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Applied Thermal Engineering 31 (2011) 2829e2840

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Applied Thermal Engineering

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Integrated biomass gasification combined cycle distributed generation plantwith reciprocating gas engine and ORC

Jacek Kalina*

Silesian University of Technology, Institute of Thermal Technology, Konarskiego 22, 44-100 Gliwice, Poland

a r t i c l e i n f o

Article history:Received 4 January 2011Accepted 5 May 2011Available online 14 May 2011

Keywords:Renewable energyInternal Combustion EnginesOrganic Rankine CycleDistributed generationCombined cycles

* Tel.: þ48 32 2372302; fax: þ48 32 2372872.E-mail address: jacek.kalina@polsl.pl.URL: http://www.itc.polsl.pl

1359-4311/$ e see front matter � 2011 Elsevier Ltd.doi:10.1016/j.applthermaleng.2011.05.008

a b s t r a c t

The paper theoretically investigates the performance of a distributed generation plant made up ofgasifier, Internal Combustion Engine (ICE) and Organic Rankine Cycle (ORC) machine as a bottoming unit.The system can be used for maximization of electricity production from biomass in the case where thereis no heat demand for cogeneration plant. To analyze the performance of the gasifier a model based onthe thermodynamic equilibrium approach is used. Performance of the gas engine is estimated on thebasis of the analysis of its theoretical thermodynamic cycle. Three different setups of the plant are beingexamined. In the first one the ORC module is driven only by the heat recovered from engine exhaust gasand cooling water. Waste heat from a gasifier is used for gasification air preheating. In the secondconfiguration a thermal oil circuit is applied. The oil transfers heat from engine and raw gas cooler intothe ORC. In the third configuration it is proposed to apply a double cascade arrangement of the ORC unitwith a two-stage low temperature evaporation of working fluid. This novel approach allows utilization ofthe total waste heat from the low temperature engine cooling circuit. Two gas engines of differentcharacteristics are taken into account. The results obtained were compared in terms of electric energygeneration efficiency of the system. The lowest obtained value of the efficiency was 23.6% while thehighest one was 28.3%. These are very favorable values in comparison with other existing small andmedium scale biomass-fuelled power generation plants.

� 2011 Elsevier Ltd. All rights reserved.

1. Introduction

Within last two decades biomass has become a very interestingfuel in electricity generation sector. This is caused mainly byemission reduction and renewable energy policies that have beenadopted in many countries [1e4]. Poland can be an example.According to legal regulations the share of electricity from renew-able resources within the total amount of electricity sold to finalconsumers must not be lower than 10.4% in 2011 and will rise to12.9% in 2017 [5]. It is difficult to satisfy this obligation as there arevery limited sources of wind, solar and water energy in the country.The co-firing of biomass in existing coal fired plants can contributeat the level of 1.6e4.6% [6]. Therefore other alternatives of biomassimplementation into the energy market are required. Currently inPoland the financial measures that promote new projects includetradable green electricity, cogeneration and CO2 emission reductioncertificates. There are also subsidies available from the National

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Fund for Environmental Protection and Water Management andother financing institutions. More often the projects tend to beattractive to investors.

On the other hand biomass is a low bulk density fuel that can becharacterized by a significant demand for energy and costs duringgrowing, harvesting, processing, transportation and storage.Investment projects are feasible if biomass is available withina specified distance from designed location of a plant. Dependingon many different factors it is typically between 25 and 100 km.Even if this condition has beenmet organizing the fuel supply chainis not an easy task in the case of utility-scale plants. Therefore therenewable fuel is suitable rather for small and medium scaledistributed energy production facilities. However in order todemonstrate emission and primary energy saving potential as wellas economic profitability of biomass to energy conversion projectsa great care should be given to an effective use of available feed-stock resources.

An important technological alternative for the relatively small-scale distributed generation plants is based on biomass gasificationand gaseous-fuel-fired electricity generation equipment. Nowadaysa typical power plant consists of gasifier, gas cleaning devices andreciprocating Internal Combustion Engine (ICE) installed

Nomenclature

BT buffer tankcoil heat capacity of thermal oil [kJ/kg K]C cycloneCO condenserCR carburetorCP, Ch, CT Power, efficiency and temperature correction factors

[e]G gasifier; electric power generatorGC1, GC2 raw gas cooler and final gas coolerGF producer gas fanFL flareF1, F2, F3 filtersEC external coolingEFQ engine fuel quality index [e]EGC exhaust gas cooler/oil heaterEV evaporatorg0jiðT ; pÞ partial free enthalpy of pure component i under

temperature T and pressure p [kJ/kmol]hOT,in, hs,OT,out specific enthalpy of working fluid at expander

inlet and after ideal expansion [kJ/kmol]hwf,in, hwf,out specific enthalpies of working fluid at heat

exchanger inlet and outlet [kJ/kmol]hhc,in, hhc,out specific enthalpies of heat carrier at heat exchanger

inlet and outlet [kJ/kmol]H1, H2 working fluid heatersJC jacket water coolerlf number of phases [e]ls number of components [e]LHVb,db lower heating value of biomass; dry basis [kJ/kg]_mb;db mass flow rate of biomass; dry basis [kg/s]MC mixture coolernji number of moles of component i in phase jn0C; n

00C number of moles of carbon per kg of dry biomass at

reactor inlet and outlet [kmol/kg]_nwf ; _nhc flow of working fluid and flow of heat carrier [kmol/s]OT ORC turbinep pressure [kPa]pCO, pEV pressure at condenser inlet and evaporator outlet [kPa]P pump; power [kW]Pe reciprocating engine effective power [kW]Pel, Pel,eng electric power of the engine module [kW]

POT, PP electric power of ORC turbine and power consumed byworking fluid pump [kW]

_Qcc available heating power of engine cooling circuit [kW]R universal gas constant 8.314 [kJ/kmol K]sOT,in, ss,OT,out specific entropy of working fluid at expander inlet

and after ideal expansion [kJ/kmol K]T temperature [K]Tex exhaust gas temperature [K]TR,cs,out temperature at regenerator cold side outlet [K]TOT,out temperature at ORC turbine outlet [K]TP,out temperature at working fluid pump outlet [K]TC turbochargerv specific volume [m3/kmol]_Vg volumetric flow of tar free producer gas [Nm3/s]VLHV volumetric lower heating value [kJ/Nm3]VS Venturi scrubberW work [kJ]WS1, WS2 water separatorszji molar concentration of component i in a mixture [e]

Greek symbolsdQ relative heat losses [e]DH total enthalpy change [kJ]3C carbon conversion efficiency3R efficiency of heat exchanger [e]hcycle theoretical cycle efficiency [e]hcold_gas gasification process cold gas efficiency [e]hel electricity generation efficiency [e]hg electric generator efficiency [e]hi engine internal efficiency [e]hi,OT isentropic efficiency of ORC expander [e]hm mechanical efficiency [e]hv engine volumetric efficiency [e]l excess oxygen coefficient [e]ra air density [kg/m3]

Subscriptscycle theoretical thermodynamic cycleex exhaust gasg dry tar free producer gasmix engine fueleair mixtureref value obtained for the reference fuel (natural gas)

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e28402830

downstream. So far many problems have been encountered in thistype of plants (Herdin et al. [7]) that limit the annual number ofrunning hours. Some vendors however claim that an achievable timeof annual operation is about 8000 h. The technology can be regardedas commercially available and it can be applied in many distributedlocations (Warren et al. [8], Bridgwater [9]).

Utilization of the waste heat from raw producer gas cooler andICE significantly improves performance of a plant in a cogenerationmode. However, the frequent problem is that in distributed loca-tions the cogeneration mode is hardly possible, what is due to thelack in heat demand. In such case it would be interesting tomaximize the electric output of a plant by application of a bot-toming Organic Rankine Cycle (ORC). The survey performed byQuoilin and Lemort [10] showed that there is a commercial offer forORC units in the power range starting from6 kWand themarket forORC technology is growing rapidly.

Using ORC modules for waste heat recovery from InternalCombustion Engines is nowadays a very interesting option for

electricity generation [11,12]. Some engine manufacturers havealready introduced the ORC as a complementary product forstationary power generation systems [13,14]. Thirunavukarasu [14]showed that a project can be fruitful. The expected payback periodfor the additional investment in ORC module is relatively short. Theeconomic attractiveness of the technology can be expected ata high level in the case of biomass utilization projects, as there isusually available a financial support.

Nowadays there is already in operation a small number of thecombined ICEeORC systems [10,15]. The bottoming ORC unit booststhe electricity generation efficiency of the plant typically by 5e7percentage points. Due to characteristics of the reciprocatingengines there is usually a problemwith full recovery of an availablewaste heat from exhaust gases and engine cooling circuit. In mostof the systems only the exhaust gas heat is recovered [11,12]. On theother hand the heat transferred into the low temperature coolingcircuit may represent a substantial portion of the engine fuelenergy input. In the range of electric power below 1000 kW the

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e2840 2831

share of the cooling circuit heat is between 40% and 70% of a totalavailable waste heat from an engine. In bigger engines the lowtemperature heat represents between 35% and 50% of a totalavailable heating power. Therefore technical and economic effec-tiveness of the systemwith the bottoming ORCmodule depends onthe characteristics of a selected engine.

The main issues of designing an efficient ICEeORC combinedcycle were addressed by Vaja et al. [15]. They presented a thermo-dynamic analysis of matching a vapour cycle to the 2098 kWelectric power stationary natural gas fuelled engine module. Threedifferent setups of the systemwere taken into account: simple andregenerated organic cycles with the use of only engine exhaustgases and simple cycle with the use of exhaust gases and enginecooling water. There were also considered three different workingfluids: benzene (dry), R11 (isentropic) and R134a (wet). The resultsrevealed that by the application of the bottoming cycle the elec-tricity generation efficiency grows from 41.8% up to 47.0%. The bestperformance of the system is achieved with benzene bottomingcycles. There were almost no differences in efficiency between theregenerated ORC cycle and the cycle with preheating of workingfluid by engine cooling water. It was demonstrated that the OrganicRankine Cycles can recover only a small fraction of the heat releasedby the engine through the cooling water.

Thirunavukarasu [14] presented the combined ICEeORC systemdeveloped by the Caterpillar Inc. The system is composed of theG3520C (2150 kW) gas engine coupled to a superheated andparallel recuperated organic cycle with dry working fluid. Aninteresting aspect of the proposed configuration is that a part of theworking fluid flow after a pump is recuperated while another partof the flow is preheated by low temperature exhaust gas. Thesystem does not make use of engine cooling water. Achievableoverall electricity generation efficiency is 47%.

Juchymenko [16] presented the system developed by the GreatNorthern Power Corporation. In this case an ORC module recoversheat from engine jacket water and exhaust gas. This is a simplecycle without regeneration of heat. Engine jacket cooling water isused for preheating of working fluid. The heat from exhaust gas istransferred into a thermal oil circuit and then used in ORC evapo-rator. The increase of electric efficiency is 12% comparing to the casewith no waste heat recovery.

Utilization of the low temperature waste heat from an enginecooling circuit requires more complicated arrangements of the ICEeORC setups. An example of such system with a double cascade ORChas been proposed and analyzed by Vaja in Ref. [17]. In the proposedconfiguration thewaste heat from engine cooling process is used forpreheating a working fluid in the bottom part of the ORC cascade.Further heating of the fluid and its evaporation are performed usingheat of condensation of a working fluid in the upper cycle of thecascade. The upper cycle is driven by the engine exhaust gas. Theresults revealed that at optimal conditions asmuch as 51%of theheatrejected through the engine cooling circuit can be recovered.

In this theoretical study there is examined a hypothetical ICEeORC combined plant fired with producer gas from a downdraftgasifier. The technology can be potentially applied for efficient useof distributed biomass resources when transportation of the feed-stock to a distant central power plant is difficult and not financiallyattractive. According to the author’s knowledge not such systemhas been examined in literature yet. There is also being proposed inthe paper a novel arrangement of the double cascade bottomingORC module. The system is configured in the way that allows theutilization of the total amount of heat available from reciprocatingengine cooling circuit. Two reciprocating engines of differentcharacteristics are taken into account. The proposed technicalsolutions are being assessed in terms of the electric energy gener-ation efficiency. This performance parameter that directly

influences energy, environmental and financial effectiveness ofa potential project if electric power is the only product of a plant.

2. Plant configuration

The proposed biomass energy conversion plant is based ontypical downdraft gasifier, gas cleaning equipment and gas engineadjusted for producer gas operation. Biomass is converted into thecombustible gas in a fixed bed reactor using atmospheric air asgasifying agent. The producer gas goes from gasifier through gascooler, cyclone (removal of solid particles), Venturi scrubber(removal of tar), water separators (removal of liquid water), fan andfilters to buffer tank where final parameters of the fuel gas arestabilized. After buffer tank and final filter the fuel gas is mixedwith atmospheric air and the mixture goes through turbochargerand mixture cooler into a spark-ignition engine.

When a bottoming waste heat driven ORC module is added tothe system three alternative configurations are taken into account.In the configuration No. 1 a typical simple cycle ORC is applied andthe ICEeORC system is the same as presented in Refs. [15,16]. TheORCmodule is driven bywaste heat from engine cooling circuit andexhaust gas. The waste heat from a raw producer gas cooler isrecovered for preheating of the atmospheric air required for thegasification process. This significantly improves performance ofgasification and heating value of the fuel gas. Thus derating of theengine power is less significant and higher efficiency of the enginemodule can be obtained.

Scheme of the configuration No. 2 is given in Fig. 1. In this case itis proposed to power up the ORC module also by the waste heatfrom the raw gas cooler GC1. A thermal oil circuit is applied totransfer the heat from engine exhaust gas and raw producer gasinto the ORC. The oil is initially heated by the engine exhaust gasand then its temperature is elevated in the raw gas cooler. Whenthe heat recovered from the raw producer gas is used for drivingthe ORC module instead of preheating the gasification air, thecalorific value of the gas is lower and thus the derating of enginepower is more significant. On the other hand the lower electricpower of the enginemodule is compensated by the higher power ofthe ORC machine. Eventually the total electric output of the plant isslightly higher than in the configuration No. 1. An additionaladvantage of this configuration is that the gasegas heat exchangeprocess is avoided so the plant can be built and operated moreeasily.

In the configuration No. 3 a novel double cascade ORCarrangement is proposed in order to efficiently recover the wasteheat from the gas engine. In the upper part of the ORC cascadea superheated and regenerated cycle is applied. In the bottom partthere is proposed a simple cycle with a low evaporation tempera-ture that allows utilization of waste heat from the engine coolingliquid. Working fluid within the bottom cycle starts to evaporate ina low temperature boiler driven by hot water. The evaporationprocess is completed in the second heat exchanger that is also thecondenser for the upper cycle. Scheme of the plant in the config-uration No. 3 is given in Fig. 2.

The analysis is performed for a small-scale plant that can becharacterized by the gasifier raw gas chemical energy output in therange of 1 MW. Two CAT gas engines [18,19] of different types aretaken into account. Both of the engines require lean fueleairmixture, that is the current trend in the case of stationarymachinesfor on-site power production. In the producer gas mode of opera-tion the coefficient of excess oxygen l in the engine is kept the sameas reference value for natural gas operation. Specifications of themachines are given in Table 1. Exhaust gas and jacket water are theonly usable sources of waste heat. An aftercooler heat is totallydissipated into the atmosphere.

Fig. 1. Alternative configuration No. 2 of downdraft gasifiereICEeORC system.

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e28402832

3. Modeling of gasification process and gas engineperformance

Nowadays many papers concerning gasification process can befound in the literature. In this study it was decided to use a modelbased on the thermodynamic equilibrium of the chemical reactorsystem. The equilibrium models are generally regarded as satisfac-tory for performance studies of energy conversion plants [20e23]. Itwas however found and also confirmed by published literature thatif unconstrained calculation of equilibrium is performed, yield andheating value of the dry gas are too optimistic [20,24,25]. Thereforereasonable assumptions have to be made in order to correct theresults and to minimize a danger of wrong conclusions froma feasibility study of an energy conversion plant. In this study theassumptions are made in order to estimate the yields of non-equilibrium products: unconverted char (pure carbon in the formof graphite), tar and methane. These products are formed from thesubstrates and are not involved in the calculation of equilibrium.

The chemical system includes carbon, hydrogen, oxygen,nitrogen and sulfur. The producer gas from the gasifier is assumedto be composed of CO, CO2, H2, CH4, H2O, N2, SO2 and tar (CxHy). Theprocess involves also ash and argon from the atmospheric air thatare regarded as inert substances. There are distinguished two zones

of the reactor. Within the first zone heating of the charge, drying,pyrolysis and oxidation take place. The second zone includes thegasification of char in the oxygen free atmosphere.

It was assumed that the gasification process achieves the state ofthermodynamic equilibrium at final temperature and pressure atthe end of the reduction zone. The amount of non-equilibrium solidcarbon leaving the reactor is determined by the carbon conversionefficiency that is defined as follows:

eC ¼ n0C � n00Cn0C

(1)

If the value of carbon conversion efficiency that results from theequilibrium is higher than 0.95 the value of 0.95 is used instead. Itwas decided basing on experimental results.

In order to estimate the amount of methane from pyrolysis theassumptionsmade by Ratnadhariya and Channiwala [26] have beeninitially adopted. According to them half of the fuel hydrogen, thatis not associated with the fuel oxygen, is released as CH4 and then itis not involved in the calculation of equilibrium. It was found duringcalibration of the model that the additional assumption thata portion of CH4 from pyrolysis is converted in the oxidation zoneleads to a good agreement between calculated and experimental

Fig. 2. Alternative configuration No. 3 of downdraft gasifiereICEeORC system.

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e2840 2833

gas compositions. Remaining methane that is not converted in theoxidation zone is inactive in reduction reactions and passes to theproducts of the process [27].

Another non-equilibrium product present at the gasifier outlet isthe tar. It can be defined as CxHy where x� 6 [28]. The tar content inthe rawgas from a downdraft gasifier is typically within the range of0.3e5.0 g/Nm3 [28]. On the other hand the yield of tar is highlyunderestimated if calculated from the equilibrium of the gasifica-tion process. Therefore in this study the tar content of 5.0 g/Nm3 isbeing arbitrarily assumed, that is the upper limit of the given range.As presented by Gerun et al. [29] the pyrolysis tar is modelled asa mixture of benzene C6H6 (35%) and naphthalene C10H8 (65%).

Parameters of the system in the state of thermodynamic equi-librium are calculated using a non-stoichiometric approach withminimization of the Gibbs free energy:

Xls

i¼1

Xlf

j¼1

nijhg0jiðT ;pÞ þ RT ln zji

i/min (2)

The approach is reported in literature as satisfactory and it waspreviously used by other authors (Schuster et al. [20], Baggio et al.[23], Ruggiero and Manfrida [30]). An in-house Fortran code waswritten to perform the calculation. The model was verified againstexperimental results available in published literature. A comparisonof modeling results with experimental data presented by Jayah et al.[31] and with theoretical results obtained by Jarungthammachoteet al. [25] using unconstrained pure equilibrium model is given inTable 2. It can be observed that the unconstrained equilibriummodelleads to relatively goodprediction of gas composition at the values ofl higher than in the real gasifier. It causes the overestimated yield ofthe product gas and thus the efficiency of the plant is to optimistic.The constraint equilibrium model (tests 3 and 4) leads to betterresults at least in the aspect of product gas yield and its calorificvalue. These parameters have a key influence on mass and energybalance as well as on economic evaluation of the whole system. Itwas also observed during test runs of themodel that if the calculatedcontent of H2 in the product gas increases, the content of COdecreases from experimental values. The total content of these two

Table 1Technical specification of CAT gas engines [18,19].

Engine manufacturer Caterpillar Inc.

Engine type G3412 TA G3412C LEBore, mm 137 137Stroke, mm 152 152Number of cylinders 12 12CR 9.7 11.4Aspiration Turbocharged-aftercooled

Nominal electric power 280 kW 360 kWFuel energy input (LHV) 976 kW 1117 kWAftercooler temperature 54 �CCombustion air flow rate 19 Nm3/min 33 Nm3/minExhaust gas flow rate 20 Nm3/min 34 Nm3/minExhaust O2 (dry) 4.0% 8.5%la 1.22 1.83Exhaust gas temperature 454 �C 356 �CHeat rejection to jacket water (90/70 �C) 360 kW 329 kWHeat rejection to exhaust (LHV to 120 �C) 161 kW 190 kWHeat rejection to aftercooler 10 kW 49 kWElectricity generation efficiencya 28.69% 32.23%Total efficiency in cogeneration modea 82.07% 78.69%Reference fuel composition,b vol% CH4¼ 98.0; C2H6¼ 0.4;

C3H8¼ 0.2; C4H10¼ 0.2;C5H12¼ 0.3, N2¼ 0.8; CO2¼ 0.1

Reference fuel LHV 36.2 MJ/Nm3

a Values calculated; only exhaust gas and jacket water heat are taken into accountfor total efficiency.

b Assumed to get the reference fuel LHV the same as stated in the technicalspecification.

Table 3Properties of spruce chips

Parameter Value

Proximate analysis, % wt (dry basis)Watera 25.0Volatiles 70.2Fixed carbon 28.3Ash, wt% 1.5

HHV, MJ/kg 20.1Ultimate analysis, wt% (dry basis)C 51.2H 6.1O 40.9N 0.3

a Water content fixed at the maximum leveldefined for downdraft gasifiers which is 20% of mass,wet basis.

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e28402834

gases in the product gas is at relatively constant level. As these gaseshave similar heating values and the same stoichiometric oxygenrequirement it can be concluded that the equilibrium models aresatisfactory for systems with gas engines.

The biomass used in the analysis is spruce in the form of chips.Properties of the fuel are given in Table 3. Primary source of oxygento the process is humid atmospheric air, with a molar composition(dry air): oxygene 20.91%, nitrogene 78.09, argone 1.00%. Relativehumidity of the air is 60%. Pressure losses in the gasifier bed havenegligible impact on the results of equilibrium calculations [27].

The results of the modeling of the gasification process are givenin Figs. 3e5. The contents of combustible components in the rawdry tar free producer gas are presented in Fig. 3. Lower heatingvalue of the gas is presented in Fig. 4. Cold gas efficiency of thegasification process is presented in Fig. 5. The efficiency is definedas follows:

Table 2Comparison of results with the data published in literature.

Parameter Experiment[31]

Model[25]

Input dataAir excess ratio l 0.364 0.433Total heat loss, % of chemical energy input n.a 0.0Tar yield, g/Nm3 of dry gas n.a 0.0Max. 3C n.a. 100.0CH4 converted in oxidation zone, %of CH4 from pyrolysis

n.a. 0.0

ResultsMax. process temperature, �C 950e980a n.aDry raw gas composition, vol%N2 52.70 52.15H2 17.00 18.04CO 18.40 17.86CH4 1.30 0.11CO2 10.60 11.84Ar n.a. 0.00Sum of H2 and CO 35.4 35.9

Calculated calorific value, kJ/Nm3 4623 4241

a Presented in a figure.

hcold gas ¼_VgVLHVg

_mb;dbLHVb;db(3)

In the configuration No. 1 of the system the temperature ofpreheated gasification process air after the raw gas cooler GC1 isassumed to be 700 K. In this case the value of l in the reactor is setto 0.25, that leads to high calorific value of the gas and high cold gasefficiency of the process. In the case the atmospheric air enters thereactor at the ambient temperature the value of l is set to 0.30 thatensures high carbon conversion efficiency. Final characteristics ofthe producer gas used as the engine fuel is given in Table 4. This isthe input data for further analysis of the power plant.

The second important issue of the analysis of the proposedsystem is themodeling of the gas engine performance. The simplestapproach to this problemwas presented by Baratieri et al. [21] whohave just assumed thermal and electrical efficiencies of a machine.In a typical gas engine that was designed for natural gas operation,the achievable electric output is reduced by about 20e30% whenthe machine is fuelled with producer gas from a downdraft gasifier.The level of reduction is almost proportional to the change ofcalorific value of the fueleair mixture. Sridhar et al. [32] presentedthe results of the tests where the derating was observed at the levelof 29.4% and 28.5% for the engines with compression ratios ofCR¼ 12 and CR¼ 10.

To predict the engine performance Tinaut et al. [33] proposed anindex called the Engine Fuel Quality (EFQ). The index is the ratio of

Present study(test 1)

Present study(test 2)

Present study(test 3)

Present study(test 4)

0.433 0.364 0.364 0.3640.0 0.0 0.0 5.00.0 0.0 5.0 5.0100.0 100.0 95.0 95.075 75 50 75

1264 1083 1113 925

52.37 46.07 48.14 47.7116.75 21.73 19.89 20.5917.93 18.06 17.26 14.260.50 0.82 1.12 1.6111.77 12.73 12.97 15.220.67 0.59 0.62 0.6134.68 39.79 37.15 34.854252 4918 4729 4599

Fig. 5. Cold gas efficiency (1 e gasification air at 298 K; 2 e gasification air at 700 K).Fig. 3. Contents of combustible components in raw dry tar free producer gas(1 e gasification air at 298 K; 2 e gasification air at 700 K).

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e2840 2835

the volumetric heating value of a mixture under specified ther-modynamic conditions Tmix and pmix to the reference air density inkg/m3 under intake manifold pressure and temperature.

EFQ ¼ VLHVmixðTmix; pmixÞra;ref

(4)

Assuming that the value of l (or equivalence ratio) in an engineairefuel gas mixture remains unchanged the effective power of themachine Pe can be estimated basing on reference power accordingto the formula:

Pe ¼ Pe;refhvhihm

ðhvhihmÞrefEFQ

EFQref(5)

Another simplified method of assessment of the engine powerwas presented by Dasappa [34]. The method is based on a set ofdimensionless correction factors that take into account change ofthe most important parameters of the engine cycle. The authorclaims that the approach, though relatively simple, gives resultsclose to experimental measurements.

Papagiannakis et al. [35] presented a comparison betweenexperimental and computed results for a conventional multi-cylinder, four-stroke, turbocharged, spark-ignition, natural gas GEJ320 engine fuelled with syngas. The problem of derating of enginepower was not directly addressed in the paper. It can be howeverconcluded from a graphical representation of results that when the

Fig. 4. Lower heating value of raw dry tar free producer gas (1 e gasification air at298 K; 2 e gasification air at 700 K).

engine was fuelled with the syngas of the relatively good calorificvalue (6.84 MJ/Nm3) the achievable power was slightly higher thanthe one computed for natural gas operation. The increase of powerwas related to the amount of heat delivered into the engine cyclewith the fueleair mixture. The heat release curves presented in Ref.[35] show that more heat was delivered with the syngaseairmixture. Additionally initiation of combustion under the syngasoperation started a little earlier. Finally the maximum cylinderpressure computed for the syngas was higher. There was reporteda little reduction of the engine efficiency.

In this paper it is proposed to estimate the performance of anengine after the change of fuel using the spark-ignition enginefueleair theoretical cycle analysis. According to Heywood [36] suchcycle can be composed of the following processes:

e reversible adiabatic compression of mixture of air, fuel andcylinder residual gas,

e combustion at constant volume without heat loss, to burnedgas in chemical equilibrium,

e reversible adiabatic expansion of the burned gas,e ideal adiabatic exhaust blowdown,e ideal intake with adiabatic mixing between residual gas and

fresh mixture.

The advantage of this approach is that it takes into account:

e variations of density of the mixture and volumetric efficiencythat result from mixing with a residual amount of gas in thecylinder,

e variations of maximum temperature and pressure of the cycle,

Table 4Characteristics of raw gas and fuel gas from downdraft gasifier.

Parameter Gasification air at 298 K Gasification air at 700 K

Raw gas Fuel gas Raw gas Fuel gas

l in gasifier 0.30 0.25_Vg= _mb;wb 2.361 2.181 2.165 2.012Tg 914 298 906 298VLHVg 4943 5261 5648 5978Molar compositionzH2 0.216 0.234 0.234 0.251zCO 0.150 0.163 0.165 0.178zCH4 0.018 0.019 0.026 0.028zCO2 0.136 0.147 0.135 0.145zN2 0.383 0.415 0.349 0.375zH2O 0.092 0.017 0.086 0.017zAr 0.005 0.005 0.004 0.005

Fig. 6. TeDH diagram in case of simple cycle ORC with direct use of exhaust gas.

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e28402836

e different properties of working fluid in compression andexpansion processes due to combustion.

The performance of the machine was estimated under theassumption that the parameters of a real machine vary propor-tionally to the parameters of the theoretical cycle. Assuming thatmechanical efficiency hm and electricity generator efficiency hgremain unchanged after the change of fuel the following formulasare used for estimation of electric power, efficiency and exhaust gastemperature of an ICE module:

Pel ¼ Pel;refWcycle

Wcycle;ref¼ Pel;refCP (6)

hel ¼ hel;refhcycle

hcycle;ref¼ hel;refCh (7)

Tex ¼ Tex;refTex;cycle

Tex;cycle;ref¼ Tex;refCT (8)

It was also assumed that the ratio of exhaust gas heat to totalavailable engine heat remains constant in the producer gas mode.Therefore the available jacket water heat _Qcc can be estimatedusing formula:

_Qcc ¼ _Qcc;ref ¼ DHexjTex393 K

DHex;ref jTex;ref393 K

(9)

where DHex is the enthalpy change of exhaust gas between Tex and393 K (engine specification outlet temperature).

In order to perform engine simulation another in-house codewas built using Fortran. Properties of working fluids were calcu-lated using JANAF tables. As there was no detailed experimentaldata available in the literature, validation of the model was per-formed using data obtained during tests of commercial GAS-250power pack system purchased from Ankur Scientific Energy Tech-nologies Ltd. (India) [37,38]. Results of the validation are given inTable 5. It was concluded that the results are in relatively goodagreement with themeasured parameters. In details themodel waspresented elsewhere [39].

Table 5Validation of the model.

Source of data Ankur Scientific Ltd. [37]

Engine type Cummins GTA-1710-GBore, mm 140Stroke, mm 152Number of cylinders 12CR 10l 1.28Nominal electric power, kW 304Nominal efficiency 30%

Reference fuel composition, vol% CH4¼ 98.0; C2H6¼ 0.6;C3H8¼ 0.2; C4H10¼ 0.2;C5H12¼ 0.1, N2¼ 0.8; CO2¼ 0.1

Producer gas composition, vol% CH4¼ 3.4; CO¼ 14.4;H2¼ 18.6; N2¼ 50.1;CO2¼ 11.0; O2¼ 0.8;H2O¼ 1.1; Ar¼ 0.6

Power in producer gas mode, kW 240e250Efficiency in producer gas mode 27.52e28.67%Relative power 0.789e0.822Relative efficiency 0.917e0.955Modeling resultsCalculated relative power, CP 0.794Calculated relative efficiency, CEFF 0.919Calculated relative exhausttemperature, CT

0.979

4. Modeling of Organic Rankine Cycle

There are two key factors influencing an increase of electricpower and efficiency of the ICEeORC system. These are efficiency ofthe bottoming cycle and amount of waste heat recovered. In typicalarrangements of the system the second factor is usually unfavor-able as themost of heat from the engine cooling circuit is dissipatedinto the atmosphere [15].

In this study a comparison is made between typical system andproposed double cascade ORC arrangement. Calculations have beenperformed using R123 as the working fluid in the case of simpleORC and pair of R123 and R245fa in the case of double cascade ORC.According to the survey made by Quoilin and Lemort [10] both ofthese fluids are among recommended ones for small-scale ORCmodules.

Organic Rankine Cycle has been studied so far by many authors.In most of the proposed cycles the working fluid has saturationparameters at the expander inlet. This assumption is also adoptedin this study in the case of simple cycle ORC. In the double cascadearrangement superheating of working fluid and regenerative heatexchange are taken into account in order to maximize the totalefficiency of the system.

In a typical arrangement of the ORC, where the heat ofcondensation of working fluid is released into the atmosphere, thedry organic fluids do not need to be superheated since the regen-erative cycle thermal efficiency remains approximately constant[40] ore even a reduction of the efficiency can be observed [41]. Inthe system that is presented in this paper the superheating isapplied only within the upper cycle of the cascade. The heatreleased from this cycle helps to evaporate the working fluid in the

Fig. 7. TeDH diagram in case of simple cycle ORC with thermal oil circuit.

Fig. 8. TeDH diagram in case of double cascade ORC with thermal oil circuit.

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e2840 2837

lower cycle to the saturation line and in the same time to make useof the total heat available from the engine cooling circuit. Thus thetotal efficiency of the system increases. Temperature profiles offluids within the proposed configurations of the system are pre-sented in Figs. 6e8.

Heat transferred into the bottoming ORC is calculated fromenergy balances of heat exchangers:

_nwf

�hwf ;out � hwf ;in

�¼ �

1� dQ�_nhc

�hhc;in � hhc;out

�(10)

Relative heat losses from heat exchanger are assumed at thelevel of 3% of hot fluid enthalpy change (dQ¼ 0.03). In the case ofraw gas cooler the value of dQ¼ 0.05 is used.

In the systemwith a single stage ORC and direct use of the engineexhaust gas (configuration No.1), the temperature of the gas leavingawaste heat boiler Tex,out and approach temperature inworkingfluidpreheater H1DTH1were specified: Tex,out¼ 373 K andDTH1¼10 K. Inthis configuration the flow of R123 is limited by temperatureconstraints within the exhaust gas heat recovery. Therefore the

Table 6Calculated characteristics of CAT G3412 TA based system.

System parameter Unit Configurati

ICE module power correction factor CP e 0.758ICE module efficiency correction factor Ch e 0.914ICE module exhaust gas temperature correction factor CT e 0.968ICE module power Pel,eng kW 212.2ICE module efficiency heng e 0.262ICE exhaust gas temperature Tex K 703.6ICE total available waste heat kW 460.2ICE cooling circuit waste heat kW 314.5Waste heat carrier e Exhaust gasType of ORC e Simple cyclEvaporation temperature K 433.0Condensation temperature K 313.0ORC turbine inlet temperature K 433.0ORC turbine outlet/inlet working fluid volume ratio e 21.1Thermal oil circuit temperature at ORC inlet/outlet K e

Thermal oil circuit temperature at engine exhaustgas waste heat boiler

K e

Engine exhaust gas outlet temperature K 373.0Engine cooling water ORC outlet temperature K 360.7ORC module heat input kW 188.5ORC module nett power kW 26.2ORC module efficiency e 0.139ICEeORC system nett power kW 238.4Dry biomass consumption kg/h 193.7System input power kW 1010System efficiency hel e 0.236System efficiency without ORC bottoming cycle e 0.210

a Upper and lower cycle of the cascade.

amount of waste heat recovered from the engine cooling circuit islow and cooling water temperature drop is insignificant.

In the configuration No. 2 there were specified: approachtemperature in working fluid preheater H1 DTH1, approachtemperature in working fluid heater H2 DTH2, pinch point temper-ature in evaporator EV DTEV and approach temperature in exhaustgas-thermal oil heat exchanger EGC DTEV. The value of 10 K wasassumed for all of these variables. In this configuration thetemperature of the raw producer gas after the gas cooler GC1 mustbe high enough to prevent condensation of tar. Therefore the valueof Tg,GC1,out¼ 500 K was assumed.

In the case of double cascade ORC the values of approachtemperatures in thermal oil cycle are the same as in the configu-ration No. 2 of the system. Due to a high value of heat transfercoefficient the temperature difference between R123 and R245fa incondenser/evaporator CO1/EV2 was assumed DTCO1/EV2¼ 5 K.

In the case of single stage ORC the evaporation temperature ofworking fluid was set at the value that causes the vapour expansionline being located on the Tes diagram within the region of the dryfluid (lack of liquid phase in the turbine). For R123 this wasTEV¼ 433 K.

In order to maximize the total efficiency of the system in thecase of double cascade ORC, the evaporation temperature of R123within the upper cycle was assumed TEV¼ 450 K, that is close to thecritical temperature T¼ 456.94 K. Then, in order to improve theprocess of heat regeneration within the cycle, the working fluid issuperheated to TSH¼ 490 K.

The efficiency of regenerative heat exchanger R in the uppercycle of ORC cascade is defined as follows:

eR ¼ TR;cs;out � TP;outTOT;out � TP;out

(11)

The value of 3R¼ 0.70 is used in calculation.The condensation temperature of working fluid in each case is

TCO¼ 313 K, that is limited by ambient conditions. Within thedouble cascade ORC the condensation temperature of the upper

on No. 1 Configuration No. 2 Configuration No. 3

0.723 0.7230.909 0.9090.940 0.940202.3 202.30.260 0.260683.5 683.5428.4 428.4292.5 292.5

, cooling water Thermal oil, cooling water Thermal oil, cooling watere single stage Simple cycle single stage Double cascade

433.0 450.0, 345.0a

313.0 350.0, 313.0a

433.0 490.0, 345.0a

21.1 8.4, 2.7a

502.2/363.2 527.6/414.9449.6 479.6

373.1 424.9359.2 343.0286.11 210.1, 460.5a

39.8 30.9, 26.1a

0.139 0.147, 0.057a

241.9 259.1194.7 194.71015 10150.238 0.2550.199 0.199

Table 7Calculated characteristics of CAT G3412C LE based system.

System parameter Unit Configuration No. 1 Configuration No. 2 Configuration No. 3

ICE module power correction factor CP e 0.861 0.829 0.829ICE module efficiency correction factor Ch e 0.935 0.929 0.929ICE module exhaust gas temperature correction factor CT e 1.027 1.004 1.004ICE module power Pel,eng kW 309.8 298.3 298.3ICE module efficiency heng e 0.301 0.299 0.299ICE exhaust gas temperature Tex K 645.8 631.9 631.9ICE total available waste heat kW 534.7 503.2 503.2ICE cooling circuit waste heat kW 340.2 320.2 320.2Waste heat carrier e Exhaust gas, cooling water Thermal oil, cooling water Thermal oil, cooling waterType of ORC e Simple cycle single stage Simple cycle single stage Double cascadeEvaporation temperature K 433.0 433.0 450.0, 345.7a

Condensation temperature K 313.0 313.0 350.7, 313.0a

ORC turbine inlet temperature K 433.0 433.0 490.0, 345.7a

ORC turbine outlet/inlet working fluid volume ratio e 21.1 21.1 8.3, 2.7a

Thermal oil circuit temperature at ORC inlet/outlet K e 502.2/363.2 527.5/415.5a

Thermal oil circuit temperature at engine exhaustgas waste heat boiler

K e 451.7 480.4

Engine exhaust gas outlet temperature K 373.0 373.1 425.5Engine cooling water ORC outlet temperature K 360.1 358.4 343.0ORC module heat input kW 256.7 382.8 273.4, 540.0a

ORC module nett power kW 35.6 53.21 40.0, 31.0a

ORC module efficiency e 0.139 0.139 0.146, 0.058a

ICEeORC system nett power kW 345.5 351.2 369.0Dry biomass consumption kg/h 246.2 250.1 250.1System input power kW 1284 1304 1304System efficiency hel e 0.269 0.269 0.283System efficiency without ORC bottoming cycle e 0.241 0.229 0.229

a Upper and lower cycle of the cascade.

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e28402838

cycle of the cascade is set at the value that allows the total utili-zation of heat of engine cooling and water circuit temperaturedrops down to 343 K (that is required by the engine).

Exhaust gas expansion in turbine OT is described by isentropicprocess equation:

ss;OT;out ¼ sOT;in þ R lnpCOpEV

(12)

Temperature Ts,OT,out and specific enthalpy hs,OT,out are calculatedfrom the value of entropy after an ideal expansion. Then the powerof ORC turbine is:

POT ¼ _nwf�hOT;in � hs;OT;out

�hi;OThmhg (13)

The ORC module nett electric power is:

PORC ¼ POT � PP (14)

The power consumption for pumping of working fluid is:

PP ¼_nwfvwf ;CO;outðpEV � pCOÞ

hi;Phm(15)

Finally the electricity generation of the system is calculatedusing formula:

hel ¼Pel;eng þ

PORC

ðPOT � PPÞel;ORC�Pel;P;oil

_mb;dbLHVb;db(16)

One of the most important problems in the analysis of ORCperformance is selection of reasonable efficiencies of expandersand pumps. Invernizzi et al. [42] assumed that the isentropic effi-ciency of 40 kWORCmachine is 0.75. They claim that the efficiencyof above 0.80 is possible only if the ratio of volume of a workingfluid after to the volume before the expander is lower than 50. Thefluids R123 and R245fa fulfil this criterion.

Leibowitz et al. [43] presented that the isentropic efficiency ofscrew expanders is above 0.70 in the 20e50 kW systems and even

higher than 0.80 in the 1 MW power range. Saleh et al. [41]assumed isentropic efficiency of turbine equal 0.85.

In this work the isentropic efficiency of ORC turbine is assumedhiOT¼ 0.75. For pumps the efficiency hiP¼ 0.80 is used.

The model of ORC was developed using the Engineering Equa-tion Solver software. Properties of working fluids are modelledusing JANAF tables. In the case the thermal oil circuit is applied theTerminol VP-1 [44] is assumed as an intermediate heat carrier. Heatcapacity of the oil is:

coil ¼ 1:498þ0:00241ðT�273Þþ5:9591

�10�6ðT�273Þ2�2:9879�10�8ðT�273Þ3þ4:4172

�10�11ðT�273Þ4 (17)

Calculations were performed for all of the analyzed configura-tions of the system. The results are given in Tables 6 and 7 for thesystem based on CAT 3412TA and CAT 3412C LE engines. Nosensitivity analysis to the main design parameters of the ORC cyclewas carried out as this would significantly extend the volume ofthe text.

5. Conclusions

The theoretical study of the biomass-fuelled small-scale powerplant with downdraft gasifier, gas engine and bottoming ORCmodule was presented in the paper. Two gas engines of differentenergy balances were considered. Three different setups of thesystem were taken into account.

The total electricity generation efficiency of the plant dependson engine selection and configuration of the system. The lowestvalue obtained was 23.6% while the highest one was 28.3%. Theseare the values that are hardly observed in existing small andmedium scale biomass-fuelled plants. Another advantage of thesystem configuration proposed in this paper is that it is based ona standard downdraft gasifier. The gasifiers of this type arecommercially available [9,37].

J. Kalina / Applied Thermal Engineering 31 (2011) 2829e2840 2839

Much better performance parameters were obtained in the caseof the CAT 3412C LE lean burn engine. The derating of engine powerand efficiency was less significant in this case. The reference effi-ciency of the engine was high what resulted in higher values ofefficiency of the gasifier-engine system. The operation of this typeof engine fuelled with producer gas has been already demonstratedby Papagiannakis et al. [35].

Without the bottoming ORCmodule the best performance of theplant was obtained if the gasification air was preheated by the rawproducer gas leaving the reactor. With the bottoming cyclehowever, the efficiency of the plant was higher if the air at ambienttemperature was used. In this case the derating of electric power ofthe enginemodulewas compensated by an increase of power of theORC system.

The proposed double cascade arrangement of the bottoming ORCmodule allows effective recovery of the engine waste heat. Eventhough in this study there had been taken into account the engineswith a relatively high thermal power of the low temperature coolingcircuit, the bottoming module led to the increase of efficiency at thelevel of 5.5 percentage points. This is caused by the application oflow evaporation temperature of the working fluid within thebottom part of the cascade. The efficiency of the lower cycle of thecascade is at the level of 5.7%what in fact is very low value. Howeverthe increase of electric power due to high thermal power inputresults in high efficiency of the whole system. An additionaladvantage comes from using R245fa in the lower cycle. In the case ofthis working fluid a lower condensation temperature is possible(limited by requirement of a slight overpressure in the condenser).This will additionally boost up the power plant performance.

In the case the double cascade arrangement of the ORC moduleis applied, the electrical efficiency of the whole system is greaterthan that of standard configuration (No. 2) by 1.4 and 1.7percentage points for both engines respectively. The increase ofelectric power is in both cases at the level of 18 kW. On the otherhand double cascade cycle leads to an increase in plant complexityand investment costs. Its application should be justified by theeconomic analysis.

A brief financial evaluation of a potential investment project inPolish conditions revealedpositivevalues of theprofitability indices.Configurations No. 2 and 3 of the CAT G3412C LE based systemwerecompared. The estimated total investment cost was at the level of1,149,000 USD and 1,220,000 USD respectively. At the cost of woodat the level of 60 USD/t and the current value of electricity that isabout 172 USD/MWh (electricity selling price 72 USD/MWh andrenewable energy certificate price 100 USD/MWh [45]) the calcu-lated values of the Discounted Payback Period (DPB) are 8.2 and 7.9years for both configurations. The values of the Internal Rate ofReturn (IRR) are 13.3% and 13.8%. No investment cost subsidiesweretaken into account. Therefore the final conclusion is that thecombined ICEeORC system is relatively attractive from theeconomic point of view. The proposed double cascade ORC bot-toming module slightly improves the economic performance.

Acknowledgements

This work was carried out within the frame of research projectno. N N513 004036, titled: Analysis and optimization of distributedenergy conversion plants integrated with gasification of biomass. Theproject is financed by the Polish Ministry of Science.

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