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LNG-FSRU cold energy recovery regasification using a zeotropic mixture of ethane and propane
Energy 173 (2019) 857e869
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LNG-FSRU cold energy recovery regasification using a zeotropic mixture of ethane and propane Lee Yoonho Mokpo National Maritime University, 91, Haeyangdaehak-ro, Mokpo-si, Jeollanam-do 58628, South Korea
a r t i c l e i n f o
a b s t r a c t
Article history: Received 14 September 2018 Received in revised form 13 February 2019 Accepted 14 February 2019 Available online 15 February 2019
This study developed a cold energy recovery regasification system for recovering and utilizing waste cold energy from liquefied natural gas floating storage regasification units, and analyzed its thermal, exergy, and economic efficiencies using a zeotropic mixture of ethane and propane. The single-stage version exhibited the highest net output, thermal efficiency, and exergy efficiency for a 6:4 ethane/propane mixture. The existing method (using only propane as working fluid) exhibited a thermal efficiency of 3.5%, exergy efficiency of 5.9% at 25 C, and thermal efficiency of 3.8%. The exergy efficiency was 6.2% because the exergy loss was reduced by 300 MJ/h compared to that for the conventional method. The highest thermal and exergy efficiencies (6.1% and 10.9%) in the two-stage version were obtained for an 8:2 ethane/propane mixture. The thermal efficiency was 6.6% and exergy loss was 16,300 MJ/h compared to the existing method, showing a 10.9% improvement. Thermal and exergy efficiencies of the two-stage version were higher than those for the one-stage system by 2.7% and 4.7%, respectively, providing an annual net income of USD 3.60 million and reduced electricity production costs by 0.0021 USD/kWh. The system could reduce exergy loss and electricity production costs while increasing the annual net income. © 2019 Elsevier Ltd. All rights reserved.
Keywords: Cold energy Exergy Liquid natural gas Regasification Zeotropic mixture
1. Introduction To facilitate its storage and sea transport, natural gas is transformed into liquefied natural gas (LNG) by decreasing its temperature to less than 162 C at atmospheric pressure, a process that requires approximately 300 kWh of energy [1]. The LNG is subsequently regasified before being supplied to customers on land, with seawater used as the main heat source in the regasification process. Although this utilization of seawater simplifies the system, the cold energy extracted from the LNG is wasted into the sea. At an air temperature of 20 C, approximately 240 kWh of cold energy is produced per ton of evaporated LNG [2], prompting numerous researchers to investigate the recovery and utilization of this currently wasted energy. Methods for recovering cold energy are commonly divided into direct expansion cycle (DEC) methods that use LNG as a working fluid, organic Rankine cycle (ORC) methods that recover cold energy based on the temperature difference between seawater (the major heat source) and the LNG (the heat sink) and the properties of the utilized hydrocarbon refrigerants, and combined cycle (CC)
E-mail address: [email protected]. https://doi.org/10.1016/j.energy.2019.02.111 0360-5442/© 2019 Elsevier Ltd. All rights reserved.
methods that combine the two former methods [3]. Recent studies investigated new cold energy recovery methods that utilize other working fluids or additional or modified processes to improve energy recovery rates [4e24]. Through mathematical modeling and analyses of five working fluids, including C4F10, Chang et al. [4] showed that an ORC that used LNG as its low-temperature heat source could produce an output higher than that of a thermal power generation system that used LNG as fuel. Szargut et al. [5] also investigated the economic optimization of LNG cold energy power generation plants. Zhang et al. [6] studied a power generation cycle that combined DEC and ORC methods with a gaseliquid separator. Meanwhile, Choi et al. [7] analyzed a multistage ORC cycle and described an optimization method for cold energy recovery from LNG. In addition, Gomez et al. [8] proposed and thermodynamically analyzed a power generation cycle that combined the Brayton and Rankine cycles. Lee [9] investigated the efficiency of a cold energy recovery regasification system that used ethane and propane as working fluids and reported that the multistage cascade Rankine cycle system exhibited the highest energy efficiency when ethane was used as the working fluid in the first stage and propane was used in the second stage. Ahmadi et al. [10] performed energy and exergy analyses according to the LNG cooling application in a power generation plant
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Nomenclature ex h s m_ Q_ T _ W
h subscripts sw wf p in out v H
exergy specific enthalpy, kJ/kg specific entropy, kJ/kg mass flow rate, kg/h heat transfer rate, kJ/s temperature, C work transfer rate, kW efficiency, % and superscripts seawater working fluid pressure, kPa inlet outlet volumetric flow, m3 =s isentropic enthalpy difference across the turbine, kJ=kg
combined with SOFC, gas turbine, and ORC. Consequently, they obtained up to 75.5% exergy efficiency. Sadaghiani et al. [11] proposed a new-concept LNG cold recovery system combining ORC and Kalina cycles in the power generation plant and found that it could improve the exergy efficiency of 32.15% over the existing method. Ahmadi et al. [12] performed thermal and economic efficiency analyses using optimization techniques in systems using flat plate collectors and LNG coolers in the transcritical CO2 power cycle. As a result, they achieved a thermal efficiency of 10.55%. Naseri et al. [13] analyzed the efficiency of the cooling utilization of LNG in the solar-driven transcritical power cycle and found a net power of approximately 26.8%. Ahmadi et al. [14] also studied optimization analyses for thermodynamics and exergy efficiency using NASA-II algorithms when LNG cooling was applied as heat sink in a transitive CO2 power cycle. These works revealed that the optimal working fluid or cycle depends on the heat source, which is an external factor. Despite the improvements achieved in the energy efficiency of LNG cold energy recovery, based on the assumption of a cryogenic LNG at 160 C and a working fluid at 20 C, significant exergy loss occurs in the heat exchanger because of the 180 C temperature difference. Among the various available methods for reducing exergy loss, the use of a zeotropic mixture instead of a single refrigerant as the working fluid in the heat exchanger has drawn much attention. A zeotropic mixture produces a Lorentz cycle, in which the refrigerant temperature varies along the flow direction during the phase change in a countercurrent heat exchanger; hence, the loss of available energy that may occur during heat exchange can be reduced to enhance the system performance. In addition, an appropriate pressure range can be maintained by changing the fluid combination or the composition of the refrigerant based on the temperature range of the heat source. Furthermore, continuous capacity control can be achieved by adjusting the composition ratio. This type of zeotropic mixture can be used as a substitute for certain freons under controlled conditions. However, the use of a zeotropic mixture requires a certain degree of difference in the boiling and dew points of the constituent fluids, and the refrigerant and heat source water must be countercurrent [15,16]. Kim et al. [17] studied a power generation facility for recovering
th CI OM N n i ∅ Acronyms SF HEX LNG NG WF C CRF EPC ATNI EP ORC
thermal initial investment total operation and maintenance costs annual system operational hours, hour system lifetime, year annual interest rate, % maintenance factor and abbreviations turbine size factor, m heat exchanger liquefied natural gas natural gas working fluid capital investment cost capital recovery factor electricity production cost annual total net income electricity price organic Rankine cycle
cold energy from LNG. They found that when the maximum output is desired from a multistage power generation system that recovers LNG cold energy at heat source temperatures of 25e85 C, the most efficient approach involved the use of an R-14/propane mixture in the first stage and an ethane/n-pentane mixture in the second and third stages. They also noted that output increased with the heat source temperature. Wang et al. [18] optimized a system that utilized a regenerative Rankine cycle, in which an ammonia/water mixture was used as the working fluid, in conjunction with LNG cold energy. Oh [19] also investigated the design of a regenerative Rankine cycle with an added regenerator that used an ammonia/ water mixture as the working fluid in a system for generating power through cold energy recovery from LNG using a low-grade heat source. Junjiang et al. [20] recently studied the optimal mixing ratio using a power generation system that uses LNG as a cool heat source by dividing 10 other working fluids into pure fluid, binary mixture, and terrier mixture. Hua Tian et al. [21] also performed a thermo-economic analysis of non-conventional refrigerants based on siloxanes to utilize engine waste heat using dual-loop ORC. Meanwhile, Yang et al. [22] performed a thermo-economic analysis according to the R1234yf/R32 mixing ratio in the transcritical organic Rankine cycle for the recovery of waste heat from low heat. In addition, Lee et al. [23] studied the optimal mix ratio of CF4, CHF3, and n-pentane for utilizing exergy in the power generation plant and improved the efficiency of the exergy by 7.7% compared to the existing method. The design of the ORC processes using LNG cold heat has also been studied. Consequently, 26.2% thermal efficiency can be achieved when the waste heat temperature was 85 C. An annual profit of 39 M$ was also achieved [24]. Until recently, studies on the application of refrigerants to ORC systems using LNG cold heat have been continued, but most have focused on land power generation plans. Cold energy recovery regasification systems that use mixed refrigerants have also been extensively studied. Nevertheless, insufficient attention has been given to the thermodynamic characteristics and exergy loss of such systems with respect to the composition ratio of the zeotropic mixture used as the working fluid [15e24]. An ORC-based cold energy recovery regasification system for the efficient recovery of the cold energy generated by LNG vaporization in a floating storage
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regasification unit (FSRU), which is a representative regasification facility, is developed herein. The system's heat transfer, energy, and exergy characteristics for the varying composition ratios of the ethane/propane zeotropic mixture used as the working fluid are then analyzed. Moreover, the annual net income and electricity production costs afforded by the system with respect to the composition ratio of the working fluid and the seawater temperature based on the assets and service lives of the devices that constitute the regasification system were examined. The better working fluid composition ratio for this system was ultimately determined.
source was set to 25 C; hence, the working fluid supplied to the turbine was assumed to transform into saturated vapor at 20 C. Table 2 lists the composition ratios of the LNG supplied to the cold energy recovery regasification system. Aspen HYSYS (HYSYS, 2004) process modeling software was used in the present process design of the cold energy recovery regasification system. The thermodynamic properties of the natural gas were calculated using the PengeRobinson state equation [26,27].
2. Process design
Fig. 1 shows the temperature gradient of the zeotropic ethane/ propane mixture with respect to the ethane mole fraction for a constant pressure of 110 kPa. The temperature variation of a zeotropic working fluid is similar to that of the heat source fluid because the temperature depends on the refrigerant's composition during a phase change under constant pressure. The curve shows no temperature gradient when pure ethane or propane is used as the working fluid. However, the temperature gradient gradually increases with the mole fraction of ethane, reaching a maximum at 60% ethane and then decreasing. Hence, a zeotropic mixture can reduce exergy loss during heat exchange better than a singlecomponent or azeotropic refrigerant. In addition, using a zeotropic mixture as the working fluid in cold energy recovery regasification enables capacity adjustments and more effective energy utilization. Table 3 details the physical properties of the working fluid with respect to its ratio of ethane and propane. Note that the molecular weight, boiling point, and critical temperature of the mixture decrease as the ratio of ethane increases. Conversely, the critical pressure increases for ratios up to 80% ethane, but slightly decreases when pure ethane is used as the working fluid.
2.1. Feed gas parameters Table 1 presents the assumptions made in designing a cold energy recovery regasification system that uses a zeotropic ethane/ propane mixture. The insulation efficiencies of all the pumps and turbines were set to 80%, which is the generally used value. A pressure drop across the system components, except for the pumps and the turbines, was not assumed. The minimum temperature difference of the working fluid in the vaporizer was set to 5 C (e.g., if the temperature of the seawater used as the heat source is 15 C, the temperature of the cold energy recovery working fluid cannot exceed 10 C). In addition, the LNG is supplied to customers on land as natural gas at 5 C through the regasification system after pressurization to 8 MPa by an LNG pump. For the process analysis, no heat or friction loss was assumed to occur in any system components, including the piping, and that the mixer, splitter, and piping had no installation cost. To compare this study's results to the conditions similar to those of an actual case, the capacity of the LNG-FSRU was set to 800 mmSCFD based on a report of the California State Lands Commission [25]. The cold energy recovery working fluid liquefies as it absorbs cold energy from the LNG in the vaporizer. The pressure of the working fluid that passes through the condenser must be maintained at or above the atmospheric pressure; otherwise, should leakage occur, the condenser's performance may decrease because of the inflow of refractory gas, possibly resulting in a large explosion caused by the combustibility of the gases involved. Therefore, the working fluid was assumed to exist as a saturated vapor at the evaporator outlet and as a saturated liquid at the condenser outlet. Thus, the pressure of the working fluid at the condenser outlet was set to 110 kPa. The temperature of the seawater used as the heat
Table 1 Process simulation assumptions [7]. Variable
Value
Adiabatic efficiency of the pumps Adiabatic efficiency of the turbines Pressure drop in the heat exchangers Pressure drop in the pump Mass flow rate of the LNG Minimum approach temperature in HEX1 Minimum approach temperature in HEX3 Minimum pressure in HEX1 Temperature of the seawater entering the seawater pumps Pressure of the seawater discharged from the pumps Temperature of the LNG entering the LNG pumps Pressure of the LNG entering the LNG pumps Pressure of the natural gas produced Temperature of the natural gas produced Quality of the inlet expansion turbine Quality of the HEX1 outlet
80% 80% 0 0 198.26 kg s1 5 C 5 C 110 kPa 20 C 750 kPa 162 C 100 kPa 8000 kPa 5 C x¼1 x¼0
2.2. Zeotropic mixtures as working fluids
2.3. Regasification Fig. 2 shows the process flow chart of the single-stage cold energy recovery regasification system. As indicated, the LNG emitted from storage tank LNG_1 was pressurized by the LNG pump to suit the application (LNG_2), being transformed into gaseous natural gas as it passes through the HEX1 and HEX2 heat exchangers. The power of the working fluid in the saturated vapor state attained by heating with seawater (WF_1) is recovered as the fluid passes through turbine WF_2, with its pressure decreased to 110 kPa. If the pressure at the turbine outlet falls below atmospheric pressure, it may cause inflow of the refractory gas or ignite a fire in the event of a leak. Therefore, the pressure at the condenser outlet must be 110 kPa or higher. The working fluid subsequently exchanges heat with the LNG in the HEX1 heat exchanger, and the working fluid in the saturated liquid state (WF_3) then passes through working fluid pump WF_4 and absorbs heat from seawater in the HEX3 heat exchanger for conversion again into the saturated vapor state
Table 2 Composition of the LNG [25]. Component
Mole fraction
Methane (CH4) Ethane (C2H6) Propane (C3H8) i-Butane (i-C4H10) n-Butane (n-C4H10) i-Pentane (i-C5H12) n-Pentane (n-C5H12) Nitrogen (N2)
0.9133 0.0536 0.0214 0.0047 0.0046 0.0001 0.0001 0.0022
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Fig. 2. Process flow chart of the single-stage cold energy recovery regasification system.
Fig. 1. Temperature gradient of the working fluid with respect to the ethane mole fraction.
(WF_1). A larger temperature difference between the condenser and the evaporator in the ORC systems enables the generation of more power; hence, the efficiency of the cold energy recovery regasification system is heavily dependent on the seawater's temperature. That is, more energy can be obtained through the resultant larger pressure difference between the turbine inlet and outlet because of the higher saturation pressure produced by the seawater temperature. The efficiency of cold energy recovery regasification can be maximized by maximizing the pressure and the temperature of the cold energy recovery working fluid supplied to the turbine. The temperature of the seawater herein was assumed to be 25 C. Fig. 3 shows the temperatureeentropy (Tes) diagram of a single-stage cold energy recovery regasification system using the zeotropic mixture as the working fluid. The WF_4ʹeWF_1 section of the evaporator and the WF_2eWF_3 section of the condenser in the figure show the mixture's characteristics. The composition ratio of the mixture continuously changed in the liquid and gaseous phases. In addition, during phase changes under constant pressure, the temperature change was similar to that of the heat source fluid because the temperature depended on the mixture's composition. A two-stage version of the system was constructed by adding another heat exchange stage to further improve the heat transfer efficiency of the HEX1 heat exchanger used to recover the cold energy from LNG during single-stage regasification. In the twostage system, the boiling point of the working fluid of the ORC of the first stage must be lower than that of the second stage. Therefore, a working fluid with a lower boiling point must be used for the first stage to use propane as the working fluid of the second stage ORC. The ethane/propane zeotropic mixture was used herein in the first stage. Five mole fractions of ethane (i.e., 20%, 40%, 60%, 80%, and 100%) were considered because of the need for the boiling point of the mixture to be lower than that of propane. The results were then analyzed.
Fig. 4 shows the process flow chart of the two-stage cold energy recovery regasification system. The two-phase LNG that passes through the HEX1 heat exchanger was transformed into natural gas while passing through the HEX2 heat exchanger. Fig. 5 shows the Tes diagram of the two-stage system, which enabled greater cold energy recovery and utilization from the LNG compared to the single-stage system.
Fig. 3. Tes diagram of the single-stage cold energy recovery regasification system.
Table 3 Physical properties of the mixture working fluid with respect to the composition ratio. Mixture
Composition
Molecular weight (g/mol)
Normal boiling point ( C)
Critical pressure (kPa)
Critical temperature ( C)
R190/R270
0.0/1.0 0.2/0.8 0.4/0.6 0.6/0.4 0.8/0.2 1.0/0.0
44.10 41.29 38.49 35.68 32.88 30.07
42.19 46.53 51.73 58.36 67.88 88.73
4257 4582 4853 5037 5073 4884
96.75 87.33 76.53 64.03 49.47 32.28
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In applying the PengeRobinson equation to the mixture, the mixing rules can be applied to the energy and size parameters, which can be expressed as Eqs. (6) and (7), respectively.
amix ¼
XX xi xj aij i
bmix ¼
(6)
j
X xi bi
(7)
i
Fig. 4. Process flow chart of the two-stage cold energy recovery regasification system.
In Eq. (6), the mixing rule of van der Waals is used, which can be used in Eq. (8), for aij that is the energy parameter for the bisector between components “i” and “j.”
aij ¼
pffiffiffiffiffiffiffiffi ai aj 1 kij
(8)
In Eq. (8), kij is a heterogeneity interaction parameter determined through regression to better estimate the basal equilibrium experimental data between each of the isomers, where the isomeric interaction parameters for ethane and propane are applied as 0.0010. 3.2. Mass and energy analysis The mass and energy balances for each component of the cold energy recovery regasification system can be calculated using Eqs. (9) and (10), respectively, with the assumption of negligible kinetic and potential energies.
X
X m_ out
m_ in ¼
X X m_ in hin m_ out hout
_ ¼ Q_ W
Fig. 5. Tes diagram of the two-stage cold energy recovery regasification system.
(9) (10)
The applied energy balance equation was derived using Eqs. (11)e(18).
Turbine ¼ w_ t ¼ mwf $ðhin hout Þ
(11)
Pump ¼ w_ p ¼ mwf $ðhout hin Þ
(12)
Condenser ¼ mLNG $ hout LNG hin LNG ¼ mwf $ hin wf hout wf
(13)
3. Analysis of system thermodynamics 3.1. Equation of state The PengeRobinson equation is applied to the equation, which can be used as shown in Eq. (1) [28].
RT aa P¼ v b vðv þ bÞ bðv bÞ
(1)
In Eq. (1), a and b are the energy and size parameters, which are expressed in Eqs. (2) and (3) as functions of critical temperature and critical pressure, respectively.
R2 T 2C a ¼ 0:45723 PC b ¼ 0:07780
(3)
In Eq. (1), a is a function of the eccentric factor of each component used to better estimate the vapor pressure depending on the temperature of the pure component as an alpha function. The formula is expressed as shown in Eqs. (4) and (5).
h
pffiffiffiffiffi i Tr
m ¼ 0:37464 þ 1:54336u 0:26992u
wf
hout
wf
¼ msw $ hout
hin
sw
sw
(14) _ net ¼ W
X _ W
turbine$i
X _ W
i
(2)
RTC PC
a¼ 1þm 1
Evaporator ¼ mwf $ hin
pump$i
i
¼
X _ W
output$i
i
X _ W
input$i
i
(15) _ output ¼ W
X
_ W turbine$i
(16)
i
_ W input ¼
X
_ W pump
i
LNG
þ
X _ W i
pump WF$i
þ
X
_ W pump
SW$i
i
(17) The thermal efficiency for a regasification system is:
(4)
.
hth ¼ W_ net Q_ 2
(5)
(18)
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3.3. Exergy analysis
EPC ¼
3600Ctotal ; Wnet
An exergy analysis was conducted on each of the system components, including the turbines, pumps, and heat exchangers. The exergy of a component is given by:
ATNI ¼ 7300ðEP EPCÞ$Wnet ;
_ ¼ mð _ _ DexÞ ¼ m½ðh h0 Þ T0 ðs s0 Þ: Ex
where EP represents the electricity price.
(19)
Exergy analysis is applied to the integrated power system to determine the irreversibilities at the various operating stages. The dead state (reference point) for all exergy evaluations is adjusted to the Tdead-state ¼ 298.15 K and the Pdead-state ¼ 101.3 kPa. The exergy losses and efficiencies of the different system components were calculated using the following equations:
_ _ ¼ mðex in exout Þ þ W pump
_ DEx loss
pump
_ DEx loss
turbine
_ DEx loss
HEX
hex ¼
_ _ ¼ mðex in exout Þ W turbine
¼ m_ 1 ðex1$in ex1$out Þ m_ 2 ðex2$out ex2$in Þ
Wnet ExLNG þ Ex SW P
Exergy output ¼1 Exergy input
hex ¼ P
(20) (21) (22) (23)
P
Exergy loss in each equipment P Exergy input (24)
3.4. Economic analysis The various performance measures of a thermodynamic system are generally assessed based on the first and second laws of thermodynamics. However, an economic evaluation of the application of such a system, including the initial investment and operating cost, is also necessary. The annual net income and the electricity production cost of the proposed cold energy recovery regasification system with respect to the composition of the zeotropic mixture and the seawater temperature were evaluated. The capital investment cost of the regasification system Ctotal is given by Eq. (25).
Ctotal ¼
X
Z_ CI þ Z_ OM k;
(25)
k
where Z_ CI$k denotes the initial investment for the system components, and Z_ OM$k represents the total operation and maintenance costs. The sum of these cost components can be expressed as follows [29]:
Zk $∅ CRF; Z_ CI;k þ Z_ OM;k ¼ N$3600
(26)
where Zk is the initial investment for each system component; ∅ is the maintenance factor; and N is the system operation time. The capital recovery factor (CRF) can be defined as follows [29]:
CRF ¼
ið1 þ iÞn ; ð1 þ iÞn 1
(27)
where n is the system lifetime, and i is the annual interest rate. The electricity production cost (EPC) and the annual total net income (ATNI) of the system are given as follows, respectively [30]:
(28) (29)
3.5. Validation The data reported by Shouguang et al. [31] were used to validate the reliability of the mathematical model herein. The literature involves a system that is the same as the single- and two-stage cold energy recovery regasification system. Seawater was taken as the heat source, while LNG was taken as the cold source. The working fluid was then applied with propane for the first stage, ethane for the second stage, and propane. The input values were according to Table 4. Table 5 shows the comparison between the results calculated by the mathematical model herein and those reported in the literature. The relative error of the single-stage exergy efficiency was only 3.7%, while that of the two-stage exergy efficiency was only 3.2%, proving the reliability of the mathematical model in this paper. 4. Results and discussion The turbine size, turbine outlet dryness, thermal efficiency, and exergy efficiency of the single-stage cold energy recovery regasification system were examined with respect to the ethane/propane composition ratio to determine the better composition ratio. Fig. 6 shows the saturation pressure with respect to the evaporator outlet temperature and the ethane/propane composition ratio. The saturation temperature was highest when pure ethane was used as the refrigerant, regardless of the temperature. The saturation pressure also increased with the evaporator outlet temperature because of the increase in the seawater temperature. However, the rate of the saturation pressure's increase slowly decreased when increasing the ethane ratio. Based only on the variation in the saturation pressure of the working fluid with the evaporator outlet temperature, the use of pure ethane as the refrigerant increased the system output because it increased the pressure of the working fluid supplied to the turbine. Fig. 7 shows the dryness at the turbine outlet with respect to the evaporator outlet temperature and the ethane/propane composition ratio. The use of pure ethane as the working fluid significantly decreased the dryness because of the increase in the evaporator outlet temperature, with the dryness falling below 0.8 for an outlet temperature of 30 C. If the dryness at the turbine outlet becomes too low, liquid would enter the turbine, and the turbine blade
Table 4 Input values for validation [31]. Parameters
Value
Working fluid LNG inlet temperature ( C) Natural gas outlet temperature ( C) Natural gas outlet pressure (MPa) Seawater inlet temperature ( C) Seawater outlet temperature ( C) Pinch temperature difference in the heat exchangers ( C) Degree of supercooling of the HEX1 outlet ( C) Isentropic efficiency of the turbine Isentropic efficiency of the pump Working medium condensing pressure (MPa)
R290/R170 162 5 8 20 15 5 2 0.8 0.75 0.11
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Table 5 Comparison of the results obtained from this paper and the literature [31]. Parameters
Literature [31]
This paper
Absolute error
Relative error (%)
Evaporation pressure (MPa) Mass flow (kg/h) Exergy efficiency of the single-stage system Exergy efficiency of the two-stage system
0.7 1.2 0.081 0.156
0.7 1.16 0.078 0.151
0.0 0.4 0.03 0.05
0.0 3.3 3.7 3.2
Fig. 6. Saturation pressure with respect to the evaporator outlet temperature and ethane/propane composition ratio.
Fig. 8. Turbine size factor with respect to the evaporator outlet temperature and ethane/propane composition ratio.
SFTurbine ¼
Fig. 7. Dryness at the turbine outlet with respect to the evaporator outlet temperature and ethane/propane composition ratio.
(rotating at high speed) may be damaged by the liquid droplets hitting it. Based on the present results, dryness should be maintained at 0.94 or higher for an ethane mole fraction of 60% or lower and an evaporator temperature of 0e30 C. Fig. 8 shows the turbine size factor with respect to the seawater temperature and the ethane/propane composition ratio, as determined by Eq. (30) [32,33]. The turbine size is an important consideration because it significantly affects the cost of the ORC system.
v0:5 H0:25
(30)
where SFTurbine is the turbine size factor ðmÞ; v is the volumetric flow through the turbine ðm3 =sÞ; and H is the isentropic enthalpy difference across the turbine ðkJ=kgÞ. As shown in Fig. 8, the turbine size gradually decreased with the increasing ethane content and seawater temperature. For the evaporator outlet temperatures of 0e5 C, the use of ethane as the working fluid required a 2.85 m turbine, which is approximately 1/2.2 of the turbine size (6.24 m) required when using propane as a working fluid. The turbine size gradually increased with the ethane content of the working fluid, indicating that an appropriate ethane/propane composition ratio can be used to reduce the turbine size and cost. Fig. 9 shows the working fluid mass flow rate with respect to the evaporator outlet temperature and the ethane/propane composition ratio. The required mass flow rate of the ORC system decreased with the increasing ethane content for two reasons: the temperature increase of the natural gas passing through the vaporizer due to the larger heat load of the propane evaporator and the change in the latent heat of vaporization of the working fluid. A higher mass flow rate increases the output of the turbine, but also requires a more powerful working fluid pump and a larger heat exchanger; hence, the net output of the system should be considered. Fig. 10 shows the system net output with respect to the evaporator outlet temperature and the ethane/propane composition ratio. For an evaporator outlet temperature of 15 C, the net output was the lowest (below 2080 kW) when pure ethane was used as the working fluid. It then significantly increased with the addition of propane. For an evaporator outlet temperature of 20 C or lower, the net output was highest for an ethane/propane composition ratio of 60:40, followed by 40:60, 20:80, 80:20, and 100:0 ratios.
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Fig. 9. Working fluid mass flow rate with respect to the evaporator outlet temperature and ethane/propane composition ratio.
Fig. 11. System thermal efficiency with respect to the evaporator outlet temperature and ethane/propane composition ratio.
Fig. 10. System net output with respect to the evaporator outlet temperature and ethane/propane composition ratio. Fig. 12. System exergy efficiency with respect to the evaporator outlet temperature and ethane/propane composition ratio.
Furthermore, for an evaporator outlet temperature of 20 C or higher, the highest net output was obtained using pure propane as the working fluid. Fig. 11 shows the thermal efficiency with respect to the evaporator outlet temperature and the ethane/propane composition ratio. For an evaporator outlet temperature of 15 C, the thermal efficiency was the lowest (less than 1.2%) when pure ethane was used as the working fluid, but significantly increased with the addition of propane. For an evaporator outlet temperature of 20 C or lower, thermal efficiency was the highest (3.8%) for an ethane/ propane composition ratio of 60:40, followed by 40:60, 20:80, 80:20, and 100:0 ratios. In contrast, for an evaporator outlet temperature higher than 20 C, the thermal efficiency was the highest (4.8%) when pure propane was used as the working fluid, producing an increase of 3.6% compared to the thermal efficiency when pure ethane was used as the working fluid. Fig. 12 shows the exergy efficiency with respect to the evaporator outlet temperature and the ethane/propane composition ratio. For an evaporator outlet temperature of 15 C, the exergy efficiency was the lowest (below 2.1%) when ethane was used as the working fluid, but significantly increased with the addition of
propane. For an evaporator outlet temperature of 20 C or lower, the exergy efficiency was the highest (5.7%) for an ethane/propane composition ratio of 60:40, followed by 40:60, 20:80, 80:20, and 100:0. Furthermore, for an evaporator outlet temperature of 20 C or higher, the exergy efficiency was the highest (7.9%) when pure propane was used as the working fluid, producing an improvement of 5.8% compared to the exergy efficiency when ethane was used as the working fluid, which was the least efficient. These results determined that the ratio of the mixture with the highest exergy efficiency varied depending on the sea temperature. Moreover, when the temperature of the sea was 20 C, the energy efficiency of ethane and propane was improved by 0.6% compared to using only propane, the existing pure fluid. Fig. 13 shows the thermal and exergy efficiencies of the singlestage cold energy recovery regasification system with respect to the ethane/propane composition ratio for a seawater temperature of 25 C. First, the 20% mixture of ethane slightly decreased to 3.3% compared to using propane only as the working fluid. The thermal efficiency has since slightly increased, with a maximum efficiency of 3.8% when ethane and propane were mixed at 6:4. Even as the
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Fig. 13. Thermal and exergy efficiencies of the single-stage cold energy recovery regasification system with respect to the ethane/propane composition ratio.
amount of ethane significantly increased, its thermal efficiency gradually decreased, thereby improving the thermal efficiency of 0.3% compared to the application of propane as a single refrigerant. In looking at the exergy efficiency, the maximum value was 6.2% when the ratio of ethane to propane was 6:4. This result confirmed that it can improve efficiency when mixed with ethane rather than only applying propane as the working fluid at a sea temperature of 25 C. Fig. 14 shows the exergy loss of each component of the singlestage cold energy recovery regasification system with respect to the ethane/propane composition ratio for a seawater temperature of 25 C. The HEX1 heat exchanger, in which the largest temperature difference occurred when pure propane was used as the working fluid, had a larger exergy loss compared with the other components. In addition, the exergy loss of the HEX1 heat exchanger gradually decreased as the proportion of ethane increased, whereas that of the HEX2 heat exchanger, in which heat exchange between the LNG and seawater occurred before the latter was supplied as natural gas, increased. The figure shows that the exergy loss in the heat exchanger varied depending on the ratio of ethane to propane, and the pump
Fig. 14. Exergy loss of each component of the single-stage cold energy recovery regasification system with respect to the ethane/propane composition ratio.
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did not change much compared to the heat exchanger. When only propane was applied as an operating fluid, HEX1 showed approximately 15,000 MJ/h of exergy, 2700 MJ/h of HEX2, 2000 MJ/h of HEX3, and a total of 19,700 MJ/h of exergy in heat exchangers. The exergy loss of HEX1 at 20% mixed with ethane was reduced by 13,200 MJ/h compared to that of propane alone, with 3700 MJ/h for HEX2 and 2900 MJ/h for HEX3 and 100 MJ/h for the heat exchanger, compared to a total of 19,800 MJ/h for propane alone. When the mixture ratio of ethane increased by 60%, HEX2 showed approximately 10,000 MJ/h at HEX1 and HEX2 at 5800 MJ/h and 3600 MJ/h at HEX3, showing a total of 19,400 MJ/h loss of extras. This was able to reduce the loss of 300 MJ/h exergy compared to the traditional method of using only propane. Fig. 15 shows the thermal and exergy efficiencies of the twostage cold energy recovery regasification system with respect to the ethane/propane composition ratio for a seawater temperature of 25 C. Propane was used as the second-stage working fluid in this system and an ethane/propane mixture with a lower boiling point was used as the first-stage working fluid. Thus, the composition ratio of ethane on the x-axis began at 0.2. When propane was mixed by 20%, the thermal efficiency was increased to 6.6% compared to using ethane as the single working fluid. The thermal efficiency later continued to decrease as the amount of propane increased. The maximum efficiency was confirmed when ethane and propane were mixed at 8:2. When looking at the exergy efficiency, the maximum value was 10.9% when the ratio of ethane to propane was 8:2. This result indicates that the efficiency can be improved when mixed with propane rather than only ethane as the operating fluid of the first stage at a sea temperature of 25 C. Fig. 16 depicts the exergy loss of each component of the twostage cold energy recovery regasification system with respect to the ethane/propane composition ratio for a seawater temperature of 25 C. The highest exergy loss occurred in the HEX1 heat exchanger, where the largest temperature difference was also observed, as in the single-stage system. In addition, the exergy loss of the HEX1 heat exchanger gradually decreased with the increasing ethane proportion. In contrast, that of the HEX2 heat exchanger, in which heat exchange between the LNG and seawater occurred before the former was supplied as natural gas, increased. However, the magnitude of the exergy loss decrease of the HEX1 heat exchanger was larger. Compared to the exergy loss of the single-stage system, the exergy loss of the HEX1 heat exchanger
Fig. 15. Exergy efficiency of the two-stage cold energy recovery regasification system with respect to the ethane/propane composition ratio.
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L. Yoonho / Energy 173 (2019) 857e869 Table 7 Economic analysis parameters. Parameter
Value
Reference
Annual system operational hours, N Annual interest rate, i System lifetime, n Maintenance factor, ∅ Processing capacity of LNG-FSRU
7300 h 14% 20 years 1.06 800 mmSCFD
[29] [34] [7] [34] [25]
Fig. 16. Exergy loss of each component of the two-stage cold energy recovery regasification system with respect to the ethane/propane composition ratio.
conditions used for the economic efficiency analysis. The annual operation time of the LNG-FSRU was assumed to be 7300 h, and the annual interest rate was set to 14%. The lifetime of the unit was set to 20 years, with a maintenance factor of 1.06. Fig. 17 shows the annual net income of the single-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio of the working fluid. The annual net income peaked at an ethane/propane composition ratio of 60:40 with the decreasing seawater temperature. In addition, based on the average results for the seawater
was approximately 14,000 MJ/h lower. The figure shows that the exergy loss in the heat exchangers and the turbines will vary depending on the ratio of ethane to propane, and as with the single stage, the pump was less variable than the heat exchanger. Therefore, if you look at the exergy loss in the heat exchanger and apply only conventional ethane as the working fluid, you lose approximately 6400 MJ/h at HEX1, 3800 MJ/h at HEX2, 3200 MJ/h at HEX3, and 1200 MJ/h at HEX3 and 1100 at HEX5 and W2 at HF. A 20% mix of propane resulted in approximately 8100 MJ/h loss at HEX1, 1800 MJ/h at HEX2, 3600 MJ/h at HEX3, approximately 800 MJ/h at HEX5, and a total loss of 800 MJ/h at the WF2 turbine (1600 MJ/h). This resulted in a reduction in the loss of 1100 MJ/h over the traditional ethane-only approach. The annual net incomes and the electricity production costs afforded by the application of the zeotropic mixture to the singleand two-stage cold energy recovery regasification systems were compared with respect to seawater temperature and interest rate. ASPEN Process Economic Analysis (HYSYS 2004.2) was used to achieve accurate cost and component comparisons of the two systems. Table 6 reveals an initial cost difference of USD 5,763,800 between the single- and two-stage systems [26]. Table 7 shows the
Fig. 17. Annual net income of the single-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio.
Table 6 Component costs of the cold energy recovery regasification systems [26]. Regasification system type
Component
Component type
Component cost (USD)
Total system cost (USD)
1-ORC
HEX1 HEX2 HEX3 Turbine_WF Pump_LNG Pump_SW1 Pump_SW2 Pump_WF HEX1 HEX2 HEX3 HEX4 HEX5 Turbine_WF1 Turbine_WF2 Pump_LNG Pump_SW1 Pump_SW2 Pump_WF1 Pump_WF2
DHE FLOAT HEAD DHE FLOAT HEAD DHE FLOAT HEAD ETURGAS DCP CENTRIF DCP CENTRIF DCP CENTRIF DCP CENTRIF DHE FLOAT HEAD DHE FLOAT HEAD DHE FLOAT HEAD DHE FLOAT HEAD DHE FLOAT HEAD ETURGAS ETURGAS DCP CENTRIF DCP CENTRIF DCP CENTRIF DCP CENTRIF DCP CENTRIF
1,989,700 748,800 1,585,200 8,159,200 865,700 2,642,700 448,600 300,600 1,292,300 1,837,400 688,300 878,000 749,000 7,219,900 5,231,100 865,700 2,851,400 449,000 284,200 158,000
16,740,500
2-ORC
22,504,300
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temperatures of 10, 20, and 30 C, the highest net income was USD 2.76 million per year for an ethane/propane composition ratio of 60:40 and USD 2.70 million per year for pure propane, indicating an annual net income of up to USD 62,200 compared with the existing methods. Fig. 18 depicts the electricity production costs using the singlestage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio. For a seawater temperature of 20 C or lower, the electricity production cost was the lowest (0.06 USD/kWh) for an ethane/propane composition ratio of 60:40. Based on the average results for the seawater temperatures of 10, 20, and 30 C, the electricity production costs were the lowest at 0.0626 USD/kWh for an ethane/ propane composition ratio of 60:40 as opposed to 0.0663 USD/kWh for pure propane, indicating savings of 0.0016 USD/kWh when compared with the existing methods. Fig. 19 shows the annual net income of the two-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio. The average results for the seawater temperatures of 10, 20, and 30 C revealed an annual net income of USD 6.36 million per year for an ethane/ propane composition ratio of 80:20 and USD 5.95 million per year for pure ethane, indicating an annual net income of approximately USD 0.4 million when compared with the existing method. The comparison of the case of an ethane/propane composition ratio of 60:40, which produced the highest annual net income in the single-stage system, and that of a composition ratio of 80:20, which produced the highest net income for the two-stage system, at a seawater temperature of 20 C revealed that a relative annual net income of up to USD 3.60 million can be achieved by the latter system. Moreover, the two-stage system afforded a more stable net income that was not significantly affected by seawater temperatures. Fig. 20 shows the electricity production cost of the two-stage system with respect to the seawater temperature and ethane/propane composition ratio. For a seawater temperature of 20 C or lower, the electricity production cost was the lowest at 0.0441 USD/ kWh at a composition ratio of 80:20. The average results for the seawater temperatures of 10, 20, and 30 C indicated that the electricity production cost when using an ethane/propane mixture as the working fluid was the lowest at 0.0451 USD/kWh for a composition ratio of 80:20, while the cost was 0.0468 USD/kWh
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Fig. 19. Annual net income of the two-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio.
Fig. 20. Electricity production cost of the two-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio.
when pure ethane was used. These findings indicated a cost reduction of 0.0017 USD/kWh when ethane/propane was used. Regarding the electricity production costs, the comparison of the ethane/propane composition ratios of 60:40, which yielded the lowest electricity production cost for the single-stage system, and those of 80:20, which yielded the lowest electricity production cost for the two-stage system, at a seawater temperature of 20 C revealed that the latter case enabled cost savings of 0.0021 USD/ kWh. In addition, the two-stage system afforded a more stable electricity production cost that was not significantly affected by seawater temperatures.
5. Conclusion
Fig. 18. Electricity production cost of the single-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio.
Herein, we developed a regasification system for utilizing the waste cold energy of an LNG-FSRU and investigated the characteristics of this system for various compositions of the zeotropic ethane/propane mixture used as the working fluid. The obtained findings can be summarized as follows:
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1. In a single-stage cold energy recovery regasification system, the use of ethane instead of propane as the working fluid allowed one to reduce the turbine size by a factor of ~2.2, decrease the mass flow rate, and increase the saturation pressure. However, considering the significant decrease in dryness at the turbine outlet and the associated risk of liquid entering the turbine, it is recommended to use mixtures with ethane contents of 60%. 2. The conventional method, wherein pure propane is employed as the working fluid, achieved a thermal efficiency of 3.5% and an exergy efficiency of 5.9% (based on a seawater temperature of 25 C). In contrast, when a 6:4 ethane/propane mixture was utilized, thermal efficiency increased to 3.8%, and exergy efficiency increased to 6.2%, since the observed exergy loss was lower than that in the conventional method by 300 MJ/h. In a two-stage version of the system, the highest thermal and exergy efficiencies (6.1 and 10.9%, respectively) were obtained for an 8:2 ethane/propane mixture. 3. Exergy loss in the cold energy recovery regasification system mostly occurred in the HEX1 heat exchanger, where heat was exchanged between the LNG and the working fluid. Based on a seawater temperature of 25 C, the exergy loss was lowest for ethane/propane ratios of 6:4 (for the single-stage system) and 8:2 (for the two-stage system). 4. Compared to the case of using pure propane as the working fluid and based on average results obtained for seawater temperatures of 10, 20, and 30 C, the single-stage cold energy recovery regasification system was calculated to provide an annual net income of up to USD 62,200 and reduce electricity production costs to 0.0015 USD/kWh when a 6:4 ethane/propane mixture was used as the working fluid. The corresponding values for the two-stage system and an 8:2 ethane/propane mixture were determined as USD 0.4 million and 0.0016 USD/kWh, respectively. For a seawater temperature of 20 C, the two-stage system afforded a relative net income of USD 3.60 million per year and reduced electricity production costs to 0.0021 USD/kWh when compared to the single-stage system. Moreover, the use of a two-stage system with a zeotropic working fluid was shown to result in a more stable electricity production cost because electricity cost was not significantly affected by seawater temperature. 5. As such, the application of ethane and propane as working fluids in an ORC-type cold energy recovery regasification system resulted in increased energy and exergy efficiency and economic synergies depending on the mixing ratio. It was also noted that although the introduction of an extra stage resulted in a slightly more complex stem, it allowed for a significant increase of energy efficiency compared to that of the single-stage version.
Funding None. Disclosure of potential conflicts of interest The authors declare that they have no conflict of interest. Acknowledgement None. References [1] Timmerhaus KE, Reed RP. Fifty-years’ development of cryogenic liquefaction processes. New York: Springer; 2007.
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LNG-FSRU cold energy recovery regasification using a zeotropic mixture of ethane and propane Lee Yoonho Mokpo National Maritime University, 91, Haeyangdaehak-ro, Mokpo-si, Jeollanam-do 58628, South Korea
a r t i c l e i n f o
a b s t r a c t
Article history: Received 14 September 2018 Received in revised form 13 February 2019 Accepted 14 February 2019 Available online 15 February 2019
This study developed a cold energy recovery regasification system for recovering and utilizing waste cold energy from liquefied natural gas floating storage regasification units, and analyzed its thermal, exergy, and economic efficiencies using a zeotropic mixture of ethane and propane. The single-stage version exhibited the highest net output, thermal efficiency, and exergy efficiency for a 6:4 ethane/propane mixture. The existing method (using only propane as working fluid) exhibited a thermal efficiency of 3.5%, exergy efficiency of 5.9% at 25 C, and thermal efficiency of 3.8%. The exergy efficiency was 6.2% because the exergy loss was reduced by 300 MJ/h compared to that for the conventional method. The highest thermal and exergy efficiencies (6.1% and 10.9%) in the two-stage version were obtained for an 8:2 ethane/propane mixture. The thermal efficiency was 6.6% and exergy loss was 16,300 MJ/h compared to the existing method, showing a 10.9% improvement. Thermal and exergy efficiencies of the two-stage version were higher than those for the one-stage system by 2.7% and 4.7%, respectively, providing an annual net income of USD 3.60 million and reduced electricity production costs by 0.0021 USD/kWh. The system could reduce exergy loss and electricity production costs while increasing the annual net income. © 2019 Elsevier Ltd. All rights reserved.
Keywords: Cold energy Exergy Liquid natural gas Regasification Zeotropic mixture
1. Introduction To facilitate its storage and sea transport, natural gas is transformed into liquefied natural gas (LNG) by decreasing its temperature to less than 162 C at atmospheric pressure, a process that requires approximately 300 kWh of energy [1]. The LNG is subsequently regasified before being supplied to customers on land, with seawater used as the main heat source in the regasification process. Although this utilization of seawater simplifies the system, the cold energy extracted from the LNG is wasted into the sea. At an air temperature of 20 C, approximately 240 kWh of cold energy is produced per ton of evaporated LNG [2], prompting numerous researchers to investigate the recovery and utilization of this currently wasted energy. Methods for recovering cold energy are commonly divided into direct expansion cycle (DEC) methods that use LNG as a working fluid, organic Rankine cycle (ORC) methods that recover cold energy based on the temperature difference between seawater (the major heat source) and the LNG (the heat sink) and the properties of the utilized hydrocarbon refrigerants, and combined cycle (CC)
E-mail address: [email protected]. https://doi.org/10.1016/j.energy.2019.02.111 0360-5442/© 2019 Elsevier Ltd. All rights reserved.
methods that combine the two former methods [3]. Recent studies investigated new cold energy recovery methods that utilize other working fluids or additional or modified processes to improve energy recovery rates [4e24]. Through mathematical modeling and analyses of five working fluids, including C4F10, Chang et al. [4] showed that an ORC that used LNG as its low-temperature heat source could produce an output higher than that of a thermal power generation system that used LNG as fuel. Szargut et al. [5] also investigated the economic optimization of LNG cold energy power generation plants. Zhang et al. [6] studied a power generation cycle that combined DEC and ORC methods with a gaseliquid separator. Meanwhile, Choi et al. [7] analyzed a multistage ORC cycle and described an optimization method for cold energy recovery from LNG. In addition, Gomez et al. [8] proposed and thermodynamically analyzed a power generation cycle that combined the Brayton and Rankine cycles. Lee [9] investigated the efficiency of a cold energy recovery regasification system that used ethane and propane as working fluids and reported that the multistage cascade Rankine cycle system exhibited the highest energy efficiency when ethane was used as the working fluid in the first stage and propane was used in the second stage. Ahmadi et al. [10] performed energy and exergy analyses according to the LNG cooling application in a power generation plant
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Nomenclature ex h s m_ Q_ T _ W
h subscripts sw wf p in out v H
exergy specific enthalpy, kJ/kg specific entropy, kJ/kg mass flow rate, kg/h heat transfer rate, kJ/s temperature, C work transfer rate, kW efficiency, % and superscripts seawater working fluid pressure, kPa inlet outlet volumetric flow, m3 =s isentropic enthalpy difference across the turbine, kJ=kg
combined with SOFC, gas turbine, and ORC. Consequently, they obtained up to 75.5% exergy efficiency. Sadaghiani et al. [11] proposed a new-concept LNG cold recovery system combining ORC and Kalina cycles in the power generation plant and found that it could improve the exergy efficiency of 32.15% over the existing method. Ahmadi et al. [12] performed thermal and economic efficiency analyses using optimization techniques in systems using flat plate collectors and LNG coolers in the transcritical CO2 power cycle. As a result, they achieved a thermal efficiency of 10.55%. Naseri et al. [13] analyzed the efficiency of the cooling utilization of LNG in the solar-driven transcritical power cycle and found a net power of approximately 26.8%. Ahmadi et al. [14] also studied optimization analyses for thermodynamics and exergy efficiency using NASA-II algorithms when LNG cooling was applied as heat sink in a transitive CO2 power cycle. These works revealed that the optimal working fluid or cycle depends on the heat source, which is an external factor. Despite the improvements achieved in the energy efficiency of LNG cold energy recovery, based on the assumption of a cryogenic LNG at 160 C and a working fluid at 20 C, significant exergy loss occurs in the heat exchanger because of the 180 C temperature difference. Among the various available methods for reducing exergy loss, the use of a zeotropic mixture instead of a single refrigerant as the working fluid in the heat exchanger has drawn much attention. A zeotropic mixture produces a Lorentz cycle, in which the refrigerant temperature varies along the flow direction during the phase change in a countercurrent heat exchanger; hence, the loss of available energy that may occur during heat exchange can be reduced to enhance the system performance. In addition, an appropriate pressure range can be maintained by changing the fluid combination or the composition of the refrigerant based on the temperature range of the heat source. Furthermore, continuous capacity control can be achieved by adjusting the composition ratio. This type of zeotropic mixture can be used as a substitute for certain freons under controlled conditions. However, the use of a zeotropic mixture requires a certain degree of difference in the boiling and dew points of the constituent fluids, and the refrigerant and heat source water must be countercurrent [15,16]. Kim et al. [17] studied a power generation facility for recovering
th CI OM N n i ∅ Acronyms SF HEX LNG NG WF C CRF EPC ATNI EP ORC
thermal initial investment total operation and maintenance costs annual system operational hours, hour system lifetime, year annual interest rate, % maintenance factor and abbreviations turbine size factor, m heat exchanger liquefied natural gas natural gas working fluid capital investment cost capital recovery factor electricity production cost annual total net income electricity price organic Rankine cycle
cold energy from LNG. They found that when the maximum output is desired from a multistage power generation system that recovers LNG cold energy at heat source temperatures of 25e85 C, the most efficient approach involved the use of an R-14/propane mixture in the first stage and an ethane/n-pentane mixture in the second and third stages. They also noted that output increased with the heat source temperature. Wang et al. [18] optimized a system that utilized a regenerative Rankine cycle, in which an ammonia/water mixture was used as the working fluid, in conjunction with LNG cold energy. Oh [19] also investigated the design of a regenerative Rankine cycle with an added regenerator that used an ammonia/ water mixture as the working fluid in a system for generating power through cold energy recovery from LNG using a low-grade heat source. Junjiang et al. [20] recently studied the optimal mixing ratio using a power generation system that uses LNG as a cool heat source by dividing 10 other working fluids into pure fluid, binary mixture, and terrier mixture. Hua Tian et al. [21] also performed a thermo-economic analysis of non-conventional refrigerants based on siloxanes to utilize engine waste heat using dual-loop ORC. Meanwhile, Yang et al. [22] performed a thermo-economic analysis according to the R1234yf/R32 mixing ratio in the transcritical organic Rankine cycle for the recovery of waste heat from low heat. In addition, Lee et al. [23] studied the optimal mix ratio of CF4, CHF3, and n-pentane for utilizing exergy in the power generation plant and improved the efficiency of the exergy by 7.7% compared to the existing method. The design of the ORC processes using LNG cold heat has also been studied. Consequently, 26.2% thermal efficiency can be achieved when the waste heat temperature was 85 C. An annual profit of 39 M$ was also achieved [24]. Until recently, studies on the application of refrigerants to ORC systems using LNG cold heat have been continued, but most have focused on land power generation plans. Cold energy recovery regasification systems that use mixed refrigerants have also been extensively studied. Nevertheless, insufficient attention has been given to the thermodynamic characteristics and exergy loss of such systems with respect to the composition ratio of the zeotropic mixture used as the working fluid [15e24]. An ORC-based cold energy recovery regasification system for the efficient recovery of the cold energy generated by LNG vaporization in a floating storage
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regasification unit (FSRU), which is a representative regasification facility, is developed herein. The system's heat transfer, energy, and exergy characteristics for the varying composition ratios of the ethane/propane zeotropic mixture used as the working fluid are then analyzed. Moreover, the annual net income and electricity production costs afforded by the system with respect to the composition ratio of the working fluid and the seawater temperature based on the assets and service lives of the devices that constitute the regasification system were examined. The better working fluid composition ratio for this system was ultimately determined.
source was set to 25 C; hence, the working fluid supplied to the turbine was assumed to transform into saturated vapor at 20 C. Table 2 lists the composition ratios of the LNG supplied to the cold energy recovery regasification system. Aspen HYSYS (HYSYS, 2004) process modeling software was used in the present process design of the cold energy recovery regasification system. The thermodynamic properties of the natural gas were calculated using the PengeRobinson state equation [26,27].
2. Process design
Fig. 1 shows the temperature gradient of the zeotropic ethane/ propane mixture with respect to the ethane mole fraction for a constant pressure of 110 kPa. The temperature variation of a zeotropic working fluid is similar to that of the heat source fluid because the temperature depends on the refrigerant's composition during a phase change under constant pressure. The curve shows no temperature gradient when pure ethane or propane is used as the working fluid. However, the temperature gradient gradually increases with the mole fraction of ethane, reaching a maximum at 60% ethane and then decreasing. Hence, a zeotropic mixture can reduce exergy loss during heat exchange better than a singlecomponent or azeotropic refrigerant. In addition, using a zeotropic mixture as the working fluid in cold energy recovery regasification enables capacity adjustments and more effective energy utilization. Table 3 details the physical properties of the working fluid with respect to its ratio of ethane and propane. Note that the molecular weight, boiling point, and critical temperature of the mixture decrease as the ratio of ethane increases. Conversely, the critical pressure increases for ratios up to 80% ethane, but slightly decreases when pure ethane is used as the working fluid.
2.1. Feed gas parameters Table 1 presents the assumptions made in designing a cold energy recovery regasification system that uses a zeotropic ethane/ propane mixture. The insulation efficiencies of all the pumps and turbines were set to 80%, which is the generally used value. A pressure drop across the system components, except for the pumps and the turbines, was not assumed. The minimum temperature difference of the working fluid in the vaporizer was set to 5 C (e.g., if the temperature of the seawater used as the heat source is 15 C, the temperature of the cold energy recovery working fluid cannot exceed 10 C). In addition, the LNG is supplied to customers on land as natural gas at 5 C through the regasification system after pressurization to 8 MPa by an LNG pump. For the process analysis, no heat or friction loss was assumed to occur in any system components, including the piping, and that the mixer, splitter, and piping had no installation cost. To compare this study's results to the conditions similar to those of an actual case, the capacity of the LNG-FSRU was set to 800 mmSCFD based on a report of the California State Lands Commission [25]. The cold energy recovery working fluid liquefies as it absorbs cold energy from the LNG in the vaporizer. The pressure of the working fluid that passes through the condenser must be maintained at or above the atmospheric pressure; otherwise, should leakage occur, the condenser's performance may decrease because of the inflow of refractory gas, possibly resulting in a large explosion caused by the combustibility of the gases involved. Therefore, the working fluid was assumed to exist as a saturated vapor at the evaporator outlet and as a saturated liquid at the condenser outlet. Thus, the pressure of the working fluid at the condenser outlet was set to 110 kPa. The temperature of the seawater used as the heat
Table 1 Process simulation assumptions [7]. Variable
Value
Adiabatic efficiency of the pumps Adiabatic efficiency of the turbines Pressure drop in the heat exchangers Pressure drop in the pump Mass flow rate of the LNG Minimum approach temperature in HEX1 Minimum approach temperature in HEX3 Minimum pressure in HEX1 Temperature of the seawater entering the seawater pumps Pressure of the seawater discharged from the pumps Temperature of the LNG entering the LNG pumps Pressure of the LNG entering the LNG pumps Pressure of the natural gas produced Temperature of the natural gas produced Quality of the inlet expansion turbine Quality of the HEX1 outlet
80% 80% 0 0 198.26 kg s1 5 C 5 C 110 kPa 20 C 750 kPa 162 C 100 kPa 8000 kPa 5 C x¼1 x¼0
2.2. Zeotropic mixtures as working fluids
2.3. Regasification Fig. 2 shows the process flow chart of the single-stage cold energy recovery regasification system. As indicated, the LNG emitted from storage tank LNG_1 was pressurized by the LNG pump to suit the application (LNG_2), being transformed into gaseous natural gas as it passes through the HEX1 and HEX2 heat exchangers. The power of the working fluid in the saturated vapor state attained by heating with seawater (WF_1) is recovered as the fluid passes through turbine WF_2, with its pressure decreased to 110 kPa. If the pressure at the turbine outlet falls below atmospheric pressure, it may cause inflow of the refractory gas or ignite a fire in the event of a leak. Therefore, the pressure at the condenser outlet must be 110 kPa or higher. The working fluid subsequently exchanges heat with the LNG in the HEX1 heat exchanger, and the working fluid in the saturated liquid state (WF_3) then passes through working fluid pump WF_4 and absorbs heat from seawater in the HEX3 heat exchanger for conversion again into the saturated vapor state
Table 2 Composition of the LNG [25]. Component
Mole fraction
Methane (CH4) Ethane (C2H6) Propane (C3H8) i-Butane (i-C4H10) n-Butane (n-C4H10) i-Pentane (i-C5H12) n-Pentane (n-C5H12) Nitrogen (N2)
0.9133 0.0536 0.0214 0.0047 0.0046 0.0001 0.0001 0.0022
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Fig. 2. Process flow chart of the single-stage cold energy recovery regasification system.
Fig. 1. Temperature gradient of the working fluid with respect to the ethane mole fraction.
(WF_1). A larger temperature difference between the condenser and the evaporator in the ORC systems enables the generation of more power; hence, the efficiency of the cold energy recovery regasification system is heavily dependent on the seawater's temperature. That is, more energy can be obtained through the resultant larger pressure difference between the turbine inlet and outlet because of the higher saturation pressure produced by the seawater temperature. The efficiency of cold energy recovery regasification can be maximized by maximizing the pressure and the temperature of the cold energy recovery working fluid supplied to the turbine. The temperature of the seawater herein was assumed to be 25 C. Fig. 3 shows the temperatureeentropy (Tes) diagram of a single-stage cold energy recovery regasification system using the zeotropic mixture as the working fluid. The WF_4ʹeWF_1 section of the evaporator and the WF_2eWF_3 section of the condenser in the figure show the mixture's characteristics. The composition ratio of the mixture continuously changed in the liquid and gaseous phases. In addition, during phase changes under constant pressure, the temperature change was similar to that of the heat source fluid because the temperature depended on the mixture's composition. A two-stage version of the system was constructed by adding another heat exchange stage to further improve the heat transfer efficiency of the HEX1 heat exchanger used to recover the cold energy from LNG during single-stage regasification. In the twostage system, the boiling point of the working fluid of the ORC of the first stage must be lower than that of the second stage. Therefore, a working fluid with a lower boiling point must be used for the first stage to use propane as the working fluid of the second stage ORC. The ethane/propane zeotropic mixture was used herein in the first stage. Five mole fractions of ethane (i.e., 20%, 40%, 60%, 80%, and 100%) were considered because of the need for the boiling point of the mixture to be lower than that of propane. The results were then analyzed.
Fig. 4 shows the process flow chart of the two-stage cold energy recovery regasification system. The two-phase LNG that passes through the HEX1 heat exchanger was transformed into natural gas while passing through the HEX2 heat exchanger. Fig. 5 shows the Tes diagram of the two-stage system, which enabled greater cold energy recovery and utilization from the LNG compared to the single-stage system.
Fig. 3. Tes diagram of the single-stage cold energy recovery regasification system.
Table 3 Physical properties of the mixture working fluid with respect to the composition ratio. Mixture
Composition
Molecular weight (g/mol)
Normal boiling point ( C)
Critical pressure (kPa)
Critical temperature ( C)
R190/R270
0.0/1.0 0.2/0.8 0.4/0.6 0.6/0.4 0.8/0.2 1.0/0.0
44.10 41.29 38.49 35.68 32.88 30.07
42.19 46.53 51.73 58.36 67.88 88.73
4257 4582 4853 5037 5073 4884
96.75 87.33 76.53 64.03 49.47 32.28
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In applying the PengeRobinson equation to the mixture, the mixing rules can be applied to the energy and size parameters, which can be expressed as Eqs. (6) and (7), respectively.
amix ¼
XX xi xj aij i
bmix ¼
(6)
j
X xi bi
(7)
i
Fig. 4. Process flow chart of the two-stage cold energy recovery regasification system.
In Eq. (6), the mixing rule of van der Waals is used, which can be used in Eq. (8), for aij that is the energy parameter for the bisector between components “i” and “j.”
aij ¼
pffiffiffiffiffiffiffiffi ai aj 1 kij
(8)
In Eq. (8), kij is a heterogeneity interaction parameter determined through regression to better estimate the basal equilibrium experimental data between each of the isomers, where the isomeric interaction parameters for ethane and propane are applied as 0.0010. 3.2. Mass and energy analysis The mass and energy balances for each component of the cold energy recovery regasification system can be calculated using Eqs. (9) and (10), respectively, with the assumption of negligible kinetic and potential energies.
X
X m_ out
m_ in ¼
X X m_ in hin m_ out hout
_ ¼ Q_ W
Fig. 5. Tes diagram of the two-stage cold energy recovery regasification system.
(9) (10)
The applied energy balance equation was derived using Eqs. (11)e(18).
Turbine ¼ w_ t ¼ mwf $ðhin hout Þ
(11)
Pump ¼ w_ p ¼ mwf $ðhout hin Þ
(12)
Condenser ¼ mLNG $ hout LNG hin LNG ¼ mwf $ hin wf hout wf
(13)
3. Analysis of system thermodynamics 3.1. Equation of state The PengeRobinson equation is applied to the equation, which can be used as shown in Eq. (1) [28].
RT aa P¼ v b vðv þ bÞ bðv bÞ
(1)
In Eq. (1), a and b are the energy and size parameters, which are expressed in Eqs. (2) and (3) as functions of critical temperature and critical pressure, respectively.
R2 T 2C a ¼ 0:45723 PC b ¼ 0:07780
(3)
In Eq. (1), a is a function of the eccentric factor of each component used to better estimate the vapor pressure depending on the temperature of the pure component as an alpha function. The formula is expressed as shown in Eqs. (4) and (5).
h
pffiffiffiffiffi i Tr
m ¼ 0:37464 þ 1:54336u 0:26992u
wf
hout
wf
¼ msw $ hout
hin
sw
sw
(14) _ net ¼ W
X _ W
turbine$i
X _ W
i
(2)
RTC PC
a¼ 1þm 1
Evaporator ¼ mwf $ hin
pump$i
i
¼
X _ W
output$i
i
X _ W
input$i
i
(15) _ output ¼ W
X
_ W turbine$i
(16)
i
_ W input ¼
X
_ W pump
i
LNG
þ
X _ W i
pump WF$i
þ
X
_ W pump
SW$i
i
(17) The thermal efficiency for a regasification system is:
(4)
.
hth ¼ W_ net Q_ 2
(5)
(18)
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3.3. Exergy analysis
EPC ¼
3600Ctotal ; Wnet
An exergy analysis was conducted on each of the system components, including the turbines, pumps, and heat exchangers. The exergy of a component is given by:
ATNI ¼ 7300ðEP EPCÞ$Wnet ;
_ ¼ mð _ _ DexÞ ¼ m½ðh h0 Þ T0 ðs s0 Þ: Ex
where EP represents the electricity price.
(19)
Exergy analysis is applied to the integrated power system to determine the irreversibilities at the various operating stages. The dead state (reference point) for all exergy evaluations is adjusted to the Tdead-state ¼ 298.15 K and the Pdead-state ¼ 101.3 kPa. The exergy losses and efficiencies of the different system components were calculated using the following equations:
_ _ ¼ mðex in exout Þ þ W pump
_ DEx loss
pump
_ DEx loss
turbine
_ DEx loss
HEX
hex ¼
_ _ ¼ mðex in exout Þ W turbine
¼ m_ 1 ðex1$in ex1$out Þ m_ 2 ðex2$out ex2$in Þ
Wnet ExLNG þ Ex SW P
Exergy output ¼1 Exergy input
hex ¼ P
(20) (21) (22) (23)
P
Exergy loss in each equipment P Exergy input (24)
3.4. Economic analysis The various performance measures of a thermodynamic system are generally assessed based on the first and second laws of thermodynamics. However, an economic evaluation of the application of such a system, including the initial investment and operating cost, is also necessary. The annual net income and the electricity production cost of the proposed cold energy recovery regasification system with respect to the composition of the zeotropic mixture and the seawater temperature were evaluated. The capital investment cost of the regasification system Ctotal is given by Eq. (25).
Ctotal ¼
X
Z_ CI þ Z_ OM k;
(25)
k
where Z_ CI$k denotes the initial investment for the system components, and Z_ OM$k represents the total operation and maintenance costs. The sum of these cost components can be expressed as follows [29]:
Zk $∅ CRF; Z_ CI;k þ Z_ OM;k ¼ N$3600
(26)
where Zk is the initial investment for each system component; ∅ is the maintenance factor; and N is the system operation time. The capital recovery factor (CRF) can be defined as follows [29]:
CRF ¼
ið1 þ iÞn ; ð1 þ iÞn 1
(27)
where n is the system lifetime, and i is the annual interest rate. The electricity production cost (EPC) and the annual total net income (ATNI) of the system are given as follows, respectively [30]:
(28) (29)
3.5. Validation The data reported by Shouguang et al. [31] were used to validate the reliability of the mathematical model herein. The literature involves a system that is the same as the single- and two-stage cold energy recovery regasification system. Seawater was taken as the heat source, while LNG was taken as the cold source. The working fluid was then applied with propane for the first stage, ethane for the second stage, and propane. The input values were according to Table 4. Table 5 shows the comparison between the results calculated by the mathematical model herein and those reported in the literature. The relative error of the single-stage exergy efficiency was only 3.7%, while that of the two-stage exergy efficiency was only 3.2%, proving the reliability of the mathematical model in this paper. 4. Results and discussion The turbine size, turbine outlet dryness, thermal efficiency, and exergy efficiency of the single-stage cold energy recovery regasification system were examined with respect to the ethane/propane composition ratio to determine the better composition ratio. Fig. 6 shows the saturation pressure with respect to the evaporator outlet temperature and the ethane/propane composition ratio. The saturation temperature was highest when pure ethane was used as the refrigerant, regardless of the temperature. The saturation pressure also increased with the evaporator outlet temperature because of the increase in the seawater temperature. However, the rate of the saturation pressure's increase slowly decreased when increasing the ethane ratio. Based only on the variation in the saturation pressure of the working fluid with the evaporator outlet temperature, the use of pure ethane as the refrigerant increased the system output because it increased the pressure of the working fluid supplied to the turbine. Fig. 7 shows the dryness at the turbine outlet with respect to the evaporator outlet temperature and the ethane/propane composition ratio. The use of pure ethane as the working fluid significantly decreased the dryness because of the increase in the evaporator outlet temperature, with the dryness falling below 0.8 for an outlet temperature of 30 C. If the dryness at the turbine outlet becomes too low, liquid would enter the turbine, and the turbine blade
Table 4 Input values for validation [31]. Parameters
Value
Working fluid LNG inlet temperature ( C) Natural gas outlet temperature ( C) Natural gas outlet pressure (MPa) Seawater inlet temperature ( C) Seawater outlet temperature ( C) Pinch temperature difference in the heat exchangers ( C) Degree of supercooling of the HEX1 outlet ( C) Isentropic efficiency of the turbine Isentropic efficiency of the pump Working medium condensing pressure (MPa)
R290/R170 162 5 8 20 15 5 2 0.8 0.75 0.11
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Table 5 Comparison of the results obtained from this paper and the literature [31]. Parameters
Literature [31]
This paper
Absolute error
Relative error (%)
Evaporation pressure (MPa) Mass flow (kg/h) Exergy efficiency of the single-stage system Exergy efficiency of the two-stage system
0.7 1.2 0.081 0.156
0.7 1.16 0.078 0.151
0.0 0.4 0.03 0.05
0.0 3.3 3.7 3.2
Fig. 6. Saturation pressure with respect to the evaporator outlet temperature and ethane/propane composition ratio.
Fig. 8. Turbine size factor with respect to the evaporator outlet temperature and ethane/propane composition ratio.
SFTurbine ¼
Fig. 7. Dryness at the turbine outlet with respect to the evaporator outlet temperature and ethane/propane composition ratio.
(rotating at high speed) may be damaged by the liquid droplets hitting it. Based on the present results, dryness should be maintained at 0.94 or higher for an ethane mole fraction of 60% or lower and an evaporator temperature of 0e30 C. Fig. 8 shows the turbine size factor with respect to the seawater temperature and the ethane/propane composition ratio, as determined by Eq. (30) [32,33]. The turbine size is an important consideration because it significantly affects the cost of the ORC system.
v0:5 H0:25
(30)
where SFTurbine is the turbine size factor ðmÞ; v is the volumetric flow through the turbine ðm3 =sÞ; and H is the isentropic enthalpy difference across the turbine ðkJ=kgÞ. As shown in Fig. 8, the turbine size gradually decreased with the increasing ethane content and seawater temperature. For the evaporator outlet temperatures of 0e5 C, the use of ethane as the working fluid required a 2.85 m turbine, which is approximately 1/2.2 of the turbine size (6.24 m) required when using propane as a working fluid. The turbine size gradually increased with the ethane content of the working fluid, indicating that an appropriate ethane/propane composition ratio can be used to reduce the turbine size and cost. Fig. 9 shows the working fluid mass flow rate with respect to the evaporator outlet temperature and the ethane/propane composition ratio. The required mass flow rate of the ORC system decreased with the increasing ethane content for two reasons: the temperature increase of the natural gas passing through the vaporizer due to the larger heat load of the propane evaporator and the change in the latent heat of vaporization of the working fluid. A higher mass flow rate increases the output of the turbine, but also requires a more powerful working fluid pump and a larger heat exchanger; hence, the net output of the system should be considered. Fig. 10 shows the system net output with respect to the evaporator outlet temperature and the ethane/propane composition ratio. For an evaporator outlet temperature of 15 C, the net output was the lowest (below 2080 kW) when pure ethane was used as the working fluid. It then significantly increased with the addition of propane. For an evaporator outlet temperature of 20 C or lower, the net output was highest for an ethane/propane composition ratio of 60:40, followed by 40:60, 20:80, 80:20, and 100:0 ratios.
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Fig. 9. Working fluid mass flow rate with respect to the evaporator outlet temperature and ethane/propane composition ratio.
Fig. 11. System thermal efficiency with respect to the evaporator outlet temperature and ethane/propane composition ratio.
Fig. 10. System net output with respect to the evaporator outlet temperature and ethane/propane composition ratio. Fig. 12. System exergy efficiency with respect to the evaporator outlet temperature and ethane/propane composition ratio.
Furthermore, for an evaporator outlet temperature of 20 C or higher, the highest net output was obtained using pure propane as the working fluid. Fig. 11 shows the thermal efficiency with respect to the evaporator outlet temperature and the ethane/propane composition ratio. For an evaporator outlet temperature of 15 C, the thermal efficiency was the lowest (less than 1.2%) when pure ethane was used as the working fluid, but significantly increased with the addition of propane. For an evaporator outlet temperature of 20 C or lower, thermal efficiency was the highest (3.8%) for an ethane/ propane composition ratio of 60:40, followed by 40:60, 20:80, 80:20, and 100:0 ratios. In contrast, for an evaporator outlet temperature higher than 20 C, the thermal efficiency was the highest (4.8%) when pure propane was used as the working fluid, producing an increase of 3.6% compared to the thermal efficiency when pure ethane was used as the working fluid. Fig. 12 shows the exergy efficiency with respect to the evaporator outlet temperature and the ethane/propane composition ratio. For an evaporator outlet temperature of 15 C, the exergy efficiency was the lowest (below 2.1%) when ethane was used as the working fluid, but significantly increased with the addition of
propane. For an evaporator outlet temperature of 20 C or lower, the exergy efficiency was the highest (5.7%) for an ethane/propane composition ratio of 60:40, followed by 40:60, 20:80, 80:20, and 100:0. Furthermore, for an evaporator outlet temperature of 20 C or higher, the exergy efficiency was the highest (7.9%) when pure propane was used as the working fluid, producing an improvement of 5.8% compared to the exergy efficiency when ethane was used as the working fluid, which was the least efficient. These results determined that the ratio of the mixture with the highest exergy efficiency varied depending on the sea temperature. Moreover, when the temperature of the sea was 20 C, the energy efficiency of ethane and propane was improved by 0.6% compared to using only propane, the existing pure fluid. Fig. 13 shows the thermal and exergy efficiencies of the singlestage cold energy recovery regasification system with respect to the ethane/propane composition ratio for a seawater temperature of 25 C. First, the 20% mixture of ethane slightly decreased to 3.3% compared to using propane only as the working fluid. The thermal efficiency has since slightly increased, with a maximum efficiency of 3.8% when ethane and propane were mixed at 6:4. Even as the
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Fig. 13. Thermal and exergy efficiencies of the single-stage cold energy recovery regasification system with respect to the ethane/propane composition ratio.
amount of ethane significantly increased, its thermal efficiency gradually decreased, thereby improving the thermal efficiency of 0.3% compared to the application of propane as a single refrigerant. In looking at the exergy efficiency, the maximum value was 6.2% when the ratio of ethane to propane was 6:4. This result confirmed that it can improve efficiency when mixed with ethane rather than only applying propane as the working fluid at a sea temperature of 25 C. Fig. 14 shows the exergy loss of each component of the singlestage cold energy recovery regasification system with respect to the ethane/propane composition ratio for a seawater temperature of 25 C. The HEX1 heat exchanger, in which the largest temperature difference occurred when pure propane was used as the working fluid, had a larger exergy loss compared with the other components. In addition, the exergy loss of the HEX1 heat exchanger gradually decreased as the proportion of ethane increased, whereas that of the HEX2 heat exchanger, in which heat exchange between the LNG and seawater occurred before the latter was supplied as natural gas, increased. The figure shows that the exergy loss in the heat exchanger varied depending on the ratio of ethane to propane, and the pump
Fig. 14. Exergy loss of each component of the single-stage cold energy recovery regasification system with respect to the ethane/propane composition ratio.
865
did not change much compared to the heat exchanger. When only propane was applied as an operating fluid, HEX1 showed approximately 15,000 MJ/h of exergy, 2700 MJ/h of HEX2, 2000 MJ/h of HEX3, and a total of 19,700 MJ/h of exergy in heat exchangers. The exergy loss of HEX1 at 20% mixed with ethane was reduced by 13,200 MJ/h compared to that of propane alone, with 3700 MJ/h for HEX2 and 2900 MJ/h for HEX3 and 100 MJ/h for the heat exchanger, compared to a total of 19,800 MJ/h for propane alone. When the mixture ratio of ethane increased by 60%, HEX2 showed approximately 10,000 MJ/h at HEX1 and HEX2 at 5800 MJ/h and 3600 MJ/h at HEX3, showing a total of 19,400 MJ/h loss of extras. This was able to reduce the loss of 300 MJ/h exergy compared to the traditional method of using only propane. Fig. 15 shows the thermal and exergy efficiencies of the twostage cold energy recovery regasification system with respect to the ethane/propane composition ratio for a seawater temperature of 25 C. Propane was used as the second-stage working fluid in this system and an ethane/propane mixture with a lower boiling point was used as the first-stage working fluid. Thus, the composition ratio of ethane on the x-axis began at 0.2. When propane was mixed by 20%, the thermal efficiency was increased to 6.6% compared to using ethane as the single working fluid. The thermal efficiency later continued to decrease as the amount of propane increased. The maximum efficiency was confirmed when ethane and propane were mixed at 8:2. When looking at the exergy efficiency, the maximum value was 10.9% when the ratio of ethane to propane was 8:2. This result indicates that the efficiency can be improved when mixed with propane rather than only ethane as the operating fluid of the first stage at a sea temperature of 25 C. Fig. 16 depicts the exergy loss of each component of the twostage cold energy recovery regasification system with respect to the ethane/propane composition ratio for a seawater temperature of 25 C. The highest exergy loss occurred in the HEX1 heat exchanger, where the largest temperature difference was also observed, as in the single-stage system. In addition, the exergy loss of the HEX1 heat exchanger gradually decreased with the increasing ethane proportion. In contrast, that of the HEX2 heat exchanger, in which heat exchange between the LNG and seawater occurred before the former was supplied as natural gas, increased. However, the magnitude of the exergy loss decrease of the HEX1 heat exchanger was larger. Compared to the exergy loss of the single-stage system, the exergy loss of the HEX1 heat exchanger
Fig. 15. Exergy efficiency of the two-stage cold energy recovery regasification system with respect to the ethane/propane composition ratio.
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L. Yoonho / Energy 173 (2019) 857e869 Table 7 Economic analysis parameters. Parameter
Value
Reference
Annual system operational hours, N Annual interest rate, i System lifetime, n Maintenance factor, ∅ Processing capacity of LNG-FSRU
7300 h 14% 20 years 1.06 800 mmSCFD
[29] [34] [7] [34] [25]
Fig. 16. Exergy loss of each component of the two-stage cold energy recovery regasification system with respect to the ethane/propane composition ratio.
conditions used for the economic efficiency analysis. The annual operation time of the LNG-FSRU was assumed to be 7300 h, and the annual interest rate was set to 14%. The lifetime of the unit was set to 20 years, with a maintenance factor of 1.06. Fig. 17 shows the annual net income of the single-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio of the working fluid. The annual net income peaked at an ethane/propane composition ratio of 60:40 with the decreasing seawater temperature. In addition, based on the average results for the seawater
was approximately 14,000 MJ/h lower. The figure shows that the exergy loss in the heat exchangers and the turbines will vary depending on the ratio of ethane to propane, and as with the single stage, the pump was less variable than the heat exchanger. Therefore, if you look at the exergy loss in the heat exchanger and apply only conventional ethane as the working fluid, you lose approximately 6400 MJ/h at HEX1, 3800 MJ/h at HEX2, 3200 MJ/h at HEX3, and 1200 MJ/h at HEX3 and 1100 at HEX5 and W2 at HF. A 20% mix of propane resulted in approximately 8100 MJ/h loss at HEX1, 1800 MJ/h at HEX2, 3600 MJ/h at HEX3, approximately 800 MJ/h at HEX5, and a total loss of 800 MJ/h at the WF2 turbine (1600 MJ/h). This resulted in a reduction in the loss of 1100 MJ/h over the traditional ethane-only approach. The annual net incomes and the electricity production costs afforded by the application of the zeotropic mixture to the singleand two-stage cold energy recovery regasification systems were compared with respect to seawater temperature and interest rate. ASPEN Process Economic Analysis (HYSYS 2004.2) was used to achieve accurate cost and component comparisons of the two systems. Table 6 reveals an initial cost difference of USD 5,763,800 between the single- and two-stage systems [26]. Table 7 shows the
Fig. 17. Annual net income of the single-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio.
Table 6 Component costs of the cold energy recovery regasification systems [26]. Regasification system type
Component
Component type
Component cost (USD)
Total system cost (USD)
1-ORC
HEX1 HEX2 HEX3 Turbine_WF Pump_LNG Pump_SW1 Pump_SW2 Pump_WF HEX1 HEX2 HEX3 HEX4 HEX5 Turbine_WF1 Turbine_WF2 Pump_LNG Pump_SW1 Pump_SW2 Pump_WF1 Pump_WF2
DHE FLOAT HEAD DHE FLOAT HEAD DHE FLOAT HEAD ETURGAS DCP CENTRIF DCP CENTRIF DCP CENTRIF DCP CENTRIF DHE FLOAT HEAD DHE FLOAT HEAD DHE FLOAT HEAD DHE FLOAT HEAD DHE FLOAT HEAD ETURGAS ETURGAS DCP CENTRIF DCP CENTRIF DCP CENTRIF DCP CENTRIF DCP CENTRIF
1,989,700 748,800 1,585,200 8,159,200 865,700 2,642,700 448,600 300,600 1,292,300 1,837,400 688,300 878,000 749,000 7,219,900 5,231,100 865,700 2,851,400 449,000 284,200 158,000
16,740,500
2-ORC
22,504,300
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temperatures of 10, 20, and 30 C, the highest net income was USD 2.76 million per year for an ethane/propane composition ratio of 60:40 and USD 2.70 million per year for pure propane, indicating an annual net income of up to USD 62,200 compared with the existing methods. Fig. 18 depicts the electricity production costs using the singlestage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio. For a seawater temperature of 20 C or lower, the electricity production cost was the lowest (0.06 USD/kWh) for an ethane/propane composition ratio of 60:40. Based on the average results for the seawater temperatures of 10, 20, and 30 C, the electricity production costs were the lowest at 0.0626 USD/kWh for an ethane/ propane composition ratio of 60:40 as opposed to 0.0663 USD/kWh for pure propane, indicating savings of 0.0016 USD/kWh when compared with the existing methods. Fig. 19 shows the annual net income of the two-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio. The average results for the seawater temperatures of 10, 20, and 30 C revealed an annual net income of USD 6.36 million per year for an ethane/ propane composition ratio of 80:20 and USD 5.95 million per year for pure ethane, indicating an annual net income of approximately USD 0.4 million when compared with the existing method. The comparison of the case of an ethane/propane composition ratio of 60:40, which produced the highest annual net income in the single-stage system, and that of a composition ratio of 80:20, which produced the highest net income for the two-stage system, at a seawater temperature of 20 C revealed that a relative annual net income of up to USD 3.60 million can be achieved by the latter system. Moreover, the two-stage system afforded a more stable net income that was not significantly affected by seawater temperatures. Fig. 20 shows the electricity production cost of the two-stage system with respect to the seawater temperature and ethane/propane composition ratio. For a seawater temperature of 20 C or lower, the electricity production cost was the lowest at 0.0441 USD/ kWh at a composition ratio of 80:20. The average results for the seawater temperatures of 10, 20, and 30 C indicated that the electricity production cost when using an ethane/propane mixture as the working fluid was the lowest at 0.0451 USD/kWh for a composition ratio of 80:20, while the cost was 0.0468 USD/kWh
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Fig. 19. Annual net income of the two-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio.
Fig. 20. Electricity production cost of the two-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio.
when pure ethane was used. These findings indicated a cost reduction of 0.0017 USD/kWh when ethane/propane was used. Regarding the electricity production costs, the comparison of the ethane/propane composition ratios of 60:40, which yielded the lowest electricity production cost for the single-stage system, and those of 80:20, which yielded the lowest electricity production cost for the two-stage system, at a seawater temperature of 20 C revealed that the latter case enabled cost savings of 0.0021 USD/ kWh. In addition, the two-stage system afforded a more stable electricity production cost that was not significantly affected by seawater temperatures.
5. Conclusion
Fig. 18. Electricity production cost of the single-stage cold energy recovery regasification system with respect to the seawater temperature and ethane/propane composition ratio.
Herein, we developed a regasification system for utilizing the waste cold energy of an LNG-FSRU and investigated the characteristics of this system for various compositions of the zeotropic ethane/propane mixture used as the working fluid. The obtained findings can be summarized as follows:
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1. In a single-stage cold energy recovery regasification system, the use of ethane instead of propane as the working fluid allowed one to reduce the turbine size by a factor of ~2.2, decrease the mass flow rate, and increase the saturation pressure. However, considering the significant decrease in dryness at the turbine outlet and the associated risk of liquid entering the turbine, it is recommended to use mixtures with ethane contents of 60%. 2. The conventional method, wherein pure propane is employed as the working fluid, achieved a thermal efficiency of 3.5% and an exergy efficiency of 5.9% (based on a seawater temperature of 25 C). In contrast, when a 6:4 ethane/propane mixture was utilized, thermal efficiency increased to 3.8%, and exergy efficiency increased to 6.2%, since the observed exergy loss was lower than that in the conventional method by 300 MJ/h. In a two-stage version of the system, the highest thermal and exergy efficiencies (6.1 and 10.9%, respectively) were obtained for an 8:2 ethane/propane mixture. 3. Exergy loss in the cold energy recovery regasification system mostly occurred in the HEX1 heat exchanger, where heat was exchanged between the LNG and the working fluid. Based on a seawater temperature of 25 C, the exergy loss was lowest for ethane/propane ratios of 6:4 (for the single-stage system) and 8:2 (for the two-stage system). 4. Compared to the case of using pure propane as the working fluid and based on average results obtained for seawater temperatures of 10, 20, and 30 C, the single-stage cold energy recovery regasification system was calculated to provide an annual net income of up to USD 62,200 and reduce electricity production costs to 0.0015 USD/kWh when a 6:4 ethane/propane mixture was used as the working fluid. The corresponding values for the two-stage system and an 8:2 ethane/propane mixture were determined as USD 0.4 million and 0.0016 USD/kWh, respectively. For a seawater temperature of 20 C, the two-stage system afforded a relative net income of USD 3.60 million per year and reduced electricity production costs to 0.0021 USD/kWh when compared to the single-stage system. Moreover, the use of a two-stage system with a zeotropic working fluid was shown to result in a more stable electricity production cost because electricity cost was not significantly affected by seawater temperature. 5. As such, the application of ethane and propane as working fluids in an ORC-type cold energy recovery regasification system resulted in increased energy and exergy efficiency and economic synergies depending on the mixing ratio. It was also noted that although the introduction of an extra stage resulted in a slightly more complex stem, it allowed for a significant increase of energy efficiency compared to that of the single-stage version.
Funding None. Disclosure of potential conflicts of interest The authors declare that they have no conflict of interest. Acknowledgement None. References [1] Timmerhaus KE, Reed RP. Fifty-years’ development of cryogenic liquefaction processes. New York: Springer; 2007.
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