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Applied Thermal Engineering 36 (2012) 279e287

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Applied Thermal Engineering

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

HD Diesel engine equipped with a bottoming Rankine cycle as a waste heatrecovery system. Part 2: Evaluation of alternative solutions

J.R. Serrano, V. Dolz, R. Novella*, A. GarcíaCMT e Motores Térmicos, Universitat Politècnica de València, Spain

a r t i c l e i n f o

Article history:Received 29 September 2010Accepted 14 October 2011Available online 24 October 2011

Keywords:Diesel engineEngine efficiencyPollutant emissionsWaste heat recoveryBottoming Rankine cycleOrganic Rankine cycle

* Corresponding author.E-mail address: rinoro@mot.upv.es (R. Novella).

1359-4311/$ e see front matter � 2011 Elsevier Ltd.doi:10.1016/j.applthermaleng.2011.10.024

a b s t r a c t

A theoretical investigation has been performed on the feasibility of introducing a waste heat recovery(WHR) system in a two-stage turbocharged HDD engine. The WHR is attained by introducing a Rankinecycle, which uses an organic substance or directly water as a working fluid depending on energeticperformance considerations. A previous research was carried out to evaluate the maximum potential ofthis WHR concept, a conventional layout was used for coupling the Rankine cycle to the thermal engine.The objective of the present research is to broad the scope of the previous analysis by considering newalternative solutions for the problems related to the coupling between the WHR Rankine cycle and thethermal engine. These solutions are based on adapting one of the turbochargers by removing its turbineand trying to recover the energy by the Rankine cycle. Finally, the turbine of the Rankine cycle suppliesthe recovered energy directly to the compressor of this turbocharger. Thus, in these layouts the couplingis simpler as it involves only two turbomachines, which are supposed to share a similar rotating speed.From the results of the global energy balance, these alternative layouts produce slight benefits in fuelconsumption but in all cases these benefits are lower compared to those attained with conventionallayouts.

� 2011 Elsevier Ltd. All rights reserved.

1. Introduction

The main advantage of direct injection Diesel engines is its highthermal efficiency. This competitive efficiency together with itshigh reliability makes the direct injection Diesel engine particularlysuitable as a power plant for heavy-duty transport applications.Promoted by the increasingly strict regulations on pollutant emis-sions [1], the Diesel engine is being object of intense research tomake it more environmentally friendly, especially regarding NOxand particulate matter. An attractive alternative for improving theoverall thermal efficiency of Diesel engines consists of recoveringthe energy lost by means of a waste heat recovery (WHR) system.

Hountalas et al. performed a theoretical analysis comparing themost common WHR systems, including mechanical and electricalturbocompounding together with a bottoming Rankine cycle [2].From the reported results, a reduction in fuel consumption up to8e9% at full engine load is feasible.

In a previous stage of the present research and in other worksavailable in the literature have been stated the benefits in terms offuel consumption produced by the bottoming Rankine cycle

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strategy [3,4]. Two different approaches were investigated, twoRankine cycles in cascade (binary cycle), and a single Rankine cycleneglecting the low temperature sources. The reduction in fuelconsumption without considering internal irreversibilities rangedfrom 16% in the first configuration to 8.5% in the second configu-ration. Similar results have been reported in the literatureregarding a potential decrease in fuel consumption of approxi-mately 10% attainable by integrating a bottoming cycle [5].

However, most of the literature limited its scope to the energyflow analysis, omitting any reference about how to reintroduce therecovered energy into the Diesel plus Rankine engine power plant.The main alternatives consist of directly linking the turbine shaftwith the crankshaft or converting the mechanical power intoelectrical power to make it suitable for its use [6]. The layout of thefirst alternative, although seemingly straightforward, is not easy todevelop due to the extreme differences in rotation speed betweenthe turbine and the Diesel engine, while the second solutionrequires an electrical generator and also a set of batteries to storethe energy [7] aside from the complexity in development therequired turbine or expander for ORC or steam bottoming cycles.

Considering the different Diesel engine subsystems, there is oneturbomachine rotating at similar speed as that of the bottomingcycle turbine, the compressor. Therefore, the power produced bythe Rankine cycle turbine could be directly used to drive the

Nomenclature

Acronymsbsfc brake specific fuel consumptionDOC Diesel oxidation catalystDPF Diesel particulate filterEGR exhaust gas recirculationHDD heavy duty dieselHP high pressureimep indicated mean effective pressureLP low pressureMP medium pressureORC Organic Rankine cycleoxycat oxidation catalystpmep pumping mean effective pressureOpenWAM� wave action modelWHR waste heat recovery

Symbolsp pressures entropyT temperatureV volume

Fig. 1. Power plant layout of the configuration without LP turbine, with MP EGR.

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compressor required to supercharge the Diesel engine, thusreplacing the conventional turbine placed at the exhaust line. Thiswould reduce the engine back pressure and the mechanicalpumping losses, which in turn could also increase the efficiency ofthe power plant. However, the authors have not found any refer-ence in the literature discussing the feasibility of this alternative.

As a continuation of previous theoretical analysis reported in[4,8,9], and with the aim of improving the knowledge about thepotential of a bottoming cycle as a WHR system for future HDDengines, this paper is focused on evaluating the energy balancebetween the energy recovered by the exhaust turbine in theconventional Diesel engine layout and the energy produced by thebottoming cycle, plus the energy recovered as pumping work fordifferent engine adapted layouts. The possibility of driving theengine compressor only with the power produced by the turbine ofthe Rankine cycle has also been evaluated. This could make thesolution easier to implement in real engines.

2. Materials and methods

The present theoretical analysis has been performed on basis ofan open-software within the category of the 1D wave action model(OpenWAM�) developed at CMT-Motores Térmicos, which hasbeen previously validated to reproduce the behavior of a state-of-the-art HD Diesel engine. The use of the model along this investi-gation is justified when considering the representative results re-ported in the literature, which were obtained with similarcomputational models [10e13].

A detailed description of OpenWAM� is provided in theprevious research work performed by the authors [14e18], wherethe methodology followed to fit this model to the engine and thequality of the fit are also provided. Contrarily to other similarstudies, in this case the model performs according to a real Dieselengine. Therefore, despite being a theoretical research, the resultsprovided by the model are expected to be qualitatively similar tothose produced by the engine.

The engine under analysis is a 6-cylinder 12-litre HDD oneequipped with two turbochargers in series. The main

characteristics of this engine are presented in [19]. As in the case ofthe OpenWAM�, further details of this engine model can be foundin reference [18].

The full load and maximum engine speed operating condition(1800 rpm), where the maximumwaste heat is produced, has beenselected for the present research. If the results obtained with thewaste heat recovery system at this operating condition are notacceptable, the solution under discussion can probably bediscarded.

Starting from the conventional two-stage turbocharged HDDengine configuration already studied in a previous part of thisresearch work, the strategy followed along this second part of thepaper consists of a stepwise evaluation of different alternativeconfigurations that link the thermal engine to the bottoming cycle.Initially, some configurations can be easily discarded, such as all ofthose which replace the high pressure turbine, since this wouldhave a negative impact on the transient response of the engine.Thus, in this research, the bottoming Rankine cycle has been linkedto the low pressure compressor, and therefore the low pressureturbine has been removed from the system, and the layout of theremaining power plant elements must be remapped. As a result ofthe previous discussion, three main solutions will be examined indetail along this paper.

In each solution, the bottoming cycle has been optimizedkeeping the injected fuel constant as in the first part of this paper.Additionally, the criterion of maximum power instead of that ofmaximum thermal efficiency has been considered, since the formeris a more convenient optimization parameter in waste heatrecovery applications [19].

3. Results and discussion

Initially, the power plant layout has been introduced andthereafter the potential of this layout for increasing the total power

Fig. 2. Pumping loop indicator diagrams for the reference configuration (top) andwithout LP turbine, with MP EGR configuration (bottom).

Fig. 3. Temperature vs. entropy diagram (condensation temperature 50 �C).

J.R. Serrano et al. / Applied Thermal Engineering 36 (2012) 279e287 281

of the thermal engine is critically discussed. Subsequently thepower required to drive the low pressure compressor and thepossibility of supplying it exclusively with the bottoming Rankinecycle is analyzed. Finally, the main drawbacks detected in thepower plant layout are discussed and after that, the powerrequirement of the low pressure compressor is evaluated. The nextstep in this analysis consists of discussing critically the potential ofthe given power plant layout to increase the total power compared

110 130 150 170 190

Evaporation Temperature (ºC)

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33Optimum

Fig. 4. Parametric study results (con

with that produced only by the thermal engine, the possibility ofsupplying the low pressure compressor only with the powerproduced by the bottoming Rankine cycle and the main drawbacksdetected.

3.1. Configuration without LP turbine, with MP EGR

The sketch of the first configuration investigated is shown inFig. 1. As shown in this figure, the low pressure turbine has beenremoved and the elements in the exhaust line have been arrangedas follows: high pressure turbine, after-treatment system thatcombines a Diesel particulate filter (DPF) and an oxidation catalyst(DOC), Rankine’s heat exchanger and muffler. This configurationmakes it difficult to reach appropriate EGR rates, due to thedecreased in cylinder outlet pressure when the LP turbine isremoved. Therefore, a medium-pressure EGR (discharging betweenthe LP compressor and the HP compressor) is needed to obtain theoriginal engine EGR rate.

The redesign of the EGR circuit involves an increase in mass flowand temperature at the intercooler, HP compressor and aftercooler.This, in turn, involves the redesign of this part of the intake line andrequires increasing the size of the heat exchangers in order toachieve similar intake temperatures to those of the referenceengine. An increase in the size of the HP turbocharger is alsoneeded to reach the same pressures in the intake line as in thereference engine.

These changes in the engine configuration produce significantchanges in the pumping loop indicator diagram. Fig. 2 shows thepumping loop for both the reference engine and this first config-uration. The shaded part of the pumping loop indicates an areawith“positive” pumping work. These changes imply a reduction inpumping mean effective pressure (pmep), from 2 bar (referenceengine) down to 0.2 bar (bottoming cycle instead of LP turbine).

This is the way how the energy recovered by the ORC is trans-ferred to the crankshaft. The imep obtained with the modelremains constant despite having changed the engine configuration.Thus, the reduction in pmep improves the engine performance.

This first configuration of the engine affects the available heatsources in the IC engine. On the one hand, the EGR cooler disap-pears. On the other hand, the temperatures in some elements willchange. Furthermore, it is noteworthy how the air mass flowthrough intercooler and aftercooler increases by 30% due to EGRflow, thus there is greater heat output from these elements. Thisfirst configuration has three main heat sources: intercooler(125 kW), aftercooler (85 kW) and exhaust gases (133 kW).

Additionally, the intercooler heat output increases due to veryhot EGR gases. However, the aftercooler is kept at a mediumelow

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Evaporation Temperature (ºC)

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0.09

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Optimum

densation temperature 50 �C).

Fig. 5. Energy scheme cycle configuration without LP turbine with MP EGR.

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temperature. Studies with and without aftercooler heat recoverywere conducted. These showed very small improvements consid-ering the aftercooler as a waste heat source for the Rankine cycledue to its low temperature. Therefore, it was decided to discard itscontribution to this solution. One objective of this configurationwas increasing the temperature of the exhaust gases for theRankine cycle heat exchanger, but this was not achieved. In fact, thetemperature of the exhaust gases for this first configuration islower than that of the reference engine. This is a consequence of the

Fig. 6. Power plant layout of the configuration without LP turbine, with MP EGR.

reduced exhaust pressure (from 6.2 bar down to 4), which impliesa general decrease in the temperatures of the exhaust line.

A parametric study of the Rankine cycle with different workingfluids has been performed. As in the first part, the investigatedworking fluids are R-245fa, FC72, FC87, HFE7000, HFE7100, R-236fa, RC-318 and water. The parametric study follows the samecriteria as the one carried out in the first part of thework. The aim isto obtain the thermodynamic cyclewith the highest possible poweroutput .The cycle presents the following hypothesis: steady stateconditions, no pressure drop in the condenser and vaporizer, andisentropic processes for the pump and expansion machine. As inthe first part of the work, the implementation of these configura-tions in industry applications implies an appropriated expandermachine selection, in order to obtain an acceptable efficiency and toconsider the most important internal irreversibilities of thesecycles. The technological and thermodynamical implementation ofthe cycle is subjected to these main two restrictions:

� Pressure ratio in the expansion machine must be lower than25.

� The heat exchangers must have a temperature differencebetween the hot and cold fluids higher than 10 �C.

The result is that water is the best suited fluid for these heatsources. Fig. 3 shows the water Rankine cycle for the first config-uration in a temperature vs. entropy diagram.

Fig. 4 shows the results of this parametric study with themaximum cycle temperature and the evaporation temperature. Theoptimum point shown in the figure is chosen for the water Rankinecycle and indicates the area where maximum power is recovered.The top-left striped area indicates the points where the outlet of theexpansion process presents a vapor title lower than 90%. This is notconsidered a valid area because a too wet steam could damage theexpansion machine. The striped area on the right shows theworking points where the pressure ratio in the expansion process ishigher than 25.

These maps show that the cycle efficiency increases withincreasing evaporation temperatures and also with increasingmaximum temperature of the cycle, but in a lower impact. In this

Fig. 7. Pumping loop indicator diagrams for the reference configuration (top) andwithout LP turbine, with LP EGR configuration (bottom).

Fig. 8. Temperature vs. entropy diagram (condensation temperature 50 �C).

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Evaporation Temperature (ºC)

HFE 7100 Output Power (kW)

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Fig. 9. Parametric study results (con

J.R. Serrano et al. / Applied Thermal Engineering 36 (2012) 279e287 283

case, it is clear that the decisive factor is the evaporation temper-ature. However, increasing the evaporation temperature results ina reduced working fluid mass flow, because the pinch-point limitsthe mass flow for a given evaporation temperature. In conclusion,the maximum net power output is obtained at the midpoint shownin Fig. 4.

Clearly, the global power of the engine plus the Rankine cycleincreases in 20 kW (32 kW saved in pumping losses plus 33 kWprovided by the steam cycle minus 45 kW required to spin theturbocompressor), thus the total mechanical power increases by6.4% considering the reference configuration. However, thisrecovered power is less than that observed for the engine config-urations with its original two-stages turbocharging architectureinvestigated in the first part of this article, but this studiedconfiguration has only twowaste heat sources to be recovered. Thiscan result into a technological simplification, due to fewer heatexchangers. Information about the amount of heat rejected hasbeen included in Fig. 5, where the energy scheme of the engineconfiguration without LP turbine and with MP EGR is shown.Finally, note how the Rankine cycle, in this configuration, has a highefficiency (about 14%) compared to the usual levels reported forbottoming cycles.

Regarding the mechanical coupling, the Rankine cycle does notprovide enough power to drive the compressor by itself (33 kWprovided by the steam cycle versus 45 kW required by the engine).Therefore, some type of shaft coupling between the IC engine andthe LP compressor should be considered in order to provide thenecessary power. Such devices are complex, but nowadays, severalelements to couple the high-speed turbomachinery to the low-speed IC engines have been developed [20]. If this is not viable,an additional electric motormust be installed to provide this powerrequired by the LP compressor.

3.2. Configuration without LP turbine, with LP EGR

Fig. 6 shows the sketch of the second configuration. As in thefirst configuration, the LP turbine has been removed and theelements in the exhaust line have been rearranged accordingly.However, in this second configuration, a low-pressure EGR line hasbeen implemented to solve the high EGR rate problem; in this casethe exhaust gases discharge to the LP compressor inlet.

For this new configuration with the low-pressure EGR, there areextra exhaust gas flows through the gas turbine, therefore the gasturbine expansion ratio can be reducedwhilemaintaining the same

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densation temperature 50 �C).

Fig. 10. Energy scheme cycle configuration without LP turbine, with LP EGR.

J.R. Serrano et al. / Applied Thermal Engineering 36 (2012) 279e287284

shaft power. This 30% extra gas mass flow also circulates throughthe exhaust gases heat exchanger, thus there is higher heat energyavailable to the bottoming cycle. Again, the inlet line should bemodified to accommodate the increased flow caused by the low-pressure EGR. In spite of cooling the exhaust gases in the heatexchanger of the Rankine cycle, the low-pressure EGR still involvesgas at a temperature as high as 135 �C through the LP compressor;this 30% of high temperature gas raises the temperature at theintake line, worsening the compressor performance. Finally, the

Fig. 11. Power plant layout of the without LP turbine, with LP EGR and high temper-ature Rankine cycle configuration.

increasedmass flow through the LP compressor implies an increasein its power consumption from 45 kW, in the configurationwithoutLP turbine and with MP EGR, up to 63 kW, in this configurationwithout LP turbine and with LP EGR.

Fig. 7 shows the pumping loop for both the reference engine andthis second configuration. These changes imply a reduction inpmep, from 2 bar (reference engine) to �0.7 bar (without LPturbine and with LP EGR). This represents a notable improvementin pumping work from the configuration without LP turbine.

As in the previous case, imep remains more or less constant.Thus, the reduction in pmep improves the engine performance. Thepower consumption in the LP compressor is higher in this secondconfiguration than in the first configuration, because now the EGRin a rate of 30% also flows through the LP compressor and at 150 �C,the advantage of extra cooling in the EGR using for example enginewater circuit at about 85 �C has not been explored in present workbut it will for sure reduce power consumption by the LPcompressor.

Regarding the available heat sources, the EGR cooler disappearsand the temperatures in the various elements will change. Thefollowing heat sources are present in the second engine configu-ration: intercooler (55 kW), aftercooler (105 kW) and exhaustgasses (182 kW). In this case, the exhaust line temperature is evenlower than in the previous case. The decrease in exhaust backpressure also reduces the temperature, limiting the availableenergy in this residual heat. Moreover, the intercooler maintainslow temperatures and low available heat. Therefore, its contribu-tion in this solution has not been considered, as the improvementsto be gained from this element would not be relevant. On the otherhand, it brings a slight aftercooler recovered heat, making itconvenient to take this heat in the Rankine cycle.

A parametric study of the Rankine cycle with different workingfluids has been performed, with the same criteria previously used.The result is that the organic fluid HFE7100 is the fluid best suitedfor these heat sources. Fig. 8 shows the Organic Rankine cycle (ORC)in a temperature vs. entropy diagram that in this case is a regen-erative cycle.

Fig. 9 shows the results of this parametric study versusmaximum cycle temperature and evaporation temperature. The

Fig. 12. Pumping loop indicator diagrams for the reference configuration (top) andwithout LP turbine, with LP EGR and high temperature Rankine cycle configuration(bottom).

Fig. 13. Temperature vs. entropy diagram (condensation temperature 50 �C).

J.R. Serrano et al. / Applied Thermal Engineering 36 (2012) 279e287 285

optimum point shown in the figure is chosen for the HFE7100 ORCand indicates the areawhere themaximumpower is recovered. Thebottom-right striped area represents the points where the evapo-ration temperature is higher than the cycle maximum temperatureand consequently, they do not constitute possible cycles.

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Fig. 14. Parametric study results (co

The maps show how the performance increases with increasingevaporation temperatures and also with increasing maximumtemperature of the cycle, but in a lower impact. In this case, it isclear that the evaporation temperature is the determining param-eter. However; the working fluid flow is reduced by increasingthese temperatures. Therefore, the maximum net power output isagain obtained at an intermediate point. The pinch-point is also thelimiting factor for this case.

In summary, the global power of the engine plus the Rankinecycle increases in 24 kW (47 kW due to the decrease of pumpinglosses plus 40 kW provided by the ORC and minus 63 kW requiredto spin the compressor). The results are similar to the configurationwith MP EGR, but in this case the total mechanical power increasesby nearly 7.7%. Although, as in the previous configuration, the use ofonly two waste heat sources will reduce the size of heat exchangesurface required for the cycle.

Fig. 10 shows the energy scheme of the engine configurationwithout LP turbine and with LP EGR. The ORC that operates onlywith medium temperature sources presents an good efficiency(15.7% in this case), with a not too large increase of heat rejected.

3.3. Configuration without LP turbine, with LP EGR and hightemperature Rankine cycle

Fig. 11 shows the sketch of the third configuration. As in theprevious configuration, the LP turbine has been removed but nowthe elements in the exhaust line have been rearranged as shown inFig. 11. The after-treatment elements were placed so that theywould work with the highest temperature followed by the Rankinecycle heat exchanger and the HP turbine; at the end of the line thelow-pressure EGR line and the muffler can be found. This thirdconfiguration seeks to recover the heat from the exhaust gasesbefore they flow through the turbine. Thus, their temperature willbe higher than in the previous configurations. The Rankine cyclewill have a higher temperature difference and therefore a higherperformance.

This configuration improves the after-treatment efficiency andthe Rankine cycle efficiency. However, the turbine performanceworsens when decreasing the turbine inlet temperature. Thisimplies an increase in the exhaust back pressure to recover thework needed to drive the HP compressor. Additionally the turbinetemperature should be high enough to prevent water condensationfrom exhaust gases.

Fig. 12 shows the pumping loop for both the reference engineand this third configuration. These changes imply a smallerreduction in pmep, from 2 bar (reference engine) to 0.9 bar (thirdconfiguration).

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ndensation temperature 50 �C).

Fig. 15. Energy scheme cycle configuration without LP turbine, with LP EGR and high temperature Rankine cycle.

Table 1Engine values for steady tests.

Configurations Mechanicalpower in ICengine

Power requiredin the compressor

Mechanicalpower cycle

Totalmechanicalpower

Increment of totalmechanical power

Total heattransferprocess

DOC inlettemperature

(kW) (kW) (kW) (kW) (kW) (kW) (�C)

Reference configuration 311 e e 311 e 357 500Configuration with all the

sources (water Rankine cycle)311 e 31 342 10.0% 868 330

Configuration with all thesources (Binary Cycle)

311 e 59 370 19.0% 938 330

Configuration with hightemperature heat sources

311 e 46 357 14.7% 729 330

Configuration without LP turbine and MP EGR 343 45 33 331 6.4% 737 310Configuration without LP turbine and LP EGR 358 63 40 335 7.7% 747 300Configuration without LP turbine, LP EGR

and high temperature Rankine Cycle334 63 19 290 �6.8% 801 500

J.R. Serrano et al. / Applied Thermal Engineering 36 (2012) 279e287286

As in the previous cases, imep remains essentially constant.Thus, the reduction in pmep improves the engine performance. Thethird configuration has these heat sources: intercooler (54 kW),aftercooler (100 kW) and exhaust gases (173 kW). The temperatureat the exhaust heat exchanger is the highest of the three studiedcases. Consequently, the Rankine cycle efficiency improves, but theexhaust gases have lower heat energy. However, the aftercooler hasenough power and high enough temperature to be used as a heatsource in the Rankine cycle, together with the exhaust gases.

A parametric study of the Rankine cycle with different workingfluids has been performed. Water is the best suited fluid forrecovering the heat from these sources, due to their hightemperature.Fig. 13shows the Rankine cycle for this configuration.

Fig. 14 shows the results of this parametric study. The optimumpoint shown in thefigure indicates the areawheremaximumpoweris recovered. The striped area on the right shows theworking pointswhere the pressure ratio in the expansion process is higher than 25.

The main restriction in this study is the expansion ratio in theexpansion process. Despite the power of the steam cycle, the totalmechanical powerdecreases in 21 kW(23 kWdue to the reduction ofpumping losses plus 19 kWprovided by the steamcycleminus 63kWrequired to spin the LP compressor), that is a power reduction of 6.8%.

Fig. 15 shows the energy scheme of this engine configuration. Inthis case, due to the limitations, the optimal solution consists ofa water steam cycle with low efficiency (7.2%). This will result ina low power output and a large increase in the total size of heatexchangers.

4. Summary

Table 1 shows a summary with all the solutions studied in boththe first and the second part of this paper. This table shows how thetotal mechanical power varies for each configuration and the totalheat transferred at the heat exchangers also increases (the size ofthe heat exchangers increases). Finally, it also shows the gastemperature at the inlet of the Diesel oxidation catalyst (DOC). Thistemperature can be related with the DOC efficiency, highertemperatures can increase DOC efficiencies.

5. Conclusions

From a global point of view, the best solution is the "configu-ration with high temperature heat sources". That is, a simple waterRankine cycle aiming to use the residual heat of the IC engine with

J.R. Serrano et al. / Applied Thermal Engineering 36 (2012) 279e287 287

high temperature heat sources. This achieves about 15% incrementin the global mechanical power, using a heat exchanger of largersurface because an increment in heat transferred is necessary. The"configuration with all the sources. binary cycle" has a powerincrement of 19% with respect to the reference engine. But it meanshigher heat transferred and technological complexity of theconfiguration.

The configurations studied, which involves removing one of thetwo turbines from the reference engine (non-conventionalconfigurations), do not show any improvement on the power overthe configurations studied in the first part of the article. However,the first and second configurations have a higher power outputthan the configuration of the reference engine due to theimprovements in the pumping loop. Despite not getting a signifi-cant increase in power, these non-conventional configurations(except the third configuration of this second part) have theadvantage of reducing pumping losses and thus directly increasethe output power in the IC engine crankshaft while still keepingHPT to deal with engine load transient response. Thus, the first andsecond configurations of this second part have direct powerincrements in the IC engine of 10% and 15% respectively withrespect the reference engine. However, these configurations havea disadvantage, since the LP compressor must be driven by anadditional power source.

Acknowledgements

This work was partially funded by “Programa de Apoyo a laInvestigación y Desarrollo de la Universidad Politécnica deValencia”.

References

[1] Emission Standards for Model Year 2007 and Later Heavy-duty HighwayEngines. U.S Environmental Protection Agency (EPA) (http://www.epa.gov/).

[2] D.T. Hountalas, C.O. Katsanos, D.A. Kouremenos, E.D. Rogdakis, Study of avail-able exhaust gas heat recovery technologies for HD Diesel engine applications,International Journal of Alternative Propulsion (IJAP) 1 (2007) 228e249.

[3] S.E.Aly, Diesel enginewaste-heat power cycle,Applied Energy29 (1988) 179e189.

[4] M.M. Bailey, Comparative Evaluation of Three Alternative Power Cycles forWaste Heat Recovery Form the Exhaust of Adiabatic Diesel Engines. NationalAeronautics and Space Administration, Washington, DC, 1985.

[5] A. Vernau, Recovery from exhaust gas on a diesel engine, VDI Berichte 539(1984) 501e514.

[6] U. Hopmann, Diesel engine waste heat recovery utilizing electric turbo-compound technology, in: Annual Diesel Engine Emissions Reduction (Deer)Conference (2004) San Diego, (California).

[7] W.M.S.R. Weerasinghea, R.K. Stobarta, S.M. Hounshama, Thermal efficiencyimprovement in high output diesel engines a comparison of a Rankine Cyclewith turbo-compounding, Applied Thermal Engineering 30 (2010)2253e2256.

[8] R. Chacartegui, D. Sánchez, J.M. Muñoz, T. Sánchez, Alternative ORC bottomingcycles for combined cycle power plants, Applied Energy 86 (2009)2162e2170.

[9] E. Doyle, L. Dinanno, S. Kramer, Installation of a diesel organic-Rankinecompound engine in a class 8 truck for a single-vehicle test, in: SAE Paper790646 (1985).

[10] J. Benajes, V. Bermúdez, J. Galindo, Efecto de la geometría del intercooler enlas prestaciones de motores de automoción. XI Congreso Nacional de Ingen-iería Mecánica, 1994.

[11] M. Reyes, Heat Transfer model for exhaust manifolds in reciprocating engines.PhD Thesis, (text in Spanish) Universidad Politécnica de Valencia 1994.

[12] F. Payri, J. Benajes, J. Galindo, One-dimensional fluid-dynamic model forcatalytic converters in automotive engines, in: SAE Paper 959785 (1995).

[13] A. Bejan, G. Tsatsaronis, M. Moran, Thermal Design and Optimization. JohnWiley, New York, 1996.

[14] F. Payri, J. Galindo, J.R. Serrano, F.J. Arnau, Analysis of numerical methods tosolve one-dimensional fluid-dynamics governing equations under impulsiveflow in tapered ducts, International Journal Mechanical Sciences 40 (2004)981e1004.

[15] R. Santos, Study of the exhaust gas energy uses in diesel engines. PhD Thesis(text in Spanish), Universidad Politécnica de Valencia 1999.

[16] J. Galindo, J. R. Serrano, J. R. Arnau and P. Piqueras, Description and analysis ofa one-dimensional gas-dynamics model with independent time discretiza-tion. ASME Internal Combustion Engine Division 2008;131:3:034504 (5pages).

[17] J.R. Serrano, F.J. Arnau, V. Dolz, A. Tiseira, C. Cervelló, A model of turbochargerradial turbines appropriate to be used in zero- and one-dimensional gasdynamics codes for internal combustion engines modelling, Energy Conver-sion and Management 49 (2008) 3729e3745.

[18] J.R. Serrano, F.J. Arnau, V. Dolz, A. Tiseira, M. Lejeune, N. Auffret, Analysis of thecapabilities of a two-stage turbocharging system to fulfil the US2007 anti-pollution directive for heavy duty Diesel engines, International Journal ofAutomotive Technology 9 (3) (2008) 277e288.

[19] T. Yamamoto, T. Furuhata, N. Ara, K. More, Design and testing of the OrganicRankine Cycle, Energy 23 (2001) 239e251.

[20] C.J. Chadwell, M. Walls, Analysis of a superturbocharged downsized engineusing 1-D CFD simulation, in: SAE Paper 2010-01-1231 (2010).