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Applied Thermal Engineering

journal homepage: www.elsevier.com/locate/apthermeng

Concept design of supercritical CO2 cycle driven by pressurized fluidizedbed combustion (PFBC) boiler

Enhui Suna, Han Hua, Hangning Lia, Jinliang Xua,b,⁎, Guohua Liua

a Beijing Key Laboratory of Multiphase Flow and Heat Transfer for Low Grade Energy Utilization, North China Electric Power University, Beijing 102206, Chinab Key Laboratory of Condition Monitoring and Control for Power Plant Equipment of Ministry of Education, North China Electric Power University, Beijing 102206, China

H I G H L I G H T S

• Two new systems are proposed to integrate PFBC boiler with S-CO2 cycle.

• System A has better performance than system B.

• Power generation efficiencies increase with increase of combustion pressures.

• Bubbling bed PFBC boiler and compact heat exchangers are recommended.

• Flue gas recirculation is suggested to decrease thermal load fraction in furnace.

A R T I C L E I N F O

Keywords:S-CO2 cyclePressurized fluidized bed combustion (PFBC)ThermodynamicsGas turbineHeat transfer

A B S T R A C T

Boiler is key to influence supercritical carbon dioxide (S-CO2) cycle performance. Here, S-CO2 cycle driven bypressurized fluidized bed combustion (PFBC) boiler is investigated. The proposed systems A and B include re-compression cycle with reheating (RC+RH), air compression-gas turbine (AC-GT), simple Brayton cycle (SRC),and RC+RH, AC-GT, respectively. A simulating code is developed for system analysis. We show that system Ahas higher efficiencies than system B. Larger flue gas pressure drop in PFBC boiler apparently worsens systemperformance. Effect of combustion pressures of PFBC boiler (Pfur) is thoroughly analyzed. Energy distributions inthe system is affected by Pfur. With Pfur increase from 1.0 MPa to 3.5 MPa, boiler efficiencies are increased due tothe decreased flue gas heat discharged to environment. Power generation efficiencies are increased but approachconstant value. To overcome the challenge of heating surface arrangement in boiler, bubbling PFBC boiler isrecommended. Compact heat exchangers are suggested in the top “pure” flue gas region. To further make iteasier for heating surface arrangement, excess air coefficient is suggested to be increased to decrease thermalload fraction in furnace αfurnace. Flue gas recirculation is a better way to decrease αfurnace, which is applicable forsystem B.

1. Introduction

Supercritical carbon dioxide (S-CO2) cycle can be driven by variousheat sources such as nuclear energy [1–3], solar energy [4–6], wasteheat [7–9], and fossil energy [10–12]. Compared with supercriticalwater-steam Rankine cycle, the benefits of S-CO2 cycle are as follows:(1) S-CO2 cycle has higher efficiency at moderate/higher main vaportemperatures [7,13,14]. (2) The system operation above the criticalpressure ensures the compact design of turbines and avoids the bladeerosion [15]. (3) Only two recuperator heat exchangers are involved inS-CO2 cycle, which are simpler compared with water-steam Rankine

cycle [14,16]. (4) The near-critical pressure operation of S-CO2 com-pressors reduces compression work, while heat transfer deterioration ofwater-cooling wall may occur when crossing the critical point forwater-steam Rankine cycle [17]. Thus, S-CO2 cycle is recognized as anefficient, compact and simple system for power generation.

Due to high demand on efficient and clean utilization of fossil en-ergy, S-CO2 cycle is a better choice to be driven by fossil energy. Thereare many types of coal-fired boilers such as pulverized coal fired boiler,fluidized bed boiler [18], pressurized boiler etc. The available studiesfocused on pulverized coal fired boiler to drive S-CO2 cycle, includingthermodynamics analysis [10–12,19,20], boiler design [21,22] and

https://doi.org/10.1016/j.applthermaleng.2019.114756Received 15 March 2019; Received in revised form 28 November 2019; Accepted 1 December 2019

⁎ Corresponding author at: Beijing Key Laboratory of Multiphase Flow and Heat Transfer for Low Grade Energy Utilization, North China Electric Power University,Beijing 102206, China.

E-mail address: xjl@ncepu.edu.cn (J. Xu).

Applied Thermal Engineering 166 (2020) 114756

Available online 03 December 20191359-4311/ © 2019 Elsevier Ltd. All rights reserved.

T

component performance [23,24]. In this paper, S-CO2 cycle driven bypressurized fluidized bed combustion (PFBC) boiler is investigated,realizing compacter system design compared with that driven by at-mospheric pressure combustion boiler.

The research and development of PFBC boiler has lasted more thanhalf century. The concept was initially proposed in 1967. Since then,many organizations, such as British Coal Corporation (formerly NCB)[25], US Department of Energy (DOE) [26], ABB Carbon [27], Ahlstrom[28], Kyushu Electric Power Company [29], have made great pro-gresses on PFBC boiler.

Compared with atmospheric pressure combustion boiler, PFBCboiler offers higher power generation efficiency [30,31], compacterdesign [30–32], decreased capital cost [33], higher efficiency in-bedsulfur retention [33,34], lower NOx emission [31,33] and flexible coalsupply [30,35,36]. Meanwhile, PFBC boiler has some disadvantagessuch as lower gas turbine inlet temperature [32], erosion of gas turbineblades and immersed heat exchanging surfaces [30,31], gas tempera-ture loss in gas cleaning sub-system [37]. Therefore, PFBC boiler needsfurther investigation.

Characterized by inlet temperature of gas turbine TGT,in, PFBC boilerhas evolved two generations development. The first generation attained~870 °C TGT, in, while the second generation reached much highervalue of ~1150 °C [30]. In a second-generation PFBC boiler, the pro-ducts of coal pyrolyzing or gasifying process are burned to form themixture of burned gas and flue gas to drive gas turbines. A combinedcycle driven by PFBC boiler includes a water-vapor Rankine cycle, anda gas turbine for power generation. Due to much higher inlet tem-perature of gas turbine, the gas turbine dominates the power genera-tion. Thus, the system driven by PFBC boiler is comparative to in-tegrated gasification combined cycle (IGCC) [38] or direct S-CO2 cycle[39]. The second generation of PFBC boiler has not been matured andneeds additional research and development [30]. Thus, the first gen-eration of PFBC boiler with the ~870 °C inlet temperature of gas tur-bine is integrated with S-CO2 cycle in this paper.

The conventional utilization of PFBC boiler includes a water-steamRankine cycle and a gas turbine to form a complicated system [30,31].In such a combined cycle, replacing the water-steam Rankine cycle by a

S-CO2 cycle significantly simplifies the system. Few works have beendone on the S-CO2 cycle driven by PFBC boiler. For carbon capturepurpose, Johnson et al. [40] proposed a zero-emission power plantdriven by PFBC boiler. The system includes a water-steam Rankinecycle and a S-CO2 cycle, but the gas turbine is not necessary. McClunget al. [41] studied an indirect-S-CO2 cycle driven by PFBC boilerwithout gas turbine involved. Combustion takes place in a ~10 MPapressure environment, which is difficult to be fulfilled for practicalapplication.

S-CO2 cycle driven by PFBC boiler introduces new coupling issuebetween cycle and boiler. Indeed, PFBC boiler is a challenge technologythat needs further investigation. The gas cleaning issue will be com-mented in Section 2. In the system level, due to distinct nature of S-CO2

cycle, it is necessary to perform a thorough thermodynamics analysis.First, water-steam Rankine cycle operates some components in highpressures but the condenser in vacuum pressure. S-CO2 cycle is a gascycle with the whole system operating above the critical pressure. Thus,different layout of heat transfer components shall be arranged to adaptto this change. Second, a water-steam Rankine cycle can absorb flue gasenergy covering a wide temperature range from ~900 °C to 120 °C. It isknown that S-CO2 cycle is suitable for heat source with temperaturelarger than 500 °C [13]. How to extract the flue gas energy over theentire temperature range is a challenge issue that is being addressed inthis paper. The originality of this paper is summarized as follows: (i)Based on the cascade energy utilization principle, two systems areproposed, which have not been reported previously. (ii) In system A, atop recompression S-CO2 cycle, a gas turbine, and a bottom simplerecuperated S-CO2 cycle form a combined cycle, which is believed tohave higher power generation efficiency. In system B, a recompressionS-CO2 cycle and a gas turbine form a combined cycle, which has simplelayout. The efficiency of the system can be increased by elevatingcombustion pressures in the furnace. (iii) Due to distinct characteristicof S-CO2 cycle, new concept is proposed for the arrangement of heattransfer components.

The paper is organized as follows. Section 2 describes the cycle ofthe two systems. Section 3 describes the simulation methodology for theproposed systems. Section 4.1 analyzes the performance characteristics

Nomenclature

AC air compressorBH bottom heaterC compressorCHE compact heat exchangerGC additional flue gas compressorGS gas cleaning sub-systemGT gas turbineh specific enthalpyHTR high temperature recuperatorLTR low temperature recuperatorm mass flow rateMH main heaterP pressurePFBC pressurized fluidized bed combustionQ heat transfer loadQf low heat value of coalRC recompression cycleRH reheaterSH superheaterS-CO2 supercritical carbon dioxideSRC simple recuperated S-CO2 cycleT turbineT temperatureW work

Subscripts

1, 2, 3… state points of top cycle1b, 2b, 3b… state points of bottom cyclear as receivedcom componentdaf dry and ash freee environmentex exhaustfg flue gasfur furnacein inletout outlets isentropic

Greek symbols

α energy distribution rateαair excess air coefficientβ proportion of recycled flue gas to total flue gasηb boiler efficiencyηe power efficiencyηg power generator efficiencyηp pipeline efficiencyηth thermal efficiencyΔP pressure drop

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of both systems under conventional combustion pressure. Section 4.2deals with the effect of combustion pressures on performances of thetwo systems. Section 4.3 presented two methods to properly arrangevarious heat transfer facilities in PFBC boiler for energy extraction offlue gas.

2. Two power generation systems driven by PFBC boiler

2.1. The system A

The two power generation systems satisfy the cascade energy uti-lization principle. System A is described first, including a top cycle(RC+RH), a bottom cycle (SRC), a PFBC boiler and an AC-GT sub-system (see Fig. 1). Here, RC, RH and SRC represent recompression S-CO2 cycle consisting of two recuperator heat exchangers, reheating, andsimple S-CO2 cycle including one recuperator heat exchanger only,respectively. AC means air-compressor and GT means gas turbine.

The whole system involves heat transfer from flue gas to S-CO2

cycle, regarded as an indirect S-CO2 cycle for heat-power conversion,

and an AC-GT for direct heat-power conversion with flue gas. The S-CO2

cycle is examined first. Fig. 1a shows the top cycle RC + RH. The S-CO2

stream at point 4 will be heated by a main-heater (MH) and then drivesturbine 1 (T1). The CO2 fluid at T1 outlet (point 4′) is reheated by areheater (RH) and then drives T2. After expansion, the CO2 fluid atpoint 6 consecutively flows through a high temperature recuperatorheat exchanger HTR and low temperature recuperator heat exchangerLTR, dissipating heat from hot side to cold side of CO2. The whole CO2

flow rate at point 8 is split into a main stream and an auxiliary stream.The former dissipates extra heat of the system to environment bycooler, and then pressurized by compressor 1 (C1). The auxiliary streamis directly pressurized by compressor 2 (C2). As shown in Fig. 1b, SRC isa single compression process, thus only one compressor (C3) is in-volved. It does not include reheating, thus only one heater BH and oneturbine (T3) are involved. Besides, SRC contains one recuperator heatexchanger LTR2 and one cooler 2.

The connected cycle concept was proposed by the present authorsrecently [11,42]. Such concept was extended in the present paper. Thecore idea is to connect the two cycles to simply the whole system

Fig. 1. The newly proposed system A (a: top cycle RC + RH, b: bottom cycle SRC, c: the connected cycle integrating with AC-GT sub-system and PFBC boiler).

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design, in which specific components in both cycles are shared. Here,top cycle and bottom cycle are combined to form the connected cycleshown in Fig. 1c. Even though the flow rate of C3 in bottom cycle isdifferent from that of C1 in top cycle, the inlet/outlet pressure andtemperature for both C3 and C1 can be identical. Thus, C3 is overlappedto C1 to use C1 only in the connected cycle. Similar simplification wasalso done for cooler 2 and cooler to use one cooler only in Fig. 1c.

Now, how the connected cycle integrated by PFBC boiler is dis-cussed. The system performs the heat-power conversion in three stages.Consecutively, flue gas in PFBC boiler drives the top S-CO2 cycle by MHand RH for first stage, GT for second stage, and then the bottom S-CO2

cycle by BH for third stage.

2.2. The system B

The system A uses three stages for heat-power conversion. However,system B uses two stages of heat-power conversion, which is simpler incycle configuration compared with system A. The system includes aRC + RH and an AC-GT sub-system (see Fig. 2). Flue gas in PFBC boilerdrives RC + RH by MH, SH and RH for first stage, where SH representssuperheater. Then, the residual heat of flue gas drives GT for secondstage.

We note that a gas cleaning sub-system (GS) is included in systems Aand B, operating in a ~870 °C flue gas temperature. The traditional gascleaning sub-system is used. Thus, the available results can be used forsystem analysis in this paper. The gas cleaning system adopts two-stageor three-stage cyclone separators. Each cyclone separator consists of atangential inlet, a separation volume, an outlet of purified gas and anoutlet of separated particles. The cyclone separator has simple structurebut can work in high temperature/pressure environment. It separatesparticles with minimum size of ~5 μm. The International EnergyAgency (IEA) once used a three-stage cyclone separator in a 45 MWtpilot plant in UK with a purification efficiency of 99.8% and a totalpressure drop of 0.03 MPa [30]. An electrostatic precipitator (ESP) or aceramic cross flow filter (CCFF) integrated with cyclone separators canseparate more fine particles to further increase separation efficiencies,which can be used for commercial operation of large-scale power plant.For an 85MWe Tomatoh-Atsuma PFBC power plant, the cyclones andceramic filter were combined to separate mineral particles from exhaustgas. At the filter outlet, the particles larger than 20 μm can be com-pletely separated, and the dust concentration of flue gas can be less than

28 mg/Nm3. The flue gas after particles separation entered the gasturbine. It was shown that the power plant could be operated stably fora long time [43–45]. The separation technique combining cyclones andceramic filter was also applied for a 360 MWe Karita PFBC power plant,which realized the commercial operation in July 2001. In such asystem, the flue gas consecutively flowed through a two-stage high-temperature cyclones separator and a particles filter. Under such con-dition, the dust content can be smaller than 5–30 mg/Nm3. After theseparation of particles, the flue gas is sufficiently clean to drive thecommercial gas turbine [29,43,46]. Moreover, the pressure drop of ESPor CCFF can be smaller than 0.006–0.007 MPa [30]. In summary, thepressure drop of a gas cleaning system is believed to be smaller than0.04 MPa. In this paper, the pressure drop in the flue gas side of PFBCboiler is set as 0.2 MPa, which is a reasonable estimation.

In system A, GT just follows GS. Thus, the inlet temperature of GTcan be ~870 °C (see Fig. 1). However, in system B, flue gas at the GSoutlet dissipates its energy to S-CO2 cycle, then enters GT. This ar-rangement yields quite lower inlet temperature of GT.

3. The simulation methodology

In both systems A and B, there is a strong coupling between S-CO2

cycle and PFBC boiler. The cycle computation method was describedfirst, followed by the analysis of the coupling.

3.1. The cycle computation

Our code for cycle computation was written in Fortran language.The computation was performed for per unit mass flow rate of CO2

[10,16]. In RC + RH, the thermal balance between hot side and coldside of HTR and LTR is written as follows

− = − −h h x h h(1 )( )7 8 3 2 (1)

− = −h h h h4 3 6 7 (2)

where h is enthalpy, x is the ratio of the mass flow rate in C2 related tothe total mass flow rate. Heat absorption in main heater MH and re-heater RH are

= −q h hMH 5 4 (3)

= −′ ′q h hRH 5 4 (4)

Fig. 2. The newly proposed system B (the cycle RC + RH integrating with AC-GT sub-system and PFBC boiler).

E. Sun, et al. Applied Thermal Engineering 166 (2020) 114756

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Heat dissipation in cooler is

= − −q x h h(1 )( )cooler 8 1 (5)

Turbine and compressor parameters are determined by assumedisentropic efficiency. The turbine T1 outlet enthalpy and power outputare

= − − = −′ ′ ′h h η h h w h h( ),s s4 5 T1, 5 4 , T1 5 4 (6)

For compressor such as C1, the compressor outlet enthalpy andpower consumption are

= +−

= − −h hh h

ηw x h h, (1 )( )s

s2 1

2, 1

C1,C1 2 1

(7)

where the subscript s represents isentropic condition.Calculations of T2 and C2 are similar to the above calculation, the

turbine inlet pressure (P5′) is [10]

=′P P P5 5 6 (8)

The thermal efficiency of RC + RH is

= + − −++η w w w w

q qth RC RHMH RH

,T1 T2 C1 C2

(9)

System A includes RC + RH and SRC, in which SRC can be calcu-lated in the similar method. System B only contains RC + RH.Calculations of gas turbine GT and air compressor AC share similarmethods to T1 and C1, which are

= − − = −h h η h h w h h( ),GT out GT in GT s GT in GT out s GT GT in GT out, , , , , , , , (10)

= +−

= −h hh h

ηw h h,AC out AC in

AC out s AC in

AC sAC AC out AC in, ,

, , ,

,, ,

(11)

where the subscripts in and out means inlet and outlet. The compressorinlet air temperature is assumed as 15 °C [47], ηGT,s and ηAC,s are 90%and 87%, respectively.

3.2. Coupling of S-CO2 cycle with PFBC boiler

Assumptions regarding the coupling between S-CO2 cycle and PFBCboiler are summarized as follows:

(1) The system operates in steady-state.(2) Processes in turbine and compressor are adiabatic but non-isen-

tropic.(3) Heat loss to environment is neglected for all facilities except for

cooler and exhaust heat in the boiler.(4) CO2 temperatures are kept identical for different CO2 streams to

minimize exergy destruction due to mixing.(5) The outlet temperature of gas cleaning sub-system (TGS) is 870 °C

[30–33]. TGS equals to TGT,in for system A, but equals to the outletflue gas temperature of SH and RH for system B.

(6) Reasonable temperature differences are maintained for various heatexchangers coupling flue gas and CO2. In system A, the inlet andoutlet temperature differences between the two fluids are 40 °C and30 °C respectively for BH in bottom cycle: T5b = TGT,out – 40 °C,Tfg,ex = T4b + 30 °C, where T5b and T4b are the outlet and inlet CO2

temperature of BH, Tfg,ex is the exhaust flue gas temperature. Insystem B, the pinch temperature of RH is 30 °C:TGT,in = T4′ + 30 °C. Because BH is not used, Tfg,ex equals to TGT,out.

After pressure/temperature parameters are determined for variousstate points, the thermal load assignment of PFBC boiler can be decided.The total thermal load is divided into following sections: heat absorp-tion by RC + RH recorded as QRC+RH, power generation by gas turbinerecorded as WGT, heat absorption by SRC recorded as QSRC, and heatrelease to environment recorded as Qex. These thermal loads are ex-pressed as follows

= −W m h h( )GT flue gas GT in GT out, , (12)

= −Q m h h( )SRC flue gas GT out fg ex, , (13)

= −Q m h h( )ex flue gas fg ex e, (14)

where the subscript e means environment, fg means flue gas, ex meansexhaust, mflue gas is the flow rate of flue gas, which is determined by coalconsumption. Finally, QRC+RH is calculated as follows

= − − − −+Q Q Q Q W W( )RC RH total SRC ex GT AC (15)

where Qtotal is the total heat input of coal. Then, the CO2 mass flowrates of RC + RH (mCO2,RC+RH) and SRC (mCO2,SRC) are

=− + −+

+

′ ′m Q

h h h h( ) ( )CO RC RHRC RH

2,5 4 5 4 (16)

=−

m Qh hCO SRC

SRC

b b2,

5 4 (17)

Eqs. (16) and (17) tell us that the CO2 flow rates depend on thethermal loads assigned to the corresponding cycle. We note that systemA contains both RC + RH and SRC, but system B has RC + RH only. Upto now, the thermal efficiency of the whole system is determined as

=+ +

−+ −η

W W WQ Qth

net RC RH net SRC net AC GT

total ex

, , ,

(18)

where the subscript net means net work, which is the power generationby turbine subtracting the compressor consumed work. The powergeneration efficiency ηe is

=η η η η ηe b th p g (19)

where ηb is the boiler efficiency which is dependent on Qex, ηp is thepipeline efficiency (ηp = 99%) and ηg is the generator efficiency(ηg = 98.5%). Table 1 summarizes various parameters in the compu-tation process. The auxiliary power rate is defined as the energy con-sumption by auxiliary components relative to the total power genera-tion. Based on the operation experience of a 360 MW PFBC powersystem, this rate is in the range of 2.2–2.9% [29]. Thus, the auxiliarypower rate is set as 2.5% in this paper.

Table 1The system design parameters.

Parameters Values

Turbine inlet temperature (T5) 590–640 °CTurbine inlet pressure (P5) 30 MPaTurbine isentropic efficiency 93%Compressor C1 inlet temperature (T1) 32 °CCompressor C1 inlet pressure (P1) 7.6 MPaCompressors isentropic efficiency 89%Pressure drops in LTR and HTR (ΔP) 0.1 MPaPinch temperature difference of recuperator 10 °CEnvironment temperature (Te) 15 °CBottom cycle compressor inlet temperature (T1b) 32 °CBottom cycle compressor inlet pressure (P1b) 7.6 MPaPressure drop of the bottom cycle heater (ΔPb) 0.2 MPaOutlet gas temperature of gas cleaning sub-system 870 °CExcess air coefficient 1.2Inlet temperature difference of BH (TGT,out-T5b) 40 °COutlet temperature difference of BH (Tfg,ex-T4b) 30 °COutlet temperature difference of RH (TGT,in-T4′) 40 °CAir compressor inlet temperature 15 °CIsentropic efficiency of air compressor 87%Isentropic efficiency of gas turbine 90%Pipeline efficiency 99%Generator efficiency 98.5%Total output 300 MWeSuperficial gas velocity 1 m/s

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3.3. The second law analysis of the system

For system analysis, the exergy of coal (ein) is [48]

⎜ ⎟= ⎛⎝

+ + + ⎞⎠

e Q HC

OC

NC

1.0064 0.1519 0.0616 0.0429in far

ar

ar

ar

ar

ar (20)

Qf means low heat value of coal (kJ/kg), Car, Har, Oar, Nar are themass fraction of carbon, hydrogen, oxygen and nitrogen in coal whichare shown in Table 2.

Exergy at each state point along cycle and exergy destruction ofeach component can be calculated based on Ref. [16]. The calculationmethods for exergy and exergy destruction can be available in manytextbooks, thus they are not repeated here. The second law efficiency ofthe system is

∑=∑

= −=ηW

eI1i

N

i

incomII

1

(21)

where Icom is the exergy destruction of each component. Physicalproperties of CO2, air and flue gas come from REFPROP [49].

4. Results and discussion

4.1. Systems A and B at conventional combustor pressure

For 15–80 MWe moderate scale power plant driven by PFBC boiler,the combustion pressure of PFBC boiler Pfur is 1.2–1.6 MPa [30]. In thispaper, Pfur is assumed to be equal to the outlet pressure of air com-pressor, which is Pfur = 1.6 MPa. Fig. 3a shows the increased thermalefficiencies ηth with increases of main vapor temperatures T5 for sys-tems A and B. When T5 is increased from 590 °C to 640 °C, system Aincreases ηth from 55.32% to 56.82%, while system B raises ηth from52.57% to 55.48%. Main vapor temperatures have significant influenceon thermal efficiencies.

Boiler efficiencies ηb are explored in Fig. 3b. Constant ηb is observedfor system A. Gas cleaning sub-system operates in a temperature of870 °C, which equals to the inlet of gas turbine: TGS = TGT,in = 870 °C.Incorporating the constant isentropic efficiency of GT yields not variedTGT,out. Further, based on the temperature difference limit of BH forbottom cycle, Tfg,ex is not changed versus main vapor temperatures T5.One shall remember that boiler efficiency is mainly dependent on Tfg,ex(see Fig. 1). The constant Tfg,ex generates the constant boiler efficiency.The situation is changed for system B. With increase of T5, T4′ is in-creased. This increase trend directly raises both inlet and outlet tem-peratures of GT (see Fig. 2). Thus, the exhaust flue gas temperatureTfg,ex is increased to worsen the boiler efficiencies. The different trendsof ηb ~ T5 achieves the crossing of the two curves for the two systems.

Power generation efficiencies ηe and exergy efficiencies are in-creased with increase of main vapor temperatures T5 for systems A andB (see Fig. 3c–d). System A has higher ηe by ~2% compared with systemB. Fig. 4 explains why system A has better performance than system B.PFBC boiler dominates exergy destruction in both the two systems.System A operates gas turbine GT in a much higher flue gas temperaturelevel than system B. In other words, GT in system A directly utilizes thehigher grade flue gas energy for heat-power conversion than that insystem B, which is the major reason of better performance for system A.

Power generation efficiencies (ηe) are also increased with increaseof main vapor pressures P5 (see Fig. 5a). As P5 increases from 22 MPa to36 MPa, ηe increased by 1.6% for system A, and 1.8% for system B.Pressure drop in flue gas side of PFBC boiler is recorded as ΔPfur, whichcan be used to calculate the inlet pressure of GT as PGT = Pfur − ΔPfur.It is found that ηe obviously decreases with increase of ΔPfur. The in-creased ΔPfur decreases PGT to lower power generation of GT. Specialattention should be paid to reduce pressure drop of flue gas in PFBCboiler.

Effect of isentropic efficiencies of CO2 turbine (ηT,s) on ηe is shown inFig. 5c. We note that ηT,s has significant effect on ηe. It is identified thatwhen ηT,s decreases by 1%, ηe decreases by ~0.3%. The decrease trendof ηe with decrease of isentropic efficiencies of gas turbine (ηGT,s) isshown in Fig. 5d. When ηGT,s decreases by 1%, ηe decreases by ~0.26%.Further, if isentropic efficiencies of CO2 compressor ηC,s and air com-pressor ηAC,s decrease by 1%, ηe decreases by ~0.14% and 0.11%, re-spectively. Our results indicate that turbomachinery for S-CO2 cycleand AC-GT apparently influence power generation efficiencies.

In summary, system A has obviously higher efficiencies than systemB. From the efficiency point of view, system A is recommended.However, system B has simpler system layout and lower operatingtemperature for gas turbine, it is easier to be applied for practical ap-plications.

4.2. Effect of combustion pressures of PFBC boiler on the performance ofsystems A and B

The system performance under fixed combustion pressure of PFBCboiler at Pfur = 1.6 MPa was discussed in Section 4.1. Because flue gasdirectly enters gas turbine, there is a strong connection between Pfurand net power output of gas turbine. The research and development ofPFBC boiler was made great progresses in 1970–90s [30]. Limited bypressure ratio of gas turbine in the range of 15–25, the operationpressure of PFBC boiler seldomly exceeded 2 MPa. Since then, thetechnology development gradually increased pressure ratios of gasturbine to 40, breaking through the high-pressure combustion barrier[50,51]. Here, we analyze how the varied Pfur influences the systemresponse. The main vapor parameters of CO2 are T5 = 620 °C andP5 = 30 MPa for RC + RH. Other parameters are identical to those inSection 4.1, but Pfur changes from 1 MPa to 3.5 MPa.

For system A, the total available heat characterized by the low heatvalue of coal is divided into four parts: heat absorption by RC+ RH, netheat-power conversion by AC-GT (gas turbine power subtracting aircompression work), heat absorption by SRC, and residual heat dis-charged to environment. Correspondingly, four ratios are defined asαRC+RH for heat absorption ratio by RC + RH, αAC-GT for net-powerratio by AC-GT, αSRC for heat absorption ratio by SRC, and αe for heatdischarged to environment ratio. Compared with system A, system Bdoes not have SRC. The energy distributions in systems A and B areshown in Fig. 6a and b, respectively. The higher-pressure-operation ofPFBC boiler increases the heat absorption ratio by RC + RH, but de-creases the heat release ratio to environment. When Pfur increases from1.0 MPa to 3.5 MPa, αRC+RH increases from 57.83% to 66.25% forsystem A, and from 79.30% to 90.96% for system B. Meanwhile, αedecreases from 14.57% to 9.12% for system A, and from 16.05% to8.63% for system B.

The decreased heat emission to environment improves boiler effi-ciencies at higher combustion pressures (see Fig. 7a). AC-GT accountsfor different energy ratios versus Pfur (see Fig. 7b). System A operatesgas turbine in higher flue gas temperature to result in higher αAC-GTthan system B. For system A, when Pfur changes from 1.0 MPa to3.5 MPa, the αAC-GT ~ Pfur curve displays parabola shape, and weakdependent of αAC-GT on Pfur is identified. System B decreases αAC-GT

Table 2Properties of the designed coal.

Car Har Oar Nar Sar Aar Mar Vdaf Qf

61.70 3.67 8.56 1.12 0.60 8.80 15.55 34.73 23,442

C (carbon), H (hydrogen), O (oxygen), N (nitrogen), S (sulfur), A (ash), M(moisture), V (Volatile).Subscripts ar, daf means as received, dry and ash free,Car + Har + Oar + Nar + Sar + Aar + Mar = 100.Qf means low heat value of coal (kJ/kg).

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covering the whole Pfur range. The high-pressure combustion atPfur = 3.5 MPa almost makes αAC-GT = 0, at which power generation bygas turbine is completely consumed by air compression to create zeronet power output for AC-GT sub-system.

Thermal efficiencies ηth for systems A and B shown in Fig. 7c behavesimilar variation trends of αAC-GT in Fig. 7b. Power generation efficiencyof system is the outcome of thermal efficiency multiplying by boilerefficiency and other efficiency components (see Eq. (19)), in whichboiler efficiency is a dominant factor to revise thermal efficiency. Thevariation trends of boiler efficiencies in Fig. 7a and thermal efficienciesin Fig. 7c directly result in the power generation variation trends inFig. 7d. For system A, with increase of Pfur from 1.0 MPa to 3.5 MPa, the

increased ηb and weakly varied ηth yields a general increase of powergeneration efficiencies to approach a constant value of 49.0% beyondPfur = 2.5 MPa. Meanwhile, system B increases power generation ef-ficiencies to approach a constant value of 46.67% beyondPfur = 2.5 MPa. Fig. 7 concludes that the combustion pressure of2.5–3.0 MPa is sufficiently high to achieve better system performance.Fig. 8 shows the distributions of exergy destruction in various compo-nents for systems A and B. The elevated combustion pressure in PFBCboiler obviously decreases the exergy destruction for PFBC boiler,which is mainly caused by the decreased residual heat discharged toenvironment at higher combustion pressure.

4.3. Concept design of PFBC boiler

Clapyron Equation Pv = RT tells us that specific volume v is in-versely proportional to pressure. With real gas assumption of flue gas,flue gas at P = 3.5 MPa has a specific volume of 2.86% of that atatmospheric pressure, indicating highly compact PFBC boiler due tohigh pressure combustion. A gas cleaning subs-system is involved in aPFBC boiler. The boiler is divided into a furnace volume and a tail fluegas volume interfaced at the end of gas-cleaning sub-system where theflue gas temperature is ~870 °C. The furnace volume behaves surfaceheating characteristic, i.e., heater surface is mainly arranged on furnaceside walls. However, heater surface can be freely arranged in the tailflue gas volume. In other words, the tail flue gas volume behaves bulkheating characteristic.

Because it is a challenge to arrange sufficient heater surface in acompact volume of furnace, it is necessary to explore the relationshipbetween the energy distribution in PFBC boiler and the heat absorptionby S-CO2 cycles and AC-GT (see Fig. 9a). For both systems A and B,

Fig. 3. The performance of systems A and B dependent on main vapor temperatures of CO2 at Pfur = 1.6 MPa.

Fig. 4. The exergy destruction distributions among various components forsystems A and B at Pfur = 1.6 MPa.

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furnace and tail account for ~66.25% and ~24.62% of total energyloads, respectively. The rest energy is discharged into environment. Thespatial energy distribution may not exactly correspond to the heat ab-sorption by S-CO2 cycle and AC-GT. For system A, RC + RH accountsfor ~66.25% of heat absorption, matching the energy load in furnace.AC-GT and SRC are responsible for the total heat absorption of tail fluegas volume. For system B, RC + RH dominates the heat absorption ofboth furnace volume and tail flue gas volume, while AC-GT only has asmall energy load.

There are two types of PFBC boilers: bubbling PFBC boiler andcirculating PFBC boiler [18,30]. Because circulating PFBC boiler con-tains gas-solid mixture of flue gas in the whole furnace volume, it ismore difficult to arrange heating surface. However, bubbling PFBCboiler contains obvious gas-solid interface, with the bottom volumecontaining gas-solid bubbling bed, and the top volume containing“pure” flue gas with few particles. It is easier to arrange heating surfacein the top volume. Thus, bubbling PFBC boiler is recommended to driveS-CO2 cycle and AC-GT. Usually, tube bundles are buried in the gas-solid bubbling bed. For present application, S-CO2 flows in tube bundlesto extract heat from bubbling bed. More attention should be paid tocontrol the temperatures of tube bundles in bubbling bed. First, S-CO2

heat transfer coefficients are in the range of 3000–5000 W/m2 K, whichare not as high as one imagines [30]. Second, the CO2 temperatureentering tube bundles is ~500 °C, which is more than ~100 °C higherthan water entering boiler for water-steam Rankine cycle [52].

In order to arrange sufficient heat transfer area in the top “pure”flue gas volume, the compact heat exchanger is proposed and shown inFig. 9b. Flue gas channel and S-CO2 channel are alternatively populatedin a compact volume. In flue gas side, plate fins can be used with

Fig. 5. The systems performance (a & b: ηe dependent on P5 and ΔPfur, respectively, c & d: ηe dependent on isentropic efficiencies of CO2 turbine and gas turbine,respectively).

Fig. 6. Energy distribution in systems A and B at varied combustion pressures ofPFBC boiler.

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hydraulic diameter larger than 10 mm. Plate fins are extended heattransfer surfaces, while larger hydraulic diameter is helpful to avoidblockage of solid particles. The S-CO2 side can use the mini-channelwith hydraulic diameter in ~mm. In fact, the printed-circuit-heat-ex-change (PCHE) concept [53] can be integrated in CO2 side for heattransfer enhancement.

In addition to arrange compact heat exchangers in furnace volume,an alternative strategy is to decrease the thermal load in such volume.For conventional coal fired boiler, the excess air coefficient is 1.2 [10].The increase of excess air coefficient obviously decreases the thermalload in furnace volume, but increases the thermal load in tail flue gasvolume, noting that the furnace volume and tail flue gas volume areinterfaced at ~870 °C temperature. This method is applicable for bothsystem A and system B (see Fig. 10a). The shortcoming of the increasedexcess air coefficient is the slightly decreased boiler efficiencies (seeFig. 10b). This is because the mass flow rate of flue gas is raised whenexcess air coefficient increases, even though the exhaust flue gas tem-perature is not changed.

Fig. 10c shows the flue gas recirculation method. An additional fluegas compressor is necessary to fulfil the flue gas recirculation. Thus, AC-GT is replaced by AC-GC-GT. The gas recirculation ratio β is defined asthe circulation flow rate divided by the total flow rate of flue gas.Fig. 10d illustrates that the thermal load fraction in furnace volume isdecreased from 66.25% to 57.67% corresponding to β from zero(without gas recirculation) to 0.35. It is noted that the modulation ofgas recirculation method is only suitable for system B. The obviouslydecreased thermal load in furnace volume makes it easier to arrangeheating surface area in a compact volume.

The present investigation involves some parameters for the analysis.

Fig. 7. Effect of combustion pressures of PFBC boiler on the performance of systems A and B.

Fig. 8. The distributions of exergy destruction in systems A and B atPfur = 1.0 MPa and 3.0 MPa.

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The topic is related to the coupling between PFBC boiler and cycle.Thus, the parameters can be classified as those for boiler and cycle. Theboiler parameters are combustion pressure, excess air coefficient andflue gas temperature at the end of gas cleaning system. The cycleparameters are the main vapor pressure and temperature, isentropicefficiencies of compressor and turbine, and pinch temperature of heatexchangers. The target parameters are exergy efficiency and powergeneration efficiency. These parameters either come from the operationexperience of practical component or system, or are performed on thesensitivity analysis to demonstrate their effects on the outcome resultsof system. For example, the 870 °C flue gas temperature at the end ofgas cleaning system is the reasonable value that a conventional PFBCboiler can reach at this stage. The combustion pressures and excess air

coefficients are varied over a wide range to explore their effects onsystem performance. In the cycle side, the main vapor pressures arevaried in the range of 22–36 MPa, in which 30 MPa pressure is themature value that can be reached for a supercritical water-steamRankine cycle. The main vapor temperatures are changed in the rangeof 590–640 °C, in which 620 °C is the value that can be easily reachedfor current water-steam Rankine cycle power plant. Besides, availablestudies focus on small scale centrifugal type compressor and turbine forS-CO2 cycle. Indeed, there are no reliable efficiencies for large scale(> 10 MWe level) axial type compressor and turbine for S-CO2 cycle.Thus, effects of both main vapor pressures/temperatures and isentropicefficiencies of turbomachines are thoroughly shown in this paper. TheS-CO2 cycle driven by PFBC boiler involves a set of parameters in the

Fig. 9. The relationship between energy spatial distribution and energy absorption ratios (a) and the proposed compact heat exchanger for flue gas-CO2 heat transfer(b).

Fig. 10. The increased excess air coefficient or flue gas recycling to adjust the energy load distribution in PFBC boiler (a: AC-GT with excess air coefficient αair, b: AC-GC-GT with flue gas recycling ratio β by GC, c: effect of αair on boiler efficiency ηb and d: effect of β on energy load fraction in furnace).

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system level. At this stage, it is difficult to present these parameters foran independent research group. The research and development of sucha new power generation system needs wide international cooperation.The perspective studies can be seen in Ref. [10]. Great effort will be puton experiments.

4.4. Concept design of the S-CO2 cycle driven by PFBC boiler

Here, the concept design of S-CO2 cycle driven by PFBC boiler waspresented, with the combustion pressure of Pfur = 3.5 MPa (see Figs. 11and 12). The main vapor parameters of CO2 are 620 °C/30 MPa. Be-cause the commercial PFBC power plant reaches the maximum powercapacity of 360 MWe, the 300 MWe power capacity was used for thepresent design. Thus, the related technologies of PFBC can be applied inS-CO2 power cycle. The bituminite coal is used as the fuel, whoseproperties can be seen in Table 2. Various parameters are marked inFigs. 11 and 12, including the temperature/pressure parameters at eachstate point, various flow rates of coal, flue gas, air and CO2, andthermal/power loads for each component.

At the design load, the system A has the power generation efficiencyof 48.95%, which is higher than 46.51% for the system B.Correspondingly, the system A has the coal consumption rate of91.77 t/h, which is smaller than 96.62 t/h for the system B. Comparedto the system A, the system B has simper structure. For example, BH(heater for bottom cycle), T3 and LTR2 are not necessary for the systemB. Besides, the flue gas temperature at the GT inlet (TGT,in) is only570.55 °C for the system B, which is more practical for engineeringapplication.

5. Conclusions

Following conclusions can be drawn.

• Two systems are proposed to integrate PFBC boiler and S-CO2 cycle.System A includes RC + RH, AC-GT and SRC to be driven by PFBCboiler, in which RC + RH and SRC are overlapped to simplify thesystem layout. System B involves RC + RH and AC-GT to be drivenby PFBC boiler.

• The increase of main vapor temperatures and/or pressures obviouslyimproves power generation efficiencies. The increase of pressuredrop of flue gas in PFBC boiler worsens the system performance.System A has higher efficiencies than system B.

• Variations of combustion pressure of PFBC boiler Pfur apparentlychange energy distributions in the system. The increase of Pfur de-creases flue gas heat discharged to environment, raising boiler ef-ficiencies. Power generation efficiencies show the increase trend butapproach a constant value when Pfur changes from 1.0 MPa to3.5 MPa.

• Bubbling PFBC boiler is recommended. Attention shall be paid tocontrol heating surface temperatures in bubbling bed region, whilecompact heat exchangers are suggested in the top “pure” flue gasregion.

• To overcome the challenge of heating surface arrangement in fur-nace volume, operation of PFBC boiler in larger excess air coeffi-cient is proposed to decrease thermal load fraction in furnaceαfurnace, deteriorating boiler efficiency. The better way recirculatesflue gas to decrease αfurnace, which is applicable for system B.

Fig. 11. S-CO2 PFBC power plant (system A).

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Declaration of Competing Interest

The authors declared that there is no conflict of interest.

Acknowledgements

The study was supported by the National Key R&D Program ofChina (2017YFB0601801), and the Natural Science Foundation ofChina (51821004).

Appendix A. Supplementary material

Supplementary data to this article can be found online at https://doi.org/10.1016/j.applthermaleng.2019.114756.

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