Energy 35 (2010) 1084–1093
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Internal Combustion Engine (ICE) bottoming with Organic Rankine Cycles (ORCs) Iacopo Vaja*, Agostino Gambarotta University of Parma, Industrial Engineering Department (IED), via G. P. Usberti 180/A, 43100 Parma, Italy
a r t i c l e i n f o
a b s t r a c t
Article history: Received 24 October 2008 Received in revised form 2 June 2009 Accepted 12 June 2009 Available online 4 July 2009
This paper describes a speciﬁc thermodynamic analysis in order to efﬁciently match a vapour cycle to that of a stationary Internal Combustion Engine (ICE). Three different working ﬂuids are considered to represent the main classes of ﬂuids, with reference to the shape of the vapour lines in the T–s diagram: overhanging, nearly isoentropic and bell shaped. First a parametric analysis is conducted in order to determine optimal evaporating pressures for each ﬂuid. After which three different cycles setups are considered: a simple cycle with the use of only engine exhaust gases as a thermal source, a simple cycle with the use of exhaust gases and engine cooling water and a regenerated cycle. A second law analysis of the cycles is performed, with reference to the available heat sources. This is done in order to determine the best ﬂuid and cycle conﬁguration to be employed, the main parameters of the thermodynamic cycles and the overall efﬁciency of the combined power system. The analysis demonstrates that a 12% increase in the overall efﬁciency can be achieved with respect to the engine with no bottoming; nevertheless it has been observed that the Organic Rankine Cycles (ORCs) can recover only a small fraction of the heat released by the engine through the cooling water. Ó 2009 Elsevier Ltd. All rights reserved.
Keywords: Organic Rankine Cycle Internal Combustion Engine Combined cycle
1. Introduction Organic Rankine Cycles (ORCs) can use different working ﬂuids in order to exploit low grade heat sources to produce useful work. An interesting application of ORCs is to couple them with other prime movers and utilize their wasted heat, thus realizing a combined power unit with the effect of enhancing the overall system efﬁciency [1–4]. Since the ORC systems generate additional power without requiring extra fuel, the speciﬁc pollutant emissions of the combined plant are reduced. Organic ﬂuids are to be preferred to water when the required power is limited and the heat source temperature is low, as these ﬂuids often have lower heat of vaporization and can better follow the heat source to be cooled, thus reducing temperature differences and therefore irreversibilities at the evaporator. Furthermore, turbines for organic cycles can provide higher efﬁciencies at part loads as well and are usually less complex (1 or 2 stages, for an axial turbine) due to the lower enthalpy drop of the ﬂuid [5,6]. Also ORC systems exhibit great ﬂexibility, high safety and low maintenance
* Corresponding author. Tel.: þ39 0521 905863; fax: þ39 0521 905705. Email address:
[email protected] (I. Vaja). 03605442/$ – see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2009.06.001
requirements in recovering lowmedium grade waste heat, from industrial processes or power plants [7]. Evaporators for ORCs are usually simple components designed as a heat exchanger with direct use of hot gases released by the thermal source, often without use of intermediary ﬂuids, such as diathermic oils. They also have one level of evaporating pressure [1,4]. Recently many power units based on Internal Combustion Engines (ICEs) are being employed as base components in Distributed Energy Systems (DES), intended as systems where power is generated in small decentralized units. DES are usually appreciated because they could help to reduce emissions, save grid capacity, provide opportunities for renewable energy and increase overall energy generation and distribution efﬁciency [8]. ICEs are often chosen for their reliability, low speciﬁc cost and high electrical efﬁciency, especially in the power range of hundreds of kW to few MW. It may happen however that many of these engines are fuelled using biomass (such as vegetal oils, biogas or others) and the existence of incentives (such as Green Certiﬁcates in Italy) for plants based on Renewable Energy Sources makes the operation of these units a viable solution even if no heat is usefully recovered. Few examples of ORCs coupled to ICEs exist and usually the ORC is conceived to efﬁciently exploit only the heat released at high
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Nomenclature
F
H H_
Subscripts C Condenser P Pump R Regenerated/regenerator T Turbine a Available appr Approach cond Condenser crit Critical dead Dead state e Engine ex Heat exchanger f Fluid ﬁn Final g Gas in Inlet out Outlet w Water
P Q Q_ S S_ T V V_ cp h m _ m p s v
Enthalpy [kJ] Enthalpy ﬂow [kW] Power [kW] Heat [kJ] Heat ﬂow [kW] Entropy [kJ/K] Entropy ﬂow [kJ/s K] Temperature [K] Volume [m3] Volume ﬂow rate [m3/s] Speciﬁc heat at constant pressure [kJ/kg K] Speciﬁc enthalpy [kJ/kg] Mass [kg] Mass ﬂow rate [kg/s] Pressure [kPa] Speciﬁc Entropy [kJ/kg K] Speciﬁc volume [m3/kg]
Greek symbols Air fuel ratio [–] Heat exchanger effectiveness [–] Efﬁciency [–] Organic Rankine Cycle efﬁciency [–] Global efﬁciency [–]
a 3 h hORC hg
temperatures such as the engine exhaust gases. In [2,3], for example, a setup is proposed where a low speed two stroke marine turbocharged Diesel engine acts as the topper of a combined cycle with exhaust gases used for a bottoming cycle based on a Rankine cycle. The paper describes a full mathematical model built to evaluate the mutual operation of the Diesel engine as source of power and waste heat and a HRSG with a steam turbine, in the whole powerspeed spectrum of the Diesel engine. While very comprehensive and detailed, the papers focus on the issue of offdesign operation of the system, which is not so important for stationary engines that are always operated at rated conditions. Furthermore in the proposed design the vapour cycle is based on a conventional water–steam cycle due to the high thermal power available from the marine engine. In [9] the retroﬁt with two different ORC system rated about 10 kWe of a small engine with a power of 200 kWe, fuelled with biogas, is considered. The bottomer, based on R134a, in this case is thermally fed using only the engine cooling water. The paper focuses on the application of scroll expanders, that may be advantageous when the power of the ORC is limited. The feasibility of the system is supported by ﬁeld tests conducted on an experimental unit. The availability of waste heat at two different levels of temperature is usually a problem for the application of ORCs onto ICEs. In [16] for example, a minihybrid solar power plant is presented, characterized by a ﬁeld of solar concentrators and a bioDiesel engine integrated to two superposed ORCs. When the system operates in hybrid mode the heat released by the engine, in series with the heat generated by the solar system, is employed to power two simple ORC cycles, the ﬁrst based on R123 and the second on R134a. Besides the use of solar thermal power, the design proposes a possible solution when heat is available at different temperature levels, with the second ORC powered both by the heat released by the ﬁrst ORC and by the heat from the engine cooling network, that may be provided either in series as liquid preheater or parallel to the evaporator. Laboratory tests made with the superposed ORCs
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Heat availability [–]
Abbreviations CC Combined Cycle ICE Internal Combustion Engine ORC Organic Rankine Cycle PP Pinch Point
conﬁrmed adequate operational characteristics with good performances over a broad range of conditions, indicating the feasibility of the design. The assessment of double cascade ORC designs in conjunction to stationary ICEs has been considered [19] and will be presented in future works. The present paper is focused on evaluating the extra power achievable by using an ICE as a heat source for just one vapour power cycle, thus with a rather simpliﬁed design of the heat recovering unit, under the hypothesis of operating the engine at full load while considering different cycles based on different organic ﬂuids and conﬁgurations. 2. The systems considered in the investigation In the analysis a commercial cogeneration engine is considered as a topping system. The engine is a 12 cylinder 4 stroke supercharged natural gas ﬁred medium speed: the main parameters of the engine are reported in Table 1. From the engine data it has been evaluated that about 1700 kWt is available by cooling the exhaust gases down to 120 C (Tg,min) and about 1000 kWt is available from the engine cooling water. It has been calculated that the air fuel ratio a is 29.2 and the excess air coefﬁcient e is 0.701 at nominal conditions, under the hypothesis of perfect combustion of pure methane. The composition
Table 1 Main engine characteristics. Electrical power output Fuel consumption Rated electrical efﬁciency Engine speed Exhaust gas temperature Exhaust mass ﬂow Combustion air mass ﬂow Engine jacket temperatures Engine jacket ﬂow
2928 7002 41.8 1000 w470 15,673 15,154 79/90 w90
kW kW – min1 C kg/h kg/h C m3/h
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of the exhaust gases on the basis of mass has been calculated at: CO2 ¼ 9.1%, H2O ¼ 7.4%, N2 ¼ 74.2%, O 2¼ 9.3%. This composition is used to evaluate the gas properties. In the study, the engine is assumed to operate at rated conditions as the aim is to determine which organic cycle would better ﬁt the overall heat available for these conditions. Organic ﬂuids used for power applications may have different characteristics in the T–s diagram and the saturation lines may be bell shaped, nearly isentropic or overhanging depending on the ﬂuid molecule complexity. Typically ﬂuids with simpler molecules are characterized by bell shaped vapour lines and lower critical temperatures and ﬂuids with more complex molecules display a overhanging vapour line and higher critical temperatures [10]. In the paper three ﬂuids commonly adopted for technical applications were chosen to represent these different behaviours in the T–s diagram (Fig. 1): benzene (overhanging), R11 (isentropic) and R134a (bell shaped). They are commonly quoted in literature as possible ﬂuids for ORCs [11,12], even though future work would lead to consider other pure ﬂuids or ﬂuid mixtures [13]. It should be noted that in the proposed analysis only thermodynamic aspects will be considered. It is important to remark however that ﬂuids for ORC applications must not only be favourable from a thermodynamic point of view, but have to satisfy conditions such as [11,14]: 

chemical stability at the operating pressures and temperatures; environmental friendliness: low ozone depletion potential (ODP), global warming potential (GWP) and atmospheric lifetime (ALT); non toxic, non corrosive and compatible with engine materials; low ﬂammability and autoignition properties.
In the analysis that follows three cycle conﬁgurations will be assumed, in order to compare not only the ﬂuids considered but also the different cycle setups and to verify the most feasible application of ORCs coupled with ICEs. The setups assumed are: 


ORC simple cycle thermally powered by engine exhaust gases (Fig. 2a); ORC simple cycle thermally powered by engine exhaust gases and engine refrigerant water (Fig. 2b); regenerated ORC thermally powered by engine exhaust gases (Fig. 2c).
Fig. 2. Scheme of different ORC cycle conﬁgurations assumed for ICE bottoming.
For all the cycles considered the following assumptions are valid: 

Fig. 1. Different shapes of coexistence curves in the T–s diagram for the considered ﬂuids.

isentropic turbine efﬁciency: hT ¼ 0.7 (this value is conservatively below usual values that may range between 0.8 and 0.88 [6,14,15]); isentropic pump efﬁciency: hP ¼ 0.8; working ﬂuid temperature at condensation: Tcond ¼ 308 K; vaporizing pressure varying between condensation pressure pcond and critical pressure pcrit; negligible pressure losses in the heat exchangers and pipes; dry expansion for all ﬂuids;
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The last hypothesis has been introduced assuming that a dry expansion would be preferable for the preservation of the expander. This is to eliminate the impingement of liquid droplets on turbine blades. Therefore a slight superheating will be assumed when the isentropic expansion line crosses the vapour line. In other words, it will be considered just the minimum superheating to keep an isentropic expansion totally in the dry zone of the diagram. Introducing high degrees of superheating, however, is not convenient from a thermodynamic point of view, as the cycle efﬁciency is a weak function of turbine inlet temperature once the evaporating pressure is chosen. In some cases, the highest cycle efﬁciencies are obtained when superheating is avoided and the ﬂuid is expanded directly from dew line, as demonstrated in [12]. In [15] it has been also shown, through a comprehensive analysis based on irreversibility calculations, that superheating organic cycles (especially if based on overhanging ﬂuids) increases cycle irreversibility and decreases the second law efﬁciency. As further assumption the extra costs of building the ORC when the turbine inlet pressure is raised have been neglected [12]. The turbine pressure therefore has been assumed as an independent variable of the problem, the only limitation being the ﬁeld of ﬂuid stability. Fluid properties have been evaluated using the REFPROPÒ database [16] that has been utilized in MatlabÒ. Speciﬁc MatlabÒ functions were deﬁned and compiled to evaluate cycle properties and performance in a parameterized way. This is in order to assess the effects of changes in one or more variables (as, for example, the pressure at the evaporator inlet p2) on the main cycle characteristics.
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It is to be noted that in Eq. (1) the organic ﬂuid enthalpies are function of the evaporating pressure chosen for the ORC and cp;g is calculated with REFPROPÒ according to the exhaust gas composition at the average temperature. A second energy balance allows to determine the exhaust gas temperature at the evaporator outlet:
Tg;fin ¼ Tg;PP
_ f ;1 ðh2 h1 Þ m _ g cp;g m
(2)
If Tg,ﬁn calculated with Eq. (2) is lower than the minimum allowed temperature for the gases, a procedure decreases the _ f ;2 is the organic ﬂuid mass ﬂow rate while Tg,ﬁn is above Tg,min. m ﬂuid mass ﬂow rate that satisﬁes the imposed condition and represents the new ﬂuid mass ﬂow rate for the cycle. The actual gas Pinch Point temperature can be calculated using:
Tg;PP ¼ Tg;out
_ f ;2 ðh30 h2 Þ m _ g cp;g m
(3)
All the organic ﬂuids (especially at relatively high pressures) considered in the analysis have a speciﬁc heat of vaporization lower in comparison to that required to warm up the ﬂuid between point 1 and 2 of Fig. 3. Therefore Tg,ﬁn calculated with Eq. (2) is always smaller than Tg,min,. This means that there are no Pinch Point limitations in the heat exchange process and that which limits the amount of heat introduced to the cycle is the need to avoid overcooling of the engine exhaust gases. The global efﬁciency of the system can be deﬁned as the net power produced by the cycle referred to the total available heat Q_ a
[17]: 3. The ORC simple cycle
P
In Fig. 3 the heat exchange diagram of the evaporator is proposed for a generic ORC simple cycle with no superheating. Heat needed to vaporize the organic working ﬂuid is provided in this case solely by the engine exhaust gases (according to Fig. 2a). The _ g ) are gas temperature at engine exhaust (Tg,out) and ﬂow rate (m deﬁned and constant. A minimum gas pinch point temperature (Tg,PP) is considered (Tg,PP ¼ T2 þ DTPP, where DTPP ¼ 30 K is the minimum temperature difference at Pinch Point to meet the gas/ ﬂuid heat exchanger performances) and this allows writing a ﬁrst energy balance referred to complete ﬂuid vaporization:
_ f ;1 ¼ m
_ g cp;g Tg;out Tg;PP m h30 h2
(1)
Fig. 3. T—Q_ diagram of the evaporator for a ORC heated with engine exhaust gases.
hg ¼ ORC ¼ 3$hORC Q_ a
(4)
where 3 is the evaporator effectiveness [18]:
cpg Tg;out Tg;fin 3¼ cpg Tg;out T1
(5)
and hORC the organic Rankine cycle efﬁciency:
hORC ¼
PORC PORC ¼ h3 h1 cpg Tg;out Tg;fin
(6)
From Eq. (4) it can be observed that not always can the maximum global efﬁciency be achieved with maximum cycle efﬁciency (hORC). It could happen that with increasing turbine inlet pressure the cycle efﬁciency increases but the ﬁnal gas temperature is also increased, thus determining a decrease in the evaporator effectiveness 3. Combination of these two terms might provide that the turbine inlet pressure that maximizes hORC is not the one that maximizes hg [1,18]. The analysis that follows therefore is a result of determining the optimal pressure at turbine inlet that would maximize hg and in turn, provide the maximum ORC power at a given value of Q_ a . In Fig. 4 the cycle efﬁciency is plotted for the reference ﬂuids in the range of turbine inlet pressures between pcond and pcrit (where pcrit is the critical pressure of each ﬂuid). As expected benzene displays higher achievable efﬁciencies. The curves are monotonic for all ﬂuids (solid lines): a consequence of the hypothesis from introducing a minor degree of superheating while the isentropic expansion is not completely dry. For higher pressures, the introduced superheating becomes higher and this slightly rises the cycle efﬁciency. The corresponding dotted curves are plotted for the saturated Rankine cycle (with no superheating) and in this case the
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Fig. 4. Simple cycle efﬁciencies for evaporation pressures between pcond and pcrit.
Fig. 6. Relative variation of net cycle power output for simple Rankine cycles at different evaporation pressures.
curves show a maximum value of the cycle efﬁciencies for pressures not far from critical: in particular a maximum hT value of 0.2146 is achieved with benzene at a pressure of 4470 kPa. _ f ðh30 h4 Þ, The curves in Fig. 5 refer to the net power (PORC ¼ m see Fig. 9 for symbols) achievable from the cycles at different pressures of vaporization. The shape of the plotted curves is similar to that of the efﬁciency curves (Fig. 4). The curves for all ﬂuids have their maximum value at the critical pressure as a consequence of pattern of the efﬁciency curves. In the analysis an optimal value of p2 will be chosen (indicated by the * symbol in the ﬁgures) as the one that maximizes the efﬁciency curves referred to saturated Rankine cycle (dotted lines of Fig. 4). Therefore the optimum pressure is chosen as 4470 kPa for benzene which returns a net power output of 376 kW, 3835 kPa for R11, with 290 kW and 3723 kPa for R134a with a net power output of 147 kW. In Fig. 6 the dimensionless cycle power output is plotted in the assumed pressure range with reference to the cycle power output, determined at the optimal pressure for each ﬂuid (* of Fig. 5). It can be noted that not only is benzene the ﬂuid with the highest power output value, but it also shows the smallest variations from the optimal power. For example at pressures of 1700 kPa, a cycle based on benzene provides about 90% of the optimal power output. This characteristic leads to consider cycles with lower expansion ratios, while simultaneously simplifying the compression and expansion phases. Fluids with bell shaped vapour lines display instead higher variability in power output with respect to vaporization pressure.
This would result in the Rankine cycle being operated at a pressure as close as possible to the optimal pressure for these ﬂuids. In Fig. 7 the estimated value of working ﬂuid mass ﬂow rates are plotted again with respect to the pressure at turbine inlet. It is possible to observe that benzene requires the lowest ﬂuid mass ﬂow rate as consequence of the highest enthalpy increase between state 3 and 1. The energy balance at the evaporator determines higher masses of ﬂuid for R123 and R134a in order to match the total energy of the ﬂue gasses. Fig. 8 reports, for each ﬂuid, the curves referring to actual volumetric ﬂow rate at the expander inlet ðV_ 3 Þ and the turbine outlet/inlet volume ﬂow ratio (v4/v3). The latter parameter is particularly signiﬁcant as it shows how much the ﬂuid volume increases through the expansion. Fig. 8a refers to benzene and shows that a high value of the ratio v4/v3 of 374 is produced when the cycle is operated at the optimal pressure of 4470 kPa (*), and V_ 3 ¼ 0:015 m3 =s. Considerations regarding the power curve for benzene (Figs. 5 and 6), suggest that a lower evaporating pressure would allow lower turbine outlet/inlet volume ﬂow ratios, while providing a net power output close to optimal. In this case, the cycle would be better operated if a lower pressure was chosen since a simpler expander could be employed. As a reference for this analysis a new optimal value of evaporating pressure for benzene is selected at 2000 kPa (marked with  in Fig. 8a) even if further and more precise considerations would require matching with an actual
Fig. 5. Simple cycle power output at different evaporation pressures.
Fig. 7. Simple cycle working ﬂuid mass ﬂow rate required at different evaporation pressures.
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The graphs plotted in Fig. 9 show the thermodynamic cycle on the T—S_ diagram for the three considered ﬂuids. The curve referring to the engine exhaust gases is superimposed on the diagrams by properly scaling the entropy axes. It is evident, the lower the critical temperature the greater the temperature difference between gases and organic ﬂuid in the heater/evaporator. This temperature difference induces irreversibilities that are the main cause for low thermodynamic efﬁciencies with R11 and R134a. Future works will include a detailed exergy analysis where the causes of exergy reduction within each component of the system will be properly quantiﬁed and discussed [19]. The proposed representation however allows to observe the entropy generation rate due to the heat transfer process between hot gases and organic ﬂuid:
_ f ðs30 s1 Þ m _ g sg;out sg;fin S_ ex ¼ m
(7)
This value, that can be gathered from Fig. 9a–c, is, as expected, the bigger the higher average temperature difference in the heat exchanging process. Particularly S_ is 0.924, 1.234 and 1.842 kJ/s K for benzene, R11 and R134a respectively. Table 2 reports the main parameters for the three cycles proposed (the values for benzene have been estimated at vaporizing pressure of 2000 kPa). It can be observed that a maximum power output of about 350 kW can be achieved from a benzene based simple cycle (hORC ¼ 0.2) while the power decreases to about 190 kW 150 kW with R11 and R134a respectively (hORC ¼ 0.17 and 0.08). It should be noted that in all cases the heat ﬂow provided to the cycles with the engine exhaust gases is the same as they are cooled down to the minimum allowed temperature Tg,min (there are no Pinch Point limitations). 4. ORC with use of heat available from engine refrigerant A further comparison of the selected cycles has been carried out under the hypothesis of employing the engine cooling water for a partial preheat of the organic ﬂuids upstream the main evaporator (Fig. 2b). The thermodynamic cycles are the same as deﬁned in the previous section (Fig. 9) and the cycle efﬁciency is unchanged. A new heat source is now considered for preheating the working ﬂuid. In the analysis a heat exchanger with a reasonable efﬁciency has been assumed for the water/ﬂuid heat exchange process, with a temperature difference of 20 K between water inlet and organic ﬂuid outlet (Tw,PP). The energy balance at the water/ﬂuid heat exchanger gives the _ f 1 is the ﬂuid mass ﬁrst value of the organic ﬂuid mass ﬂow rate: m ﬂow rate that allows complete exploitation of energy available from the cooling water: Fig. 8. Volume ﬂow rate at turbine inlet ðV_ 3 Þ and turbine outlet/inlet volume ﬂow ratio (v4/v3) for different ﬂuids at different evaporation pressures.
commercial expander. At this new value of evaporating pressure the net power output from the cycle becomes 349 kW, the ratio v4/ v3 decreases to 107 while V_ 3 increases to 0.052 m3/s (Table 2). On the other hand, R11 and R134 show lower turbine outlet/inlet volume ﬂow ratios at the chosen evaporating pressure, suggesting that a simple expander could be employed with these ﬂuids even at optimal pressures. It should be noted that the ratio v4/v3 can change signiﬁcantly depending on the characteristics of the working ﬂuid. Several ﬂuids can achieve values up to 550 and when v4/v3 is smaller than 50, expansion efﬁciencies higher than 0.8 can be achieved via a single stage axial turbine [1].
_ f ;1 ¼ m
_ w cp;w Tw;out Tw;in m hA h1
(8)
A new energy balance should be written for the evaporator to match the available energy from the engine exhaust gases and the energy required to fully vaporize and superheat the working ﬂuid (if superheating is considered):
_ f ;2 ¼ m
_ g cp;g Tg;out Tg;PP m h30 h2
(9)
_ f 1 is apparently higher than m _ f 2 . This For the cycles analyzed, m condition would in fact not allow complete exploitation of energy available from the cooling water. A further energy balance would deﬁne the ﬁnal temperature of _ f ;1 ; m _ f ;2 Þ it gives: _ f ;3 ¼ minðm the gases leaving the exchanger. If m
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Table 2 Comparison of ORC cycles for benzene, R11 and R134a.
Benzene R11 R134a
PORC [kW]
hORC [–]
pcond [kPa]
pvap [kPa]
Tvap [K]
_ f [kg/s] m
V_ 3 [m3/s]
v4/v3 [–]
Dh30 4 [kJ/kg]
Dh01 [kJ/kg]
349.3 290.3 147.5
0.1986 0.1658 0.0852
19.6 147.9 883.3
2000 3835.9 3723.4
494.5 461 369.9
2.737 7.487 8.9667
0.052 0.030 0.041
107 32 5
130.5 41.9 19.4
0.286 0.316 0.302
Tg;fin ¼ Tg;PP
_ f ;3 ðh2 hA Þ m _ g cp;g m
(10)
Whenever Tg,ﬁn should be smaller than the rated ﬁnal gas temperature (Tg,min), the procedure implemented decreases the ﬂuid mass ﬂow rate in order to match this condition. Fig. 10 describes the T—Q_ diagram for the three cycles that have been considered. The heat rate entering the cycle, 0 Q_ in ¼ Q_ g þ DH_ 1A , is represented by the sum of the heat rate from hot exhaust gases ðQ_ g Þ and the heat rate from the cooling water actually used to preheat the working ﬂuid ðDH_ 1A Þ. It is to be noted (see Table 3 and Fig. 10) that the heat rate introduced into the cycles is not the same for all ﬂuids but it increases with decreasing critical temperature of the working ﬂuid. For R134a the heat rate introduced into the cycle from the engine cooling water is about 26.1% of 0 the total heat to the cycle ðDH_ 1A =Q_ in Þ and 57.9% of the total heat _ available from the cooling water ðDH 1A =Q_ w Þ. The same parameters has lower values for R11 (12.7% and 24.2%) and benzene (9.5% and 17.5%). The increased heat rate transferred to working ﬂuid leads to a signiﬁcant increase in net power output of the cycle for ﬂuids with low critical temperature (since the net efﬁciency of the cycle is unchanged). Table 3 shows that R134a allows a net increase in power output, calculated with respect to the cycle without heat 0 PORC Þ=PORC Þ, of 34.8%, while recovery from coolant water ððPORC a cycle with benzene and water preheating increases its power output only by about 10%. R11 displays intermediate characteristics with an increase in power output of 14.5%. Under these conditions R11 allows to reach mechanical power outputs close to that of benzene (332.5 vs. 386 kW). The new system layout therefore brings less advantages for cycles designed to utilize ﬂuids with high critical temperatures. Fluids with low critical temperatures and bell shaped vapour lines can signiﬁcantly increase their capability to produce useful work, but the net power output still remains lower than that achievable by overhanging ﬂuids like benzene even without heat recovery from cooling water. 5. ORC regenerated cycle ORCs modules available in commerce often utilize working ﬂuids with overhanging vapour lines. This condition is favourable as it allows dry expansions without superheating and regenerate the cycle by subcooling the vapour at the end of expansion (direct regeneration) without vapour extraction. For the cycles considered in the present work it is apparent that only benzene is suitable for direct regeneration. It is considered that the recuperator is a counterﬂow type heat exchanger and requires a DTappr,R of 15 K. Under these hypothesis the temperature of the vapour at the heat exchanger outlet can be evaluated, as well as the available energy for the internal heat exchange process ðQ_ R Þ:
Fig. 9. Optimal organic cycles and engine exhaust gases cooling curve plotted on two superimposed T S_ diagrams.
T4R ¼ T1 þ DTappr
(11)
_ f ;R ðh4 h4R Þ Q_ R ¼ m
(12)
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possible to see that, unlike Fig. 9a, hot engine exhaust gases should be used to vaporize a ﬂuid at a higher temperature than T1; this condition decreases the entropy generation rate, that now accounts for 0.869 kJ/s K, lower than the value calculated for the same ﬂuid in the simple cycle case. Also, in analogy with the engine cooling water case, the mass ﬂow rate of the working ﬂuid will be higher, thus providing a higher power output (the cycle efﬁciency is also increased in this case). Table 4 reports the main ﬁgures of the regenerated cycle and it can be observed that the working ﬂuid mass ﬂow rate is 3.076 kg/s (compared to 2.737 kg/s of the simple cycle and 3.024 kg/s with preheating). The net power output is thus increased by 12.4% with 00 PORC Þ=PORC Þ and is now rated at respect to the simple cycle ððPORC 392.6 kW, with a net cycle efﬁciency of 22.3%. From Table 4 it can also be observed that the internal heat exchange provides 218 kWt, which is about 10% of the total intro00 duced heat rate ðQ_ R =Q_ in Þ. Fig. 12 describes the T—Q_ diagram for the regenerated cycle, while Fig. 13 shows the cycle power output vs. turbine inlet pressure. It is evident that a maximum power of about 430 kW can be achieved if the regenerated cycle were operated at a pressure of 4180 kPa, resulting in a cycle efﬁciency of 0.246. In Fig. 14 the three conﬁgurations assumed for benzene are compared at varying expander efﬁciencies. In particular hT has been assumed in the range of 0.4 (accounting for expanders with very low performance) to 1 (assuming ideal expansion). The expander efﬁciency of hT ¼ 0.7 assumed in the analysis presented in this paper, is in the middle of this range. It can be observed that the simple cycle thermally powered only by engine exhaust gases always displays lower power output values than the other two conﬁgurations. Also observed that an expander efﬁciency value of about 0.75 marks the break even between a regenerated and water cooling cycle. Therefore turbines with low adiabatic efﬁciency should be employed in cycles that adopt the regenerated design. In fact a low efﬁciency expander leads to a temperature and
Fig. 10. T Q_ diagrams for benzene, R11 and R134a with preheating using engine coolant.
The state of the liquid at the recuperator outlet can be calculated, since:
_ f ;R h2R ¼ h1 þ Q_ R =m
(13)
The T—S_ regenerated cycle is presented in Fig. 11, with the superimposed line referring to engine exhaust gas cooling. It is
Fig. 11. Optimal benzene regenerated cycle and engine exhaust gases cooling curve plotted on two superimposed T S_ diagrams.
Table 3 Comparison of ORC cycles for benzene, R11 and R134a with preheating using engine coolant.
Benzene R11 R134a
0 [kW] PORC
0 Q_ in [kW]
Dh30A [–]
DH_ 1A =Q_ in [–]
0
DH_ 1A =Q_ w [–]
_ 0f [kg/s] m
0 V_ 3 [m3/s]
0 ððPORC PORC Þ=PORC Þ [–]
386.0 332.5 199.5
1943.1 2005.6 2342.5
581.6 204.3 142.8
0.095 0.127 0.261
0.175 0.242 0.579
3.024 8.573 12.129
0.058 0.035 0.056
0.099 0.145 0.348
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Table 4 Main parameters of benzene regenerated cycle.
Benzene
00 [kW] PORC
h00ORC [–]
Q_ R [kW]
00 Q_ R =Q_ in [–]
pvap [kPa]
_ 00f [kg/s] m
00 V_ 3 [m3/s]
00 ððPORC PORC Þ=PORC Þ [–]
392.6
0.223
218
0.099
2000
3.076
0.059
0.124
Table 5 Combined ICE–ORC efﬁciencies.
hCC
Fig. 12. T Q_ diagrams for regenerated benzene cycle.
Simple cycle
Simple cycle with preheat
Reg. cycle
Benzene R11 R134a
0.466 0.458 0.438
0.471 0.463 0.445
0.471 – –
Benzene R11 R134a
0.114 0.095 0.048
(hCC he)/he 0.126 0.108 0.065
0.128 – –
enthalpy increase of the superheated vapour at the turbine outlet and this energy is partly recovered in the recuperator. If performing expanders with efﬁciencies higher than 0.75 were employed, the design with water preheating should be preferred. It can be observed that when hT / 1 the overall power achieved by the benzene regenerated cycle approximates the power generation capability of a simple cycle with no preheating. At very high expander efﬁciencies, direct thermal regeneration is almost fruitless as the vapour state at the expander outlet nears saturation. 6. Conclusions
Fig. 13. Regenerated benzene cycle power output at different evaporation pressures.
Fig. 14. Cycle power versus turbine efﬁciency for the three cycle conﬁgurations assumed for benzene.
Table 5 summarizes the net efﬁciency estimated for the combined ICE–ORC cycle (hCC). These values have been calculated assuming 95% efﬁciency of the electric generator coupled to ORC expander. It is possible to appreciate a signiﬁcant increase in efﬁciency from the rated value for the engine (he ¼ 0.418, see Table 1). The highest values are achieved with benzene cycles and interestingly the value is nearly the same in the regenerated and preheated cases, with hCC ¼ 0.47. This is because the value of hT adopted in the analysis is close to the values that makes indifferent the adoption of the water preheating or regeneration in the design of the cycle (Fig. 14). The increase of the efﬁciency is about 12.5% with respect to engine with no bottoming ((hCC he)/he) in both cases. It has to be considered that the regenerative preheating requires a liquidgas heat exchanger and furthermore it has been assumed a quite low temperature difference at the exchanger outlet (DTappr,R). This requires very complex heat exchangers with high exchange surfaces, which leads to a component that needs a critical design. When engine coolant is used for preheating instead, a liquid–liquid exchange process is performed and the high availability of ﬂuid at high temperatures makes it possible to use simpler components. Therefore, if expanders with average performances are available (as hT ¼ 0.7 assumed in the analysis) and the heat released through the engine refrigerant is not otherwise utilized, it could be preferable to employ the scheme proposed in Fig. 2b, thus simplifying the design of the heat exchanger required. While cycles based on R134a give small contributions in both conﬁgurations (the increase of engine efﬁciency may reach 6.5%), R11 can provide good performances especially when water preheating is adopted. In this case the increase in engine efﬁciency is almost 11% and hCC ¼ 0.46. R11 or similar ﬂuids may therefore be preferable than overhanging ﬂuids like benzene for the smaller turbine outlet/inlet
I. Vaja, A. Gambarotta / Energy 35 (2010) 1084–1093
volume ﬂow ratio. From Fig. 8 in fact it can be observed that v4/v3 is always smaller than 50, which would allow the use of simple expanders such as single stage turbines. In conclusion the use of ORC has been assessed in conjunction with ICEs and it has been shown how interesting thermodynamic advantages can be achieved (when heat released from an ICE is not used for cogeneration purposes). Further ORC cycle solutions will be analysed in future works as well as dynamical model will be introduced to evaluate the behaviour of the system in offdesign conditions. The effects of using diathermic oils as heat transfer media between engine gases and organic ﬂuid will also be considered in order to reduce risks related to the ﬂammability of some of the ﬂuids that may be employed and to ensure higher stability for the operation of the ORC due to the oil thermal inertia. This design would however introduce further irreversibilities in the main heat exchange process that are cause for a reduction in the global ORC system efﬁciency. A detailed exergy analysis will numerically quantify the entity of these losses.
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