Analysis of Exhaust Waste Heat Recovery From a Dual Fuel Low Temperature Combustion Engine Using an Organic Rankine Cycle

Energy 35 (2010) 2387e 2387 e2399

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Analysis of exhaust waste heat recovery from a dual fuel low temperature combustion engine using an Organic Rankine Cycle Kalyan K. Srinivasan, Pedro J. Mago, Sundar R. Krishnan

*

Department of Mechanical Engineering, Mail Stop 9552, 210 Carpenter Building, Mississippi State University, Mississippi State, MS 39762, USA

a r t i c l e

i n f o

 Article history: Received 9 November 2009 Received in revised form 9 February 2010 Accepted Accepted 10 February 2010 2010 Available online 8 April 2010 Keywords: Waste heat recovery recovery Organic Rankine Cycle Dual fuel engines Low temperature temperature combustion

a b s t r a c t

This paper examines the exhaust waste heat recovery potential of a high-ef  �ciency, low-emissions dual fuel low temperature combustion engine using an Organic Rankine Cycle (ORC). Potential improvements in fuel conversion ef �ciency (FCE) and speci �c emissions (NO x  and CO2) with hot exhaust gas recirculation (EGR) and ORC turbocompounding were quanti �ed over a range of injection timings and engine loads. With hot EGR and ORC turbocompounding, FCE improved by an average of 7 percentage points for all injection timings and loads while NO x  and CO2  emissions recorded an 18 percent (average) decrease. From pinch-point analysis of the ORC evaporator, ORC heat exchanger effectiveness ( 3), percent EGR, and exhaust exhaust manifold manifold pressure pressure were identi identi�ed as import important ant design design param paramete eters. rs. Highe Higherr pinch pinch point point temperature differences (PPTD) uniformly yielded greater exergy destruction in the ORC evaporator, irrespect irrespective ive of engine engine operating operating conditions. conditions. Increasin Increasing g percent percent EGR yielded yielded higher higher FCEs and stable stable engine operation but also increased exergy destruction in the ORC evaporator. It was observed that hot EGR can prevent water condensation in the ORC evaporator, thereby reducing corrosion potential in the exhaust piping. Higher  3  values yielded lower PPTD and higher exergy ef �ciencies while lower  3   values decreased post-evaporator exhaust temperatures below water condensation temperatures and reduced exergy ef �ciencies.   2010 Elsevier Ltd. All rights reserved.

1. Introduction

National National energy energy security security,, rising energy prices, prices, increasingl increasingly y comp compet etit itiv ive e glob global al mark market ets, s, and and stri string ngen entt envi enviro ronm nmen enta tall emissions regulations are primary driving forces in the search for sustainable and economically viable technologies for ef �cient and clean approaches approaches to energy energy conversi conversion on and utilization utilization.. Internal Internal combustion engines (IC) are prime movers of choice when high power densities and ef �ciencies are desirable. Due to relatively cheap cheap fuel fuel availa availabil bility ity in past past decade decades, s, IC engine enginess had been been optimized optimized for high power power densities densities and low-emiss low-emissions. ions. However However,, in recent recent years, years, with escalating escalating fuel prices prices and sustainabi sustainability lity concerns concerns,, engine ef �ciency has assumed greater importance. Over the past 150 years since the invention of the IC engine great great strides strides have have been made in improving improving its fuel conversion conversion ef �ciencyand ciencyand reduc reducing ing its emissi emissions ons.. As a testam testament ent to this this fact, fact, the average peak brake thermal ef �ciency of on-road spark ignition power power trains trains is about about 30% while while that that of onon-ro road ad compr compress ession ion igniti ignition on powe powerr trains trains is about about 41% 41% [1] [1].. Furth Further er impro improvem vement entss in fuel fuel

*

 Corresponding author. Tel.:  þ 1 662 325 1544; fax:  þ 1 662 325 7223. E-mail address:   [email protected]  (S.R. Krishnan).

0360-5442/$ e 0360-5442/$ e  see front matter    2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2010.02.018

conversion ef �ciency require a system-level analysis of the various losses encountered in the IC engine. This analysis can begin with thermodynamic modeling of the IC engine. Traditional �rst lawbase based d ther thermo mody dyna nami micc mode models ls faci facilit litat ate e accu accura rate te ener energy gy accounting, i.e., they are useful in estimating the net losses associated ciated with with the combus combustio tion n proce process. ss. But the these se models models do not provide estimates of how much of the wasted energy is actually recoverable as useful work or exergy. This requires second lawbased based model modelss that that track track the irrev irrevers ersibi ibiliti lities es associ associate ated d with with various processes that destroy fuel chemical exergy. 1.1. 1. Selective review of exergy destruction in combustion processes

Examin Examinati ation on of exerg exergy y destru destructi ction on in intern internal al combus combustio tion n engine processes has been actively studied by many researchers to identify, and possibly recover some of the lost available work. Lior and Rudy [2] [2] performed  performed a second second law analysis of an ideal Otto cycle and they concluded that a simple  � rst law analysis does not reveal the magnitude of work-potential lost in the combustion process. The exergy analysis results showed that the magnitude of these losses were comparable to that of useful shaft work. They recommended the employment of bottoming cycles to recover recover the exergy destroyed in the exhaust. Dunbar and Lior  [3]  examine the exergy

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Nomenclature

ALPING LTC BTDC WHR ORC CCHP PPTD COV EGR BMEP IC P  T  H  DH    h _ E  h   C  p   R g _ Q  _   m 3   q

Advanced Low Pilot Ignited Natural Gas Low Temperature Combustion Before Top Dead Center Waste Heat Recovery Organic Rankine Cycle Combined Cooling heating and Power Pinch Point Temperature Difference Coef�  cient of Variation Exhaust Gas Recirculation Brake Mean Effective Pressure Internal Combustion pressure (kPa) temperature ( C or K) enthalpy (kJ) enthalpy change (kJ) speci�c enthalpy (kJ/kg) exergy (kJ) ef �ciency (%) speci�c heat at constant pressure (kJ/kg-K) speci�c gas constant for air (kJ/kg-K) ratio of speci�c heats rate of heat transfer (kW) mass  � ow rate (kg/s) heat exchanger effectiveness speci�c heat transfer (kJ/kg)

destruction in constant pressure adiabatic combustion of hydrogen and methane. They conclude that the most important mechanism for exergy destruction in adiabatic combustion is internal thermal energy exchange. They de�ned internal thermal energy exchange as the internal heat transfer between products that are at a high temperature and still unreacted reactants, which are at a lower temperature. This temperature difference provides a �nite temperature gradient for internal heat transfer, and hence, exergy destruction. Caton  [4]   presents a comprehensive review of literature that utilized second law of thermodynamics to study IC engines. All of the literature reviewed used phenomenological combustion models that basically consisted of mass, energy and exergy balance equations to model the gas-exchange and combustion processes. Some signi�cant conclusions from this review are: 1. The available energy or the energy that can be used to obtain useful work increases with increasing combustion temperature. 2. Exergy is destroyed through irreversibilities in the combustion process, heat transfer across a �nite temperature difference, and exhaust processes. Most of the studies examined conclude that about 20e30% of fuel chemical exergy is destroyed due to irreversible combustion. 3. Combustion in the presence of excess air or lean combustion resulted in greater availability destruction than stoichiometric and rich combustion 4. Potential for utilizing exhaust waste heat to produce additional useful power to complement shaft power. Clearly, there emerge two pathways for better utilization of the chemical exergy of the fuel. The  � rst focuses on minimizing exergy destruction in the combustion process. The second pathway is to tap the exhaust exergy to obtain further improvements in thermal ef �ciency of the prime mover. Waste heat recovery (WHR) using Rankine bottoming cycles involves the utilization of the sensible

_ W  FCE LHV BSNOx BSCO2 NOx CO2 CO HC PM EPA ppm

power (kW) Fuel Conversion Ef �ciency (%) Lower Heating Value (kJ/kg) Brake Speci�c Nitrogen Oxide emissions (g/kWh) Brake Speci�c Carbon Dioxide emissions (g/kWh) oxides of nitrogen emissions Carbon dioxide emissions carbon monoxide emissions hydrocarbon emissions Particulate Matter Environmental Protection Agency parts per million

Subscripts: 1, 2, 20 thermodynamic states of exhaust gas 3e6 thermodynamic states of Organic Rankine Cycle � uid in intake exh exhaust TT turbocharger turbine s isentropic end state t ORC turbine p ORC pump max maximum orc Organic Rankine Cycle eng engine ng natural gas eng-orc engine-ORC turbocompounded system

enthalpy of the hot exhaust from the IC engine to heat a suitable �uid, preferably to saturated/superheated vapor, and then the sensible enthalpy of the vapor is used to obtain useful work from a turbine. This paper focuses on recovering exhaust waste heat using Organic Rankine Cycles (ORC). 1.2. Selective review of WHR technologies using Rankine bottoming  cycles

Rankine bottoming cycles have been explored in the automotive and power generation industry since the late 1970s [5e10] and more recently by automotive companies [11e13], and others [14e21]. The most signi�cant common-denominator in all the aforementioned WHR research efforts is the unanimous demonstration of 10%e15% improvements in fuel economy from status quo. A selective summary of various WHR literature using Rankine bottoming cycles is presented in Table 1. The ef �ciency of the Rankine cycle depends on the choice of  the working  � uid. As shown in Fig. 1, Rankine cycle working  � uids can be classi�ed as wet, dry and isentropic, depending on the slope of the saturated vapor line on a T-S diagram. A classical example of a wet �uid is water. Although water has been used as a Rankine cycle � uid in recent WHR efforts [12,13], some problems associated with wet �uids include the need for superheating, formation of condensate (for example, water) in the late-stage expansion process, and the possibility of high velocity impact on turbine blades from the small droplets of water posing potential damage to these blades   [15]. Therefore, there is widespread preference for dry or isentropic �uids. Many researchers have discussed the characteristics of the ideal organic solvent for WHR  from a Rankine bottoming cycle [21e26]. From these studies, it can be gleaned that the ideal organic �uid is one that results in maximum waste heat recovery ef �ciency for the particular ORC application. In addition, this  � uid should be safe and have minimal environmental impact if there were any leaks. Typically dry or

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 Table 1 Selective Review of research on WHR technologies using Rankine bottoming cycles A: Analytical, S: Simulation, T: Theoretical, E: Experimental, R: Review.

Ref.

Nature of study

Method

Salient conclusions

[6]

E

WHR accomplished in a six cylinder, 14.5 L Cummins NTC-400 diesel engine rated at 298 kW at 2100 rpm by turbocompounding

Demonstrated 12.5% improvement in power for the same fuel input, 14.8% net improvement in fuel economy and 4.6% improvement in fuel economy purely due to WHR by Rankine cycle turbocompounding

[8]

R

Reviewed many strategies to improve engine ef �ciency, such as advanced thermal cycles including multi-stage turbocompounding using Rankine cycles.

Addressed the possibility of a multi-stage Rankine cycle, 1 st stage operating on water to recover high temperature exhaust waste heat, followed by a 2 nd stage operating on R-11 (Organic solvent) to enable low temperature exhaust WHR. Predicted 15% improvement in ef �ciency through exhaust WHR.

[9]

E

An Organic Rankine Cycle System (ORCS) operating on tri�uoroethanol designedfor use witha 288-bhp, Class8, long-haul vehiclediesel engine was used to assess fuel economy improvements through both laboratory and on-road testing.

Demonstrated 12.5% improvement in highway fuel economy.

[11]

A/E

Proposes maximization of exergy in automotive engines. Design changes including cylinder head and cooling passages were made to the base 2.0 L Honda Stream SI engine to maximize exhaust energy recovery.

Demonstrated a net improvement in thermal ef �ciency of 13.2%. Notable achievements included the development of a novel compact steam expander, which had very good transient response to load changes.

[12]

S

A simulation model was developed for a (SI) hybrid electric vehicle (HEV).This simulation analyzed theWHR potential using Rankine cycles at a variety of speeds and loads.

Transient 10/15 mode Japanese drive cycle simulations indicate that the WHR con�guration using the novel steam expander is effective even at low speed urban driving conditions. Predicted an average overall improvement of 6.1% in thermal ef �ciency over the simulated 10/15 mode drive cycles

[13]

E/A

Two basic con�gurations of WHR using Rankine cycles from exhaust (con�guration A), and exhaust and coolant (con �guration B), are investigated. Dymola  simulations are performed to studythe effects of usingvarious �uids for Rankine cycles.

Simulations revealed that water would be the ideal  � uid for WHR for con�guration A; however, an organic solvent such as ethanol would be ideal for con�guration B. Experiments revealed that 0.7 e2 kW of extra power could be produced from WHR. This corresponded to a 10% improvement in engine power under transient conditions.

[15]

A

Analyze a supercritical ORC system of WHR from heavy duty diesel engines. The organic  � uid also serves as the coolant for the charge air and EGR coolers. Their analysis involved both  � rst and second law perspectives of exhaust WHR.

Predict that upto 20% improvement in engine power is possible using exhaust WHR using supercritical ORCs.

[17]

S

Utilizes PSAT software to review different architectures for exhaust WHR in HEVs.

Initial simulation results indicate that fuel economy advantages of upto 30% could be realized with exhaust WHR techniques.

[19]

S/T

Presents a concept to recover waste heat from high temperature exhaust and low temperature engine coolant. Also reviews the Rankine cycle ef �ciency of various wet, dry and isentropic  � uids.

Operating the Rankine cycles on dry or isentropic  � uids eliminates the need for superheating. Simulations indicate that using ORCs for waste heat recovery is bene �cial since both exhaust waste heat and engine coolant heat can be recovered, and therefore up to 32% improvement in fuel economy is possible.

[20]

A

Perform a thermodynamic analysis to investigate WHR from stationary IC engines.

Up to 12% improvement in thermal ef �ciency is possible with WHR  using Rankine cycles from both exhaust and engine coolant.

isentropic �uids are chosen because they do not need to be superheated to enable superior waste heat recovery ef �ciencies. For the analysis presented in this paper, a dry �uid, R113, was chosen due to its excellent low temperature waste heat recovery properties  [25,26].

popular choice due to their low setup and operational costs and relative ease of operation and maintenance   [27]. Therefore, investigation of technologies such as WHR that can signi�cantly improve the overall ef �ciency of IC engine powered CCHP systems is of paramount interest.

1.3. Relevance of WHR technologies to combined cooling, heating  and Power (CCHP) applications

1.4. Objectives

The speci�c objectives of this paper are, Due to soaring electricity costs concepts such as CCHP, which provide on-site electricity and satisfy additional cooling and heating requirements, have generated substantial public interest. Micro-scale CCHP [27e31], a small-scale application of CCHP that caters to the power requirements of residential and of �ce spaces, is a potentially viable, decentralized and cheaper alternative to produce electricity (and heating and cooling) rather than obtaining it from the central grid. It has been reported that the fuel utilization ef �ciency of cogeneration with waste heat recovery is about 70 percent [31]. Although cogeneration in these units can be accomplished by a variety of prime movers and energy sources such as microturbines and fuel cells, IC engines are the most

1. To analytically examine the exhaust waste heat recovery potential from the novel low temperature, low-emissions and high-ef �ciency Advanced LowPilot Ignited Natural Gas (ALPING) combustion using a bottoming Organic Rankine Cycle (ORC). In particular, quantify the fuel conversion ef �ciency and brake speci�c emissions (oxides of nitrogen, NOx  and carbon dioxide, CO2) bene�ts from ORC turbocompounding. 2. To obtain a second law-based understanding of the nature of  irreversibilities in ORC components, and to establish basic design criteria for ORC evaporator design using pinch-point analysis. “

”

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Fig. 1.  Schematic representation of (a) isentropic, (b) wet and (c) dry  � uids.

2. Advanced Low Pilot Ignited Natural Gas (ALPING) combustion

Improving engine fuel ef �ciencies while simultaneously ensuring very lowexhaust emissions continues to be the primarychallenge for engine designers. To meet this challenge, advanced combustion concepts such as low temperature combustion (LTC) [32] are being explored. In contrast to conventional diesel combustion, LTC offers signi�cant emissions bene�ts, especially in oxides of nitrogen (NOx) and particulate matter (PM), while maintaining diesel-equivalent fuel ef �ciencies. However, many LTC concepts have drawbacks that are potentially dif �cult to surmount. Despite their NOx   and PM emissions bene�ts, LTC strategiesyield signi�cantly higher unburned hydrocarbon (HC) and carbon monoxide (CO) emissions. More importantly, loss of combustion control at high loads prohibits most LTC concepts from achieving engine brake mean effective pressures (BMEP) (or loads) greater than 5e6 bar. Consequently, more advanced strategiesare necessary to improve engine operating range and exhaust emissions (HC & CO) with LTC. In recent experimental investigations by the authors [33,34], the ALPING LTC concept was demonstrated to yield excellent fuel conversion ef �ciencies ( 40 percent) and extremely low NOx emissions (<0.25 g/kWh). In ALPING combustion, very small pilot diesel sprays are injected relatively early in the compression stroke ( 60 Before Top Dead Center [BTDC]) to ignite lean, premixed natural gaseair mixtures and achieve partially premixed, controlled LTC. On an energy basis, the pilot fuel (diesel) contributes only 2e5 percent of the total fuel chemical energy, with the remaining 95e98 percent provided by natural gas. Early pilot injection with ALPING LTC leads to long ignition delays, allowing suf �cient time for the pilot diesel sprays to mix with the surrounding natural gaseair mixture and form an appropriately w

w

“

premixed diesel-natural gaseair mixture. Later in the compression process, when local temperatures and pressures become conducive for ignition, diesel auto-ignites and subsequently initiates partially premixed combustion of natural gas. Partially premixed ALPING LTC leads to relatively low local temperatures, thus reducing NOx formation rates signi�cantly. Further, locally lean conditions imply that PM emissions will likely be negligible. Compared to other LTC concepts such as HCCI  [35], ALPING LTC affords controlled engine operation with all bene�ts of LTC at relatively higher engine power outputs. On a single-cylinder engine, the ef �ciency and emissions bene�ts of ALPING LTC were experimentally demonstrated at BMEPs as high as 12 bar. As with other LTC concepts, ALPING LTC led to high HC and CO emissions, which were particularly severe at low engine loads. In addition, high coef �cient of variation (COV) of indicated mean effective pressure (IMEP) indicated that engine operation was unstable at low loads. To improve engine stability, HC, and CO emissions, hot (un-cooled) exhaust gas recirculation (EGR) was investigated at low loads   [36,37]. These hot EGR experiments yielded 70 percent HC emissions reduction, ef �ciency improvement by 5 percentage points, and more stable engine operation with virtually no NOx  penalty at half and quarter loads. Also, the most advanced injection timing for ALPING LTC was extended from 60 BTDC without EGR to 70 BTDC with hot EGR. In view of the NOx  and PM emissions bene�ts of ALPING LTC, it is desirable to investigate the WHR potential of the engine exhaust for conditions with and without EGR. ”

3. Experimental setup

The experimental results used in the current investigation were obtained on a single-cylinder research engine operating on

K.K. Srinivasan et al. / Energy 35 (2010) 2387 e 2399  Table 2 Engine Details.

Parameter

Speci �cation

Engine type Bore (mm) Stroke (mm) Compression ratio Combustion system Diesel injection system Diesel injector Natural gas fueling Engine speed, rev/min Brake power, kW (engine load) BMEP, bar (engine load) Intake pressure, kPa (engine load) Intake temperature,   C (engine load)

1-cylinder, 4-stroke 137 165 14.5:1 Direct injection, Mexican hat Electronic, common-rail Pencil-type, four-holes Manifold fumigation 1700 21 (half), 10.5 (quarter) 6 (half), 3 (quarter) 181 (half), 151 (quarter) 75 (half), 115 (quarter)

ALPING LTC over a wide range of injection timings. Important engine details and operating conditions are presented in  Table 2 and a schematic of the overall experimental setup is shown in Fig. 2. The engine was coupled to a direct current (d.c.) electric dynamometer via an inline torque meter. Intake, exhaust, coolant, and oil temperatures were measured using Type-K thermocouples. The engine used for the ALPING LTC experiments was coupled to intake and exhaust surge tanks to simulate turbocharging by the regulation of intake and exhaust manifold pressures (see Fig. 2). Air�ow rate was measured with a laminar � ow element (LFE) and natural gas �ow rate was measured using a thermal mass �ow meter. Emissions were sampled near the exhaust port and sent

 

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through a heated sample line to an integrated emissions bench with Rosemount Analytical NGA 2000 analyzers. Total hydrocarbon and NOx   measurements were performed on the hot, undried exhaust sample using a heated �ame ionization detector and two chemiluminescent detectors, respectively. The two NOx instruments provided several full-scale ranges from 0 to 10 ppm up to 0e10,000 ppm. Carbon monoxide and carbon dioxide were measured using non-dispersive infrared analyzers and oxygen using a paramagnetic analyzer. All engine instruments (thermocouples, pressure sensors, �ow meters, etc.) were periodically calibrated while the emissions analyzers were calibrated daily following standard procedures. The error in all experimental measurements is expected to be less than 2 percent. An electronically-controlled common-rail diesel injection system was used to achieve ALPING LTC. This injection system was capable of consistently injecting a minimum diesel quantity of  4.6 mm3/stroke and a maximum diesel quantity of 20 mm3/stroke. City natural gas containing 98.3% methane, 1.3% nitrogen, and less than 1% of ethane and carbon dioxide was fumigated along with air in the engine intake manifold. 4. Simulated turbocharging analysis

In the experiments, compressed air was provided in the intake manifold from an external source. The exhaust manifold pressure (P ex) was appropriately regulated (based on intake manifold pressure) using a gate valve to simulate the back pressure that would be experienced by the engine if an actual turbocharger

Fig. 2.  Schematic of Experimental Setup.

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were present. To determine exhaust gas properties at the inlet of  the ORC evaporator (see Fig. 3 between the dotted line 2e20 ), it is necessary to calculate the temperature (T 20 ) and enthalpy (h20 ) downstream of the turbine of an actual turbocharger (if it were present). For this purpose, the isentropic ef �ciency (hTT) of the turbocharger turbine is assumed to be 80 percent. Also, it is reasonable to expect that the downstream pressure (P 20 ) of the turbocharger turbine is just above atmospheric pressure to enable expulsion of exhaust gases through any exhaust aftertreatment devices; in this work, P 20   is assumed to be 1 psig. Assuming constant speci�c heat (c  p) across the turbocharger turbine, T 20 is determined from the following equation:



T 20 ¼ T 2   hTT T 2   T 20 s



(1)

where, T 2 is the temperature upstream of the turbocharger turbine and T 20 s   is the isentropic post-turbocharger turbine temperature that is determined from the pressure ratio across the turbine:

T 20 s ¼ T 2

1g g

 ! P 2 P 2'

(2)

The ratio of speci�c heats (g) of the exhaust gas was assumed to be 1.37 (for air at an average temperature of 400  C) for these calculations. 5. Organic Rankine Cycle analysis

This section presents the equations used to estimate the exhaust WHR potential using ORC turbocompounding from actual engine data. Measured exhaust parameters such as temperatures, pressures and  � ow rates, at quarter load (10.5 kW) and half load (21 kW) from the ALPING LTC engine (see Experimental Setup  section) have been used as inputs for the

bottoming ORC model to obtain a realistic idea of the possible bene�ts of WHR. A schematic of the engine-ORC con�guration is shown in   Fig. 3   for a typical data set for quarter load engine operation, 50 BTDC injection timing, and 0% EGR. The ORC is modeled as a closed thermodynamic system that uses R113 as the working �uid. The isentropic ef �ciencies of the ORC pump and turbine are, h p ¼  80% and hT  ¼  82%, respectively. The effectiveness (3) of the ORC evaporator is used as a modeling parameter. The sensible enthalpy of the engine exhaust (between state 20 and state 3) is used to heat the organic � uid, R113, from state 5 to state 6 or saturated vapor in the evaporator. It is to be noted that the properties of the exhaust gas at state 20 were obtained according to the procedure outlined in the simulated turbocharging section described earlier. The R113 vapor is expanded in a turbine followed by heat transfer to the environment in a condenser (state 7estate 4). The saturated liquid at state 4 is then pumped to a higher pressure, state 5, and fed into the evaporator to close the ORC circuit. The maximum energy that can be extracted from the exhaust can be determined as: _ _ Q  max ¼ mexh C  pexh T 20   T 5





(3)

_ where C  pexh  is the constant speci�c heat of the exhaust gases, m exh  is the mass �ow rate of the exhaust, T 5   is the temperature of the organic working �uid entering the evaporator, and T 20   is the temperature of the exhaust gases leaving the engine and entering the evaporator. The actual energy that can be used in the evaporator as well as the temperature of the exhaust leaving the evaporator (T 3) can be estimated using the heat exchanger effectiveness as follows:

_ _ _ Q  actual ¼ 3Q max ¼ m exh c  pexh T 20   T 3





where, 3  is the heat exchanger effectiveness.

Fig. 3.  Schematic of Engine-ORC con �guration showing a typical data set for quarter load engine operation, 50 BTDC injection timing, and 0% EGR. 

(4)

K.K. Srinivasan et al. / Energy 35 (2010) 2387 e 2399

82 80 1 2 303

FCEeng ¼

_ W  eng _ _ m LHV  ng LHV ng diesel diesel  þ m

FCEengorc ¼ The mass �ow rate of the organic working �uid and the net power output from the ORC can be estimated as: _ m orc ¼

_ _ Q  Q  actual actual ¼ qevap ðh6   h5 Þ

_ _ _ W  orc ¼ W t  þ W  p

(5) (6)

_ where m orc  is the organic  � uid mass  � ow rate,  q evap  is the evaporator speci�c heat transfer,  h 5  and  h 6  are the speci�c enthalpies of  the organic  � uid entering and leaving the evaporator, respectively, _ _ W  t  and W  p  are the ORC turbine and pump power, respectively, and they can be determined using Eqs. (7) and (8)

_ _ _ _ W  orc ðh6   h7s Þ ¼ m orc ðh6   h7 Þ t  ¼ ht W t ;  ideal ¼ ht m

_ W   p ¼

_ _ W  m  p;  ideal orc ðh4   h5s Þ _ ¼ ¼ m orc ðh4   h5 Þ h p h p

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The fuel conversion ef �ciency (FCE) of the engine and the engine-ORC con�guration can be estimated as:

 Table 3 Values used to simulate the ORC.

Turbine Ef �ciency (%) Pump Ef �ciency (%) Quality of organic  � uid leaving evaporator Evaporator Pressure (MPa) Condenser Temperature (K)

 

(7) (8)

_ _ where W  t ;   ideal  and W  p;   ideal  are the ideal power of the turbine and pump, respectively,  h t  and  h p  are the isentropic ef �ciencies of the turbine and pump, h6 and h7s  are the enthalpies of the working �uid at the inlet and outlet of the turbine for the ideal case, and h 4 and h5s   are the enthalpies of the working �uid at the inlet and outlet of the pump for the ideal case. The ORC parameters used in this analysis are speci�ed in Table 3.

(9)

_ _ W  eng  þ W orc _ _ m ng LHV ng diesel LHV diesel  þ m

(10)

where, FCEeng   is the fuel conversion ef �ciency of the engine, FCEengorc   is the fuel conversion ef �ciency of the turbocompounded engine and ORC system, LHV diesel  is the lower heating value of diesel and LHV ng  is the lower heating value of natural gas, _ _ m ng   are the mass �ow rates of diesel injected and diesel and m natural gas inducted, respectively. 6. Pinch point analysis and exergy ef �ciency 

Pinch point temperature difference (PPTD) and exergy ef �ciency are important parameters that affect combined engine-orc system performance. Pinch point temperature difference (PPTD) is de�ned as the difference between the exhaust gas temperature and the temperature at which the organic �uid, R113 in this case, �rst begins to vaporize (see   Fig. 4). This is the smallest temperature difference in the evaporator (ORC heat exchanger), and it de�nes the performance limits of the ORC heat exchanger. The pinch-point technique outlined in this section is based on the analysis in Ref. [38]. The  T eDH   diagram used for the pinch-point analysis is illustrated in Fig. 4 for a typical data set for quarter load engine operation, 50 BTDC injection timing, and 0% EGR. The pinch-point temperature can be estimated as:

T pinch ¼ b



DH 550   DH 56



 þ T 20 ¼

T 20   T 3 DH 56



DH 550   DH 56



Fig. 4.   Schematic of  T eDH   diagram (not-to-scale) used for the pinch-point analysis in the evaporator, showing a typical data set for quarter load engine operation, 50 injection timing, and 0% EGR; Note:  D H is the product of the speci �c enthalpy of the ORC working  � uid at a given state and its mass  � ow rate.

 þ T 20

(11)



BTDC

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K.K. Srinivasan et al. / Energy 35 (2010) 2387 e 2399

where  b  is the slope of Line 20 -3, DH 56   is the enthalpy difference between Points 6 and 5, and DH 550   is the enthalpy difference between Points 50 and 5. The PPTD can be calculated as:

PPTD ¼ T pinch   T 50

 

(12)

The evaporator exergy ef �ciency can be expressed as:

hexergy;  evap ¼

_ E  useful _ E  available

¼

_ _ E  6   E 5

(13)

_ _ E  20   E 3

_ _ where E  useful and E available   are the actual exergy used and the _ _ theoretically available exergy at the evaporator, E  5 and E 6  are the exergy  � ow rates of the organic working � uid entering and leaving _ _ the evaporator, respectively, and E  2  and E 3  are the exergy �ow rates of the exhaust gases entering and leaving the evaporator, respectively. The exergy �ow rate of the organic working  � uid at any point of  the ORC can be determined as:

_ _ E  orc ½ðhi   ho Þ  T o ðsi   so Þ i ¼ m

 

(14)

Similarly, the exergy of the exhaust gases (assumed to be an ideal gas) entering and leaving the ORC can be calculated as: _

_

 

E  j ¼ mexh c  pexh

T  j T  j   T o  c  pexh ln T o

  

P  j  Rexh ln P o

 

  (15)

In this study, the dead state is speci �ed by T o ¼   298 K and P o ¼ 1 atm, and the characteristic gas constant of the exhaust (Rexh) was assumed to be approximated by the characteristic gas constant of air (Rair ), and the speci�c heat of the exhaust gases at constant pressure was calculated using

c  pexh ¼

gR ; g ¼ 1:37; Rexh g1

Rair ¼ 0:287 kJ=kg:K

z

(16)

7. Results and discussion

This section presents engine-ORC results obtained for ALPING LTC at different engine operating conditions. Two different engine loads at a constant engine speed of 1700 rev/min are considered in this paper: half load operation at 21 kW (BMEP ¼ 6 bar) and quarter load operation at 10.5 kW (BMEP ¼ 3 bar). Engine performance and emissions results were obtained with and without exhaust gas recirculation (EGR) over a wide range of pilot injection timings between 20 BTDCe70 BTDC. The effects of ORC turbocompounding on engine fuel conversion ef �ciency (FCE), PPTD, exergy ef �ciency, and brake speci�c emissions are discussed. 7.1. Fuel conversion ef   �ciency: effects of EGR and ORC  turbocompounding 

In this section, FCE behavior is compared forconditions with and without waste heat recovery using ORC turbocompounding. Two aspects are compared. First, the effect of EGR on fuel conversion ef �ciency is examined; second, the improvement in fuel conversion ef �ciency with EGR and exhaust waste heat recovery using ORC is discussed. Percent EGR or EGR ratio is de �ned as follows:

EGR ratio ¼

_ ½CO2 EGR  m EGR  _ ½CO2 exh m int z

(17)

where,  ½ CO2 EGR  and  ½ CO2 exh  are CO2  concentrations in the intake _ and exhaust manifolds, respectively, m int  is the total mass �ow rate _ (of air and EGR) in the intake manifold, and m EGR  is the mass  � ow

Fig. 5.  Variation of the engine (E) and combined engine-ORC (E-ORC) fuel conversion ef �ciency with the injection timing for different EGR for quarter load operation (Power  10.5 kW, ORC evaporator effectiveness, 3  0.7). ¼

¼

rate of EGR in the intake manifold. Since it is not convenient to measure EGR mass �ow rates without considerable modi�cations to the EGR loop in the engine, the EGR ratio in this study was determined from the CO2   concentrations measured in the intake and exhaust manifolds using the integrated emissions bench. Fig. 5 shows the variation of FCE with pilot injection timing at quarter load operation for various EGR concentrations. The baseline for comparison across various EGR ratios is the case with no EGR or 0% EGR. For this baseline (0% EGR), stable engine operation could be maintained between 20 BTDC and 60 BTDC injection timings. The FCE is observed to increase with increasing percent EGR across all injection timings investigated. For instance, baseline FCE at 20 BTDC is 19 percent; whereas with the addition of 22% EGR, it increases to 22 percent. Similarly at 60 BTDC, the baseline FCE is approximately 20 percent, and with 22% EGR this value is approximately 28 percent. This trend can be explained from the fact that the EGR used in this study was un-cooled or hot, and a primary effect of using hot EGR is a net increase in the intake charge temperature. This increase in intake charge temperature translates to higher charge temperatures at the start of combustion, which leads to faster combustion rates and higher FCEs [36]. Furthermore, the addition of EGRextends the range of injection timings for stable operation from 20 BTDC e 60 BTDC to 20 BTDC e 70 BTDC. For all injection timings with or without EGR, exhaust waste heat recovery using ORC turbocompounding clearly results in increased FCEs. This is due to the fact that the exhaust gas enthalpy, which is otherwise unutilized, is used to heat the pressurized organic  � uid (R113) from compressed liquid to saturated vapor, which is then expanded in the ORC turbine (see Fig. 3). This expansion generates additional useful power without the expenditure of excess fuel, thereby contributing to a net increase in FCE of the engine-ORC (EORC) con�guration. For example, at 20 BTDC the FCE for baseline operation (0% EGR) is approximately 19 percent. With waste heat recovery using ORC turbocompounding, the FCE increases to approximately 22 percent, and  � nally with ORC and EGR addition (22% EGR  þ  ORC), the FCE increases to approximately 26 percent. Again, similar observations can be made for 60 BTDC, where the baseline FCE of 20 percent is observed to increase to 28 percent with 22% EGR addition, and a further increase in FCE to 32 percent is observed with ORC turbocompounding and EGR addition (22% EGR  þ  ORC).

K.K. Srinivasan et al. / Energy 35 (2010) 2387 e 2399

 

2395

7.2. Pinch Point temperature difference (PPTD) and exergy ef  �ciency: effects of EGR and ORC turbocompounding 

Fig. 6.  Variation of the engine and combine engine-ORC fuel conversion ef �ciency with the injection timing for different EGR for half load operation (Power  21 kW, ORC evaporator effectiveness, 3  0.7). ¼

 ¼

As shown in Fig. 6, for half load operation (21 kW) the maximum amount of EGR was limited to 8%. Higher amounts of EGR led to engine knock and unstable engine operation. At the representative injection timings of 20 BTDC and 60 BTDC, addition of EGR (8% EGR) results in FCE increasing from baseline values of 26 percent and 30 percent, respectively, to 27 percent and 35 percent, respectively. Moreover, with the addition of EGR, as observed for the quarter load case, the range of injection timings for stable engine operation was extended from 20 BTDC e  60  BTDC to 20 BTDC e  75  BTDC. Exhaust waste heat recovery using ORC turbocompounding led to further improvements in FCE values: from 26 percent to 32 percent at 20 BTDC and from 30 percent to 40 percent at 60 BTDC. Overall, the results show that exhaust waste heat recovery using ORC turbocompounding seems to be more bene�cial for quarter load operation in comparison to half load operation. This is attributable to the lower exhaust gas temperatures in the quarter load case, which led to higher exergy ef �ciencies in the ORC evaporator.

Fig. 7.   Variation of the PPTD with the injection timing for different EGR for quarter load and half load operation ( 3  0.7). ¼

Fig. 7   below discusses pinch-point temperature differences (PPTD) for quarter and half load operation with exhaust waste heat recovery using ORC turbocompounding at various EGR substitutions and for injection timings between 20 BTDC and 75 BTDC. Since the quarter load and half load trends are similar, only quarter load trends are discussed for brevity. For quarter load operation at all injection timings, as the EGR substitution is increased from 0% to 22%, the PPTD is observed to increase. This suggests that with increasing EGR substitution the potential to achieve higher heat transfer rates between the exhaust gases and the organic �uid (R113) increases; however, the heat transfer now has to occur across a greater temperature difference. This presents an interesting cost-ef �ciency tradeoff in heat exchanger design. Pinch point temperature difference is de �ned as the difference between the exhaust gas temperature within the ORC evaporator and the temperature at which the organic  � uid  � rst begins to vaporize (see Fig. 4). This is the smallest temperature difference in the evaporator (ORC heat exchanger) and it de�nes the performance limits of the ORC heat exchanger. The heat transfer rate across the ORC heat exchanger is proportional to the PPTD. As PPTD increases, the mass �ow rate of the organic �uid decreases, which results in poor utilization of the exhaust energy. On the other hand, to accomplish heat transfer across smaller PPTD values, larger heat exchanger areas are required. This leads to larger and more expensive heat exchangers. But, as mentioned above, the exergy ef �ciency of heat transfer across a smaller temperature difference is much higher (i.e., this leads to lower exergy destruction). This fact is clearly illustrated in Fig. 8, which shows the exergy ef �ciency versus PPTD for all injection timings at quarter and half load operation across various EGR substitutions. This �gure indicates that the exergy ef �ciency decreases (almost linearly) with increasing PPTD and EGR ratios across the entire injection timing range for both half and quarter load operation. It is particularly interesting to observe that despite the wide range of injection timings, exhaust gas substitutions, and loads, all curves seem to collapse into one unifying trend signifying that the exergy ef �ciency is only a function of PPTD across the evaporator. Despite the greater exergy destruction with increasing EGR substitution, addition of EGR has been demonstrated to improve combustion stability, improve fuel conversion

Fig. 8.   Variation of the evaporator exergy ef �ciency with the PPTD for different injection timings and EGR for quarter load operation and half load operation ( 3  0.7).  ¼

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K.K. Srinivasan et al. / Energy 35 (2010) 2387 e 2399

Fig. 9.   Variation of the exhaust gases temperature after the evaporator with the injection timing for different EGR for quarter load operation and half load operation (3  0.7).  ¼

ef �ciency and reduce emissions in a previous study conducted by the authors   [37]. Moreover, addition of hot EGR also helps in resolving a major practical issue e  condensation of water in the evaporator tubing, which is discussed below. Fig. 9   shows the post-evaporator exhaust temperature versus injection timing at a constant evaporator effectiveness of  3  ¼  0.7. It is to be noted that the condensation line , the temperature below which condensation of water would occur in the evaporator tubing, is at approximately 380 K. This temperature correspondsto the dew point temperature for water at 124 kPa, which is the minimum exhaust manifold pressure corresponding to the quarter load engine operation with 0% EGR. Therefore, this pressure was set as the minimum operating pressure of the ORC evaporator. It is important to prevent water condensation in the exhaust to reduce the potential for corrosion in the exhaust manifold. FromFig. 9, it is evident that the evaporator temperatures are above the condensation temperature for all injection timings. Also, at both quarter and half load operation, as EGR increases so does the postevaporator temperature. This indicates that despite the lower exergy ef �ciencies with increasing EGRsubstitutions,it is bene�cial to operate the engine with modest amounts of EGR to prevent water condensation in the evaporator tubing, and achieve higher FCEs while maintaining lower emissions. “

”

Fig. 10.   Variation of the evaporator exergy ef �ciency with the PPTD for 22% EGR for quarter load operation for different evaporator effectiveness ( 3).

At this juncture, it is instructive to discuss the practical consideration of evaporator effectiveness and its impact on exergy ef �ciency and PPTD. Fig.10 shows the variation of exergy ef �ciency and PPTD versus injection timing at the highest EGR substitution of  22% for quarter load. In general, PPTD decreases and then increases slightly, whereas exergy ef �ciency is observed to increase and then decrease slightly when the injection timing is advanced from 20 BTDC to 70 BTDC. In particular, at any injection timing, the lowest PPTD and highest exergy ef �ciency are obtained for the highest evaporator effectiveness of 80%. These trends favor higher evaporator effectiveness that will allow for exergetically ef �cient heat transfer across a given temperature difference. However, another important feature is that for the injection timing range between 40 and 65 BTDC, the PPTD exhibits little variation. To further examine the effects of evaporator effectiveness on PPTD and exergy ef �ciency,   Fig. 11   shows the exergy ef �ciency versus PPTD for various evaporator effectiveness values at the highest EGR substitution of 22% at quarter load operation. For all injection timings, it is clear that as the evaporator effectiveness increases the PPTD decreases. These trends also indicate that higher exergy ef �ciencies are possible by choosing higher evaporator effectiveness values. However, as discussed before, this entails higher cost due to the fact that larger evaporators are needed to facilitate heat transfer across smaller temperature differences. On the other hand, favoring lower evaporator effectiveness presents a practical problem of water condensation in the evaporator tubing. For instance, as discussed previously in  Fig. 9, the post-evaporator exhaust temperatures are just above the condensation temperature of 380 K at 124 kPa at an effectiveness of 0.7; but these evaporator temperatures were observed to fall below the condensation line for an evaporator effectiveness of 0.6. Therefore, evaporator effectiveness is an important design parameter that needs to be selected carefully to achieve good exergy ef �ciencies while simultaneously preventing water condensation in the evaporator tubing. From the aforementioned discussion it is clear that the exhaust back pressure, percentage EGR substitution, and ORC evaporator effectiveness are important design parameters to achieve a practical and thermodynamically ef �cient engine-ORC design to recover exhaust waste heat from an ALPING combustion engine operating at constant speed and different loads.

Fig. 11.   Variation of the evaporator exergy ef �ciency with the PPTD for 22% EGR for quarter load operation for different evaporator effectiveness ( 3).

K.K. Srinivasan et al. / Energy 35 (2010) 2387 e 2399  Table 4 Quarter Load e Speci�c Emissions.

Injection Timing

Engine (E)

Engine-ORC (E-ORC)

W (kW)

NOx (g/kWh)

CO2 (g/kWh)

W (kW)

NOx (g/kWh)

CO2 (g/kWh)

EGR  ¼  0% 20 30 40 50 55

10.47 10.64 10.68 10.59 10.53

0.65 2.13 1.59 0.21 0.08

568.10 536.77 531.63 536.63 554.61

12.73 12.72 12.69 12.60 12.64

0.53 1.75 1.31 0.18 0.07

467.59 441.80 437.57 441.69 456.49

EGR  ¼  11% 20 30 40 50 55 60 65 70

10.78 10.71 10.49 10.61 10.68 10.74 10.64 10.70

0.33 1.53 2.50 1.09 0.30 0.08 0.07 0.05

642.52 576.74 575.51 576.55 581.63 588.65 601.80 622.22

13.12 12.73 12.42 12.52 12.59 12.74 12.70 12.86

0.27 1.26 2.06 0.90 0.25 0.06 0.05 0.04

528.85 474.70 473.69 474.55 478.73 484.51 495.33 512.14

EGR  ¼  21% 20 30 40 50 55 60 65 70

10.55 10.64 10.46 10.46 10.58 10.63 10.54 10.73

0.18 0.81 1.87 0.88 0.41 0.09 0.05 0.04

646.40 581.12 593.70 589.86 585.39 588.27 595.03 598.48

12.56 12.34 12.06 12.06 12.19 12.28 12.23 12.48

0.15 0.67 1.54 0.72 0.34 0.08 0.04 0.03

532.04 478.31 488.66 485.51 481.83 484.20 489.76 492.60

 

2397

quarter load are equally valid for half load operation; therefore, only quarter load trends will be discussed. With the addition of  EGR, both speci�c NOx   and CO2   emissions were observed to increase. Addition of hot EGR led to higher intake temperatures, and consequently, the combustion temperatures were higher, resulting in increased NOx  emissions. However, higher CO2  emissions with increasing EGR re�ect better fuel conversion ef �ciencies. In other words, EGR addition favors faster combustion rates, thereby burning the fuel (in this case natural gas and very small diesel quantities) more completely and leading to higher CO2 emissions. In general, total engine power output registered an average increase of 20 percent with ORC turbocompounding (EORC). Similarly, both speci�c NOx  and speci�c CO 2  emissions were observed to decrease by 18 percent on an average with E-ORC. It is important to note that without any expensive aftertreatment technologies, it is possible to achieve a signi�cant reduction in engine-out   brake speci�c emissions by employing exhaust waste heat recovery techniques such as ORC turbocompounding. This is a consequence of accounting for total (engine plus ORC) power output in the calculation of brake speci �c emissions as explained below.

ðBSNOx Þeng-orc ¼

ðBSCO2 Þeng-orc ¼

_ m NOx _ W  eng

_ m CO2 _ W  eng

(18)

(19)

7.3. Brake speci �c emissions: effects of EGR and ORC  turbocompounding 

With ORC turbocompounding, brake speci�c emissions are calculated as follows:

Tables 4 and 5 are summary comparisons of engine performance and speci�c NOx  and CO2  emissions with and without ORC turbocompounding at quarter and half load operation for all injection timings and EGR substitutions investigated. The observations for

ðBSNOx Þeng-orc ¼

 Table 5 Half Load e Speci�c Emissions.

Injection Timing

Engine (E)

Engine-ORC (E-ORC)

W (kW)

NOx (g/kWh)

CO2 (g/kWh)

W (kW)

NOx (g/kWh)

CO2 (g/kWh)

EGR  ¼  0% 20 30 40 50 55 60 65

21.13 21.02 21.07 21.03 21.07 21.09 21.01

0.65 1.61 1.75 0.36 0.10 0.04 0.03

461.90 437.77 441.05 442.24 443.24 444.45 453.23

24.97 24.49 24.42 24.33 24.38 24.50 24.54

0.53 1.32 1.44 0.29 0.08 0.03 0.02

380.19 360.32 363.02 364.00 364.83 365.82 373.05

EGR  ¼  5% 20 30 40 50 55 60 65 70

20.84 20.82 20.73 20.77 21.08 21.00 20.96 20.70

0.53 1.68 1.94 0.56 0.22 0.05 0.03 0.03

495.93 466.94 472.68 467.98 467.27 464.74 475.20 483.70

24.70 24.29 24.05 24.07 24.39 24.31 24.39 24.30

0.43 1.39 1.59 0.46 0.18 0.04 0.02 0.02

408.19 384.33 389.05 385.19 384.60 382.52 391.13 398.12

EGR  ¼  8% 20 30 40 50 55 60 65 70 75

21.20 21.17 21.00 20.74 20.96 20.89 20.95 20.82 20.96

0.29 1.08 1.89 0.98 0.35 0.17 0.05 0.03 0.03

536.10 478.02 479.13 480.40 475.47 478.26 480.26 484.08 491.00

25.35 24.61 24.17 23.83 24.08 24.00 24.09 23.99 23.99

0.23 0.89 1.56 0.81 0.28 0.09 0.04 0.02 0.02

441.26 393.45 394.36 395.41 391.35 393.65 395.29 398.44 404.14

ðBSCO2 Þeng-orc ¼

_ m NOx _ _ W  eng  þ W orc

_ m CO2 _

_ W eng  þ W  orc

(20)

(21)

As seen in Eqs. (20) and (21), the total brake power, which is the denominator of Eqs.   (18) and (19)   increases with ORC turbocompounding, whereas the numerator remains the same. Therefore, the brake speci�c emissions are reduced without any FCE penalty or other changes in the combustion process. Further, it is to be noted that the lowest speci�c NOx emissions at quarter load operation of 0.03 g/kWh and 0.02 g/kWh at half load operation are well below the US EPA NOx   emissions standards for off-highway heavy duty diesel engines. Clearly, exhaust waste heat recovery in addition to the novel ALPING LTC strategy discussed in this study help achieve very low NOx   emissions while maintaining high thermal ef �ciencies. Moreover, the reduction in speci�c CO2 emissions with exhaust waste heat recovery using ORC turbocompounding indicates its potential for reducing greenhouse gas emissions. 8. Conclusions

In this paper, the exhaust waste heat recovery potential from the novel low temperature, low-emissions and high-ef �ciency Advanced Low Pilot Ignited Natural Gas (ALPING) low temperature combustion (LTC) using a bottoming Organic Rankine Cycle (ORC) was examined. The potential improvements in fuel conversion ef �ciency (FCE) and brake speci�c emissions (NOx and CO2) with hot EGR addition and waste heat recovery (WHR) using ORC turbocompounding were quanti �ed. In addition, the nature of irreversibilities in the ORC bottoming cycle was analyzed from a second law perspective and the basic design

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criteria for the ORC evaporator were established using a pinchpoint   analysis. The following conclusions can be drawn from the analyses: “

”

1. With hot EGR addition, FCE improved for all injection timings at half and quarter loads, and ORC turbocompounding along with hot EGR led to even higher ef �ciencies. For example, at 60 BTDC injection timing and quarter load, FCE increased from the baseline (0% EGR) value of 20 percent to 28 percent with 22% EGR  alone, and eventually to 32 percent with 22% EGR and ORC turbocompounding. Half load FCE improvements at 60 BTDC injection timing were more modest (from 30 percent at 0% EGR  to 40 percent at 8% EGR and ORC turbocompounding) and therefore exhaust WHR with ORC turbocompounding appears more bene�cial at quarter load. 2. Engine-out speci�c emissions of NOx  and CO2  can be decreased substantially by adopting ORC turbocompounding without any FCE penalty, combustion modi �cations, or the need for expensive aftertreatment devices. With ORC turbocompounding, both NOx and CO2 emissions were reduced by about 18 percent (on an average). 3. For all injection timings, the pinch-point temperature difference (PPTD) between exhaust gas in the ORC evaporator and the saturation temperature of the organic �uid (R113) increased with increasing EGR, indicating a greater potential for heat transfer. However, irrespective of engine load, injection timing, or percent EGR, higher PPTD values uniformly led to lower exergy ef �ciencies of the ORC evaporator, signifying greater exergy destruction. 4. With increasing hot EGR, exergy destruction in the ORC evaporator increased. However, EGR facilitated more stable engine operation, higher FCEs, and lower emissions. Further, the addition of hot EGR ensured that post-evaporator exhaust temperatures were above the condensation line,  i.e., hot EGR  will prevent water condensation in the ORC evaporator tubing, thus reducing the potential for corrosion in the exhaust manifold. 5. The heat exchanger effectiveness (3) of the ORC evaporator was identi�ed as an important design parameter. In general, higher 3 values (3 ¼   0.8) yielded the lowest PPTD and highest exergy ef �ciencies while lower 3   values (3 ¼   0.6) decreased postevaporator exhaust temperatures below water condensation temperatures (at a given exhaust manifold pressure) and reduced exergy ef �ciencies. However, resorting to higher 3 values may increase heat exchanger costs. Consequently, the ORC heat exchanger effectiveness, exhaust manifold pressures, and percent EGR must be chosen carefully to ensure the best tradeoffs between exergy ef �ciency of the ORC, incremental system cost, system reliability, exhaust emissions, and overall fuel conversion ef �ciencies. “

”

 Acknowledgments

The authors gratefully acknowledge the �nancial support provided by the Micro-CHP and Biofuel Center and the Sustainable Energy Research Center (SERC) of Mississippi State University. References [1] Advanced combustion and emission control technical roadmap for light-duty powertrains. US Department of Energy,   http://www1.eere.energy.gov/ vehiclesandfuels/pdfs/program/acec_roadmap.pdf  ; May 2006. [2] Lior N, Rudy GJ. Second law analysis of an ideal Otto cycle. Energy Convers Manage 1988;28(4):327e34. [3] Dunbar WR, Lior N. Sources of combustion irreversibility. Combust Sci Technol 1994;41:41e61.

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