is generally much lower than water-steam mixture. Therefore, only a singlestage expander is usually used for an ORC power system, whereas a multi-stage expander is needed for the expansion of a watervapor cycle. This makes an ORC system more economical in terms of costs and maintenance [16]. Unlike water, most organic fluids suffer chemical decomposition and deterioration at high temperatures and pressures. Therefore, the ORC’s must be operated below the temperatures and pressures at which the fluids are chemically unstable. Most of the organic fluids have relatively low critical pressure and temperature compared to water. Therefore, they are usually operated at low pressures and heat capacity differences compared to watervapor cycle. A suitable organic fluid must have a relatively high
Different power systems are used on ships. Most modern ships use reciprocating diesel engines for propulsion due to their operating simplicity, fuel economy compared to most other power systems, and robustness. Gas turbines are other choices for propulsion on ships, but they are used especially on warships due to their low efficiency compared to the diesel engines. Nuclear power is another alternative on ships, but it is rare because of low prices of the diesel oil. There are usually two different engine systems on ships; a main propulsion system for cruising, and an auxiliary system for generating electric power. The case naval ship has a gas turbine system for main propulsion system. Since most of the ships have diesel engines for cruising, a diesel engine of the ship (auxiliary engine system) has been selected to design the ORC power system. The case naval surface ship has four auxiliary diesel generators to produce electric power, but usually one of them is being used on cruising. The specification of each diesel engine on the naval ship is given in Table 2 [17]. Electric power consumption of the case ship is about 500 kW on cruising whereas it is less when the ship is at the port. The average exhaust temperatures at different generator loads have been measured and are presented in Table 3. The specific fuel consumption of the diesel engine is given in the technical manual and presented in Table 4 [17]. 5. System design The simple ORC system has four main parts, as seen in Fig. 1: Evaporator, turbine, condenser, and pump. All irreversible losses in the components and pressure drops in the piping have been neglected for the design. The turbine and the pump have been considered as isentropic. Rotational speed of generator shaft and diesel engine crankshaft is constant since generators produce electricity at a fixed frequency. Therefore, mass flow rate of combustion air intake can be assumed constant. Exhaust gases volumetric flow rate which
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Fig. 3. T-s diagrams of ‘‘wet”, ‘‘isentropic”, and ‘‘dry” fluids.
is 340 m3/min at 460 °C is given in Table 2. The mass flow rate of a fluid stream is expressed as:
_ ¼ qV_ m
ð1Þ
where q is density, and V_ is volumetric flow rate. Mass flow rate of the exhaust gases is calculated about 160 kg/min. The pressure of the exhaust gases has been assumed atmospheric. The effect of fuel consumption on the exhaust mass flow rate is less than 2% at different generator loads, as seen in Table 4. The mass flow rate of the exhaust gases including the effect of fuel consumption is presented in Fig. 5. Therefore, the effect of fuel consumption on the mass flow of exhaust gases has been neglected and mass flow rate of exhaust gases has been assumed 160 kg/min at the calculations. The condenser temperature of the working fluid depends on the temperature of the heat sink, which is seawater for this case. According to Loydu’s [18] Rules for the Classification of Naval Ships, the selection, layout and arrangement of all shipboard machinery, equipment and appliances should ensure faultless continuous operation under the seawater temperature of 2 °C to +32 °C [19]. Therefore, the condensation temperature of the working fluid has been assumed to be 35 °C for the analysis. After the condenser, the ORC fluid enters the pump as saturated liquid. The pressure is increased isentropically in the pump which means a constant entropy process. For the simple ideal ORC seen in Fig. 1, after the pumping process, the ORC fluid enters the evaporator to absorb waste heat from
exhaust gases and becomes saturated vapor. Superheating is not required for ORC’s, so the fluid leaves the evaporator as saturated vapor in the design. Then it enters the isentropic turbine, produces power and expands to condenser pressure. Afterwards, it enters the condenser and becomes saturated liquid again. When a dry fluid is used as the ORC fluid, a regenerator is needed to increase the power and the efficiency of the system. For the ideal ORC with regenerator seen in Fig. 6, after the pumping process, the ORC fluid enters regenerator and absorbs heat from the stream leaving the turbine. Then, it enters the evaporator and becomes saturated vapor. After that, it enters the turbine, produces power, and enters the regenerator. An ideal regenerator has been assumed in the design process where the temperatures of the points 5 and 2 are the same. The exhaust temperature of the diesel engine at different loads is presented in Table 3. The temperature of the exhaust flowing through the evaporator decreases as the waste heat is absorbed by the ORC fluid. But, if the outlet temperature of the exhaust gases decreases too much, condensate of highly corrosive compounds may form, especially sulfuric acid. Fuel sulfur level effects the corrosion in the evaporator. Most of the sulfur in the fuel is converted into gaseous SO2 that reacts with oxygen in the exhaust to form SO3. The SO3 reacts with the water vapor to form sulfuric acid (H2SO4). The sulfuric acid in the exhaust gases condensates at near 150 °C [20]. Therefore, the exhaust gases exit temperature should not be lower than this value. Minimum exhaust gases temperature after the turbocharger is 158 °C during the idle operation of the
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627
Fig. 4. T-s diagrams of selected ORC fluids.
diesel engine, as seen in Table 3. From routine observations of the exhaust manifold and the funnel, it has been observed that idle engine exhaust temperature, 158 °C is safe for the ducts not to
form acidic corrosion. Therefore, the exit temperature of the exhaust gases from the evaporator has been selected as 158 °C as a design criterion. Air properties have been used for exhaust during
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Table 1 Thermodynamic properties of selected working fluids. Working fluids
R123
R141b
Isopentane R-601a
n-Pentane R-601
n-Hexane
Benzene
Toluene
Type Critical temp. (°C) Critical pressure (kPa) Boiling Temp@1 atm (°C) ASHRAE 34 safety group Maximum operative temperature (°C)
Isentropic 183.7 3668 27.79 B1 150
Isentropic 204.2 4249 32.07 – 250
Dry 187.2 3370 27.86 A3 300
Dry 196.5 3364 35.87 – 300
Dry 234.7 3058 69.28 – 300–400
Dry 288.9 4894 80.07 – 350–400
Dry 318.6 4126 110.4 – 350–400
6. Thermodynamic analysis
Table 2 The specifications of the case diesel generator [17]. Diesel generator set
Value
Engine type No. of cylinders Bore & stroke –(cm) Compression ratio Combustion air flow, (m3/min) Exhaust gas flow, (m3/min) Exhaust outlet temperature (°C) Frequency @ rpm Max. kW Rating
Two cycle 16 14.6 14.6 16:1 148.7 340 460 60 Hz. @ 1800 1000
Table 3 Average exhaust temperature at different generator loads. Generator load (kW)
Exhaust gas temperature after turbocharger (°C)
Idle 200 400 600
158 212 259 328
X dEcv V2 _ cv þ _ i hi þ i þ gzi m ¼ Q_ cv W dt 2 i ! 2 X V _ e he þ e þ gze m 2 e
!
ð2Þ
_ cv are the heat and work transfer at unit time with where Q_ cv and W the environment, respectively. Neglecting kinetic end potential energy changes between the inlet and the exit of the control volume, the waste heat that is transferred to the working fluid in the evaporator from exhaust gases can be written as:
qH ¼ h3 h2
ð3Þ
The isentropic work produced by the turbine per unit mass of the fluid, neglecting kinetic and potential energies can be written as:
Table 4 Specific fuel consumption of diesel engine at different loads. Generator load (kW)
Specific fuel consumption, lb/kW/h (kg/kW/h)
4/4 3/4 2/4 1/4
0.587 0.608 0.632 0.844
(1000) (750) (500) (250)
All the components of the designed ORC power systems are taken to be steady flow open system. The 1st law of Thermodynamics for a steady flow open system is: [21]
(0.266) (0.277) (0.287) (0.383)
wT ¼ h3 h4
ð4Þ
The heat transferring to the sea water in the condenser is equal to:
qL ¼ h1 h4
ð5Þ
The isentropic pump work for unit mass of working fluid can be expressed as:
wp ¼ tðP2 P1 Þ ¼ h2 h1
ð6Þ
where t is specific volume and P is pressure. The net work produced by the system per unit mass of the working fluid is:
wnet ¼ wT wP ¼ qH qL
ð7Þ
The thermal efficiency of a heat engine is expressed as [22]:
gth ¼
wnet q ¼1 L qH qH
ð8Þ
The thermal power absorbed from exhaust gases by the ORC fluid can be written as:
_ a1 ha2 Þ Q_ exhaust ¼ mðh
ð9Þ
The thermal power absorbed from exhaust and transferred into the fluid in the evaporator are equal to each other for an ideal evaporator. Therefore, the net power for an ORC can be written as:
_ net ¼ g Q_ exhaust W th
ð10Þ
Thermal efficiency of a reversible engine can be written as: Fig. 5. Mass flow rate of exhaust gases at different generator loads.
the calculations, because the ratio of nitrogen inside the air is about 80% and the form of nitrogen does not change in combustion. The summary of the design criteria is presented in Table 5.
grev ¼ 1
TL TH
ð11Þ
where TH and TL are the temperatures of hot and low temperature heat sources, respectively.
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Fig. 6. Organic Rankine cycle with regenerator.
Table 5 The design criteria. Type of power cycle
Organic Rankine cycle
Energy source Diesel engine fuel type Working fluids
Diesel engine exhaust gases NATO F-76 Diesel R123, R141b, isopentane, n-pentane, n-hexane, benzene, toluene Isentropic pump and turbine, no irreversible losses in the components Saturated vapor Saturated liquid 35 °C
Components of the cycle Evaporator exit Condenser exit Condenser saturation temperature Exhaust exit temperature from evaporator Regenerator effectiveness
158 °C
e¼
The difference between the reversible work Wrev and the useful work Wnet is due to the irreversibilities present during the process which is called as irreversibility I. It is equivalent to the exergy destroyed and is expressed as:
ð12Þ
where Sgen is the entropy generated during the process. Exergy is defined as the useful work potential of a given amount of energy at some specified state. Neglecting the kinetic and potential energy of the fluid, the flow exergy of a flowing fluid is expressed as:
w ¼ ðh ho Þ T o ðs so Þ
ð14Þ
The overall exergy efficiency is a measure of the performance of a device which is relative to the performance and is given by
_ W
gex ¼ _ net W rev
ð15Þ
Exergy input to the ORC is the exergy transferred by the exhaust gases. So that the overall exergy efficiency of the cycle can be written as [23]:
_ W
net gex ¼ _ m½ðha1 ha2 Þ T o ðsa1 sa2 Þ
qreg h4 h5 ¼ qreg;max h4 hP5; T 2
ð17Þ
ð16Þ
ð18Þ
In this equation, h5 is the enthalpy of the ORC fluid at the inlet of the condenser. If the regenerator is ideal, the temperature of 5 is equal to that of the point 2 where the effectiveness is 1, and the maximum heat transfer between the two streams becomes possible. The regenerator effectiveness has been assumed 1 for the case system, and the temperature of the point 5 has been assumed equal to the temperature of the point 2.The pinch point (PP) temperature of exhaust gases shown in Fig. 2 can be calculated as follows:
_ ORC ðhf ;T H h6 ÞORC _ exhaust ðhPP ha2 Þair ¼ m m
ð13Þ
where ho and so are the enthalpy and entropy at atmospheric pressure and temperature, respectively. Therefore, the reversible work per unit mass, wrev which is the exergy change of the exhaust gases in the evaporator becomes:
wrev ¼ wa1 wa2 ¼ ðha1 ha2 Þ T o ðsa1 sa2 Þ
h6 h2 ¼ h4 h5 Regenerator effectiveness is defined as [21]:
1
I ¼ X destroyed ¼ T 0 Sgen ¼ W rev W net
‘‘Dry” and ‘‘isentropic” fluids are usually preferred in organic Rankine cycles. In this case, superheating before the turbine inlet is not needed in an ORC. But, if the working fluid is a ‘‘dry” fluid, it may be needed an extra heat recovery system, a regenerator, between the turbine and the condenser to increase the thermal efficiency. A simple ORC with a regenerator and T-s diagram of the cycle is seen in Fig. 6. The heat transfer inside the cycle between 2 and 6 before entering the evaporator is equal to the enthalpy change between 4 and 5. Therefore, for an ideal regenerator assuming no heat loss to the environment:
hPP ¼
_ ORC ðhf ;T H h6 ÞORC m þ ha2 _ exhaust m
T PP ¼ T air;hpp
ð19Þ ð20Þ ð21Þ
where hf,TH is the enthalpy of the saturated ORC liquid at the temperature TH, and Tair,hpp is the air temperature at the enthalpy of air, hpp. In the heat transfer analysis of heat exchangers, it is convenient to express the heat transfer rate between the hot and cold streams as: [21]
Q ¼ UAs DT lm
DT lm ¼
DT 1 DT 2 ln ðDT 1 =DT 2 Þ
ð22Þ ð23Þ
where As is the heat transfer area, U is the overall heat transfer coefficient, DTlm is log mean temperature difference between hot and
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cold fluid, and DT1 and DT2 are the temperature differences of the fluids at the inlet and the outlet, respectively. For an unfinned double pipe heat exchanger, it can be expressed as:
ð24Þ
where Ai and Ao are the internal and external surface areas of the pipe, Rf,i and Rf,o are fouling resistances of the surfaces, Di and Do are the internal and external diameters, and hi and ho are the internal and external convection heat transfer coefficients, respectively. For an unfinned tubular heat exchanger, this equation reduces to [24]:
Uo ¼ r
1
0.25
0.2
0.15
0.1 Turbine inlet is saturated vapor
ð25Þ
0.05
If the heat exchanger is unfinned and tubular, and the thickness of the tubes are very thin, then ro and ri are about the same, and the equation reduces to:
0
0 1 r i hi
Uo ¼
1 hi
þ rr0i Rf ;i þ r0 lnðrk 0 =ri Þ þ Rf ;o þ h1o
1 þ Rf ;i þ Rf ;o þ h1o
R123 isopentane n-pentane R141b n-hexane benzene toluene
0.3
Efficiency
Rf ;i lnðD0 =Di Þ Rf ;o 1 1 1 1 1 ¼ ¼ ¼ þ þ þ þ UAs U i Ai U o Ao hi Ai Ai 2pkL Ao ho Ao
0.35
0
100
150
200
250
Turbine Inlet Temperature, C
ð26Þ
7. Results
R123 isopentane n-pentane R141b n-hexane benzene toluene
0.25
Efficiency
The efficiency of a reversible heat engine (Eq. (11)) can be increased in two ways: by decreasing TL, the condensation temperature in condenser, or by increasing TH, the boiling temperature in evaporator. The same is also valid for a Rankine cycle. But, the condensation temperature in the condenser cannot be lower than the temperature of cooling fluid which is sea water for this case. That means sea water temperature limits the efficiency of the system. The condensation temperature for the ORC fluid in the condenser is taken to be 35 °C as a design criteria regarding to the temperature of the sea. But, high temperature of the ORC fluid in the evaporator can be changed depending on the evaporator pressure and evaporator exit temperature. In the beginning, the pump exit pressure which changes the boiling temperature in the evaporator has been changed and the efficiency of the simple ORC without regenerator has been calculated using Eq. (8) and plotted in Fig. 7. It is apparent from the figure that the efficiency of the cycle gets higher as the temperature at the turbine inlet increases. But critical temperature limits the increase of turbine inlet temperature. Therefore, the critical temperature of the ORC fluid should be suitable for the temperature of waste heat source to reach high efficiencies. The efficiencies are very close to each other for all the fluids at low turbine inlet temperatures as shown in the figure, because the outlet of the turbine for all the fluids are close to saturated vapor where a regenerator is not needed. But as the turbine inlet temperature increases, the efficiencies differ from each other as seen in the figure. The figure shows that the efficiency of the ORC with the working fluid benzene is largest compared to the other fluids. Benzene is a fluid between ‘‘dry” and ‘‘isentropic”, like the other fluids, R123 and R141b that have relatively high efficiencies without regenerator. The exit from the turbine for these fluids is close to saturated vapor, and regeneration does not affect the efficiency so much. But if the ORC fluid is a dry fluid, as seen in Figs. 3 and 4, the energy of the fluid, which has much higher temperature than the condenser, is wasted in the condenser. Therefore, regeneration is required for the ORC with dry fluids. The efficiencies of the ORC with the same turbine inlet temperature are relatively low for isopentane, n-pentane, and n-hexane since these fluids are dry fluids as seen in the figure. Regeneration
50
0.2
0.15 100
150
200
250
Turbine Inlet Temperature, C Fig. 7. The efficiency of the simple ORC without regeneration.
is required for these fluids to use the excess energy of the fluid at the exit of the turbine. Following these calculations, an ideal regenerator whose effectiveness 1 has been added to the case design and the efficiency of the ORC has been calculated again to observe the effect of the regenerator. The results of the calculations are presented in Fig. 8. The efficiencies of the ORC’s with and without regenerator are seen in the figure. It is observed from the figure that the efficiencies of dry fluids increase dramatically as a regenerator is used. At relatively low evaporator temperatures, the efficiency of n-hexane with regenerator is highest. But as the evaporator temperature increases, since the critical temperature of toluene is highest among all, toluene with regenerator reaches maximum efficiency. So both the regenerator and the turbine inlet temperature affect the efficiency. Consequently, if the temperature of the heat source is high, the fluids with high critical point temperatures should be preferred for higher thermal efficiencies, and regenerator should be used for dry fluids. The effects of regeneration effectiveness, pump and turbine efficiencies on the cycle efficiency are presented in Fig. 9, using Toluene as the working fluid. It is seen from the Fig. 9 that the pump efficiency does not affect the cycle efficiency if it is more than about 0.25. The cycle efficiency increases as the regenerator effectiveness and turbine efficiency get larger. The pinch point temperature, as seen in Fig. 2, has been calculated for the toluene, benzene and n-hexane cycles. The exhaust
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I. Girgin, C. Ezgi / Energy Conversion and Management 148 (2017) 623–634 0.35 isopentane isopentane with regenerator n-pentane n-pentane with regenerator n-hexane n-hexane with regenerator benzene benzene with regenerator toluene toluene with regenerator
0.3
Efficiency
0.25
0.2
0.15
0.1 Turbine inlet is saturated vapor 0.05
0
0
50
100
150
200
250
Turbine Inlet Temperature, C
0.35
Efficiency
0.3
isopentane isopentane with regenerator n-pentane n-pentane with regenerator n-hexane n-hexane with regenerator benzene benzene with regenerator toluene toluene with regenerator
0.25
0.2
0.15 150
160
170
180
190
200
210
220
230
240
250
Turbine Inlet Temperature, C Fig. 8. The efficiency of the ORC with and without regeneration.
are higher than the other case fluids. The pinch point temperatures for these fluids are presented in Table 6. The other ORC cycles do not have pinch point problems since their evaporator temperatures are much lower. Pinch point temperature difference with toluene (with regenerator) in the evaporator is about 10 °C. This is the least temperature difference between two fluids. The average electric power consumption of the case ship is about 500 kW while the average exhaust temperature for this power production has been measured approximately as 294 °C. The thermal power absorbed by ORC working fluid from exhaust gases, which has a mass flow rate 160 kg/min, has been calculated as 372 kW for cruising conditions of the case naval ship. The power that the ORC system can produce with the working fluids is seen in Table 7 and Fig. 10. The temperature and pressure information on the table is the conditions of the ORC working fluid at the turbine inlet. Regeneration has not been calculated for R123 and R141b since both they are isentropic fluids. Maximum power that can be recovered from waste heat of exhaust gases and maximum cycle efficiency was calculated as 118 kW and 0.32, respectively with the fluid toluene using an ideal regenerator. Overall exergy efficiency and irreversibility are presented for the case ORC cycles in Table 7, Fig. 11, and Fig. 12. It is seen on the figures that the cycle with toluene and regenerator has the maximum power, maximum exergy efficiency and minimum irreversibility among the selected case fluids. The navies of NATO countries use naval distillate fuel (NATO symbol F-76) on power systems. Some values of the logistic fuel NATO F-76 are presented in Table 8 [19,25].The efficiency of the case diesel generator can be calculated as:
_ W
g ¼ _ generator Q combustion
ð27Þ
_ generator is 500 kW for the cruising conditions, and Q_ combustion where W is:
_ u Q_ combustion ¼ mH
ð28Þ
_ is the mass flow rate of diesel fuel given in Table 4. Therwhere m mal efficiency of the diesel generator without ORC is calculated (gDG ) as 0.294. The overall efficiency of combined diesel generator-ORC system can be defined as:
g¼
_ ORC _ generator þ W W _ Q combustion
ð29Þ
In case of using an ideal ORC with regenerator and toluene working fluid, the total power generated is 618 kW and the overall efficiency of the combined system becomes 0.364. The efficiency increases from 0.294 to 0.364 using the ideal ORC system. However, an ideal ORC with isentropic turbine is not realistic. The isentropic efficiencies of various types of prototype expansion machines have been reported in [4] where presented efficiencies ranges from 10% to 85% depending on the expander types, working fluids, and the working conditions.
Table 6 Pinch point (PP) temperatures.
Fig. 9. The effects of regenerator effectiveness, turbine and pump efficiencies on the cycle efficiency.
temperature inside the evaporator should not be less than that of the ORC fluid near it. The pinch point temperature is critical especially for toluene and benzene since their evaporator temperatures
Working fluids
Evaporation temperature (°C)
PP temperature difference (°C)
Toluene (no regenerator) Toluene (with regenerator) Benzene (no regenerator) Benzene (with regenerator) n-Hexane (no regenerator) n-Hexane (with regenerator)
225 225 225 225 200 200
15 10 17 14 55 45
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Table 7 The efficiency, power, overall exergy efficiency, and irreversibility of the ORC on cruising. R123a T3 = 160 °C P3 = 2496 kPa
Efficiency Power, kW Overall exergy efficiency % Irreversibility, kW a
Isopentane T3 = 160 °C P3 = 2202 kPa
n-Pentane T3 = 160 °C P3 = 1892 kPa
n-Hexane T3 = 200 °C P3 = 1787 kPa
Benzene T3 = 225 °C P3 = 2108 kPa
Toluene T3 = 225 °C P3 = 1137 kPa
No regen.
With regen.
No regen.
With regen.
No regen.
With regen.
No regen.
With regen.
No regen.
With regen.
53.2
0.229 85.2 57.8
0.201 74.8 50.7
0.235 87.4 59.3
0.205 76.3 51.7
0.237 88.2 59.8
0.231 85.9 58.2
0.29 107.9 73.2
0.287 106.8 72.4
0.304 113.1 76.7
0.284 105.7 71.7
0.316 117.6 79.7
69.1
62.2
72.7
60
71.3
59.3
61.7
39.5
40.7
34.4
41.8
29.9
0.211
R141b T3 = 175 °C P3 = 2695 kPa
The temperature and pressure information on the table is the conditions of the ORC working fluid at the turbine inlet.
Fig. 10. The power that can be generated from ORC on cruising.
Fig. 11. Exergy efficiency of the cycles.
The isentropic efficiency of a turbine is defined as:
w gt ¼ t ws
ð30Þ
where wt is actual turbine work per for unit mass, and ws is isentropic turbine work. Bala et al. studied the overall efficiency of sliding-vane refrigerant pumps; the highest reported efficiency was about 20% [26]. Melotte performed an experimental study on a centrifugal pump and obtained an efficiency varying between 10% and 20% [27]. The isentropic efficiency of a pump is defined as:
gp ¼
ws wp
ð31Þ
where wp is actual pump work per for unit mass, and ws is isentropic pump work. To be realistic, assuming an isentropic efficiency for the turbine of the case ORC as 0.75, an isentropic efficiency for the pump as 0.20, and neglecting the losses at the ORC electric generator, the electric power output of the cycle with toluene and regenerator becomes 92 kW. In this case the power of the combined system becomes 592 kW while the efficiency of the combined system is equal to 0.349. If the selected ORC cycle is used to generate power from waste heat, 25,500 L of diesel fuel (US$24,870) can be saved and 67.2 tons of CO2 emissions reduced at the end of 1000 operating hours a year of a naval surface ship [19].
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633
Fig. 12. Irreversibility of the cycles.
hexhaust 103 W=m2 K
Table 8 Values of the logistic fuel NATO F-76.
htoluene 3000 W=m2 K
Molecular formula (Average)
C14.8H26.9
Molecular weight Sulfur content, wt.% (max) Density, at 15 °C, kg m3 (max) Fuel price, US$ gallon1, (2016) Lower heating value, Hu, kJ kg1
205 0.1 876 2.97 42,700
It is reported that the cost of some working ORC systems of different manufacturers ranges from $1800/kW to $2857/kW [28]. Assuming an average cost, $2300/kW for the selected ORC system, the overall cost of the system will be approximately $210,000. The system saves about $24,870 at the end of 1000 operating hours a year. It means that, the system will be profitable after 8 years. But if the operating hours of the ship will be more than 1000 h in a year, the system will be profitable before 8 years. Log mean temperature difference at the evaporator is calculated 64 °C for toluene with regenerator. The typical average convection heat transfer coefficients and fouling resistances can be taken from Ref. [24] as:
Rf ;exhaust 0:000308 m2 K=W Rf ;benzene 0:000352 m2 K=W Uo is calculated as 93.44 W/m2 K with the Eq. (26). The heat transfer in the evaporator is taken to be 372 kW. The required heat transfer area at the evaporator for toluene can be calculated as Arequired 60 m2. But it has been assumed that there is no fin in the evaporator, and the size of the evaporator can be reduced using fins inside the heat exchanger. Fig. 8 shows that the efficiency of n-hexane with regenerator is higher than other case ORC fluids when the turbine inlet temperature is less than 190 °C, but for the higher temperature values, the efficiency of toluene with regenerator is highest compared to the other working fluids. For the changing exhaust temperatures of the ship, different working fluids can be used to reach a higher efficiency. But it is not possible only a single ORC system. For such a case, that means, when different fluids give better efficiencies at different temperatures, a combined ORC system can be used. The proposed system is seen in Fig. 13. In this system, two different turbines are connected to the same generator via two hydraulic couplings. When the temperature is
Fig. 13. Combined organic Rankine cycle.
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below 190 °C, ORC system 1 runs with n-hexane. As the exhaust temperature gets higher than 190 °C, the ORC system 1 stops and the second system starts working. So higher efficiencies can be reached with this system compared to a single ORC system. Addition to the turbines, there are two pumps, a three-fluid heat exchanger for condenser, and a three-fluid heat exchanger for the evaporator.
Acknowledgments The views and conclusions contained herein are those of the author and should not be interpreted as necessarily representing official policies or endorsements, either expressed or implied, of any affiliated organization or government. References
8. Conclusions A waste heat powered ORC power system for a naval surface ship was designed and thermodynamically analyzed. Seven different working fluids at different turbine inlet temperatures were used as the working fluids and the efficiencies were calculated. A regenerator after the turbine was used for the dry fluids to improve the cycle efficiency. The study estimated the efficiency between 0.201 and 0.316 depending on the type of the fluid, the inlet temperature of the turbine, and regenerator. The efficiency, power, overall exergy efficiency, and irreversibility of the ORC with benzene on cruising highly good. Yet, benzene is a human carcinogen and most non-industrial applications have been limited. Therefore, benzene should be used carefully on ships. The electrical load of the case naval ship is about 500 kW with 294 °C average exhaust temperature at mass flow rate of 160 kg/ min. The exit temperature of exhaust from the regenerator was taken 158 °C for the calculations not to cause sulfuric acid formation in exhaust manifolds and funnel. The powers of the ideal ORC’s were calculated and compared. The results show that a net power can be obtained from the cycle between 76.3 and 117.6 kW depending on the selected fluids and the operating conditions. The exergy efficiencies and irreversibilities were calculated. Among the selected working fluids, the ORC cycle working with toluene with regenerator was found to have the maximum generated power and exergy efficiency. Assuming an isentropic efficiency for the turbine and the pump of the case ORC equal to 0.75 and 0.20, respectively, and neglecting the losses at the ORC electric generator, the electric power output of the cycle with toluene and regenerator becomes 92 kW. In this case the power of the combined system is calculated as 592 kW while the efficiency of the combined system is equal to about 0.349. If the selected ORC cycle is used to generate power from waste heat, 25,500 L of diesel fuel (US$24,870) can be saved and 67.2 tons of CO2 emissions reduced at the end of 1000 operating hours a year of a naval surface ship. The system will be profitable after 8 years assuming 1000 h cruising time for a year. The excess of production is stored in Li-ion batteries because of high energy densities, good cycle life and high charge/discharge efficiency The results show that the ORC generates power from waste heat of the exhaust energy emitted to atmosphere and thus reduce the fuel that is used on ship and global CO2 emissions. It has been proposed a combined ORC system to work at different exhaust temperatures to reach higher efficiencies. Author contributions _ Ibrahim Girgin conceived and designed the research, analyzed _ the data. Ibrahim Girgin and Cüneyt Ezgi worked out the theory, and wrote the manuscript, read and approved the final manuscript. Conflicts of interest The author declares no conflict of interest.
[1] Atmaca M. Efficiency analysis of combined cogeneration systems with steam and gas turbines. Energy Sources 2011;33:360–9. [2] Tchanche BT, Lambrinos G, Franoudakis A, Papadakis G. Low-grade heat conversion into power using organic Rankine cycles – a review of various applications. Renew Sustain Energy Rev 2011;15:3963–79. [3] Atmaca M, Gumus M, Demir A. Comparative thermodynamic analysis of dual cycle under alternative conditions. Therm Sci 2011;15:953–60. [4] Bao J, Zhao L. A review of working fluid and expander selections for organic Rankine cycle. Renew Sustain Energy Rev 2013;24:325–42. [5] Guo C, Du X, Yang L, Yang Y. Organic Rankine cycle for power recovery of exhaust flue gas. Appl Therm Eng 2015;75:135–44. [6] Liu H, Shao Y, Li J. A biomass-fired micro-scale CHP system with organic Rankine cycle (ORC) – thermodynamic modelling studies. Biomass Bioenerg 2011;35:3985–94. [7] Eyidogan M, Kilic FC, Kaya D, Coban V, Cagman S. Investigation of Organic Rankine Cycle (ORC) technologies in Turkey from the technical and economic point of view. Renew Sustain Energy Rev 2016;58:885–95. [8] Desai NB, Bandyopadhyay S. Thermo-economic analysis and selection of working fluid for solar organic Rankine cycle. Appl Therm Eng 2016;95:471–81. [9] Hung TC, Shai TY, Wang SK. A review of Organic Rankine Cycles (ORCs) for the recovery of low-grade waste heat. Energy 1997;22(7):661–7. [10] Shu G, Liang Y, Wei H, Tian H, Zhao J, Liu L. A review of waste heat recovery on two-stroke IC engine aboard ships. Renew Sustain Energy Rev 2013;19:385–401. [11] International Maritime Organization (IMO). Report of the marine environment protection committee on its sixty-seventh session. MEPC 67/6; IMO: London, UK; 2014. [12] Ezgi C. Design and thermodynamic analysis of an H2O-LiBr AHP system for naval surface ship application. Int J Refrig 2014;48:153–65. [13] Larsen U, Sigthorsson O, Hagling F. A comparison of advanced heat recovery power cycles in a combined cycle for large ships. Energy 2014;74:260–8. [14] Carcasci C, Ferraro R, Miliotti E. Thermodynamic analysis of an organic Rankine cycle for waste heat recovery from gas turbines. Energy 2014;65:91–100. [15] Khaljani M, Saray RK, Bahlouli K. Thermodynamic and thermoeconomic optimization of an integrated gas turbine and organic Rankine cycle. Energy 2015;93:2136–45. [16] Chen H, Goswami DY, Stefekanos EK. A review of thermodynamic cycles and working fluids for the conversion of low-grade heat. Renew Sustain Energy Rev 2010;14:3059–67. [17] Detroit Diesel Engines, standby electric set models, 16V–149T 1070 kW. Detroit Diesel Allison Division of General Motors Corporation. [18] Loydu T. Rules for the classification of naval ships, ship operation installations and auxiliary systems. Türk Loydu: Istanbul, Turkey; 2015. [19] Ezgi C, Girgin I. Design and thermodynamic analysis of a steam ejector refrigeration/heat pump system for naval surface ship applications. Entropy 2015;17(12):8152–73. [20] Xin Q. Diesel engine system design. Cambridge, UK: Woodhead Publishing; 2013. [21] Cengel YA, Boles MA. Thermodynamics: an engineering approach. Boston, MA, USA: McGraw-Hill; 2001. [22] Atmaca M, Gumus M, Inan AT, et al. Int J Thermophys 2009;30:1724. http://dx. doi.org/10.1007/s10765-009-0621-3. [23] Kaska O. Energy and exergy analysis of an organic Rankine for power generation from waste heat recovery in steel industry. Energy Convers Manage 2014;11:108–17. [24] Kakaç S, Liu H. Heat exchangers, selection, rating, and thermal design. Washington, D.C., USA: CRC Press; 1997. [25] Steinfeld G, Sanderson R, Ghezel-Ayagh H, Abens S, Cervi MC. Distillate fuel processing for marine fuel cell applications. In: Proceedings of the AICHE Spring Meeting, Atlanta, GA, USA; 5–9 March 2000. [26] Bala E, O0 Callaghan P, Probert S. Influence of organic working fluids on the performance of a positive-displacement pump with sliding vanes. Appl Energy 1985;20:153–9. [27] Melotte N. Development and optimization of Organic Rankine Cycle control strategies. Master thesis. University of Liège; 2012. [28] Arvay P, Muller M, Ramdeen V. Economic implementation of the organic Rankine cycle in industry. ACEEE Summer Study on Energy Efficiency in Industry; 2011.
Contents lists available at ScienceDirect
Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Design and thermodynamic and thermoeconomic analysis of an organic Rankine cycle for naval surface ship applications Ibrahim Girgin ⇑, Cüneyt Ezgi Mechanical Engineering Department, Beykent University, Istanbul 34396, Turkey
a r t i c l e
i n f o
Article history: Received 28 February 2017 Received in revised form 23 May 2017 Accepted 11 June 2017
Keywords: Organic Rankine cycle Waste heat recovery Ship Cogeneration
a b s t r a c t This paper presents the thermodynamic modeling of an organic Rankine cycle (ORC) that uses waste exhaust energy of a 1000 kW diesel generator on a naval ship. Seven different working fluids have been selected as the ORC fluids. The commercial software (EES) has been used to predict the thermodynamic properties of the selected fluids. The efficiency of the ORC goes up to 32% with toluene. For the needed generator power of 500 kW on cruising, the ideal ORC can produce 118 kW power with the working fluid toluene. Assuming an isentropic efficiency for the turbine and the pump of the case ORC to be 0.75 and 0.20, respectively, and neglecting the losses at the ORC electric generator, the electric power output of the ORC cycle becomes 92 kW. The power of the diesel-ORC system becomes 592 kW while the combined efficiency is calculated as 0.349. The ORC saves 25,500 L of diesel fuel (US$24,870) and reduces 67.2 tons of CO2 emissions at the end of 1000 operating hours. ORC working fluids may result different efficiencies at different temperatures. Therefore, a combined ORC system is proposed to get higher efficiencies at different thermal loads. The exergy efficiencies and irreversibilities were calculated. Ó 2017 Elsevier Ltd. All rights reserved.
1. Introduction Energy is ability of a system to do work. High cost of energy and the environmental effects of energy consumption cause enormous stresses on the world. Fossil fuel has commonly been used to produce energy. But, the consumption of fossil fuels has caused many environmental problems including air pollution, acid rain, and global warming. Rankine Cycle using fossil fuel is still a dominant power supply method. Low-grade waste heat accounts for 50% or more of the total heat generated in industry, and it has generally been discarded due to lack of efficient recovery methods. Therefore, recovery of lowtemperature waste heat and renewable energy is attracting much research attentions [1]. Rankine cycle using water as working fluid does not allow efficient recovery of waste heat below 370 °C. Many problems are encountered in steam Rankine cycle when water is used as the working fluid: Superheating is needed to prevent condensation during expansion, risk of erosion of turbine blades, high pressure in the evaporator, and complex and expensive turbines [2,3]. Various thermodynamic cycles can be used for conversion of low-grade heat to work, such as supercritical Rankine cycle,
⇑ Corresponding author. E-mail addresses: [email protected] (I. Girgin), cuneytezgi@beykent. edu.tr (C. Ezgi). http://dx.doi.org/10.1016/j.enconman.2017.06.033 0196-8904/Ó 2017 Elsevier Ltd. All rights reserved.
organic Rankine cycle, and Kalina cycle. But organic Rankine cycle has the characteristic of simple structure, easy maintenance and high reliability [4]. Therefore, it has many uses including industrial waste heat [5], biomass energy [6], geothermal energy [7], and solar energy [8]. Guo et al. studied thermal efficiency, influence of recuperator and exergy destruction for a 240 MW pulverized coal-fired power plant. The analytical results show that the mixture that matches with heat sink has the greatest efficiency and the mixture that matches with heat source has the lowest superheat degree. There exists no optimal working fluid for all indicators (thermal efficiency, heat exchanger area, mass flow and volumetric flow etc.) [5]. Liu et al. presented the results of thermodynamic modeling studies of a 2 kW biomass-fired system with organic Rankine cycle. They used three different environmentally friendly refrigerants to predict the efficiency of the system [6]. Desai and Bandyopadhyay reported thermo-economic comparisons of organic Rankine and steam Rankine cycles powered by linefocusing concentrating solar collectors. They also proposed a simple selection methodology, based on thermo-economic analysis, and a comparison diagram for working fluids of power generating cycles [8]. Hung et al. investigated the effects of turbine inlet temperature and condenser outlet temperature at different pressures on the efficiency of the Rankine Cycle for different working fluids [9]. Bao and Zhao have reviewed the influence of working fluid
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Nomenclature A h I _ m
t
P R s S T U V_ w
heat transfer area (m2) enthalpy (kJ kg1) irreversibility mass flow rate (kg s1) specific volume (m3 kg1) pressure (Pa) resistance (m2 KW1) specific entropy (J kg1 K1) entropy (J K1) temperature (K) overall heat transfer coefficient (Wm2 K1) volumetric flow rate (m3 s1) specific work (J kg1)
g
efficiency (%)
Subscripts ex exergy f fouling i internal lm log mean o external ORC organic Rankine cycle P pump PP pinch point rev reversible s isentropic t turbine
Greek letters q density (kg m3) 2 effectiveness of heat exchanger
properties on organic Rankine cycle, the screening of working fluid, and the comparison of different types of expansion turbines [4]. Different types of waste heat recovery technologies can be used onboard ships, which can be turbo charging of air into the engine, absorption refrigeration, thermoelectric generation, or combined power cycles [10]. The International Maritime Organization has developed a global CO2 reduction index known as the Energy Efficiency Design Index for new ships and the Ship Energy Efficiency Management Plan for all ships. The new chapter added to MARPOL ANNEX VI Regulations for the prevention of air pollution from ships, which was implemented on January 1, 2013, aims to reduce the emission of greenhouse gases, specifically CO2 emissions [11]. Implementing CO2 reduction measures will result a significant reduction in fuel consumption. Therefore, searching for new energy conservation methods that can be applied onboard ships is necessary [12]. The potential of waste heat recovery is among the most important technologies to lower fuel consumption. Larsen et al. compared ORC, the Kalina cycle, and the steam Rankine Cycle in a combined cycle application with a large marine two-stroke diesel engine. It is concluded that the ORC has the greatest potential for increasing the combined thermal efficiency [13]. Carcasci et al. illustrated the results of the simulations of an ORC combined with a gas turbine in order to convert waste heat into electric power. Four different ORC working fluids were compared to identify best choice. Rankine cycle was optimized by varying the main pressure of the fluid at different temperatures. The possible use of a superheater was also investigated in order to increase electrical power [14]. Khaljani et al. optimized a cogeneration system consisting a gas turbine and an ORC with selected decision parameters. It is reported that the gas turbine inlet temperature has important role on the trade-off between exergy efficiency and cost criteria [15]. In this study, seven different ORC fluids have been compared on an ORC design which uses the thermal energy of a diesel engine of a naval ship. A combined ORC cycle using two different ORC fluids has been proposed to reach higher efficiencies. 2. Organic Rankine cycle (ORC) Organic Rankine cycle (ORC) applies the principles of Rankine cycle using an organic fluid that has a low boiling point to recover heat from low temperature heat sources. A simple ORC converting waste heat from exhaust gases into useful work, and a typical T-s diagram is seen in Fig. 1.
There are four main components and ideal processes of this system: An evaporator for recovering waste heat from exhaust gases, a turbine for expansion of the working fluid and producing work, a condenser for transferring heat to the environment, and a pump for increasing the pressure of the working fluid. Evaporator is the component to recover waste heat which may be in various types, such as solar heat, waste heat of flue gas, and geothermal heat. The fluid is heated in the evaporator and a phase change occurs from compressed liquid to saturated vapor. After the evaporator, the working fluid at high pressure expands in the turbine and produces electrical work at the generator connected to the turbine. According to the 2nd law of Thermodynamics, a heat engine should run between at least two heat sources to produce net work. Therefore, some amount of heat should be transferred to outside environment in condenser to change phase of the working fluid to saturated liquid at an ideal Rankine cycle. Then, the liquid is pumped into the evaporator to absorb waste heat. Even though the points 1 and 2 seem to be the same points in T-s diagrams as seen in Fig. 1, actually they are two different points. The point 2 is after the pump where the point 1 is before the pump. A higher evaporator pressure gives better efficiencies in ORC power systems. But one restriction for evaporator pressure is pinch point (PP) temperature which shall be kept above zero. Pinch point is defined as the point at which the temperature difference between the hot and the cold fluid is minimum. This point is a fundamental parameter when designing a practical ORC. Pinch value must always be positive, in order to make the heat exchange possible. A small value of pinch corresponds to a very difficult heat transfer and therefore requires more heat exchange area. The minimum temperature difference occurs in evaporator at the point of saturated liquid as seen in Fig. 2. 3. Working fluids The saturation curve is the most crucial characteristic of an ORC working fluid. This characteristic affects the cycle efficiency, fluid applicability, and the arrangement of the associated equipment in an ORC power-generation system [9]. The slope of the saturation vapor curve of a fluid in T-s diagram can be negative (e.g. water), vertical (e.g. R11), or positive (e.g. n-hexane), and the fluids are called ‘‘wet”, ‘‘isentropic”, and ‘‘dry”, respectively. Typical T-s diagrams of wet, isentropic, and dry fluids are seen in Fig. 3. Wet fluids like water need to be superheated, because as they enter the turbine in saturated vapor phase, the percentage of satu-
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625
Fig. 1. The simple ideal organic Rankine cycle.
boiling point to reach a higher efficiency [9]. Recommended fluids for different applications, working conditions, performance indicators, and heat source temperature levels have been presented in [4]. Based on these features, R-123, R-141b, isopentane, npentane, n-hexane, benzene, and toluene have been selected as working fluids for thermodynamic analysis. Since the shape of the saturation curve is so important for ORC fluids, T-s diagrams and some thermodynamic properties of the selected fluids are presented in Fig. 4 and Table 1, respectively. 4. System selection for naval surface ships
Fig. 2. Pinch points in an ORC.
rated liquid increases at the exit and it may be harmful for turbine blades, as shown in Fig. 3a. But many organic fluids, which are ‘‘isentropic (Fig. 3b) or dry (Fig. 3c), do not need superheating. The saturated vapor phase of a dry fluid becomes superheated after the expansion (Fig. 3c). The exit temperature of the turbine is higher than the saturation temperature of the condenser. So, a regenerator can be used to recover some heat. But an isentropic fluid which is saturated vapor at the turbine inlet remains saturated throughout the expansion without condensation. Therefore, a regenerator is not needed after the expansion. The features of persistent saturation during the expansion and the fact that a regenerator after the turbine is not needed make the isentropic fluid ideal for ORC’s. Another advantage of the organic fluids is that the enthalpy drop during expansion for isentropic and dry fluids is generally much lower than water-steam mixture. Therefore, only a singlestage expander is usually used for an ORC power system, whereas a multi-stage expander is needed for the expansion of a watervapor cycle. This makes an ORC system more economical in terms of costs and maintenance [16]. Unlike water, most organic fluids suffer chemical decomposition and deterioration at high temperatures and pressures. Therefore, the ORC’s must be operated below the temperatures and pressures at which the fluids are chemically unstable. Most of the organic fluids have relatively low critical pressure and temperature compared to water. Therefore, they are usually operated at low pressures and heat capacity differences compared to watervapor cycle. A suitable organic fluid must have a relatively high
Different power systems are used on ships. Most modern ships use reciprocating diesel engines for propulsion due to their operating simplicity, fuel economy compared to most other power systems, and robustness. Gas turbines are other choices for propulsion on ships, but they are used especially on warships due to their low efficiency compared to the diesel engines. Nuclear power is another alternative on ships, but it is rare because of low prices of the diesel oil. There are usually two different engine systems on ships; a main propulsion system for cruising, and an auxiliary system for generating electric power. The case naval ship has a gas turbine system for main propulsion system. Since most of the ships have diesel engines for cruising, a diesel engine of the ship (auxiliary engine system) has been selected to design the ORC power system. The case naval surface ship has four auxiliary diesel generators to produce electric power, but usually one of them is being used on cruising. The specification of each diesel engine on the naval ship is given in Table 2 [17]. Electric power consumption of the case ship is about 500 kW on cruising whereas it is less when the ship is at the port. The average exhaust temperatures at different generator loads have been measured and are presented in Table 3. The specific fuel consumption of the diesel engine is given in the technical manual and presented in Table 4 [17]. 5. System design The simple ORC system has four main parts, as seen in Fig. 1: Evaporator, turbine, condenser, and pump. All irreversible losses in the components and pressure drops in the piping have been neglected for the design. The turbine and the pump have been considered as isentropic. Rotational speed of generator shaft and diesel engine crankshaft is constant since generators produce electricity at a fixed frequency. Therefore, mass flow rate of combustion air intake can be assumed constant. Exhaust gases volumetric flow rate which
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Fig. 3. T-s diagrams of ‘‘wet”, ‘‘isentropic”, and ‘‘dry” fluids.
is 340 m3/min at 460 °C is given in Table 2. The mass flow rate of a fluid stream is expressed as:
_ ¼ qV_ m
ð1Þ
where q is density, and V_ is volumetric flow rate. Mass flow rate of the exhaust gases is calculated about 160 kg/min. The pressure of the exhaust gases has been assumed atmospheric. The effect of fuel consumption on the exhaust mass flow rate is less than 2% at different generator loads, as seen in Table 4. The mass flow rate of the exhaust gases including the effect of fuel consumption is presented in Fig. 5. Therefore, the effect of fuel consumption on the mass flow of exhaust gases has been neglected and mass flow rate of exhaust gases has been assumed 160 kg/min at the calculations. The condenser temperature of the working fluid depends on the temperature of the heat sink, which is seawater for this case. According to Loydu’s [18] Rules for the Classification of Naval Ships, the selection, layout and arrangement of all shipboard machinery, equipment and appliances should ensure faultless continuous operation under the seawater temperature of 2 °C to +32 °C [19]. Therefore, the condensation temperature of the working fluid has been assumed to be 35 °C for the analysis. After the condenser, the ORC fluid enters the pump as saturated liquid. The pressure is increased isentropically in the pump which means a constant entropy process. For the simple ideal ORC seen in Fig. 1, after the pumping process, the ORC fluid enters the evaporator to absorb waste heat from
exhaust gases and becomes saturated vapor. Superheating is not required for ORC’s, so the fluid leaves the evaporator as saturated vapor in the design. Then it enters the isentropic turbine, produces power and expands to condenser pressure. Afterwards, it enters the condenser and becomes saturated liquid again. When a dry fluid is used as the ORC fluid, a regenerator is needed to increase the power and the efficiency of the system. For the ideal ORC with regenerator seen in Fig. 6, after the pumping process, the ORC fluid enters regenerator and absorbs heat from the stream leaving the turbine. Then, it enters the evaporator and becomes saturated vapor. After that, it enters the turbine, produces power, and enters the regenerator. An ideal regenerator has been assumed in the design process where the temperatures of the points 5 and 2 are the same. The exhaust temperature of the diesel engine at different loads is presented in Table 3. The temperature of the exhaust flowing through the evaporator decreases as the waste heat is absorbed by the ORC fluid. But, if the outlet temperature of the exhaust gases decreases too much, condensate of highly corrosive compounds may form, especially sulfuric acid. Fuel sulfur level effects the corrosion in the evaporator. Most of the sulfur in the fuel is converted into gaseous SO2 that reacts with oxygen in the exhaust to form SO3. The SO3 reacts with the water vapor to form sulfuric acid (H2SO4). The sulfuric acid in the exhaust gases condensates at near 150 °C [20]. Therefore, the exhaust gases exit temperature should not be lower than this value. Minimum exhaust gases temperature after the turbocharger is 158 °C during the idle operation of the
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Fig. 4. T-s diagrams of selected ORC fluids.
diesel engine, as seen in Table 3. From routine observations of the exhaust manifold and the funnel, it has been observed that idle engine exhaust temperature, 158 °C is safe for the ducts not to
form acidic corrosion. Therefore, the exit temperature of the exhaust gases from the evaporator has been selected as 158 °C as a design criterion. Air properties have been used for exhaust during
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Table 1 Thermodynamic properties of selected working fluids. Working fluids
R123
R141b
Isopentane R-601a
n-Pentane R-601
n-Hexane
Benzene
Toluene
Type Critical temp. (°C) Critical pressure (kPa) Boiling Temp@1 atm (°C) ASHRAE 34 safety group Maximum operative temperature (°C)
Isentropic 183.7 3668 27.79 B1 150
Isentropic 204.2 4249 32.07 – 250
Dry 187.2 3370 27.86 A3 300
Dry 196.5 3364 35.87 – 300
Dry 234.7 3058 69.28 – 300–400
Dry 288.9 4894 80.07 – 350–400
Dry 318.6 4126 110.4 – 350–400
6. Thermodynamic analysis
Table 2 The specifications of the case diesel generator [17]. Diesel generator set
Value
Engine type No. of cylinders Bore & stroke –(cm) Compression ratio Combustion air flow, (m3/min) Exhaust gas flow, (m3/min) Exhaust outlet temperature (°C) Frequency @ rpm Max. kW Rating
Two cycle 16 14.6 14.6 16:1 148.7 340 460 60 Hz. @ 1800 1000
Table 3 Average exhaust temperature at different generator loads. Generator load (kW)
Exhaust gas temperature after turbocharger (°C)
Idle 200 400 600
158 212 259 328
X dEcv V2 _ cv þ _ i hi þ i þ gzi m ¼ Q_ cv W dt 2 i ! 2 X V _ e he þ e þ gze m 2 e
!
ð2Þ
_ cv are the heat and work transfer at unit time with where Q_ cv and W the environment, respectively. Neglecting kinetic end potential energy changes between the inlet and the exit of the control volume, the waste heat that is transferred to the working fluid in the evaporator from exhaust gases can be written as:
qH ¼ h3 h2
ð3Þ
The isentropic work produced by the turbine per unit mass of the fluid, neglecting kinetic and potential energies can be written as:
Table 4 Specific fuel consumption of diesel engine at different loads. Generator load (kW)
Specific fuel consumption, lb/kW/h (kg/kW/h)
4/4 3/4 2/4 1/4
0.587 0.608 0.632 0.844
(1000) (750) (500) (250)
All the components of the designed ORC power systems are taken to be steady flow open system. The 1st law of Thermodynamics for a steady flow open system is: [21]
(0.266) (0.277) (0.287) (0.383)
wT ¼ h3 h4
ð4Þ
The heat transferring to the sea water in the condenser is equal to:
qL ¼ h1 h4
ð5Þ
The isentropic pump work for unit mass of working fluid can be expressed as:
wp ¼ tðP2 P1 Þ ¼ h2 h1
ð6Þ
where t is specific volume and P is pressure. The net work produced by the system per unit mass of the working fluid is:
wnet ¼ wT wP ¼ qH qL
ð7Þ
The thermal efficiency of a heat engine is expressed as [22]:
gth ¼
wnet q ¼1 L qH qH
ð8Þ
The thermal power absorbed from exhaust gases by the ORC fluid can be written as:
_ a1 ha2 Þ Q_ exhaust ¼ mðh
ð9Þ
The thermal power absorbed from exhaust and transferred into the fluid in the evaporator are equal to each other for an ideal evaporator. Therefore, the net power for an ORC can be written as:
_ net ¼ g Q_ exhaust W th
ð10Þ
Thermal efficiency of a reversible engine can be written as: Fig. 5. Mass flow rate of exhaust gases at different generator loads.
the calculations, because the ratio of nitrogen inside the air is about 80% and the form of nitrogen does not change in combustion. The summary of the design criteria is presented in Table 5.
grev ¼ 1
TL TH
ð11Þ
where TH and TL are the temperatures of hot and low temperature heat sources, respectively.
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Fig. 6. Organic Rankine cycle with regenerator.
Table 5 The design criteria. Type of power cycle
Organic Rankine cycle
Energy source Diesel engine fuel type Working fluids
Diesel engine exhaust gases NATO F-76 Diesel R123, R141b, isopentane, n-pentane, n-hexane, benzene, toluene Isentropic pump and turbine, no irreversible losses in the components Saturated vapor Saturated liquid 35 °C
Components of the cycle Evaporator exit Condenser exit Condenser saturation temperature Exhaust exit temperature from evaporator Regenerator effectiveness
158 °C
e¼
The difference between the reversible work Wrev and the useful work Wnet is due to the irreversibilities present during the process which is called as irreversibility I. It is equivalent to the exergy destroyed and is expressed as:
ð12Þ
where Sgen is the entropy generated during the process. Exergy is defined as the useful work potential of a given amount of energy at some specified state. Neglecting the kinetic and potential energy of the fluid, the flow exergy of a flowing fluid is expressed as:
w ¼ ðh ho Þ T o ðs so Þ
ð14Þ
The overall exergy efficiency is a measure of the performance of a device which is relative to the performance and is given by
_ W
gex ¼ _ net W rev
ð15Þ
Exergy input to the ORC is the exergy transferred by the exhaust gases. So that the overall exergy efficiency of the cycle can be written as [23]:
_ W
net gex ¼ _ m½ðha1 ha2 Þ T o ðsa1 sa2 Þ
qreg h4 h5 ¼ qreg;max h4 hP5; T 2
ð17Þ
ð16Þ
ð18Þ
In this equation, h5 is the enthalpy of the ORC fluid at the inlet of the condenser. If the regenerator is ideal, the temperature of 5 is equal to that of the point 2 where the effectiveness is 1, and the maximum heat transfer between the two streams becomes possible. The regenerator effectiveness has been assumed 1 for the case system, and the temperature of the point 5 has been assumed equal to the temperature of the point 2.The pinch point (PP) temperature of exhaust gases shown in Fig. 2 can be calculated as follows:
_ ORC ðhf ;T H h6 ÞORC _ exhaust ðhPP ha2 Þair ¼ m m
ð13Þ
where ho and so are the enthalpy and entropy at atmospheric pressure and temperature, respectively. Therefore, the reversible work per unit mass, wrev which is the exergy change of the exhaust gases in the evaporator becomes:
wrev ¼ wa1 wa2 ¼ ðha1 ha2 Þ T o ðsa1 sa2 Þ
h6 h2 ¼ h4 h5 Regenerator effectiveness is defined as [21]:
1
I ¼ X destroyed ¼ T 0 Sgen ¼ W rev W net
‘‘Dry” and ‘‘isentropic” fluids are usually preferred in organic Rankine cycles. In this case, superheating before the turbine inlet is not needed in an ORC. But, if the working fluid is a ‘‘dry” fluid, it may be needed an extra heat recovery system, a regenerator, between the turbine and the condenser to increase the thermal efficiency. A simple ORC with a regenerator and T-s diagram of the cycle is seen in Fig. 6. The heat transfer inside the cycle between 2 and 6 before entering the evaporator is equal to the enthalpy change between 4 and 5. Therefore, for an ideal regenerator assuming no heat loss to the environment:
hPP ¼
_ ORC ðhf ;T H h6 ÞORC m þ ha2 _ exhaust m
T PP ¼ T air;hpp
ð19Þ ð20Þ ð21Þ
where hf,TH is the enthalpy of the saturated ORC liquid at the temperature TH, and Tair,hpp is the air temperature at the enthalpy of air, hpp. In the heat transfer analysis of heat exchangers, it is convenient to express the heat transfer rate between the hot and cold streams as: [21]
Q ¼ UAs DT lm
DT lm ¼
DT 1 DT 2 ln ðDT 1 =DT 2 Þ
ð22Þ ð23Þ
where As is the heat transfer area, U is the overall heat transfer coefficient, DTlm is log mean temperature difference between hot and
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cold fluid, and DT1 and DT2 are the temperature differences of the fluids at the inlet and the outlet, respectively. For an unfinned double pipe heat exchanger, it can be expressed as:
ð24Þ
where Ai and Ao are the internal and external surface areas of the pipe, Rf,i and Rf,o are fouling resistances of the surfaces, Di and Do are the internal and external diameters, and hi and ho are the internal and external convection heat transfer coefficients, respectively. For an unfinned tubular heat exchanger, this equation reduces to [24]:
Uo ¼ r
1
0.25
0.2
0.15
0.1 Turbine inlet is saturated vapor
ð25Þ
0.05
If the heat exchanger is unfinned and tubular, and the thickness of the tubes are very thin, then ro and ri are about the same, and the equation reduces to:
0
0 1 r i hi
Uo ¼
1 hi
þ rr0i Rf ;i þ r0 lnðrk 0 =ri Þ þ Rf ;o þ h1o
1 þ Rf ;i þ Rf ;o þ h1o
R123 isopentane n-pentane R141b n-hexane benzene toluene
0.3
Efficiency
Rf ;i lnðD0 =Di Þ Rf ;o 1 1 1 1 1 ¼ ¼ ¼ þ þ þ þ UAs U i Ai U o Ao hi Ai Ai 2pkL Ao ho Ao
0.35
0
100
150
200
250
Turbine Inlet Temperature, C
ð26Þ
7. Results
R123 isopentane n-pentane R141b n-hexane benzene toluene
0.25
Efficiency
The efficiency of a reversible heat engine (Eq. (11)) can be increased in two ways: by decreasing TL, the condensation temperature in condenser, or by increasing TH, the boiling temperature in evaporator. The same is also valid for a Rankine cycle. But, the condensation temperature in the condenser cannot be lower than the temperature of cooling fluid which is sea water for this case. That means sea water temperature limits the efficiency of the system. The condensation temperature for the ORC fluid in the condenser is taken to be 35 °C as a design criteria regarding to the temperature of the sea. But, high temperature of the ORC fluid in the evaporator can be changed depending on the evaporator pressure and evaporator exit temperature. In the beginning, the pump exit pressure which changes the boiling temperature in the evaporator has been changed and the efficiency of the simple ORC without regenerator has been calculated using Eq. (8) and plotted in Fig. 7. It is apparent from the figure that the efficiency of the cycle gets higher as the temperature at the turbine inlet increases. But critical temperature limits the increase of turbine inlet temperature. Therefore, the critical temperature of the ORC fluid should be suitable for the temperature of waste heat source to reach high efficiencies. The efficiencies are very close to each other for all the fluids at low turbine inlet temperatures as shown in the figure, because the outlet of the turbine for all the fluids are close to saturated vapor where a regenerator is not needed. But as the turbine inlet temperature increases, the efficiencies differ from each other as seen in the figure. The figure shows that the efficiency of the ORC with the working fluid benzene is largest compared to the other fluids. Benzene is a fluid between ‘‘dry” and ‘‘isentropic”, like the other fluids, R123 and R141b that have relatively high efficiencies without regenerator. The exit from the turbine for these fluids is close to saturated vapor, and regeneration does not affect the efficiency so much. But if the ORC fluid is a dry fluid, as seen in Figs. 3 and 4, the energy of the fluid, which has much higher temperature than the condenser, is wasted in the condenser. Therefore, regeneration is required for the ORC with dry fluids. The efficiencies of the ORC with the same turbine inlet temperature are relatively low for isopentane, n-pentane, and n-hexane since these fluids are dry fluids as seen in the figure. Regeneration
50
0.2
0.15 100
150
200
250
Turbine Inlet Temperature, C Fig. 7. The efficiency of the simple ORC without regeneration.
is required for these fluids to use the excess energy of the fluid at the exit of the turbine. Following these calculations, an ideal regenerator whose effectiveness 1 has been added to the case design and the efficiency of the ORC has been calculated again to observe the effect of the regenerator. The results of the calculations are presented in Fig. 8. The efficiencies of the ORC’s with and without regenerator are seen in the figure. It is observed from the figure that the efficiencies of dry fluids increase dramatically as a regenerator is used. At relatively low evaporator temperatures, the efficiency of n-hexane with regenerator is highest. But as the evaporator temperature increases, since the critical temperature of toluene is highest among all, toluene with regenerator reaches maximum efficiency. So both the regenerator and the turbine inlet temperature affect the efficiency. Consequently, if the temperature of the heat source is high, the fluids with high critical point temperatures should be preferred for higher thermal efficiencies, and regenerator should be used for dry fluids. The effects of regeneration effectiveness, pump and turbine efficiencies on the cycle efficiency are presented in Fig. 9, using Toluene as the working fluid. It is seen from the Fig. 9 that the pump efficiency does not affect the cycle efficiency if it is more than about 0.25. The cycle efficiency increases as the regenerator effectiveness and turbine efficiency get larger. The pinch point temperature, as seen in Fig. 2, has been calculated for the toluene, benzene and n-hexane cycles. The exhaust
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I. Girgin, C. Ezgi / Energy Conversion and Management 148 (2017) 623–634 0.35 isopentane isopentane with regenerator n-pentane n-pentane with regenerator n-hexane n-hexane with regenerator benzene benzene with regenerator toluene toluene with regenerator
0.3
Efficiency
0.25
0.2
0.15
0.1 Turbine inlet is saturated vapor 0.05
0
0
50
100
150
200
250
Turbine Inlet Temperature, C
0.35
Efficiency
0.3
isopentane isopentane with regenerator n-pentane n-pentane with regenerator n-hexane n-hexane with regenerator benzene benzene with regenerator toluene toluene with regenerator
0.25
0.2
0.15 150
160
170
180
190
200
210
220
230
240
250
Turbine Inlet Temperature, C Fig. 8. The efficiency of the ORC with and without regeneration.
are higher than the other case fluids. The pinch point temperatures for these fluids are presented in Table 6. The other ORC cycles do not have pinch point problems since their evaporator temperatures are much lower. Pinch point temperature difference with toluene (with regenerator) in the evaporator is about 10 °C. This is the least temperature difference between two fluids. The average electric power consumption of the case ship is about 500 kW while the average exhaust temperature for this power production has been measured approximately as 294 °C. The thermal power absorbed by ORC working fluid from exhaust gases, which has a mass flow rate 160 kg/min, has been calculated as 372 kW for cruising conditions of the case naval ship. The power that the ORC system can produce with the working fluids is seen in Table 7 and Fig. 10. The temperature and pressure information on the table is the conditions of the ORC working fluid at the turbine inlet. Regeneration has not been calculated for R123 and R141b since both they are isentropic fluids. Maximum power that can be recovered from waste heat of exhaust gases and maximum cycle efficiency was calculated as 118 kW and 0.32, respectively with the fluid toluene using an ideal regenerator. Overall exergy efficiency and irreversibility are presented for the case ORC cycles in Table 7, Fig. 11, and Fig. 12. It is seen on the figures that the cycle with toluene and regenerator has the maximum power, maximum exergy efficiency and minimum irreversibility among the selected case fluids. The navies of NATO countries use naval distillate fuel (NATO symbol F-76) on power systems. Some values of the logistic fuel NATO F-76 are presented in Table 8 [19,25].The efficiency of the case diesel generator can be calculated as:
_ W
g ¼ _ generator Q combustion
ð27Þ
_ generator is 500 kW for the cruising conditions, and Q_ combustion where W is:
_ u Q_ combustion ¼ mH
ð28Þ
_ is the mass flow rate of diesel fuel given in Table 4. Therwhere m mal efficiency of the diesel generator without ORC is calculated (gDG ) as 0.294. The overall efficiency of combined diesel generator-ORC system can be defined as:
g¼
_ ORC _ generator þ W W _ Q combustion
ð29Þ
In case of using an ideal ORC with regenerator and toluene working fluid, the total power generated is 618 kW and the overall efficiency of the combined system becomes 0.364. The efficiency increases from 0.294 to 0.364 using the ideal ORC system. However, an ideal ORC with isentropic turbine is not realistic. The isentropic efficiencies of various types of prototype expansion machines have been reported in [4] where presented efficiencies ranges from 10% to 85% depending on the expander types, working fluids, and the working conditions.
Table 6 Pinch point (PP) temperatures.
Fig. 9. The effects of regenerator effectiveness, turbine and pump efficiencies on the cycle efficiency.
temperature inside the evaporator should not be less than that of the ORC fluid near it. The pinch point temperature is critical especially for toluene and benzene since their evaporator temperatures
Working fluids
Evaporation temperature (°C)
PP temperature difference (°C)
Toluene (no regenerator) Toluene (with regenerator) Benzene (no regenerator) Benzene (with regenerator) n-Hexane (no regenerator) n-Hexane (with regenerator)
225 225 225 225 200 200
15 10 17 14 55 45
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Table 7 The efficiency, power, overall exergy efficiency, and irreversibility of the ORC on cruising. R123a T3 = 160 °C P3 = 2496 kPa
Efficiency Power, kW Overall exergy efficiency % Irreversibility, kW a
Isopentane T3 = 160 °C P3 = 2202 kPa
n-Pentane T3 = 160 °C P3 = 1892 kPa
n-Hexane T3 = 200 °C P3 = 1787 kPa
Benzene T3 = 225 °C P3 = 2108 kPa
Toluene T3 = 225 °C P3 = 1137 kPa
No regen.
With regen.
No regen.
With regen.
No regen.
With regen.
No regen.
With regen.
No regen.
With regen.
53.2
0.229 85.2 57.8
0.201 74.8 50.7
0.235 87.4 59.3
0.205 76.3 51.7
0.237 88.2 59.8
0.231 85.9 58.2
0.29 107.9 73.2
0.287 106.8 72.4
0.304 113.1 76.7
0.284 105.7 71.7
0.316 117.6 79.7
69.1
62.2
72.7
60
71.3
59.3
61.7
39.5
40.7
34.4
41.8
29.9
0.211
R141b T3 = 175 °C P3 = 2695 kPa
The temperature and pressure information on the table is the conditions of the ORC working fluid at the turbine inlet.
Fig. 10. The power that can be generated from ORC on cruising.
Fig. 11. Exergy efficiency of the cycles.
The isentropic efficiency of a turbine is defined as:
w gt ¼ t ws
ð30Þ
where wt is actual turbine work per for unit mass, and ws is isentropic turbine work. Bala et al. studied the overall efficiency of sliding-vane refrigerant pumps; the highest reported efficiency was about 20% [26]. Melotte performed an experimental study on a centrifugal pump and obtained an efficiency varying between 10% and 20% [27]. The isentropic efficiency of a pump is defined as:
gp ¼
ws wp
ð31Þ
where wp is actual pump work per for unit mass, and ws is isentropic pump work. To be realistic, assuming an isentropic efficiency for the turbine of the case ORC as 0.75, an isentropic efficiency for the pump as 0.20, and neglecting the losses at the ORC electric generator, the electric power output of the cycle with toluene and regenerator becomes 92 kW. In this case the power of the combined system becomes 592 kW while the efficiency of the combined system is equal to 0.349. If the selected ORC cycle is used to generate power from waste heat, 25,500 L of diesel fuel (US$24,870) can be saved and 67.2 tons of CO2 emissions reduced at the end of 1000 operating hours a year of a naval surface ship [19].
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Fig. 12. Irreversibility of the cycles.
hexhaust 103 W=m2 K
Table 8 Values of the logistic fuel NATO F-76.
htoluene 3000 W=m2 K
Molecular formula (Average)
C14.8H26.9
Molecular weight Sulfur content, wt.% (max) Density, at 15 °C, kg m3 (max) Fuel price, US$ gallon1, (2016) Lower heating value, Hu, kJ kg1
205 0.1 876 2.97 42,700
It is reported that the cost of some working ORC systems of different manufacturers ranges from $1800/kW to $2857/kW [28]. Assuming an average cost, $2300/kW for the selected ORC system, the overall cost of the system will be approximately $210,000. The system saves about $24,870 at the end of 1000 operating hours a year. It means that, the system will be profitable after 8 years. But if the operating hours of the ship will be more than 1000 h in a year, the system will be profitable before 8 years. Log mean temperature difference at the evaporator is calculated 64 °C for toluene with regenerator. The typical average convection heat transfer coefficients and fouling resistances can be taken from Ref. [24] as:
Rf ;exhaust 0:000308 m2 K=W Rf ;benzene 0:000352 m2 K=W Uo is calculated as 93.44 W/m2 K with the Eq. (26). The heat transfer in the evaporator is taken to be 372 kW. The required heat transfer area at the evaporator for toluene can be calculated as Arequired 60 m2. But it has been assumed that there is no fin in the evaporator, and the size of the evaporator can be reduced using fins inside the heat exchanger. Fig. 8 shows that the efficiency of n-hexane with regenerator is higher than other case ORC fluids when the turbine inlet temperature is less than 190 °C, but for the higher temperature values, the efficiency of toluene with regenerator is highest compared to the other working fluids. For the changing exhaust temperatures of the ship, different working fluids can be used to reach a higher efficiency. But it is not possible only a single ORC system. For such a case, that means, when different fluids give better efficiencies at different temperatures, a combined ORC system can be used. The proposed system is seen in Fig. 13. In this system, two different turbines are connected to the same generator via two hydraulic couplings. When the temperature is
Fig. 13. Combined organic Rankine cycle.
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below 190 °C, ORC system 1 runs with n-hexane. As the exhaust temperature gets higher than 190 °C, the ORC system 1 stops and the second system starts working. So higher efficiencies can be reached with this system compared to a single ORC system. Addition to the turbines, there are two pumps, a three-fluid heat exchanger for condenser, and a three-fluid heat exchanger for the evaporator.
Acknowledgments The views and conclusions contained herein are those of the author and should not be interpreted as necessarily representing official policies or endorsements, either expressed or implied, of any affiliated organization or government. References
8. Conclusions A waste heat powered ORC power system for a naval surface ship was designed and thermodynamically analyzed. Seven different working fluids at different turbine inlet temperatures were used as the working fluids and the efficiencies were calculated. A regenerator after the turbine was used for the dry fluids to improve the cycle efficiency. The study estimated the efficiency between 0.201 and 0.316 depending on the type of the fluid, the inlet temperature of the turbine, and regenerator. The efficiency, power, overall exergy efficiency, and irreversibility of the ORC with benzene on cruising highly good. Yet, benzene is a human carcinogen and most non-industrial applications have been limited. Therefore, benzene should be used carefully on ships. The electrical load of the case naval ship is about 500 kW with 294 °C average exhaust temperature at mass flow rate of 160 kg/ min. The exit temperature of exhaust from the regenerator was taken 158 °C for the calculations not to cause sulfuric acid formation in exhaust manifolds and funnel. The powers of the ideal ORC’s were calculated and compared. The results show that a net power can be obtained from the cycle between 76.3 and 117.6 kW depending on the selected fluids and the operating conditions. The exergy efficiencies and irreversibilities were calculated. Among the selected working fluids, the ORC cycle working with toluene with regenerator was found to have the maximum generated power and exergy efficiency. Assuming an isentropic efficiency for the turbine and the pump of the case ORC equal to 0.75 and 0.20, respectively, and neglecting the losses at the ORC electric generator, the electric power output of the cycle with toluene and regenerator becomes 92 kW. In this case the power of the combined system is calculated as 592 kW while the efficiency of the combined system is equal to about 0.349. If the selected ORC cycle is used to generate power from waste heat, 25,500 L of diesel fuel (US$24,870) can be saved and 67.2 tons of CO2 emissions reduced at the end of 1000 operating hours a year of a naval surface ship. The system will be profitable after 8 years assuming 1000 h cruising time for a year. The excess of production is stored in Li-ion batteries because of high energy densities, good cycle life and high charge/discharge efficiency The results show that the ORC generates power from waste heat of the exhaust energy emitted to atmosphere and thus reduce the fuel that is used on ship and global CO2 emissions. It has been proposed a combined ORC system to work at different exhaust temperatures to reach higher efficiencies. Author contributions _ Ibrahim Girgin conceived and designed the research, analyzed _ the data. Ibrahim Girgin and Cüneyt Ezgi worked out the theory, and wrote the manuscript, read and approved the final manuscript. Conflicts of interest The author declares no conflict of interest.
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