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A two-stage combustion system for burning lean gasoline mixtures in a stationary spark ignited engine
Applied Energy 105 (2013) 271–281
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A two-stage combustion system for burning lean gasoline mixtures in a stationary spark ignited engine Stanislaw Szwaja ⇑, Arkadiusz Jamrozik 1, Wojciech Tutak 2 Institute of Thermal Machinery, Czestochowa University of Technology, 42-200 Czestochowa, ul. Dabrowskiego 69, Poland
h i g h l i g h t s " A two-stage combustion system was implemented to a spark ignited gasoline engine. " Extremely lean gasoline mixtures (lambda up to 2.0) were burnt effectively. " The NOx emission from the engine was found to be below the EURO-5 limit. " Effective burning ultra lean mixtures at partial load allow engine throttleless run. " 3D modeling was satisfactory verified with the experiments.
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Article history: Received 28 September 2012 Received in revised form 28 December 2012 Accepted 31 December 2012 Available online 4 February 2013 Keywords: Combustion Stationary engine Two-stage combustion system Pre-chamber Heat release rate NOx
a b s t r a c t The paper mainly focuses on applying the two-stage combustion system with a pre-chamber into the stationary internal combustion spark ignited engine. It especially concentrates on applying throttle less operation at partial load and reduction of the NOx emission. Considerations conducted in the paper are based on the in-cylinder combustion progress analysis. Additionally, analysis of tailpipe toxic emission, with particular focus on the NOx formation in the engine equipped with the pre-chamber, is also performed. The paper presents both results of 3-D combustion modeling in the SI engine and results conducted on a test SI engine. The 3-D modeling was performed in the KIVA-3V code. Next, results from modeling were compared with results obtained from tests. Finally, satisfactory good consistency between modeled and experimental courses of both pressure, temperature and NOx were obtained. Thus, the engine model with the proposed two-stage combustion system properly simulates engine working conditions on the test bed. Results from both analyses confirmed that the two-stage combustion system significantly shortens combustion duration of an ultra lean gasoline–air mixture and contributes to reduction in NOx. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction Combustion a lean gasoline–air mixture in a spark ignited (SI) engine is one method to reduce nitric oxide (NOx) emissions and increase engine efficiency by decreasing a peak combustion temperature [1–4]. However, while a lean mixture is combusted, combustion process gets longer and incomplete that causes engine work unstable and decrease in engine brake efficiency [4–6]. The paper presents a two-stage combustion system, in which combustion starts in a pre-chamber (1 stage) and further flames jet from a pre-chamber initiate lean mixture combustion in the engine
⇑ Corresponding author. Tel.: +48 343250524; fax: +48 343250555. E-mail addresses: [email protected] (S. Szwaja), [email protected]. czest.pl (A. Jamrozik), [email protected] (W. Tutak). 1 Tel.: +48 343250543; fax: +48 343250555. 2 Tel.: +48 343250541; fax: +48 343250555. 0306-2619/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2012.12.080
cylinder (2 stage). The system is proposed as an effective remedy to increase combustion rate and make combustion more stable, what should contribute to complete combustion inside the engine cylinder during engine expansion stroke. Among numerous environmentally harmful components in exhaust gases, emission of NOx is the most difficult to be limited. High temperature and excess of air in the engine cylinder provide favorable conditions for their formation. It was found that nitric oxides formation rate increases with increase in combustion temperature, especially above 1600 K [7]. To decrease temperature of combustion process, burning lean air–fuel mixture is proposed. It effectively affects reduction in NOx emission. Although, the EGR strategy [8] can be also considered as effective measure for NOx reduction, but investigation presented here is focused on improvement in lean mixture combustion. Regarding the excess air ratio k (defined as reciprocal of equivalence ratio) of combustible mixture, conventional spark-ignited engines should work in its narrow
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range from 0.995 to 1.005, that is around stoichiometric air to fuel ratio. This condition is required by a 3-way catalytic converter, which at this k range works within its maximal effectiveness with respect to NOx reduction. Burning rich mixtures (k < 1) is economically unjustified due to presence of combustible content in the exhaust gases, hence higher fuel consumption is observed. On the other hand, burning lean mixture (k > 1) leads to longer combustion duration [9] and higher instability in engine work as well as increase in HC and CO emission [4,5]. Therefore, a two-stage combustion system with a pre-chamber was proposed as potential solution to overcome these drawbacks [2,10–13]. In such a system a combustion chamber consists of two parts: the main chamber inside the engine cylinder and the pre-chamber in the engine head. The pre-chamber is connected to the main chamber with a channel with its diameter several times lower than the size of the prechamber. Very lean mixture characterized with k above 2.0 is prepared in the engine intake port and aspirated to the cylinder, while the nearly stoichiometric mixture is formed in the pre-chamber. Fuel–air mixture in the pre-chamber is enriched to nearly stoichiometric ratio by injecting additional amounts of fuel there. Next, the mixture in the pre-chamber is ignited by spark discharge. While combustion is taking place in the pre-chamber, pressure rapidly rises, what forces burning content in the pre-chamber to be pushed out through the connecting channel into the cylinder. As result both flame and hot exhaust gases can be treated as sources for igniting ultra-lean mixture in the cylinder. Hence, plenty of ignition kernels have been generated in the cylinder space. Consequently, lean mixture in the cylinder is simultaneously ignited in many regions, but the process cannot be considered as the HCCI, due to unburnt and burnt zones are separated from each other and temperature in these zones also differs significantly. Although, leaning the combustible mixture contributes to remarkable lengthening combustion [5,9], but the combustion goes fast enough and is completed before expansion stroke is fully developed and the exhaust valve is open. It is worth mentioning that the idea of two-stage combustion is not new. A similar combustion system with a pre-chamber was in common use in compression ignition diesel fueled engines installed in cars in past. The diesel fuel was injected directly onto hot surface of a glow plug located in a pre-chamber. Unlike direct diesel fuel injection into the cylinder, the diesel system with a prechamber featured itself with lower in-cylinder pressure rate during combusting that contributed to soft run of an engine [14–18]. Main reason for applying this system to a diesel fueled engine was to slow down combustion rate of diesel fuel, in particular at its first combustion phase. Here, the two-stage combustion system with a pre-chamber is proposed as a method for accelerating combustion rate of ultra-lean gasoline–air combustible mixture and making the combustion more stable. Even though, gasoline combustion is remarkably accelerated within the aid of this two-stage combustion system, but following general correlations between knock intensity and fuel combustion rate [19,20] anyone can conclude that the knock, which is specific for gasoline abnormal combustion, should not be present while lean mixture is burnt. One of the first attempts to study effectiveness of lean mixtures ignition and combustion in an engine with a pre-chamber was the system called the pilot flame torch ignition system. This system was patented in 1963 by Gussak et al. [10]. The two-stage combustion system with a small chamber for ignition was also subject of studies conducted by Oppenheim’s group [21]. Their research resulted in development of a controlled burning system, there lean mixture ignition was generated by the pilot flame generator (Pulsed Jet Combustion). The system was beneficial in terms of satisfactory good thermal efficiency of the engine as well as low emission of CO and HC. Roubaud and Favrat [2] investigated a combustion system with a pre-chamber to stationary biogas fueled
engines. They applied the KIVA-3V code to calculate and further to optimize pre-chamber shape and find best working conditions for the 6-in-line, turbocharged, intercooled, heavy duty engine. They concluded the pre-chamber volume should have been approximately 3% the whole cylinder volume. The pre-chamber was water-cooled and located in a cylinder head between two valves. A spark plug was located in the bottom of this pre-chamber (no fuel injection in the pre-chamber) instead of being directly installed at the top of the cylinder as is typical for the SI engine. The ignition in the main cylinder was triggered by hot gas jet from the pre-chamber as a result of combustion there. This study showed that usage of an unscavenged pre-chamber in a biogas fueled engine is particularly beneficial in terms of improvement in fuel conversion efficiency and low toxic content in the exhaust gases. Another conception for lean mixture combustion system was developed by Robinet et al. [11]. The pre-chamber was fed with rich mixture by additional fuel supply system and the pre-chamber had four holes with diameter of 1 mm each faced to the cylinder. Results from this investigation showed positive impact of the two-stage combustion on improvement in both engine work cycles stability, resistance to knock and exhaust toxic content in comparison with the conventional spark ignited engine. Further studies on improving the in-cylinder combustion process by moving the ignition point from the main in-cylinder combustion chamber to a small pre-chamber were carried out by Roethlisberger and Favrat [12]. They also confirmed positive impact of a pre-chamber, even though their pre-chamber was not fueled. Majority of automotive companies have been conducting works on a two-stage system for stratified mixture combustion. In some cases, the research resulted in implementation of new engine design into mass production. The most popular and known is the Compound Vortex Controlled Combustion (CVCC) system developed by Honda [22]. However, since gasoline direct injection into a cylinder was introduced, intensive development in combustion of stratified mixtures in engines equipped with pre-chambers was postponed [23]. On the other hand, the two-stage system for stratified mixtures combustion is mainly applied in stationary medium and high power supercharged gaseous engines operating at fixed rotational speed. The two-stage lean mixture combustion system with a pre-chamber was applied to stationary gaseous SI engines with cylinder diameter of more than 200 mm by such manufactures as Jenbacher, MAN, Wärtsilä, Dresser and Caterpillar [24]. The main goal of investigation presented in the manuscript was to prove that ultra lean gasoline–air mixtures with k up to 2.1 for in-cylinder charge (overall k = 2.0) can be effectively combusted in the engine equipped with the two-stage system, that would make it possible to apply throttle less engine operation. Furthermore, NOx emission can be found below the EURO-5 regulation limit [25].
2. The test bench description The test engine for the investigation has been modified on the basis of the four-stroke compression-ignition engine S320 ER by ANDORIA Diesel Engine Factory. The engine after several modifications was designated to work as a spark-ignition one. The reason for taking a CI engine for the investigation was to obtain high turbulence in the main in-cylinder chamber for better mixing flame and the unburnt mixture to be completely burnt. Shape of the combustion chamber in the SI engine does not provide conditions for such high in-cylinder turbulence. The main engine component which was remarkably modified was the engine head (Fig. 1).
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Fig. 1. A cross-section of the test engine head with the pre-chamber.
It was equipped with a pre-chamber, that was made of Nimonic 90 (alloy with enhanced strength at high temperatures). In the pre-chamber (its volume stands for 4.5% the total combustion volume) a spark plug and a piezo-ceramic pressure transducer were installed. The pre-chamber was connected with the cylinder by a single channel with optionally 3, 6 or 9 mm diameter. The diameter size was assumed on the basis of [12,26,27]. Second pressure transducer was installed in the main cylinder combustion chamber. A gasoline–air lean mixture pushed from the cylinder into the pre-chamber was enriched by injecting small amounts of gaseous LPG at the end of the compression stroke. Enriching the lean mixture in the pre-chamber to nearly stoichiometric ratio was required with respect to its easy and effective ignition. The LPG applied to both modeling and the tests, consisted of 50% propane and 50% butane. Volumetric fraction of the LPG injected into the pre-chamber to provide nearly-stoichiometric combustion in the pre-chamber was 20% and 36% while the overall k was 1.4 and 2.0, respectively. Similar LPG content was found in investigation conducted by others on combusting LPG–gasoline in a SI engine [6,28,29]. Although, the LPG content is relatively high, but it is associated with the pre-chamber volume. The LPG fraction referred to total amounts of gasoline is around 2.5% only. Thus, from this point of view, this phenomena cannot be managed as LPG–gasoline co-combustion. The LPG in this case is used to reinforce combustion in the pre-chamber. Other gases as hydrogen or CNG can be also applied instead of the LPG. However, LPG can be easily stored. Unlike gaseous fuels, gasoline cannot be used to enrich lean mixture in the pre-chamber. Gasoline provides several problems with proper evaporation when it is injected into a small-size pre-chamber, particularly during engine cold start. As mentioned, modeling and tests were performed with overall k of 1.4, 1.6, 1.8 and 2.0. The main parameters of the test engine are shown in Table 1. The study was conducted on a test stand consisting of the following measurement apparatus: – piezo-ceramic pressure transducer – Kistler 6061 SN 298131 – installed in the main combustion chamber, – piezo-ceramic pressure transducer – PCB Piezotronics M112B10 SN 20761 – installed in the pre-chamber,
– data acquisition system – National Instruments USB-6251 – 16 bits resolution, sampling frequency 20 kHz, – crank angle encoder – Kistler CAM 2611 – resolution 1024 pulses/rev, – UHC, CO, NOx gas analyzer – Signal 4000VM. 3. Introduction to KIVA computational code As found in the up-to-date literature [3,12,30–32] modeling combustion process inside a SI engine is satisfactory inline with results from tests due to employing sophisticated mathematical tools and calibrating models before solving it in three dimensional space. Although, the KIVA code was worked out several years ago, but it is known for its open-source feature. That provides possibilities of various modifications in its source code as far as the researcher is skillful in it. Modeling with aid of the KIVA takes into account as follows: – non-reactive flows (model cannot be modified), – reactive flows (the Chemkin solver was added and a combustion mechanism was changed). As known, the KIVA-3V code [33] allows calculation of 3-D flows in the engine cylinder including effects of turbulence and heat transfer to walls. As far as the mathematical description in details is provided by the KIVA manual [33], there is no necessity to insert all the equations here. The default model of combustion process in KIVA usually is simple and based on a single global reaction. Both non-reactive and reactive flows are modeled on the basis on the laws as follows: – – – –
continuity of mass, momentum conservation, energy conservation, species conservation.
After rearranging, these equations are in form of the threedimensional Navier–Stokes equations for a compressible fluid. Turbulence phenomena can be modeled using one of the three
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Table 1 Engine specification and fuel composition. Engine specification Displacement volume Pre-chamber volume Number of cylinders Cylinder configuration Cylinder bore Connecting-rod length Stroke Geometric compression ratio Effective compression ratio (including pre-chamber volume) Intake valve opening (before TDC) Intake valve closure (after BDC) Exhaust valve opening (before BDC) Exhaust valve closure (after TDC) Engine rotational speed
1810 cm3 10.6 cm3 1 Horizontal 120 mm 275 mm 160 mm 9 8.6 23° 40° 46° 17° 1000 rpm
Relative equivalence ratio k Fuel composition Test no. 1 Overall k = 1.4 Test no. 2 Overall k = 2.0
Fuel
BSFC (g/kW h)
Main cylinder chamber Pre-chamber
kc = 1.45 kp = 0.95
Gasoline Gasoline + LPG (2.5% by mass)
245
Main cylinder chamber Pre-chamber
kc = 2.05 kp = 1.1
Gasoline Gasoline + LPG (2.5% by mass)
240
sub-models: SGS (Sub-Grid Scale), k–e or RNG k–e (ReNormalisation Group). The k–e model was used in case of turbulence modeling, in which the kinematic viscosity depends on the turbulent kinetic energy k and the rate of this energy dissipation e. Ignition in the KIVA-3V software is realized by delivering specific amount of energy to a cell or block of cells. Heat transfer to cylinder walls was modeled using the heat transfer sub-model based on the so-called turbulent wall law (law-of-the-wall). The computational algorithm was based on the Arbitrary Lagrangian–Eulerian Method (ALE), in which a mesh of fluid was made up of arbitrary cuboids. The mesh moves with the fluid (Lagrangian scheme) and is fixed in space (Eulerian scheme). This method makes it possible to conduct calculations in two phases.
convergence in iteration process and unsatisfactory accuracy of results were reported. Thus, finally another mechanism, the C8–C16 n-alkanes [36], was taken into computations. This mechanism is based on the mechanism for the primary reference fuels PRF and was recommended by others [37]. Preliminary results obtained from the KIVA based model and the KIVA-Chemkin model, presented here, do not remarkably differ from each other as far as the basic combustion products (CO2, H2O) are considered. Finally, authors came to conclusion, that the simpler model, that uses the KIVA mechanism only, can also satisfactory determine the NOx formation due to its origin based on the thermal mechanism. However, several important radicals as e.g. H, OH, CHO playing crucial role in occurrence of abnormal combustion knock can be evaluated only in the model extended with the mechanism taken from [36] and solved with aid of the Chemkin software.
3.1. Chemical kinetic mechanisms for gasoline combustion There are several works done on improvement the combustion mechanism implicitly implemented in the KIVA software. Among others interesting approach was proposed by Kong and Reitz [34]. They implemented the Chemkin II solver into the KIVA code. Thus, following their successful achievements, the similar approach was applied into this computational simulation of combustion. As mentioned, the default mechanism implemented in the KIVA code is based on the single global chemical reaction for gasoline combustion [33].
3.2. Modeling the cylinder and the pre-chamber mesh The mesh for fluid filling both the cylinder and the pre-chamber was generated in the KIVA-3V code in accordance with the test engine geometry as depicted in Fig. 2. The mesh consisted of 24,500 cells and 27,000 nodes. It included these two combustion chambers: the pre-chamber in the engine head and the main chamber in the engine cylinder. The total volume of those chambers is 237 cm3. The pre-chamber volume of 10.6 cm3 stands for approximately 4.5% the minimal volume of the main cylinder combustion
4C8 H17 þ 49O2 ! 32O2 þ 34H2 O This reaction is split into four sub-reactions that form a simple combustion submodel. Additionally this combustion submodel is enriched with the extended Zeldovich mechanism [7] for NO formation. This is a relatively simple chemical kinetic mechanism, which makes it difficult to model combustion of the two fuels: gasoline and LPG at various equivalence ratios in two different combustion chambers, unless it was precisely calibrated. Hence, by coupling the KIVA code with the Chemkin solver, made it possible to model combustion of both gasoline and the LPG fuel simultaneously present in both the main in-cylinder combustion chamber and the pre-chamber. The each cell in the mesh was treated as the perfectly stirred reactor (PSR) for which the combustion mechanism for alkanes was applied. At first, the detailed mechanism for iso-octane [35] was used, but several problems with
Fig. 2. Geometric mesh of combustion chambers of the engine model corresponded to piston location at the TDC; look from the top-shape of the cylinder head, channel diameter = 6 mm.
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chamber, when the piston is at its TDC position. As depicted in Fig. 2 it is located asymmetrically regarding the cylinder vertical axis. The pre-chamber is connected with the main chamber by cylindrical channel with diameter optionally equaled 3, 6 or 9 mm. Within confines of the research, the 3-D model of gasoline–air mixture preparation and its combustion in a SI engine with the pre-chamber fueled with both the LPG and lean gasoline–air mixture was analyzed in the KIVA-3V code. The results from numerical modeling were compared with measurements conducted on a test engine with this two-stage combustion system built in.
4. Result and discussion 4.1. Results from modeling Computing the model in KIVA started from the IVC with a computation step of 0.06 CA deg until the EVO, but the step for data recording was 1 CA deg. At 45° before the TDC additional amounts of LPG were injected to the mixture in the pre-chamber to enrich it to nearly stoichiometric ratio. Ignition by spark discharge took place in the pre-chamber at 12° before the TDC. As result of pressure increase in the pre-chamber, the burning mixture and hot exhaust gases were pushed out through the connecting channel to the main cylinder chamber, where they ignited the lean mixture. Investigation was conducted with various excess air ratio k from 1.4 to 2.0. Results presented in the manuscript concern excess air ratio of 1.4 and 2.0 as lower and higher limit of the k in investigation. These lambdas were determined over the total fuel filling the main chamber and the pre-chamber. Computational analysis was carried out with various connecting channel diameter of as follows: 3 mm, 6 mm and 9 mm. The additional fuel injection to the pre-chamber caused the pre-chamber mixture composition was close to stoichiometric. The mean kp in the pre-chamber equaled 0.95 and 1.1 while the overall excess air ratio k was 1.4 and 2.0 respectively. Fig. 3 shows 3-D distribution of air–fuel composition expressed by the local k in the combustion chambers while the overall k was 2.0 and the mean kp in the pre-chamber was 1.1. As seen in Fig. 3, the mixture in the pre-chamber prepared during the compression stroke was not perfectly homogeneous. Excess air ratio k of the pre-chamber mixture varies from 0.715 to 2.03 while the mean kp equals 1.1 there. Computational analysis was performed with three different diameters of 3, 6 and 9 mm for the channel connecting the prechamber with the main cylinder chamber (Fig. 4). Lean mixture combustion in a classic SI engine lengthens with excess air ratio increase [5] that contributes to decrease in cylinder combustion pressure that, in consequence, decreases indicated
Fig. 4. Focus on the connecting channel diameter at the mesh of the total combustion space.
mean effective pressure (IMEP), hence causes deterioration in engine efficiency. The engine equipped with the pre-chamber characterized with faster combustion rate than a classic engine. Thus, leaner gasoline–air mixture up to k = 2.0 could be applied. Figs. 5 and 6 present combustion pressure and temperature histories in the main (in-cylinder) combustion chamber with overall k = 1.4 and 2.0, respectively. Unlike the in-cylinder pressure, that can be treated as uniform all over the cylinder space, the spatial temperature distribution cannot be. Thus, the combustion temperature history presented in Figs. 5 and 6 is the mean (over volume) in-cylinder temperature during combustion. To make comparison between the combustion events starting at the same initial pressure–temperature–volume conditions, all the computations and further tests were performed at constant spark timing of 12 CA deg before the TDC. As plotted, the highest combustion pressure was observed in case the connecting channel diameter equaled 6 mm. The phenomena of flame propagation over these both combustion chambers is depicted in Fig. 7 presenting temperature distribution in these chambers for these three cases with three various diameters for the connecting channel. As known, the highest gradient at temperature profiles corresponds to flame zone. When the channel diameter is reduced to 3 mm, then flame propagation through the channel into the main combustion chamber was more throttled. Thus, it could make higher combustion pressure and, as result, higher temperature in the pre-chamber, that led to faster combustion rate there and made combustion completed, so the flame extinguished in the pre-chamber and was unable to ignite the lean mixture in the main chamber (Fig. 7a). In case of 9 mm diameter for the channel, the flame propagated through the channel more easily, but lean mixture back flow from
Fig. 3. Distribution of excess air ratio k over combustion chambers at ignition point while the overall k = 2.0 and in the pre-chamber kp = 1.1, channel diameter = 6 mm.
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Fig. 5. Pressure (a) and temperature (b) histories at k = 1.4.
Fig. 6. Pressure (a) and temperature (b) histories at k = 2.0.
Fig. 7. Spatial temperature distribution in both combustion chambers during combustion progress at k = 2.0.
the cylinder to the pre-chamber was also intensified (Fig. 7c) due to the piston squeezing effect at compression stroke. Hence, the back flow generated a swirl in the pre-chamber with higher momentum. Thus, significantly higher amounts of lean mixture in the pre-chamber were burnt at lower combustion rate, that caused the overall combustion process slower and incomplete. As depicted in Fig. 7 the fastest combustion history is presented in case b (6 mm channel diameter). Furthermore, temperature inside
the pre-chamber dropped faster in this case when compared with the channel of 9 mm. The faster temperature drop stands for lower heat transfer to walls. As known from the diesel engine knowledge, the pre-chamber is responsible for remarkably (up to 8% – [14]) high heat losses there. Thus, lower heat losses, then higher engine overall efficiency. It should be also noticed, that the diameter of 6 mm was found as the best one, but it should not be treated as the optimal one. The optimization function with respect to channel
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diameter was not worked out. These three values 3, 6 and 9 mm proposed for the channel diameter were presented to show influence of the channel diameter on progress in combustion. Although, these diameters were taken arbitrarily without any optimization works done before, but, as analyzed, the diameter of 6 mm was found as close to its optimal size. Further analysis concerns heat release rate (HRR) and accumulative heat released during combustion process. Figs. 8 and 9 depict heat release rate during combustion and cumulative heat released against crank angle for the model with these three different diameters of the connecting channel at k = 1.4 and 2.0, respectively. Plots presented in Figs. 8 and 9 confirm that the most favorable conditions for the fastest combustion were provided in the engine equipped with the channel diameter of 6 mm. In this case peak in the heat release rate was the highest among all these three cases (Figs. 8 and 9a). That results in earlier end of combustion as depicted in graphs for cumulative heat (Figs. 8 and 9b). As evidenced by now, the numerical analysis provides justification for implementing the pre-chamber into the spark ignited gasoline engine. The Fig. 10 shows temperature distribution in both the conventional engine and the engine with the pre-chamber with the channel diameter of 6 mm. Results were computed at k = 2.0, under which such the ultra-lean gasoline–air mixture is very difficult to be ignited and completely burnt in a conventional spark ignited engine due to high instabilities in flames and requirements for extremely high energy to create ignition kernel and initiate combustion. As far as the combustion model in KIVA does not contain any sub-models for flames instabilities, the combustion history for the conventional engine (left column in Fig. 10) should be treated as an ideal process with regard to flame propagation. In case the pre-chamber is installed, then combustion starts inside it at stoichiometric ratio. Thus, energy of spark discharge is high enough to initiate combustion there. Next, high energy in flames thrown from the pre-chamber into the cylinder can easily ignite the lean mixture. Additionally, higher turbulence by flame jet accelerates combustion in the cylinder chamber. Thus, flame instabilities associated with lean mixture combustion are less crucial and might be neglected in the engine equipped with the pre-chamber. As seen in Fig. 10, the combustion process goes faster in the engine equipped with the pre-chamber, even though temperature of the intake mixture at the ignition is lower due to lower compression ratio (the conventional engine CR = 9, the engine with the prechamber CR = 8.6). Computed both the HRR and the cumulative heat released during combustion at k of 1.4 and 2.0 are plotted in Figs. 11 and 12, respectively. As depicted in these figures the thesis on faster combustion process with aid of the pre-chamber can be confirmed. As seen in Figs. 11 and 12a, peak in the HRR in
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the engine with the pre-chamber is significantly advanced with respect to the HRR peak in the conventional engine. Additionally, profile of the cumulative heat (Figs. 11 and 12b) also appears earlier. Unfortunately, heat losses were not quantified. Furthermore, authors did not also break down the overall heat losses into prechamber and main chamber contribution. Such studies were conducted by Rakopoulos and Giakoumis [14], however they worked with a compression ignition engine. They found the contribution from a pre-chamber is approximately 8% of the total heat losses to walls although the total pre-chamber surface constituted 5% of the total surface of the combustion chambers. 4.2. Comparison between modeling results and results from experimental studies As mentioned, there were several difficulties in the conventional spark ignited engine during combustion tests with the air– gasoline mixtures at k of 1.4 and higher up to 2.0. The engine work was unsteady with lot of misfiring events. Thus, results presented here concern the engine equipped with the pre-chamber. The primary target of these tests was to verify the numerical model. The model evaluation was determined on the basis of comparison the combustion pressure traces from the test engine and from the model. Additionally, temperature courses from these both works were also analyzed. As plotted in Fig. 13 these courses are in good consistency with each other and provide premises for applying 3-D modeling to analyze the two-stage combustion in the real engine. Difference between the results obtained from modeling and from testing the real engine can be used to validate the numerical model. 4.3. Toxic emission As discussed, combustion goes faster when the pre-chamber is applied, hence, following the thermal NOx mechanism by Zeldovich [7] higher NOx emission is expected. However, the highest temperature was observed in the pre-chamber, in which the mixture was stoichiometric or even was slightly rich, thus, there were oxygen amounts not enough to form high NOx concentration there. Mixture in the cylinder chamber was highly lean, that resulted in lower combustion temperature and lower rate of NOx formation. The exhaust NOx emission from the engine and from modeling is shown in Fig. 14. As seen, at k = 1.4 the NOx from the test engine without and with the pre-chamber is compared. Additionally, the modeled NOx is also included. As found, the NOx from the conventional engine was approximately 2.5 times higher than the NOx from the engine with the pre-chamber.
Fig. 8. Heat release rate (a) and cumulative heat released (b) during combustion vs crank angle at k = 1.4.
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Fig. 9. Heat release rate (a) and cumulative heat released (b) during combustion vs crank angle at k = 2.0.
Fig. 10. Spatial temperature distribution in a conventional engine and in the engine with the pre-chamber (channel diameter of 6 mm) at overall k = 2.0.
Among the toxic compounds in the engine exhaust gases, only the NOx emission was significantly reduced through applying the two-stage combustion system with the pre-chamber. Other toxic emission as unburnt hydrocarbons (UHCs) and carbon monoxide
(CO) were not reduced, even they were slightly increased due to pre-chamber cooling effect. Moreover, small pre-chamber size contributes to higher number of three-body reactions that lead to terminate combustion in the pre-chamber. Summing up, as far as
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Fig. 11. Heat release rate (a) and cumulative heat released (b) during single and two stage combustion vs crank angle at k = 1.4 for the channel diameter of 6 mm.
Fig. 12. Heat release rate (a) and cumulative heat released (b) in the conventional engine and the engine with the pre-chamber vs crank angle at k = 2.0 for the channel diameter of 6 mm.
Fig. 13. Pressure and temperature histories from the KIVA-3 V model and the test engine at k = 1.4 and 2.0 for the channel diameter of 6 mm.
4.4. Engine stability and fuel consumption Instabilities in engine work concern combustion process in its macro-scale and they can be measured and expressed by the COV of IMEP. The COV is determined from computing the IMEP from consecutive engine work cycles. It was defined as follows:
COVIMEP ¼ STDIMEP =IMEPMEAN 100%
Fig. 14. Comparison between the NOx emission from the model and the test engine for the channel diameter of 6 mm.
the NOx can be limited below the EURO 5 level, only the oxidation catalytic converter can be applied for CO and UHC reduction, what reduces costs of the exhaust gas after-treatment. Hence, there was no need for the CO and UHC to be addressed.
where STDIMEP is the standard deviation of the indicated mean effective pressure (IMEP) and IMEPMEAN is the mean of the IMEP over a single test series at fixed excess air ratio. As plotted in Fig. 15 the COV for the test series from the engine with the pre-chamber is relatively low and is below the limit of 10% recommended by engine manufacturers and research centers [24,38]. The Qin in Figs. 15 and 16 is energy in fuel filing the pre-chamber, Qtot stands for energy stored in the total fuel burnt. With regard to fuel consumption, Fig. 16 shows brake specific fuel consumption (BSFC) against excess air ratio. As seen, the BSFC for the engine with the pre-chamber is at the same level as fuel consumption for the conventional engine at excess air ratio of
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COVIMEP, %
9 Q in / Q tot = 2.5% Q in / Q tot = 8.0%
7
Engine with a prechamber
Engine without a prechamber
5 3 1 0,7
0,9
1,1
1,3
1,5
1,7
1,9
2,1
4. As found from modeling, diameter of the channel connecting the pre-chamber with the cylinder chamber is important with respect to flow processes and strongly affects combustion rate and duration in the main cylinder chamber. 5. Applying ultra lean mixtures can be useful at engine partial load and makes it possible to introduce throttle less engine operation. 6. The combustion system with the pre-chamber is used in highpower stationary gaseous spark ignited engines, thus, all the conclusions can be found as useful hints in further developing this technology.
λ Fig. 15. COV of IMEP against excess air ratio.
References 370 Q in / Q tot = 2.5% Q in / Q tot = 8.0%
BSFC [g/kWh]
340 Engine without a prechamber
310
Engine with a prechamber
280 250 220
0,7
0,9
1,1
1,3
1,5
1,7
1,9
2,1
λ Fig. 16. BSFC vs excess air ratio.
approximately 1.2. At this k the conventional engine works at its maximal indicated efficiency. 5. Conclusions 1. The spark ignited engine equipped with the pre-chamber provides better conditions for burning lean gasoline–air mixtures than the classic engine. This two-stage combustion system ensures not only burning ultra-lean gasoline–air mixtures with excess air ratio k up to 2.0, but also it remarkably speeds up combustion process and improves stability in engine work. Results from tests show that the engine was working steadily without any misfiring events. The COV of IMEP was approximately 3% for engine running on ultra lean mixture with k = 2.0. Additionally, specific fuel consumption of 245 g/kW h was almost the same if compared to the engine without a pre-chamber working at k close to stoichiometric ratio. 2. Results from modeling the two-stage combustion were in satisfactory good correlation with results from tests. The coefficient of multiple determination (CoMD) for combustion pressure was above 0.93. Hence modeling combustion process in the engine with the pre-chamber can be considered as useful tool in the engine computational analysis. Although, modeling combustion process of ultra-lean mixture in a conventional engine can be charged with relatively high error as far as sub-models for flames instabilities are not involved. With regard to applying the pre-chamber, these instabilities can be neglected. Fuel ignition in the main cylinder chamber can be treated as ignition in large space by flame released from the pre-chamber. Due to low volume in the pre-chamber, its impact for overall combustion rates can be treated as negligible. 3. The exhaust NOx emission was in the range 0.37–1.71 g/kW h while the engine was operating at k of 1.8 and higher. This emission is below the EURO-5 threshold for HDV motor vehicles powered by liquid fuel (2.0 g/kW h), even though the NOx catalytic reactor was not applied.
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Glossary BDC: bottom dead centre BSFC: brake specific fuel consumption CA: crank angle, (°) CNG: compressed natural gas CO: carbon monoxide CR: compression ratio EVC: exhaust valve closure EVO: exhaust valve open EGR: exhaust gases recirculation HCCI: homogeneous charge compression ignition HRR: heat release rate, J/° IVC: intake valve closure IVO: intake valve open LPG: liquefied petroleum gas NOx: nitric oxides SI: spark ignition TDC: top dead centre UHC: unburned hydrocarbons p: pressure, Pa T: temperature, K V: volume, m3 t: time, s k: excess air ratio (lambda) averaged over the main chamber and the pre-chamber kp: excess air ratio in the pre-chamber
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
A two-stage combustion system for burning lean gasoline mixtures in a stationary spark ignited engine Stanislaw Szwaja ⇑, Arkadiusz Jamrozik 1, Wojciech Tutak 2 Institute of Thermal Machinery, Czestochowa University of Technology, 42-200 Czestochowa, ul. Dabrowskiego 69, Poland
h i g h l i g h t s " A two-stage combustion system was implemented to a spark ignited gasoline engine. " Extremely lean gasoline mixtures (lambda up to 2.0) were burnt effectively. " The NOx emission from the engine was found to be below the EURO-5 limit. " Effective burning ultra lean mixtures at partial load allow engine throttleless run. " 3D modeling was satisfactory verified with the experiments.
a r t i c l e
i n f o
Article history: Received 28 September 2012 Received in revised form 28 December 2012 Accepted 31 December 2012 Available online 4 February 2013 Keywords: Combustion Stationary engine Two-stage combustion system Pre-chamber Heat release rate NOx
a b s t r a c t The paper mainly focuses on applying the two-stage combustion system with a pre-chamber into the stationary internal combustion spark ignited engine. It especially concentrates on applying throttle less operation at partial load and reduction of the NOx emission. Considerations conducted in the paper are based on the in-cylinder combustion progress analysis. Additionally, analysis of tailpipe toxic emission, with particular focus on the NOx formation in the engine equipped with the pre-chamber, is also performed. The paper presents both results of 3-D combustion modeling in the SI engine and results conducted on a test SI engine. The 3-D modeling was performed in the KIVA-3V code. Next, results from modeling were compared with results obtained from tests. Finally, satisfactory good consistency between modeled and experimental courses of both pressure, temperature and NOx were obtained. Thus, the engine model with the proposed two-stage combustion system properly simulates engine working conditions on the test bed. Results from both analyses confirmed that the two-stage combustion system significantly shortens combustion duration of an ultra lean gasoline–air mixture and contributes to reduction in NOx. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction Combustion a lean gasoline–air mixture in a spark ignited (SI) engine is one method to reduce nitric oxide (NOx) emissions and increase engine efficiency by decreasing a peak combustion temperature [1–4]. However, while a lean mixture is combusted, combustion process gets longer and incomplete that causes engine work unstable and decrease in engine brake efficiency [4–6]. The paper presents a two-stage combustion system, in which combustion starts in a pre-chamber (1 stage) and further flames jet from a pre-chamber initiate lean mixture combustion in the engine
⇑ Corresponding author. Tel.: +48 343250524; fax: +48 343250555. E-mail addresses: [email protected] (S. Szwaja), [email protected]. czest.pl (A. Jamrozik), [email protected] (W. Tutak). 1 Tel.: +48 343250543; fax: +48 343250555. 2 Tel.: +48 343250541; fax: +48 343250555. 0306-2619/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2012.12.080
cylinder (2 stage). The system is proposed as an effective remedy to increase combustion rate and make combustion more stable, what should contribute to complete combustion inside the engine cylinder during engine expansion stroke. Among numerous environmentally harmful components in exhaust gases, emission of NOx is the most difficult to be limited. High temperature and excess of air in the engine cylinder provide favorable conditions for their formation. It was found that nitric oxides formation rate increases with increase in combustion temperature, especially above 1600 K [7]. To decrease temperature of combustion process, burning lean air–fuel mixture is proposed. It effectively affects reduction in NOx emission. Although, the EGR strategy [8] can be also considered as effective measure for NOx reduction, but investigation presented here is focused on improvement in lean mixture combustion. Regarding the excess air ratio k (defined as reciprocal of equivalence ratio) of combustible mixture, conventional spark-ignited engines should work in its narrow
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range from 0.995 to 1.005, that is around stoichiometric air to fuel ratio. This condition is required by a 3-way catalytic converter, which at this k range works within its maximal effectiveness with respect to NOx reduction. Burning rich mixtures (k < 1) is economically unjustified due to presence of combustible content in the exhaust gases, hence higher fuel consumption is observed. On the other hand, burning lean mixture (k > 1) leads to longer combustion duration [9] and higher instability in engine work as well as increase in HC and CO emission [4,5]. Therefore, a two-stage combustion system with a pre-chamber was proposed as potential solution to overcome these drawbacks [2,10–13]. In such a system a combustion chamber consists of two parts: the main chamber inside the engine cylinder and the pre-chamber in the engine head. The pre-chamber is connected to the main chamber with a channel with its diameter several times lower than the size of the prechamber. Very lean mixture characterized with k above 2.0 is prepared in the engine intake port and aspirated to the cylinder, while the nearly stoichiometric mixture is formed in the pre-chamber. Fuel–air mixture in the pre-chamber is enriched to nearly stoichiometric ratio by injecting additional amounts of fuel there. Next, the mixture in the pre-chamber is ignited by spark discharge. While combustion is taking place in the pre-chamber, pressure rapidly rises, what forces burning content in the pre-chamber to be pushed out through the connecting channel into the cylinder. As result both flame and hot exhaust gases can be treated as sources for igniting ultra-lean mixture in the cylinder. Hence, plenty of ignition kernels have been generated in the cylinder space. Consequently, lean mixture in the cylinder is simultaneously ignited in many regions, but the process cannot be considered as the HCCI, due to unburnt and burnt zones are separated from each other and temperature in these zones also differs significantly. Although, leaning the combustible mixture contributes to remarkable lengthening combustion [5,9], but the combustion goes fast enough and is completed before expansion stroke is fully developed and the exhaust valve is open. It is worth mentioning that the idea of two-stage combustion is not new. A similar combustion system with a pre-chamber was in common use in compression ignition diesel fueled engines installed in cars in past. The diesel fuel was injected directly onto hot surface of a glow plug located in a pre-chamber. Unlike direct diesel fuel injection into the cylinder, the diesel system with a prechamber featured itself with lower in-cylinder pressure rate during combusting that contributed to soft run of an engine [14–18]. Main reason for applying this system to a diesel fueled engine was to slow down combustion rate of diesel fuel, in particular at its first combustion phase. Here, the two-stage combustion system with a pre-chamber is proposed as a method for accelerating combustion rate of ultra-lean gasoline–air combustible mixture and making the combustion more stable. Even though, gasoline combustion is remarkably accelerated within the aid of this two-stage combustion system, but following general correlations between knock intensity and fuel combustion rate [19,20] anyone can conclude that the knock, which is specific for gasoline abnormal combustion, should not be present while lean mixture is burnt. One of the first attempts to study effectiveness of lean mixtures ignition and combustion in an engine with a pre-chamber was the system called the pilot flame torch ignition system. This system was patented in 1963 by Gussak et al. [10]. The two-stage combustion system with a small chamber for ignition was also subject of studies conducted by Oppenheim’s group [21]. Their research resulted in development of a controlled burning system, there lean mixture ignition was generated by the pilot flame generator (Pulsed Jet Combustion). The system was beneficial in terms of satisfactory good thermal efficiency of the engine as well as low emission of CO and HC. Roubaud and Favrat [2] investigated a combustion system with a pre-chamber to stationary biogas fueled
engines. They applied the KIVA-3V code to calculate and further to optimize pre-chamber shape and find best working conditions for the 6-in-line, turbocharged, intercooled, heavy duty engine. They concluded the pre-chamber volume should have been approximately 3% the whole cylinder volume. The pre-chamber was water-cooled and located in a cylinder head between two valves. A spark plug was located in the bottom of this pre-chamber (no fuel injection in the pre-chamber) instead of being directly installed at the top of the cylinder as is typical for the SI engine. The ignition in the main cylinder was triggered by hot gas jet from the pre-chamber as a result of combustion there. This study showed that usage of an unscavenged pre-chamber in a biogas fueled engine is particularly beneficial in terms of improvement in fuel conversion efficiency and low toxic content in the exhaust gases. Another conception for lean mixture combustion system was developed by Robinet et al. [11]. The pre-chamber was fed with rich mixture by additional fuel supply system and the pre-chamber had four holes with diameter of 1 mm each faced to the cylinder. Results from this investigation showed positive impact of the two-stage combustion on improvement in both engine work cycles stability, resistance to knock and exhaust toxic content in comparison with the conventional spark ignited engine. Further studies on improving the in-cylinder combustion process by moving the ignition point from the main in-cylinder combustion chamber to a small pre-chamber were carried out by Roethlisberger and Favrat [12]. They also confirmed positive impact of a pre-chamber, even though their pre-chamber was not fueled. Majority of automotive companies have been conducting works on a two-stage system for stratified mixture combustion. In some cases, the research resulted in implementation of new engine design into mass production. The most popular and known is the Compound Vortex Controlled Combustion (CVCC) system developed by Honda [22]. However, since gasoline direct injection into a cylinder was introduced, intensive development in combustion of stratified mixtures in engines equipped with pre-chambers was postponed [23]. On the other hand, the two-stage system for stratified mixtures combustion is mainly applied in stationary medium and high power supercharged gaseous engines operating at fixed rotational speed. The two-stage lean mixture combustion system with a pre-chamber was applied to stationary gaseous SI engines with cylinder diameter of more than 200 mm by such manufactures as Jenbacher, MAN, Wärtsilä, Dresser and Caterpillar [24]. The main goal of investigation presented in the manuscript was to prove that ultra lean gasoline–air mixtures with k up to 2.1 for in-cylinder charge (overall k = 2.0) can be effectively combusted in the engine equipped with the two-stage system, that would make it possible to apply throttle less engine operation. Furthermore, NOx emission can be found below the EURO-5 regulation limit [25].
2. The test bench description The test engine for the investigation has been modified on the basis of the four-stroke compression-ignition engine S320 ER by ANDORIA Diesel Engine Factory. The engine after several modifications was designated to work as a spark-ignition one. The reason for taking a CI engine for the investigation was to obtain high turbulence in the main in-cylinder chamber for better mixing flame and the unburnt mixture to be completely burnt. Shape of the combustion chamber in the SI engine does not provide conditions for such high in-cylinder turbulence. The main engine component which was remarkably modified was the engine head (Fig. 1).
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Fig. 1. A cross-section of the test engine head with the pre-chamber.
It was equipped with a pre-chamber, that was made of Nimonic 90 (alloy with enhanced strength at high temperatures). In the pre-chamber (its volume stands for 4.5% the total combustion volume) a spark plug and a piezo-ceramic pressure transducer were installed. The pre-chamber was connected with the cylinder by a single channel with optionally 3, 6 or 9 mm diameter. The diameter size was assumed on the basis of [12,26,27]. Second pressure transducer was installed in the main cylinder combustion chamber. A gasoline–air lean mixture pushed from the cylinder into the pre-chamber was enriched by injecting small amounts of gaseous LPG at the end of the compression stroke. Enriching the lean mixture in the pre-chamber to nearly stoichiometric ratio was required with respect to its easy and effective ignition. The LPG applied to both modeling and the tests, consisted of 50% propane and 50% butane. Volumetric fraction of the LPG injected into the pre-chamber to provide nearly-stoichiometric combustion in the pre-chamber was 20% and 36% while the overall k was 1.4 and 2.0, respectively. Similar LPG content was found in investigation conducted by others on combusting LPG–gasoline in a SI engine [6,28,29]. Although, the LPG content is relatively high, but it is associated with the pre-chamber volume. The LPG fraction referred to total amounts of gasoline is around 2.5% only. Thus, from this point of view, this phenomena cannot be managed as LPG–gasoline co-combustion. The LPG in this case is used to reinforce combustion in the pre-chamber. Other gases as hydrogen or CNG can be also applied instead of the LPG. However, LPG can be easily stored. Unlike gaseous fuels, gasoline cannot be used to enrich lean mixture in the pre-chamber. Gasoline provides several problems with proper evaporation when it is injected into a small-size pre-chamber, particularly during engine cold start. As mentioned, modeling and tests were performed with overall k of 1.4, 1.6, 1.8 and 2.0. The main parameters of the test engine are shown in Table 1. The study was conducted on a test stand consisting of the following measurement apparatus: – piezo-ceramic pressure transducer – Kistler 6061 SN 298131 – installed in the main combustion chamber, – piezo-ceramic pressure transducer – PCB Piezotronics M112B10 SN 20761 – installed in the pre-chamber,
– data acquisition system – National Instruments USB-6251 – 16 bits resolution, sampling frequency 20 kHz, – crank angle encoder – Kistler CAM 2611 – resolution 1024 pulses/rev, – UHC, CO, NOx gas analyzer – Signal 4000VM. 3. Introduction to KIVA computational code As found in the up-to-date literature [3,12,30–32] modeling combustion process inside a SI engine is satisfactory inline with results from tests due to employing sophisticated mathematical tools and calibrating models before solving it in three dimensional space. Although, the KIVA code was worked out several years ago, but it is known for its open-source feature. That provides possibilities of various modifications in its source code as far as the researcher is skillful in it. Modeling with aid of the KIVA takes into account as follows: – non-reactive flows (model cannot be modified), – reactive flows (the Chemkin solver was added and a combustion mechanism was changed). As known, the KIVA-3V code [33] allows calculation of 3-D flows in the engine cylinder including effects of turbulence and heat transfer to walls. As far as the mathematical description in details is provided by the KIVA manual [33], there is no necessity to insert all the equations here. The default model of combustion process in KIVA usually is simple and based on a single global reaction. Both non-reactive and reactive flows are modeled on the basis on the laws as follows: – – – –
continuity of mass, momentum conservation, energy conservation, species conservation.
After rearranging, these equations are in form of the threedimensional Navier–Stokes equations for a compressible fluid. Turbulence phenomena can be modeled using one of the three
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Table 1 Engine specification and fuel composition. Engine specification Displacement volume Pre-chamber volume Number of cylinders Cylinder configuration Cylinder bore Connecting-rod length Stroke Geometric compression ratio Effective compression ratio (including pre-chamber volume) Intake valve opening (before TDC) Intake valve closure (after BDC) Exhaust valve opening (before BDC) Exhaust valve closure (after TDC) Engine rotational speed
1810 cm3 10.6 cm3 1 Horizontal 120 mm 275 mm 160 mm 9 8.6 23° 40° 46° 17° 1000 rpm
Relative equivalence ratio k Fuel composition Test no. 1 Overall k = 1.4 Test no. 2 Overall k = 2.0
Fuel
BSFC (g/kW h)
Main cylinder chamber Pre-chamber
kc = 1.45 kp = 0.95
Gasoline Gasoline + LPG (2.5% by mass)
245
Main cylinder chamber Pre-chamber
kc = 2.05 kp = 1.1
Gasoline Gasoline + LPG (2.5% by mass)
240
sub-models: SGS (Sub-Grid Scale), k–e or RNG k–e (ReNormalisation Group). The k–e model was used in case of turbulence modeling, in which the kinematic viscosity depends on the turbulent kinetic energy k and the rate of this energy dissipation e. Ignition in the KIVA-3V software is realized by delivering specific amount of energy to a cell or block of cells. Heat transfer to cylinder walls was modeled using the heat transfer sub-model based on the so-called turbulent wall law (law-of-the-wall). The computational algorithm was based on the Arbitrary Lagrangian–Eulerian Method (ALE), in which a mesh of fluid was made up of arbitrary cuboids. The mesh moves with the fluid (Lagrangian scheme) and is fixed in space (Eulerian scheme). This method makes it possible to conduct calculations in two phases.
convergence in iteration process and unsatisfactory accuracy of results were reported. Thus, finally another mechanism, the C8–C16 n-alkanes [36], was taken into computations. This mechanism is based on the mechanism for the primary reference fuels PRF and was recommended by others [37]. Preliminary results obtained from the KIVA based model and the KIVA-Chemkin model, presented here, do not remarkably differ from each other as far as the basic combustion products (CO2, H2O) are considered. Finally, authors came to conclusion, that the simpler model, that uses the KIVA mechanism only, can also satisfactory determine the NOx formation due to its origin based on the thermal mechanism. However, several important radicals as e.g. H, OH, CHO playing crucial role in occurrence of abnormal combustion knock can be evaluated only in the model extended with the mechanism taken from [36] and solved with aid of the Chemkin software.
3.1. Chemical kinetic mechanisms for gasoline combustion There are several works done on improvement the combustion mechanism implicitly implemented in the KIVA software. Among others interesting approach was proposed by Kong and Reitz [34]. They implemented the Chemkin II solver into the KIVA code. Thus, following their successful achievements, the similar approach was applied into this computational simulation of combustion. As mentioned, the default mechanism implemented in the KIVA code is based on the single global chemical reaction for gasoline combustion [33].
3.2. Modeling the cylinder and the pre-chamber mesh The mesh for fluid filling both the cylinder and the pre-chamber was generated in the KIVA-3V code in accordance with the test engine geometry as depicted in Fig. 2. The mesh consisted of 24,500 cells and 27,000 nodes. It included these two combustion chambers: the pre-chamber in the engine head and the main chamber in the engine cylinder. The total volume of those chambers is 237 cm3. The pre-chamber volume of 10.6 cm3 stands for approximately 4.5% the minimal volume of the main cylinder combustion
4C8 H17 þ 49O2 ! 32O2 þ 34H2 O This reaction is split into four sub-reactions that form a simple combustion submodel. Additionally this combustion submodel is enriched with the extended Zeldovich mechanism [7] for NO formation. This is a relatively simple chemical kinetic mechanism, which makes it difficult to model combustion of the two fuels: gasoline and LPG at various equivalence ratios in two different combustion chambers, unless it was precisely calibrated. Hence, by coupling the KIVA code with the Chemkin solver, made it possible to model combustion of both gasoline and the LPG fuel simultaneously present in both the main in-cylinder combustion chamber and the pre-chamber. The each cell in the mesh was treated as the perfectly stirred reactor (PSR) for which the combustion mechanism for alkanes was applied. At first, the detailed mechanism for iso-octane [35] was used, but several problems with
Fig. 2. Geometric mesh of combustion chambers of the engine model corresponded to piston location at the TDC; look from the top-shape of the cylinder head, channel diameter = 6 mm.
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chamber, when the piston is at its TDC position. As depicted in Fig. 2 it is located asymmetrically regarding the cylinder vertical axis. The pre-chamber is connected with the main chamber by cylindrical channel with diameter optionally equaled 3, 6 or 9 mm. Within confines of the research, the 3-D model of gasoline–air mixture preparation and its combustion in a SI engine with the pre-chamber fueled with both the LPG and lean gasoline–air mixture was analyzed in the KIVA-3V code. The results from numerical modeling were compared with measurements conducted on a test engine with this two-stage combustion system built in.
4. Result and discussion 4.1. Results from modeling Computing the model in KIVA started from the IVC with a computation step of 0.06 CA deg until the EVO, but the step for data recording was 1 CA deg. At 45° before the TDC additional amounts of LPG were injected to the mixture in the pre-chamber to enrich it to nearly stoichiometric ratio. Ignition by spark discharge took place in the pre-chamber at 12° before the TDC. As result of pressure increase in the pre-chamber, the burning mixture and hot exhaust gases were pushed out through the connecting channel to the main cylinder chamber, where they ignited the lean mixture. Investigation was conducted with various excess air ratio k from 1.4 to 2.0. Results presented in the manuscript concern excess air ratio of 1.4 and 2.0 as lower and higher limit of the k in investigation. These lambdas were determined over the total fuel filling the main chamber and the pre-chamber. Computational analysis was carried out with various connecting channel diameter of as follows: 3 mm, 6 mm and 9 mm. The additional fuel injection to the pre-chamber caused the pre-chamber mixture composition was close to stoichiometric. The mean kp in the pre-chamber equaled 0.95 and 1.1 while the overall excess air ratio k was 1.4 and 2.0 respectively. Fig. 3 shows 3-D distribution of air–fuel composition expressed by the local k in the combustion chambers while the overall k was 2.0 and the mean kp in the pre-chamber was 1.1. As seen in Fig. 3, the mixture in the pre-chamber prepared during the compression stroke was not perfectly homogeneous. Excess air ratio k of the pre-chamber mixture varies from 0.715 to 2.03 while the mean kp equals 1.1 there. Computational analysis was performed with three different diameters of 3, 6 and 9 mm for the channel connecting the prechamber with the main cylinder chamber (Fig. 4). Lean mixture combustion in a classic SI engine lengthens with excess air ratio increase [5] that contributes to decrease in cylinder combustion pressure that, in consequence, decreases indicated
Fig. 4. Focus on the connecting channel diameter at the mesh of the total combustion space.
mean effective pressure (IMEP), hence causes deterioration in engine efficiency. The engine equipped with the pre-chamber characterized with faster combustion rate than a classic engine. Thus, leaner gasoline–air mixture up to k = 2.0 could be applied. Figs. 5 and 6 present combustion pressure and temperature histories in the main (in-cylinder) combustion chamber with overall k = 1.4 and 2.0, respectively. Unlike the in-cylinder pressure, that can be treated as uniform all over the cylinder space, the spatial temperature distribution cannot be. Thus, the combustion temperature history presented in Figs. 5 and 6 is the mean (over volume) in-cylinder temperature during combustion. To make comparison between the combustion events starting at the same initial pressure–temperature–volume conditions, all the computations and further tests were performed at constant spark timing of 12 CA deg before the TDC. As plotted, the highest combustion pressure was observed in case the connecting channel diameter equaled 6 mm. The phenomena of flame propagation over these both combustion chambers is depicted in Fig. 7 presenting temperature distribution in these chambers for these three cases with three various diameters for the connecting channel. As known, the highest gradient at temperature profiles corresponds to flame zone. When the channel diameter is reduced to 3 mm, then flame propagation through the channel into the main combustion chamber was more throttled. Thus, it could make higher combustion pressure and, as result, higher temperature in the pre-chamber, that led to faster combustion rate there and made combustion completed, so the flame extinguished in the pre-chamber and was unable to ignite the lean mixture in the main chamber (Fig. 7a). In case of 9 mm diameter for the channel, the flame propagated through the channel more easily, but lean mixture back flow from
Fig. 3. Distribution of excess air ratio k over combustion chambers at ignition point while the overall k = 2.0 and in the pre-chamber kp = 1.1, channel diameter = 6 mm.
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Fig. 5. Pressure (a) and temperature (b) histories at k = 1.4.
Fig. 6. Pressure (a) and temperature (b) histories at k = 2.0.
Fig. 7. Spatial temperature distribution in both combustion chambers during combustion progress at k = 2.0.
the cylinder to the pre-chamber was also intensified (Fig. 7c) due to the piston squeezing effect at compression stroke. Hence, the back flow generated a swirl in the pre-chamber with higher momentum. Thus, significantly higher amounts of lean mixture in the pre-chamber were burnt at lower combustion rate, that caused the overall combustion process slower and incomplete. As depicted in Fig. 7 the fastest combustion history is presented in case b (6 mm channel diameter). Furthermore, temperature inside
the pre-chamber dropped faster in this case when compared with the channel of 9 mm. The faster temperature drop stands for lower heat transfer to walls. As known from the diesel engine knowledge, the pre-chamber is responsible for remarkably (up to 8% – [14]) high heat losses there. Thus, lower heat losses, then higher engine overall efficiency. It should be also noticed, that the diameter of 6 mm was found as the best one, but it should not be treated as the optimal one. The optimization function with respect to channel
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diameter was not worked out. These three values 3, 6 and 9 mm proposed for the channel diameter were presented to show influence of the channel diameter on progress in combustion. Although, these diameters were taken arbitrarily without any optimization works done before, but, as analyzed, the diameter of 6 mm was found as close to its optimal size. Further analysis concerns heat release rate (HRR) and accumulative heat released during combustion process. Figs. 8 and 9 depict heat release rate during combustion and cumulative heat released against crank angle for the model with these three different diameters of the connecting channel at k = 1.4 and 2.0, respectively. Plots presented in Figs. 8 and 9 confirm that the most favorable conditions for the fastest combustion were provided in the engine equipped with the channel diameter of 6 mm. In this case peak in the heat release rate was the highest among all these three cases (Figs. 8 and 9a). That results in earlier end of combustion as depicted in graphs for cumulative heat (Figs. 8 and 9b). As evidenced by now, the numerical analysis provides justification for implementing the pre-chamber into the spark ignited gasoline engine. The Fig. 10 shows temperature distribution in both the conventional engine and the engine with the pre-chamber with the channel diameter of 6 mm. Results were computed at k = 2.0, under which such the ultra-lean gasoline–air mixture is very difficult to be ignited and completely burnt in a conventional spark ignited engine due to high instabilities in flames and requirements for extremely high energy to create ignition kernel and initiate combustion. As far as the combustion model in KIVA does not contain any sub-models for flames instabilities, the combustion history for the conventional engine (left column in Fig. 10) should be treated as an ideal process with regard to flame propagation. In case the pre-chamber is installed, then combustion starts inside it at stoichiometric ratio. Thus, energy of spark discharge is high enough to initiate combustion there. Next, high energy in flames thrown from the pre-chamber into the cylinder can easily ignite the lean mixture. Additionally, higher turbulence by flame jet accelerates combustion in the cylinder chamber. Thus, flame instabilities associated with lean mixture combustion are less crucial and might be neglected in the engine equipped with the pre-chamber. As seen in Fig. 10, the combustion process goes faster in the engine equipped with the pre-chamber, even though temperature of the intake mixture at the ignition is lower due to lower compression ratio (the conventional engine CR = 9, the engine with the prechamber CR = 8.6). Computed both the HRR and the cumulative heat released during combustion at k of 1.4 and 2.0 are plotted in Figs. 11 and 12, respectively. As depicted in these figures the thesis on faster combustion process with aid of the pre-chamber can be confirmed. As seen in Figs. 11 and 12a, peak in the HRR in
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the engine with the pre-chamber is significantly advanced with respect to the HRR peak in the conventional engine. Additionally, profile of the cumulative heat (Figs. 11 and 12b) also appears earlier. Unfortunately, heat losses were not quantified. Furthermore, authors did not also break down the overall heat losses into prechamber and main chamber contribution. Such studies were conducted by Rakopoulos and Giakoumis [14], however they worked with a compression ignition engine. They found the contribution from a pre-chamber is approximately 8% of the total heat losses to walls although the total pre-chamber surface constituted 5% of the total surface of the combustion chambers. 4.2. Comparison between modeling results and results from experimental studies As mentioned, there were several difficulties in the conventional spark ignited engine during combustion tests with the air– gasoline mixtures at k of 1.4 and higher up to 2.0. The engine work was unsteady with lot of misfiring events. Thus, results presented here concern the engine equipped with the pre-chamber. The primary target of these tests was to verify the numerical model. The model evaluation was determined on the basis of comparison the combustion pressure traces from the test engine and from the model. Additionally, temperature courses from these both works were also analyzed. As plotted in Fig. 13 these courses are in good consistency with each other and provide premises for applying 3-D modeling to analyze the two-stage combustion in the real engine. Difference between the results obtained from modeling and from testing the real engine can be used to validate the numerical model. 4.3. Toxic emission As discussed, combustion goes faster when the pre-chamber is applied, hence, following the thermal NOx mechanism by Zeldovich [7] higher NOx emission is expected. However, the highest temperature was observed in the pre-chamber, in which the mixture was stoichiometric or even was slightly rich, thus, there were oxygen amounts not enough to form high NOx concentration there. Mixture in the cylinder chamber was highly lean, that resulted in lower combustion temperature and lower rate of NOx formation. The exhaust NOx emission from the engine and from modeling is shown in Fig. 14. As seen, at k = 1.4 the NOx from the test engine without and with the pre-chamber is compared. Additionally, the modeled NOx is also included. As found, the NOx from the conventional engine was approximately 2.5 times higher than the NOx from the engine with the pre-chamber.
Fig. 8. Heat release rate (a) and cumulative heat released (b) during combustion vs crank angle at k = 1.4.
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Fig. 9. Heat release rate (a) and cumulative heat released (b) during combustion vs crank angle at k = 2.0.
Fig. 10. Spatial temperature distribution in a conventional engine and in the engine with the pre-chamber (channel diameter of 6 mm) at overall k = 2.0.
Among the toxic compounds in the engine exhaust gases, only the NOx emission was significantly reduced through applying the two-stage combustion system with the pre-chamber. Other toxic emission as unburnt hydrocarbons (UHCs) and carbon monoxide
(CO) were not reduced, even they were slightly increased due to pre-chamber cooling effect. Moreover, small pre-chamber size contributes to higher number of three-body reactions that lead to terminate combustion in the pre-chamber. Summing up, as far as
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Fig. 11. Heat release rate (a) and cumulative heat released (b) during single and two stage combustion vs crank angle at k = 1.4 for the channel diameter of 6 mm.
Fig. 12. Heat release rate (a) and cumulative heat released (b) in the conventional engine and the engine with the pre-chamber vs crank angle at k = 2.0 for the channel diameter of 6 mm.
Fig. 13. Pressure and temperature histories from the KIVA-3 V model and the test engine at k = 1.4 and 2.0 for the channel diameter of 6 mm.
4.4. Engine stability and fuel consumption Instabilities in engine work concern combustion process in its macro-scale and they can be measured and expressed by the COV of IMEP. The COV is determined from computing the IMEP from consecutive engine work cycles. It was defined as follows:
COVIMEP ¼ STDIMEP =IMEPMEAN 100%
Fig. 14. Comparison between the NOx emission from the model and the test engine for the channel diameter of 6 mm.
the NOx can be limited below the EURO 5 level, only the oxidation catalytic converter can be applied for CO and UHC reduction, what reduces costs of the exhaust gas after-treatment. Hence, there was no need for the CO and UHC to be addressed.
where STDIMEP is the standard deviation of the indicated mean effective pressure (IMEP) and IMEPMEAN is the mean of the IMEP over a single test series at fixed excess air ratio. As plotted in Fig. 15 the COV for the test series from the engine with the pre-chamber is relatively low and is below the limit of 10% recommended by engine manufacturers and research centers [24,38]. The Qin in Figs. 15 and 16 is energy in fuel filing the pre-chamber, Qtot stands for energy stored in the total fuel burnt. With regard to fuel consumption, Fig. 16 shows brake specific fuel consumption (BSFC) against excess air ratio. As seen, the BSFC for the engine with the pre-chamber is at the same level as fuel consumption for the conventional engine at excess air ratio of
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COVIMEP, %
9 Q in / Q tot = 2.5% Q in / Q tot = 8.0%
7
Engine with a prechamber
Engine without a prechamber
5 3 1 0,7
0,9
1,1
1,3
1,5
1,7
1,9
2,1
4. As found from modeling, diameter of the channel connecting the pre-chamber with the cylinder chamber is important with respect to flow processes and strongly affects combustion rate and duration in the main cylinder chamber. 5. Applying ultra lean mixtures can be useful at engine partial load and makes it possible to introduce throttle less engine operation. 6. The combustion system with the pre-chamber is used in highpower stationary gaseous spark ignited engines, thus, all the conclusions can be found as useful hints in further developing this technology.
λ Fig. 15. COV of IMEP against excess air ratio.
References 370 Q in / Q tot = 2.5% Q in / Q tot = 8.0%
BSFC [g/kWh]
340 Engine without a prechamber
310
Engine with a prechamber
280 250 220
0,7
0,9
1,1
1,3
1,5
1,7
1,9
2,1
λ Fig. 16. BSFC vs excess air ratio.
approximately 1.2. At this k the conventional engine works at its maximal indicated efficiency. 5. Conclusions 1. The spark ignited engine equipped with the pre-chamber provides better conditions for burning lean gasoline–air mixtures than the classic engine. This two-stage combustion system ensures not only burning ultra-lean gasoline–air mixtures with excess air ratio k up to 2.0, but also it remarkably speeds up combustion process and improves stability in engine work. Results from tests show that the engine was working steadily without any misfiring events. The COV of IMEP was approximately 3% for engine running on ultra lean mixture with k = 2.0. Additionally, specific fuel consumption of 245 g/kW h was almost the same if compared to the engine without a pre-chamber working at k close to stoichiometric ratio. 2. Results from modeling the two-stage combustion were in satisfactory good correlation with results from tests. The coefficient of multiple determination (CoMD) for combustion pressure was above 0.93. Hence modeling combustion process in the engine with the pre-chamber can be considered as useful tool in the engine computational analysis. Although, modeling combustion process of ultra-lean mixture in a conventional engine can be charged with relatively high error as far as sub-models for flames instabilities are not involved. With regard to applying the pre-chamber, these instabilities can be neglected. Fuel ignition in the main cylinder chamber can be treated as ignition in large space by flame released from the pre-chamber. Due to low volume in the pre-chamber, its impact for overall combustion rates can be treated as negligible. 3. The exhaust NOx emission was in the range 0.37–1.71 g/kW h while the engine was operating at k of 1.8 and higher. This emission is below the EURO-5 threshold for HDV motor vehicles powered by liquid fuel (2.0 g/kW h), even though the NOx catalytic reactor was not applied.
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Glossary BDC: bottom dead centre BSFC: brake specific fuel consumption CA: crank angle, (°) CNG: compressed natural gas CO: carbon monoxide CR: compression ratio EVC: exhaust valve closure EVO: exhaust valve open EGR: exhaust gases recirculation HCCI: homogeneous charge compression ignition HRR: heat release rate, J/° IVC: intake valve closure IVO: intake valve open LPG: liquefied petroleum gas NOx: nitric oxides SI: spark ignition TDC: top dead centre UHC: unburned hydrocarbons p: pressure, Pa T: temperature, K V: volume, m3 t: time, s k: excess air ratio (lambda) averaged over the main chamber and the pre-chamber kp: excess air ratio in the pre-chamber