Waste heat recovery of andiesel engine propelling a

D. Di Battista , M. Mauriello, R. CDepartment of Industrial and Information Engineering a

h i g h l i g h t s

d powe(backprvehicled expaent re

gine and vehicle needs is one of the smartest ways to participate to reduce fuel

eat recoverwas evaluas produced

engine operating point variations. In fact, exhaust gas ow rate and temperature variations lead thorator, even though properly designed, to severe off-design conditions which modify the inlet wuid conditions till to make the mechanical recovery impossible. Under the hypothesis that thepropels a light duty vehicle, the effect of the extra weight is discussed re-evaluating the propulsion powerincrease in terms of fuel consumption.

2015 Elsevier Ltd. All rights reserved.

Corresponding author.E-mail addresses: davide.dibattista@univaq.it (D. Di Battista), marco.mauriello@graduate.univaq.it (M. Mauriello), roberto.cipollone@univaq.it (R. Cipollone).

Applied Energy 152 (2015) 109120

Contents lists available at ScienceDirect

journaconsumption is presented.Finally, once the engine performances were veried, the contribution due to the h

the ORC-based power unit fed by the exhaust gases in terms of mechanical powerexperimentally veried in some points, considering the strong off design conditionhttp://dx.doi.org/10.1016/j.apenergy.2015.04.0880306-2619/ 2015 Elsevier Ltd. All rights reserved.y insideted andby thee evap-orkingengineTurbocharged diesel engineBackpressureORCPlate heat exchangerSpecic fuel consumption

consumption and, therefore, CO2 emissions. In particular, Organic Rankine Cycle (ORC)-based power unitsfed by the exhaust gases are promising and technologically ready, but they have a signicant impact onthe exhaust line and engine behavior.A trade-off between energy recovered in mechanical form and energy lost due to the engine back pres-

sure, vehicle weight increase, discharge energy at the condenser, and the management of the strongoff-design operating conditions is a key point which could denitively open the way to this technologyor limit it to particular applications.The paper discusses the effects of the pressure losses produced by an ORC-based power unit mounted

on the exhaust line of a turbocharged IVECO F1C engine, operated on a test bench. The interactions pro-duced on the turbocharged engine have been experimentally investigated: the presence of an Inlet GuideVane (IGV) system to manage the turbocharger makes the effect of the back pressure not straightforwardto be predicted. The IGV opening and closure degree, in fact, can compensate the effect of the back pres-sure which intrinsically tends to increase specic fuel consumption. A wide experimental testing of theturbocharging group in order to understand its reaction and the net effect in terms of engine specic fuelKeywords:Waste heat recovery

energy and re-use it for en Experimental campaign on a ORC-base Evaluation of the effect on the engine Evaluation of the effect on a light duty Investigation on the heat exchanger an Assessment of the net power improvem

a r t i c l e i n f o

Article history:Received 21 February 2015Received in revised form 20 April 2015Accepted 21 April 2015Available online 15 May 2015ORC-based power unit in a turbochargedlight duty vehicle

ipollonend Economics, University of LAquila, Italy

r unit bottomed an IVECO F1C engine.essure increase).(weight increase).

nder off-design working conditions.lated to ORC unit on light duty vehicle.

a b s t r a c t

The present technologies in internal combustion engines for transportation purposes clearly demonstratethe room for improvement still achievable. In a recent past, harmful emission reduction was the maingoal: wonderful technologies were developed which strongly reoriented the interest and the use of suchengines. Actually, CO2 is the most important driver: it calls for fuel consumption reduction (energy sav-ing) and energy recovery from that usually wasted.Considering that about one third of the fuel energy is in the ue gases, the possibility to recover this

Applied Energy

l homepage: www.elsevier .com/locate /apenergy

km exhaust gas mass ow rate calibration coefcientkp pressure drop calibration coefcientkwall thermal conductivity of the heat exchanger wallsm generic mass ow ratemgas exhaust gas mass ow ratenvanes expander vanes numberp pressuret thicknessv uid velocityDp HRVG pressure dropDT temperature differenceg efciencyn permeability scale factorq density

Subscriptsin inletout outlet

d Energy 152 (2015) 109120mate reasons behind the actual technological research in this cooling optimization (3050 /gCO ), tires and transmissions opti-1. Introduction

Currently, the major global concern of the transportation sectoron the road is the environmental aspect related to propulsion sys-tems. In fact, fuel economy and emissions reduction are the ulti-

Nomenclature

A cross sectionBSFC brake specic fuel consumptionECU electronic control unitEGR exhaust gas recirculatingGWP global warming potentialH enthalpyHRVG heat recovery vapor generatorIGV inlet guided vanesN engine speedNexp expander speedODP ozone depletion potentialOEM original equipment manufacturerORC organic rankine cycleP powerQexch heat exchangedQexhaust exhaust thermal powerS heat exchange surfaceVGT variable geometry turbineVsuc expander suction volumeh heat transfer coefcientkc,1 rst BSFC calibration coefcientkc,2 second BSFC calibration coefcientkc,3 third BSFC calibration coefcient

110 D. Di Battista et al. / Applieeld, and global vehicle efciency is one of the best ways to reachthese goals.

In particular, Euro regulation and carbon dioxide targets areconstraining the recent and future developments in vehicle tech-nologies. The rst one is aimed to reduce pollutants (in particularparticulate matter and nitrogen oxides), while the second one setsat 95 g/km by 2021, the CO2 emission level for passenger cars andlight-duty vehicles (with a reference mass of 1392 kg); these limitsreach 147 gCO2/km for vans (with reference mass of 1706 kg) by2020. More recently, the European Parliament proposed a targetof 68 gCO2/km by 2025. This is very challenging, considering thatthey have to be reached without modifying the traditional drivingexpectations (torque, power, acceleration, fun-to-drive, etc. . .)breaking the relationships between engine and environmental per-formances. Fig. 1 shows how the increase in engine power (also perdisplacement unit) has been accompanied by a fuel consumptionimprovement, leading to more efcient vehicles also keeping agood fun-to-drive feature (0-to-100 km/h time).

The constraints on the CO2 emissions (specic fuel consump-tion) which are rapidly approaching, will push the technologicaladvancement even further: they involve several engine and vehicleaspects. Each of them is characterized by a different carbon inten-sity benet per unit of distance traveled (gCO2 saved/km).

Among all the available technologies, the rst ones which willappear on the market will have a lower specic cost increase perunit of CO2 saved (Euro/gCO2 saved) and minor on-board modica-tions on the engine and vehicle. Downsizing, transmission opti-mization, start and stop [2] are yet settled in vehicles; engineefciency improvement [3,4], thermal management [5,6] andexhaust heat recovery [7,8] have reached a level of developmentto be almost market ready. Knowing that the penalties imposedby the EU regulation is of 95 /gCO2 exceeding the limits, a tech-nology extra cost of about 6070 /gCO2 would be acceptable.This is the main reason why important technologies, likehybridizations, still have difculties to be extensively introducedon the market. On the other hand, technological options like engine

sl saturated liquidsv saturated vaporORC organic rankine cycleWF working uid (R236fa)2

mization (3040 /gCO2) and waste heat recovery (6080 /gCO2)are considered acceptable and a stronger interest is reversed tothem in order to achieve the international targets [9].

With reference to the heat recovery, it is known that the energyin the gases has a magnitude almost equal to the engine mechan-ical energy, representing a big challenge for recovery. Energyrecovery performed with turbo-compounding [1012] and withthermoelectric devices [13,14] are signicant examples but they

Fig. 1. Engine performances trends in transportation sector [1].

d Enrequire important engine modications or cost increase, today toohigh and the reference to technologies not completely proven.Turbo-compounding has been the subject of a strong interestbecause it is able to recover the enthalpy of gases due to a higherpressure as it is actually done to drive the compressor, [15,16].For the intrinsic type of recovery, this technology can be addressedas direct recovery [17]. Main concern of this direct recovery isrepresented by very high speed of revolution of the turbine (dueto the high ow rates) which partially prevents the mechanicaltransformation into electric energy or a direct use in mechanicalform. The suitability is, therefore, limited to the mild or fullyhybrid propulsion systems [18] and to new electric generatortechnologies.

For these reasons, scientic and technical literature has concen-trated a greater attention on the indirect recovery systems basedon Organic Rankine Cycles (ORC) [19,20] power units. The highenthalpy of the exhaust gases (in terms of high temperature) isrecovered inside an evaporator which produces a mediumhighpressure vapor of an organic uid. It is expanded converting theenthalpy in useful work; if expander is of a rotary volumetric type,revolution speeds are low and fully compatible with conventionalelectric generators or with a mechanical use. After the condensa-tion of the exhausted vapor, it is re-pressurized by a pump. Theadvantages of an ORC-based unit for vehicle applications are out-standing. In fact, for many aspects the power unit could be down-sized thanks to suitable technologies for the pump and theexpander [21,22], even though the heat exchangers size still repre-sent a crucial aspect and deserve specic developments. The costincrease can also be reduced thanks to component integration,leading to a value cheaper than other technologies.

In literature, several studies have been concentrated on uidchoice [23,24] for which low ODP and GWP seem to be the mostimportant constraints; suitable operating temperature range havebeen treated considering uid degradation and component simpli-cation. A great attention has been also focused on thermodynamicanalysis of the cycle [25,26]: high efciency is searched, includinga regeneration phase and different thermodynamic options, likedual-loopor transcritical cycles [27,28]. Components choice and siz-ing are, also, investigated in literature: inparticular, heat exchangers[29] and expanders [30,31] are themost worthy of care: their sizingis really important but, in particular in transportation sector, theirresilience to off-design behavior appears as one of the most impor-tant aspects. Engine operations, in fact, are characterized by highow rates and temperature variations which strongly inuence theheat really recovered in theuidor themechanical conversion insidethe expander. Dynamic modeling and optimal control strategy areusually used to predict waste heat recovery performances [32,33],but they still need to be further investigated.

More recently, the integration of an ORC-based power unit withan internal combustion engine in the transportation sector has beenpresented [34]. Main drawbacks are represented by: (a) weightincrease, which produces a propulsion energy increase which canmake nil the energy recovered; (b) back pressure produced by theheat exchanger on the exhaust gases [35]; (c) the frequent off designconditions of all the components (mainly, expander, condenser andevaporator)which can sensibly reduce the energy recovered; (d) theheat discharge devices at the condenser.

As per the authors, all these aspects have never been consideredin a whole system analysis: it is certain that a trade-off betweenmechanical energy recovered and different drawbacks calls for asuitable choice of the design point and for an intelligent (modelbased) control. Single aspects, in reality, have been extensivelytreated. The effects of weight increase on propulsion power and

D. Di Battista et al. / Applieheat exchanger and expanders off design conditions are mattersof basic mechanical engineering. On the other hand, back pressureeffects strongly inuence engine performances [36,37]. Excessiveback pressure in the exhaust system create excessive heat, tur-bocharger problems, increased pumping work, lower engine powerand fuel penalty, that may cause damage of the engine parts andpoor performance [38]. The amount of power lost due to back pres-sure depends on many factors, but a good rule-of-thumb is that30 mbar back pressure causes about 1.0% loss of maximum enginepower [39]. Recent studies on turbocharged engines assumed afuel penalty of 2% having 100 mbar back pressure [40].

In a turbocharged diesel engine, as back pressure increases, thepressure ratio across the turbine decreases and the engine has topumpthegasesoutof the cylinderagainst ahigherpressure, increas-ing the pumping. This would immediately produce a fuel consump-tion increase. In reality, the situation in real engines is morecomplex: the presence of a turbocharging system, which is usuallyoverdesigned in most part of engine operation, separates engineexhaust immediately after the exhaust valves from engine exhaustdownstream the turbine. Modifying back pressure at turbineexhaust and leaving boost pressure unchanged (engine load set bythe ECU which operates on boost pressure), would require a newequilibrium of the turbo-compressor. This behaves differently if awaste gate system is used to control turbine power or if an IGV isoperated.

For this reason, a stronger scientic interest is needed on thisrecovery technology focusing the attention on the important inter-actions produced on the engine and on the vehicle.

The Authors already developed an ORC-based power unit whichis under an extensive testing [41]. It has a unique peculiarityrelated to the very low high temperature source (8090 C) towiden the potential applications and a novel technology for theexpander based on rotary vane sliding machines; rated mechanicalpower recovered is close to 2 kW.

In this paper, an IVECO F1C Turbocharged Diesel Engine 3.0 hasbeen tested on a dynamometer test bench with the presence at theexhaust (after the catalyzer) of a Heat Recovery Vapor Generator(HRVG) which fed an ORC-based recovery unit. Main features ofthe power unit was the use of plate heat exchangers at the HRVGand at the condenser side and the use of rotary vanesmachines bothfor the pump and the expander, as a result of a great research onthese machines [42,43]. Sliding vane rotary machines, in fact, havesame intrinsic advantages and they should be reconsidered in thepanorama of technological choices: they are noiseless, compact,exible from a geometrical point of view (diameter/length ratio),very reliable and do not require important maintenance actions[44,45]. They rotate at conventional revolution speeds (15003000 RPM), suitable for electricity generation and an eventual speedincrease produces a proportional size reduction and weight.

The engine is fully equipped in order to measure engine proper-ties and related relevant components (turbo-compressor speed,rack position, etc.). In fact, the engine presents a variable geometryturbine and its turbocharging group controlled by a variable vane(IGV) in order to adjust the pressure in the exhaust manifold[46,47]: this actuation is much more effective on the turbocharginggroup with respect to the by-pass of a part of the exhaust gasthrough a waste gate. At low gases ow rates (engine speed), theclosure of the inlet guide vane speed up the gases themselves,increasing the reaction produced on the turbine wheel.

In the tested engine, steady working conditions are consideredin terms of torque at given revolution speeds. When a high backpressure occurs at turbine outlet (due to the plate type HRVG),the VGT actuator changes the position of the IGV and a new tur-bocharging speed is reached. Therefore, the boost pressure on itsturn is modied and, denitively, engine characteristics in termsof fuel consumption, emission levels, etc. Moreover, the effects of

ergy 152 (2015) 109120 111the extra weight of the recovery unit was considered, when theengine considered is used to propel a light duty vehicle, in termof Brake Specic Fuel Consumption (BSFC) increase.

In the paper; furthermore, the off design of the HRVG was stud-ied occurring when the engine runs at a different steady point. A1D model was developed in order to predict the thermodynamicconditions of the HRVG at the gas and working uid sides, so eval-uating the inlet conditions of the expander which is operated at offdesign conditions too. This knowledge was useful to preliminarilyevaluating the role of uid ow rate and its capability to insureheat recovery. From a previous experimental activity on anORC-based power unit [21], which had the same inlet conditions

sented by the back pressure produced by the HRVG mounted justafter the catalyst (Fig. 2) in order to preserve the operating temper-

112 D. Di Battista et al. / Applied Enature of the catalyst itself. The ORC-based plant is also composed ofa sliding vane rotary expander (connected to an electric generator),a pump, a water cooled condenser and a regenerator/economizerused to improve ORC efciency (Fig. 2). Working uid used isR236fa operating in the range of pressure of 312 bar.

Table 1Specications of the diesel engine used (IVECO F1C).

Displaced volume 2998 ccStroke 104 mmBore 95.8 mmConnecting rod 255 mmCompression ratio 19:1Number of valves 16Number of cylinders 4 in lineat the expander as those estimated by the HRVGs off design, themechanical power recovered by the expander has been assessed.

In this way, the net effect on the fuel consumption was dis-cussed, contributing to redene the real expectations of thistechnology.

2. The experimental apparatus

The experimental campaign was performed on a dynamicengine test-bench system which consists of a combustion enginedirectly coupled with an electric dynamometer (AVL APA 100)and controlled by a real-time system (PUMA 5.6). The enginetested is the IVECO F1C 3.0 L. It is a turbocharged direct-injectiondiesel engine equipped to comply with the regulatory constraintsEURO IV. The engine is equipped with a short route EGR systemand its turbocharging group includes a variable geometry turbine(VGT), a stator stage that allows to modify the turbocharging speedin order to fulll the boost pressure required by the engine.

The electronic control unit (ECU) is a Bosch ETK P7 type. Thedetailed characteristics of the engine IVECO F1C are listed inTable 1.

The engine was equipped with a series of sensors adapted todetect relevant engine variables, in order to monitor the behaviorof the engine during the experimental tests. Thermocouples, ther-mistors and pressure sensors are installed in all the characteristicpoints of the engine intake and exhaust lines. A gravimetric bal-ance is used to measure the fuel consumption and an air owmeter BOSCH EH-10523; the characteristic parameters of the tur-bocharging (boost pressure vs. engine speed and load), EGR andinjection can be read from the ECU. The turbocharging compressorwas equipped with an AVL Turbo-Speed Sensor TS350 in order tomeasure the speed of the compressor turbine. The measurementsof the inlet and outlet pressure of the compressor and of theturbine allow the denition of the equilibrium point of the group.

The aim of this project is to study the effect of the HRVGmounted on the exhaust line on the engine performances. The linkbetween the ORC-based plant and combustion engine is repre-Maximum power 130 kW @ 3250 RPMMaximum torque 400 Nm @ 2000 RPMAll the heat exchangers (and, in particular, HRVG) are of a plateheat exchangers type. They present a good compactness, the abilityto be modulated (simply adding one or more plates), ease of main-tenance, high exchange efciency, simplicity of assembly andtransport, good resistance to high temperatures and corrosion,suitable shaping in order to simplify the on board installation.

Fig. 3 shows the performances of the HRVG chosen in terms ofpressure losses when it is crossed by a gas: pressure lossesincreases and reach 350 mbar when the gas ow rate reachedabout 500 kg/h. Overall dimensions are 243 393 130 mm witha volume of about 2.7 L.

The high value of back pressure (more than 250 mbar) invites toa choice of an evaporative heat exchanger with greater permeabil-ity. In particular, it is noticeable the quadratic trend of the backpressure with gas ow rate which is closely linked to the speedrotation of the engine.

The quadratic law (Fig. 3), can be expressed as:

Dp n kp m2gas 1

where kp = 1.316 104 kPa/(kg/h)2 represents the best t of theexperimental data of Fig. 3 (n = 1) and n is an adimensional sizeparameter which allows to shape differently the pressure vs. owrate curve. In fact, n = 1 represents the tested plate type HRVGassumed as a reference case. A value of n < 1 represents a familyof more permeable (lower pressure losses for the same gas owrate) HRVG designs which is suitable. This parametric dependencygiven by n will be useful for discerning the fuel consumptionincrease as a function of HRVG back pressure (and type) on thesame engine.

At design conditions, the organic uid considered (R236fa) has apressure of 12 bar, entering in the HRVG at 45 C and exiting at111 C after a superheating. Its mass ow rate is about 131 g/s.

At the same time, the engine operating point considered has aow of gas equal to 0.072 kg/s and a temperature of the gases atthe catalyst of 330 C, this corresponds to an engine operating con-dition equal to 100 Nm @ 2500 RPM. This engine working point,being a mediumlow one, is a characteristic of a homologationdriving cycle of a vehicle (passenger car or light duty). In orderto keep the heat exchanger dimensions under control, a suitablepressure loss has been considered about 60 mbar at designconditions at the working uid side.

The heat exchangers are sized according to the method ofLogarithmic Mean Difference Temperatures which permitted todetermine the heat exchange surface of the heat exchanger. Theheat exchanger was produced by the company EmmeGi S.p.A. withthe characteristics shown in Table 2.

As already observed, the Variable Geometry Turbocharger (VGT)raises the boost pressure even at lower engine speeds, togetherwith the reduction of engine pumping losses at higher enginespeeds, compared with a waste-gated turbocharger. VGT is adevice that can vary the ow area and ow angle between the tur-bine volute and rotor channel. At low engine speeds, making theow passage between turbine nozzle vanes narrower, the exhaustgas approaching the turbine rotor channel is accelerated. By thisacceleration, the boost pressure and, therefore, the charge air owrate is increased. At high engine speeds, the ow rate of the gas ishigher and it is not necessary to speed up the ow to move thecompressor; therefore, the area of the ow passage between nozzlevanes can be larger and it is realized by manipulating the IGV angleadequately. This results in the reduced pumping loss, which is theprimary reason of the lower fuel consumption with VGT at highengine speeds. In a medium speed range, with raising the boost

ergy 152 (2015) 109120pressure and increasing the charge air mass, higher torque can alsobe obtained. But, the performance level in this region is alreadylimited to the maximum cylinder pressure [48].

d EnD. Di Battista et al. / Applie3. Engine experimental campaign and results

The experimental tests on IVECO F1C engine were conducted insteady conditions reproducing the engine characteristic curve at

Fig. 2. Engine scheme with heat recovery

Fig. 3. Experimental curve of HRVG pressure drops vs. gas ow rate mgas.

Table 2Characteristics of the heat exchanger tested.

Hot side Cold side

Number of channels 25 24Capacity 16 kWOverall heat exchange surface 4.46 m2

Heat transfer coefcient 456/117 W/m2 CPlates 50Fouling 6.19 m2 C/kWPinch point 55 Cergy 152 (2015) 109120 11350% of maximum load for different revolution speed: Table 3shows all the engine point tested and, in particular, exhaust massow rates and temperatures of the gas at the inlet of the HRVG.First, tests were carried out in OEM condition and, then, they wereconducted again with the HRVG placed after the catalyst in theexhaust line. Focus has been given at the thermodynamic condi-tions measured in this point. Temperature has been measured byK-Thermocouple and pressure by membrane transducers. Table 3summarizes the increase of the pressure measured in this pointof the exhaust line due to the presence of the heat exchanger: itreaches values above 300 mbar. In particular, this pressureincrease is correlated to the gas ow rate that crosses the exhaustpipeline, which depends on the engine speed and intake manifoldair pressure. Table 3 shows how the back pressure rises accordingto the engine speed, except to the last operating point considered,where the lower torque is related to a lower charge air pressureand, consequently, to a lower air mass ow rate. Therefore, pres-sure increase is reduced, even though engine speed increases.

It is expected that this back pressure increase would imply anextra fuel consumption, in order to keep the mechanical power

vapor generator on the exhaust line.

Table 3Back pressure increase due to the HRVG on the exhaust line in the working pointsconsidered.

Enginetorque(Nm)

Enginespeed(RPM)

Exhaustmass owrate (kg/h)

Exhaust gasestemperature aftercatalyst (C)

Exhaust backpressure (aftercatalyst) (mbar)

112 835 76.3 243.1 12127 1000 91.7 275.6 25187 1250 133.3 336.5 37191 1500 166.8 341.7 52198 1800 239.8 337.1 90202 2000 275.1 340.3 107199 2250 310.1 360.1 125202 2500 336.6 371.8 168200 2750 369.8 373.7 188198 3000 384.7 415.7 209186 3250 417.5 439.9 235173 3500 455.5 460.7 279153 3750 512.7 430.8 335110 3900 454.5 400.2 266

of the engine. An average value of fuel consumption increase isdemonstrated of about 2% (3.95 g/kW h), but it can be very high(till to 5%) at higher engine speeds. The propensity for the BSFCto increase (Fig. 4), which was observed at all operating conditions,was attributed to ineffective blow-down of the engine exhaust,causing more residual gas that remains in the cylinders [49].Consequently, the charging efciency decreases, and the enginemust consume more fuel to give the same mechanical power.

Data on Fig. 3 and on Fig. 4 can be analytically processed inorder to correlate the BSFC with the gas ow rate, thus, ultimately,with the pressure losses across the HRVG (back pressure).Therefore, different families of HRVG represented by different val-ues of the size parameter n can be considered and studied in termsof the BSFC increase that they would produce on the tested engine.

Fig. 5 reports the exhaust ow rates variations as a function ofthe engine speed. Best t of these variations is given by a straightline (Eq. (2)). Exhaust back pressure does not signicantly affectthe mass ow rate.

mgas km N 2In this way, the dependency of the BSFC vs. back pressure can be

analytically expressed as in Eq. (3), and, so, related to the perme-ability parameter n (Eq. (1)).

DBSFC kc;1 Dp2 kc;2 Dp kc;3 3Table 4 species the parameters in Eqs. (3) and (4).n parameter reveals itself very useful because it simply reshapes

pressure losses and it represents a more permeable HRVG, Fig. 6.Taking as reference the tested plate-type HRVG (for which n = 1,Eq. (1)), the n < 1 range reports the effect of a more permeableHRVG on the engine BSFC due to the back pressure produced.The n > 1 range represents the behavior of a less permeableHRVG (greater pressure losses produced with respect to the actualdesign). It is evident that for engine speeds less than 2000 RPM, theeffect of the back pressure on the BSFC is below 1.5%. For enginespeeds greater than this value, the back pressure effect starts tobe great and makes nil the benets of the energy recovered bythe ORC-based unit. Fig. 6, therefore, can be used to evaluate thenegative effect produced by the back pressure on the engine testedand should be taken into consideration to evaluate the net effect ofthe energy recovery unit ORC based.

The back pressure produced by the HRVG rises back through allthe exhaust line from the catalyst till to the turbine (Fig. 2). Thisback pressure increase, however, seems to affect to a lesser extentthe turbine inlet (Fig. 7): relative pressure differences between thetwo cases are very low till to 1800 RPM, only at high speed pointthe pressure increases signicantly. This signies that, in someway, the VGT is able to rearrange its working point and thepressure increase in the engine exhaust manifold is not as highas the back pressure after the turbine. The VGT mitigates the fuel

114 D. Di Battista et al. / Applied Energy 152 (2015) 109120Fig. 4. BSFC increase in presence of the heat exchanger for the operating conditions(50% of the maximum engine load).Fig. 5. Exhaust mass ow rates vs. engine speed: experimental values with andwithout HRVG compared with a linear variation.consumption increase due to the back pressure, but it cannotcompletely compensate this negative effect.

Therefore, the repositioning of the IGV at the inlet of the turbine(just after the exhaust manifold) denes a new equilibrium pointof the engine which corresponds to a different boost pressure ofthe air entering in the cylinders. Fig. 8 illustrates the boost pressureincrease due to the HRVG inside the intake common manifold forthe operating conditions analyzed.

This behavior is, however, so complex: at low engine speeds theboost pressure is quite the same in both cases (with and without

Table 4Calibration parameters of Eqs. (2) and (3).

km 0.1183 kg/h/RPMkc,1 4.993 105 g/kW h/mbar2kc,2 4.687 103 g/kW h/mbarkc,3 1.1005 g/kW hFig. 6. Computed BSFC of a different permeable HRVG based on measured value onthe IVECO F1 engine.

Fig. 9. Comparison between turbocharger revolution speeds difference and rackposition difference between the two cases (positive rack position differencesindicates a more open VGT, while negative ones a more closed VGT).

d Energy 152 (2015) 109120 115Fig. 7. Comparison between exhaust manifold pressure in the two cases for theoperating conditions examined (50% of the maximum engine load).D. Di Battista et al. / ApplieHRVG presence), while at higher engine speeds the systemturbine-compressor increases the boost pressure of the engine,restoring the air mass inside the cylinder when the engine backpressure tends to grow up. This is conrmed by the value of theposition of the VGT actuator (rack): till to 1800 RPM the VGT isfully closed and it is not inuenced by the presence of the HRVG(Fig. 9); therefore, the higher back pressure, due the HRVG, leadsto a lower turbine specic power, slowing the turbocharger group(Fig. 9). On the other hand, at higher engine speeds, the VGT ismore closed, when the HRVG is present, and turbocharger is accel-erated (Fig. 9), producing a signicant boost pressure increase(Fig. 8).

Denitely, when a plate heat exchanger (characterized by thepressure losses as in Fig. 3) is present on the exhaust, engine boostpressure increases till to 100 mbar (depending on the enginespeed), in order to restore the engine torque, turbocharger speedincreases in the most part of the engine speed range even though,due to the IGV rack position, till to 1800 RPM it slightly decreases.

3.1. Weight increase

An additional important issue related to the ORC-based powerunit presence on a vehicle is its not-negligible weight. This is

related to the component themselves (expander, pump, heat

Fig. 8. Boost pressure in the two cases for the operating conditions examined (50%of the maximum engine load).exchangers, piping, uid inside, etc.) but also to the additional radi-ator which is needed in order to deliver outside the heat removedat the condenser. The recovery unit, in fact, re-introduces insidethe engine system the heat that would have been reversedtoward the environment.

Therefore, the estimation of the weight increase depends onmany factors; among them, the size of the power unit and the naldestination of the heat removed from the condenser.

Fig. 10 shows the vehicle propulsive power increase due to theadditional weight of a ORC-based plant placed on board, for thetypical data of a light duty vehicle (Table 5). Propulsive powercan be easily estimated considering the equilibrium of the forceson a vehicle during motion in steady conditions. Vehicle drivingconditions are calculated from the engine working point ofTable 3 (engine torque and speed), considering the proper gearratio of the vehicle considered. The inuence of the weight increaseis much higher at low engine speeds, which are denitely the mostused in real driving cycles.Fig. 10. Propulsive power increase due to the additional ORC-plant weight on thevehicle for a light duty vehicle (engine torque is 50% of the maximum one). Data onthe legend are referred to propulsive power considered for each curve.

For an additional weight of 50 kg, for instance, the propulsivepower increase is of about 0.7% at 3500 RPM and reaches 1.25%

(Eq. (4)). This was done both at the gas and at the working uid

used in single phase uids, while Kandlikar specic plate correla-

can be evaluated and how the temperature difference DT betweenthem proceeds inside the heat exchanger. Therefore, the overallheat exchanged Qexch can be evaluated.

Calculation is completed by pressure drops: at the gas side theyhave been experimentally evaluated, while at the uid side theyhave been calculated according to homogeneous ow Friedelmodel [51]. Hence, the deviations with respect to the pressure dropat design conditions have been evaluated.

Outlet quality x of the refrigerant is dened as in Eq. (5):

x Hout Hsl 5

Table 5Light-duty vehicle reference data.

Vehicle weight 3350 kgDrag coefcient 0.5Frontal area 4.3 m2

Tires radius 0.372 mTires friction coefcient 0.017

116 D. Di Battista et al. / Applied Energy 152 (2015) 109120tion is used for evaporation [50]. Following in counterow thetwo uids, for each elementary piece, the temperature of the uids

Table 6HRVG operating conditions in all the test effectuated.sides, considering a counterow heat exchanger. The cross sectionof the uids passages were known being the heat exchangeralready sized for a specic design ow rate and temperature inlet(at the gas side and at the working uid side).

m qinAinv in qoutAoutvoutm

v2in2 mHin Qexch m

v2out2 mHout

mv in pinAin mvout poutAoutpout pin f m;qin;qoutQexch hgas kwallt hfluid S DT

8>>>>>>>>>>>>>:

4

Starting from the inlet thermo-uid-dynamic conditions (pres-sure pin, density qin, enthalpy Hin, temperature Tin, mass ow ratem) and knowing the heat exchanger geometry and material used(cross sections Ain and Aout, thickness t, thermal conductivity kwall,heat exchange surface S), the convective heat transfer coefcientshgas and huid can be calculated for each HVRG elementary pieceaccording to explicit correlations: DittusBoelter correlations areat 1000 RPM. This is paid in terms of fuel consumption and emis-sions increase.

3.2. Heat recovery vapor generator off-design analysis

A correct evaluation of the power recoverable from the gasesmust consider the off design behavior of the recovery unit, mainlyof the HRVG and of the expander. This is because of the strong owrate and temperature variations of the gases during engineoperation.

In order to take into account these principal effects (gas owrate and gas temperature), a 1D model of the HVRG was consideredmaking reference to a plate heat exchanger properly sized: Theuid path has an evident 1D nature and it was divided into ele-mentary pieces in which mass, momentum and energy conserva-tion equations in steady conditions were simultaneously solvedTest # 1 2 3 4 5

Exhaust mass ow rate g/s 21.2 25.5 37.0 46.3 66Exhaust pressure drop bar 0.283 0.284 0.255 0.256 0.Exhaust inlet temperature C 242.9 268.0 348.0 342.5 34Exhaust outlet temperature C 46.3 46.1 51.0 51.4 54Exhaust thermal power exchanged kW 4.6 6.3 12.3 15.1 21R236fa mass ow rate g/s 131.2 131.2 131.3 131.3 13R236fa inlet temperature C 45 45 45 45 45R236fa outlet temperature C 67.4 74.3 78.7 78.7 78R236fa outlet quality Fraction 0.15 0.06 0.40 0.55 0.R236fa outlet density kg/m3 1202 1168 194.5 146.8 91Heat exchanger effectiveness % 96.7 93.5 98.8 97.1 95Hsv Hsl

where subscript sl is saturated liquid and sv is saturated vapor.Therefore, negative quality means subcooled uid, while x > 1means a superheated vapor.

The HRVG off design was calculated using:

(a) From the gas side, the results of the experimental character-ization of the engine in terms of ow rate and temperature(fourteen points, Table 6); the design conditions were72 g/s and 330 C as ow rate and inlet temperature.

(b) From the refrigerant side the design conditions were 45 Cand 131 g/s as temperature inlet and ow rate, respectively.They have been considered constant during the off designconditions at the gas side; evaporating pressure is about12 bar.

Table 6 summarizes the HRVG off-design behavior when therefrigerant mass ow rate is equal to 131 g/s. In particular, the out-let quality of the R236fa gives the status of vaporization of theorganic uid.

At the lowest conditions (tests #1 and #2) in terms of gas owrate and temperature, the vaporization does not take place: only apart of the economization of the uid is reached (without reachingthe saturation liquid state). When ow rates increase and temper-ature too (tests #3 and #4), a liquidvapor saturated mixture isproduced while for a further increase (tests #6 to #10 and #14)a good range of superheating is reached. For an additional increase(tests #11 to #13) the superheating becomes too high, incompati-ble with the degradation of the organic uid (temperature limit isclose to 230 C).

In particular, Fig. 11 shows the thermal exchange diagram ofthe HRVG close to the design conditions (test #6): the organic uidis completely evaporated and reaches a superheating of 10 C.

In Fig. 12 a low thermal power exchanged case is represented(test #4): the evaporation is just started when the organic uidexits from the HRVG.

Fig. 13 shows a case where the power exchanged by the gas isvery high: the result is that the organic uid reaches a great valueof superheating, very close to the degradation point of the uid.

6 7 8 9 10 11 12 13 14

.6 76.4 86.1 93.5 102.7 106.9 116.0 126.5 142.4 126.3265 0.258 0.267 0.286 0.289 0.314 0.348 0.362 0.379 0.3256.0 358.0 370.0 383.0 385.0 436.0 456.8 470.1 423.0 389.0.0 64.5 69.3 72.9 77.6 83.2 91.4 101.0 102.3 88.6.9 25.4 29.5 33.2 36.3 43.7 49.4 54.8 53.5 44.11.3 131.3 131.3 131.3 131.3 131.2 131.3 131.2 131.2 131.2

45 45 45 45 45 45 45 45 45.7 111.2 140.0 158.8 185.5 229.5 232.2 232.2 232.2 230.894 1.34 1.63 1.84 2.10 2.58 2.61 2.61 2.61 2.59

.6 70.0 61.6 57.5 52.8 46.8 46.4 46.4 46.4 46.6.0 99.4 99.4 100 99.1 98.0 98.6 98.5 97.7 97.5

d EnD. Di Battista et al. / ApplieIt is evident how the designed refrigerant mass ow rate(131 g/s) is suitable only at mediumlow load and speed condi-tions. Data in Table 6 highlights three ranges of working conditionsamong the fourteen points considered: tests from #1 to #4 have alow thermal power exchanged by the exhaust gases and workinguid ow rate is unsuitable, tests from #4 to #8 have a goodbehavior with the designed uid mass ow rate, and tests from#9 to #14 which have a very high thermal power exchanged wouldrequire a greater working uid ow rate.

Hence, three suitable values of the refrigerant mass ow ratehave been detected in order to introduce the need of controllingthe working uid ow rate: for the rst range of working points(#1 to #4) it was reduced by 50%, for the second range (#4 to#8) it was kept as it was designed and, nally, for the third range(#9 to #14) it was increased by 50%. This ow rate, on the otherhand, can be easily changed because it is imposed by the pump

pump speed. Table 7 summarizes the results with these three dif-

Fig. 11. Thermal exchange curves in design conditions (test #6).

Fig. 12. Thermal exchange curves in low power conditions (test #4).ferent ow rates: uid degradation is always avoided and workinguid at the outlet of the evaporator is almost always superheated,so suitable for expansions. Only for the tests #1 and #2 (at verylow engine speed) refrigerant quality is too low for a good expan-sion: an even lower ow rate should be considered but the thermalpower recovered by the gas would be really low.

Finally, results obtained demonstrate that a model basedcontrol strategy applied to the working uid ow rate is indeedsuitable to manage the recovery unit.

3.3. Expander off-design analysiswhich is of a rotary volumetric type so linearly correlated to the

Fig. 13. Thermal exchange curves in high power conditions (test #14).

ergy 152 (2015) 109120 117Also the expander is strongly affected by a variable behavior atoff-design conditions. In particular, if it is entrained at xed speedimposed by an electrical generator, the mass ow rate of the work-ing uidmWF has to change according to the expander inlet densityqWF but this could not match with the HRVG requirements. For theapplication considered, working uid ow rate is known (imposed)and the expander revolution speed has to change according to:

Nexp mWF 60qWFVsucnvanes6

for an ideal vane lling. Inlet uid density remains the onlyvariable to match. Fig. 14 shows how expander speed has tochange in order to match inlet density, at different working uidow rates.

Fig. 14 allows to observe that expander speed should changesensibly in order to match the requirements of the HRVG and thoseof a perfect lling of the expander. At low thermal engine powerrecovered, expander speed should be in the range of 2501500 RPM while at rated design conditions it increases to 20003500 RPM. For a higher thermal power recovered, best expanderspeed is in the range of 40006000 RPM. Dynamic expandersappear not immediately suitable to allow this requirement whilevolumetric machines are more exible. Among them, rotary vanesintroduce further additional advantages in terms of uid manage-ment, robustness, noise, shape factor sizing, reliability and ease ofmaintenance.

mass ow rate (low, design value, high).

5 6 7 8 9 10 11 12 13 14

66.6 76.4 86.1 93.5 102.7 106.9 116.0 126.5 142.4 126.30.085 0.105 0.127 0.145 0.169 0.182 0.216 0.262 0.393 0.261346.0 358.0 370.0 383.0 385.0 436.0 456.8 470.1 423.0 389.054.0 64.5 69.3 72.9 67.7 72.7 77.8 83.8 85.7 76.021.9 25.4 29.5 33.2 37.6 45.2 51.6 57.8 56.6 46.11345780.99195

d Energy 152 (2015) 109120Table 7HRVG operating conditions in all the test effectuated with a three suitable refrigerant

Test # 1 2 3 4

Exhaust mass ow rate g/s 21.2 25.5 37.0 46.3Exhaust pressure drop bar 0.015 0.018 0.034 0.048Exhaust inlet temperature C 242.9 268.0 348.0 342.5Exhaust outlet temperature C 47.4 46.8 50.0 52.5Exhaust thermal power exchanged kW 4.6 6.2 12.2 14.8R236fa mass ow rate g/s 65.6 65.6 65.6 65.6R236fa inlet temperature C 45 45 45 45R236fa outlet temperature C 77.7 78.7 78.7 78.7R236fa outlet quality Fraction 0.15 0.36 1.11 1.39R236fa outlet density kg/m3 409.9 210.7 79.7 68.1Heat exchanger effectiveness % 95.4 95.8 94.6 93.3

118 D. Di Battista et al. / ApplieGas ow rate by pass as well as other strategies which considera multivariable control can match the HRVG and the expander con-straints. Data from HRVG off design conditions in Table 6 wereused to evaluate the performances of an ORC-based power unitpreviously developed by the Authors [21].

Among the wide available experimental data set, those whichmatch closer the conditions presented in Table 6 regarding owrates and inlet thermodynamic conditions at the expander havebeen considered. This procedure appears to be consistent becausethe performances of the power unit are the same, regardless of thehigh temperature source, if ow rate and inlet expander conditionsare kept equal. In this way the mechanical power really recover-able could be estimated as the HRVG was fed with the exhaustgases of the F1C IVECO engine.

Finally, Fig. 15 shows the compared effects on the BSFC of theback pressure, the effect of an extraweight of 50 kg of the ORCplant and the mechanical power recovered by the expander. Therst produced an increase of the BSFC (Fig. 4) as for the second,while the third has a positive effect: the overall mechanical poweris the sum of the engine shaft power and the ORC mechanicalpower recovered. Mechanical power recovered by ORC-plant isevaluated considering the exhaust thermal energy Qexhaust andexperimental overall ORC efciency, considering the expander inoff-design conditions [21] (Eq. (7)).

PORC gORC Qexhaust 7In Fig. 15 is evident how the BSFC reduction due to the mechan-

ical power recovered by the ORC plant is always higher than theBSFC increase due to back pressure effect on the engine. The net

Fig. 14. Expander speed ad a function of the uid density at the expander inlet(expander vanes are served and suction volume is Vsuc is 5.9 cm3).1.3 131.3 131.3 131.3 196.9 196.9 196.9 196.9 196.9 196.945 45 45 45 45 45 45 45 45

.7 111.2 140.0 158.8 124.6 163.5 194.3 222.6 216.9 167.54 1.34 1.63 1.84 1.33 1.71 2.01 2.29 2.23 1.74.6 70.0 61.6 57.5 70.3 59.8 54.2 50.1 50.9 59.2.0 99.4 99.4 100 99.7 100.0 99.9 99.7 99.7 99.8positive effect depends on the engine speed which reports theeffect of the thermal power recovered. Data represented inFig. 15 have been reported in Table 8 in terms of BSFC percentageincrease (or decrease), vs. engine speed. BSFC reference values arethose of the original engine equipment.

From values reported in Table 8 it is clear that the weight of theORC determines an increase of the fuel consumption because of thepropulsion power increase. It tends to decrease with the increaseof the engine speed: Table 8 shows a value close to 1% but a moreprecise estimation depends on the weight increase produced bythe power unit, with respect to the original vehicle weight.

The effect of the back pressure produced on the exhaust line ofthe engine causes a variable trend of the BSFC increase and it isstrongly engine dependent, as discussed previously. Averageincrease value for the testing activity done is close to 3.6% but thisis related to the permeability of the plate heat exchanger tested.Other HRVG types could produce lower values: a reference idealsituation can be estimated considering nil this contribution.

Table 8 demonstrates denitively that a big difference existsbetween potential recovery evaluated by only thermodynamicconsiderations and the real one, which accounts for the present

Fig. 15. BSFC variations due to back pressure increase, 50 kg weight increase onboard vehicle and mechanical power recovered by ORC-based plant.

Table 8BSFC percentage increase/decrease vs. engine speed.

RPM Back pressure(%)

Weightincrease (%)

Mechanicalrecovery (%)

Net (%)

1000 0.19 1.20 1.56 0.171500 0.60 1.06 2.02 0.372000 1.46 0.92 2.70 0.312500 0.89 0.87 3.00 1.243000 0.90 0.78 4.08 2.403500 2.82 0.72 6.32 2.78

loss equal to 350 mbar at 500 kg/h (similar to the onetested), engine boost pressure increases till to 100 mbar

d Endepending on the engine speed. This is required by theengine in order to restore previous engine torque, tur-bocharger speed increases in the most part of the enginespeed range even though, due to the IGV rack position, tillto 1500 RPM it slightly decreases.

(c) The tested counterow HRVG, even though properlydesigned. produces an unacceptable specic fuel consump-tion increase in the range of 25%. Greater values corre-sponds to higher gas ow rates.technology available and for the interferences produced on theengine: the poor values of the recovery are also associated to thestrong off design suffered by the ORC-based power unit whose ini-tial design point must be chosen with care. Considering a negligiblecontribution due to the back pressure, a maximum net benet isclose to 5.5%.

Finally, an additional loss must be considered represented bythe energy required to drive the fan which would guarantee therequired heat exchange toward the environment, if needed.

4. Conclusions

The heat recovery from the exhaust gases in ICE has an interest-ing potential and it is characterized by a cost increase per gram ofCO2 saved very promising with respect to the other technologies.When an ORC-based power unit is considered, four contributionsneed to be evaluated to state the net recoverys benet: (a) theeffect of the back pressure produced on the engine by the HRVG;(b) the weight increase of the vehicle due to the power unit; (c)the power recovered by the unit taking into consideration the offdesign conditions of the HRVG and that of the expander; (d) theadditional power required to discharge the energy extracted atthe condenser toward the environment.

The nal use of the net power recovered as in electrical form orin mechanical form is an additional issue. Making reference to anIVECO F1C 3.0 L turbocharged diesel engine operated in an enginetest bed, the paper discussed:

(a) The effect of the back pressure produced by a real HRVGwhich fed an ORC-based power unit. The presence of anIGV turbine control makes the prediction of these effectsnot straightforward.

(b) The effect produced by the weight increase due to theORC-based power unit.

(c) The recovery done by an existing power unit taking intoaccount the off design situations produced on the HRVGand on the expander by engine speed variation.

The engine was fully sensorized in a way to understand how itreacts to a back pressure, mainly in terms of the turbocharging sys-tem with and IGV control. Main conclusions are:

(a) Engine back pressure has a crucial role and can represent themost sensible interference of the heat recovery system onthe engine. Plate heat exchangers are not suitable for thisapplication: more permeable components are suggested,i.e. shell and nned tube. Limiting the back pressure at175 mbar for a ow rate equal to 500 kg/h, a negligible effectis done on the BSFC (less than 1% in the worst case), thanksalso to the IGV system.

(b) When a plate heat exchanger is characterized by a pressure

D. Di Battista et al. / Applie(d) Considering other heat exchanger types (for instance, shelland nned tubes) the BSFC increase can be kept below 1%at all engine speeds (i.e. engine ow rates).(e) The variation of the exhaust gas ow rate and inlet temper-ature at the inlet of the HRVG produces strong off-designconditions which must be managed with care. If the workinguid ow rate remains constant (at a suitable design value),vaporization could not occur as well as maximum allowabletemperature could be reached implying uid stabilityproblems.

(f) The HRVG off design conditions at the working uid sideentrain the expander to a similar off design which inuencesvolumetric efciency, reducing the mechanical power recov-erable. The effects of both HRVG and expander off-designcan be minimized thanks to a multivariable control in whichthe working uid density, the mass ow rate, the expanderspeed as well as the exhaust gas ow rate can be changed.

(g) 50 kg as extra-weight compared to the original equipmentvehicle, whose weight was 3350 kg (light duty vehicle)appears correctly evaluated standing the present technol-ogy: this issue produces an additional power request whichis equivalent to a loss of about 1% in terms of fuel consump-tion. Extra weight assumed considers also the bigger radia-tor needed to reverse toward the environment the heatexchanged at the condenser.

(h) Gross benet of the ORC-based unit power is of level of 4%5%; additional improvement is possible by an appropriatechoice of the design point of the unit and model basedcontrol.

A nal consideration which further limits the net recovery isrelated to the need to discharge toward the environment the heatremoved at the condenser. The presence of an additional fan mustbe taken into account and when it is operated (low vehicle speedand high external temperature) an additional loss has to be consid-ered. A further limitation is related to the condenser temperaturewhich depends, denitely, on external temperature.

A potential improvement is under development concerning: (a)the use of a more permeable HRVG; (b) the downsizing of thepower unit and its weight reduction: (c) the design of more ef-cient expander; (d) the reduction of the off design effects throughthe choice of a more suitable design point and through a modelbased control of the recovery unit.

Acknowledgements

Meccanotecnica Umbra S.p.A. and Ing. Enea Mattei S.p.A. areacknowledged for continuous technological support and funding.

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Waste heat recovery of an ORC-based power unit in a turbocharged diesel engine propelling a light duty vehicle1 Introduction2 The experimental apparatus3 Engine experimental campaign and results3.1 Weight increase3.2 Heat recovery vapor generator off-design analysis3.3 Expander off-design analysis

4 ConclusionsAcknowledgementsReferences