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Study on combustion and emission of a dimethyl ether-diesel dual-fuel premixed charge compression ignition combustion engine with LPG (liquefied petroleum gas) as ignition inhibitor
Energy 96 (2016) 278e285
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Study on combustion and emission of a dimethyl ether-diesel dualfuel premixed charge compression ignition combustion engine with LPG (liquefied petroleum gas) as ignition inhibitor Ying Wang a, *, Hong Liu b, Zhiyong Huang a, Zhensheng Liu a a b
School of Energy and Power Engineering, Xi'an Jiaotong University, Xi'an 710049, PR China School of Energy and Power Engineering, Dalian University of Technology, Dalian 116024, PR China
a r t i c l e i n f o
a b s t r a c t
Article history: Received 26 June 2015 Received in revised form 31 October 2015 Accepted 15 December 2015 Available online xxx
A scheme of the DME-diesel dual-fuel PCCI (premixed charge compression ignition) combustion is an attractive option of relatively high thermal efficiency, low emission and operating costs as well as availability of wide DME (dimethyl ether) resource. However, one of the main problems in the DMEdiesel dual-fuel PCCI operation is that an early start of combustion easily leads to detonation due to high cetane number and better auto-ignition property of DME, and above-mentioned phenomenon is especially pronounced with a large DME energy ratio or at higher loads. Thus, LPG (liquefied petroleum gas), as a kind of ignition inhibitor, is added into DME for changing the property of port premixed fuel, and the effects of LPG quantity on the combustion and exhaust performance of a DME-diesel dual-fuel PCCI engine is investigated in this paper. Experimental results show that the appearing positions of the maximum cylinder pressure and mass-averaged temperature shift towards later crank angles and their peak values decline slightly when LPG mass fraction in DME/LPG mixture (fL) increases. The start of combustion postpones, and the combustion duration shortens with a rise of fL. Simultaneously, brake thermal efficiency and NOx emission reduce slightly; whereas the number concentration of particles firstly increases but then decreases with a rise of fL. © 2015 Elsevier Ltd. All rights reserved.
Keywords: Dual-fuel DME-diesel engine LPG (liquefied petroleum gas) Combustion Emission
1. Introduction The application of renewable and clean fuels for vehicles becomes more desirable due to the promulgation of increasingly stringent emission regulations and great attentions to petroleum safety. Furthermore, massive attempts have been made on various combustion approaches to decline NOx and PM (particulate matter) from CI (compression ignition) engines. In the conventional CI engines combustion starts with an injection event, and it originates as a spatially heterogeneous distribution of temperature and species concentration, which is regarded as one of the main reason of NOx and PM formation. In this respect, a scheme of dual-fuel engines with diesel and gas fuels is an attractive option of low exhaust pollutions because the gas fuel is a kind of clean fuel with the availability of wide resource. The gas fuel is normally introduced through an intake-pipe, and a
* Corresponding author. Tel.: þ86 29 82668726; fax: þ86 29 82668789. E-mail address: [email protected] (Y. Wang). http://dx.doi.org/10.1016/j.energy.2015.12.056 0360-5442/© 2015 Elsevier Ltd. All rights reserved.
pre-mixture is spontaneously formed in it. The dual-fuel engines of this type have the merits such as limited modifications in the diesel engines, the possibility of use the gas fuel with a low heating value and a relatively high brake thermal efficiency at a high load. The dual-fuel combustion engine also produces relatively low NOx and PM emissions, although it typically results in relatively high HC (hydrocarbon) and CO (carbon monoxide) emissions. The studies on this type of dual-fuel operation have been widely reported in the past years [1e3]. Barik et al. [1] studied the combustion and emission characteristics of biogas-diesel duel-fuel combustion engine. They found that the maximum cylinder pressure for the dualfuel operation was higher by about 11 bar than that of diesel operation, and NO and smoke emissions from the dual-fuel engine were lower by about 39% and 49%, compared to that of diesel engine. Lounici et al. [2] carried out the performance research in a CNG-diesel dual-fuel engine. Results showed a simultaneous reduction of NOx and soot could be realized in the CNG-diesel dualfuel engine; moreover, it showed the possibility to obtain lower brake specific fuel consumption over the certain operating conditions. Chintala et al. [3] dealt with the effect of different
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Nomenclature DME LPG HCCI DICI BMEP HRR LTR HTR TDC NOx PM SOC
he rp fL CA10 CA50 CA90
dimethyl ether liquefied petroleum gas homogeneous charge compression ignition direct injection compression ignition brake mean effective pressure heat release rate low temperature reaction high temperature reaction top dead center nitrogen oxide particulate matter start of combustion brake thermal efficiency energy ratio of port premixing fuel LPG mass fraction in port DME/LPG mixture crank angle where an accumulated heat release rate reaches 10% crank angle where an accumulated heat release rate reaches 50% crank angle where an accumulated heat release rate reaches 90%
compression ratios on the hydrogen energy share, emissions and thermal efficiency in a H2-diesel dual-fuel engine. It was found that reducing the compression ratio was a promising method for improving the hydrogen energy share and thermal efficiency along with lower emissions. DME (dimethyl ether) is also regarded as one kind of fuel choice in the view of reducing the dependence upon petroleum and realizing relatively low emissions [4e6]. DME can be made from various sources such as coal, natural gas and biomass, and it is an ideal diesel fuel substitute for CI engines because of its high cetane number and smoke-free combustion characteristic. In addition, the characteristic of the DME fuel itself containing oxygen is advantageous to the decrease in PM emission. Chapman et al. [7] studied a DME-diesel combustion mode in a turbo-diesel engine. They found that fumigation DME ignited first during the combustion process, thus resulting in both an early SOC (start of combustion) and elevated pressure peak. Wang et al. [8] studied the DME-diesel dual-fuel combustion in a natural aspirated engine, and the results also indicated that the ignition timing of DME was prior to that of diesel. As the port DME quantity increased, the start of combustion advanced, and the peak values of cylinder pressure and mass-averaged temperature increased. Meanwhile, the maximum heat release rate of diffusion-controlled combustion decreased, and smoke emissions decreased obviously. Zhao et al. [9] studied the effects of EGR (exhaust gas recirculation) on combustion and emission characteristics of the DME-diesel dual-fuel PCCI (premixed charge compression ignition) engine. They found that the maximum values of in-cylinder pressure and calculated massaveraged cylinder temperature reduced with a rise of EGR rate. Simultaneously, the crank-angle positions corresponding to the maximum heat release rate lagged with a rise of EGR rate. It was also observed that the decreasing values in the peak of in-cylinder pressure, mass-averaged temperature and heat release rate, as well as delaying phase in crank-angle degree resulted from the adoption of EGR, gradually weakened with a rise of DME energy ratio. Furthermore, with an increase in EGR rate, NOx emission decreased sharply.
279
Based on the previous studies, it was observed that DME-diesel dual-fuel combustion is an attractive option, but one of the main problems in the DME-diesel dual-fuel operation is that an early SOC easily leads to denotation, especially with a relatively large DME quantity and at a relatively high load. Hence, in this paper a kind of high antiknock-property fuel-LPG, as an ignition inhibitor, is blended with DME for better controlling the combustion phase. LPG (liquefied petroleum gas) is mainly composed of propane and butane. The demand for LPG as a kind of promising vehicle alternative fuel is increasing each year around the world [10e12]. There are numerous papers about LPG as vehicle fuel now. Jemni et al. [13] investigated the effects of intake manifold design on in-cylinder flow field and engine performance in a bus diesel engine converted to LPG fueled engine by CFD (computational fluid dynamics) analysis and experiments. The results affirmed that the power output and brake thermal efficiency increased when adopting an optimal design of the intake manifold. Tira et al. [14] carried out the LPG-diesel dual-fuel combustion research to understand the effects of the properties of the direct injection fuels, such as RME (rapeseed methyl ester) and GTL (gas-to-liquid), on engine performance, combustion and emissions characteristics. The results indicated that up to 60% of liquid fuel replacement by LPG was achieved while remaining engine combustion variability within the acceptable range as well as the PM-NOx trade-off. However, the amount of LPG was limited by adverse impacts in brake thermal efficiency, HC and CO emissions. Surawski et al. [15] studied the performance and emissions of a LPG fumigated diesel compression ignition engine. The results showed that LPG fumigation coerced the combustion into premixed mode, and increased the peak value of in-cylinder pressure. NOx and PM emissions decreased, but CO and HC emissions increased. Ashok et al. [16] reviewed the LPG/diesel dual-fuel engines and concluded that such kind of engines had relatively high thermal efficiency at high loads but poor performance at part loads due to the bad utilization of charges, but poor performance at parts could be overcome by adjusting diesel fuel quantity, injection timing, gaseous fuel composition and intake charge conditions. In addition, the handling characteristic of LPG is similar to that of DME, so a mixture of DME and LPG can be easily prepared in advance. Some researchers also reported the experiment results about DME/LPG combustion in the engines. Lee et al. [17]experimentally studied the combustion characteristics of a spark ignition engine operated with DME-LPG blend fuel. They found that the knock occurrence area increased virtually when DME was blended into LPG. However, the stable combustion can be observed in all test conditions with DME (20% mass fraction)-LPG (80% mass fraction) blend fuel. Jang et al. [18] evaluated the controllability of combustion phase and performance improvement of the LPG-DME HCCI (homogeneous charge compression ignition) engine. The results indicated that in the case of LPG direct injection with DME port injection, the engine performance was limited by early combustion of DME, whereas DME direct injection with LPG port injection was the better way to increase the IMEP (indicated mean effective pressure) and reduce emissions. Now the studies on the fuel effects on engine performance are somewhat engine dependent, and a great number of studies report the fact that it is very hard to attribute the changes in the performance and emission of CI engines to the change of single fuel property due to correlations among the fuel properties [19,20]. However, it is expected that dual-fuel PCCI combustion can exhibit a greater correlation with the fuel property than that of the conventional diesel CI combustion because the compression ignition characteristic of a dilute fuel/ air mixture strongly rests with the fuel vaporization, the fuel/air mixing process and the fuel chemical reactivity [21e23]. Although there have been many studies concerning DME-diesel dual-fuel combustion [7e9], LPG-diesel dual-fuel operation
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[14e16], and DME/LPG combustion [17,18], there is little information about combustion and emission characteristics of a DMEdiesel dual fuel PCCI engine with LPG as ignition inhibitor. Therefore, the purposes of this paper are to examine the effects of LPG addition on the combustion and exhaust emission characteristics of a DME-diesel dual-fuel engine and to determine whether the LPG addition is feasible in the combustion phase control in the DMEdiesel dual-fuel engine. 2. Test method and fuels Fig. 1 exhibits a schematic diagram of the experimental bench. The experiments were conducted in a two-cylinder, four-stroke, naturally-aspirated DICI (direct injection compression ignition) offroad engine. The test engine specifications are shown in Table 1. Four kinds of the different mass proportions of the DME/LPG mixture are prepared in advance. In the experiment, fL is used to denote as a LPG mass fraction in DME/LPG mixture, and it was 0, 20%, 33% and 50% respectively. Then the DME/LPG mixture enters into the intake port to form an approximately homogeneous fuel/ air charge prior to combustion, as shown as in Fig. 1. In order to avoid the influence of intake temperature variation, an intake temperature control system is used to maintain the intake temperature constant. The mass flow rate of the DME/LPG mixture is controlled through a valve located up-stream of the heater. The mass change is monitored by an electric balance. The traditional diesel direct-injection fuel system is still retained for the diesel DICI combustion process, and the direct injection timing is 7 CA BTDC. The diesel, DME and LPG fuels used in this study are the commercial products and their main properties are shown in Table 2. In the tests, the pressure in the combustion chamber is obtained with a Kistler 6125A pressure transducer, and then the output of this transducer is converted to an amplifying voltage through a Kistler 5011. Crank angle signals are obtained from a Kistler 2629B mounted on a crankshaft. The acquisition process of the signals covers 100 completed cycles, and the average value of these 100 consecutive cycles is used as the pressure data. The rates of heat release are calculated according to the First Law of Thermodynamics [24]. NOx emissions are monitored with HORIBA MEXA720NOx. PM is collected and analyzed by Dekati ELPI (electrical low-pressure impactor), which can measure real-time size distributions in a size range of 7 nm10 mm. The dilution ratio remained constant for comparisons in all the experimental conditions. ELPI and aerosol sampling device were thoroughly cleaned, and all the apparatus were calibrated before the experiments. The real measurement was not started until the engine warmed up for the stable particle size distribution. During any switch between different
Data acquisiƟon In-cylinder pressure transducer Exhaust
Diesel injector
Temperature control system
Intake valve
DME/LPG mixture
Table 1 The test engine specifications. Bore/mm Stroke/mm Cylinder number Displacement/liter Compression ratio Aspiration mode injection timing injector opening pressure/MPa Nozzle (hole number hole diameter)
105 120 2 2.078 17 Naturally aspirated 7 CA BTDC 19 4 0.32 (mm)
Table 2 Physical and chemical property of DME, LPG and diesel. Property
Diesel
DME
LPG Propane
Butane
Liquid density (kg/m3) Low heating value (MJ/kg) Cetane number Octane number Stoichiometric air fuel ratio % Wt of oxygen
840 42.5 40e55 20e30 14.5 0
668 28.43 55e60 low 9 34.8
500 46.3 5 105 15.9 0
578 45.7 10 92 15.5 0
Note: LPG consists of 60% propane and 40% butane in this paper.
operational modes, the system would stabilize for 10e15 min for aerosol stability. The cooling water temperature remains at approximately 80 C and the oil temperature is kept at 90 C during all the experiments. The accuracies of the engine speed and torque are 1r/min and 0.01 N$m respectively. The experiment is carried out at 1700r/min and two BMEPs (brake mean effective pressure): 0.24 MPa and 0.48 MPa. Overall excess air ratio is about 4.2 at BMEP ¼ 0.24 MPa and 7.8 at BMEP ¼ 0.48 MPa. Because there is obvious difference in the lower heating value of DME, LPG and diesel, the port fuel energy ratio (rp) is introduced in order to make the comparison clearly, and its calculating equation is as follows.
rp ¼
Bmix Humix 100% Bmix Humix þ Bdiesel Hudiesel
(1)
where Bmix and Bdiesel specify the mass consumption rate of DME/ LPG mixture and diesel fuel respectively. Hudiesel specifies a low heating value of the diesel fuel. Hu mix (MJ/kg) specifies a total low heating value of the DME/LPG mixture, which is approximately calculated by the following equation.
Humix ¼ HuLPG fL þ HuDME ð1 fL Þ
(2)
where fL is a LPG mass fraction in DME/LPG mixture. Because early combustion easily occurs under relatively high DME premixed ratio, rp ¼ 30% and rp ¼ 40% are selected in this experiment. LPG mass fractions (fL) are 0, 25%, 33% and 50% respectively. 3. Results and discussions 3.1. Influence of LPG addition on heat release rate, maximum cylinder pressure and mass-averaged cylinder temperature
Gas analyzer ELPI Crank-angle encoder
Electric balance
Fig. 1. The schematic diagram of the experimental bench.
The curves of heat release rate of a DME-diesel dual-fuel PCCI engine with various LPG mass fractions under 0.24 MPa and 0.48 MPa are presented in Fig. 2. It can be observed that the curve shapes of the heat release rate of the DME-diesel dual-fuel combustion are not typical of diesel DICI combustion and are somewhat complex.
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Fig. 2. Effects of LPG addition on the heat release rate of the engine.
It is well-known that DME homogeneous compression ignition combustion exhibits typical LTR (low temperature reaction) and HTR (high temperature reaction) behavior [25,26]. The LTR taking place in the range 700e750 K promotes the auto-ignition. The main reaction scheme for DME LTR is the H-atom abstraction, CH3OCH2O2 formation and intermolecular isomerization. The combustible mixture can be ignited when the heat release in the LTR raises the mixture temperature to the region of HTR (high temperature reaction) above 1100 K. During the stage of DME HTR, the large energy releases. At the same time, H2O, CO and CO2 form. It is also known that LPG homogeneous compression ignition combustion exhibits single-stage heat release behavior without cool flame [27]. Thus, it can be expected that DME/LPG premixed homogeneous compression ignition combustion also shows two-stage behavior. The first-stage heat release is still DME LTR combustion. The second-stage heat release is the HTR combustion of the mixture of DME and LPG. Fig. 2(a) gives the heat release process of the dual-fuel PCCI combustion engine with different fL at the condition of BMEP ¼ 0.24 MPa and rp ¼ 30%. It can be observed that the heat release process of the dual-fuel combustion is made up of DME LTR and conventional diffusion-controlled CI combustion. During this process, DME LTR combustion firstly proceeds, then the conventional CI combustion composed of premixed and the diffusion combustion follows. In Fig. 2(a) the appearing timing of the LTR almost keeps invariable (about 24e25 CA BTDC) regardless of the ratio of premixed charge. This is related to the DME LTR combustion mechanism. The initial combustion temperature of DME LTR combustion is normally about 700e750 K [28], and the ignition timing of DME LTR combustion is mainly dependent on the cylinder temperature; whereas at that times the rise of temperature is mainly influenced by charge compression, thus the ignition timing of DME LTR combustion little affected by the quantity of premixed charge. The invariable appearing timing of DME LTR was also reported by Refs. [29] and [30]. In Fig. 2(a) it is also observed that the maximum value of heat release rate during the DME LTR process decreases slightly with an increase in fL because the DME quantity available for auto-ignition reduces. It can be seen clearly
that the start of CI combustion time postpones with a rise of fL because premixed LPG prevents DME from auto-ignition, and part of the energy or radical released from DME cool-flame is moved to ignite LPG when the quantity of LPG in the mixture increases. Furthermore, the maximum value of heat release rate increases a little with a rise of fL due to higher low heating value of LPG than that of DME. Fig. 2(b) gives the heat release process of the dual-fuel PCCI combustion engine at the condition of BMEP ¼ 0.24 MPa and rp ¼ 40%. The two figures (Figs. 2(a) and (b)) show the similar curve shape except the maximum values of heat release rate and the combustion phase. During the conventional CI combustion stage, the start of combustion advances, but the maximum value of heat release rate decreases as the port fuel energy ratio increases from 30% to 40%. Fig. 2(c) gives the heat release process of the dual-fuel PCCI combustion engine with different fL at the condition of BMEP ¼ 0.48 MPa and rp ¼ 30%. It can be observed that due to an increase in the amount of port fumigation DME, HTR becomes obvious compared with that at the condition of BMEP ¼ 0.24 MPa and rp ¼ 30%, and the heat release process of the dual-fuel combustion is made up of LTR, HTR and conventional diffusioncontrolled CI combustion. Same to the Figs. 2(a) and (b), the appearing timing of the LTR stage is nearly invariable, and the maximum value of heat release rate during the DME LTR process decreases slightly with a rise of fL. Meanwhile, the starting time of HTR postpones and the maximum value of heat release rate decreases a little with a rise of fL due to the suppression of LPG addition on DME auto-ignition. The start of diffusion-controlled combustion also postpones with a rise of fL, but the maximum value of heat release rate during this stage increases a little due to higher heating value of LPG. Fig. 2(d) gives the heat release process of dual-fuel PCCI combustion engine at the condition of BMEP ¼ 0.48 MPa and rp ¼ 40%. Similar to Fig. 2(c), the same variation trend during the LTR, HTR and diffusion-controlled CI combustion stages are showed in Fig. 2(d). However, due to an increase in port premixing quantity, the values of heat release rate increase during HTR stage
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and combustion phases advance a little when rp increases from 30% to 40%. Fig. 3 and 4 exhibit the effects of LPG on the maximum values of in-cylinder pressure and mass-averaged cylinder temperature. Based on the analysis of heat release process, the ignition timing of DME is prior to that of diesel and the initial burning occurs in the form of DME HCCI for DME-diesel dual-fuel PCCI operation. Consequently, an increased peak pressure and mass-averaged maximum cylinder-temperature as well as an advanced position of pressure and temperature peak are shown in DME-diesel dualfuel combustion. As mentioned as above, when higher octane number of LPG is blended into DME, the amount of DME becomes less and thus the start of combustion time postpones. Therefore, the onset of an obvious rise in the cylinder-pressure and mass-averaged cylinder-temperature shifts to a later crank angle, and the maximum values of the cylinder-pressure and the mass-averaged cylinder-temperature both decline slightly with a rise of fL. 3.2. Influence of LPG addition on combustion phase An interval crank angle between CA10 and CA90 is defined as combustion duration. The crank angle where an accumulated heat release rate is 10% of a total heat release rate is defined as the CA10, and the crank angle where an accumulated heat release rate is 90% of a total heat release rate is defined as the CA90. CA10 and CA90 are also separately used to indicate the start and the end of combustion per cycle. CA50, the most frequently used indicator of the combustion phasing, is denoted as the crank angle where an accumulated heat release rate reaches 50% of a total heat release rate. Fig. 5 shows the effects of LPG addition on the CA10, CA50 and CA90 of the DME-diesel dual fuel engine under the various operation conditions. It can be seen that the CA10 is postponed with a rise of fL, and such a delay effect is the most obvious in the cases of fL ¼ 50%. At the BMEP ¼ 0.24 MPa, CA10 is delayed about 6.7 CA with rp ¼ 30% and delayed about 7.3 CA with rp ¼ 40%; whereas at the BMEP ¼ 0.48 MPa, CA10 is delayed about 7 CA with rp ¼ 30% and delayed about 8.2 CA with rp ¼ 40%. Meanwhile, the LPG
addition almost has no great influence on the CA90 for all of operating conditions. For both 0.24 MPa and 0.48 MPa, there is only slight increment in CA90 with the different fL in each figure. Consequently, the combustion duration correspondingly shortens with a rise of fL, and this phenomenon becomes more remarkable as fL increases. In the case of fL ¼ 50%, at the BMEP ¼ 0.24 MPa, the combustion duration shortens about 4.9 CA with rp ¼ 30% and shortens about 6.48 CA with rp ¼ 40%; at the BMEP ¼ 0.48 MPa, the combustion duration shortens about 6.2 CA with rp ¼ 30% and shortens about 6.5 CA with rp ¼ 40%. It can also be found in Fig. 5 that the CA50 also slightly delays with a rise of fL, but the delay of CA50 is less than that of CA10. 3.3. Influence of LPG addition on the BTE (brake thermal efficiency) BTE (brake thermal efficiency) is the brake work per cycle divided by the fuel chemical energy indicated as the low heating value. The calculation equation of BTE in this paper is as follows:
het ¼
Pe Pe ¼ B$Hu Bdiesel Hudiesel þ Bmix Humix
(3)
where Pe (kW) indicates the power output. It can be known from this equation BTE is inversely proportional to the fuel consumption, and the low fuel consumption denotes better fuel economy. Combustion is mainly controlled by a diffusion process in a conventional diesel DICI engine, whereas the combustion of the diesel-DME dual-fuel engine is firstly controlled by chemical kinetics and then mainly controlled by the diffusion combustion. As indicated in many HCCI researches [31,32], one of the main merits of the kinetically-controlled HCCI combustion is a fairly high BTE, which leads to the low fuel consumption compared to the classical diffusion-controlled combustion. Therefore, it can be expected that with an increase in LPG mass fraction in DME/LPG mixture, BTE reduces slightly due to the decrease in the amount of the fuel burned in kinetically-controlled DME HCCI combustion, as shown as Fig. 6. When fL reaches 50%, the decrease in BTE is the most
Fig. 3. Effects of LPG addition on the cylinder maximum pressure of the engine.
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283
Fig. 4. Effects of LPG addition on the maximum mass-averaged cylinder temperature of the engine.
Fig. 5. Effects of LPG addition on combustion duration and CA50 of the engine.
obvious. Under this condition, because larger amount of LPG exists in DME/LPG mixture, ignition delay prolongs as well as combustion phase postpones. Thus more combustible mixture is formed during the period of ignition delay, finally leading to the BTE reduction. 3.4. Influence of LPG addition on NOx and PM emission It is widely accepted that the diesel engine is significant source of PM and NOx emissions, mainly owing to its heterogeneous combustion process. The particles destined to be emitted as PM are
born as soot nuclei in the highly oxygen-deficient core of fuel sprays and form in the locally rich regions of an inhomogeneous combustion. NOx is formed in the high temperature burned gases during the combustion process. The burned gas temperature, the amount of oxygen in the burned gases and the resident time at the high temperatures all are the primary variables affecting NOx formation. Fig. 7 gives the effects of LPG addition on the NOx emission of the DME-diesel dual fuel engine. It can be observed that NOx emission shows a very slight decrease with a rise of fL. It is due to that the
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Fig. 6. Effects of LPG addition on brake thermal efficiency of the engine.
Fig. 7. Effects of LPG addition on NOx emission of the engine.
onset of an obvious rise in the cylinder temperature shifts to the relatively late crank angle, and the peak cylinder temperature declines slightly with a rise of fL. Relatively low temperatures may somewhat prohibit the NOx formation. It can also be seen from Fig. 7 that NOx emission with rp ¼ 40% is a little lower than that with rp ¼ 30%. This can be attributed that the ignition delay of injected diesel fuel shortens slightly via the port induction of DME/ LPG mixture. In addition, NOx emission at low load (0.24 MPa) is lower than that at high load (0.48 MPa) because in-cylinder temperature becomes higher at high loads. Fig. 8 shows the effects of LPG addition on the particle number size distribution of the DME-diesel dual-fuel engine. In Fig. 8, the maximum number concentration of the particles locates almost at a diameter of 0.055 mm for different fL; whereas the variation of the number concentration of particles in the range of 0.3 mme0.7 mm is not significant. As shown in Fig. 8, as fL ¼ 20% and fL ¼ 33%, the LPG addition results in higher number concentration of particles over the 0.02e0.1 mm aerodynamic diameter measured compared with particle number concentration of DME-diesel dual-fuel engine without LPG addition. The main reason of higher particle number concentration is that the oxygen content in the charge decreases since LPG partly substitutes air in the intake pipe. LPG, unlike DME, does not contain any oxygen itself. It was reported by some researchers [33,34] that oxygen could decrease the particulate number concentrations effectively. However, when LPG mass fraction increases further, particle number concentration instead decreases, and particle number concentration for fL ¼ 50% operation is comparable to that of fL ¼ 0 operation (DME-diesel dual-fuel combustion without LPG). Based on Fig. 4, the cylinder massaveraged temperature decreases sharply when fL reduces to 50% so that the high temperature condition required for the formation of particulate matter lacks to some extent, thus leading to the reduction in PM emission. In the case of the high load condition, the increased amount of fuel is responsible for the increased particulate emissions. The increased amount of fuel used per cycle results in an increase of fuel droplets in the combustion chamber and the production of more
Fig. 8. Effects of LPG addition on the particle number size distribution of the engine.
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particles. Moreover, it can also be found in Fig. 8 that the increased amount of port fuel/air mixture is also responsible for the decreased particulate emissions. The increased amount of port mixture leads to an increased proportion of the premixed homogenous combustion and a decreased quantity of diesel droplets in the diffusion-controlled combustion, both of which suppresses the particulate formation. 4. Conclusion The effects of LPG on the combustion and emission characteristics in a DMEediesel dual-fuel engine are examined. The main conclusions are as follows: 1) The onset of an obvious rise in the cylinder-pressure and massaveraged cylinder-temperature shifts to a later crank angle and the maximum values of cylinder pressure and mass-averaged temperature decline slightly with a rise of fL. The diffusioncontrolled combustion begins at a later crank angle, and the maximum value of heat release rate during this stage rises a little as fL increases. 2) CA10 and CA50 both delay, but the combustion duration shortens when LPG is added into DME. Furthermore, such the delay effect of LPG on the combustion phase becomes more obvious as fL increases. 3) With a rise of fL, brake thermal efficiency reduces slightly due to a decrease in the amount of kinetically-controlled DME HCCI combustion. 4) NOx emission from diesel-DME dual-fuel engine reduces a little with a rise of fL. However, the addition of the LPG results in the first increased number concentration of particles but the decreased number concentration of particles over the aerodynamic diameters measured of 0.02e0.2 mm. Acknowledgment The authors wish to express thanks to Project of National Natural Science Foundation of China (91541118 and 51376038), Open Fund of Key Laboratory of Low-grade Energy Utilization Technologies and Systems, Chongqing University (LLEUTS-201506) and Scientific Research Foundations for the Returned Overseas Chinese Scholars, State Education Ministry. References [1] Barik D, Murugan S. Investigation on combustion performance and emission characteristics of a DI (direct injection) diesel engine fueled with biogase diesel in dual fuel mode. Energy 2014;72:760e71. [2] Lounici MS, Loubar K, Tarabet L, Balistrou M, Niculescu DC, Tazerout M. Towards improvement of natural gas-diesel dual fuel mode: an experimental investigation on performance and exhaust emissions. Energy 2014;64: 200e11. [3] Chintala V, Subramanian KA. Experimental investigations on effect of different compression ratios on enhancement of maximum hydrogen energy share in a compression ignition engine under dual-fuel mode. Energy 2015;87:448e62. [4] Wang Y, Xiao F, Zhao YW, Li DC, Lei X. Study on cycle-by-cycle variations in a diesel engine with dimethyl ether as port premixing fuel. Appl Energy 2015;143:58e70. [5] Wang Y, Zhou LB. Experimental study on exhaust emissions from a multicylinder DME engine operating with EGR and oxidation catalyst. Appl Therm Eng 2008;28:1589e95. [6] Park SH, Lee CS. Combustion performance and emission reduction characteristics of automotive DME engine system. Prog Energy Combust Sci 2013;39: 147e68.
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[7] Chapman EM, Boehman AL. Pilot ignited premixed combustion of dimethyl ether in a turbodiesel engine. Fuel Process Tech 2008;89(12):1262e71. [8] Wang Y, Zhao YW, Yang ZL. Dimethyl ether energy ratio effects in a dimethyl ether -diesel dual fuel premixed charge compression ignition engine. Appl Therm Eng 2013;54:481e7. [9] Zhao YW, Wang Y, Li DC, Lei X, Liu SH. Combustion and emission characteristics of a DME (dimethyl ether)- diesel dual fuel premixed charge compression ignition engine with EGR (exhaust gas recirculation). Energy 2014;72: 608e17. [10] Ayhan V, Parlak A, Cesur I, Boru B, Kolip A. Performance and exhaust emission characteristics of a diesel engine running with LPG. Int J Phys Sci 2011;6(8): 1905e14. [11] Elnajjar E, Hamdan MO, Selim MYE. Experimental investigation of dual engine performance using variable LPG composition fuel. Renew Energy 2013;56: 110e6. [12] Erkus B, Surmen A, Karamangil MI. A comparative study of carburation and injection fuel supply methods in an LPG-fuelled SI engine. Fuel 2013;107: 511e7. [13] Jemni MA, Kantchev G, Abid MS. Influence of intake manifold design on incylinder flow and engine performances in a bus diesel engine converted to LPG gas fuelled, using CFD analyses and experimental investigations. Energy 2011;36:2701e15. [14] Tira HS, Herreros JM, Tsolakis A, Wyszynsk ML. Characteristics of LPG-diesel dual fuelled engine operated with rapeseed methyl ester and gas-to-liquid diesel fuels. Energy 2012;47:620e9. [15] Surawski NC, Miljevic B, Bodisco TA, Situ R, Brown RJ, Ristovski ZD. Performance and gaseous and particle emissions from a liquefied petroleum gas (LPG) fumigated compression ignition engine. Fuel 2014;133:17e25. [16] Ashok B, Ashok SD, Kumar CR. LPG diesel dual fuel engine e a critical review. Alex Eng J 2015;54:105e26. [17] Lee S, Oh S, Choi Y, Kang KY. Effect of n-Butane and propane on performance and emission characteristics of an SI engine operated with DME-blended LPG fuel. Fuel 2011;90:1674e80. [18] Jang J, Yang K, Bae C. The effect of injection location of DME and LPG in a dual fuel HCCI engine. SAE paper no.2009-01-1847. 2009. [19] Zannis TC, Hountalas DC, Papagiannakis RG, Lavendis YA. Effect of fuel chemical structure and properties on diesel engine performance and pollutant emissions: review of the results of four European research programs. SAE paper no.2008-01-0838. 2008. [20] Dumitrescu CE, Neill WS, Guo HS, Hosseini V, Chippior WL. Fuel property effects on PCCI combustion in a heavy-duty diesel engine. J Eng Gas Turb Power 2012;134(5):052801. [21] Park SH, Youn IM, Lim Y, Lee CS. Fuel reactivity effects on the efficiency and operational window of dual-fuel compression ignition engines. Fuel 2014;118:163e75. [22] Hosseini V, Neill WS, Guo H, Dumitrescu CE, Chippior WL, Fairbridge C, et al. Effects of cetane number, aromatic content and 90% distillation temperature on HCCI combustion of diesel fuels. SAE paper 2010-01-2168. 2010. [23] Lv XC, Han D, Huang Z. Fuel design and management for the control of advanced compression- ignition combustion modes. Prog Energ Combust 2011;37:741e83. [24] Heywood JB. Internal combustion engine fundamentals. McGraw-Hill Book Company; 1998. [25] Curren HJ, Fischer SL, Dryer FL. The reaction kinetics of dimethyl ether.I: high-temperature oxidation in flow reactors. Int J Chem Kinet 2000;32: 713e40. [26] Curren HJ, Fischer SL, Dryer FL. The reaction kinetics of dimethyl ether.Ⅱ:Hightemperature oxidation in flow reactors. Int J Chem Kinet 2000;32:741e59. [27] Dagaut K, Hadj A. Kinetics of oxidation of a LPG blend mixture in a JSR: experimental and modeling study. Fuel 2003;82:475e80. [28] Sato S, Iida N. Analysis of DME homogeneous charge compression ignition combustion. SAE paper no. 2003-01-1825. 2003. [29] Fang Q, Fang JH, Zhuang J, Huang Z. Influences of pilot injection and exhaust gas recirculation (EGR) on combustion and emissions in a HCCI-DI combustion engine. Appl Therm Eng 2012;48:97e104. [30] Kim DS, Lee CS. Improved emission characteristics of HCCI engine by various premixed fuels and cooled EGR. Fuel 2006;85:695e704. [31] Christensen M, Johansson B. Supercharged homogeneous charge compression ignition (HCCI) with exhaust gas recirculation and pilot fuel. SAE paper no. 2000-01-1835. 2000. [32] Kobayashi K, Sako T, Sakaguchi Y, Morimoto S, Kanematsu S, Suzuki K, et al. Development of HCCI natural gas engines. J Nat Gas Sci Eng 2011;3: 651e6. [33] Wang XG, Cheung CS, Di YG, Huang ZH. Diesel engine gaseous and particle emissions fueled with diesele oxygenate blends. Fuel 2012;94:317e23. [34] Balamurugan T, Nalini R. Experimental investigation on performance, combustion and emission characteristics of four stroke diesel engine using diesel blended with alcohol as fuel. Energy 2014;78:356e63.
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Energy journal homepage: www.elsevier.com/locate/energy
Study on combustion and emission of a dimethyl ether-diesel dualfuel premixed charge compression ignition combustion engine with LPG (liquefied petroleum gas) as ignition inhibitor Ying Wang a, *, Hong Liu b, Zhiyong Huang a, Zhensheng Liu a a b
School of Energy and Power Engineering, Xi'an Jiaotong University, Xi'an 710049, PR China School of Energy and Power Engineering, Dalian University of Technology, Dalian 116024, PR China
a r t i c l e i n f o
a b s t r a c t
Article history: Received 26 June 2015 Received in revised form 31 October 2015 Accepted 15 December 2015 Available online xxx
A scheme of the DME-diesel dual-fuel PCCI (premixed charge compression ignition) combustion is an attractive option of relatively high thermal efficiency, low emission and operating costs as well as availability of wide DME (dimethyl ether) resource. However, one of the main problems in the DMEdiesel dual-fuel PCCI operation is that an early start of combustion easily leads to detonation due to high cetane number and better auto-ignition property of DME, and above-mentioned phenomenon is especially pronounced with a large DME energy ratio or at higher loads. Thus, LPG (liquefied petroleum gas), as a kind of ignition inhibitor, is added into DME for changing the property of port premixed fuel, and the effects of LPG quantity on the combustion and exhaust performance of a DME-diesel dual-fuel PCCI engine is investigated in this paper. Experimental results show that the appearing positions of the maximum cylinder pressure and mass-averaged temperature shift towards later crank angles and their peak values decline slightly when LPG mass fraction in DME/LPG mixture (fL) increases. The start of combustion postpones, and the combustion duration shortens with a rise of fL. Simultaneously, brake thermal efficiency and NOx emission reduce slightly; whereas the number concentration of particles firstly increases but then decreases with a rise of fL. © 2015 Elsevier Ltd. All rights reserved.
Keywords: Dual-fuel DME-diesel engine LPG (liquefied petroleum gas) Combustion Emission
1. Introduction The application of renewable and clean fuels for vehicles becomes more desirable due to the promulgation of increasingly stringent emission regulations and great attentions to petroleum safety. Furthermore, massive attempts have been made on various combustion approaches to decline NOx and PM (particulate matter) from CI (compression ignition) engines. In the conventional CI engines combustion starts with an injection event, and it originates as a spatially heterogeneous distribution of temperature and species concentration, which is regarded as one of the main reason of NOx and PM formation. In this respect, a scheme of dual-fuel engines with diesel and gas fuels is an attractive option of low exhaust pollutions because the gas fuel is a kind of clean fuel with the availability of wide resource. The gas fuel is normally introduced through an intake-pipe, and a
* Corresponding author. Tel.: þ86 29 82668726; fax: þ86 29 82668789. E-mail address: [email protected] (Y. Wang). http://dx.doi.org/10.1016/j.energy.2015.12.056 0360-5442/© 2015 Elsevier Ltd. All rights reserved.
pre-mixture is spontaneously formed in it. The dual-fuel engines of this type have the merits such as limited modifications in the diesel engines, the possibility of use the gas fuel with a low heating value and a relatively high brake thermal efficiency at a high load. The dual-fuel combustion engine also produces relatively low NOx and PM emissions, although it typically results in relatively high HC (hydrocarbon) and CO (carbon monoxide) emissions. The studies on this type of dual-fuel operation have been widely reported in the past years [1e3]. Barik et al. [1] studied the combustion and emission characteristics of biogas-diesel duel-fuel combustion engine. They found that the maximum cylinder pressure for the dualfuel operation was higher by about 11 bar than that of diesel operation, and NO and smoke emissions from the dual-fuel engine were lower by about 39% and 49%, compared to that of diesel engine. Lounici et al. [2] carried out the performance research in a CNG-diesel dual-fuel engine. Results showed a simultaneous reduction of NOx and soot could be realized in the CNG-diesel dualfuel engine; moreover, it showed the possibility to obtain lower brake specific fuel consumption over the certain operating conditions. Chintala et al. [3] dealt with the effect of different
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Nomenclature DME LPG HCCI DICI BMEP HRR LTR HTR TDC NOx PM SOC
he rp fL CA10 CA50 CA90
dimethyl ether liquefied petroleum gas homogeneous charge compression ignition direct injection compression ignition brake mean effective pressure heat release rate low temperature reaction high temperature reaction top dead center nitrogen oxide particulate matter start of combustion brake thermal efficiency energy ratio of port premixing fuel LPG mass fraction in port DME/LPG mixture crank angle where an accumulated heat release rate reaches 10% crank angle where an accumulated heat release rate reaches 50% crank angle where an accumulated heat release rate reaches 90%
compression ratios on the hydrogen energy share, emissions and thermal efficiency in a H2-diesel dual-fuel engine. It was found that reducing the compression ratio was a promising method for improving the hydrogen energy share and thermal efficiency along with lower emissions. DME (dimethyl ether) is also regarded as one kind of fuel choice in the view of reducing the dependence upon petroleum and realizing relatively low emissions [4e6]. DME can be made from various sources such as coal, natural gas and biomass, and it is an ideal diesel fuel substitute for CI engines because of its high cetane number and smoke-free combustion characteristic. In addition, the characteristic of the DME fuel itself containing oxygen is advantageous to the decrease in PM emission. Chapman et al. [7] studied a DME-diesel combustion mode in a turbo-diesel engine. They found that fumigation DME ignited first during the combustion process, thus resulting in both an early SOC (start of combustion) and elevated pressure peak. Wang et al. [8] studied the DME-diesel dual-fuel combustion in a natural aspirated engine, and the results also indicated that the ignition timing of DME was prior to that of diesel. As the port DME quantity increased, the start of combustion advanced, and the peak values of cylinder pressure and mass-averaged temperature increased. Meanwhile, the maximum heat release rate of diffusion-controlled combustion decreased, and smoke emissions decreased obviously. Zhao et al. [9] studied the effects of EGR (exhaust gas recirculation) on combustion and emission characteristics of the DME-diesel dual-fuel PCCI (premixed charge compression ignition) engine. They found that the maximum values of in-cylinder pressure and calculated massaveraged cylinder temperature reduced with a rise of EGR rate. Simultaneously, the crank-angle positions corresponding to the maximum heat release rate lagged with a rise of EGR rate. It was also observed that the decreasing values in the peak of in-cylinder pressure, mass-averaged temperature and heat release rate, as well as delaying phase in crank-angle degree resulted from the adoption of EGR, gradually weakened with a rise of DME energy ratio. Furthermore, with an increase in EGR rate, NOx emission decreased sharply.
279
Based on the previous studies, it was observed that DME-diesel dual-fuel combustion is an attractive option, but one of the main problems in the DME-diesel dual-fuel operation is that an early SOC easily leads to denotation, especially with a relatively large DME quantity and at a relatively high load. Hence, in this paper a kind of high antiknock-property fuel-LPG, as an ignition inhibitor, is blended with DME for better controlling the combustion phase. LPG (liquefied petroleum gas) is mainly composed of propane and butane. The demand for LPG as a kind of promising vehicle alternative fuel is increasing each year around the world [10e12]. There are numerous papers about LPG as vehicle fuel now. Jemni et al. [13] investigated the effects of intake manifold design on in-cylinder flow field and engine performance in a bus diesel engine converted to LPG fueled engine by CFD (computational fluid dynamics) analysis and experiments. The results affirmed that the power output and brake thermal efficiency increased when adopting an optimal design of the intake manifold. Tira et al. [14] carried out the LPG-diesel dual-fuel combustion research to understand the effects of the properties of the direct injection fuels, such as RME (rapeseed methyl ester) and GTL (gas-to-liquid), on engine performance, combustion and emissions characteristics. The results indicated that up to 60% of liquid fuel replacement by LPG was achieved while remaining engine combustion variability within the acceptable range as well as the PM-NOx trade-off. However, the amount of LPG was limited by adverse impacts in brake thermal efficiency, HC and CO emissions. Surawski et al. [15] studied the performance and emissions of a LPG fumigated diesel compression ignition engine. The results showed that LPG fumigation coerced the combustion into premixed mode, and increased the peak value of in-cylinder pressure. NOx and PM emissions decreased, but CO and HC emissions increased. Ashok et al. [16] reviewed the LPG/diesel dual-fuel engines and concluded that such kind of engines had relatively high thermal efficiency at high loads but poor performance at part loads due to the bad utilization of charges, but poor performance at parts could be overcome by adjusting diesel fuel quantity, injection timing, gaseous fuel composition and intake charge conditions. In addition, the handling characteristic of LPG is similar to that of DME, so a mixture of DME and LPG can be easily prepared in advance. Some researchers also reported the experiment results about DME/LPG combustion in the engines. Lee et al. [17]experimentally studied the combustion characteristics of a spark ignition engine operated with DME-LPG blend fuel. They found that the knock occurrence area increased virtually when DME was blended into LPG. However, the stable combustion can be observed in all test conditions with DME (20% mass fraction)-LPG (80% mass fraction) blend fuel. Jang et al. [18] evaluated the controllability of combustion phase and performance improvement of the LPG-DME HCCI (homogeneous charge compression ignition) engine. The results indicated that in the case of LPG direct injection with DME port injection, the engine performance was limited by early combustion of DME, whereas DME direct injection with LPG port injection was the better way to increase the IMEP (indicated mean effective pressure) and reduce emissions. Now the studies on the fuel effects on engine performance are somewhat engine dependent, and a great number of studies report the fact that it is very hard to attribute the changes in the performance and emission of CI engines to the change of single fuel property due to correlations among the fuel properties [19,20]. However, it is expected that dual-fuel PCCI combustion can exhibit a greater correlation with the fuel property than that of the conventional diesel CI combustion because the compression ignition characteristic of a dilute fuel/ air mixture strongly rests with the fuel vaporization, the fuel/air mixing process and the fuel chemical reactivity [21e23]. Although there have been many studies concerning DME-diesel dual-fuel combustion [7e9], LPG-diesel dual-fuel operation
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[14e16], and DME/LPG combustion [17,18], there is little information about combustion and emission characteristics of a DMEdiesel dual fuel PCCI engine with LPG as ignition inhibitor. Therefore, the purposes of this paper are to examine the effects of LPG addition on the combustion and exhaust emission characteristics of a DME-diesel dual-fuel engine and to determine whether the LPG addition is feasible in the combustion phase control in the DMEdiesel dual-fuel engine. 2. Test method and fuels Fig. 1 exhibits a schematic diagram of the experimental bench. The experiments were conducted in a two-cylinder, four-stroke, naturally-aspirated DICI (direct injection compression ignition) offroad engine. The test engine specifications are shown in Table 1. Four kinds of the different mass proportions of the DME/LPG mixture are prepared in advance. In the experiment, fL is used to denote as a LPG mass fraction in DME/LPG mixture, and it was 0, 20%, 33% and 50% respectively. Then the DME/LPG mixture enters into the intake port to form an approximately homogeneous fuel/ air charge prior to combustion, as shown as in Fig. 1. In order to avoid the influence of intake temperature variation, an intake temperature control system is used to maintain the intake temperature constant. The mass flow rate of the DME/LPG mixture is controlled through a valve located up-stream of the heater. The mass change is monitored by an electric balance. The traditional diesel direct-injection fuel system is still retained for the diesel DICI combustion process, and the direct injection timing is 7 CA BTDC. The diesel, DME and LPG fuels used in this study are the commercial products and their main properties are shown in Table 2. In the tests, the pressure in the combustion chamber is obtained with a Kistler 6125A pressure transducer, and then the output of this transducer is converted to an amplifying voltage through a Kistler 5011. Crank angle signals are obtained from a Kistler 2629B mounted on a crankshaft. The acquisition process of the signals covers 100 completed cycles, and the average value of these 100 consecutive cycles is used as the pressure data. The rates of heat release are calculated according to the First Law of Thermodynamics [24]. NOx emissions are monitored with HORIBA MEXA720NOx. PM is collected and analyzed by Dekati ELPI (electrical low-pressure impactor), which can measure real-time size distributions in a size range of 7 nm10 mm. The dilution ratio remained constant for comparisons in all the experimental conditions. ELPI and aerosol sampling device were thoroughly cleaned, and all the apparatus were calibrated before the experiments. The real measurement was not started until the engine warmed up for the stable particle size distribution. During any switch between different
Data acquisiƟon In-cylinder pressure transducer Exhaust
Diesel injector
Temperature control system
Intake valve
DME/LPG mixture
Table 1 The test engine specifications. Bore/mm Stroke/mm Cylinder number Displacement/liter Compression ratio Aspiration mode injection timing injector opening pressure/MPa Nozzle (hole number hole diameter)
105 120 2 2.078 17 Naturally aspirated 7 CA BTDC 19 4 0.32 (mm)
Table 2 Physical and chemical property of DME, LPG and diesel. Property
Diesel
DME
LPG Propane
Butane
Liquid density (kg/m3) Low heating value (MJ/kg) Cetane number Octane number Stoichiometric air fuel ratio % Wt of oxygen
840 42.5 40e55 20e30 14.5 0
668 28.43 55e60 low 9 34.8
500 46.3 5 105 15.9 0
578 45.7 10 92 15.5 0
Note: LPG consists of 60% propane and 40% butane in this paper.
operational modes, the system would stabilize for 10e15 min for aerosol stability. The cooling water temperature remains at approximately 80 C and the oil temperature is kept at 90 C during all the experiments. The accuracies of the engine speed and torque are 1r/min and 0.01 N$m respectively. The experiment is carried out at 1700r/min and two BMEPs (brake mean effective pressure): 0.24 MPa and 0.48 MPa. Overall excess air ratio is about 4.2 at BMEP ¼ 0.24 MPa and 7.8 at BMEP ¼ 0.48 MPa. Because there is obvious difference in the lower heating value of DME, LPG and diesel, the port fuel energy ratio (rp) is introduced in order to make the comparison clearly, and its calculating equation is as follows.
rp ¼
Bmix Humix 100% Bmix Humix þ Bdiesel Hudiesel
(1)
where Bmix and Bdiesel specify the mass consumption rate of DME/ LPG mixture and diesel fuel respectively. Hudiesel specifies a low heating value of the diesel fuel. Hu mix (MJ/kg) specifies a total low heating value of the DME/LPG mixture, which is approximately calculated by the following equation.
Humix ¼ HuLPG fL þ HuDME ð1 fL Þ
(2)
where fL is a LPG mass fraction in DME/LPG mixture. Because early combustion easily occurs under relatively high DME premixed ratio, rp ¼ 30% and rp ¼ 40% are selected in this experiment. LPG mass fractions (fL) are 0, 25%, 33% and 50% respectively. 3. Results and discussions 3.1. Influence of LPG addition on heat release rate, maximum cylinder pressure and mass-averaged cylinder temperature
Gas analyzer ELPI Crank-angle encoder
Electric balance
Fig. 1. The schematic diagram of the experimental bench.
The curves of heat release rate of a DME-diesel dual-fuel PCCI engine with various LPG mass fractions under 0.24 MPa and 0.48 MPa are presented in Fig. 2. It can be observed that the curve shapes of the heat release rate of the DME-diesel dual-fuel combustion are not typical of diesel DICI combustion and are somewhat complex.
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281
Fig. 2. Effects of LPG addition on the heat release rate of the engine.
It is well-known that DME homogeneous compression ignition combustion exhibits typical LTR (low temperature reaction) and HTR (high temperature reaction) behavior [25,26]. The LTR taking place in the range 700e750 K promotes the auto-ignition. The main reaction scheme for DME LTR is the H-atom abstraction, CH3OCH2O2 formation and intermolecular isomerization. The combustible mixture can be ignited when the heat release in the LTR raises the mixture temperature to the region of HTR (high temperature reaction) above 1100 K. During the stage of DME HTR, the large energy releases. At the same time, H2O, CO and CO2 form. It is also known that LPG homogeneous compression ignition combustion exhibits single-stage heat release behavior without cool flame [27]. Thus, it can be expected that DME/LPG premixed homogeneous compression ignition combustion also shows two-stage behavior. The first-stage heat release is still DME LTR combustion. The second-stage heat release is the HTR combustion of the mixture of DME and LPG. Fig. 2(a) gives the heat release process of the dual-fuel PCCI combustion engine with different fL at the condition of BMEP ¼ 0.24 MPa and rp ¼ 30%. It can be observed that the heat release process of the dual-fuel combustion is made up of DME LTR and conventional diffusion-controlled CI combustion. During this process, DME LTR combustion firstly proceeds, then the conventional CI combustion composed of premixed and the diffusion combustion follows. In Fig. 2(a) the appearing timing of the LTR almost keeps invariable (about 24e25 CA BTDC) regardless of the ratio of premixed charge. This is related to the DME LTR combustion mechanism. The initial combustion temperature of DME LTR combustion is normally about 700e750 K [28], and the ignition timing of DME LTR combustion is mainly dependent on the cylinder temperature; whereas at that times the rise of temperature is mainly influenced by charge compression, thus the ignition timing of DME LTR combustion little affected by the quantity of premixed charge. The invariable appearing timing of DME LTR was also reported by Refs. [29] and [30]. In Fig. 2(a) it is also observed that the maximum value of heat release rate during the DME LTR process decreases slightly with an increase in fL because the DME quantity available for auto-ignition reduces. It can be seen clearly
that the start of CI combustion time postpones with a rise of fL because premixed LPG prevents DME from auto-ignition, and part of the energy or radical released from DME cool-flame is moved to ignite LPG when the quantity of LPG in the mixture increases. Furthermore, the maximum value of heat release rate increases a little with a rise of fL due to higher low heating value of LPG than that of DME. Fig. 2(b) gives the heat release process of the dual-fuel PCCI combustion engine at the condition of BMEP ¼ 0.24 MPa and rp ¼ 40%. The two figures (Figs. 2(a) and (b)) show the similar curve shape except the maximum values of heat release rate and the combustion phase. During the conventional CI combustion stage, the start of combustion advances, but the maximum value of heat release rate decreases as the port fuel energy ratio increases from 30% to 40%. Fig. 2(c) gives the heat release process of the dual-fuel PCCI combustion engine with different fL at the condition of BMEP ¼ 0.48 MPa and rp ¼ 30%. It can be observed that due to an increase in the amount of port fumigation DME, HTR becomes obvious compared with that at the condition of BMEP ¼ 0.24 MPa and rp ¼ 30%, and the heat release process of the dual-fuel combustion is made up of LTR, HTR and conventional diffusioncontrolled CI combustion. Same to the Figs. 2(a) and (b), the appearing timing of the LTR stage is nearly invariable, and the maximum value of heat release rate during the DME LTR process decreases slightly with a rise of fL. Meanwhile, the starting time of HTR postpones and the maximum value of heat release rate decreases a little with a rise of fL due to the suppression of LPG addition on DME auto-ignition. The start of diffusion-controlled combustion also postpones with a rise of fL, but the maximum value of heat release rate during this stage increases a little due to higher heating value of LPG. Fig. 2(d) gives the heat release process of dual-fuel PCCI combustion engine at the condition of BMEP ¼ 0.48 MPa and rp ¼ 40%. Similar to Fig. 2(c), the same variation trend during the LTR, HTR and diffusion-controlled CI combustion stages are showed in Fig. 2(d). However, due to an increase in port premixing quantity, the values of heat release rate increase during HTR stage
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and combustion phases advance a little when rp increases from 30% to 40%. Fig. 3 and 4 exhibit the effects of LPG on the maximum values of in-cylinder pressure and mass-averaged cylinder temperature. Based on the analysis of heat release process, the ignition timing of DME is prior to that of diesel and the initial burning occurs in the form of DME HCCI for DME-diesel dual-fuel PCCI operation. Consequently, an increased peak pressure and mass-averaged maximum cylinder-temperature as well as an advanced position of pressure and temperature peak are shown in DME-diesel dualfuel combustion. As mentioned as above, when higher octane number of LPG is blended into DME, the amount of DME becomes less and thus the start of combustion time postpones. Therefore, the onset of an obvious rise in the cylinder-pressure and mass-averaged cylinder-temperature shifts to a later crank angle, and the maximum values of the cylinder-pressure and the mass-averaged cylinder-temperature both decline slightly with a rise of fL. 3.2. Influence of LPG addition on combustion phase An interval crank angle between CA10 and CA90 is defined as combustion duration. The crank angle where an accumulated heat release rate is 10% of a total heat release rate is defined as the CA10, and the crank angle where an accumulated heat release rate is 90% of a total heat release rate is defined as the CA90. CA10 and CA90 are also separately used to indicate the start and the end of combustion per cycle. CA50, the most frequently used indicator of the combustion phasing, is denoted as the crank angle where an accumulated heat release rate reaches 50% of a total heat release rate. Fig. 5 shows the effects of LPG addition on the CA10, CA50 and CA90 of the DME-diesel dual fuel engine under the various operation conditions. It can be seen that the CA10 is postponed with a rise of fL, and such a delay effect is the most obvious in the cases of fL ¼ 50%. At the BMEP ¼ 0.24 MPa, CA10 is delayed about 6.7 CA with rp ¼ 30% and delayed about 7.3 CA with rp ¼ 40%; whereas at the BMEP ¼ 0.48 MPa, CA10 is delayed about 7 CA with rp ¼ 30% and delayed about 8.2 CA with rp ¼ 40%. Meanwhile, the LPG
addition almost has no great influence on the CA90 for all of operating conditions. For both 0.24 MPa and 0.48 MPa, there is only slight increment in CA90 with the different fL in each figure. Consequently, the combustion duration correspondingly shortens with a rise of fL, and this phenomenon becomes more remarkable as fL increases. In the case of fL ¼ 50%, at the BMEP ¼ 0.24 MPa, the combustion duration shortens about 4.9 CA with rp ¼ 30% and shortens about 6.48 CA with rp ¼ 40%; at the BMEP ¼ 0.48 MPa, the combustion duration shortens about 6.2 CA with rp ¼ 30% and shortens about 6.5 CA with rp ¼ 40%. It can also be found in Fig. 5 that the CA50 also slightly delays with a rise of fL, but the delay of CA50 is less than that of CA10. 3.3. Influence of LPG addition on the BTE (brake thermal efficiency) BTE (brake thermal efficiency) is the brake work per cycle divided by the fuel chemical energy indicated as the low heating value. The calculation equation of BTE in this paper is as follows:
het ¼
Pe Pe ¼ B$Hu Bdiesel Hudiesel þ Bmix Humix
(3)
where Pe (kW) indicates the power output. It can be known from this equation BTE is inversely proportional to the fuel consumption, and the low fuel consumption denotes better fuel economy. Combustion is mainly controlled by a diffusion process in a conventional diesel DICI engine, whereas the combustion of the diesel-DME dual-fuel engine is firstly controlled by chemical kinetics and then mainly controlled by the diffusion combustion. As indicated in many HCCI researches [31,32], one of the main merits of the kinetically-controlled HCCI combustion is a fairly high BTE, which leads to the low fuel consumption compared to the classical diffusion-controlled combustion. Therefore, it can be expected that with an increase in LPG mass fraction in DME/LPG mixture, BTE reduces slightly due to the decrease in the amount of the fuel burned in kinetically-controlled DME HCCI combustion, as shown as Fig. 6. When fL reaches 50%, the decrease in BTE is the most
Fig. 3. Effects of LPG addition on the cylinder maximum pressure of the engine.
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Fig. 4. Effects of LPG addition on the maximum mass-averaged cylinder temperature of the engine.
Fig. 5. Effects of LPG addition on combustion duration and CA50 of the engine.
obvious. Under this condition, because larger amount of LPG exists in DME/LPG mixture, ignition delay prolongs as well as combustion phase postpones. Thus more combustible mixture is formed during the period of ignition delay, finally leading to the BTE reduction. 3.4. Influence of LPG addition on NOx and PM emission It is widely accepted that the diesel engine is significant source of PM and NOx emissions, mainly owing to its heterogeneous combustion process. The particles destined to be emitted as PM are
born as soot nuclei in the highly oxygen-deficient core of fuel sprays and form in the locally rich regions of an inhomogeneous combustion. NOx is formed in the high temperature burned gases during the combustion process. The burned gas temperature, the amount of oxygen in the burned gases and the resident time at the high temperatures all are the primary variables affecting NOx formation. Fig. 7 gives the effects of LPG addition on the NOx emission of the DME-diesel dual fuel engine. It can be observed that NOx emission shows a very slight decrease with a rise of fL. It is due to that the
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Fig. 6. Effects of LPG addition on brake thermal efficiency of the engine.
Fig. 7. Effects of LPG addition on NOx emission of the engine.
onset of an obvious rise in the cylinder temperature shifts to the relatively late crank angle, and the peak cylinder temperature declines slightly with a rise of fL. Relatively low temperatures may somewhat prohibit the NOx formation. It can also be seen from Fig. 7 that NOx emission with rp ¼ 40% is a little lower than that with rp ¼ 30%. This can be attributed that the ignition delay of injected diesel fuel shortens slightly via the port induction of DME/ LPG mixture. In addition, NOx emission at low load (0.24 MPa) is lower than that at high load (0.48 MPa) because in-cylinder temperature becomes higher at high loads. Fig. 8 shows the effects of LPG addition on the particle number size distribution of the DME-diesel dual-fuel engine. In Fig. 8, the maximum number concentration of the particles locates almost at a diameter of 0.055 mm for different fL; whereas the variation of the number concentration of particles in the range of 0.3 mme0.7 mm is not significant. As shown in Fig. 8, as fL ¼ 20% and fL ¼ 33%, the LPG addition results in higher number concentration of particles over the 0.02e0.1 mm aerodynamic diameter measured compared with particle number concentration of DME-diesel dual-fuel engine without LPG addition. The main reason of higher particle number concentration is that the oxygen content in the charge decreases since LPG partly substitutes air in the intake pipe. LPG, unlike DME, does not contain any oxygen itself. It was reported by some researchers [33,34] that oxygen could decrease the particulate number concentrations effectively. However, when LPG mass fraction increases further, particle number concentration instead decreases, and particle number concentration for fL ¼ 50% operation is comparable to that of fL ¼ 0 operation (DME-diesel dual-fuel combustion without LPG). Based on Fig. 4, the cylinder massaveraged temperature decreases sharply when fL reduces to 50% so that the high temperature condition required for the formation of particulate matter lacks to some extent, thus leading to the reduction in PM emission. In the case of the high load condition, the increased amount of fuel is responsible for the increased particulate emissions. The increased amount of fuel used per cycle results in an increase of fuel droplets in the combustion chamber and the production of more
Fig. 8. Effects of LPG addition on the particle number size distribution of the engine.
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particles. Moreover, it can also be found in Fig. 8 that the increased amount of port fuel/air mixture is also responsible for the decreased particulate emissions. The increased amount of port mixture leads to an increased proportion of the premixed homogenous combustion and a decreased quantity of diesel droplets in the diffusion-controlled combustion, both of which suppresses the particulate formation. 4. Conclusion The effects of LPG on the combustion and emission characteristics in a DMEediesel dual-fuel engine are examined. The main conclusions are as follows: 1) The onset of an obvious rise in the cylinder-pressure and massaveraged cylinder-temperature shifts to a later crank angle and the maximum values of cylinder pressure and mass-averaged temperature decline slightly with a rise of fL. The diffusioncontrolled combustion begins at a later crank angle, and the maximum value of heat release rate during this stage rises a little as fL increases. 2) CA10 and CA50 both delay, but the combustion duration shortens when LPG is added into DME. Furthermore, such the delay effect of LPG on the combustion phase becomes more obvious as fL increases. 3) With a rise of fL, brake thermal efficiency reduces slightly due to a decrease in the amount of kinetically-controlled DME HCCI combustion. 4) NOx emission from diesel-DME dual-fuel engine reduces a little with a rise of fL. However, the addition of the LPG results in the first increased number concentration of particles but the decreased number concentration of particles over the aerodynamic diameters measured of 0.02e0.2 mm. Acknowledgment The authors wish to express thanks to Project of National Natural Science Foundation of China (91541118 and 51376038), Open Fund of Key Laboratory of Low-grade Energy Utilization Technologies and Systems, Chongqing University (LLEUTS-201506) and Scientific Research Foundations for the Returned Overseas Chinese Scholars, State Education Ministry. References [1] Barik D, Murugan S. Investigation on combustion performance and emission characteristics of a DI (direct injection) diesel engine fueled with biogase diesel in dual fuel mode. Energy 2014;72:760e71. [2] Lounici MS, Loubar K, Tarabet L, Balistrou M, Niculescu DC, Tazerout M. Towards improvement of natural gas-diesel dual fuel mode: an experimental investigation on performance and exhaust emissions. Energy 2014;64: 200e11. [3] Chintala V, Subramanian KA. Experimental investigations on effect of different compression ratios on enhancement of maximum hydrogen energy share in a compression ignition engine under dual-fuel mode. Energy 2015;87:448e62. [4] Wang Y, Xiao F, Zhao YW, Li DC, Lei X. Study on cycle-by-cycle variations in a diesel engine with dimethyl ether as port premixing fuel. Appl Energy 2015;143:58e70. [5] Wang Y, Zhou LB. Experimental study on exhaust emissions from a multicylinder DME engine operating with EGR and oxidation catalyst. Appl Therm Eng 2008;28:1589e95. [6] Park SH, Lee CS. Combustion performance and emission reduction characteristics of automotive DME engine system. Prog Energy Combust Sci 2013;39: 147e68.
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