Experimental Study of a Lab-Scale Organic Rankine Cycle System for Heat and Water Recovery from Flue Gas in Thermal Power Plants †

energies

Article Experimental Study of a Lab-Scale Organic Rankine Cycle System for Heat and Water Recovery from Flue Gas in Thermal Power Plants †

Young-Min Kim 1,*, Assmelash Negash 1,2, Syed Safeer Mehdi Shamsi 2,3, Dong-Gil Shin 1 and Gyubaek Cho 1

1 Research Division of Environmental and Energy Systems, Korea Institute of Machinery and Materials, 156 Gajeongbuk-ro, Yuseong-gu, Daejeon 34104, Korea; [email protected] (A.N.); [email protected] (D.-G.S.); [email protected] (G.C.) 2 Environment & Energy Mechanical Engineering, Korea University of Science and Technology, 217 Gajeong-ro, Yuseong-gu, Daejeon 34113, Korea; [email protected] 3 Department of Mechanical Engineering, Aerospace and Aviation Campus, Air University, Kamra 43570, Pakistan * Correspondence: [email protected]; Tel.: +82-42-868-7377 † This article is based on a paper presented at the 22nd Conference on Process Integration, Modeling and Optimization for Energy Saving and Pollution Reduction held from 20 to 23 October 2019.

Abstract: Fossil fuel power plants can cause numerous environmental issues, owing to exhaust emissions and substantial water consumption. In a thermal power plant, heat and water recovery

from flue gas can reduce CO2 emissions and water demand. High-humidity flue gas averts the diffusion of pollutants, enhances the secondary transformation of air pollutants, and leads to smog weather; hence, water recovery from flue gas can also help to lessen the incidence of white plumes Citation: Kim, Y.-M.; Negash, A.; and smog near and around the power plant. In this study, a lab-scale system for heat and water Shamsi, S.S.M.; Shin, D.-G.; Cho, G. recovery from flue gas was tested. The flue gas was initially cooled by an organic Rankine cycle Experimental Study of a Lab-Scale Organic Rankine Cycle System for (ORC) system to produce power. This gas was further cooled by an aftercooler, using the same Heat and Water Recovery from Flue working fluid to condense the water and condensable particulate matter in the flue gas. The ORC ◦ Gas in Thermal Power Plants. system can produce approximately 220 W of additional power from flue gas at 140 C, with a thermal ◦ Energies 2021, 14, 4328. https:// efficiency of 10%. By cooling the flue gas below 30–40 C, the aftercooler can recover 60% of the doi.org/10.3390/en14144328 water in it.

Academic Editors: Xingxing Cheng Keywords: waste heat recovery; organic rankine cycle; ﬂue gas; water condensation; thermal and Dimitris S. Manolakos efﬁciency

Received: 2 June 2021 Accepted: 15 July 2021 Published: 18 July 2021 1. Introduction Power plants that use fossil fuels—such as coal-fired power plants—exhibit problems Publisher’s Note: MDPI stays neutral in fuel consumption, eventually resulting in high CO emissions, substantial water con- with regard to jurisdictional claims in 2 published maps and institutional affil- sumption, and harmful gas emissions. In a thermal power plant, heat and water recovery iations. from flue gas can reduce CO2 emissions and water requirements. Over the years, the organic Rankine cycle (ORC) has been used for waste heat recovery (WHR) from various heat sources, including automobile exhaust gas [1] and power plant flue gas. The sensitivity of the power generation industry to water availability limits is widespread, and increasing; this has increased the burden on power plant operators Copyright: © 2021 by the authors. to save water [2]. In the Netherlands, a typical 400-MW coal-fired power plant equipped Licensee MDPI, Basel, Switzerland. with a wet flue gas desulfurization unit utilizes 30 ton/h of makeup water in the boiler This article is an open access article distributed under the terms and feedwater. Meanwhile, 150 ton/h of water in the flue gas is released through the stack; conditions of the Creative Commons moreover, 90–120 ton/h of water is released by plants without flue gas desulfurization Attribution (CC BY) license (https:// units [3]. Therefore, water recovery from flue gas, and its reuse, can reduce water con- creativecommons.org/licenses/by/ sumption in thermal power plants [4]. Furthermore, the low temperature of the latent heat 4.0/). recovered can be utilized by heat pumps for district heating [5].

Energies 2021, 14, 4328. https://doi.org/10.3390/en14144328 https://www.mdpi.com/journal/energies Energies 2021, 14, 4328 2 of 13

Condensing the moisture in the flue gas after it has passed through the pollution control system can reduce emissions even further. In the late 1980s, Haldor Topsoe A/S created and patented wet sulfuric acid, in which SOx and water vapor react to form liquid sulfuric acid (H2SO4) at high concentrations by cooling the exhaust gas to lower than the dew point of sulfuric acid [6]. Zukeran et al. [7] studied the removal rate of SO2 and particulate matter (PM) reduction in marine engines via water condensation in flue gas. Mist particles were formed as a result of water condensation, and they were collected using an electrostatic precipitator. Jeong et al. [8] used an integrated condensing heat exchanger system composed of a heat exchanger, a cooling water storage tank, and an acid removal system. The PM produced by fossil fuel combustion is composed of filterable particulate matter (FPM) and condensable particulate matter (CPM). CPM is a gaseous material at flue gas temperature prior to discharge; however, it forms a particulate substance (PM2.5) after cooling in the plume. Because FPM emissions have significantly reduced due to the rapid development of FPM technology, CPM discharge has recently emerged as an important issue [9]. Qi et al. [10] showed that a low-low temperature electrostatic precipitator could reduce the amount of CPM considerably by cooling the exhaust gas below the dew point. Furthermore, Shuangchen et al. [11] studied the environmental impact of high- humidity flue gas discharged by power plants, and proposed that flue gas emissions of high humidity promoted the secondary transformation of air pollutants and reduced atmospheric visibility. In the case of the ORC for waste heat recovery, the system must be optimized to produce the maximum power output from the given heat source. Many researchers have conducted experiments on small-scale ORC systems for various purposes. Malavolta et al. [12] built and tested a small-scale ORC system with a scroll expander for a domestic combined heat and power plant (electric power output of 1–3 kW). Lecompte et al. [13] presented the experimental results for an 11-kWe ORC prototype with a twin-screw expander, which was utilized to calibrate and validate off-design models. Shao et al. [14] experimentally studied the performance and features of a kW-scale ORC system under various cooling conditions. For low-grade heat recovery, Bianchi et al. [15] varied the heat source temperature and organic fluid feed pump velocity in a micro-ORC driven by a piston expander. Yang et al. [16] experimentally investigated a 3-kW ORC with a scroll expander for low-grade waste heat under different operating parameters. Rosset et al. [17] experimentally examined the performance of a novel turbogenerator (TG) supported by gas-lubricated bearings for small-scale ORC systems, generating an electrical power of up to 2.3 kW. In this study, a lab-scale system for heat and water recovery from flue gas was devel- oped and tested. The system comprises an ORC for power generation with the integration of a pumped heat pipe cooling (HPC) cycle that uses a single working fluid to execute heat extraction for power output and water recovery.

2. Waste Heat and Water Recovery System A waste heat and water recovery (WHWR) system comprises an ORC (power generation unit) and cooling cycles (water recovery unit) using a single working fluid R134a accompanied by a phase change, as shown in Figure1[ 18]. The flue gas is initially cooled through heat extraction by the power unit (ORC), and then further cooled below the dew point of the water in the flue gas by the water recovery unit based on a pumped HPC mechanism, in which a circulation pump and low-temperature (LT) evaporator were used. The vaporized working fluid flows into the common condenser to dissipate heat together with the ORC unit. In this system, the evaporation temperature in the LT evaporator is close to the condensation temperature of the condenser. Therefore, the cooling temperature of the pumped HPC system was restricted by the ambient temperature conditions. Energies 2021, 14, 4328 3 of 14

Energies 2021, 14, 4328 3 of 13 temperature of the pumped HPC system was restricted by the ambient temperature conditions.

Figure 1. Schematic of a heat and water recovery system [18]. Figure 1. Schematic of a heat and water recovery system [18]. The ORC unit was optimized for maximum power output. Subsequently, water The ORC unit was optimized for maximum power output. Subsequently, water con- condensation commences in the water recovery unit at the dew point—that is, when densation commences in the water recovery unit at the dew point—that is, when the par- the partial pressure of the water vapor is equal to the saturation pressure of the water. tial pressure of the water vapor is equal to the saturation pressure of the water. According According to Dalton’s law, the partial pressure ratio of the water vapor must be equal to theto Dalton’s partial volume law, the ratio. partial In the pressure case of ratio a 600-MW of the coal-fired water vapor power must plant be withequal a to water the partial vapor volumevolume ofratio. 16% In in the flue case gas of at a 150 600-MW◦C, the coal dew-fired point power is 55.3 plant◦C, and with the a water WHWR vapor system volume can produceof 16% in 6.6 flue MW gas with at 150 a thermal°C, the dew efficiency point ofis 55.3 8%, and°C, and allow the for WHWR water system recovery can of produce 50% by cooling6.6 MW the with flue a gasthermal to 40 efficiency◦C at an ambientof 8%, and temperature allow for ofwater 20 ◦ C[recovery18]. of 50% by cooling the flue gas to 40 °C at an ambient temperature of 20 °C [18]. 3. Experimental Setup 3. ExperimentalIn this study, Setup a lab-scale system for heat and water recovery from flue gas was devel- opedIn and this tested. study, Figure a lab-scale2 shows system a schematic for heat of and the water WHWR recovery system. from In the flue experiment, gas was devel- flue gasoped from and a steamtested. boiler Figure using 2 shows natural a schematic gas as fuel of was the used. WHWR R134a system. was used In the as theexperiment, working fluid,flue gas and from coolant a steam water boiler was usedusing to natural condense gas theas fuel working was fluid.used. AfterR134apassing was used through as the theworking condenser fluid, andand reservoir,coolant water the working was used fluid to condense was split, the with working one stream fluid. suppliedAfter passing to a high-temperaturethrough the condenser (HT) evaporatorand reservoir, by the a high-pressure working fluid pump was split, for the with ORC one part, stream and sup- the otherplied streamto a high-temperature supplied to the LT(HT) evaporator evaporator by by a low-pressure a high-pressure pump pump for thefor pumpedthe ORC HPCpart, unit.and the The other flue gas stream from supplied the steam to boiler the LT above evaporator 200 ◦C was by firsta low-pressure cooled by the pump steam for from the thepumped steam HPC boiler unit. to below The flue 150 gas◦C from in order the tostea simulatem boiler the above lower 200 temperature °C was first of cooled flue gas by Energies 2021, 14, 4328 4 of 14 fromthe steam the fossil from fuel the powersteam boiler plant. to Two below parallel 150 °C heat in order exchangers to simulate were the used lower for each temperature HT and LTof flue evaporator, gas from in the order fossil to maximizefuel power the plant. heat andTwo water parallel recovery heat exchangers performance. were used for each HT and LT evaporator, in order to maximize the heat and water recovery performance.

Figure 2. SchematicFigure of 2. theSchematic experimental of the experimental setup of the setup WHWR of the system. WHWR system.

A scroll expander was used for power generation, and a load bank was used to dissipate the power output and adjust the rotation speed of the expander. The torque (model: UNIPULSE UTM II-50 NM, 0–50 Nm ± 0.03% FS) and speed (3600 pulses per rotation) of the expander were measured to determine the mechanical power output. The mass flow rate of the working fluid of the ORC part and the pumped HPC part was measured using three mass flow meters (model: RHEONIK RHM 08, 0.3~50 kg/min ± 0.1% RV). To evaluate the water recovery efficiency, two relative humidity sensors (model: EE310, 0–100% RH ± 1.3% RH) were used to measure the relative humidity before and after the WHWR system. Pressure sensors (model: Sensys PSH, 0–50 bar and 0–10 bar ± 0.15% FS) and temperature sensors (K-type thermocouple, 0–150 °C ± 0.5 °C) were used to measure the state of the fluid. The heat exchangers for the HT and LT evaporators were shell- and tube-type, as shown in Figure 3. The flue gas was routed through the tubes of the heat exchangers, while R134a flowed through the shell side. Sight glasses were used to inspect the state of the working fluid after it had passed through the evaporators.

Figure 3. Shell- and tube-type heat exchangers used for the HT and LT evaporators.

In the previous experiment, the thermal efficiency of the ORC part was too low, due to the difficulty in optimizing the operating conditions of the ORC system, owing to a large mismatch between the capacity of the scroll expander and the low capacity of the burner in the laboratory. Therefore, an additional working fluid (R134a), heated by steam from the boiler to the same state as the working fluid from the waste heat recovery of the flue gas, was added to the inlet of the expander. In Figure 2, the red line depicts the flow of the steam circulation for this purpose. However, the net power from the waste heat recovery of the flue gas in the increased power was calculated by using the portion of the

Energies 2021, 14, 4328 4 of 14

Energies 2021, 14, 4328 4 of 13

Figure 2. Schematic of the experimental setup of the WHWR system. A scroll expander was used for power generation, and a load bank was used to dissi- pateA the scroll power expander output was and used adjust for power the rotation generation, speed and of a theload expander. bank was used The torqueto dis- (model: UNIPULSEsipate the power UTM output II-50 and NM, adjust 0–50 the Nm rotati±on0.03% speed FS)of the and expander. speed (3600The torque pulses (model: per rotation) ofUNIPULSE the expander UTM II-50 were NM, measured 0–50 Nm to ± 0.03% determine FS) and the speed mechanical (3600 pulses power per rotation) output. of The mass flowthe expander rate of thewere working measured fluid to determine of the ORC the partmechanical and the power pumped output. HPC The part mass was flow measured usingrate of the three working mass fluid flow of meters the ORC (model: part and RHEONIK the pumped RHM HPC part 08, 0.3~50was measured kg/min using± 0.1% RV). Tothree evaluate mass flow the meters water (model: recovery RHEONIK efficiency, RHM two 08, relative0.3~50 kg/min humidity ± 0.1% sensors RV). To (model: evalu- EE310, 0–100%ate the water RH ±recovery1.3% RH) efficiency, were usedtwo relati to measureve humidity the relativesensors (model: humidity EE310, before 0–100% and after the WHWRRH ± 1.3% system. RH) were Pressure used to sensors measure (model: the relative Sensys humidity PSH, before 0–50 bar and and after 0–10 the WHWR bar ± 0.15% FS) andsystem. temperature Pressure sensors sensors (model: (K-type Sensys thermocouple, PSH, 0–50 bar 0–150 and 0–10◦C ± bar0.5 ± 0.15%◦C) were FS) and used tem- to measure perature sensors (K-type thermocouple, 0–150 °C ± 0.5 °C) were used to measure the state the state of the fluid. The heat exchangers for the HT and LT evaporators were shell- and of the fluid. The heat exchangers for the HT and LT evaporators were shell- and tube-type, tube-type, as shown in Figure3. The flue gas was routed through the tubes of the heat as shown in Figure 3. The flue gas was routed through the tubes of the heat exchangers, exchangers,while R134a flowed while R134athrough flowed the shell through side. Sight the glasses shell side. were Sight used glassesto inspect were the state used of to inspect the working state of thefluid working after it had fluid passed after through it had passedthe evaporators. through the evaporators.

Figure 3. 3. Shell-Shell- and and tube-type tube-type heat heat exchangers exchangers used usedfor the for HT the and HT LT and evaporators. LT evaporators.

InIn the the previous previous experiment, experiment, the the thermal thermal efficiency efficiency of the of ORC the ORCpart was part too was low, too due low, due to theto the difficulty difficulty in in optimizing optimizing thethe operatingoperating conditions conditions of of the the ORC ORC system, system, owing owing to a to a large mismatchlarge mismatch between between the the capacity capacity of of the the scroll scroll expander expanderand and thethe lowlow capacity of of the the burner inburner the laboratory.in the laboratory. Therefore, Therefore, an additionalan additional working working fluidfluid (R134a), heated heated by by steam steam from thefrom boiler the boiler to the to the same same state state as as the the working working fluidfluid from from the the waste waste heat heat recovery recovery of the of the flue gas,flue gas, was was added added to theto the inlet inlet of of the the expander. expander.In In FigureFigure 22,, thethe red line line depicts depicts the the flow flow of the steamof the steam circulation circulation for this for purpose.this purpose. However, However, the the net net power power from from the the waste waste heat heat recovery Energies 2021, 14, 4328 recovery of the flue gas in the increased power was calculated by using the5 ofportion 14 of the of the flue gas in the increased power was calculated by using the portion of the working fluid (flow meter 1 and flow meter 3) flowing into the expander. This method to optimize workingthe operating fluid (flow conditions meter 1 and flow of the meter ORC 3) flowing is not into practical the expan forder. full-scale This method ORC to systems in field

optimizeapplications. the operating However, conditions in a of lab-scale the ORC ORCis not system,practical for it isfull-scale temporary, ORC systems and can be used to solve inthe field mismatch applications. problem However, between in a lab-scale the ORC obtainable system, it components is temporary, and and can demonstrate be used the potential toperformance solve the mismatch of the problem system. between Figure the 4obtainable shows thecomponents photograph and demonstrate of the experimental the setup of potential performance of the system. Figure 4 shows the photograph of the experimental setupthe WHWRof the WHWR system. system.

FigureFigure 4. Photograph 4. Photograph of the experimental of the experimental setup of the setup WHWR of system. the WHWR system.

4. Experimental Procedure After completely heating the steam boiler, the temperature and mass flow rate of the flue gas flowing into the WHWR system were maintained at constant values. To begin with, the mass flow rate of the working fluid and the speed of the expander of the ORC part were optimized by adjusting the speed of the HP pump and setting the load bank to maximize the power output of the ORC unit from the given waste heat source. Then, to maximize the water recovery efficiency, the mass flow rate of the working fluid in the pumped HPC unit was increased stepwise; however, it was limited to maintaining the state of the working fluid after the LT evaporator was close to the saturated gas state, and avoiding unnecessary circulation. During the experiment, the temperature, pressure, and flow rate data from the installed sensors were used to calculate the thermodynamic properties of the working fluid, using NIST-REFPROP software [19], and subsequently used to analyze the experimental data. The mechanical power output (𝑊) was measured by the torque (T) and rotation speed (N) of the expander from Equation (1). The isentropic efficiencies of the expander and the pump are expressed in Equations (2) and (3), respectively. The pump work was not measured because the low pump efficiency was not well matched in the lab-scale ORC system, which severely distorts the potential efficiency of the full-scale ORC system in the application; however, it was estimated by the assumed isentropic efficiency of the pump to be 0.7, by considering the general ORC pump efficiency between 0.65 and 0.85 [20], using Equation (3).

𝑊 =2𝜋𝑁𝑇 (1)

𝑊 𝜂 = (2) 𝑚 (ℎ −ℎ,)

𝑚 (ℎ, −ℎ) 𝜂 = (3) 𝑊

Energies 2021, 14, 4328 5 of 13

4. Experimental Procedure After completely heating the steam boiler, the temperature and mass flow rate of the flue gas flowing into the WHWR system were maintained at constant values. To begin with, the mass flow rate of the working fluid and the speed of the expander of the ORC part were optimized by adjusting the speed of the HP pump and setting the load bank to maximize the power output of the ORC unit from the given waste heat source. Then, to maximize the water recovery efficiency, the mass flow rate of the working fluid in the pumped HPC unit was increased stepwise; however, it was limited to maintaining the state of the working fluid after the LT evaporator was close to the saturated gas state, and avoiding unnecessary circulation. During the experiment, the temperature, pressure, and flow rate data from the installed sensors were used to calculate the thermodynamic properties of the working fluid, using NIST-REFPROP software [19], and subsequently used to analyze the experimental data. . The mechanical power output (We) was measured by the torque (T) and rotation speed (N) of the expander from Equation (1). The isentropic efficiencies of the expander and the pump are expressed in Equations (2) and (3), respectively. The pump work was not measured because the low pump efficiency was not well matched in the lab-scale ORC system, which severely distorts the potential efficiency of the full-scale ORC system in the application; however, it was estimated by the assumed isentropic efficiency of the pump to be 0.7, by considering the general ORC pump efficiency between 0.65 and 0.85 [20], using Equation (3). . We = 2πNT (1) . We ηe = . (2) m f 1(h7 − h7,s) . m f 1(h3,s − h1) ηp = . (3) W p . where h, s, and m f are the specific enthalpy, exit state for the isentropic process, and mass flow rate of the working fluid, respectively. The net power output of the system was calculated by subtracting the pump work input from the work output of the expander, as expressed in Equation (4).

. . . Wnet = We − W p (4)

The thermal efﬁciency of the ORC was calculated as the ratio of the net power output to the heat recovery of the ORC, as expressed in Equation (5).

. Wnet ηth = . (5) m f 1(h5 − h3)

The water recovery efﬁciency was calculated as the ratio of the water vapor pressure difference after passing through the WHWR system and the initial water vapor pressure before passing through the WHWR system, as expressed in Equation (6).

Pw,in − Pw,out RH(in)Psat@T,in − RH(out)Psat@T,out ηWR = = (6) Pw,in RH(in)Psat@T,in

where Pw, RH, and Psat@T are the water vapor pressure, relative humidity, and saturation water vapor pressure at a given temperature, respectively.

5. Results and Discussion Although the ORC unit was optimized for maximum power from the given ﬂue gas during the experiment, the thermal efﬁciency of the ORC unit was substantially lower than Energies 2021, 14, 4328 6 of 13

expected. The operating conditions of the ORC were investigated in order to improve the thermal efficiency of the ORC unit. Figure5 (solid lines) depicts the operating conditions of the ORC and flue gas in the temperature and entropy diagrams. The high-pressure side was approximately 11 bar, with an evaporation temperature of 43 ◦C, while the low-pressure side was 6 bar, with a condensation temperature of 22 ◦C. Furthermore, the isentropic efficiency of the expander was low, because the operating pressure ratio of 1.8 was too low Energies 2021, 14, 4328 for the design pressure ratio of 3.5. To improve the thermal efficiency of the ORC unit, the 7 of 14 pressure of the high-pressure side must be increased for higher evaporation temperatures, as indicated by the red dashed lines in Figure5.

FigureFigure 5.5. ImprovementImprovement of of thermal thermal efficiency efficiency of the of ORC the ORC unit. unit. However, optimizing the operating conditions of the ORC unit is challenging, because the capacityThe amount of the of scroll additional expander wo (model:rking fluid E15H022A-SH) heated by builtthe steam by Air from Squared, the boiler Inc. is into the expander1–2 kW, despite was increased the fact that stepwise; the maximum it was power then from decreased the given in fluereverse gas was in order 200 W. to It check is the reproducibilitydifficult to match of the the capacity experiment. of the scroll As shown expander in withFigure the 6, capacity increasing of the the burner amount in the of addi- tionallaboratory. working As a result,fluid into to increase the expander the pressure increa onsed the the high-pressure expander side,inlet anpressure additional from 10 to 22working bar, while fluid (R134a)the expander heated byinlet the temperature steam from the slightly boiler todecreased the same statefrom as 107 the °C working to 80 °C, and remainedfluid from thealmost waste constant. heat recovery The high of the expander flue gas was inlet added temperature to the inlet in of thethe expander.early stage was Despite the fact that the power of the expander was increased due to the additional heat attributed to the thermal inertia of the heat exchangers, because the stored thermal energy input, the net power from the waste heat recovery of the flue gas was calculated using the ofportion the heat of the exchanger working fluidcould that be wasused allowed to heat tothe pass working into the fluid. expander. Figure 7 shows that increasing theThe expander amount of inlet additional pressure working increased fluid the heated expander by the speed steam from from the870 boilerto 2840 into rpm, and thethe expander waspower increased from 204 stepwise; to 1790 it wasW (with then decreaseda maximum in reverse uncertainty in order of to0.67%). check the reproducibility of the experiment. As shown in Figure6, increasing the amount of

additional working ﬂuid into the expander increased the expander24 inlet pressure from 10 to 22 bar, while the expander inlet temperature slightly decreased from 107 ◦C to 80 ◦C, 100 and remained almost constant. The high expander inlet temperature22 in the early stage was attributed to the thermal inertia of the heat exchangers, because the stored thermal

energyC) of80 the heat exchanger could be used to heat the working20 ﬂuid. Figure7 shows that o increasing the expander Expander Inlet inlet Temp. pressure increased the expander speed from 870 to 2840 rpm, Expander Inlet Pressure 18 and the60 expander power from 204 to 1790 W (with a maximum uncertainty of 0.67%). 16

40 14

20 12 Expander Inlet Temp. ( Temp. Inlet Expander Expander Inlet Pressure (bar) Pressure Inlet Expander 10 0 0 500 1000 1500 2000 2500 3000 Time (sec)

Figure 6. Expander inlet pressure and temperature.

Energies 2021, 14, 4328 7 of 14

Figure 5. Improvement of thermal efficiency of the ORC unit.

The amount of additional working fluid heated by the steam from the boiler into the expander was increased stepwise; it was then decreased in reverse in order to check the reproducibility of the experiment. As shown in Figure 6, increasing the amount of additional working fluid into the expander increased the expander inlet pressure from 10 to 22 bar, while the expander inlet temperature slightly decreased from 107 °C to 80 °C, and remained almost constant. The high expander inlet temperature in the early stage was attributed to the thermal inertia of the heat exchangers, because the stored thermal energy of the heat exchanger could be used to heat the working fluid. Figure 7 shows that increas- Energies 2021, 14, 4328 ing the expander inlet pressure increased the expander speed from 870 to 2840 rpm, and7 of 13 the expander power from 204 to 1790 W (with a maximum uncertainty of 0.67%).

100 22

C) 80 20 o Expander Inlet Temp. Expander Inlet Pressure 18 60 16

40 14

20 12 Expander Inlet Temp. ( Temp. Inlet Expander Expander Inlet Pressure (bar) Pressure Inlet Expander 10 0 0 500 1000 1500 2000 2500 3000 Energies 2021, 14, 4328 Time (sec) 8 of 14

FigureFigure 6. Expander 6. Expander inlet inlet pressu pressurere and and temperature. temperature.

2.0 3500

1.8 3000 1.6 Expander Power Expander Speed 1.4 2500

1.2 2000 1.0 1500 0.8

0.6 1000 Expander Power (kW) Power Expander 0.4 (rpm) Speed Expander 500 0.2

0.0 0 0 500 1000 1500 2000 2500 3000 Time (sec)

FigureFigure 7. Expander 7. Expander speed speed and and power. power.

TheThe test test results results are are shown shown for for a relatively a relatively short short operating operating time time of of approximately approximately 3000 3000 s s (1(1 h) h) under under different different operating conditions.conditions. TheThe performance performance of of the the ORC ORC and and water water recovery re- system was dependent on the temperature of the cooling water from the cooling tower, covery system was dependent on the temperature of the cooling water from the cooling which fluctuated during the day. Therefore, to keep the temperature of the cooling water tower, which fluctuated during the day. Therefore, to keep the temperature of the cooling as cold as possible in order to ensure the system’s high performance, the overall operating water as cold as possible in order to ensure the system’s high performance, the overall time to achieve the optimal conditions of the system was kept short during the start of the operating time to achieve the optimal conditions of the system was kept short during the day. All variables at each operating point were measured stepwise until they stabilized. The start of the day. All variables at each operating point were measured stepwise until they amount of additional working fluid heated by the steam into the expander was decreased stabilized. The amount of additional working fluid heated by the steam into the expander in reverse in a stepwise manner, in order to check the repeatability of each condition. was decreased in reverse in a stepwise manner, in order to check the repeatability of each Figure8 shows that with the increase and decrease in the evaporation (expander inlet) condition. pressure, between 14 and 22 bar (between 750 s and 3000 s), the exhaust gas temperature Figure 8 shows that with the increase and decrease in the evaporation (expander in- after passing through the HT evaporator of the ORC unit (Texh2) with an inlet temperature let) pressure, between 14 and 22 bar (between 750 s and 3000 s), the exhaust gas tempera- (Texh1) of 143 ◦C increased and decreased symmetrically between 64 ◦C and 81 ◦C. In turegeneral, after passing if the evaporation through thepressure HT evaporator of the ORC of the for ORC waste unit heat (Texh2) recovery with from an inlet the fluetem- gas peratureincreases, (Texh1) the of outlet 143 °C temperature increased and of the decreased flue gas symmetrically from the ORC between also increases, 64 °C and and 81 the °C.quantity In general, of heatif the recovery evaporation decreases pressure [21,22 of]. the ORC for waste heat recovery from the flue gas increases, the outlet temperature of the flue gas from the ORC also increases, and the quantity of heat recovery decreases [21,22].

140

Texh1 120 Texh2

C) Texh3 o 100

60 Exhaust Temp. (

0 500 1000 1500 2000 2500 3000 Time (sec)

Figure 8. Temperature decrease of the exhaust gas.

Energies 2021, 14, 4328 8 of 14

2.0 3500

1.8 3000 1.6 Expander Power Expander Speed 1.4 2500

1.2 2000 1.0 1500 0.8

0.6 1000 Expander Power (kW) Power Expander 0.4 (rpm) Speed Expander 500 0.2

0.0 0 0 500 1000 1500 2000 2500 3000 Time (sec)

Figure 7. Expander speed and power.

The test results are shown for a relatively short operating time of approximately 3000 s (1 h) under different operating conditions. The performance of the ORC and water recovery system was dependent on the temperature of the cooling water from the cooling tower, which fluctuated during the day. Therefore, to keep the temperature of the cooling water as cold as possible in order to ensure the system’s high performance, the overall operating time to achieve the optimal conditions of the system was kept short during the start of the day. All variables at each operating point were measured stepwise until they stabilized. The amount of additional working fluid heated by the steam into the expander was decreased in reverse in a stepwise manner, in order to check the repeatability of each condition. Figure 8 shows that with the increase and decrease in the evaporation (expander inlet) pressure, between 14 and 22 bar (between 750 s and 3000 s), the exhaust gas temperature after passing through the HT evaporator of the ORC unit (Texh2) with an inlet temperature (Texh1) of 143 °C increased and decreased symmetrically between 64 °C and 81 Energies 2021, 14, 4328 °C. In general, if the evaporation pressure of the ORC for waste heat recovery from the 8 of 13 flue gas increases, the outlet temperature of the flue gas from the ORC also increases, and the quantity of heat recovery decreases [21,22].

140

Texh1 120 Texh2

C) Texh3 o 100

60 Exhaust Temp. (

0 500 1000 1500 2000 2500 3000 Time (sec)

FigureFigure 8. Temperature 8. Temperature decrease decrease of the exhaust of the gas. exhaust gas.

To minimize the effect of the stored thermal energy of the heat exchanger during the early stages of the experiment, and check the reproducibility of the experiment, the experimental results, as shown in Figure9, were selected for further analyses in the range of 750 s to 3000 s. Figure 10 shows the operating conditions of the system, with an increasing evaporation pressure in the temperature and entropy diagrams. The outlet temperature of the exhaust gas from the ORC unit (Texh2) linearly increased with an increase in the evaporation pressure between 14 and 22 bar (from conditions 1 to 5). Figure 11 shows the relationship between Texh2 and the evaporation temperature of the ORC (Tevap_orc) with an increasing evaporation pressure, in which Texh2 increased and decreased reversibly with the evaporation temperature of the ORC, along with a variation in the evaporation pressure. Figure 11 clearly shows the decrease in the heat input from the HT evaporator with an increase in the evaporation pressure of the ORC. Figure 12 shows the analyses of the expander inlet temperature with an increase in the evaporation pressure between 14 and 22 bar. If the heat input into the ORC is constant, the expander inlet temperature (Te_Qcst) increases with an increase in the evaporation pressure. However, if the decreased heat input from the HT evaporator with an increase in the evaporation pressure of the ORC is applied with a constant mass flow of the ORC Energies 2021, 14, 4328 9 of 14 working fluid, the expander inlet temperature (Te_MFcst) decreases with an increase in the evaporation pressure. In the experiment, when the pump speed was kept constant, the measuredTo minimize massthe effect flow of the rate stored of thethermal ORC energy working of the heat fluid exchanger slightly during decreased the with an increase earlyin stages the of evaporation the experiment, pressure. and check the The reproducibility calculated of expanderthe experiment, inlet the temperatureexper- (Te_cal) with imentalthe results, decrease as shown in the inheat Figure input 9, were from selected the for HT further evaporator analyses was in the plotted range of between Te_Qcst and 750 sTe_MFcst. to 3000 s. Figure In addition, 10 shows the the operating actual expander conditions of inlet the system, temperature with an withincreasing the increase and decrease evaporation pressure in the temperature and entropy diagrams. The outlet temperature in the evaporation pressure (Tact1 and Tact2, respectively) was plotted with the calculated of the exhaust gas from the ORC unit (Texh2) linearly increased with an increase in the evaporationexpander pressure inlet between temperature 14 and 22 (Te_cal), bar (from in conditions which there 1 to 5). were some deviations observed due to the measurement errors (temperature, pressure, mass flow rate, etc.) and the unsteady state.

100 Expander Inlet Temp. Expander Inlet Pressure 22

C) 80 20 o

18 60 16

40 14

20 12 Expander Inlet ( Temp. Expander Inlet Pressure (bar) 10 0 1000 1500 2000 2500 Time (sec)

FigureFigure 9. Expander 9. Expander inlet pressure inlet and pressure temperat andure over temperature selected time over interval. selected time interval.

Figure 10. Increase in the evaporation pressure in temperature–entropy diagram.

Figure 11 shows the relationship between Texh2 and the evaporation temperature of the ORC (Tevap_orc) with an increasing evaporation pressure, in which Texh2 increased and decreased reversibly with the evaporation temperature of the ORC, along with a variation in the evaporation pressure. Figure 11 clearly shows the decrease in the heat input from the HT evaporator with an increase in the evaporation pressure of the ORC. Figure 12 shows the analyses of the expander inlet temperature with an increase in the evaporation pressure between 14 and 22 bar. If the heat input into the ORC is constant, the expander inlet temperature (Te_Qcst) increases with an increase in the evaporation pressure. However, if the decreased heat input from the HT evaporator with an increase

Energies 2021, 14, 4328 9 of 14

To minimize the effect of the stored thermal energy of the heat exchanger during the early stages of the experiment, and check the reproducibility of the experiment, the experimental results, as shown in Figure 9, were selected for further analyses in the range of 750 s to 3000 s. Figure 10 shows the operating conditions of the system, with an increasing evaporation pressure in the temperature and entropy diagrams. The outlet temperature of the exhaust gas from the ORC unit (Texh2) linearly increased with an increase in the evaporation pressure between 14 and 22 bar (from conditions 1 to 5).

100 Expander Inlet Temp. Expander Inlet Pressure 22

C) 80 20 o

18 60 16

40 14

20 12 Expander Inlet ( Temp. Expander Inlet Pressure (bar) Energies 2021, 14, 4328 10 10 of 14 Energies 2021, 14, 4328 10 of 14 0 1000 1500 2000 2500 Time (sec) Energies 2021, 14, 4328 9 of 13

inin the the evaporation evaporation pressure pressure of of the the ORC ORC is is applied applied with with a a constant constant mass mass flow flow of of the the ORC ORC workingFigureworking 9. Expanderfluid, fluid, the the inlet expander expander pressure inlet inlet and temperature temperattemperatureure over (Te_MFcst) (Te_MFcst) selected time decreases decreases interval. with with an an increase increase in in thethe evaporation evaporation pressure. pressure. In In the the experiment, experiment, when when the the pump pump speed speed was was kept kept constant, constant, the the measuredmeasured mass mass flow flow rate rate of of the the ORC ORC working working fl fluiduid slightly slightly decreased decreased with with an an increase increase in in thethe evaporation evaporation pressure. pressure. The The calculated calculated expand expanderer inlet inlet temperature temperature (Te_cal) (Te_cal) with with the the de- de- creasecrease inin thethe heatheat inputinput fromfrom thethe HTHT evaporatorevaporator waswas plottedplotted betweenbetween Te_QcstTe_Qcst andand Te_MFcst.Te_MFcst. InIn addition,addition, thethe actualactual expanderexpander inletinlet temperaturetemperature withwith thethe increaseincrease andand de-de- creasecrease inin thethe evaporationevaporation pressurepressure (Tact1(Tact1 andand Tact2,Tact2, respectively)respectively) waswas plottedplotted withwith thethe calculatedcalculated expander expander inlet inlet temperature temperature (Te_cal) (Te_cal), ,in in which which there there were were some some deviations deviations ob- ob- servedserved duedue toto thethe measurementmeasurement errorserrors (tempe (temperature,rature, pressure,pressure, massmass flow flow rate,rate, etc.) etc.) and and thethe unsteady unsteady state. state. AsAs shown shown in in Figure Figure 13, 13, with with an an increase increase in in the the evaporation evaporation pressure pressure between between 10 10 and and 2222 bar, bar, the the thermal thermal efficiency efficiency increased increased co continuouslyntinuously from from 3% 3% to to 10% 10% (with (with a a maximum maximum uncertaintyuncertainty of of ±0.15%); ±0.15%); however, however, the the heat heat in inputput from from the the HT HT evaporator evaporator of of the the ORC ORC unit unit decreaseddecreased from from 3.5 3.5 kW kW to to 2.1 2.1 kW kW (with (with a a maxi maximummum uncertainty uncertainty of of ±0.05 ±0.05 kW), kW), as as shown shown in in FigureFigure 14. 14. The The abnormal abnormal temperature temperature decrease decrease of of Texh2 Texh2 during during the the early early stage stage of of the the ex- ex- perimentperiment between between 0 0 and and 750 750 s s was was attributed attributed to to the the thermal thermal inertia inertia of of the the heat heat exchangers, exchangers, becausebecause theirtheir storedstored thermalthermal energyenergy couldcould bebe usedused toto heatheat thethe workworkinging fluid.fluid. Thus,Thus, thethe heatheat exchangersexchangers couldcould bebe cooled,cooled, becausebecause thethe abnormalabnormal temperaturetemperature decreasedecrease ofof Texh2Texh2 duringduring the the early early stages stages of of the the experiment experiment in in the the range range of of 0 0 to to 750 750 s s did did not not increase increase the the heatFigureFigureheat input input 10. 10. Increase Increasefrom from the the in in theHT theHT evaporation evaporationevaporatorevaporator pressure pressureof of the the ORC, ORC,in in temperature–entropy temperature–entropy as as shown shown in in Figure Figure diagram. diagram. 14. 14.

Figure 11 shows the relationship between Texh2 and the evaporation temperature of 80 the ORC80 (Tevap_orc) with an increasing evaporation pressure, in which Texh2 increased

and decreased75 reversibly with the evaporation temperature of the ORC, along with a var- 75 iation in the evaporation pressure. Figure 11 clearly shows the decrease in the heat input 70

fromC) the70 HT evaporator Texh2 with an increase in the evaporation pressure of the ORC. o

C) Texh2 o Figure 12 shows the analyses of the expander inlet temperature with an increase in 65 65 the evaporation pressure between 14 and Tevap_orc 22 bar. If the heat input into the ORC is constant, Tevap_orc the expander60 inlet temperature (Te_Qcst) increases with an increase in the evaporation 60

pressure. ( Temperature However, if the decreased heat input from the HT evaporator with an increase Temperature ( Temperature 55 55

50 50

12 13 14 15 16 17 18 19 20 21 22 23 12 13 14 15 16 17 18 19 20 21 22 23 Evaporation Pressure (bar) Evaporation Pressure (bar) FigureFigure 11. 11. Exhaust Exhaust outlet outlet temperature temperature with with va variationvariationriation in in evaporation evaporation pressure. pressure.

100 100

95 Te_Qcst 95 Te_Qcst C) o

C) 90 o 90

85 Te_act2 85 Te_act2

80 Te_act1 80 Te_act1

75 75 Te_cal Expander InletTemp. Expander ( Te_MFcst Te_cal Expander InletTemp. Expander ( Te_MFcst 70 70

13 14 15 16 17 18 19 20 21 22 23 13 14 15 16 17 18 19 20 21 22 23 Evaporation Pressure (bar) Evaporation Pressure (bar) Figure 12. Expander inlet temperature with variation in evaporation pressure. Figure 12. Expander inlet temperature with variation in evaporation pressure.

As shown in Figure 13, with an increase in the evaporation pressure between 10 and 22 bar, the thermal efﬁciency increased continuously from 3% to 10% (with a maximum

Energies 2021, 14, 4328 10 of 13

uncertainty of ±0.15%); however, the heat input from the HT evaporator of the ORC unit decreased from 3.5 kW to 2.1 kW (with a maximum uncertainty of ±0.05 kW), as shown in Figure 14. The abnormal temperature decrease of Texh2 during the early stage of the experiment between 0 and 750 s was attributed to the thermal inertia of the heat exchangers, because their stored thermal energy could be used to heat the working ﬂuid. Thus, the heat Energies 2021, 14, 4328 11 of 14 Energies 2021, 14, 4328 exchangers could be cooled, because the abnormal temperature decrease of Texh211 duringof 14

the early stages of the experiment in the range of 0 to 750 s did not increase the heat input from the HT evaporator of the ORC, as shown in Figure 14.

12 80 12 80 Thermal Efficierncy 11 Thermal Efficierncy 11 Water Recovery Efficiency 70 10 Water Recovery Efficiency 70 10 9 60 9 60 8 8 50 7 50 7 6 40 6 40 5 5 30 4 30 4 3 20 3 20 Thermal Efficiency (%)

Thermal Efficiency (%) 2 2

10 Water Recovery Efficiency (%)

1 10 Water Recovery Efficiency (%) 1 0 0 0 0 0 500 1000 1500 2000 2500 3000 0 500 1000 1500 2000 2500 3000 Time (sec) Time (sec) Figure 13. Thermal efficiency of ORC and water recovery efficiency. FigureFigure 13. 13.ThermalThermal efficiency efﬁciency of ORC of ORC and and water water recovery recovery efficiency. efﬁciency.

5.0 0.30 5.0 0.30 Heat Recovery ORC 4.5 Heat Recovery ORC 4.5 Net Power ORC Net Power ORC 0.25 4.0 0.25 4.0 3.5 3.5 0.20 0.20 3.0 3.0 2.5 0.15 2.5 0.15 2.0 2.0 0.10 1.5 0.10 1.5

1.0 Net Power ORC (kW) Net Power ORC (kW)

Heat Recovery ORC (kW) 1.0 0.05

Heat Recovery ORC (kW) 0.05 0.5 0.5 0.0 0.00 0.0 0.00 0 500 1000 1500 2000 2500 3000 0 500 1000 1500 2000 2500 3000 Time (sec) Figure 14. Heat recovery and netTime power (sec) of ORC unit. Figure 14. Heat recovery and net power of ORC unit. Figure 14.Therefore, Heat recovery the netand powernet power of theof ORC ORC unit. unit increased from 115 to 220 W in the range of 10 to 18 bar (evaporation pressure), and then decreased to 200 W in the range of 18 to Therefore,Therefore, the the net net power power of ofthe the ORC ORC unit unit increased increased from from 115 115 to to220 220 W Win inthe the range range of of 10 to 2218 bar (evaporation (condensation pressure), pressure). and The then optimal decreased evaporation to 200 W in pressure the range for of the 18 maximumto 22 10 powerto 18 bar obtained (evaporation from the pressure), experimental and then results decreased was approximately to 200 W in 18 the bar. range Consequently, of 18 to 22 the barbar (condensation (condensation pressure). pressure). The The optimal optimal ev evaporationaporation pressure pressure for for the the maximum maximum power power obtainedmaximum from the power experimental output from results the givenwas approximately waste heat source 18 bar. can Consequently, be obtained by the incorporating max- obtainedboth the from heat the recovery experimental efficiency results (heat was input approximately from the given 18 bar. waste Consequently, heat source) andthe max- thermal imumimum power power output output from from the the given given waste waste heat heat source source can can be be obtained obtained by by incorporating incorporating both theefficiency heat recovery (net power efficiency from the (heat heat input input) from of thethe ORCgiven [21 waste,22]. heat source) and ther- both theWith heat anrecovery increase efficiency and decrease (heat in input the evaporation from the given (expander waste heat inlet) source) pressure, and as ther- shown malmal efficiency efficiency (net (net power power from from the the heat heat input) input) of ofthe the ORC ORC [21,22]. [21,22]. Within Figure an increase 15, the waterand decrease recovery in efficiency the evap decreasesoration (expander and increases inlet) from pressure, 58% to as 62%, shown because theWith temperature an increase of theand exhaust decrease gas in after the evap the LToration evaporator (expander of the inlet) water pressure, recovery as unit shown (Texh3) in Figure 15, the water recovery efficiency decreases and increases from 58% to 62%, be- in Figureincreased 15, andthe water decreased recovery in the efficiency range of 36decreases◦C to 38 and◦C, owingincreases to thefrom increase 58% to and 62%, decrease because the temperature of the exhaust gas after the LT evaporator of the water recovery causein thethe exhausttemperature gas afterof the the exhaust HT evaporator gas after the of theLT ORCevaporator (Texh2), of asthe shown water inrecovery Figure 8. unit (Texh3) increased and decreased in the range of 36 °C to 38 °C, owing to the increase unitThe (Texh3) relative increased inlet humidity and decreased at 143 ◦ inC the was rang 3.6%e of (the 36 partial°C to 38 pressure °C, owing ratio to ofthe water increase vapor and decrease in the exhaust gas after the HT evaporator of the ORC (Texh2), as shown in andbeing decrease 14.4%), in the and exhaust the relative gas after outlet the humidityHT evaporator at 37 ◦ofC the was ORC 90.0% (Texh2), (the partial as shown pressure in FigureFigure 8. The8. The relative relative inlet inlet humidity humidity at at143 143 °C °C was was 3.6% 3.6% (the (the partial partial pressure pressure ratio ratio of ofwater water vaporvapor being being 14.4%), 14.4%), and and the the relative relative outlet outlet humidity humidity at at37 37°C °C was was 90.0% 90.0% (the (the partial partial pres- pres- suresure ratio ratio of ofwater water vapor vapor being being 5.8%). 5.8%). Therefor Therefore, thee, the calculated calculated water water recovery recovery efficiency efficiency waswas 60.0% 60.0% (maximum (maximum uncertainty, uncertainty, ±6.6%). ±6.6%). The The flue flue gas, gas, with with an an initial initial temperature temperature of of143 143 °C,°C, was was cooled cooled to to79 79°C °C through through the the HT HT evaporator, evaporator, and and further further cooled cooled to to37 37°C °C using using an an LTLT evaporator. evaporator. However, However, the the heat heat recovery recovery through through the the HT HT and and LT LT evaporators evaporators was was 2.3 2.3

Energies 2021, 14, 4328 11 of 13

Energies 2021, 14, 4328 ratio of water vapor being 5.8%). Therefore, the calculated water recovery efﬁciency12 of 14 was 60.0% (maximum uncertainty, ±6.6%). The ﬂue gas, with an initial temperature of 143 ◦C, was cooled to 79 ◦C through the HT evaporator, and further cooled to 37 ◦C using an LT evaporator. However, the heat recovery through the HT and LT evaporators was 2.3 kW kW (max uncertainty of ±0.05 kW) and 18.2 kW (max uncertainty of ±0.12 kW), respec- (max uncertainty of ±0.05 kW) and 18.2 kW (max uncertainty of ±0.12 kW), respectively, tively, because the temperature drop was small through the LT evaporator; however, the because the temperature drop was small through the LT evaporator; however, the latent latent heat owing to water condensation accounted for the large heat recovery. heat owing to water condensation accounted for the large heat recovery.

0.9 4.0 3.8 Expander Efficinency 0.8 Pressure Ratio 3.6 3.4 3.2 0.7 3.0 2.8 0.6 2.6 2.4 0.5 Ratio Pressure 2.2 Expander Efficiency Efficiency Expander 2.0 0.4 1.8 1.6 0 500 1000 1500 2000 2500 3000 Time (sec)

FigureFigure 15. 15.PressurePressure ratio ratio and and expander expander efficiency. efﬁciency.

WithWith an anincrease increase in the in the evaporation evaporation (expander (expander inlet) inlet) pressure, pressure, the the thermal thermal efficiency efficiency of the ORC increased significantly, from 3% to 10%. This was partly due to the increase in of the ORC increased significantly, from 3% to 10%. This was partly due to the increase in the evaporation pressure of the ORC, and as shown in Figure 11, the expander efficiency the evaporation pressure of the ORC, and as shown in Figure 11, the expander efficiency increased significantly from 45% to 80% (max uncertainty of ±1.3%) with an increase in increased significantly from 45% to 80% (max uncertainty of ±1.3%) with an increase in the pressure ratio from 1.8 to 3.2, approaching the design pressure ratio of 3.5. the pressure ratio from 1.8 to 3.2, approaching the design pressure ratio of 3.5. 6. Conclusions 6. Conclusions Heat and water recovery from the flue gas by condensing the flue gas moisture in a Heat and water recovery from the flue gas by condensing the flue gas moisture in a thermal power plant can help to reduce CO2 emissions, water requirements, condensable thermalPM, andpower the plant occurrence can help of to white reduce plumes CO2 emissions, and smog water around requirements, the power condensable plant. After a PM,preliminary and the occurrence study, a WHWRof white system plumes composed and smog of around an ORC the and power cooling plant. cycles After that a pre- uses a liminarysingle study, working a WHWR fluid accompanied system composed by a phase of an change ORC and was cooling proposed cycles and that optimized uses a sin- for a gle600-MW working powerfluid accompanied plant with water by a vaporphase withchange a volume was proposed of 16% and in flue optimized gas at 150 for◦ C.a 600- MW powerIn this plant study, with a water lab-scale vapor system with fora volume heat and of 16% water in recoveryflue gas at from 150 the°C. flue gas of a steamIn this boiler study, was a built lab-scale and tested. system First, for heat the low and thermal water recovery efficiency from of the the ORC flue unit gas resulted of a steamfrom boiler the mismatch was built and between tested. the First, low the flow low rate thermal of the efficiency flue gas and of the the ORC large unit capacity resulted of the fromexpander the mismatch of the ORC between unit. the Therefore, low flow an rate additional of the flue working gas and fluid the (R134a) large capacity heated byof steamthe expanderfrom the of boilerthe ORC was unit. added Therefore, to the inlet an addi of thetional expander working in fluid order (R134a) to achieve heated high-efficiency by steam fromconditions the boiler of was the ORCadded unit, to the where inlet the of the power expander from the in wasteorder to heat achieve recovery high-efficiency of the flue gas conditionswas calculated of the ORC by the unit, portion where of the the power working from fluid the that waste passed heat intorecovery the expander. of the flue gas was calculatedWith an by increase the portion in the of evaporation the working (expander fluid that inlet)passed pressure into thefrom expander. 10 to 22 bar, the thermalWith an efficiency increase ofin thethe ORCevaporation increased (expander significantly inlet) frompressure 3% tofrom 10%. 10 Thisto 22 wasbar, partlythe thermaldue to efficiency the increase of the in theORC evaporation increased significantly pressure of thefrom ORC, 3% to and 10%. the This significant was partly increase due in to thethe increase expander in efficiency, the evaporation approaching pressure the of design the ORC, pressure and the ratio significant of the expander. increase However, in the expanderthe heat efficiency, input from approaching the HT evaporator the design of thepressure ORC unitratio decreasedof the expander. from 3.5 However, to 2.1 kW. the The heatoptimal input from evaporation the HT pressureevaporator for of the the maximum ORC unit power decreased from from the experimental3.5 to 2.1 kW. result The is optimalapproximately evaporation 18 bar,pressure because for the maximummaximum power outputfrom the from experimental the given wasteresult heatis approximatelysource can be 18 obtained bar, because by incorporating the maximum both power the heat output recovery from efficiency the given (heat waste input heat from sourcethe givencan be waste obtained heat by source) incorporating and thermal both the efficiency heat recovery (net power efficiency from (heat theheat input input) from of thethe given ORC. waste heat source) and thermal efficiency (net power from the heat input) of the ORC. At the optimized conditions, the flue gas with an initial temperature of 143 °C was cooled to 77 °C through the ORC evaporator, with a heat recovery of 2.3 kW, generating an additional power of 220 W, with a thermal efficiency of 9.4%. Subsequently, the flue

Energies 2021, 14, 4328 12 of 13

At the optimized conditions, the flue gas with an initial temperature of 143 ◦C was cooled to 77 ◦C through the ORC evaporator, with a heat recovery of 2.3 kW, generating an additional power of 220 W, with a thermal efficiency of 9.4%. Subsequently, the flue gas was further cooled to 37 ◦C through the pumped HPC evaporator, with a heat recovery of 18.2 kW, producing condensed water with a water recovery efficiency of 58%.

Author Contributions: All authors contributed to this study. Y.-M.K. designed the experimental apparatus and procedure. A.N. and S.S.M.S. performed the experiments. D.-G.S. and G.C. discussed the results and commented on the manuscript at all stages. All authors have read and agreed to the published version of the manuscript. Funding: This research was funded by Korea Institute of Machinery and Materials Research Fund, 2020 (NK225A). Institutional Review Board Statement: Not applicable. Informed Consent Statement: Not applicable. Data Availability Statement: Not applicable. Conﬂicts of Interest: The authors declare no conﬂict of interest.

Nomenclature

h Specific enthalpy: kJ/kg . m Mass flow rate, kg/s N Rotational speed, rev/s P Pressure, kPa RH Relative humidity, % s Specific entropy, kJ/kg·K T Torque, kNm . W Power, kW Special characters η Efficiency Subscripts e Expansion i Inlet net Net power o Outlet p Pump s Isentropic process th Thermal sat Saturation w Water vapor WR Water recovery

References 1. Bufi, E.A.; Camporeale, S.M.; Cinnella, P. Robust optimization of an Organic Rankine Cycle for heavy-duty engine waste heat recovery. Energy Procedia 2017, 129, 66–73. [CrossRef] 2. Carpenter, A.M. Water Conservation in Coal-Fired Power Plants; IEA Clean Coal Centre: London, UK, 2017. 3. Daal, L. How the power industry can contribute to reducing the global water shortage. Power Eng. Int. 2011, 19, 62–69. 4. Jeong, K.; Kessen, M.J.; Bilirgen, H.; Levy, E.K. Analytical modeling of water condensation in condensing heat exchanger. Int. J. Heat Mass Transf. 2010, 53, 2361–2368. [CrossRef] 5. Li, Y.; Yan, M.; Zhang, L.; Chen, G.; Cui, L.; Song, Z.; Chang, J.; Ma, C. Method of flash evaporation and condensation—heat pump for deep cooling of coal-fired power plant flue gas: Latent heat and water recovery. Appl. Energy 2016, 172, 107–117. [CrossRef] 6. Nielsen, R.F.; Haglind, F.; Larsen, U. Design and modeling of an advanced marine machinery system including waste heat recovery and removal of sulphur oxides. Energy Convers. Manag. 2014, 85, 687–693. [CrossRef] 7. Zukeran, A.; Ninomiya, K.; Ehara, Y.; Yasumoto, K.; Kawakami, H.; Inui, T. SOx and PM Removal Using Electrostatic Precipitator with Heat Exchanger for Marine Diesel. In Proceedings of the Annual Meeting of the Electrostatic Society of America, Cocoa Beach, FL, USA, 11–13 June 2013; p. 1. Energies 2021, 14, 4328 13 of 13

8. Jeong, K. Waste heat recovery and saving fresh water consumption in power plants. ASME Int. Mech. Eng. Congr. Expo. 2012, 45226, 751–756. 9. Feng, Y.; Li, Y.; Cui, L. Critical review of condensable particulate matter. Fuel 2018, 224, 801–813. [CrossRef] 10. Qi, Z.; Li, J.; Wu, D.; Xie, W.; Li, X.; Liu, C. Particulate matter emission characteristics and removal efficiencies of a low-low temperature electrostatic precipitator. Energy Fuels 2017, 31, 1741–1746. [CrossRef] 11. Shuangchen, M.; Jin, C.; Kunling, J.; Lan, M.; Sijie, Z.; Kai, W. Environmental influence and countermeasures for high humidity flue gas discharging from power plants. Renew. Sustain. Energy Rev. 2017, 73, 225–235. [CrossRef] 12. Malavolta, M.; Beyene, A.; Venturini, M. Experimental implementation of a micro-scale ORC-based CHP energy system for domestic applications. ASME Int. Mech. Eng. Congr. Expo. 2010, 44298, 323–332. 13. Lecompte, S.; Gusev, S.; Vanslambrouck, B.; Paepe, M.D. Experimental results of a small-scale organic Rankine cycle: Steady state identification and application to off-design model validation. Appl. Energy 2018, 226, 82–106. [CrossRef] 14. Shao, L.; Ma, X.; Wei, X.; Hou, Z.; Meng, X. Design and experimental study of a small-sized organic Rankine cycle system under various cooling conditions. Energy 2017, 130, 236–245. [CrossRef] 15. Bianchi, M.; Branchini, L.; Casari, N.; Pascale, A.D.; Melino, F.; Ottaviano, S.; Pinelli, M.; Spina, P.R.; Suman, A. Experimental analysis of a micro-ORC driven by piston expander for low-grade heat recovery. Appl. Therm. Eng. 2019, 148, 1278–1291. [CrossRef] 16. Yang, S.C.; Hung, T.C.; Feng, Y.Q.; Wu, C.J.; Wong, K.W.; Huang, K.C. Experimental investigation on a 3 kW organic Rankine cycle for low-grade waste heat under different operation parameters. Appl. Therm. Eng. 2017, 113, 756–764. [CrossRef] 17. Rosset, K.; Pajot, O.; Schiffmann, J. Experimental investigation of a small-scale organic Rankine cycle turbo-generator supported on gas-lubricated bearings. ASME J. Eng. Gas. Turbines Power. 2021, 143, 051015. [CrossRef] 18. Shamsi, S.S.M.; Negash, A.A.; Cho, G.B.; Kim, Y.M. Waste heat and water recovery system optimization for flue gas in thermal power plants. Sustainability 2019, 11, 1881. [CrossRef] 19. Lemmon, E.W.; Mclinden, M.O.; Huber, M.L. N.I.S.T. Standard Reference Database 23; Version 10.0, Reference Fluid Thermodynamic and Transport Properties-REFPROP; National Institute of Standards and Technology (NIST): Boulder, CO, USA, 2018; Volume 23. 20. Borsukiewicz-Gozdur, A. Pumping work in the organic Rankine cycle. Appl. Therm. Eng. 2013, 51, 781–786. [CrossRef] 21. Kim, Y.M.; Sohn, J.L.; Yoon, E.S. Supercritical CO2 Rankine cycles for waste heat recovery from gas turbine. Energy 2017, 118, 893–905. [CrossRef] 22. Negash, A.; Kim, Y.M.; Shin, D.G.; Cho, G.B. Optimization of organic Rankine cycle used for waste heat recovery of construction equipment engine with additional waste heat of hydraulic oil cooler. Energy 2018, 143, 797–811. [CrossRef]