Potential for exhaust gas energy recovery in a diesel passenger car under European driving cycle

Applied Energy 174 (2016) 201–212

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Potential for exhaust gas energy recovery in a diesel passenger car under European driving cycle Andrés F. Agudelo a, Reyes García-Contreras b, John R. Agudelo a, Octavio Armas b,⇑ a b

Mechanical Engineering Department, Universidad de Antioquia, Medellín, Colombia Escuela de Ingeniería Industrial de Toledo, Universidad de Castilla-La Mancha, Toledo, Spain

h i g h l i g h t s
Potential of waste thermal energy recovery from a diesel passenger car was evaluated.
Tests were carried out following the NEDC at ambient temperatures of 7 °C and 20 °C.
Thermal conditions of the gas were characterized at different points on the exhaust system.
Significant differences on recovery potential were found by energy and exergy analyses.
Feasible recovery potential had a peak value of 6% of the exergy supplied by fuel.

a r t i c l e

i n f o

Article history: Received 25 December 2015 Received in revised form 19 March 2016 Accepted 24 April 2016 Available online xxxx Keywords: Diesel passenger car Waste heat energy recovery Driving cycle Exergy analysis

a b s t r a c t This work addresses the potential for waste energy recovery from exhaust gases in a diesel passenger car mounted in a chassis dynamometer. The New European Driving Cycle was followed, while recording relevant operating variables. Tests were performed under three temperature conditions, and exergy analysis was included to find the potential of exhaust gases to produce useful work at six points in the exhaust system. Results include mean temperature at each point, as well as the energy quality index, which was lower than 33%, meaning that less than one-third of the energy of exhaust gases can be converted into useful work in a recovery system. In general, the highest exergy losses were found in the muffler. Although the greatest recovery potential corresponds to the highest temperature of gases, environmental regulations for vehicles restrict waste energy recovery to be performed downstream after-treatment devices, which, in the present work, was the outlet of the diesel particle filter. Temperature of gases at this location varied in the range 115–320 °C, and potential fuel saving varied between 8% and 19% for the complete driving cycle. Ó 2016 Published by Elsevier Ltd.

1. Introduction World industry is challenged by primary energy availability and environmental concerns, given that about 80% of the world’s energy comes from the combustion of coal, oil, and natural gas [1,2]. The European Commission adopted a target of limiting anthropogenic global warming to 2 °C above preindustrial levels, Abbreviations: DOC, diesel oxidation catalyst; DPF, diesel particle filter; EGR, exhaust gas recirculation; ICE, internal combustion engine; NEDC, new European driving cycle; ORC, organic rankine cycle; PFS, potential fuel savings; PM, particulate matter; SI, spark ignition. ⇑ Corresponding author at: Escuela de Ingeniería Industrial de Toledo, Universidad de Castilla-La Mancha, Av. Carlos III s/n., Real Fábrica de Armas, 45071 Toledo, Spain. E-mail address: [email protected] (O. Armas). http://dx.doi.org/10.1016/j.apenergy.2016.04.092 0306-2619/Ó 2016 Published by Elsevier Ltd.

which implies emission reductions of greenhouse gases of near 50% by 2050, relative to 1990 levels [3]. Energy efficiency is of capital importance among the strategies for mitigating carbon emissions, given its potential contribution to the reduction of growth in energy demand and the consequent reduction of pollutant emissions [4–6]. In particular, transport is fundamental for welfare and economic development of a society, and it represents a significant share in the energy consumption of nations, as well as a major concern in regard to environmental pollution [7–9]. In this context, the search for improvements in energy efficiency of passenger cars is undeniable [10–13], given the great number of cars in urban areas [7,8,14]. In this direction, Euro regulations aim to reach a CO2 emissions target of 95 g/km by 2021 and 68 g/km by 2025 for passenger cars and light duty vehicles (through reduction of fuel consumption) [12].

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Nomenclature a cp e E E_ F Fe Fr g h H m _ m p R s t T V Y Z

coefficient for the calculation of specific heat specific heat (J/kg K) specific exergy (J/kg) exergy (J) exergy rate (W) fuel/air ratio stoichiometric fuel/air ratio fuel/air equivalence ratio acceleration of gravity (m/s2) specific enthalpy (J/kg) enthalpy–energy (J) mass (kg) mass flow rate (kg/s) pressure (kPa) gas constant (J/kg K) specific entropy (J/kg K) time (s) temperature (K) speed (m/s) mass fraction elevation from ground level (m)

Subscripts 0 dead state conditions, initial conditions a air

Since a typical passenger car uses about 35% of the energy available in the fuel [10,15–17], there is a potential for energy saving if part of the heat losses from cooling and exhaust systems [9] is recovered. In diesel engines, about 30% of fuel energy is wasted with exhaust gases [16,18–27]. There are several feasible alternatives to recover energy from the exhaust. The most promising are the production of mechanical power and electricity by Rankine cycles and thermoelectric generators, respectively [9,11,19,23, 25,28–32]. Unfortunately, heat recovery systems might increase fuel consumption, mainly due to back pressure in the exhaust [12,33]. Rankine cycles are more suitable for large vehicles, such as buses and trucks where there is a higher absolute energy potential, and the additional weight and space required are less critical. The study of Rankine cycle technology to recover exhaust energy in diesel engines dates back to beginning of 1970 decade [11,13,19]. This technology requires complex control and is not well suited for transient operation [33]. Reported improvements in fuel consumption by using Rankine cycles vary significantly, in the range 1–16% [11–13,18,19,21,25,34,35], depending on the highest temperature of the cycle and on engine operating conditions. The higher the temperature of gases and engine load, the better the impact of this technology on fuel economy [13]. Thermoelectric generators (TEG) are a promising technology despite their low efficiency compared to conventional power cycles, due to their small weight and size, low maintenance costs, silent operation and high reliability [9,11,26,36–40]. A limitation for the implementation of TEG in passenger cars is the bigger radiator needed to dissipate the additional low temperature heat flow [11]. It is estimated that this technology could drive to a reduction of about 5% in fuel consumption, which might be significantly higher with the use of more advanced thermoelectric materials [11]. Most reported energy recovery studies are carried out under steady state engine conditions [10,12,15,17,19,27,34,41–43], which are not representative of real driving conditions [16]. The final design of the recovery system depends heavily on the actual amount and variability of energy flows [24]. Some works use the

b CV d f F g i j k Loss m Q W

burned gases (combustion products) control volume destruction (exergy) final conditions fuel exhaust gases location of measurement points in the exhaust system constant for calculation of properties generic component losses in the components mean (time average) heat transfer work

Superscripts ch chemical (exergy) F fuel Greek characters d exergy to energy ratio e ratio of exergy of the gases to that supplied with fuel D difference

new European driving cycle (NEDC) to simulate engine operation [33,44], or to obtain steady engine operation points [16], finding fuel savings in the range of 1–4% depending on the recovery technology employed. Fuel savings are estimated to increase significantly for steady highway driving [44]. Location of a recovery system in the exhaust system of a vehicle is relevant, given that heat transfer may affect the operation of pollutant emissions after-treatment devices [20], as they need specific temperature ranges to operate efficiently [42,45,46]. Among reported investigations on waste energy recovery, it is usual to take hot gases immediately after the turbocharger [10,18], or directly from the exhaust manifold in naturally aspirated diesel engines [15,19,27,41,43]. Studies carried out using spark-ignition engines usually take gases after the catalyst [15–17,33,47]. Energy and exergy characterization of exhaust gases at the outlet of the turbine has resulted in significant differences between energy and exergy values [10]. Measurements from steady state [19,27,43,48] and dynamic [49] operating conditions had been used to characterize energy recovery from exhaust gases by means of heat exchangers, finding recovery potentials up to 15% of fuel’s exergy at high engine loads. One reference [50] presents the effect of ambient temperature on exergy efficiency of the engine, finding that this parameter decreases as the temperature increases from 5 °C to 30 °C. Although several researchers have used exergy analysis to study diesel engines, few are focused on energy recovery and none was found to use experimental data from driving conditions. Exergy analysis has proven to be a useful tool to extend the meaningfulness of energy recovery analysis. In general, it is found that there is a significant fuel saving potential with the different waste energy recovery alternatives, mainly at high engine load and speed. This work focuses on determining the potential for waste energy recovery from the exhaust system of a typical diesel passenger car under the current European driving cycle. Tests were carried out at two different ambient temperatures in a climatic chamber, according to homologation conditions, and two initial engine thermal conditions, cold and

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warm start up. In this work the exergy analysis is used to determine the potential of exhaust gases to produce useful work for all exhaust system components. Under the test conditions studied, it was identified that the outlet of the diesel particle filter (DPF) were the most adequate locations for energy recovery, providing a fuel potential saving from 8% to 19%. Also it was determined how much of the fuel’s potential to produce work is transferred to exhaust gases for each exhaust system component and it was quantified how this potential is affected by actual driving cycle at different ambient and engine temperatures.

2. Methodology The experimental work was carried out in a Euro 4 Nissan Qashqai 2.0 dCi light-duty vehicle equipped with a four-cylinder, fourstroke, turbocharged, intercooled, common-rail direct-injection diesel engine (model M1D). Although the vehicle was tuned for Euro 4, it was equipped with the same emission control devices as Euro 5 engines such as cooled exhaust gas recirculation (EGR), diesel oxidation catalyst (DOC), and regenerative wallflow-type diesel particle filter (DPF). The main characteristics of both the vehicle and the engine are shown in Table 1. Tests were carried out with the vehicle placed on the chassis dynamometer reproducing the New European Driving Cycle conditions. The chassis dynamometer is located inside an AVL climatic chamber (see Fig. 1). It consists of a single roller (2WD) with a diameter of 157.48 cm (62 in.). The roller is connected to a nominal power Schenck direct current 65-kW system, which has a base inertia of approximately 1386 kg. Differences between base and vehicle inertias were reproduced electrically by this machine. The rolling and wind resistances, both included in the coast down procedure, were also reproduced electrically by the same machine. Following the procedure of Regulation 83 [51], a wind simulation fan was placed in front of the vehicle to blow air over the car for cooling. The wind simulator controller operates to achieve blower speeds consistent with simulated vehicle velocity. The climatic chamber can set air temperature between 20 °C and 35 °C. Additionally, relative humidity can be set between 30% and 80% with temperatures above 20 °C. Relative fuel-air equivalence ratio was calculated from measured values of air and fuel mass flow rates, both registered by means of the INCA PC software. Inlet air was measured with the Table 1 Main Vehicle and Engine Characteristics. Property

Value

Vehicle Effective frontal area (Cd  A) Weight

0.83 m2 2025 kg

Engine Cylinders Displacement Bore Stroke Max. Power Max. Torque

4 1994 cm3 84 mm 90 mm 110 kW at 4000 min1 323 Nm at 2000 min1

Transmission Type First gear ratio Second gear ratio Third gear ratio Fourth gear ratio Fifth gear ratio Sixth gear ratio Differential ratio

Manual, six gears 3.727 2.043 1.322 0.947 0.723 0.596 4.266

Tyres Code

215/65 R16

203

engine’s hot wire flow meter, and fuel mass flow rate by means of the Electronic Control Unit – ECU of the engine, after its calibration with an AVL 733 s fuel gravimetric system, according to the method developed by Broatch et al. [52]. As Fig. 2 shows, thermocouples type K were installed at the inlet and outlet of each component of the exhaust system (Fig. 2), starting at the turbine outlet. Point 1 corresponds to the maximum temperature of exhaust gases. The exhaust system of the vehicle tested is composed by the following components: oxidation catalyst (DOC), diesel particulate filter (DPF), and muffler (MUF). At point 3, temperature was registered from the ECU. Gas pressure was measured at all points, except at point 3. All measurements were recorded with a frequency of 1 Hz during the NEDC, which lasts for 1180 s (see Fig. 3). This cycle is stablished by the current European regulation for the homologation of passergar cars [51], and it consists of four identical urban cycles (U1 to U4) and one extra urban cycle (EU). Table 2 shows the temperature conditions used in the tests. An initial temperature of 20 °C for both, test cell and engine was selected since this is the lower temperature established in the current European standard for homologation tests, and additionally it is representative of warm weather countries. The other temperature of 7 °C, was selected because, on one hand, it may be representative of cold weather countries and, on the other hand, although it is currently included in the homologation cycle for spark ignition engines, it could be standardized for diesel engines. A third condition was included (test cell at 7 °C, and vehicle at 20 °C) because it represents a warm engine in cold weather. 2.1. Determination of energy recovery potential Energy recovered from exhaust gases in a vehicle is most interesting in the form of useful work or electricity. These high quality energy forms are best addressed by means of exergy analysis. This analysis provides two advantages compared to energy analysis: (i) it allows to quantify irreversibilities, pointing out the sources of inefficiency [10], and (ii) it reveals the true potential for improvement on an energy system [53,54]. Diesel exhaust gases can be considered a mixture of stoichiometric combustion products and air [55]. They can be considered as ideal gases at the pressure levels in the exhaust system [56]. Properties of air depend on its chemical composition and on atmospheric temperature, pressure and relative humidity. Ambient temperature is fixed by the test cell conditions, while ambient pressure is about 95 kPa at the location of the laboratory. The relative humidity (/) was about 10% at T0 = 7 °C, and about 45% at T0 = 20 °C, leading to two possible compositions for air: dry air composition taken from reference [57] and humid air, both calculated from the respective atmospheric condition. Specific heat of air is very similar for both cases (in the range from 250 to 1200 K, it exhibited a maximum difference of 1.2%). Therefore, there is no significant error in taking any of them for the calculations. The resulting polynomial function of temperature is presented in Eq. (1), in which temperature is given in K.

cp;a ¼

4 X aa;j T j

½J=kg K

ð1Þ

j¼0

Constants on this equation are given in Table 3. The resulting gas constant for air is Ra = 288.3 J/kg K. The composition of stoichiometric combustion products is found by means of the CEA (Chemical Equilibrium with Application) software [58,59]. The software was run using the same air composition as for Eq. (1), using T0 = 7 °C, and / = 45%, with commercial diesel fuel, which contains about 5 v/v% of biodiesel. Properties were calculated between 250 K and 1200 K, at 101.0325 kPa.

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AMBIENT AIR TEMPERATURE AND HUMIDITY CONTROL EXHAUST PIPE OF CLIMATIC CHAMBER CLIMATIC CHAMBER

ETAS 591.1

EVAPORATOR

ECU

WIND SIMULATION FAN INCA PC T MUFFLER

DPF

DATA ACQUISITION SYSTEM p, T

DOC

CHASSIS DYNAMOMETER

VEHICLE VELOCITY WHEEL TORQUE

Fig. 1. Scheme of the experimental installation.

Table 2 Temperature conditions used in the tests. Test

Test cell temperature, T0

Engine temperature, TE

1 2 3

7 °C 7 °C 20 °C

7 °C 20 °C 20 °C

Fig. 2. Points of temperature and pressure measurement in the exhaust system. Table 3 Constants for the calculation of specific heat of air. Constant

Value

Units

aa,0 aa,1 aa,2 aa,3 aa,4

1063 0.44603 1.2505  103 9.8802  107 2.7444  1010

J/kg K J/kg K2 J/kg K3 J/kg K4 J/kg K5

polynomial, as presented in Eq. (2). The resulting gas constant is Rb = 288.47 J/kg K.

cp;b ¼

4 X ab;j T j

½J=kg K

ð2Þ

j¼0

Fig. 3. Vehicle speed during the NEDC.

It is reported that the effect of pressure on the properties of equilibrium products is negligible for temperatures below 2000 K [55]. Twelve species were considered for chemical equilibrium. The resulting specific heat of stoichiometric combustion products as a function of temperature (in K) is fitted to a fourth degree

Specific heat and gas constant of exhaust gases are obtained from those of air and combustion products, by means of their mass fraction of burned products, Yb,g.

cp;g ¼ ð1  Y b;g Þ cp;a þ ðY b;g Þ cp;b

ð3Þ

Rg ¼ ð1  Y b;g Þ Ra þ ðY b;g Þ Rb

ð4Þ

This mass fraction depends on the instantaneous values of fuel/air equivalence ratio, Fr, as expressed by Eq. (5):

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Y b;g

  Fe þ 1 ¼ Fr Fþ1

ð5Þ

Fe is the stoichiometric fuel/air ratio of the fuel used, which has a value of 1/14.5, and F is the absolute fuel/air ratio, which is calculated from measured fuel and air mass flow rates. The specific heat of exhaust gases can be written itself as a fourth degree polynomial on temperature (in K):

cp;g ¼

4 X

ag;j T j

contribution is estimated to be lower than 0.6% [64]. These simplifications result in Eq. (9).

eg;i ¼ ðh  h0 Þg;i  T 0 ðs  s0 Þg;i ¼ Dhg;i  T 0 Dsg;i

Under the ideal gas hypothesis, this equation can be written as Eq. (10).

Z

Ti

eg;i ¼

Z

½J=kg K

ð6Þ

The constants in the previous equation can be obtained from Tables 3 and 4, as in Eq. (3):

ag;j ¼ ð1  Y b;g Þ aa;j þ ðY b;g Þ ab;j

ðj ¼ 0 to 4Þ

ð7Þ

Ti

cp;g dT  T 0

T0

j¼0

ð9Þ

cp;g T0

  pg;i dT ½J=kg  Rg ln T p0

ð10Þ

The integrals in this expression can be computed from the specific heat of gases given in Eq. (6). Instantaneous exergy rate of exhaust gases at each location is defined in Eq. (11), by using the measured mass flow:

_ g eg;i E_ g;i ¼ m

½W

ð11Þ

In order to perform an exergy analysis, it is necessary to define previously the dead or reference state. For the case under study, atmospheric pressure is held constant, but atmospheric temperature takes two values, therefore, there will be two dead states, one for each condition: p0 = 95 kPa, T0 = 7 °C, 20 °C. Specific exergy of exhaust gases at location i in a given instant is defined by Eq. (8):

Mean values representing average conditions along the driving cycle are useful to quantify the thermal characteristics of the different measurement points, or parts of the driving cycle. Given that measurements were recorded with a frequency of 1 Hz, mean values of interest variables can be determined by means of time integrals [65–68]. Mean energy and exergy of exhaust gases at each location are calculated by means of Eqs. (12) and (13), respectively.

1 2 eg;i ¼ ðh  h0 Þg;i  T 0 ðs  s0 Þg;i þ ech g;i þ V g;i þ gZ g;i 2

Hm;g;i ¼

ð8Þ

Z

tf

_ g Dhg;i dt m

½J

ð12Þ

t0

The two firs terms on the right of this equation are the thermomechanical exergy, which depend on temperature and pressure differences between the system and the environment. The third term is the specific chemical exergy, which depends only on differences in chemical composition of the system relative to the reference environment. The chemical contribution (ech) is neglected for waste heat recovery systems due to physical and technological limitations. For example, using a representative composition of exhaust gases for a diesel engine (67% N2, 9.123% O2, 12% CO2, 11% H2O, 0.852% Ar, 50 ppm CO, 150 ppm NOx, and 50 ppm THC), the chemical exergy is around 35 kJ/kg, which appears to be significant compared to the specific thermochemical exergy which can reach up to around 200 kJ/kg. Nevertheless, this contribution to total exergy cannot be used in waste energy recovery systems, since it requires the reversible separation of every component from the gas mixture, and the subsequent individual reversible expansion from its partial pressure in the mixture to that of the environment (or vice versa), until chemical equilibrium is reached. Therefore, it would not be realistic to include chemical exergy of exhaust gases in the evaluation of the waste energy recovery potential [60–63]. In regard to kinetic exergy term (1/2 V2), the maximum velocity of exhaust gases is around 25 m/s, which is estimated from instantaneous mass flow and exhaust pipe geometry, leading to a specific kinetic exergy of gases of 0.31 kJ/kg. The third term (Potential exergy, gZ) is associated with the height between the system and the dead state condition (Z0 = 0). The exhaust system of the passenger car used in this work is less than 0.5 m respect to the ground, resulting in a specific potential exergy of around 0.005 kJ/kg. For this, the last two terms in Eq. (8) are also neglected, their

Table 4 Constants for the calculation of specific heat of combustion products.

Subscripts 0 and f for time, indicate start and end of the measurement process, respectively, and allow separating results for any part of the driving cycle, or to obtain whole cycle mean values. Time intervals for integration are presented in Table 5.

Z Em;g;i ¼

tf

E_ g;i dt

½J

ð13Þ

t0

Mean energy at each location is defined with the ambient temperature as reference, in order to be consistent with the definition of exergy. Mean values of exergy give a direct value of the potential for waste heat recovery as useful work from exhaust gases. It shows the maximum theoretical value of useful work (or electricity) that could be obtained from the stream of hot gases, which is lower than the energy associated with them. Exergy to energy ratio is defined using mean values of energy and exergy at each location in the exhaust system, according to Eq. (14).

dm;g;i ¼

Em;g;i  100 ½% Hm;g;i

ð14Þ

This ratio quantifies the fraction of the energy associated with exhaust gases that is of high quality, i.e., the portion that can be converted completely into useful work by an ideal device. Eq. (15) represents the portion of the exergy of the fuel that is available in exhaust gases at a given location. Fuel’s chemical exergy is determined from its lower heating value (LHV), and its chemical composition [69,70]. For the commercial diesel fuel used in this work (contains 5.8 v/v of biodiesel), the chemical exergy was about 7% higher than its LHV.

Table 5 Time intervals used for integration of time-averaged variables.

Constant

Value

Units

Part of the cycle

Time interval (s)

ab,0 ab,1 ab,2 ab,3 ab,4

1029.5 0.018538 5.7214  104 3.9167  107 7.855  1011

J/kg K J/kg K2 J/kg K3 J/kg K4 J/kg K5

U1 U2 U3 U4 EU

0–195 s 196–390 s 391–585 s 586–780 s 781–1180 s

206

eFg;i ¼

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Em;g;i  100 ½% mF  ech F

ð15Þ

Location i in this equation is relevant. If we choose point 1, corresponding to the turbine’s outlet, this index will show the greatest potential for recovery in the exhaust system, since this is the point of maximum gas temperature. However, the after-treatment devices downstream need specific temperature levels to work properly. Therefore, the most suitable location in the exhaust system to place a recovery system is the outlet of the DPF (point 4). This location provides the maximum waste energy recovery potential from the vehicle without compromising its environmental requirements.In the same way, it is possible to define the potential fuel savings (Eq. (16)) associated with exhaust stream exergy. If this useful work were to be produced by the engine, it is necessary to consider its associated exergy efficiency (ge ). A representative value of diesel engine exergy efficiency of 32% can be used [10,50]. In this way, potential fuel savings – PFS, can be determined as follows: F

PFSi ¼

eg;i Em;g;i ½% ¼ ½% ch ge ge mF  eF

ð16Þ

Finally, it is interesting to address how each device in the exhaust system affects the exergy of the flowing gases, because this indicates how they affect the potential for recovery. Exergy balance for a generic device k is expressed by Eq. (17).

  X X dECV _ k _ k þ E_ Q;k þ E_ W;k  E_ d;k ¼ ðmeÞ ðmeÞ dt k out in

ð17Þ

The term on the left of Eq. (17) is the rate of accumulation of exergy inside the control volume, i.e., the rate of exergy transfer from exhaust gases to the device. The first two terms on the right side of this equation account for the rate of exergy transfer associated with mass entering and leaving the control volume. The next two terms on the right are the rate of exergy transfer associated with heat and work transfer of the device, respectively. The latter is non-existent, given that the whole exhaust system is a rigid body and there are no work interactions on it. Finally, the last term on the right is the rate of exergy destruction due to irreversibilities inside the control volume, associated with heat transfer, duct flow, chemical reactions, and mixing. Eq. (17) can be rearranged to obtain the instantaneous rate of exergy loss in the devices of the exhaust system:

  X X dECV _ k _ k ¼ DE_ Loss;k  E_ Q ;k þ E_ d;k ¼ ðmeÞ ðmeÞ dt k out in

ð18Þ

This variable encompasses the rates of exergy accumulation, destruction and that associated with heat transfer of the control volume. Recalling Fig. 2, and by using Eqs. (11) and (13) in the terms of Eq. (18), mean exergy loss in the components can be calculated as expressed by Eqs. (19)–(21):

DELoss;DOC ¼ Em;g;1  Em;g;2

½J

ð19Þ

DELoss;DPF ¼ Em;g;3  Em;g;4

½J

ð20Þ

DELoss;MUF ¼ Em;g;5  Em;g;6

½J

ð21Þ

The previous equations allow determining the quantity of exergy lost in a given device during the driving cycle, which may contribute to optimize the design of energy recovery systems.

3. Results and discussion Fig. 4 shows the fuel mass flow rate for the three tests conditions. At the initial phase of the driving cycle the fuel consumption corresponding to the lower test cell and engine temperatures was higher. This is because at colder ambient temperature the engine needs more energy for the warming up process due to its higher mechanical losses and the less favorable conditions for combustion process. It is also observed that there are no significant differences beyond 400 s of testing time. By integrating the curves in this figure over time, it is possible to calculate the total fuel mass consumption for the cycle. The same procedure allows to obtain the distance traveled by the vehicle by integration of the speed-time profile of the NEDC from Fig. 3. By using both results it is possible to calculate the fuel economy (Table 6) for tests 1, 2, and 3 showed in Table 2. The test at colder weather exhibited the worst fuel economy. It is relevant that the two tests with the engine at 20 °C have similar fuel economy regardless test cell temperature, which proves that engine warm up is a high energy demanding process. Urban driving has worst fuel economy than extra urban driving, as reported in previous works [71,72], because of the colder conditions (higher mechanical losses) and the variability of transient conditions under urban cycles. Instantaneous fuel/air equivalence ratio (used to calculated exhaust gas composition, from Eq. (5)) is presented in Fig. 5. Tests 1 and 2 (with test cell at 7 °C) exhibit similar values beyond 400 s, while test 3, with warmer test cell and engine, exhibits higher values through the driving cycle due to a lesser air mass intake. It was explained by the opening of the EGR valve in this test condition, which reduced air intake to the engine. EGR valve was closed during tests at 7 °C ambient temperature. Mass flow rate of exhaust gases was lower for warmer conditions because of EGR valve opening, as shown in Fig. 6. This impacts energy and exergy rates of exhaust gases. Test for lower test cell temperature exhibit similar mass flow rate of gases. Fig. 7 presents instantaneous temperature values for the six measurement locations in the exhaust system for test 3. Below 200 s the engine is warming up and temperatures are higher for points nearest to the turbine outlet (point 1). As the engine warms up, exhaust system devices accumulate heat, which is transferred to gases during lower temperature periods, leading to eventual higher temperature at the outlet than the inlet of exhaust components. As observed in Fig. 7, the temperature at point 4 (DPF outlet) is ranged from 100 to 200 °C along the time. The instantaneous exergy rate at point 4, calculated by means of Eq. (11), is presented in Fig. 8. It is observed that test 3 has lower potential for exergy recovery than tests 1 and 2, as a consequence of the lower mass flow rate of exhaust gases (see Eq. (4) and Fig. 6). Values vary from less than 1 kW to about 3 kW during urban driving, while during extra urban driving, peak exergy rate of exhaust gases amount to about 10 kW. The greater recovery potential during extra urban driving is due to higher mass flow rate and temperature of exhaust gases (see Figs. 6 and 7). Similar results have been reported by other researchers [16,44]. Exhaust gas temperature is a relevant parameter to select and design a waste heat recovery system. Extreme temperatures at a given point in the exhaust system determine the operating range of a component. This information was identified for the different parts of the driving cycle, and it is presented in Fig. 9 for point 4 (outlet of DPF). Fig. 9 shows the heating process of the engine during the first stages of the total driving cycle, which is more significant for the first urban cycle (U1). After the second urban cycle, the engine is

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207

Fig. 4. Fuel mass flow rate during the driving cycle.

Table 6 Fuel economy during tests. Test

1 2 3

Total driving cycle

Urban driving

Extra urban driving

L/100 km

L/100 km

L/100 km

8.16 7.07 7.28

11.7 9.21 9.34

6.12 5.85 6.1

warmer and exhaust gases leave the DPF in the range of 120– 160 °C. Lower mass flow rate for test 3 (see Fig. 6) results in higher cooling of gases, as they flow through the components of the exhaust system. Fig. 10 presents the mean exergy during the driving cycle at the different measurement points (see Eq. (14)). Exergy of exhaust gases decreases as they flow through the exhaust system as a result of irreversibilities, heating of devices, and heat transfer to the environment. At the outlet of the DPF (point 4), there is a mean potential for recovery as work or electricity of 800 kJ for test 3, while for tests at 7 °C this potential goes up to about 1800 kJ, due to higher mass flow rate. These values represent the true potential for recovery from exhaust gases in the form of mechanical work or electricity. They are the theoretical limit for reversible recovery systems. Differences between the outlet from the particle filter and inlet to the muffler (points 4 and 5, respectively), are a consequence of a particular characteristic of the vehicle used in the tests: the

particle filter is located near the engine, at the front part of the car, while the muffler is located at the rear end of the car. For this reason there is a connecting pipe between points 4 and 5, that is about 2 m long, resulting in significant heat and pressure losses. The exergy to energy ratio of exhaust gases (Eq. (14)) is presented in Fig. 11. It is observed that this ratio remains below 33%, meaning that about one third of the energy of exhaust gases can be converted into useful work, which agrees with the results of previous works [10]. These results indicate that exergy analysis is more suitable than energy analysis to account for the recovery potential: the former shows directly the true amount of energy recoverable as work. In consequence, results from point 4, show that about 22–27% of the energy of exhaust gases can be converted into useful work in a recovery system. Fig. 12 presents the ratio of mean exergy of the gases, at the different locations in the exhaust system, with respect to the fuel exergy in a complete driving cycle. These results are obtained by means of Eq. (15). Results show that less than 10% of the exergy supplied by the fuel is transferred to exhaust gases at the outlet of the turbine, similar results have been reported by other researchers [10,73]. As exhaust gases flow to the atmosphere, they lose the ability to produce useful work because of the irreversibilities associated with heat transfer to the components of the exhaust system, due to flow friction, and to chemical reactions at the catalyst. Test 3 exhibits lower values because of its smaller mass flow rate of gases. Values of test 2 are higher than those of

Fig. 5. Fuel/air equivalence ratio during the driving cycle.

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Fig. 6. Mass flow rate of exhaust gases during the driving cycle.

Fig. 7. Temperature at the measurement points for the test at 20 °C.

Fig. 8. Instantaneous exergy rate at the outlet of the DPF.

test 1, due to the warm engine condition which gives room to more energy to be transferred to exhaust gases. At the outlet of the particle filter, it is observed that there is a potential to produce useful work between 2.6% and 6% with respect to the original potential in the fuel. The value of this ratio on an energy basis is shown in brackets. It can be calculated by using the exergy to energy ratio at each location (see Fig. 11) and the ratio of chemical exergy to LHV of the fuel (1.07). Results on an energy basis are about four

times those on exergy basis, and even greater near the end of the exhaust system, which might suggest that the potential for recovery is very high. Nevertheless, exergy analysis shows that only a fraction of this potential is available to produce useful work, which is essential for the design of a recovery system. Fig. 13 presents the potential fuel savings calculated by Eq. (16). These values express the fuel mass percentage that would be saved if exhaust waste heat was recovered by a reversible device. This

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Fig. 9. Maximum and minimum temperatures at the outlet of the particle filter.

Fig. 10. Mean exergy at the measurement points.

Fig. 11. Exergy to energy ratio of exhaust gases.

parameter represents the maximum theoretical fuel saving, serving as a reference for the evaluation of waste energy recovery. It must be considered that existent devices, such as thermoelectric generators or organic Rankine cycles have irreversibilities and; consequently, actual fuel savings are lower.

Potential fuel savings decrease as gases flow to the atmosphere, because of flow irreversibilities and lower gas temperatures. At the DPF outlet, there are potential fuel savings around 17–19% for tests 1 and 2, and a lower value of 8.1% for test 3 associated with its lesser mass flow rate. In the case of test 2, the potential fuel savings

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Fig. 12. Ratio of the exergy of exhaust gases at different locations along the exhaust system.

Fig. 13. Potential fuel savings.

Fig. 14. Mean exergy loss in the components of the exhaust system.

reached was 18.7% at point 4, where mean exergy of exhaust gases is 1799 kJ (Fig. 10). By using the exergy to energy ratio of Fig. 11, it is found that the mean energy of gases is 6788.7 kJ. Fig. 9 shows that temperature at this location varies from about 160 °C to about 316 °C. If this amount of energy is to be recovered by a thermoelectric generator (efficiency of 4%), then the device would produce

271.5 kJ of electricity, resulting in exergy efficiency of about 15%. The fuel saved in this hypothetical situation can be calculated by the ratio of the useful work and the engine exergy efficiency (32%), resulting in 848.4 kJ of fuel exergy or 18.5 g of mass fuel with chemical exergy of 45,836 kJ/kg. Since the total fuel consumption for the whole driving cycle in test 2 was 655.1 g, then a 2.8% of

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fuel saving is expected. This value can be determined directly as the product of the potential fuel saving (18.7%) and exergy efficiency of the recovery device. Eqs. (19)–(21) are used to obtain the mean exergy lost at the components of the exhaust system, as shown in Fig. 14. The muffler has the greatest exergy loss for tests at the low ambient temperature, reinforcing the idea of recovering exhaust gas energy before this component (at the outlet of the DPF). Exergy loss of the muffler is proportional to the pressure drop of the exhaust gases as they flow through it. Mean pressure drop for tests 1 and 2 is of 1164 Pa and 1261 Pa, respectively, while for test 3, the pressure drop is 656 Pa, mainly due to the lower exhaust gas mass flow rate at these conditions. Pressure drop is enhanced by a drastic change in the flow direction of gases (90°) as they enter the muffler, which is located at the rear end of the car, parallel to the rear wheels shaft. The exergy loss in the catalyst and particle filter is unavoidable to guarantee pollutant emissions reduction. Nevertheless, if these components are thermally isolated, the gases would reach the outlet of the DPF at a high temperature, enhancing the energy recovery potential. 4. Conclusions In this work, the potential for energy recovery from the exhaust gases of a diesel passenger car was determined. Tests were carried out following the current European driving cycle in a climatic chamber, using ambient and engine temperatures of 7 °C and 20 °C. Six points along the exhaust system were instrumented, corresponding to inlet and outlet of the exhaust gas components: oxidation catalyst, particle filter, and muffler. Exergy analysis was used to determine the potential of exhaust gases to produce useful work. EGR valve (activated at 20 °C) reduced air mass flow rate, causing a significant reduction of the potential for waste energy recovery. The best fuel economy was reached under cold environment with the engine warmed up, where the EGR was deactivated. The hottest point in the exhaust system was the outlet of the turbine, where the maximum exergy to energy ratio was about 33%. The fraction of fuel’s exergy transferred to exhaust gases at this location was below 10%. Although the highest potential for exhaust gas energy recovery is at the turbine’s outlet, the most suitable location for a recovery system should be at the outlet of the last emissions after-treatment device to avoid any interference with the pollutant emissions reduction process. According to the specific testing conditions of this work, it was found that the most promising location for waste energy recovery is the particle filter outlet (temperature above 115 °C after engine warm up). At this point, the potential for waste energy recovery as useful work ranged from 2.6% to 6% of the fuel exergy. The exergy to energy ratio of exhaust gases ranged between 22% and 27%, while the potential fuel savings varied from 8% to 19%. The muffler exhibited the highest exergy losses indicating that the implementation of any waste heat recovery system should be located before it. The impact of a colder engine on waste energy recovery is stronger than that of a colder environment. This suggests that under cold weather conditions, it might be more energy efficient to warm up the exhaust system of the vehicle before driving. Acknowledgements Authors acknowledge the financial support of the Spanish Ministry of Economy and Competitiveness to project ENE2014-57043R.

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