Energy and exergy analyses of an externally fired gas turbine (EFGT) cycle integrated with biomass gasifier for distributed power generation

Energy 35 (2010) 341–350 Contents lists available at ScienceDirect Energy journal homepage: www.elsevier.com/locate/energy Energy and exergy analyses of an externally fired gas turbine (EFGT) cycle integrated with biomass gasifier for distributed power generation Amitava Datta*, Ranjan Ganguly, Luna Sarkar Department of Power Engineering, Jadavpur University, Salt Lake Campus, Kolkata 700098, India a r t i c l e i n f o a b s t r a c t Article history: Received January 2009 Received in revised form August 2009 Accepted 25 September 2009 Available online 28 October 2009 Biomass based decentralized power generation using externally fired gas turbine (EFGT) can be a technically feasible option In this work, thermal performance and sizing of such plants have been analyzed at different cycle pressure ratio (rp ¼ 2À8), turbine inlet temperature (TIT ¼ 1050–1350 K) and the heat exchanger cold end temperature difference (CETD ¼ 200–300 K) It is found that the thermal efficiency of the EFGT plant reaches a maximum at an optimum pressure ratio depending upon the TIT and heat exchanger CETD For a particular pressure ratio, thermal efficiency increases either with the increase in TIT or with the decrease in heat exchanger CETD The specific air flow, associated with the size of the plant equipment, decreases with the increase in pressure ratio This decrease is rapid at the lower end of the pressure ratio (rp < 4) but levels-off at higher rp values An increase in the TIT reduces the specific air flow, while a change in the heat exchanger CETD has no influence on it Based on this comparison, the performance of a 100 kW EFGT plant has been analyzed for three sets of operating parameters and a trade-off in the operating condition is reached Ó 2009 Elsevier Ltd All rights reserved Keywords: Gas turbine External firing Biomass Gasifier Introduction Small scale decentralized power generation is gaining importance for distributing electricity in the remote areas far from the centralized grid [1–4] The delivery of grid power to the remote areas, particularly in the hilly terrain, is extremely uneconomic [5] On the contrary, the installation of small capacity plants catering to the local needs using the local resource can be an attractive alternative for remote places Biomass is one of the important available primary resources, which generally exists in abundance in the villages and already serves as the source of energy e.g in cooking Energy from the biomass can be thermochemically recovered for the generation of electricity either through direct combustion or through gasification and subsequent combustion of the producer gas In large scale, biomass gasification can be used for power generation in a combined cycle [6,7] On the other hand, piston engines or micro gas turbines are suitable for small capacity distributed generation Producer gas can be used in conventional diesel engines in the dual fuel mode or in producer gas engines for the generation of power [8] However, such engines having * Corresponding author Tel.: þ91 33 23355813; fax: þ91 33 23357254 E-mail address: amdatta_ju@yahoo.com (A Datta) 0360-5442/$ – see front matter Ó 2009 Elsevier Ltd All rights reserved doi:10.1016/j.energy.2009.09.031 reciprocating components require more maintenance and abundance of cooling water, which make them unsuitable for remote locations The use of biomass as fuel in conventional (internally fired) gas turbine engines entails various problems [9] Firstly, the gas turbines are sensitive machines that require extremely clean gas to avoid damage to the turbine blades (such as erosion, incrustation, and corrosion) and blockage of filters and fuel injectors This requires installation of expensive gas clean up system, consisting of scrubbers, ceramic filters, cyclones etc., at the gasifier outlet Secondly, the low calorific value of the producer gas, obtained from biomass gasification, necessitates a high fuel flow It calls for a design modification in the combustor and the turbine inlet guide vanes, otherwise the change in the mass balance between the compressor and the turbine moves the compressor operating point towards surge [9] These problems are resolved, if the biomass can be conveniently used as a fuel in an externally fired gas turbine (EFGT) engine In an EFGT cycle [9], the high pressure air from the compressor is heated in a heat exchanger before admitting to the turbine The turbine essentially handles clean air and the turbine exhaust air is subsequently used to burn the fuel in a combustion chamber The combustion product is employed as the hot stream of the heat exchanger, before being released from the power cycle The cycle can employ dirty and low cost fuels, as the combustion products 342 A Datta et al / Energy 35 (2010) 341–350 Nomenclature AHE AFSt CETD Cp ei ech EFGT EN h hf K Mj Pi DP rp R s Ti TIT U w W X Z Heat exchanger surface area Stoichiometric Air-fuel ratio Cold End Temperature Difference of heat exchanger Specific heat at constant pressure Specific thermomechanical flow exergy at state i Specific chemical exergy Externally Fired Gas Turbine Energy released with exhaust gas Enthalpy Enthalpy of formation Equilibrium constant Molecular weight of species j Pressure at state i or Partial pressure for species i Pressure drop Pressure ratio Universal gas constant Entropy Temperature at state i Turbine Inlet Temperature Overall heat transfer coefficient of heat exchanger Specific work Work Number of moles Moisture content in the as-fired biomass (by mass) not enter the turbine Although the presence of ash in the products may cause erosion and fouling of the heat exchanger tubes, while corrosive products eats away the tube material, maintenance of the heat exchanger is much less troublesome than that for the turbine Anheden [10] presented thermodynamic and economic analyses of closed and open cycle externally fired gas turbine plants with direct combustion of biomass in a circulating fluidized bed furnace It is found that the efficiency reaches a maximum value at an optimum pressure ratio of the cycle Ferreira and Pilidis [9] compared the thermodynamic performance of an externally fired gas turbine cycle with direct combustion of biomass against an internally fired cycle firing either natural gas or producer gas from biomass gasification The study was performed for the simple gas turbine cycle as well as for the combined cycle operation with a steam based Rankine cycle at the bottom The results showed promising performance for the EFGT plant particularly considering the renewable and environment-friendly attributes of the biomass fuel Bram et al [11] reviewed the technological and economic feasibility of the external firing of biomass in gas turbines The authors concluded that cogeneration based on EFGT on the scale of 100–200 kWe offers good prospects from both economic and technical aspects Cocoa et al [12] evaluated the performance of a 100 kW externally fired gas turbine plant fuelled with biomass and having an integral dryer for biomass The influence of parameters like pressure ratio, turbine inlet temperature and temperature difference in the heat exchanger on the thermal efficiency for electrical generation was analyzed It was found that the dry biomass produces efficiency in the range of 22–33% and the integration of the dryer improves flexibility in the plant operation Traverso et al [13] presented the steady state and transient performance of an externally fired micro gas turbine pilot plant of 80 kW capacity fired with natural gas The paper demonstrated the feasibility of operation and control of the gas turbine plant of small capacity Greek Letters Equivalence ratio Ratio of specific heats Isentropic efficiency of compressor Isentropic efficiency of turbine Efficiency g hc,isen ht,isen h Sub-scripts a Air B Producer gas after gasification of biomass C Compressor CC Combustion chamber f Fuel G Gasifier g Product gas HE Heat exchanger in Input i Index for thermodynamic state point o Reference state T Turbine w Water Super-scripts c Cold side of the heat exchanger h Hot side of the heat exchanger All the literatures on EFGT universally claim that one of the biggest challenges in the design lies in developing the high temperature heat exchanger that is capable of achieving high turbine inlet temperature and at the same time withstands the stresses imposed by the working conditions and the constituent of the combustion product [9–12] The size of the heat exchanger and the cost of material are the two important considerations that decide the economy of the plant The use of nickel based super alloys in the heat exchanger allows the turbine inlet temperature to  reach 800–825 C, while more advanced oxide dispersion (ODS)  alloys withstand temperature up to 1100 C at the turbine inlet [10]  The turbine inlet temperature may be as high as 1300 C with ceramic heat exchanger materials [14], but prolonged operation with such exchangers is yet to be firmly tested Increase in the turbine inlet temperature is favorable towards achieving higher plant efficiency but it complicates the equipment design An uncooled micro gas turbine can sustain a maximum turbine inlet  temperature of 950 C, while further increase in the temperature requires turbine blade cooling arrangement [13] Since all these modifications towards performance improvement bear considerable cost implications, such modifications always needs a priori evaluation, based on energy and exergy based performance analysis of the cycle In the present work, we have conducted the energy and exergy based performance analysis of an externally fired gas turbine cycle running on biomass as fuel The effects of operating parameters, like pressure ratio, turbine inlet temperature, heat exchanger cold end temperature difference, on the thermal efficiency and specific air flow for the cycle have been analyzed The main focus of the present study is to identify the ideal operating parameters for the use of a EFGT plant for decentralized power generation supplying the local needs in the remote areas, where extending the grid power is uneconomic Accordingly, the performance parameters for a 100 kW gas turbine plant have been evaluated with selective sets A Datta et al / Energy 35 (2010) 341–350 of operating conditions An integral gasifier has been considered in the cycle for the gasification of the biomass fuel prior to its combustion This is because of the fact that the operation and control of a direct biomass combustor (like a CFB combustor as in [10]) at a small scale (as considered here) involves additional complexities and more number of skilled personnel that is unavailable in remote areas at low cost On the other hand, there is developed technology of biomass gasifier at small scale [15] which can be integrated with the proposed gas turbine plant An exergy based accounting has been performed for the cycle to find out the major irreversibilities in the cycle The exergetic efficiencies of the individual plant equipment are also compared at different cycle operating conditions Theoretical formulation 2.1 Description of the proposed cycle Fig 1(a) illustrates the schematic description of the externally fired gas turbine cycle analyzed, while Fig 1(b) shows the different processes on a temperature-entropy (T-s) plane In the power cycle, the ambient air is compressed in a centrifugal compressor over the pressure ratio (rp) of the cycle A part of the air is extracted from an intermediate stage of the compressor for the gasification of the Biomass a G Exhaust Gas HE B CC T C Air c P2- ΔPHE T5 T3 (=TIT) P2 Air is admitted to the compressor (state 1, refer Fig 1b) at atmospheric condition, P1 ¼101.325 kPa, T1 ¼ 300 K The compression process is adiabatic with an isentropic efficiency of 87% [9] The gasification process is adiabatic and chemical equilibrium is reached in the producer gas at the gasifier exit A total pressure drop (DPG) of 16 mm Hg column (i.e 2.13 kPa) [16] is considered across the gasifier The ultimate analysis of the dry biomass fuel (wood) shows a gravimetric composition of C: 50%, H: 6% and O: 44%, while the calorific value of the biomass (on dry basis) is 449568 kJ/ kmol (i.e 18732 kJ/kg) [17] The moisture content in the biomass is 20% on mass basis The pressure drop in heat exchanger cold side is 3% of the inlet pressure, while on the hot side the pressure drop is 1.5% of the inlet pressure [12] The expansion process in the turbine is adiabatic with an isentropic efficiency of 89% [9] Complete combustion takes place in the combustion chamber under adiabatic condition A pressure drop of 0.5% of the inlet pressure takes place across the combustion chamber  The reference temperature To and pressure Po are 25 C and 101.325 kPa, respectively 2.2 Energy analysis of the cycle b biomass stock, while the remaining air undergoes the full compression The compressed air is then heated in an indirect heat exchanger before entering the turbine After expansion in the turbine, the air is fed into the combustion chamber, where the producer gas, generated from gasification of the biomass, is burnt The high temperature products gas of combustion is then passed through the heat exchanger in order to heat the air, and finally released into the atmosphere The following assumptions have been made for the analysis of the cycle: A 2.2.1 Air compressor The compressor delivery pressure (P2) is evaluated using the cycle pressure ratio (rp), which is varied in the range of 2–8 The temperature of air (T2s) at pressure P2 for the isentropic compression is calculated considering a third order polynomial variation of the molar specific heat of air with temperature as, Cpair ¼ aair þ bair T þ cair T þ dair T [18] The actual work done on the compressor per kmol of admitted air (wC) is calculated using the isentropic efficiency of the compressor (hc,isen) The specific compressor work can be expressed as: wC ¼ CETD T h P1+ ΔPHE +ΔPCC h P1+ ΔPHE P1 T1 343 s Fig (a) Schematic diagram and (b) Temperature-Entropy diagram of the EFGT cycle C-Compressor, CC- Combustion Chamber, HE-Heat Exchanger, G-Gasifier, T-Turbine, TIT-Turbine Inlet Temperature, CETD-Cold End Temperature Difference of Heat Exchanger ZT2 Cpair dT T1    c   bair  T2 À T12 þ air T23 À T13 ¼ aair ðT2 À T1 Þ þ    dair  T2 À T14 þ (1) Eq (1) is solved by Newton–Raphson method for the actual temperature of air (T2) at the compressor outlet The air, extracted at the intermediate state point ‘A’ from the compressor, is used for the gasification of the biomass stock The pressure PA should be sufficient to overcome the pressure drop across the gasifier and feed the producer gas to the combustion chamber Therefore, 344 A Datta et al / Energy 35 (2010) 341–350   h PA ¼ P1 þ DPHE þ DPCC þ DPG (2) h , DP where, DPHE CC DPG are the pressure drops across the heat exchanger hot side, combustion chamber and gasifier respectively The temperature of extracted air (TA) is obtained using a similar approach for T2 described above 2.2.2 Gasifier The biomass feedstock (wood) is fed to the gasifier in a substoichiometric environment The gasifier environment is described by the equivalence ratio (f), which is defined as the ratio of stoichiometric air-fuel ratio to actual air-fuel ratio We have considered f ¼ 3.33 for our calculation A representative chemical formula is considered for the dry biomass fuel as CHQOR, using the mass percentage of carbon, hydrogen and oxygen, respectively, from the ultimate analysis of the fuel [19] The number of moles of oxygen for the gasification of kmol of dry biomass (CHQOR) is calculated as XO AFSt ,Mf ¼ 4:76,Ma f (3) where, AFSt is the stoichiometric air-fuel ratio for the fuel used and Mf and Ma are the molecular weights of fuel and air, respectively The amount of moisture (in kmol) fed with every kmol of dry feedstock is   B ¼ ðZ=ð100 À ZÞÞ, Mf =Mw (4) where, Z is the moisture content (mass percentage) in the biomass (as-fired) and Mw is the molecular weight of the water vapour The global gasification reaction can be expressed as follows: CHQ OR þ BH2 O þ XO O2 þ 3:76XO ,N2 ¼ X1 H2 þ X2 CO þ X3 CO2 þ X4 H2 O þ X5 CH4 þ 3:76XO ,N2 (5) where X1, X2, X3, X4 and X5 are the number of moles of H2, CO, CO2, H2O, CH4, respectively, produced on gasification of one kmol of wood The values of X1 through X5 are solved considering the carbon, hydrogen and oxygen balances from the chemical reaction (Eq (5)) and the chemical equilibrium in the product gas following the methanation reaction and water gas shift reaction [17,19] as below: C þ 2H2 4CH4 (6) CO þ H2 O4CO2 þ H2 (7) The equilibrium constants for methanation reaction (K1) and water gas shift reaction (K2) are expressed as follows: K1 ¼ À PCH4 =Po Á2 ¼ PH2 =Po X5 Po X Xj j¼1 X1 P4 À K2 ¼ ÁÀ Á PCO2 =Po PH2 =Po X X À Á ¼ X2 X4 ðPCO =Po Þ PH2 O =Po (8) (9) In Eqs (8) and (9), Pi represents the partial pressure of species i, while Po is the reference pressure P4 is the pressure at the gasifier exit (which is equal to the combustor pressure) The equilibrium constants K1 and K2 depend on the gasification temperature An energy balance equation is drawn to evaluate the gasification temperature (Tg) (assuming no heat loss from the gasifier) as follows: hfwood þ BhfH 2O B þ XO @ ZTA To B ¼ X1 @hfH þ ZTg To B þ X3 @hfCO þ ZTg To B þ X5 @hfCH þ ZTg C B CpO2 dT A þ 3:76XO @ 2O C B CpCH4 dT A þ 3:76XO @ To C CpN2 dT A ZTg C CpCO dT A To C B CpCO2 dT A þ X4 @hfH 1 To C B CpH2 dT A þ X2 @hfCO þ ZTA þ ZTg C CpH2 O dT A To ZTg C CpN2 dT A ð10Þ To where hfwood, hfH O , hfH , hfCO , hfCO and hfCH represent enthalpies of 2 formation of wood, moisture, hydrogen, carbon monoxide, carbon dioxide and methane, respectively The enthalpy of formation of wood has been derived from the heating value of the fuel The composition and temperature of the producer gas are obtained by solving for the values of X1 through X5 and Tg simultaneously Table shows the product gas concentration on gasification of rubber wood using the present gasifier model under two different moisture content and air-fuel ratio The composition of the biomass is taken from the earlier work of Jayah et al [20] and Sharma [21], who worked with gasification of biomass The corresponding results from the experiments of Jayah et al [20] and the equilibrium model of Sharma [21] under the same conditions are also given for comparison The results show that the present gasifier model predicts the gas composition fairly well 2.2.3 Turbine The turbine inlet pressure P3 is calculated considering a 3% pressure drop from the compressor exit (P2) in the cold side of the heat exchanger The temperature of air at the inlet to the turbine (T3) is an input parameter for the analysis The actual work done by the turbine per kmol of air (wT) is evaluated considering the variable specific heat and the isentropic turbine efficiency (ht,isen) The actual temperature of air at the turbine outlet (T4) is found by solving the following equation using Newton-Raphson method Table Product gas composition from gasification of rubber wood using the present model and from the works of Jayah et al [20] and Sharma [21] Dry Gas Composition Jayah et al [20] Experiment Sharma [21] Equilibrium Model Present Model Moisture content ¼ 16%, A-F Ratio ¼ 2.2 18.3 H2 CO 20.2 9.7 CO2 CH4 1.1 50.7 N2 19.35 19.34 11.18 0.25 50.19 18.97 24.75 8.01 0.39 47.88 Moisture content ¼ 18.5%, A-F Ratio ¼ 2.03 17.2 H2 CO 19.6 9.9 CO2 1.4 CH4 51.9 N2 19.85 19.64 11.01 0.26 49.26 20.91 23.79 9.25 0.99 45.06 A Datta et al / Energy 35 (2010) 341–350 wT ¼ ZT3 Cpair dT ¼ aair ðT3 ÀT4 Þþ T4   d þ air T34 ÀT44    c   bair  T3 ÀT42 þ air T33 ÀT43 (11) 2.2.4 Combustion chamber The combustion chamber is fed with the air from the turbine exhaust and the producer gas from the gasifier Complete combustion has been assumed in the combustion chamber following the chemical equation þ XO0 ðO2 þ 3:76N2 Þ/X6 CO2 þ X7 H2 O þ X8 O2 À Á þ 3:76 XO þ XO0 N2 considered in the analysis The thermomechanical exergy is defined with respect to a restricted dead state, which is characterized by the reference pressure and temperature of the dead state The specific thermomechanical flow exergy at any state is calculated from the generalized equation given as follows: ei ¼ hi À ho À To ðsi À so Þ (12) XO0 denotes the kmoles of O2 entering the combustor from the turbine for each kmol of wood fired in the cycle Assuming an adiabatic condition in the combustor, the energy balance is given as:- Xjðhfj þ j ZTg To ¼ X    Cpj dTÞ  Xj ðhfj þ j ZT5 To þ Xj ðhfj þ j Fuel    Cpj dTÞ  X ZT4 To    Cpj dTÞ  Air (13) Products where, Xj represents the number of mole of the jth component in fuel (the producer gas), air or product gas mixture and hfj and Cpj are the heat of formation and temperature dependent specific heat of that component Putting the number of moles of different components from Eq (12), it is found that Eq (13) reduces to one involving T5 and XO0 2.2.5 Heat exchanger The hot combustion gases leaving the combustor enters the heat exchanger at state and leaves at state 6, heating the compressed air from state to state without any heat loss to the surrounding (See Fig 1) A heat balance across the heat exchanger gives 4:76XO0 ZT3 T2 Cpair dT ¼ Xg ZT5 Cpg dT (14) T6 where, Xg represents the number of moles of hot exhaust gases leaving the combustor Following the reaction Eq (12), À Á Xg ¼ X6 þ X7 þ X8 þ 3:76 XO þ XO0 ZTi Cp dT (17) To s i À so ¼ X (16) where ‘i’ represents the state point (e.g through and A, as given in Fig 1) at which the exergy is evaluated and ‘o’ is the state point at the exergy reference environment hi À ho ¼ ðX1 H2 þ X2 CO þ X3 CO2 þ X4 H2 O þ X5 CH4 þ 3:76XO N2 Þ 345 (15) Cpg represents the molar specific heat of the exhaust gas mixture entering the heat exchanger Equations (13) and (14), representing the energy balance of the combustor and the heat exchanger, are simultaneously solved using an iterative technique to obtain the values of T5 and XO0 ZTi To Cp dT P À Rln i Po T (18) The chemical exergy is defined with respect to the true dead state, which considers the chemical composition of the reference environment in addition to the reference pressure and temperature [22] In true thermodynamic sense a multicomponent system possesses chemical exergy at restricted dead state when the partial pressure of the components in the system differs from the partial pressure of the same components in the reference environment However, in the combustion literature chemical availability is associated when useful work could be extracted through chemical reaction [23,24] at reference temperature and pressure conditions We have followed the latter concept in evaluating chemical exergy in this work ) is obtained from its lower The chemical exergy of the wood (ech wood heating value using a multiplication factor b [25], which is given by 1:044 þ 0:0160 b¼   H O H þ 0:0531 À 0:34493 C C C O À 0:4124 C (19) The producer gas from the gasifier possesses chemical exergy (ech B ) in addition to the thermomechanical exergy (eB, which is due to the elevated pressure and temperature of the gas and the mixing of the constituents) The specific chemical exergy of the producer gas is given by ech B ¼ X1 ech X1 þ X þX3 þ X4 þ X5 þ 3:76XO H2 þ X2 ech X1 þ X2 þ X3 þ X4 þ X þ3:76XO CO þ X5 ech X1 þ X2 þ X3 þ X4 þ X5 þ 3:76XO CH4 (20) ch ch where, ech H2 eCO and eCH4 represent the specific chemical exergy of H2, CO and CH4, respectively [22] 2.4 Performance parameters 2.3 Exergy analysis of the cycle Since the power cycle involves the gasifier and the combustor, both the thermomechanical exergy and chemical exergy are Finally, the cycle performance parameters have been evaluated based on one kmole of dry biomass fed to the plant The actual work done on the compressor is expressed as, 346 A Datta et al / Energy 35 (2010) 341–350 WC ¼ 4:76XO0 ZT2 Cpair dT þ 4:76XO T1 ZTA Table Parameters for the analysis of EFGT cycle in the present work Cpair dT; (21) Biomass Analysis (by mass) on dry basis [17] T1 Carbon Hydrogen Oxygen Calorific Value While the actual work done by the turbine is WT ¼ 4:76XO0 ZT3 Cpair dT Moisture content in the biomass by mass (22) Properties of Ambient Air Pressure Temperature Composition (by vol.) Nitrogen Oxygen T4 The thermal efficiency (hth) of the EFGT cycle is obtained using the turbine and compressor work and the calorific value of the fuel The energy delivered with the exhaust gas from the cycle, which can be subsequently recovered as waste heat in a downstream process is EN ¼ X Xj j ZT6 Cpj dT Equipment performance Isentropic efficiency of compressor (hc,isen) Isentropic efficiency of turbine (ht,isen) Pressure drop across the gasifier (DPG) Pressure drop at heat exchanger c Þ as % of inlet pressure cold side ðDPHE Pressure drop at heat exchanger h Þ as % of inlet pressure hot side ðDPHE Pressure drop across combustor (DPCC) as % of inlet pressure (23) To where, Xj is the number of moles of the jth species on the product side of Eq (12) and Cpj is the respective specific heat The exergy input into the plant (ein) for every mole of biomass þ 4:76fXO þ XO0 ge1 Þ A part of the fed to the cycle is given as ðech wood input exergy is actually converted into useful work, while the other parts are lost with the exhaust gas and are destroyed due to the irreversibilities in different components of the plant The useful exergy and the exergy lost as fractions of the input exergy are ðWT ÀWC Þ ein and EN ein respectively The former also represents the exer- getic efficiency of the cycle In addition to these, exergy has been destroyed in each of the components of the cycle due to process irreversibilities The expressions of exergy destruction in the individual components of the cycle are presented in Table In addition to the exergy destruction, the expressions of the exergetic efficiency of the individual components are also evaluated as indicators of their deviation from ideality, while operating between the corresponding thermodynamic states The expressions of exergetic efficiency of the individual plant components are also shown in Table Results and discussion 3.1 Influence of the key operating parameters on cycle performance The integrated model has been used to evaluate the performance of an EFGT cycle at different operating conditions A performance comparison is eventually made with reference to a 100 kW unit for distributed power generation Simple operation Operating parameters with range Equivalence ratio at the gasifier (4) Compressor Pressure ratio (rp) Turbine inlet temperature (TIT) Heat exchanger cold end temperature difference (CETD) 50% 6% 44% 449568 kJ/kmol (18732 kJ/kg) 20% 101.325 kPa 300 K 79% 21% 87% 89% 16 mm Hg column 1.5 0.5 3.33 2–8 1050–1350 K 200–300 K and low cost are the two key factors in choosing the plant operating parameters for distributed generation in remote areas In this effort, both the thermal performance and sizing of the plant are taken into account The former is represented by the thermal efficiency of the plant and is an indicator of the operating cost (fuel cost) for a particular plant capacity The plant size is compared on the basis of specific air flow (i.e air flow per unit energy output) through the turbine Lower value of the specific air flow indicates smaller size of the plant equipment and lower capital cost Table summarizes the operating parameters based on which the performance analysis has been performed here The estimated producer gas temperature from the gasifier at the corresponding conditions is 1006 K The influence of three salient operating parameters, viz., the pressure ratio of the cycle (rp), turbine inlet temperature (TIT) and the heat exchanger cold end temperature difference (CETD) are the three critical operating parameters, whose influence on the cycle performance are investigated Fig shows the variation in the cycle thermal efficiency with the pressure ratio at three different turbine inlet temperatures, viz 1050 K, 1200 K and 1350 K Still higher turbine inlet temperature is Table Exergy destruction and second law efficiency of individual component of the EFGT plant Component Exergy Destruction Exergetic efficiency Compressor 4:76½ðXO þ XO0 Þe1 À XO eA À XO0 e2 ŠþWC 4:76½XO eA þ XO0 e2 À ðXO þ XO0 Þe1 Š WC Gasifier jwood þ 4:76XO eA À XB ðech B þ eB Þ; where, XB ¼ X1 þ X2 þ X3 þ X4 þ X5 þ 3.76XO Turbine 4.76X O (e3 À e4)ÀWT Combustor 4:76XO0 e4 þ XB ðech B þ eB Þ À Xg e5 Heat Exchanger Xg (e5 À e6) þ 4.76X O (e2 À e3) XB ðech B þ eB Þ ech þ 4:76XO eA wood WT 4:76XO0 ðe3 À e4 Þ Xg e5 4:76XO0 e4 þ XB ðech B þ eB Þ Xg e6 þ 4:76XO0 e3 Xg e5 þ 4:76XO0 e2 A Datta et al / Energy 35 (2010) 341–350 a 0.5 TIT=1050 K 50 TIT=1050 K Specific air flow by mass (kg/kWh) TIT=1200 K 0.4 347 TIT=1350 K 0.3 0.2 0.1 TIT=1200 K 40 TIT=1350 K 30 20 10 0 10 possible in today’s gas turbine technology [26] However, it requires expensive turbine materials and extensive cooling arrangement for the turbine blades, thereby increasing the capital cost as well as the complexity of operation It is observed that for a particular turbine inlet temperature, the efficiency first increases with the increase in pressure ratio to attain a maximum value and then decreases with further increase in the pressure ratio On the other hand, higher turbine inlet temperature ensures higher thermal efficiency at all pressure ratios As a result, the efficiency peaks of 24.3%, 29.7% and 34.4% are obtained for TITs of 1050 K, 1200 K and 1350 K, respectively The maximum efficiency is reached in the pressure ratio range of 3–4 for the three TITs considered here The variation in the pressure ratio influences the specific air consumption in the cycle and therefore the size of the plant components The pressure at the inlet to the turbine is different at different pressure ratios, while at the exit of the turbine the pressure remains the same for all the cases At the high pressure end of the turbine the size at different conditions are compared using the specific air consumption by volume, while for the low pressure end the specific consumption by mass determines the size Also, an increase in the pressure increases the metal thickness of the equipment casing walls, increasing their weight and cost Fig 3a shows the variation of the specific air flow by mass (kg/kWh) entering the turbine against the pressure ratio (rp) at different turbine inlet temperature It is observed that at a particular turbine inlet temperature the specific air mass flow first decreases with the increase in the pressure ratio The decrease in mass flow is found to be rapid at the lower end of pressure ratio and gradually decreases as the pressure ratio is increased Beyond a pressure ratio value the mass flow begins to increase with further increase in pressure ratio The pressure ratio at which the reversal in the trend of mass flow variation occurs is lower at lower value of turbine inlet temperature (for TIT ¼ 1050 K the reversal occurs at rp ¼ 7.0, while for TIT ¼ 1200 K and 1350 K this reversal is not observed till rp ¼ 8.0) It is also important to note from Fig 3a that the increase in the turbine inlet temperature decreases the specific air consumption at the turbine inlet The variation in the specific air consumption by volume (m3/ kWh) at the inlet to the turbine with changing pressure ratio, at constant turbine inlet temperature, is shown in Fig 3b The decrease in the specific air consumption by volume is monotonic in this case with the increase in rp While the decrease is very rapid at the lower end of the pressure ratio range, the incremental b 70 Specific air flow by volume (m3/kWh) Fig Variation of thermal efficiency (h) of the EFGT cycle with the pressure ratio (rp) at different turbine inlet temperatures (TIT) 10 rp rp 60 TIT=1050 K TIT=1200 K TIT=1350 K 50 40 30 20 10 0 10 rp Fig (a) Variation of specific air flow by mass (kg/kWh) with pressure ratio (rp) for the EFGT cycle at different turbine inlet temperatures (TIT) (b) Variation of specific air flow by volume (m3/kWh) with pressure ratio (rp) for the EFGT cycle at different turbine inlet temperatures (TIT) change becomes less at higher rp The specific volume flow of air at the turbine inlet is guided by the pressure and the specific mass flow rate at a given TIT At low values of rp the pressure remains low and the specific mass flow is high, both contributing to the fact that in the small rp regime a reduction in pressure ratio sharply increases the specific volume flow of air At higher values of rp, the changes are much flatter since the specific mass flow curves are nearly flat, and the pressure is high For TIT ¼ 1050 K and 1200 K, although the mass flow of air observes a gradual increase with the increase in rp, such an increase is masked the effect of increasing pressure, and the volume flow continues to decrease (though only at a slow rate) Therefore, as observed from Figs 3a and b, the specific mass and volume flow rates of air are high at low values of rp (e.g at rp ¼ 2.0) Both the values decrease rapidly till rp increases to about 5.0 Further increase in rp levels-off the specific mass flow of air and leads to a marginal decrease in the specific volume flow, but the increased pressure warrants thicker walls for the high pressure components of the cycle Therefore, though there may be a marginal advantage in the reduction in volume flow rate (and hence the plant size) beyond a particular pressure ratio, the higher wall thickness will offset the cost benefit 348 A Datta et al / Energy 35 (2010) 341–350 Table Performance parameters of 100 kW Biomass fired EFGT plant at different operating conditions 0.4 0.3 0.2 CETD=200 K 0.1 CETD=250 K CETD=300 K 0 10 rp Fig Variation of thermal efficiency (h) of the EFGT cycle with the pressure ratio (rp) at different heat exchanger cold end temperature difference (CETD) The heat exchanger is one of the most critical equipment in the EFGT cycle Considering the cost of the material for the high temperature heat exchanger its size requires to be optimized However, the design of the heat exchanger also influences the thermal performance of the power cycle, by influencing the exhaust gas loss from the cycle Fig shows the variation in the cycle thermal efficiency with pressure ratio at different cold end temperature difference (CETD) of the heat exchanger for a particular turbine inlet temperature The results show the same trend in the variation of efficiency with pressure ratio at all the CETD values, with the maximum efficiency reached at an optimum pressure ratio The optimum pressure ratio is found to be 4.0 for the three different CETD values of 200 K, 250 K and 300 K considered However, with the increase in the CETD at a particular pressure ratio, the efficiency is found to decrease When the CETD is high more amount of the energy is wasted through the exhaust gas stream, reducing the net work produced in the cycle In fact for particular rp and TIT, the state points 2, and shown in Fig not change with the variation of CETD However, the variation in the temperatures across the heat exchanger changes the quantity of air flow governed by the energy balance across the heat exchanger It is observed that the number of moles of air flowing through the turbine per unit mole of dry biomass feed (XO0 ) decreases with the increase in the CETD The variation in CETD does not change the specific air flow rate through the turbine as the corresponding state points remain identical Based on the above discussion, it can be proclaimed that the cycle thermal efficiency is maximized in the rp range of 3–4, depending on the TIT and CETD At the low pressure ratio of 2–3, the size of the equipment will be large because of the high value of the specific air flow Conversely, a high pressure ratio increases the wall thickness of the equipment, thereby increasing the cost and weight Considering all these facts, we have chosen rp ¼ 4.0 as the optimum value of the pressure ratio for the EFGT cycle Two different turbine inlet temperatures (1200 K and 1350 K) and two different CETD values (200 K and 300 K) are chosen to compare the performance Accordingly, three sets of cycle operating conditions with different turbine inlet temperatures (TIT) and heat exchanger cold end temperature differences (CETD) have been identified (as given in Table 4) to compare the performance of a 100 kW EFGT plant 3.2 Energy and exergy based analysis of a 100 kW biomass fired EFGT plant Table lists the important performance parameters for the three cases for a 100 kW EFGT based micro gas turbine plant running at full Fuel (biomass) flow rate, kg/s Air flow rate, kg/s Thermal Efficiency, % Exhaust heat, kW Rate of heat exchange across heat exchanger, kW Heat Exchanger hot end temperature difference, K LMTD, K (UAHE)overall for heat exchanger, W/K Case 1: rp ¼ 4, TIT ¼ 1200 K CETD ¼ 200 K Case 2: rp ¼ 4, TIT ¼ 1350 K CETD ¼ 200 K Case 3: rp ¼ 4, TIT ¼ 1200 K CETD ¼ 300 K 0.0216 0.0186 0.0265 0.5867 29.68 238.92 450.5 0.4712 34.33 193.09 440.03 0.5942 24.18 315.65 450.72 80.0 44.2 147.0 131.0 3.44 103.2 4.26 214.5 2.10 load It is seen from Table that when the TIT is 1350 K and CETD is 200 K (Case 2), the thermal efficiency of the plant attains the highest value of 34.33% Accordingly, the fuel flow rate and the exhaust heat loss are the lowest The air flow rate is also the lowest for this case, indicating smaller size of the components, like compressor and turbine On the other hand, the logarithmic mean temperature difference (LMTD) of the heat exchanger based on the temperature differences at the hot and cold ends is low, giving a high overall (UAHE) value for the heat exchanger, where U and AHE are the overall heat transfer coefficient and the heat transfer surface area of the heat exchanger, respectively If we consider a nearly constant value of the overall heat transfer coefficient (U) for all the cases, then case performance data calls for the largest size of the heat exchanger The higher turbine inlet temperature and the increased size of the heat exchanger required for this case is indicative of a high cost of the plant In case when the TIT is 1200 K and the CETD 300 K (Case3) the thermal efficiency of the plant is the lowest (24.18%) The fuel flow rate and the exhaust heat loss are the maximum in this case The corresponding air flow rate is also the highest among the three sets compared indicating a larger size of the turbine and compressor While the heat exchanger LMTD for this case is high (214.5 K) indicating a smaller sized heat exchanger The operating parameters in case offer a performance tradeoff in terms of thermal efficiency and the heat exchanger size A thermal efficiency of 29.68% has been achieved in this case The heat exchanger LMTD is 131 K giving overall UA as 3.44 Therefore, considering the capital and operating cost of the plant, case is the better choice of plant operating condition A second law based performance analysis for the three cases has been presented in Fig 5, where the complete exergy balance has been made as fractions of the exergy input to the cycle The fraction of the input exergy converted into useful work determines the exergetic efficiency of the cycle The remaining part of the input exergy is either lost in the exhaust heat or destroyed through irreversibilities in various components It is observed from the results of the three cases that the maximum exergetic efficiency is attained in case 2, where the turbine inlet temperature is the highest The exergy loss in the exhaust is the highest in case 3, where the exhaust gas leaves the cycle at the maximum temperature (because of the highest CETD) Table shows a comparison of the exhaust heat for the three cases The major exergy destruction takes place in the gasifier, combustor and the heat exchanger, while the exergy destruction A Datta et al / Energy 35 (2010) 341–350 Case 1: rp=4, TIT=1200 K, CETD= 200 K 349 Case 2: rp =4, TIT=1350 K, CETD=200 K Heat Exchanger 8.45% Heat Exchanger 8.94% Useful 28.01% Useful 32.40% Combustor 17.57% Combustor 19.60% Turbine 2.11% Turbine 2.28% Gasifier 15.39% Gasifier 15.39% Exergy out 23.54% Compressor 2.07% Compressor 2.24% Exergy out 22.01% Case 3: rp =4, TIT=1200 K, CETD= 300 K Heat Exchanger 10.08% Useful 22.82% Combustor 18.98% Turbine 1.85% Exergy out 29.05% Gasifier 15.39% Compressor 1.83% Fig Exergy balance of the EFGT cycle for the three different cases described in Table of the compressor and turbine are only a little The fraction of exergy destructed in the gasifier is the same in the three cases, since the operating parameters of the gasifier has been considered identical A sizeable amount (15.39%) of the input exergy is destructed in the gasifier owing to the gasification reactions that take place there The exergy destruction in the combustor is the highest in all the three cases, amounting to 19.6%, 17.57% and 18.98% of the input exergy, respectively The destruction of exergy in the combustion chamber is due to the heat exchange between the streams and chemical reactions that take place Operating the combustor at higher temperature and higher temperature of the air fed to the combustion chamber decrease the exergy destruction in the combustor Exergy destruction in the heat exchanger increases when the temperature difference between the two streams exchanging heat increases Accordingly, the maximum fraction of the exergy destruction in heat exchanger occurs in case 3, where the LMTD is also the highest More than 10% of the input exergy is destroyed in the heat exchanger for this case For the other two cases (i.e case and case 2) the exergy destroyed in the heat exchanger are 8.94% and 8.45% of the input exergy respectively Fig shows the exergetic efficiencies for the individual components for the three cases The individual exergetic efficiency value of the equipment indicates the deviation from ideality for the equipment operating across its respective thermodynamic states It is observed that the exergetic efficiency of the compressor, turbine and heat exchanger remain above 90%, while those of the gasifier and the combustion chamber are less The relatively lower exergetic efficiency in the gasifier and the combustion chamber is attributed to the irreversibility pertaining to the chemical reactions occurring there The exergetic efficiency of the compressor is identical for all the three cases (91.5%), since it operates at same pressure ratio and isentropic efficiency Similarly, the exergy efficiencies of the gasifier are the same for the three cases as the operating pressure, gasifier equivalence ratio and the properties of the biomass are considered to be the same The exergetic efficiency for the turbine is the highest (96.4%) in case where the turbine operates with the highest inlet temperature For this condition, the air temperature at the turbine outlet also remains higher than the other conditions As a result, the combustion chamber operates with the maximum air preheat in case The flame temperature in the combustor also becomes the maximum in this case As the chemical reaction occurs at high temperature the associated irreversibility becomes less and the combustion chamber exergetic efficiency attains the maximum value for case The exergetic efficiency of the heat exchanger largely depends on the mean temperature difference between the streams across the exchanger Lower mean temperature difference is indicative of lower irreversibilities This is evident in the result as the heat exchanger in case (the case with lowest LMTD among the three) shows the highest exergetic efficiency 350 A Datta et al / Energy 35 (2010) 341–350 1.0 performance of a 100 kW plant running on EFGT cycle The thermal performance and sizing have been compared based on the thermal efficiency, air flow rate and heat transfer area of the heat exchanger Moreover, an exergy balance has been carried out for each of the cases to account the useful exergy, exergy loss and exergy destruction Major exergy destruction is found to occur in the gasifier, combustor and the heat exchanger Though the parameters in Case (TIT ¼ 1350 K, CETD ¼ 200 K) offer a higher thermal efficiency and exergetic efficiency and a lower air flow rate, the heat exchanger size for this case is found to be large On the other hand, the heat exchanger size for the Case (TIT ¼ 1200 K, CETD ¼ 300 K) is small, but it gives the lowest thermal and exergetic efficiencies A trade-off in performance is observed for Case (TIT ¼ 1200 K, CETD ¼ 200 K) 0.8 ε 0.6 0.4 References 0.2 0.0 r so s pre m Co r ine ifie s Ga rb Tu Co rp=4, TIT=1200 K, CETD=200 K r sto u mb at He Ex er ng a ch rp=4, TIT=1350 K, CETD=200 K rp=4, TIT=1200 K, CETD=300 K Fig Exergetic efficiency of individual components in the EFGT cycle for the three different cases described in Table 4 Conclusions A thermodynamic analysis has been performed for an externally fired gas turbine (EFGT) cycle with an integrated biomass gasifier The effects of operating parameters like pressure ratio (rp), turbine inlet temperature (TIT) and cold end temperature difference (CETD) of the heat exchanger on the thermal efficiency and specific air flow have been studied The thermal efficiency of the cycle is found to be within 16–34% for the range of operating parameters under investigation The cycle thermal efficiency is the maximum at an optimum pressure ratio of the cycle (in the range of 3–4) for a particular turbine inlet temperature and cold end temperature difference across the heat exchanger At a particular pressure ratio of the cycle the thermal efficiency increases either with the increase in the turbine inlet temperature or with the decrease in the cold end temperature difference of the heat exchanger The specific air flow at the turbine inlet is evaluated to compare the size of the plant equipment It is found that the specific air flow by volume decreases with the increase in pressure ratio sharply at the lower end of rp, while the incremental change is marginal at high values of rp The specific air flow by mass exhibits a rapid decrease with the increase in rp at the lower end of 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