Studying combustion and cyclic irregularity of diethyl ether as supplement fuel in diesel engine

Fuel xxx (2013) xxx–xxx

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Studying combustion and cyclic irregularity of diethyl ether as supplement fuel in diesel engine D.C. Rakopoulos, C.D. Rakopoulos ⇑, E.G. Giakoumis, A.M. Dimaratos Internal Combustion Engines Laboratory, Department of Thermal Engineering, School of Mechanical Engineering, National Technical University of Athens (NTUA), Zografou Campus, 9 Heroon Polytechniou St., 15780 Athens, Greece

h i g h l i g h t s " Experimental diesel engine fueled on 24% DEE supplement in diesel, at various loads. " HRR diagrams delayed, pressures, temperatures, heat loss reduced, leaner operation. " Stochastic techniques showed combustion stability with random cyclic irregularity. " Moreover, no effect on cyclic irregularity of injection process or DEE/diesel blend.

a r t i c l e

i n f o

Article history: Received 13 December 2012 Received in revised form 6 January 2013 Accepted 7 January 2013 Available online xxxx Keywords: Diesel engine Diethyl ether blend Combustion Cyclic irregularity Heat release and stochastic analysis

a b s t r a c t An experimental study is conducted to evaluate the effects of using diesel fuel blend with diethyl ether (DEE) 24% by vol., a promising fuel that can be produced from biomass (bio-DEE), on the combustion behavior of a standard, direct injection, ‘Hydra’ diesel engine. Combustion chamber and fuel injection pressure diagrams are obtained at four loads, using a high-speed, data acquisition and processing system. A heat release analysis of the experimentally obtained cylinder pressure diagrams and plots of histories in the combustion chamber of the gross heat release rate (HRR) and other related parameters, reveal some interesting features of the combustion mechanism when using DEE blend. Cylinder pressures and temperatures are reduced, HRR diagrams are delayed, and the engine runs overall a little ‘leaner’ at reduced heat losses, with the DEE blend compared to neat diesel fuel for all loads. Moreover, given the shown low ignition quality of DEE/diesel fuel blend and reports for unstable engine operation at high DEE blending ratios, the strength of cyclic (combustion variation) irregularity is examined as reflected in the pressure indicator diagrams, by analyzing for the maximum pressure and rate as well as dynamic injection timing and ignition delay, using stochastic analysis for averages, coefficients of variation, probability density functions, auto-correlations, and cross-correlation coefficients. The stochastic analysis reveals the randomness of fluctuation phenomena observed in the engine, and the cross-correlation coefficients showed that neither the injection process nor the DEE/diesel fuel blend had practical effect on cyclic irregularity. Ó 2013 Elsevier Ltd. All rights reserved.

1. Introduction Stringent imposed emissions regulations have forced researchers to focus their interest on the domain of engine- or fuel-related techniques [1–4]. Moreover, the ever increasing energy demands in the energy generation and transport sectors, coupled with the limited availability of fossil fuels and their detrimental environmental effects, has guided research to seek alternative fuels for gradually substituting conventional ones [5–7]. Among those, bio-fuels have received increasing attention due to their attractive ⇑ Corresponding author. Tel.: +30 210 7723529; fax: +30 210 7723531. E-mail address: [email protected] (C.D. Rakopoulos).

features of being renewable in nature and reducing the net CO2 emissions, and have been used in both conventional diesel and gasoline engines [8–12]. The share of bio-fuels in the automotive fuel market is expected to grow rapidly in the next decade. In 2009, the new European regulation (Directive 2009/28/EC) introduced new targets for the European Union member states (among those Greece), stating that each state shall ensure that the share of energy from renewable sources in all forms of transport in 2020 is at least 10% of the corresponding final energy consumption [13,14]. In the USA, the environmental protection agency renewable fuel standard version 2 (EPA-RFS2) and the Californian low-carbon fuel standard are driving the US market [15]. The most promising bio-fuels for fossil

0016-2361/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.fuel.2013.01.012

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Nomenclature cv h he m N p Q r R br R

specific heat capacity under constant volume (J/kg K) sampling time interval (s) (sensible) specific enthalpy (J/kg) cylinder charge mass (kg), or maximum lag number number of raw data values pressure (Pa) heat (J) lag number specific gas constant (J/kg K) auto-correlation function of time record

liquid fuels substitutes/supplements are: bio-alcohols and bioethers primarily used for spark-ignition engines, and vegetable oils [16], bio-diesels [17], bio-ethanol [18–20] and bio-butanol [21–24] mixed in small proportions with diesel fuel for diesel engines. Works originating from this laboratory studied the performance and emissions behavior of the present single-cylinder, standard diesel engine, fueled with blends of diesel fuel with the most promising of those bio-fuels, such as: vegetable oils and bio-diesels of various origins [13,25], ethanol [26], n-butanol [27], or diethyl ether (DEE) [28], and with blends of cottonseed oil and its bio-diesel with either n-butanol or DEE with no diesel fuel at all [29]. The above investigations were extended on a six-cylinder, turbocharged, direct injection, ‘Mercedes-Benz’ bus diesel engine used by the Athens Urban Transport Organization, fueled with blends of diesel fuel with vegetable oils and bio-diesels [30,31], ethanol [32], or n-butanol [33]. The lowest carbon-chain ether, dimethyl ether (DME), CH3OCH3, has been experimented as an ignition-improving additive or replacement in diesel engines with success for lowering smoke and nitrogen oxides emissions [34,35]. However, as DME is a gaseous fuel, its use in vehicles requires some engine fuel injection system modifications [36], while the corresponding fuel delivery infrastructure is not currently suitable for distributing large quantities of gaseous fuels. Thus, a more appropriate fuel (ether) may be diethyl ether (DEE), CH3CH2OCH2CH3, which is a fuel with similar attractive properties to DME for use in diesel engines but in liquid form (at ambient conditions). It can be produced from ethanol, which is produced itself from biomass [26], via a dehydrating process, thus being also a bio-fuel (bio-DEE). DEE has several favorable properties for diesel engines [36], including exceptional cetane number, reasonable energy density for on-board storage, high oxygen content, low autoignition temperature, broad flammability limits, and high miscibility with diesel fuel. Bailey et al. [36] had reported a review of the subject up to 1997 to identify the potential of DEE as a transportation fuel. Even up to date the testing of DEE in diesel engines performanceand emissions-wise is limited to very few works [37–41], which were reviewed by the authors [28]. Thus, it is made obvious that a gap exists for the study of combustion mechanism of this bio-fuel when fueling diesel engines, with the relevant information being rare and incomplete, and with some works reporting adverse behavior at higher DEE/diesel fuel blend ratios or loads. Unlike works [37,39] that did not report any engine stability problems though working up to high DEE/diesel fuel blends (30%) and loads, two works [40,41] reported unstable and heavy smoke engine operation with higher than 15% DEE/ diesel fuel blends. In the light of the above and especially the always shown low ignition quality (higher ignition delay) behavior of DEE/blends (despite the DEE high cetane number [36]) that may give rise to unstable operation [19], a pertinent investigation

b xy R t T V

cross-correlation function between time records x(t) and y(t) time (s) absolute temperature (K) cylinder volume (m3)

Greek symbols H fuel lower calorific value (J/kg) q density (kg/m3) q^ xy sample cross-correlation coefficient u crank angle (deg)

is called for the detailed combustion mechanism and strength of its cyclic irregularity (variability), by examining any cause and effect relationships. Therefore, this work reports the results of systematic experimental investigation on a standard, experimental, four-stroke, single-cylinder, ‘Hydra’, Ricardo/Cussons, naturally aspirated diesel engine, which possesses high versatility and control over the variation of its operating parameters. It is a continuation of previous work [28], where performance and emissions results were presented using various blends of diesel fuel with DEE, examining the influence of varying the DEE/diesel fuel blending ratio (92/8, 84/16 and 76/24). The current work examines the influence of load, the detailed combustion characteristics and the possible driving to unstable engine operation, at various loads, for the highest blending ratio that is more likely prone to cyclic irregularity. Two strong ‘tools’ are used here for treating the experimentally obtained cylinder pressure diagrams, viz. heat release analysis [42] and stochastic techniques [43], which are reviewed briefly in later sections. The stochastic techniques of auto- and cross-correlation functions are powerful, objective, scientific tools for removing the ‘noise’ from signals and uncover any useful harmonics, thus disclosing information on any cause and effect relationship, e.g. here any instability due to fuel low ignition quality or erratic pump operation. Concluding this section, it is to be noted that DEE is an isomer of butanol (the ‘counterpart’ of ethanol), a very promising fuel for which extensive research is carried out at present. It may then be worth stating a brief comparison of the emission-wise behavior for the same conditions and engine, fueled with the same percentages (in diesel fuel) of either n-butanol, reported in [27], or DEE, reported in [28], both by the present group. With increasing percentage of either n-butanol or DEE in the blends, it was reported [27,28] decrease of emitted smoke, nitrogen oxides and carbon monoxide, and increase of unburned hydrocarbons, with no fuel penalty. This is a noteworthy similar behavior of those isomer bio-fuels, showing a remarkable simultaneous decrease in both emitted soot and nitrogen oxides.

2. Experimental engine test facilities, measuring apparatus and procedure Facilities to monitor and control engine variables such as speed, load, water and lube oil temperatures, fuel and air flows, are installed on a fully automated test bed, single-cylinder, four-stroke, water cooled, Ricardo/Cussons, ‘Hydra’, high-speed, experimental standard engine. It has the ability to operate on the Otto (sparkignition) or direct injection (DI) diesel or indirect injection (IDI) diesel, four-stroke principle. Here, it is used as a naturally aspirated, DI diesel engine having a re-entrant, bowl-in-piston

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combustion chamber. It has a cylinder bore of 80.26 mm, a piston stroke of 88.90 mm, a compression ratio of 19.8:1, and a speed range of 1000–4500 rpm. The ‘Bosch’ fuel injection pump has an 11 mm diameter plunger, and the ‘Bosch’ injector nozzle has four holes, 0.25 diameter each. The injector opening pressure is 250 bar, and the injection advance (at pump spill) can be varied from 0° to 40° crank angle (°CA). The engine is mounted on a fully automated test bed and coupled to a ‘McClure’ DC motoring dynamometer, equipped with a load cell for engine torque measurements. Full details can be found in past publications by the authors, e.g. [25,26]. For measuring the cylinder pressure, a ‘Kistler’ 6125B miniature piezoelectric transducer is used, flush mounted to the cylinder head and connected to a ‘Kistler’ 5008 charge amplifier. Also, a ‘Kistler’ 4067A2000 piezoelectric transducer connected to a ‘Kistler’ 4618A2 charge amplifier is fitted on the injector side of the pipe linking the injection pump and injector, to provide the fuel pressure signal. A ‘Tektronix’ TDC (Top Dead Center) magnetic pick-up marker is used for time reference. These output signals are routed to the input of a ‘Keithley’ DAS-1801ST A/D board installed on a Pentium III PC, which can acquire input data at a total throughput rate of 312.5 ksamples/s from up to eight differential analogue inputs, utilizing also dual-channel Direct Memory Access operation. Control of this high-speed data acquisition system is achieved by a developed computer code based on the ‘TestPoint’ control software. The conventional diesel fuel was supplied by Aspropyrgos Refineries of the ‘Hellenic Petroleum SA’, representing the typical, Greek automotive, low sulfur (0.035%) diesel fuel (gas oil). The diethyl ether (DEE) (otherwise called ‘ethyl ether’ or more simply ‘ether’) was purchased from local commercial representatives certified to a purity of 99.7% (analytical grade), and was blended with the normal diesel fuel. Preliminary solubility evaluation tests with blending ratios up to 30/70 proved that the mixing was excellent with no phase separation for a period of days, thus requiring no emulsifying agent. The properties of diesel fuel and DEE are shown in Table 1. The density of the 24% DEE blend used was measured at 0.810 kg/m3. It is true that addition of a low viscosity fuel (cf. values in Table 1) to diesel fuel, such as DEE or ethanol, can reduce lubricity and create potential wear problems in sensitive fuel pump designs [20]. Thus, reduction of lubricity is one of the reasons for keeping low their percentage in the blends, apart from the effect of reduced viscosity on spray. In previous work [28], performance (brake specific fuel consumption and thermal efficiency) and regulated emissions results were reported at full load, for blends of diesel fuel with 8%, 16% and 24% (by vol.) of DEE. Here, detailed combustion analysis and stability results are presented for the highest 24% blend, denoted hereafter and in the figures as DEE24-D. The engine is working at the same speed of 2000 rpm and static (pump spill) injection timing of 29 °CA before TDC, at various loads, viz. no-load, low load, medium load and high load, corresponding to brake mean effective pressures (b.m.e.p.) of 0.00, 1.40, 2.57 and 5.37 bar, respectively. Owing to the differences among the lower calorific values and oxygen contents of the fuels, the comparison is effected at the same b.m.e.p., i.e. load, and not injected fuel mass or air–fuel ratio. Combustion chamber (indicator) and injector pressure diagrams are obtained, where pressures are measured with accuracy better than within ±1% of full-scale output, while the accuracy of the analogue input readings of the data acquisition system is within ±0.01%. These pressures are directly measured quantities (generic) possessing inherently the inaccuracy of the piezoelectric transducers stated, which form the ‘seeds’ for the computations of the various heat release and stochastic analysis parameters. The present test engine installation is a standard, versatile, experimental one with very accurate instruments and controls to keep

Table 1 Properties of diesel fuel and diethyl ether (DEE). Fuel properties

Diesel fuel

Diethyl ether CH3CH2OCH2CH3

Density at 20 °C (kg/m3) Cetane number Lower calorific value (MJ/kg) Kinematic viscosity (mm2/s) Bulk modulus of elasticity (bar) Boiling point (°C) Latent heat of evaporation (kJ/kg) Oxygen (% weight) Stoichiometric air/fuel ratio

837 50 43 2.6 (at 40 °C) 16,000 180–360 250 0 15.0

713 >125 33.9 0.23 (at 20 °C) 13,000est. 35 355 21.6 11.2

the same speed and load conditions, having also the capabilities of keeping constant the temperatures (lube oil, cooling water, etc.). Then, for experiments conducted in the same day, the repeatability is expected to be very good for the various fuels tested. 3. Background of experimental data heat release analysis In the study of combustion process in diesel engines, an important means to analyze combustion characteristics is the calculation and analysis of heat release rates (HRRs) according to actual measurements of pressures in the combustion chamber [42–44], with a corresponding diagram of the fuel injection pressure assisting towards this side. The experimental cylinder pressure (indicator) diagrams are here directly processed in connection with the pertinent application of the energy and state equations. The results of the analysis for the HRR and other related parameters in the combustion chamber reveal some interesting features, which aid the interpretation of the combustion mechanism associated with the use of DEE/diesel fuel blend in the diesel engine. Towards that side assist also the widely differing physical and chemical properties of DEE against the normal diesel fuel, which forms the ‘baseline’ case. The method of processing the experimental cylinder pressure diagrams and their analysis for heat release has been reported in detail in previous publications, e.g. [26,44]. Thus, only a brief outline will be given below. A recording is made of the cylinder pressure data for ten cycles in a contiguous file, with a sampling rate corresponding to 0.5 °CA. A signal from a magnetic pick-up, simultaneously recorded, indicates the position of the TDC in each cycle. Then, the ‘mean’ of the cylinder (indicator) and the fuel pressure diagrams are obtained, while a ‘light smoothing’ for the pressure signals is applied that is based on performing a four-data points weighted smoothing. This seems to offer reasonable compromise between no-loss of valuable signal information and relatively smooth values for the first derivative of pressure with respect to crank angle. The measured pressure data processed for the heat release analysis concern the closed part of the thermodynamic cycle. A spatial uniformity of pressure, temperature and composition in the combustion chamber (single-zone model), at each instant of time or during a crank angle step or instantaneous cylinder volume, is assumed. By combining the first law of thermodynamics and the perfect gas state equation in differential form for the cylinder gas content, the net heat release rate dQn/du (with respect to crank angle) is derived as [45–47]:

  dV dp pV dm dV dm þp þV  þ he p du du m du du du

ð1Þ

with the perfect gas equation of state pV ¼ mRT

ð2Þ

dQ n cv ¼ du R

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D.C. Rakopoulos et al. / Fuel xxx (2013) xxx–xxx

Thus, the corresponding gross heat release rate dQg/du, which is the energy released from the combustion of fuel is given by:

dQ g dQ n dQ w ¼ þ du du du

ð3Þ

Term dQw/du stands for the rate of heat transferred to the combustion chamber walls, which is calculated by using the formula of Annand [48]. By knowing the fuel lower calorific value, the fuel burned mass rate dmfb/du is computed as:

dmfb 1 dQ g ¼ du H du

ð4Þ

If the differential equations are integrated [26] from the point of inlet valve closing event up to any crank angle, one can obtain the respective cumulative values in the chamber of Qg and mfb. The specific internal energies (sensible part) of the components are given [49] as fourth order polynomial expressions of T. Similar expressions are then derived for the specific enthalpies, heat capacities and their ratio, by applying the thermodynamic relations connecting these quantities for a perfect gas [46]. The mixture properties are then computed by knowing the prevailing gas composition, as calculated by knowing the air and the fuel mass burned mfb up to the point in question [26,49] and the temperature T calculated from Eq. (3). 4. Background of experimental data stochastic analysis An internal combustion engine may display variations in the cylinder pressure from one cycle to another, even under nominally constant operating conditions [50]. Any deviation in the pressure time development reduces the efficiency and reliability of the engine, increases its noise and exhaust-gas emissions, and is one of causes of power fluctuations [51]. Measurements and analysis of cycle-by-cycle variations in spark-ignition engines have been made by many investigators [52,53]. However, it seems that corresponding analyzes for diesel engines have not kept pace, though randomness in the cylinder pressure was known to exist, probably because of the lower strength cycle-by-cycle pressure variations occurring in diesel engines. A short literature review for this phenomenon in diesel engines has been presented in [19], which deals with the engine in hand with ethanol/diesel fuel blends. Wing [54] was the first to deal, in depth, with this aspect of diesel engine operation. His experimental study concerned a multi-cylinder, four-stroke, DI diesel engine having a rotary distributor fuel injection pump, which was suspect and proved to be the culprit of cyclic pressure variations (irregularity). Sczomak and Henein [55] in an extensive experimental investigation on a CFR pre-chamber diesel engine operating with various low-ignition quality fuels, correlated cyclic pressure variations with ignition delay and dynamic injection timing, and pointed out that low cetane number fuels can cause cyclic irregularity in diesel engines. Following the heat release analysis above, the present work focuses on the study of cyclic combustion variations in the engine running with DEE/diesel fuel blend at the same operating conditions. The need for such a complementary study emanates from the reporting in some works (stated in the Introduction) of diesel engine unstable operation with DEE/diesel fuel blends, and more generally motivated by the always reported behavior of those blends presenting higher ignition delay than the neat diesel fuel (cf. also next section), despite the much higher cetane number of DEE [36]. Thus, by showing a low-ignition quality fuel behavior they need to be investigated in that respect according to the find-

ings of Sczomak and Henein [55]. The combustion cyclic variability (irregularity) is tackled here in the way it is reflected in the pressure indicator diagrams, by analyzing for the maximum pressure and pressure rate, dynamic injection timing and ignition delay, using stochastic analysis techniques. For the stochastic analysis a recording is made of the cylinder and fuel measured pressure data for 480 cycles in a contiguous file, with a sampling rate corresponding to 0.5 °CA. In contrast to the previously described HRR analysis, for the stochastic analysis the 480 pressure diagrams (cycles) are used separately (the ‘mean’ is meaningless here), again with ‘light smoothing’, since by definition the parameters drawn from them will form the data record values to be statistically processed. For assessing the errors involved with the number of cycles chosen [43], the variations of the mean value and the standard deviation of the maximum pressures and pressure rates were plotted against the number of cycles, revealing that a number of cycles greater than 400 form a safe limit. By processing the fuel (injection) pressure diagram, the static injection timing (at the injector) was determined at the crank angle where this pressure rises above the almost constant residual in the connecting pipe pressure value, after the (pump spill) injection timing event. The dynamic injection timing was assumed to coincide with the crank angle where the fuel pressure reaches the value of the injector nozzle opening pressure, immediately following the event of static injection timing [26]. The difference between dynamic injection timing and pump spill timing forms the injection delay. By processing the cylinder pressure diagram, the ignition timing was located at the crank angle u where the first derivative of pressure with respect to u changes slope, immediately following the event of dynamic injection timing, going from a negative to a positive value and so presenting a local minimum. The ignition timing was then determined either by using this condition, or by locating the corresponding ‘zeroing’ crank angle of the second derivative of pressure with respect to u, assuming that this signal is ‘smooth’ enough. Note that with every differentiation of the pressure signal the noise-to-signal ratio increases, while if over-smoothing is applied this ‘zero’ point might disappear as being ‘ill conditioned’. The difference between the ignition and dynamic injection timing forms the ignition delay. From the first and second derivatives of cylinder pressure diagrams with respect to u, the crank angles of maximum values of the first derivative of pressure and the pressure itself can be computed, bearing also in mind that they immediately follow the ignition timing and in that order. The following statistical quantities are used for the analysis of the N raw data values ui (i = 1, 2, ... , N) of a time record: averages, standard deviations, and probability density functions, with the Gaussian (or normal) probability density function with the same mean value and standard deviation as that of the data record also computed [56]. For computation of the auto- and cross-correla has been subtions of the parameters involved, the mean value u tracted from each value ui, i.e. the new time history record is  ði ¼ 1; 2; . . . ; NÞ where h is considered xðtÞ ¼ xðt 0 þ nhÞ ¼ ui  u the sampling time interval and n = 1, 2, ... , N. The auto-correlation function is estimated by direct computation after any linear trend removal. For N data values xi (i = 1, 2, ... , N), from a transformed record x(t), the estimated autocorrelation function at the time displacement rh is defined by the formula [56,57]:

br ¼ R

Nr 1 X xi xiþr N  r i¼1

r ¼ 0; 1; 2; . . . ; m

ð5Þ

br where r is the lag number, m the maximum lag number, and R the estimate of the true value Rr at lag r, corresponding to the

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D.C. Rakopoulos et al. / Fuel xxx (2013) xxx–xxx

displacement rh. A normalized value for the auto-correlation b r by R b 0 ; where function is obtained by dividing R

b0 ¼ R b x ð0Þ ¼ 1 R N

N X

x2i ¼ x2

Nr 1 X ¼ xi y N  r i¼1 iþr

and

b yx R

Nr 1 X ¼ y xiþr N  r i¼1 i

ð7Þ

All pressure diagrams in this section are ‘mean-smooth’, which are then processed to produce the other related parameters. They are presented below in the two fourfolded Figs. 1 and 2. Fig. 1a shows at the four loads considered the fuel (injection) pressure against crank angle diagrams for the neat diesel fuel and the DEE24-D blend. First it can be seen that with increasing engine load the injection duration increases (as more fuel is injected)

80

Cylinder pressure (bar)

Fuel pressure (bar)

5. Discussion of the heat release analysis combustion results

400

60

40 Diesel, b.m.e.p.=5.37 bar DEE24-D, b.m.e.p.=5.37 bar Diesel, b.m.e.p.=2.57 bar DEE24-D, b.m.e.p.=2.57 bar Diesel, b.m.e.p.=1.40 bar DEE24-D, b.m.e.p.=1.40 bar Diesel, b.m.e.p.=0. bar DEE24-D, b.m.e.p.=0. bar

200 20

0

0 -40

-20

0

-20

20

-10

0

10

20

30

40

Degrees crank angle

Degrees crank angle

(b)

(a) 2000 Diesel, b.m.e.p.=5.37 bar DEE24-D, b.m.e.p.=5.37 bar Diesel, b.m.e.p.=2.57 bar DEE24-D, b.m.e.p.=2.57 bar Diesel, b.m.e.p.=1.40 bar DEE24-D, b.m.e.p.=1.40 bar Diesel, b.m.e.p.=0. bar DEE24-D, b.m.e.p.=0. bar

30

20

1600

Temperature (K)

40

Gross heat release rate (J/deg.)

ð8Þ

x2 y 2

100

Diesel, b.m.e.p.=5.37 bar DEE24-D, b.m.e.p.=5.37 bar Diesel, b.m.e.p.=2.57 bar DEE24-D, b.m.e.p.=2.57 bar Diesel, b.m.e.p.=1.40 bar DEE24-D, b.m.e.p.=1.40 bar Diesel, b.m.e.p.=0. bar DEE24-D, b.m.e.p.=0. bar

600

b x ð0Þ R b y ð0Þ R

ð6Þ

The maximum value of r should normally be [57] less than 10% of N. The normalization of the cross-correlation function defines the sample cross-correlation coefficient:

800

b R

i¼1

The cross-correlation between two time records x(t) and y(t) at lag numbers r = 0, 1, 2, . . . , m is:

b xy R

b R

xy xy ffi ¼ qffiffiffiffiffiffiffiffiffiffi q^ xy ðrhÞ ¼ qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi

1200

800

Diesel, b.m.e.p.=5.37 bar DEE24-D, b.m.e.p.=5.37 bar Diesel, b.m.e.p.=2.57 bar DEE24-D, b.m.e.p.=2.57 bar Diesel, b.m.e.p.=1.40 bar DEE24-D, b.m.e.p.=1.40 bar Diesel, b.m.e.p.=0. bar DEE24-D, b.m.e.p.=0. bar

10 400

0

0 -10

0

10

20

Degrees crank angle

(c)

30

40

-20

-10

0

10

20

30

40

Degrees crank angle

(d)

Fig. 1. Fuel (injection) pressure (a), cylinder pressure (b), gross heat release rate (c), and cylinder temperature (d) against crank angle diagrams, at the four loads, for the neat diesel fuel and the 24% diethyl ether blend cases.

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D.C. Rakopoulos et al. / Fuel xxx (2013) xxx–xxx Diesel, b.m.e.p.=5.37 bar DEE24-D, b.m.e.p.=5.37 bar Diesel, b.m.e.p.=2.57 bar DEE24-D, b.m.e.p.=2.57 bar Diesel, b.m.e.p.=1.40 bar DEE24-D, b.m.e.p.=1.40 bar Diesel, b.m.e.p.=0. bar DEE24-D, b.m.e.p.=0. bar

0.6

600

Equivalence ratio

Cumulative gross heat release (J)

800

Diesel, b.m.e.p.=5.37 bar DEE24-D, b.m.e.p.=5.37 bar Diesel, b.m.e.p.=2.57 bar DEE24-D, b.m.e.p.=2.57 bar Diesel, b.m.e.p.=1.40 bar DEE24-D, b.m.e.p.=1.40 bar Diesel, b.m.e.p.=0. bar DEE24-D, b.m.e.p.=0. bar

400

0.4

0.2

200

0

0 -20

0

20

40

60

-40

80

0

40

(b)

5000

300

4000

200

Cumulative heat loss (J)

Heat transfer coefficient (W/m2 K)

(a)

3000

Diesel, b.m.e.p.=5.37 bar DEE24-D, b.m.e.p.=5.37 bar Diesel, b.m.e.p.=2.57 bar DEE24-D, b.m.e.p.=2.57 bar Diesel, b.m.e.p.=1.40 bar DEE24, b.m.e.p.=1.40 bar Diesel, b.m.e.p.=0. bar DEE24, b.m.e.p.=0. bar

2000

80

Degrees crank angle

Degrees crank angle

100

Diesel, b.m.e.p.=5.37 bar DEE24-D, b.m.e.p.=5.37 bar Diesel, b.m.e.p.=2.57 bar DEE24-D, b.m.e.p.=2.57 bar Diesel, b.m.e.p.=1.40 bar DEE24, b.m.e.p.=1.40 bar Diesel, b.m.e.p.=0. bar DEE24, b.m.e.p.=0. bar

0

1000

-100

-20

0

20

40

-80

-40

0

40

Degrees crank angle

Degrees crank angle

(c)

(d)

80

120

Fig. 2. Cumulative gross heat release (a), equivalence (fuel–air) ratio (b), heat transfer coefficient (c), and cumulative heat loss (d) against crank angle diagrams, at the four loads, for the neat diesel fuel and the 24% diethyl ether blend cases.

for both fuels and the same holds true for the injection pressures. Further, for each load considered, the DEE fuel pressure diagram is distorted with respect to the corresponding neat diesel fuel one. Specifically, its uprising leg acquires a lower gradient, which is translated into a delay of the dynamic injection timing, and furthermore its maximum value is slightly reduced and its final falling leg delayed.

The different densities ql and bulk moduli of elasticity Kbm of blends influence the whole injection process [58,59] following the simplified analysis of Obert [60]. For a jerk pump when its plunger begins to compress the fluid, a pressure wave is propagated down the connecting pipe, at essentially the speed of sound as = (Kbm/ql)1/2, reaching eventually the injector needle in order to open it. Thus, depending on the values of these properties the

Please cite this article in press as: Rakopoulos DC et al. Studying combustion and cyclic irregularity of diethyl ether as supplement fuel in diesel engine. Fuel (2013), http://dx.doi.org/10.1016/j.fuel.2013.01.012

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Fig. 3. Cyclic variation, as a function of load, expressed as mean values and coefficients of variation (COV) of the maximum cylinder pressure (a), maximum rate of cylinder pressure rise (b), dynamic injection timing (c), and ignition delay (d), for the neat diesel fuel and the 24% diethyl ether blend cases.

dynamic injection timing is affected despite that the pump spill timing is kept constant, as here for all fuel samples tested. The bulk modulus of elasticity of DEE is not known, but is expected to be much lower than the diesel fuel one and near to the ethanol value at around 13,000 bar [28,61]. Using the values of ql and Kbm from Table 1, as is computed as 1382.6 m/s and 1350.3 m/s for the diesel fuel and the DEE, respectively, showing indeed a relatively later arrival of the pressure pulse at the injector needle for the DEE case. Fig. 1b shows, at the four loads considered, the cylinder pressure against crank angle diagrams for the neat diesel fuel and the DEE24-D blend, focusing on their part around ‘hot’ TDC. First it can be seen that the pressures increase with load (with the compression lines remaining the same), while the ignition delay decreases with engine load for both fuels due to the increasing gas temperatures with load. One can observe that for each load considered, the DEE blend start of combustion occurs later (the pressure rise due to combustion starts later) with respect to the corresponding neat diesel fuel one, while its maximum pressure

falls and occurs later. The start of combustion is delayed as a consequence of synergy of the lower dynamic injection timing (cf. Fig. 1a) and increased ignition delay. It is worth explaining this behavior also in conjunction with Fig. 1c, which shows the corresponding gross heat release rate (HRR) diagrams. First it can be seen that the ignition delay decreases with engine load for both fuels (since temperatures increase), while the heat release rate values become higher. For the higher loads, both parts of combustion, i.e. the premixed combustion (the part under the first ‘sharp’ peak) and the diffusion combustion (the last part under the second ‘rounded’ peak), are apparent with the diffusion combustion diminishing with load decrease. One can again observe that for each load considered, the ignition delay for the DEE24-D blend is higher than the corresponding one for the neat diesel fuel case. The increase of ignition delay of DEE when blended with diesel fuel has also been reported early in [35] and by later investigators despite its much higher cetane number than diesel fuel, with possible explanations

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1.2

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provided in [62,28], while the decreasing dynamic injection timing and the higher latent heat of evaporation of DEE (see Table 1) here contribute also towards this side (injection into a lower temperature environment). Further, it is observed that the premixed combustion (area under the first ‘sharp’ peak) of the DEE blend seems to decline against the corresponding neat diesel fuel case, thus leading to lower pressures and temperatures during the initial part of combustion process. Fig. 1d shows the corresponding cylinder temperature diagrams. First it can be seen that there is a temperature increase with engine load [45] for both fuels. One can observe that for each load considered, with respect to the neat diesel fuel case, the temperatures for the DEE24-D blend are lower up to around their maximum values and appear delayed (cf. previous paragraph for the premixed part of combustion), while later on during expansion they seem to recover and even slightly switch over the corresponding diesel fuel ones. The latter is due to the delayed and prolonged (last) part of diffusion combustion (area under the second ‘rounded’ peak in the HRR diagrams). It is reminded here that this is a computed ‘mixed’ temperature due to the inherent single-zone assumptions of the heat release analysis followed. The observed above increase of delay of the fuel pressure and heat release rate diagrams (and consequent fall in cylinder pressures and temperatures) with the use of DEE in the diesel fuel blend, points to the influence on the combustion and emissions formation processes [59,61]. This is effected through a later and slower spray development with possible impingement on the combustion chamber walls [58], apart from any possible poor fuel injection (and so atomization) due to vapor locks because of the high volatility of DEE as mentioned in [40,41]. Fig. 2a shows the corresponding cumulative gross heat release diagrams. One can observe that for each load considered, the cumulative gross heat release curve for the DEE24-D blend lies, at the beginning, a little lower than the corresponding one for the neat diesel fuel case and catches up later on into the expansion stoke, thus revealing the slower rate of combustion as also explained with reference to Fig. 1c above. Then, the corresponding final (almost) equal cumulative gross heat release values are translated into the same brake thermal efficiency, given the constant engine speed and load. Fig. 2b shows the corresponding fuel–air

equivalence ratio (i.e. the actual fuel–air ratio divided by its stoichiometric value) diagrams. One can observe that for each load considered, the fuel–air equivalence ratio curve for the DEE24-D blend lies a little lower than the corresponding one for the neat diesel fuel case. This proves that the engine runs overall a little ‘leaner’ with the DEE24-D blend, at least at the beginning, for the same engine load and speed conditions, noting that the calculation of fuel–air equivalence ratio was made by considering all the fuelbound oxygen. Fig. 2c shows the corresponding gas side heat transfer coefficient (from the cylinder gas to the combustion chamber walls) diagrams. One can observe that these diagrams follow in shape closely the corresponding ones of (cylinder) temperatures (cf. Fig. 1d). This is explained as the gas side heat transfer coefficients are computed from the relevant formula of Annand [48], which is an increasing monotonic function of gas temperature T. It can be easily proved by assuming, for example, variation laws [45] of gas thermal conductivity kgas = T0.75, and dynamic viscosity lgas = T0.62. Fig. 2d shows the corresponding cumulative heat loss (to the combustion chamber walls) diagrams. One can observe that for each load considered, the cumulative heat loss curve for the DEE24-D blend lies a little lower than the corresponding one for the neat diesel fuel case. This is due to the lower cylinder temperatures and heat transfer coefficients encountered with the DEE blend case (cf. Figs. 1d and 2c), as the cumulative heat loss is effectively the integral, over the cycle, of the product of these two quantities.

6. Discussion of the stochastic analysis results of combustion parameters In the figures to follow, results are presented at all four loads considered, and for the neat diesel fuel and the blend of 24% (by vol.) diethyl ether (DEE) in diesel fuel. From the large amount of data collected at each operating condition, only representative sample plots are presented owing to imposed conservation of space. Preliminary tests to determine the extent of cyclic variation in combustion over the load range examined, used both the maximum cylinder pressure and the maximum cylinder pressure rate as measures of the cyclic variation (the ‘effect’). Their variations are

Please cite this article in press as: Rakopoulos DC et al. Studying combustion and cyclic irregularity of diethyl ether as supplement fuel in diesel engine. Fuel (2013), http://dx.doi.org/10.1016/j.fuel.2013.01.012

D.C. Rakopoulos et al. / Fuel xxx (2013) xxx–xxx

1 0.8

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b.m.e.p. (bar) Fig. 5. Correlation coefficients between dynamic injection timing (dyn. inj.) and maximum rate of pressure rise (pr. rate), dynamic injection timing (dyn. inj.) and ignition delay (ign. del.), ignition delay (ign. del.) and maximum rate of pressure rise (pr. rate), and maximum cylinder pressure (pr.) and maximum rate of pressure rise (pr. rate), as a function of load, for the neat diesel fuel and the 24% diethyl ether blend cases.

distinct and obviously have a close reference to the combustion process itself but, in any case, a rather strong degree of correlation exists between those as will be shown in last Fig. 5. The dynamic injection timing was chosen [54] as potential ‘cause’ of any influence of the injection process on the cyclic variation, while the ignition delay was chosen as corresponding potential ‘cause’ of any influence of the fuel [55]. Fig. 3a and b presents the cyclic variation of the maximum cylinder pressure and the maximum rate of cylinder pressure rise, respectively, expressed as mean values and coefficients of variation (COV), i.e. standard deviation divided by the mean value, as a function of the engine b.m.e.p. (load) for the cases of the neat diesel fuel and the 24% addition of DEE in the blend. Fig. 3c and d presents the corresponding cyclic variations of the dynamic injection timing and the ignition delay, respectively. The observed variation (mean values) with either the load or the addition of DEE in the blend has already been discussed with reference to Fig. 1a–c. From Fig. 3a–d, one can conclude, by observing the coefficients of variation (COV) values, that the addition of DEE in the blend, at least for up to 24% DEE, does not practically affect the cyclic variability (irregularity) with respect to the neat diesel fuel case, which in any case is already small. The probability density functions of the experimental maximum cylinder pressure, pressure rate, dynamic injection timing and ignition delay, for the neat diesel fuel and the 24% addition of DEE in the blend cases, followed quite closely the corresponding Gaussian ones (computed) having the same mean value and standard deviation. They showed a slightly different skewness (in the range 0.1 to +0.1) and kurtosis (in the range 0.2 to 0.6) against the corresponding values of zero for the Gaussian ones. Hence, the error of the analysis will be insignificant if a normal distribution is assumed for the purpose of determining the statistical nature of the above four parameters, as has already been tacitly assumed in previous Fig. 3. This implies that the cause of the fluctuations of these parameters is rather random (stochastic) and does not depend on its value of any other cycle, i.e. on any residual effects of previous combustions taken place in the cylinder [43,54].

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Fig. 4a and b shows sample normalized auto-correlation functions of the maximum rate of pressure rise and the ignition delay, respectively, for the cases of the neat diesel fuel and the 24% addition of DEE in the blend, at the high engine load (b.m.e.p. = 5.37 bar). The auto-correlation function for the other engine loads and the other parameters were similar, not exceeding the critical value (0.20) at the 1% significance level. From observation of the autocorrelation values, it is concluded that there is no correlation between the fluctuations of different cycles, thus confirming the same conclusion as of the sample probability density functions discussed above. For examining the influence of the injection process (potential ‘cause’) and the kind of fuel used via its cetane number (another potential ‘cause’) on the cyclic pressure variation, a crosscorrelation analysis was carried out. This computed the degree of correlation between the dynamic injection timing and the maximum rate of pressure rise, between the dynamic injection timing and the ignition delay, and between the ignition delay and the maximum rate of pressure rise. Also, the degree of correlation between the maximum cylinder pressure and the maximum rate of pressure rise is presented only for reference. The reason is that the values of the maximum rate of pressure rise were selected as the measure of cyclic variation (the ‘effect’) in the combustion chamber. Thus, Fig. 5 presents all these correlation coefficients (Eq. (8) with r = 0) for the cases of the neat diesel fuel and the 24% addition of DEE in the blend, as a function of load. It can be observed that there is a minimal to slight correlation of these parameters (absolute values much less than 0.5), with the exception of the expected rather strong (positive) correlation between the maximum cylinder pressure and the maximum cylinder rate of pressure rise [43] that seems to be decreasing with load. All the results of the above analysis indicate clearly that neither the injection process (through the dynamic injection timing), nor the kind of DEE/diesel fuel blend used (through the shown low ignition quality) have any practical effect on the above cyclic variations (irregularity). Therefore, there is no unstable operation of the engine at least for up to 24% addition of DEE. These findings are in accord with works [37,39] that did not report any stability problems though working up to high DEE blending ratios (30%) and loads, thus not encountering the findings of the two works [40,41], reporting unstable and heavy smoke engine operation with higher than 15% (up to 25%) of DEE in its blends with diesel fuel. The latter researchers (working on essentially the same engine) attributed this behavior to erratic combustion, possibly due to phase separation of the blends that resulted in cavitations (vapor locks because of the high volatility of DEE) in the fuel line and injector nozzle, thus leading eventually to poor fuel injection (large droplets) in the combustion chamber. It is noticed that their injection system was already operating in (or over) the limit for the neat diesel fuel with high smoking at the high load points, and thus deteriorating its performance when a different fuel (DEE blends) was tried.

7. Conclusions An extended experimental study is conducted to evaluate and compare the use of DEE, a promising bio-fuel, as supplement to the conventional diesel fuel in a high-speed, direct injection diesel engine, operating at four loads. A heat release analysis of the experimentally obtained pressure diagrams revealed that with the use of DEE blend against neat diesel fuel, at all loads, the fuel injection pressure diagrams are delayed (with the uprising leg inclined), dynamic injection timing decreased, ignition delay increased, maximum cylinder pressures

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D.C. Rakopoulos et al. / Fuel xxx (2013) xxx–xxx

and temperatures decreased, while the engine runs overall a little ‘leaner’ with reduced heat losses. The acquired data were statistically analyzed and shown in this paper for the maximum pressure and its maximum rate of pressure rise, the dynamic injection timing, and the ignition delay. The cycle-by-cycle variation was expressed as the mean and coefficient of variation of these parameters. The analysis of probability density and auto-correlation functions of the various parameters, revealed the randomness (stochastic nature) of fluctuation phenomena observed in the engine. Cross-correlation coefficients showed clearly that neither the injection process (through the dynamic injection timing) nor the DEE/diesel fuel blend used (through the cetane number) have any practical effect on the above cyclic variations (irregularity). Thus, there is no unstable operation of the engine at least for up to 24% addition of DEE in its blend with diesel fuel.

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