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Ethanol and Biogas Fuels

Comparison of Performance and Pollution Emission of Engine
Fueled with Gasoline Ethanol Blended Fuels and Biogas

 
Bui Van Ga, Tran Van Nam, Nguyen Van Đong, Bui Van Tan
University of Science and Technology, The University of Danang

Journal of Science and Technology 112 (2016), pp. 93-99

 

 
Abstract
Indicated cycle work of Daewoo engine fueled with E15 at speed 5000rpm increases 5.3% as engine compression ratio increased from 9.5 to 10.3. However the indicated cycle work decreased by 20% when the engine speed increases from 2000rpm to 5000rpm with a given compression ratio. If advance ignition timming is kept constant, the engine cycle work slightly decreases while rising ethanol concentration in gasoline. Indicated engine cycle work decreases about 3% as fueled with E30 and decreases 17% as fueled with biogas containing 95% CH4 in comparison with gasoline case.
As a given engine speed, as engine compression ratio increased from 9.5 to 10.3, the concentration of NOx in the exhaust gas rose 7% and the concentration of CO increased 1%. As the engine speed increases from 2000rpm to 5000rpm NOx concentration in exhaust gas decreases 78.5% while the concentration of CO increased by 5%. As the engine fueled with E5 and E30, NOx in exhaust gas concentrations increase, respectively 3% and 15% while reducing the corresponding CO levels were 17% and 87% compared to gasoline. When running on biogas containing 95% CH4, NOx concentration in the exhaust gas decreased to 43% and CO concentration decreased to 39% compared to gasoline. NOx concentration decreased slightly but CO concentration is almost unchanged while reducing engine load. NOx concentration increases slightly with admission mixture temperature but decreases strongtly with exhaust gas recycled.
Keywords: Alternative Fuels, Ethanol, Biogas, Pollution Emission, Engine Performance

1. Introduction

COP21 summit on climate change organized by the United Nations in Paris has made historic decision after 20 years of negotiations. 195 countries agreed to take actions to reduce the level of gas emissions causing the greenhouse effect for the Earth's temperature does not rise to 20C than average temperatures from 1889 to 1899 period. Increased use of renewable fuel derived from solar energy is one of the fundamental solutions to achieve COP21
Vietnam is a tropical country with nearly 80% of the population live in rural areas. Waste from agriculture and animal husbandry are very good raw material for the production of bio-fuels. Ethanol and biogas are renewable fuels that can be produced from organic substances. Their use does not increase fuel gases causing the greenhouse effect in the atmosphere.
Ethanol blended gasoline has been widely used in many countries around the world with different mixing ratio. According to the regulations of the Government of Vietnam, gasoline 5% volume ethanol (called E5) are widely used across the country from the date 1-12-2015. Ethanol is organic compound, located in the homologous series of ethyl alcohol, has a high octane number than gasoline. Ethanol can therefore be used to raise the octane of the fuel to improve the efficiency of the combustion process in internal combustion engines [1].
Mustafa Koc studied the effects of E50 and E85 ethanol blended gasoline to the engine functionality and pollution emission levels at compression ratio 10, 11 and engine speed ranges from 1500 to 5000 rpm. The results showed that when ethanol is mixed into gasoline, motor torque and fuel consumption increased, but the level of pollution emissions reduction [2]. This study also showed that ethanol blended gasoline allows increase compression ratio of engine without detonation occurs. Due to the latent heat of evaporation and burning temperature of ethanol are higher than gasoline, latency burn time of ethanol is prolonged. Therefore, to increase the efficiency of ethanol blended gasoline-powered engines we need to increase advance ignition angle of the engine as a function of ethanolcontent.Richie Daniel studied empirically the effects of advance ignition angle to performance of ethanol and gasoline engines [3]. Results showed that at a speed 1500 rpm, angle-ignition is about 7-80 for gasoline and about 220 for ethanol [3]. Phuangwongtrakul tested gasohol with different content of ethanol. Results showed that, at 5000 rpm, the largest torque achieved is with ignition angles are 300, 350 and 400 with E10, E30 and E85, respectively [4].
In this paper we study the combustionprocess and performance ofethanol blended gasoline-powered enginewithdifferent concentration of ethanol. Effects of ignition angle and the engine compression ratio are also examined and evaluated.The experiment was conducted at Engine and Automobile Testing Center, University of Science and Technology, The University of Danang.The experiment systemis equipped with AVL’s laboratory equipment. Detailed description of the engine testing system is presented in [5]. Daewoo engine A16DMS was renovated to experiment with a mixture of petrol and ethanol. Thatis a 4-stroke engine, 4 cylinders, cylinder diameter79mm, stroke81,5mm, total cylinder volume 1598cm3, compression ratio 9.5. The maximum powerof the engine withgasoline is 78kW at 5800rpm.
Experimental fuel mixture is E15whichcomposes15% ethanol and 85% RON92trade gasoline, by volume.

2. Simulation of burning process

Figure 1 presents the cylinder and the engine combustion chamber Daewoo wasmeshedto simulate with ANSYS FLUENT software. In the simulationwe used the model k-ε turbulence model and the partially premixed combustion model. Modellingand setting dynamic nets duringthe piston displacement insideengine cylinder is presented in [6]. The fuel’scomponents areadjustedand chemical thermodynamic characteristics of burning mixedare set in a map in order to access quickly during simulationof combustion in the cylinder. In the following simulation, equivalence coefficientof the mixture is selected f = 1.1, ignition angle is 250before TDC.






 


Figure 2 shows the development of burningarea on the cross section which is 8 mm from the top of the combustion chamber. On the cross section, it can be seen that flame is acircle with the radius increases with the crankshaft angle. Figure 3apresentspressureand temperaturevariationin the engine cylinder; Figure 3b presents pressure variations and concentrations of O2, gasoline and ethanol withthe crankshaft angle. At the start of burning, concentration of the substancesdecreases whiletemperature and pressure in the cylinder  increases rapidly.Temperature peak is later than the peak of pressure about 200. Figure 4 presents the fuel concentration variation at speed of 3000 rpm with E15 and with biogas containing 95% CH4. This result shows that the consumption rate of CH4 is approximate gasoline consumption rate. Meanwhile ethanol consumption rate is much lower than the two aforementioned fuel types. Figure 5 presents the influenceof the engine speed to the variationofaverage gasolineconcentration in the combustion mixtureduring burning process. It can be seen thatwhen the enginespeedincreases, the rate of fuel consumption decrease.The correlation between the temperature of the gas mixture and concentrations of CO, NOx when the engine runs on E15 is presented in Figure 6a and 6b.  The mass concentrations of pollutants were averaged over the entire volume of the combustion chamber.It can be seen thatafter the launch of burning process, pollutant concentrations increased rapidly with the increase in temperature and reaches a maximum value before the temperature reaches its peak. CO formation rate depends on the concentration and temperature of the fuel. MaximumCO concentration is achieved atthe high temperature and fuel concentration.




 

Meanwhile, the rate of formation of NOx depends on the oxygen concentration and temperature.Above results show that CO and NOx reach their peaks almost at the same position of crankshaft with different engine speeds(Figure 6a, b). For biogas fuels, peak of COconcentrationis achievedlater than one of NOx (Figure 7). After reaching the maximum value, CO and NOx oxidized with residual oxygen in the mixture and therefore their concentrations reduceduring expansion.

3. Analysis of engine performance
Figure 8a, b introduce the influence of engine speed to work diagram when engine fueled by E15 with compression ratio 9.5 and 10.3.It can be seen thatin thesameoperating conditionsindicated work of the engine decreases as the engine speed increases. This can be explained by the time for burning process decrease whenengine speedincreases. A fire in the fuel department expansion phase, reducing the fertility.A portion of the fuel burns in the expansion phasereducing the fertilityof work.When increaseengine speed from 2000 rpmto 5000 rpm with compression ratio 9.5, the indicatedcycleworkdecreases from 574J/cycle to 457J/cycle(about 20%).Similarly, with compression ratio of 10.3, the indicated cycle work decreases from 588J/cycle to 481J/cycle (about 18%) with the same level of engine speed (Figure 9). When the engine compression ratio increases from 9.5 to 10.3, the indicated cycle work increases by 2.4% at 2000 rpm and by 5.3% at 5000 rpm.
 


 
Figure 10 compares the work diagrams of the motor fueled by gasoline, ethanol blended gasoline with different concentrations and biogas at 3000 rpm. It can be seen that with the same advance angle ignition the indicated cycle work in case of ethanol blended gasoline decreases slightly in comparison to case of gasoline. Specifically, when fueled by traditional gasoline, E5, E15, E30, the indicated cycle works are 511 J /cycle, 510 J /cycle, 506 J/cycle and 493 J /cycle, respectively. Simulationresults also showed that when powered by biogas containing 95% and 100% CH4indicatedcyclework is only425J/cycle and 426J/cycle. Thus the indicated cyclework decreases by 17% when switching from petrol to biogas with CH4 concentrations greater than 95%. It means that when running on E5 and E15, engine power is almost unchanged in comparision to running on gasoline. In case of E30, indicated cycle workdecreasesby 3% in comparision to when running on gasoline and this reduction can be overcome by adjust the ignition angle appropriately.


 
4. Analysis of emission pollutants
Figure 11 introduces the variations ofNOx concentrationin combustion chamber withthe enginecrankshaft rotation, corresponding tocompression ratio 9.5 and 10.3 and engine speed range from 2000 rpmto 5000 rpm.It can be seen thatNOxconcentration reaches a maximum value athigh temperature, then gradually decreases and maintainstability tothe end of the expansion.
This stable concentrations are considered as the concentration of NOx in engine exhaust. At a given engine speed, when the compression ratio increased from 9.5 to 10.3, the NOx concentration increased by 7%. This can be explained by the increasing of solvent temperature which accelerates the formation speed of NOx. As engine speed increases, the concentration of NOx fell sharply (Figure 12). NOx concentration in the exhaust gas decreases to 78.5% when the engine speed increases from 2000 rpmto 5000 rpm. The reason is that when the engine speed increases leading to the increasing of turbulent movement and burns the roots forming NOx.
Variations of CO concentrations withcrankshaft angle, is similar to NOx variation but steep sharply at high temperature areas (Figure 13). Increasingcompression ratio slightly increasesCO concentration in the exhaust gas.When the compression ratio increasesfrom 9.5 to 10.3, the concentration of CO in the exhaust gas increases only 1%.






 

Although there are significant differences in the concentrationof CO in the maximumarea,at the end of the expansion the level of difference of CO concentration is not large and the effect of compression ratio to concentrationof CO in emissions is negligible. As the engine speed increases from 2000 rpmto 5000 rpmconcentrations of CO in exhaust gas increased by 5% (Figure 14).
Influencesof fuel to the NOx concentration variation in the engine combustion chamber isintroduced in Figure 15. The simulation results are calculated with compression ratio of 9.5 and atspeed of3000 rpm. NOx concentration in the exhaust gas increases with content of ethanol blended with gasoline. When fueled by E5and E30,the concentration of NOx increase50 ppm(3%)and 250 ppm (15%) respectively, in comparison to the case of running on traditional gasoline(Figure 16).Meanwhile, in comparison with the gasoline, when powered by biogas containing 95% and 100% CH4, NOx concentration reduces to43% and 41%, respectively.
Unlike NOx, CO concentration decreases with increasing content of ethanol in petrol. Figure 17 introduced CO concentrationsvariation withcrank angle in case ofrunning on biogas and ethanol-blended gasoline atthe different components.

 

It can be seen that when content of ethanol increases, theconcentration of CO in the exhaust gas rapidly decrease. CO concentration reduction,compared to traditional gasoline,whenfueled byE5, E15 and E30 are 17%, 51% and 87%, respectively (Figure 18).Calculation results also showed that when content of CH4 in the biogas increases from 95% to 100% concentration of CO in the exhaust gas does not significantly change. Concentration of CO when running on biogas decreases 39% in comparison to running on gasoline and equivalent to running on E15.
Load also significantly affect concentrations of NOx in the exhaust gas. When the engine load is reduced,NOx emissions also decreased (Figure 19). This can be explained by the temperature of the mixture dropped reducing NOx creationspeed. Meanwhile the load changes almost does not affect level of CO emissions (Figure 20).



Intake air temperature slightly affected the level of NOx emissions. Figure 21 introduction of NOx variation with crank angle when engine is fueled by E15, compression ratio of 9.5 with intake air temperature is assumed 303K, 352K and 373K. It can be seen thatan increasingintake air temperature causes increasing of NOxconcentration in combustion products. This can be explained by the temperature of the combustion mixture increases with temperature intake air.These results indicate that NOx concentration depends on the temperature of the mixture. To reduce NOx concentrations, especially when the engine is running at low load, the modern engine is equipped with exhaust gas recirculation systems. After mixing the new intake air and exhaust gases (mainly CO2), the temperature of the combustion mixture decrease leads to lower levels of NOx in the exhaust gas.






 


5. Conclusion

The above research results allow us to draw the following conclusions:
1.      When engine is fueled by E15 with compression ratio increasing from 9.5 to 10.3, indicated cycle workincreases by 2.4% at 2000 rpmand by 5.3% at 5000 rpm. Conversely the motor indicated cycle workdecreasesby 20% when the engine speed increases from 2000 rpmto 5000 rpm.
2.      Without adjust the ignition angle, in case of using ethanol blended gasoline with ethanol content below 15%, the indicated cycle workis equivalent to the case of using traditional gasoline. When ethanolcontent increased to 30%, the indicated cycle workdecreased by 3% compared to the case of using traditional gasoline.When powered by biogas containingmore than95% CH4 the indicated cycle workdecreased by 17% compared to when running on traditional gasoline.
3.      Ata given engine speed, when compression ratio increasesfrom 9.5 to 10.3, the NOx concentration increasesby 7% and concentration of CO increases 1%. When engine speed increases from 2000 rpmto 5000 rpm, the concentration of NOx in exhaust gas decreasesby 78.5% while the concentration of CO increasesby 5%.
4.      Whenrunning onE5 and E30 , the NOx concentration in the exhaust gas increasesby3% and 15%, respectively,and CO concentration reducesby17% and 87%, respectively,compared to when running on gasoline. Meanwhile fueledby biogascontaining 95% CH4, NOx concentrations decreasesby43% and concentrations of CO reduction 39%in comparison to case ofrunning on traditional gasoline.
5.      When reducing the engine load, the concentration of NOx in exhaust gas decreasesbut CO concentration almost less varies with load. The concentration of NOx in the exhaust gas increase slightly with the temperatureof intake air, but decreases dramatically with content of emissiongas recirculatedinto the manifold..


References

[1].    Huseyin Serdar Yucesu, Tolga Topgul, Can Cinat, Melih Okur: Effect of ethanol–gasoline blends on engine performance and exhaust emissions in different compression ratios. Applied Thermal Engineering, Volume 26, Issues 17-18, December 2006, Pages 2272–2278

[2].    Mustafa Koc, Yakup Sekmen, Tolga Topgül, Hüseyin Serdar Yücesu: The effects of ethanol–unleaded gasoline blends on engine performance and exhaust emissions in a spark-ignition engine. Renewable Energy, Volume 34, Issue 10, October 2009, Pages 2101–2106

[3].    Richie Daniel, Guohong Tian, Hongming Xu, Shijin Shuai: Ignition timing sensivities of oxygenated biofuels compared to gasoline in direct-injection SI engine. Fuel 99 (2012), pp. 72-82

[4].    S. Phuangwongtrakul, K.Wannatong, T. Laungnarutai and W. Wechsatol: Suitable Ignition Timing and Fuel Injection Duration for Ethanol-Gasoline Blended Fuels in a Spark Ignition Internal Combustion Engine. Proc. of the Intl. Conf. on Future Trends in Structural, Civil, Environmental and Mechanical Engineering, FTSCEM 2013, ISBN: 978-981-07-7021-1, pp. 39-42

[5].    Bui Van Ga, Nguyen Viet Hai, Nguyen Van Anh, Vo Anh Vu, Bui Van Hung: in cylinder pressure analysis in biogas-diesel dual fuel engine by simulation and experiment. UD’s Journal of Science and Technology, Vol.01(86), 2015, pp.24-29

[6].    Bui Van Ga, Tran Van Nam, Tran Thanh Hai Tung: A Simulation of Effects of Compression Ratios on the Combustion in Engines Fueled With Biogas with Variable CO2 Concentrations. Journal of Engineering Research and Application www.ijera.com Vol. 3, Issue 5, Sep-Oct 2013, pp.516-523 (IF: 1,69).

Diễn biến trong buồng cháy

Diễn biến trong buồng cháy động cơ xe gắn máy chạy bằng biogas nén

Clips diễn biến nhiệt độ trong buồng cháy



Clips diễn biến nồng độ O2 trong buồng cháy



Clips diễn biến nồng độ CH4 trong buồng cháy

 

hybrid

Biogas Combustion in Motorcycle Engine

Turbulent Burning Velocity in Combustion Chamber
of SI Engine Fueled with Compressed Biogas  


Bui Van Ga, Nguyen Van Dong, Bui Van Hung
GATEC, University of Science and Technology, The University of Danang
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Vietnam Journal of Mechanics, Volume 37, Number 3, pp 205-216, 2015

 

Abstract

Turbulentburningvelocity is the most important parameter in analyzing pre-mixed combustion simulation of spark ignition engines. It depends on the laminar burning velocity and turbulence intensity in the combustion chamber. The first term can be predicted if one knows fuel composition, physico chemical properties of the fluid. The second term strongly depends on the geometry of the combustion chamber and fluid movement during the combustion process. One cannot suggest a general expression for different cases of engine. Thus, for accuracy modeling, one  should determine turbulent burning velocity in the combustion chamber of each case of engine individually.
In this study,the turbulent burning velocity is defined by a linear function of laminar burning velocity in which the proportional constant is defined as the turbulent burning velocity coefficient. This coefficient was obtained by analyzing the numerical simulation results and experimental data and this is applied to a concrete case of a Honda Wave motorcycle engine combustion chamber that fueled with compressed biogas.
The results showed that the turbulent burning velocity coefficient in this case is around 1.3 when the average engine revolutions is in the range of 3000 rpm to 6000 rpm with biogas containing 80% Methane. We can then predict the effects of different parameters on the performance of the engine fueled with compressed biogas by simulation.

Keywords:Turbulent burning velocity, Combustion simulation, Biogas engine, Spark ignition engine, Biogas fuel

Nomenclature:
ST :      Turbulent burning velocity (m/s)
SL :      Laminar burning velocity (m/s)
SLo :     Laminar burning velocity at atmospheric condition (m/s)
AT :      Wrinkled flame surface area (m2)
AL:      Flow cross section area (m2)
ff :        Turbulent burning velocity coefficient
u' :       Turbulent intensity (m/s)
k :        Turbulent kinetic energy (m2/s2)
e :        Turbulent kinetic energy dissipation rate (m2/s3)
T :       Fluid temperature (K)
P :        Pressure (Pa)
To, Tu: Temperature at atmosphericcondition and temperature of unburnt mixture (K)
Po, Pu: Pressure at atmosphericcondition and pressure of unburnt mixture (Pa)
n:        Engine speed (rpm)
ao :      Constant
f:        Fuel-air equivalence ratio
js :       Advance spark timing angle (°)
TDC:  Top dead center
CA:    Crank angle (°)

1. Introduction

            Biogas is a renewable energy which is interchangeable with natural gas. However, the problem which has existed until today is that all of the biogases yielded by different biogas digestion tanks are of low pressure, low specific gravity and large specific volume. The large quantity of CO2 present in biogas lowers its calorific value, flame velocity and flammability range compared with natural gas. Therefore, the biogas needs to be purified and compressed before it can be used in engines, especially in vehicle engines.
            Vietnam is tropical country and waste from agriculture production is abundant to produce biogas. Besides, most of individual vehicles used in the country are motorcycles, so the application of biogas on this kind of vehicle will be a good way for fossil fuel saving and climate change mitigation. Research group GATEC of the University of Danang is the pioneer in developing technology of gaseous fuel application on the motorcycle [1]. The research is initially carried out on LPG fuel [2] and now it is shifted to compressed biogas [3,4].
            In order to convert a gasoline motorcycle engine to be a biogas engine, we should carry out a theoretical study on engine performance before doing technical modification. One of the most important unresolved problems of this study is the determination of the turbulent burning velocity in the combustion chamber of the engine. There is no consensus in literature whether the turbulent burning velocity is a characteristic quantity that can be defined unambiguously for different geometries.
            Turbulent premixed flame propagation was first investigated by Damköhler (1940). He observed that the burning speed increases as Reynolds number increases and it was affected by two different scales of turbulence: low intensity u'/SL, large scale turbulence (weakly wrinkled flames); and high intensity u'/SL, small scale turbulence (strongly wrinkled flames) [5,6]. To develop mathematical models, Damköhler assumed that the flamelet propagates with a constant velocity in a one-dimensional plane. Thus, the only effect of turbulence is the wrinkling of the flame front which results in an increase in flame surface area but internal structure and SL are unchanged. The turbulent burning velocity was originally defined as follows [7, 8]
                                    ST/S L= AT/AL                                                                 (1)
            Where SL is the laminar burning velocity, AT is the wrinkled flame surface area and AL is the flow cross section area [5,9,10]. With such assumption, Damköhler developed the first following model:
                                                                                           (2)
            Where u' is the turbulence intensity.
            The high pressure conditions have an effect on the thermo-physical properties of the air-fuel mixture and the turbulent structures become finer as the pressure increases [11], as well as a decrease of laminar burning velocity and the thickening of the laminar flame front [12]. Kobayashi [11] has reported an increasing in turbulent burning velocity as pressure increases and he suggested empirical correlation for methane/air flames:
                                                                              (3)
            Where P is the operating pressure and Po is the atmospheric pressure.
            In these above expressions, laminar burning velocity SLdepends on the physico chemical characteristics of the mixture and it can be calculated if one knows the details of rate of chemical reactions taking place in the combustion process.In general, this data is established for mixture of pure fuel and air. M. Elia et al. [13] found a relationship between laminar burning velocity and physical-chemical behavior of the mixture before combustion. R. Stone and A. Clarke [14] conducted experiments to determine the laminar burning velocity CH4-air mixture diluted by CO2 atmosphere. Laminar burning velocity of methane-air mixture can be established by empirical expressions as follows:
SL=0.366 T1.42P- 0.297 (m/s)                                        (4)
Where T=Tu/To is dimensionless temperature and P=Pu/Po is dimensionless pressure.
According to Rallis and Garforth [15],laminar burning velocity of methane-air stoichiometric mixture can be expressed by:
                                                (5)
Where aois in interval of 1.37 and 2.33
Biogas-air mixture couldbe considered as methane-air mixture diluted by CO2. Laminar burning velocity in this case is presented in [16]. Thus, with a given cylinder pressure, mixture temperature, and composition of biogas, we can calculate laminar burning velocity. Turbulent burning velocity depends not only on laminar burning velocity but also on fluid movement in the combustion chamber. In this work,we simplify the relationship between turbulent burning velocity and laminar burning velocity by an assumptionof ST=ff.SL. We try to determine the turbulent burning velocity coefficient ffin the combustion chamber of a Honda Wave motorcycle engine fueled with compressed biogas.
            In view of the above, the specific aims of this study were defined as follows:
(1) To carry out experimental measurement of the indicated cylinder pressure
(2) To carry out simulation calculation with identical experimental condition
(3) To compare the simulation results with experimental data in order to determine the turbulent burning velocity coefficient in the combustion chamber.

2. Method of study

2.1 Experimental setup
The experiment wasconducted at the research laboratory of internal combustion engines of the Hanoi University of Science and Technology. The experimental facilities include a CD20" Chassis Dynamometer Test Bed AVL for motorcycles that is controlled by Zoller software. Equipment Indiset 620 including a computer with the IndiWin620 software for data acquisition from different sensors such as the pressure in the combustion chamber, detonation, advance spark timing angle, TDC position, etc. The cylinder pressure is obtained by means of a piezoelectric transducer. A synchronizationoptical crank angle encoder 364C was used to acquire the cylinder pressure on the basic of crank angle rather than the time.Operational parameters such as speed, acceleration and power of the motorcycle are displayed on the control screen. Experimental data is treated by Concerto software. A schematic layout of the experimental facilities is shown in Fig. 1. Fig. 2 illustrates pictures of instruments for biogas motorcycle testing at the laboratory.


            The following mass flow rates were determined: (1) air, by measurement of the pressure drop across an orifice; (2) biogas, by means of rotameters; (3) petrol, by weighing and timing. Hence, we can estimate the equivalence ratio of the mixture.
Compressed biogas from pressure cylinders is supplied to the engine with help of conversion kit GATEC25 [18]. This is a special gas carburetor that ensures stable equivalence ratio at any regime of the engine which was fitted to the engine's air intake upstream of the petrol carburetor. Biogas flow rate could be adjusted by a needle valve in the gas inlet port. A pointer and scale arrangement indicating percentage of full throttle opening was retrofitted to the butterfly valve. The fuel was inducted into the throat of the venturi and the mixture flow rate was controlled by the throttle.
2.2 Numerical simulation      
            The combustionof biogas-air mixture in the combustion chamber of the engine is simulated using Ansys FLUENT computational fluid dynamics software. The structural and operational parameters of the engine are introduced into the program via the dynamic mesh option. Thermodynamic properties of the working fluid are set in PrePDF table integrated into the FLUENT software. Turbulent combustion is simulated via the k-eturbulence model and the partially premixed combustion model with laminar burning velocity SLwhich is determined empirically with fuel containing two major components, which are CH4 and CO2. The turbulent burning velocity is determined via SL and a given turbulent burning velocity coefficient by simplified relationship ST=ff.SL.
Figure3a illustrates the dimensions of the combustion chamber and cylinder of a typical single-cylinder Honda Wave motorcycle engine. It is a four-stroke engine with bore, D = 50mm, stroke, S = 49.50mm, and a rated power output of 5.1kW at 8000rpm.The compression ratio of the engine is 9:1. The ignition system is powered by a CDI with an essentially fixed spark timing of 30° before TDC. The combustion chamber of the engine is hemispherical in shape with a bowl-shaped cylinder head and a flat piston top.
Figure 3b shows the meshingof the computational space. To avoid the occurrence of negative volume elements caused by element deformation during piston displacement, the combustion chamber and the cylinder are meshed separately.

3. Results and Discussion
3.1. Simulation results
            In the following section, the turbulent burning velocity coefficient is fixed at f= 1.3 and then we predict the effects of the equivalence ratio and composition of biogas to the performance of the Honda Wave motorcycle engine. The effect of advance spark timing angle and engine speed has been published in previous works [17].
Figure4 presents flame propagation at different crank angle and variation of temperature and pressure as result in the combustion chamber of the engine fueled with biogas containing 85% CH4. We observe that flame front initially has a spherical shape and then it is deformed during spreading out in space of the combustion chamber away from the  spark plug. Apeak of pressure occurred at approximately13 degreesafter TDCand a peak of combustion temperature occurredat about 5 degrees later.

            Figure 5a and Figure 5b introduce the variation of concentrations of CH4 and O2 in the combustion chamber with engine revolution speed of 3000rpm, advance spark timing angle 30°and equivalence ratio of 0.9, 1,0, and 1,5. The higher the slope of the curve, the higher rate of fuel and oxidizer consumption. The results show that the highest fuel consumption rate is achieved at an equivalence ratio of f= 1. When f= 0.9, in the early phase of combustion, fuel consumption rate is not different with f= 1 case but at the end of combustion process, the difference is more evident on O2 consumption curve.  The slope of the curve for variation of CH4 and O2 with f=1.5 is significantly lower than the previous two cases.
            Figure6a and 6b show the indicated cylinder pressure diagrams and the indicated cycle work diagrams for various equivalence ratioswith engine revolution speed of 3000rpm, advance spark timing angle of 30°and biogas fuel containing 85% CH4. The indicated cycle work of the engine is represented by the area bounded in the compression and expansion curves. Variation of the indicated cycle work versus fis presented in Fig. 7. We observe that the indicated cycle work of the engine reaches its maximum value with fin range from 1 to 1.1 corresponding to the zone with the highest value of combustion rate.

            The following section presents the effects of biogas components on engine performance. The calculations are carried out with advance spark timing angle 35°, equivalence ratio f= 1 and engine speed n = 5000rpm.  Biogas fuel contains 60%, 70%, and 80% CH4. Figure 8a and 8b show the variation of indicated pressure and indicated cycle work versus CH4 component in biogas as engine runs at 5000rpm, advance spark timing angle of 35°, equivalence ratio f= 1. When the concentration of CH4 in biogas increases, the maximum pressure of the engine is increased leading to an increasing of the indicated cycle work. Figure 8a illustratesthat at the fixed equivalence ratio, the combustion chamber peak pressure decreases gradually with the introduction of carbon dioxide into the mixture, due to the lower reactive charge inducted and the thermal release rate with the increase of CO2 giving rise to the above observations. The result shows that the indicated cycle work increases linearly with CH4 composition in biogas, as shown in Fig. 9.


3.2. Experimental measurements      
The experiment was carried out firstly with gasoline RON92 and then mainly with compressed biogas. The full load curves of the engine were established as the throttle valve was fully opened. Figure 10a and 10b present the comparison of the indicated cylinder pressure and the indicated work diagram of the Honda Wave motorcycle engine fueled with gasoline RON92 and fueled with biogas containing 80% CH4 at the same operating conditions: engine revolution speed of 3000rpm, stoichiometric mixture, advance spark timing angle 30o. The results showed that when using biogas, the peak cylinder pressure is 35bar which is lower than when gasoline was used (57bar). When switching from gasoline to biogas, the maximum cylinder pressure drops and it can be explained by two reasons: firstly, volume efficiency decreases because of gas fuel, and secondly, reduction in burning velocity and heat value of the mixture caused by the dilution of CH4 with CO2 in biogas. As a result, the indicated cycle work of the engine fueled with compressed biogas containing 80% CH4 presents only 72% value of that fueled with gasoline RON92. It confirms the observation of Jawurek et al. The authors observed that the engine operates smoothly on gases containing up to 23% CO2, slightly noisily at 31% CO2 and harshly at 42% CO2. Maximum power output was 17% lower with CH4 than with petrol. Increased CO2 content of the gas led to further losses, with a 45% loss (compared with petrol) at 41% CO2 [19].

            The presence of carbon dioxide in the biogas reduces the burning velocity which ultimately affects the performance of the engine. According to Bari [20], engine power is lower when compared with that obtained in a diesel engine fumigated with natural gas, while Neyloff found out that HC, NOx, and CO emissions from a CFR engine were reduced [21]. Though the quantity of fuel admitted can be increased to ensure approximately the same thermal loading [22], the indicated power output and cyclic variation generally deteriorate with the increased proportion of carbon dioxide mixed with the methane.
            In the following section, we compare the indicated cylinder pressure given by the simulation model and the experiment at different speed regimes in order to identify the turbulent burning velocity coefficient. Compressed biogas contains 85% CH4, advance spark timing angle of the engine is fixed at 27°, equivalence ratio f = 1at full throttle opening.Using each experimental result, we adjust the computational model’s pressure diagram to match the experimental data.
            The comparison of indicated pressures given by experiment and by simulation model at an engine revolution speed of 3000 rpm and 3620 rpm is shown in Figure 11a and 11b with three turbulent burning velocity coefficient ff of 1.2; 1.3 and 1.5. The results showed that with the turbulent burning coefficient ff = 1.3, the indicated pressures given by simulation are close to the experimental data. This coefficient is also consistent with the case of n = 4070 rpm (Fig. 11c). As the engine revolution speed increases to 5360 rpm, maximum indicated pressure decreases rapidly. If using the same coefficient ff = 1.3 as the above cases, the maximum indicated cylinder pressure given by the computational model is higher than the experimental results by approximately 10%, as shown in Fig. 11d.

4. Conclusions
The resultsof this study allow us to draw the following conclusions:
(1) The turbulent burning velocity coefficient of biogas-air mixture in the combustion chamber of a Honda Wave motorcycle engine fueled with compressed biogas containing 85% CH4 is approximately 1.3.
(2) The indicated cycle work of a 110cc Honda Wave motorcycle engine reduces 28% when switching from gasoline RON92 to compressed biogas containing 80% CH4.
(3) For a given equivalence ratio, the indicated cycle work of the engine increases almost linearly with CH4 composition in the biogas fuel.
 
References
[1]. www.dongcobiogas.com/en/
[2]. Bui Van Ga, Tran Van Nam, Tran Thanh Hai Tung: LPG Motorcycles. ICAT 2002, PROCEEDINGS International conference on automotive technology, paper 031, pp. 1-6. Science and Technics publishing house
[3]. Bui Van Ga, Tran van Nam, Tran Thanh Hai Tung: Motorcycle fueled by compressed biogas. The 2009 International Forum on Strategic Technologies IFOST2009, Section Renewable Energy and Energy Conservation, pp. 17-24, HoChiMinh City, October 21-23, 2009
[4]. Ga Bui Van, Tung Tran Thanh Hai and Dong Nguyen Van: Simulation and experimental studies of performance of 110cc motorcycle engine running on biogas. The 4" AUN/SEED-Net Regional Conference in Mechanical and Aerospace Technology. HoChiMinh City, January 10-11, 2012, pp. 182-190
[5]. K. K. Kuo: Principles of Combustion. John Wiley & Sons, Inc, New York, 1986
[6]. G. R. Inger: Scaling Nonequilibrium-Reacting Flows. The Legacy of Gerhard Damkohler. J. Spacecraft Rockets, 38(2):185-190, 2001.
[7]. O. L. Gulder: Turbulent premixed flame propagation models for different combustion regimes. Proc. Combust. Inst., 23:743-750, 1990
[8]. N. Peters: The turbulent burning velocity for large-scale and small scale turbulence. J. Fluid Mech., 384:107-132, 1999
[9]. S. R. Turns: An Introduction to Combustion: Concepts and Applications. McGraw Hill Companies, Inc., U.S.A., 2nd edition, 2000.
[10]. O. L. Gulder and G. J. Smallwood: Do turbulent premixed flame fronts in spark ignition engines behave like passive surfaces? SAE Transactions - Journal of Engines, 109-3:1823-1832, 2001
[11]. H. Kobayashi: Experimental study of high-pressure turbulent premixed flames. Exp. Therm. Fluid Sci., 26:375-387, 2002
[12]. F. Halter, C. Chauveau, and I. Gokalp: Investigation on the flamelet inner structure of turbulent premixed flames. Combust. Sci. Tech., 180:713-728, 2008.
[13]. M. Elia, M. Ulinski, M. Metghalchi: Laminar Burning Velocity of Methane-Air-Diluent Mixtures. Journal of Engineering for Gas Turbines and Power JANUARY 2001, Vol. 123, pp 190-196
[14]. B. Galmiche, F. Halter, F. Foucher, P. Dagaut: Effects of Dilution on Laminar Burning Velocity of Premixed Methane/Air Flames. Energy Fuels 2011, 25, 948-954
[15]. R. Stone, A. Clarke, B Beckwith: Correlations for the Laminar-Burning Velocity of Methane/Diluent/Air Mixtures Obtained in Free-Fall Experiments. Combustion and Flame 114:546–555 (1998)
[16]. Cracknell (2010): High pressure laminar burning velocity measurements and modeling of methane and n-butane. Combustion Theory and Modeling, 14:4, 519-540
[17]. Bui Van Ga, Tran Van Nam, Tran Thanh Hai Tung, Nguyen Van Dong: Simulation of effects of operation parameters to combustion process of SI engine fueled with biogas. Vietnam Review of Mechanics, No 01, pp. 4-9 (2011) (in Vietnamese)
[18]. Bui Van Ga: Patent No 6643 “3-valve system for gas fuel supplying to LPG/gasoline motorcycle". National Intellectual Property Office, 2007 (in Vietnamese)
[19]. H. H. Jawurek, N. W. Lane and C. J. Rallis: Biogas/Petrol Dual Fuelling of Sl Engine for Rural Third World Use. Biomass13 (1987) 87-103.
[20]. Bari, S.: Effect of carbon dioxide on the performance of biogas/diesel dual-fuel engine. Renewable Energy, 1996, 9, 1007–1010.
[21]. Neyeloff, S. and Gunkel, W.: Performance of a CFR engine burning simulated anaerobic digester’s gas. ASAE Publication, 1981, 2, 324–329.
[22]. Karim, G. A. and Wierzba, I.: Methane–carbon dioxide mixtures as a fuel. SAE Special Publications, 1992, No. 927, No 921557, pp. 81–91
 
 
 

BUỒNG CHÁY PHÙ HỢP VỚI ĐỘNG CƠ BIOGAS

BUỒNG CHÁY PHÙ HỢP VỚI ĐỘNG CƠ BIOGAS ĐÁNH LỬA CƯỠNG BỨC
GATEC

            Bài báo trình bày kết quả tính toán mô phỏng ảnh hưởng của dạng buồng cháy và tỉ số nén đến tính năng công tác của động cơ biogas đánh lửa cưỡng bức được cải tạo từ động cơ Diesel có buồng cháy nguyên thủy dạng omega.

1. Giới thiệu

           Động cơ Diesel có thể chuyển sang chạy bằng biogas bằng 2 cách: động cơ nhiên liệu kép (dual fuel) và động cơ đánh lửa cưỡng bức. Động cơ nhiên liệu kép hoạt động theo nguyên lý động cơ hỗn hợp hòa trộn trước nhưng được đánh lửa bằng tia diesel phun mồi thay cho bougie. Động cơ biogas đánh lửa cưỡng bức được cải tạo từ động cơ Diesel, phương án cải tạo này có thể tận dụng được tỉ số nén cao của động cơ diesel nguyên thủy để cải thiện hiệu suất nhiệt nhờ biogas có tính chống kích nổ tốt hơn nhiên liệu truyền thống. Mặt khác tốc độ cháy của hỗn hợp biogas-không khí thấp hơn nhiên liệu lỏng nên rất phù hợp với thiết kế của động cơ diesel. Bài báo này phân tích việc lựa chọn dạng buồng cháy và tỉ số nén phù hợp khi cải tạo động cơ Diesel thành động cơ biogas đánh lửa cưỡng bức.

2. Thiết lập mô hình tính toán

          Nghiên cứu được thực hiện trên động cơ Diesel ZH1115 do Trung Quốc chế tạo. Động cơ có đường kính xi lanh D=115mm, hành trình piston S=115mm, tỉ số nén e=17 đạt công suất cực đại 24HP ở tốc độ định mức 2200 vòng/phút. Đây là chủng loại động cơ dùng khá phổ biến hiện nay ở nước ta.
          Trong nghiên cứu này chúng ta sẽ so sánh tính năng của động cơ biogas khi sử dụng lại buồng cháy xoáy lốc có sẵn của động cơ diesel và khi cải tạo buồng cháy này thành buồng cháy phẳng đơn giản. Trong trường hợp thứ nhất, piston nguyên thủy của động cơ được tiện bỏ phần đỉnh một lớp để giảm tỷ số nén. Phần lõm dạng omega của buồng cháy không thay đổi. Trong trường hợp thứ hai, phần lõm của piston được hàn đắp sau đó tiện đỉnh piston một lớp để đạt các tỉ số nén cần thiết.

            Xác lập không gian tính toán đối với 2 dạng buồng cháy nói trên, chia lưới và đặt điều kiện biên cho bài toán được thực hiện trong phần mềm GAMBIT (hình 1 và hình 2). Áp dụng Dynamic Mesh cho phép chúng ta cài đặt các thông số kết cấu động cơ trước khi thực hiện việc tính toán bằng phần mềm động lực học thủy khí FLUENT [2].

3. Diễn biến quá trình cháy

                                                                                 

Hình 3: Trường tốc độ ở vị trí 345°của môi chất công tác trong động cơ biogas đánh lửa cưỡng bức có buồng cháy Omega (a) và buồng cháy phẳng (b) (e=11,63; n=1500 vòng/phút; js=50°; f=1,08; nhiên liệu chứa 70% thể tích CH4)

            Kết quả tính toán quá trình cháy động cơ biogas đánh lửa cưỡng bức cho thấy màng lửa có dạng chỏm cầu, lan dần từ vị trí đánh lửa đến khu vực xa nhất của buồng cháy. Do trục buồng cháy omega lệch so với trục xi lanh nên cuối quá trình cháy vẫn còn một bộ phận hỗn hợp ở khu vực xa trục buồng cháy chưa cháy hết. Tuy nhiên do vận động xoáy lốc mạnh của hỗn hợp trong buồng cháy nên màng lửa lan tràn rất nhanh.
            Hình 3 thể hiện trường tốc độ dòng khí dưới dạng vector trong buồng cháy trong quá trình cháy. Chúng ta thấy trong trường hợp buồng cháy omega, vận động xoáy lốc của dòng khí mạnh hơn rất nhiều so với trường hợp buồng cháy phẳng. Khi tính toán buồng cháy phẳng, mặc dầu vị trí đặt nến đánh lửa đã giả định đặt trên đỉnh buồng cháy đối xứng hoàn toàn nhưng do buồng cháy không xoáy lốc nên quá trình cháy diễn ra chậm. Cuối quá trình cháy, một bộ phận hỗn hợp phía culasse vẫn không cháy hết.
            Kết quả bước đầu này cho thấy nên tận dụng khả năng xoáy lốc có sẵn trong động cơ diesel nguyên thủy để tăng tốc độ cháy khi chuyển sang chạy bằng biogas.

4. Ảnh hưởng của dạng buồng cháy đến tính năng động cơ

 

Hình 4: Ảnh hưởng của dạng buồng cháy đến biến thiên nồng độ O2 (a) và CH4 (b) trong quá trình cháy của động cơ biogas đánh lửa cưỡng bức (e=11,63; n=2200 vòng/phút; js=50°; f=1,08; nhiên liệu chứa 70% thể tích CH4)

            Tốc độ tiêu thụ hỗn hợp thể hiện qua tốc độ giảm nồng độ O2 và CH4 trong quá trình cháy. Hình 4a và b giới thiệu biến thiên nồng độ của chất oxy hóa và nhiên liệu khi động cơ sử dụng buồng cháy omega và buồng cháy phẳng có tỉ số nén 11,63 chạy ở tốc độ 2200 vòng/phút với biogas chứa 70% thể tích CH4. Đường cong càng dốc thì tốc độ tiêu thụ hỗn hợp càng cao. Kết quả này cho thấy tốc độ tiêu thụ hỗn hợp của động cơ có buồng cháy omega cao hơn đáng kể so với động cơ có buồng cháy phẳng. Điều này dẫn đến tốc độ tỏa nhiệt trong buồng cháy omega cao làm cho nhiệt độ cực đại của môi chất trong buồng cháy này lớn hơn nhiệt độ của chúng trong buồng cháy phẳng  (hình 5). Ngược lại, trong trường hợp buồng cháy phẳng do màng lửa lan tràn với tốc độ thấp nên quá trình cháy tiếp tục diễn ra trong quá trình dãn nở khiến cho nhiệt độ khí thải tăng so với trường hợp buồng cháy omega. Trong điều kiện vận hành nêu trên với hỗn hợp có độ đậm đặc f=1,08, nhiệt độ cực đại của môi chất trong trường hợp buồng cháy omega lớn hơn nhiệt độ cực đại trong trường hợp buồng cháy phẳng khoảng 350K nhưng nhiệt độ khí thải thấp hơn khoảng 100K.

Hình 5: Ảnh hưởng của dạng buồng cháy đến biến thiên nhiệt độ trong quá trình cháy của động cơ biogas đánh lửa cưỡng bức           

         Do tốc độ tỏa nhiệt và nhiệt độ tăng cao nên áp suất cực đại trong trường hợp buồng cháy omega cao hơn đáng kể so với áp suất cực đại trong trường hợp buồng cháy phẳng (hình 6a). Cùng điều kiện vận hành với biogas chứa 70% thể tích CH4 ở tốc độ động cơ 2200 vòng/phút, áp suất chỉ thị cực đại trong trường hợp động cơ có buồng cháy omega tăng gần 20 bar so với trường hợp động có có buồng cháy phẳng. Điều này dẫn đến áp suất trên đường dãn nở của động cơ buồng cháy omega cao hơn áp suất tương ứng của động cơ buồng cháy phẳng trên đồ thị công (hình 6b). Công chỉ thị chu trình được tính trên diện tích đồ thị công trong 2 trường hợp: Wiomega=1235J, Wiphẳng=956J. Như vậy công chỉ thị chu trình giảm khoảng 22% khi chuyển buồng cháy omega sang buồng cháy phẳng.

Hình 6: Ảnh hưởng của dạng buồng cháy đến đồ thị áp suất chỉ thị (a) và đồ thị công chỉ thị (b) của động cơ biogas đánh lửa cưỡng bức (e=11,63; n=2200 vòng/phút; js=50°; f=1,08; nhiên liệu chứa 70% thể tích CH4)

  5. Ảnh hưởng của tỉ số nén

            Hình 7 giới thiệu biến thiên nồng độ CH4 trong quá trình cháy của động cơ có buồng cháy omega sử dụng tỉ số nén 9,11; 12,82 và 16,10. Chúng ta thấy tỉ số nén động cơ càng nhỏ thì đường cong biến thiên nồng độ CH4 càng dốc nghĩa là tốc độ lan tràn màng lửa càng cao [3].  

            Ảnh hưởng rõ rệt nhất của tỉ số nén động cơ đến tính năng công tác của nó thể hiện qua đồ thị áp suất chỉ thị. Các hình 8 a, b giới thiệu biến thiên áp suất chỉ thị của động cơ có tỉ số nén 9,11 và 16,10 với góc đánh lửa sớm 20°, 30°, 40°. Kết quả cho thấy ở giá trị góc đánh lửa sớm cho trước, khi tăng tỉ số nén, áp suất chỉ thị cực đại của động cơ tăng. Tuy nhiên công chỉ thị của động cơ không tăng tỉ lệ với áp suất chỉ thị cực đại hay tỉ số nén. Khi tỉ số nén tăng cao thì phần công nén cũng tăng theo. Do đó nếu phần tăng công dãn nở không bù đắp được giá trị tăng của công nén thì công chỉ thị chu trình của động cơ lại giảm. Vì vậy việc tăng tỉ số nén vượt qua một giá trị ngưỡng tối ưu không có lợi về công suất cũng như tuổi thọ của động cơ.

            Hình 9a giới thiệu biến thiên công chỉ thị chu trình theo tỉ số nén động cơ ứng với các giá trị góc đánh lửa sớm khác nhau khi động cơ chạy ở tốc độ định mức 2200 vòng/phút. Hình 9b giới thiệu kết quả tương tự khi động cơ chạy ở tốc độ 1500 vòng/phút. Kết quả này cho thấy khi góc đánh lửa sớm bé, công chỉ thị chu trình của động cơ tăng gần như tuyến tính với tỉ số nén. Ở góc đánh lửa sớm 40° và động cơ chạy với tốc độ 2200 vòng/phút thì công chỉ thị đạt giá trị cực đại ở tỉ số nén khoảng 12. Để đạt được giá trị công chỉ thị cực đại này khi động cơ có góc đánh lửa sớm 30° thì tỉ số nén của động cơ phải có giá trị khoảng 16,5. Khi động cơ chạy ở tốc độ 1500 vòng/phút, công chỉ thị của động cơ đạt được ở giá trị tỉ số nén khoảng 14-16 (hình 9b).
            Khi tăng tỉ số nén của động cơ thì tổn thất ma sát tăng đồng thời công cần thiết để nén môi chất trong xi lanh cũng tăng. Những yếu tố này làm giảm công có ích của động cơ. Do đó mặc dù khi tăng tỉ số nén, hiệu suất nhiệt động cơ có được cải thiện nhưng nếu sự cải thiện này không đáng kể so với tổn thất công vừa nêu thì hiệu quả công tác của động cơ bị giảm. Do đó khi chuyển động cơ Diesel thành động cơ biogas đánh lửa cưỡng bức chúng ta cần lựa chọn tỉ số nén phù hợp nhất. Kết quả tính toán mô phỏng này cho thấy trong trường hợp động cơ ZH1115 chạy bằng biogas ở tốc độ định mức 2200 vòng/phút và góc đánh lửa sớm 40°, ta nên chọn tỉ số nén động cơ trong khoảng từ 11,5 đến 12,5.

 

5. Kết luận

       Từ các kết quả tính toán mô phỏng trên đây chúng ta có thể rút ra các kết luận sau đây:
1.    Khi chuyển động cơ diesel thành động cơ đánh lửa cưỡng bức chạy bằng biogas ta nên giữ các điều kiện tạo xoáy lốc trong buồng cháy để nâng cao hiệu quả công tác.
2.    Công chỉ thị của động cơ ZH1115 khi sử dụng buồng cháy phẳng nhỏ hơn công chỉ thị của nó khi sử dụng buồng cháy omega khoảng 22% khi chạy bằng biogas ở tốc độ định mức 2200 vòng/phút
3.    Ở tốc độ định mức với góc đánh lửa sớm 40°, tỉ số nén tối ưu của động cơ ZH1115 khi chuyển thành động cơ biogas đánh lửa cưỡng bức nằm trong khoảng từ 11,5 đến 12,5 

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