A Computational Investigation of Double Injection Strategy for Ethanol

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A Computational Investigation of Double Injection Strategy for Ethanol PPCI Ragavendra Prasad Panakarajupally, and Gaurav Mittal Energy Fuels, Just Accepted Manuscript • DOI: 10.1021/acs.energyfuels.7b00382 • Publication Date (Web): 31 Aug 2017 Downloaded from http://pubs.acs.org on September 4, 2017

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A Computational Investigation of Double Injection Strategy on Ethanol PPCI Ragavendra P Panakarajupally*1, Gaurav Mittal2 1

Department of Mechanical Engineering, University of Akron, Akron, OH 44325, USA 2 Department of Mechanical Engineering, Graphic Era University, Dehradun, India

Corresponding author email: [email protected]

ABSTRACT Ethanol PPCI (Partially Premixed Compression Ignition) was computationally investigated to understand the effect of injection strategy on efficiency, emissions, and noise for a heavy-duty engine having displacement of 2123 cm3 and a compression ratio of 17.3 at medium load. CFD modelling with detailed chemistry was done using the CONVERGE CFD package for a single sector from the IVC (intake valve closure) to EVO (exhaust valve opening). A double injection strategy was used and the Start of First Injection (SOFI), Start of Second Injection (SOSI) and the % Mass Injected in the First Injection (MIFI) were varied while keeping all other parameters constant. Results show that the variation in SOFI has a non-monotonic effect on the combustion phasing and the peak pressure rise rate, whereas SOSI can be used to control the combustion phasing. In addition, variation in the MIFI is an effective way to control combustion phasing and the peak rate of pressure rise. By an optimum choice of SOFI, SOSI and MIFI, it is possible to attain high efficiency and low peak pressure rise rate and emissions. The role of the delayed combustion in the cooler squish zone in moderating the heat release rate is also highlighted. KEYWORDS: Ethanol, PPCI, Injection strategy

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1. INTRODUCTION Compression Ignition (CI) engines are more efficient than Spark Ignition (SI) engines because of higher compression ratio and lack of throttling1. Unfortunately, conventional CI engines using diesel as a fuel, although efficient, are a source of Particulate Matter (PM) and NOx emissions, which have serious effects on human health and environment. The concern about these emissions has led to the implementation of stringent emission regulations, which have necessitated the development of advanced engine concepts that are clean and efficient. Especially, in the last decade, significant research efforts have been expended towards practical realization of advanced engine concepts such as HCCI (Homogeneous Charge Compression Ignition), which offers advantages of high efficiency and ultra-low NOx and soot emissions. However, control of combustion timing and burn rate have been problematic due to the lack of direct means to initiate combustion and power densities are lower for naturally aspirated HCCI engines2, due to highly diluted or fuel lean operation. At high loads, HCCI becomes especially problematic due to excessive pressure rise rates. Turbocharging or supercharging has to be applied to overcome low power density.3,4 Efforts have been made to extend the load limits and achieve better combustion control by resorting to HCCI-like combustion through charge stratification as in PPCI (Partially Premixed Compression Ignition), reactivity stratification as in RCCI (Reactivity Controlled Compression Ignition)5, and use of an external ignition source, such as a spark plug or laser

6-8

. All these strategies have been shown to increase the load limits and

improve coefficient of variance (COV) of the indicated mean effective pressure (IMEP). These strategies share the common denominator of dilute, high pressure combustion relative to the traditional CI engines. This combustion regime is different from the traditional SI and CI engines and is referred to as the advanced combustion regime or low temperature combustion.


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In the recent years, the PPCI concept has been studied by many researchers because of its potential for providing combustion control while achieving low soot and NOx emissions without compromising the engine efficiency.9,10 The single fuel PPCI and the dual fuel PPCI (also called RCCI) are promising concepts that can potentially allow controlled combustion with the desired rate of heat release and the pressure rise rate.9 The dual fuel PPCI allows an extra control knob by allowing for a change in the reactivity of the in-cylinder mixture, but is relatively more complicated due to the requirement of two fuels on board.9 The pioneering work in the PPCI was done by Kalghatgi et al.11 In order to find the best fuel for the PPCI, Kalghatgi et al.11 compared RON 95 gasoline with a diesel fuel in a CI engine and studied the effect of change in EGR levels, load and start of injection timings in a single cylinder heavy duty engine. The results showed that gasoline fuel resulted in lower levels of NOx and smoke than diesel at all operating conditions. Due to its higher resistance to autoignition and consequent increase in the ignition delay, gasoline allowed longer time for fuel and air to mix and achieve lower NOx and smoke. Use of ethanol, a higher octane number and an oxygenated fuel, in place of gasoline led to further reduction in soot in the PPCI mode.12 In the recent years, extensive work on the PPCI using ethanol has been done at the Lund University.10, 12-15 Manente et al.12,13 compared gasoline, ethanol and diesel in the PPCI mode and showed that the best fuel for a CI engine has to have high octane number. They also argued that when high octane number fuels are used at high load in a CI engine, the start of combustion should not be completely separated from the end of injection for mitigating the heat release rate. A complete separation of the start of combustion from the end of injection may result in high pressure oscillations after combustion. Manente et al.14 further studied the behavior of ethanol and seven other fuels in the boiling point range of gasoline with an octane number spanning from 3 ACS Paragon Plus Environment

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69 to 99 in the PPCI mode. In order to allow high load operation and achieve sufficient mixing, a lower compression ratio of 14.3 was used and the maximum pressure rise rate below 15 bar/CAD was achieved with most of the fuels.14 Kaiadi et al.15 performed extensive experimental investigations on a single-cylinder heavyduty engine to investigate a suitable injection strategy for PPCI combustion fueled with ethanol. The number of injections for each cycle, timing of the injections and the ratio between different injection pulses was varied one at a time and the combustion behavior was investigated at medium and low loads. The operating range of the engine showed to be very limited with the single injection strategy. The results indicated that the double injection strategy offered good combustion controllability and combustion performance and should be preferred for the ethanol PPCI. Kaiadi et al.10 further performed a sensitivity analysis to understand the effect of lambda, EGR rate, injection pressure and inlet temperature on ethanol PPCI in terms of controllability, stability, emissions and efficiency. The results showed that the combustion is not sensitive to the changes in the injection pressure but the adjustments in lambda, EGR and inlet temperature should be controlled carefully as there are clear limitations for each of these parameters.10 In contrast to the experiments, simulations allow relatively easy exploration of the effects of various geometrical and operating parameters on the advanced engine concepts. For instance, Kokjohn et al.9 presented a comparison of the sample experimental data for diesel, ethanol and dual fuel PPCI with CFD predictions with the aim of realizing accurate predictive capabilities for the PPCI combustion. Although various experimental investigations with the PPCI strategy have demonstrated that low NOx and soot levels can be achieved, the issues of the sensitivity of the combustion process, control of combustion phasing, and high peak pressure rise rate limit the use of this strategy. Since the PPCI combustion process is known to be sensitive to the temperature – 4 ACS Paragon Plus Environment

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equivalence ratio stratification, it might be possible to arrive at a controllable strategy by computationally exploring the effect of geometrical and operating parameters as well as the fuel effects. As the first step in our computational investigations in this direction, ethanol PPCI is studied to understand the effect of the injection strategy on efficiency, emissions, and noise. The CI engine selected for this study is a heavy-duty engine having a displacement of 2.123 L and a compression ratio of 17.3, as used by the various researchers from the Lund University.10, 12-15 The CFD modelling with detailed chemistry is done using the CONVERGE CFD package16. The injection strategy used in this study is a double injection strategy and the start of the first and the second injection as well as the % mass injected in each injection is varied to understand their effects and arrive at an optimum injection strategy. Some of the parametric variations studied here (presented later) are adapted from the work of Kaiadi et al.l5; however our motivation is to gain deeper insights into the ethanol PPCI and initiate a framework for further investigations of the geometrical, operating parameters and fuel effects. In the following, the computational details, validation and results are presented sequentially.

2. COMPUTATIONAL MODELLING The computational modelling is performed by using the CONVERGE CFD package16. It is a multipurpose engine simulation package that eliminates the creation of the manual mesh and provides flexibility in generating orthogonal structured mesh based on simple grid generation parameters. The chemical kinetic mechanism for ethanol was taken from Mittal et al.17, which was based on the mechanism by Metcalfe et al.18 who presented a C1–C2 sub-mechanism validated over a large range of experimental conditions for both hydrocarbon and oxygenated species. Specifically, the mechanism has been validated for ethanol against ignition delay data


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from shock tubes and rapid compression machines (800-1400 K, 1-40 atm), speciation data from jet stirred reactor and flow reactor up to 12 bar as well as flame speed and speciation data. 17 The NOx sub-mechanism consisting of additional species and reactions taken from Sivaramakrishan et al.19 was added to the primary ethanol mechanism. Soot formation was predicted using the Hiroyasu phenomenological soot model, which uses acetylene as the soot inception species.20 The kinetic model17 had been shown to perform very well against the acetylene concentration targets from jet stirred reactor, flow reactor and flame speciation. In addition, the ignition delay as well as acetylene concentration predictions from the merged mechanism were confirmed to remain unchanged in comparison to the original ethanol mechanism.17 The finite volume method was used in the CFD calculations and the relation between the gaseous and liquid phase was described by the Eulerian Lagrangian method 16,21. The RNG − turbulence model was used in the simulations and the Kelvin Helmholtz – Rayleigh Taylor (KH-RT) models were used to model the primary and secondary spray breakup. The droplet collision was modelled by using the O’Rourke Collison model

22

, which includes

collision and coalescence. Droplet drag was described using the Spherical drop drag model and evaporation using the Frossling drop evaporation model. The Drop/Wall interactions were modelled by using the wall film model.23 The boundary heat transfer was modelled by giving constant temperatures to the physical boundaries and the law of the wall model. The temperatures considered for the cylinder (490 K), piston (600 K) and the cylinder head (580 K) were based on the accepted practices in the published literature.21,24,25 The engine geometry considered was a single cylinder heavy-duty Scania D13 engine used by Kaiadi et al.15 The detailed engine specifications are given in Table 1. In order to reduce the computational time, the simulations were run for a single sector from the IVC (Inlet valve 6 ACS Paragon Plus Environment

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Closing) to the EVO (Exhaust Valve Opening). A realistic flow field for the sector at the IVC was initialized as follows. The initial axial velocity in the cylinder dropped linearly from the piston speed at the piston face to zero at the cylinder head. The other components of the velocity field at the IVC were specified based on the swirl ratio.26 In addition, the turbulent kinetic energy at the IVC was specified based on the values suggested by the CONVERGE technical support from the full cycle and complete geometry simulations on engine having similar compression ratio. The sector grid for the engine was taken as 45° since the injector consisted of 8 orifices. The engine sector geometry is shown in Figure 1. Simulations were conducted for the sector mesh with a base mesh size of 1.4 mm with an adaptive time stepping. At the piston, the cylinder head and near the injector nozzle, a fixed embed scale of 2 was used, resulting in a grid refinement factor of 4 (22) in each direction. An AMR (Adaptive Mesh Refinement) scale of 2 was used to refine the grid automatically based on the fluctuating velocity and temperature conditions. The grid independence was ascertained by conducting sample simulations on coarser and finer grids and a sample case is shown in Figure 2. Further, the grid used is in agreement with other RANS based modeling using CONVERGE 21, 27. All the simulations were conducted for an EGR of 50 % at 1300 rpm at a medium load of ≈ 12 bar IMEP. The molar composition at the IVC was O2/N2/CO2/H2O = 0.1478/ 0.7388/ 0.0703/ 0.0431. The common conditions for all simulations are shown in Table 2. Table 1. Engine Specifications Displacement Volume Bore

2123 cm3 130 mm

Stroke

160 mm

Connecting Rod length

255 mm


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40o aBDC (after bottom dead center)

IVC (inlet valve closing)

50o bBDC (before bottom dead center)

EVO (exhaust valve opening) CR (compression ratio)

17.3:1

Orifices

8

Orifice diameter

0.19 mm

Umbrella angle

148 deg Table 2. Operating Conditions Temperature at IVC (K) 400 Pressure at IVC (bar) 2.5 Injection Pressure (bar) 1500 Fuel Injected (mg/cycle) 164

The simulated cases presented below differ only in the timing for the Start of First Injection (SOFI), Start of Second Injection (SOSI) and the % Mass Injected in the First Injection (MIFI). For each scenario, only one parameter was changed at a time. The operating parameters for simulations are given in Tables 3-5. For the cases studying the effect of the SOFI variation, SOFI was varied from -60o to -20o aTDC while keeping SOSI constant at -5o aTDC and MIFI at 52%. The effect of the SOSI variation was studied by varying SOSI from -10o to -2o aTDC while keeping the SOFI constant at -50o aTDC. Finally, the effect of the MIFI variation was studied by varying MIFI from 60% to 90% while keeping SOFI and SOSI at -50o and -10o aTDC, respectively.

SOFI (CAD aTDC) -60 -50 -40 -30 -20

Table 3. Operating Parameters for SOFI variation FI duration SOSI MIFI (%) SI Duration (CAD) (CAD) (CAD aTDC) 9.5 52 -5 8.7 9.5 52 -5 8.7 9.5 52 -5 8.7 9.5 52 -5 8.7 9.5 52 -5 8.7

Table 4. Operating Parameters for SOSI variation 8 ACS Paragon Plus Environment

MISI (%) 48 48 48 48 48

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SOFI (CAD aTDC) -50 -50 -50 -50 -50

FI duration (CAD) 9.5 9.5 9.5 9.5 9.5

MIFI (%) 52 52 52 52 52

SOSI (CAD aTDC) -10 -8 -6 -4 -2

SI duration (CAD)

MISI (%)

8.7 8.7 8.7 8.7 8.7

48 48 48 48 48

Table 5. Operating Parameters for MIFI variation SOFI (CAD aTDC) -50 -50 -50 -50

FI duration (CAD) 10.9 12.8 14.6 16.5

MIFI (%) 60 70 80 90

SOSI (CAD aTDC) -10 -10 -10 -10

SI duration (CAD)

MISI (%)

7.3 5.5 3.5 1.82

40 30 20 10

3. RESULTS AND DISCUSSION Validation The physical and chemical models described above are based on the accepted practices in the literature. The computational model was further validated by comparing predicted pressure profiles with experiments from Kaiadi et al.15 Sample comparison for heat release rate for a double injection case is shown in Figure 3. The agreement in terms of the combustion phasing and duration is pretty good, especially considering the fact that the heat release in PPCI is sensitive to the temperature – equivalence ratio stratification and requires adequate modeling of the injection, mixing and chemical processes. The discrepancy in the rate of heat release could be due to the kinetic model, initial or boundary conditions or various sub-models. Although differences are noted, the simulations show a reasonable performance in replicating the overall features of the heat release curve. The focus of the present study is to understand the effect of


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double injection on engine performance. The results for the variation in the SOFI, SOSI and MIFI are presented next.

Effect of SOFI variation As mentioned previously, SOFI was changed from -60o to -20o aTDC while SOSI was kept constant at -5o aTDC. The simulated pressure curves, heat release rates, peak pressure rise rates (PPRR) and CA50 are shown in Figure 4. The combustion phasing, CA50, is defined as the crank angle at which 50 % of the heat release has happened. The peak pressure rise rate is an indication of the noise from the combustion. Shahlari et al.28 have shown that the PPRR is not always perfectly correlated with the combustion generated noise, and provided an algorithm for deducing noise from the pressure data over the 720 CAD cycle. In this work, due to the lack of the pressure data over the entire cycle, we have reported the PPRR as a metric of the combustion noise. Generally, PPRR above 20 bar/CAD is considered high and impractical for engines. The heat release curve for SOFI of -60o aTDC shows that combustion occurs with high rate of heat release. The cases for SOFI -50o and -40o aTDC are distinct from the previous case as two spikes are observed in the heat release rate curves. When SOFI is retarded to -20o aTDC, combustion occurs close to the TDC with a higher rate of heat release and a long tail in the heat release curve. The effect of change in SOFI on the combustion phasing is non-monotonic and consistent with the experimental results of Kaiadi et al.15 However, valuable insights into the details of autoignition can be obtained from the temperature and equivalence ratio distributions as discussed below. The equivalence ratio for a multicomponent mixture is defined as = (2 # +

#

)/#

where # , # and # are the numbers of carbon, hydrogen and oxygen atoms, respectively. 10 ACS Paragon Plus Environment

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Based on this definition, the intake mixture has an equivalence ratio of 0.377 even prior to the injection of ethanol, owing to the presence of and attributed to the EGR. Alternatively, a progress equivalence ratio (also referred to as the reaction ratio) can be defined as = # (2 +

#

# )/ , where the calculation of the carbon, hydrogen and oxygen atoms

is based on excluding and . The progress equivalence ratio, , tracks the progress of the reaction, whereas the equivalence ratio, , remains unchanged due to the reaction and characterizes the mixing of the fuel inside the combustion chamber. In order to understand the details of the combustion process, the temperature, , and distributions for SOFI -60o aTDC are shown in Figure 5. Each row corresponds to a specific CAD. The temperature distributions are shown in the plane of the injector as well as in a plane at 22.5 from the plane of the injector. Higher reactivity is observed in the later plane, and is, therefore, considered better suited for presenting the onset and the progress of autoignition. All subsequent distributions are shown for the plane at 22.5 from the injector. In addition, a plot of the maximum instantaneous temperature and the local at the location of the maximum instantaneous temperature is shown in Figure 6. From Figure 5 at the TDC, it is observed that the region with the higher equivalence ratios has lower temperature and vice versa, partially due to the cooling produced by the vaporization of the injected ethanol. The first injection at -60o aTDC impinges on the cylinder wall and the second injection at -5o aTDC occurs inside the bowl. Ignition begins in the region of high temperature even though the equivalence ratio happens to be lower there. This is also supported by Figure 6 wherein the maximum instantaneous temperature is noted to occur at low equivalence ratios, and gradually shifts to the higher equivalence ratios as combustion progresses. This observation is consistent with Kokjohn et al.9, who reported that the significant 11 ACS Paragon Plus Environment

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charge cooling during the injection event caused the richest regions of the chamber to be the coldest and subsequently the ignition occurred in the leaner region near the center of the combustion chamber. An inspection of the and distributions at 4o and 6o aTDC reveals that heat release happens rapidly in a large region of the bowl as well as the squish zone, where drops to near zero at 6o aTDC. The heat release rate curve also shows most of the heat release manifesting during 4o to 6o aTDC. In the fuel rich regions inside the bowl, increases due to combustion as available oxygen is consumed to form and , and some unburned hydrocarbons remain which burn relatively gradually. In summary, the case with SOFI of -60o aTDC is characterized by an HCCI like heat release behavior. The distributions for SOFI -20o aTDC are shown in Figure 7. In this case, both injections occur inside the bowl, resulting in relatively higher equivalence ratios in the bowl and negligible diffusion of ethanol to the squish zone. Ignition begins in the region above the bowl as noted from the temperature and reaction ratio fields at the TDC. The mixture in the bowl burns rapidly during 2o to 4o CAD aTDC, leading to high rate of heat release. Subsequently, the unburned hydrocarbons in the fuel rich region of the bowl, where is large, burn slowly and lead to the long tail in the heat release rate curve. This also leads to lower combustion efficiency (shown later). The distributions for SOFI -40o aTDC are shown in Figure 8. The distributions for this case compare with the SOFI -60o aTDC case in the following manner. Both have similar maximum temperature at the TDC. However, SOFI -40o aTDC leads to lower temperature and higher equivalence ratios in the squish zone and relatively lower equivalence ratios inside the bowl. This results in relatively lower reactivity in the squish zone in comparison to the SOFI -60o aTDC case. The differences in the mixing time available for the two cases alter the equivalence


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ratio - temperature stratification and influence the onset and progress of combustion. Ignition in the bowl is noted at 4o aTDC and major heat release takes place during 6o to 8o aTDC. In contrast to the SOFI -60o aTDC, where heat release in the bowl and the squish zone occurred concurrently, the squish zone burns later (during 8o to 10o aTDC) for SOFI -40o aTDC and leads to the second peak (or plateau) in the heat release rate curve. In summary, whereas the combustion for SOFI -60o aTDC is more HCCI like and rapid, that with SOFI -40o aTDC is slower with greater stratification of the ignition delays owing to the lower reactivity in the squish zone that moderates the rate of heat release as well as the peak pressure rise rate. However, the PPRR still remains above 20 bar/CAD with SOFI -40o aTDC as shown in Figure 4, and is too high for a practical engine. The indicated thermal efficiency, combustion efficiency and emissions for SOFI variation are shown in Figure 9. The combustion efficiency assesses the conversion of the fuel chemical energy into the thermal energy and the indicated thermal efficiency is a measure of the conversion of the thermal energy into mechanical work. The combustion efficiency is particularly low for SOFI -20o aTDC due to inadequate fuel/air mixing. Peak combustion and overall efficiency occurs in the range of SOFI -50o to -40o aTDC. Soot emissions increase rapidly as SOFI is retarded to -20o aTDC as there is less time for mixing. In addition, slightly higher NOx is obtained when combustion occurs closer to the TDC.

Effect of SOSI variation From the simulations on the SOFI variation, it was noticed that higher efficiency with lower NOX and soot emissions and lower PPRR are realized for SOFI of -50o to -40o aTDC. Therefore, in the following simulations, SOFI was fixed at -50o aTDC and the effect of SOSI was


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investigated by varying it from -10o to -2o aTDC. The simulated pressure curves, heat release rates, PPRR and CA50 are shown in Figure 10. The peak pressure decreases as SOSI is retarded. It is also noted that the heat release curve is divided into two parts due to the reasons explained previously with reference to Figure 8; namely the first peak is primarily due to the heat release in the bowl and the second due to the subsequent heat release in the squish zone. The heat release curves move down as SOSI is retarded. CA50 increases as SOSI moves from -10o to -2o aTDC and the PPRR is also noted to increase moderately with retardation in SOSI. Form these curves it is evident that the second injection could be used to control the combustion phasing. The combustion efficiency, shown in Figure 11, decreases slightly as SOSI is retarded. Soot and NOx emissions, also shown in Figure 11, are very low and NOx decreases slightly as SOSI is retarded. Effect of MIFI variation For this study, the SOFI and SOSI times were taken as -50o and -10o aTDC, respectively. SOSI of -10o aTDC was chosen despite of slightly higher NOx due to the lower PPRR and the highest combustion efficiency. The mass of fuel injected in the first injection was increased from 50 % to 90 %. The pressure profiles, heat release rate curves and the PPRR are shown in Figure 12. It is noteworthy that the MIFI variation has no effect on the start of combustion. At 60% MIFI, two distinct peaks in the heat release rate curves are noted. As MIFI is increased, the first peak decreases and the second increases. The second peak is not due to the mixing controlled combustion but attributed to the thermal stratification. In fact, the mass injected in the second injection burns first. A comparison of the temperature and equivalence ratio distributions for MIFI 90% and 60% are shown in Figure 13. At the TDC, MIFI 90% yields much higher equivalence ratios and lower temperatures in the squish zone than MIFI 60%. However, similar local equivalence 14 ACS Paragon Plus Environment

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ratio/temperature distributions are obtained for both cases in a portion of the bowl region where ignition occurs. This can be discerned from the temperature and equivalence ratio distributions at 2o and 4o aTDC. Subsequently, rapid heat release occurs in the bowl for MIFI 60% followed by heat release in the squish zone, leading to the two distinct peaks in the heat release rate curve. For MIFI 90%, the first peak due to the combustion of the mass injected inside the bowl during the second injection leads to less heat release, and a slower and long drawn burning occurs in the squish zone during the expansion stroke. Consistent with the heat release rate curve, CA 50 shown in Figure 12 shows rapid increase with an increase in MIFI. As shown in Figure 14, both combustion efficiency and the indicated thermal efficiency decrease with an increase in MIFI due to late combustion. NOx decreases with an increase in MIFI and soot emissions are negligible. It is evident that the variation in MIFI provides an effective way of controlling the combustion phasing and the PPRR. The results also indicate that it is possible to get low PPRR and high efficiency even when combustion takes place after the end of both injections, such as with MIFI 60-70% at the considered load. This is due to the thermal and equivalence ratio stratification due to the direct injection, which leads to a staged auto-ignition that progress from the leaner-hotter region to the richer-cooler regions. In contrast, Manente et al.12,13 recommended a non-separation of the end of injection and the start of combustion to achieve acceptable pressure rise rate; this might be necessary, though, at higher loads.

4. CONCLUDING REMARKS CFD simulations were performed in order to investigate the effect of the injection strategy on the performance of ethanol PPCI at medium load. Specifically, the effect of various injection 15 ACS Paragon Plus Environment

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parameters on combustion and indicated efficiency, PPRR and emissions were assessed. The simulations were conducted by varying the Start of First Injection (SOFI), Start of Second Injection (SOSI) and the % Mass Injected in the First Injection (MIFI). In the SOFI simulations, the start of first injection was varied from -60o to -20o aTDC while keeping the SOSI constant at -5o aTDC. Consistent with the experimental work15, the results indicate that the SOFI timing cannot be used to control combustion. The PPRR were higher for SOFI -60o aTDC and also for SOFI retarded to more than -40o aTDC. The optimum SOFI was found to be around -50o to -40o aTDC that resulted in relatively lower PPRR, higher efficiency and lower NOx and soot emissions. In the simulations studying the effect of the SOSI variation, SOSI was varied from -10o to -2o aTDC while keeping SOFI constant at -50o aTDC. The results show that as SOSI was retarded, combustion phasing (CA50) was delayed and the PPRR increased moderately. The combustion efficiency deteriorated slightly as the SOSI was retarded. These simulations indicated that the variation in SOSI can be used to control the combustion phasing. Lastly, % MIFI was changed from 50% to 90%. The results showed that as MIFI is increased, combustion duration increases significantly even though the start of combustion remains unchanged. The effect is an increase in CA50, and decrease in the PPRR. Variation of the MIFI provides an effective means for controlling the combustion phasing and the PPRR. The combustion efficiency, however, deteriorates with an increase in MIFI. At the conditions investigated here, 60 to 70% MIFI provided high efficiency with low emissions and low peak pressure rise rate. The simulations also demonstrate that acceptable heat release rate for ethanol PPCI is possible by the thermal stratification achieved from fuel injections even when ignition occurs after the end of fuel injections. This is due to the staged autoignition that progresses from the leaner-hotter regions to


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Energy & Fuels

the richer-cooler regions. The delayed combustion in the cooler squish zone plays a vital role in moderating the heat release rate. Further work is required to understand the effects of the intake temperature, pressure and EGR levels on the sensitivity of the combustion process as well as exploration at high load and use of a second reactive fuel, as in dual fuel PPCI, to propose a robust practical strategy that might provide good combustion control.


Energy & Fuels

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

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REFERENCES (1) Heywood, J. B. An introduction to IC engines. 1988. (2) Hiltner, J.; Fiveland, S. B.; Agama, R.; Willi, M. System efficiency issues for natural gas fueled HCCI engines in heavy-duty stationary applications. SAE Paper. 2002, 2002-010417. (3) Olsson, J. O.; Tunestal, P.; Haraldsson, G.; Johansson, B. A turbo charged dual fuel HCCI engine. SAE Paper. 2001, 2001-01-1896. (4) Olsson, J. O.; Tunestal, P.; Johansson, B. Boosting for high load HCCI. SAE Paper. 2004, 2004-01-0940. (5) Christensen, M.; Johansson, B. Supercharged homogeneous charge compression ignition (HCCI) with exhaust gas recirculation. SAE Paper. 2000, 2000- 01-1835. (6) Christensen, M.; Johansson, B.; Amneus, P.; Mauss, F. Supercharged homogeneous charge compression ignition. SAE Paper. 1988, 980787. (7) Morimoto, S.; Kawabata, Y.; Sakurai, T.; Amano, T. Operating characteristics of a natural gas-fired homogeneous charge compression ignition engine (performance improvements using EGR). SAE Paper. 2001, 2001-01-1034. (8) Reitz, R. D.; Duraisamy, G. Review of high efficiency and clean reactivity controlled compression ignition (RCCI) combustion in internal combustion engines. Progress in Energy and Combustion Science. 2015, 46, 12-71. (9) Kokjohn, S. L.; Splitter, D. A.; Hanson, R. M.; Reitz, R. D.; Manente, V.; Johansson, B. Modeling charge preparation and combustion in diesel fuel, ethanol, and dual-fuel PCCI engines. ILASS-Americas 22nd Annual Conference on Liquid Atomization and Spray Systems. 2010. (10) Kaiadi, M.; Johansson, B.; Lundgren, M.; Gaynor, J. A. Sensitivity analysis study on ethanol partially premixed combustion. SAE Int. J. Engines. 2013, 6(1), 120-131. (11) Kalghatgi, G.; Risberg, P.; Ångström, H. Partially pre-mixed auto-ignition of gasoline to attain low smoke and low NOx at high load in a compression ignition engine and comparison with a diesel fuel. SAE Technical Paper. 2007, 2007-01-0006. (12) Manente, V.; Johansson, B.; Tunestal, P. Partially premixed combustion at high load using gasoline and ethanol, a comparison with diesel. SAE Technical Paper. 2009, 200901-0944


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Energy & Fuels

(13) Manente, V.; Johansson, B.; Tunestal, P. Characterization of partially premixed combustion with ethanol: EGR sweeps, low and maximum loads. J. Eng. Gas Turbines Power. 2009, 132(8), 1-7. (14) Manente, V.; Tunestal, P.; Johansson, B.; Cannella, W. Effects of ethanol and different type of gasoline fuels on partially premixed combustion from low to high load. SAE Technical Paper. 2010, 2010-01-0871. (15) Kaiadi, M.; Johansson, B.; Lundgren, M.; Gaynor, J. A. Experimental investigation on different injection strategies for ethanol partially premixed combustion. SAE Technical Paper. 2013, 2013-01-0281. (16) Converge theory manual 2.1.0. 2013. (17) Mittal, G.; Burke, S. M.; Davies, V. A.; Parajuli, B.; Metcalfe, W. K.; Curran, H. J. Autoignition of ethanol in a rapid compression machine. Combustion and Flame. 2014, 161 (5), 1164–1171. (18) Metcalfe, W. K.; Burke, S. M.; Ahmed, S. S.; Curran, H. J. A hierarchical and comparative kinetic modeling study of c1 − c2 hydrocarbon and oxygenated fuels. Int. J. Chem. Kinetics. 2013, 45, 638–675. (19) Sivaramakrishnan, R.; Brezinsky, K.; Dayma, G.; Dagaut, P. High pressure effects on the mutual sensitization of the oxidation of NO and CH4–C2H6 blends. Physical Chemistry Chemical Physics. 2007, 9, 4230-4244. (20) Hiroyasu, H.; Kadota, T. Models for combustion and formation of nitric oxide and soot in DI diesel engines. SAE Technical Paper. 1976, 760129. (21) Huang, M.; Gowdagiri, S.; Cesari, X. M.; Oehlschlaeger, M. A. Diesel engine cfd simulations: influence of fuel variability on ignition delay. Fuel. 2016, 181, 170-177. (22) Rourke, O. P.; Collective drop effects on vaporizing liquid sprays. Ph.D. Thesis, Princeton University, 1981. (23) Rourke, O. P.; Amsden, A. A spray/wall interaction submodel for the KIVA-3 wall film model. SAE Technical Paper. 2000, 2000-01-0271. (24) Emami, S.; Jafarmadar, S. Multidimensional modeling of the effect of fuel injection pressure on temperature distribution in cylinder of a turbocharged DI diesel engine. Propul Power Res 2013. 2013, 2, 162–175. (25) Mohammadi, A.; Yaghoubi, M. Estimation of instantaneous local heat transfer coefficient in spark-ignition engines. Int J Therm Sci. 2010, 49, 1309–17.


Energy & Fuels

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(26) Amsden, A. KIVA-3V: A block structured KIVA program for engines with vertical or canted valves. Los Alamos National Laboratory Report No. LA-13313-MS, 1997. (27) Senecal, P. K.; Richards, K. J.; Pomraning, E.; Yang, T.; Dai, M. Z.; McDavid, R. M. A new parallel cut-cell cartesian CFD code for rapid grid generation applied to in-cylinder diesel engine simulations. SAE Technical Paper. 2007, 2007-01-0159. (28) Shahlari, A.; Hocking, C.; Kurtz, E.; Ghandhi, J. Comparison of compression ignition engine noise metrics in low-temperature combustion regimes. SAE Int. J. Engines. 2013, 6, 541-552.


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Figure 1. Engine sector Geometry.

20 1.4 mm 1.8 mm 1 mm

Pressure (MPa)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

Energy & Fuels

15

10

5 -20

-10

0 10 Crank Angle (deg)

20

Figure 2. Sample grid independence study with different base grid sizes. SOFI = -50o aTDC, SOSI = -5o aTDC, MIFI = 52%


Energy & Fuels

700 Experiment

600 HRR (J/CAD)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

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Simulation

500 400 300 200 100 0 -20

-10


20

30

Figure 3. Simulated and experimental (Kaiadi et al.15) heat release rate profiles. SOFI = -35o aTDC, SOSI = 0o aTDC, MIFI = 52%


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Pressure (MPa)

20 FI -60 CAD aTDC FI -50 CAD aTDC FI -40 CAD aTDC FI -30 CAD aTDC FI -20 CAD aTDC

15 10 5 0 -40

-20


40

(a)

HRR (J/CAD)

3000 FI -60 CAD aTDC FI -50 CAD aTDC FI -40 CAD aTDC FI -30 CAD aTDC FI -20 CAD aTDC

2000

1000

0 -10

-5

0

5

10

15

Crank Angle (deg) 10

60

8

50 40

6

30

4

20

2

CA50

PPRR (bar/CAD)

(b)

CA50 (CAD aTDC)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

Energy & Fuels

10

PPRR

0

0

-60

-50

-40

-30

-20

SOFI (CAD aTDC)

(c) Figure 4. SOFI variation (a) Pressure curves (b) Heat release rate (c) Peak pressure rise rate and CA50 23 ACS Paragon Plus Environment

Energy & Fuels

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

Temperature (injector plane)

Temperature (22.5 to the injector plane)

(22.5 to the injector plane)

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(22.5 to the injector plane)

Figure 5. Temperature, equivalence ratio and reaction ratio distributions. SOFI = -60o aTDC, SOSI = -5o aTDC, MIFI = 52%


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1.1

2500 Temperature

2200

1

Equivalence Ratio

0.9 0.8

1900

0.7

1600

0.6

1300

0.5

1000

Equivalence Ratio

Temperature (K)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

Energy & Fuels

0.4

0

2

4

6

8

Crank Angle (deg)

Figure 6. Maximum instantaneous temperature and the local equivalence ratio at the location of the maximum instantaneous temperature. SOFI = -60o aTDC, SOSI = -5o aTDC, MIFI = 52%


Energy & Fuels

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

Temperature (K)

Equivalence Ratio,

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Reaction Ratio,

Figure 7. Temperature, equivalence ratio and reaction ratio distributions. SOFI = -20o aTDC, SOSI = -5o aTDC, MIFI = 52%


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Energy & Fuels

Temperature (K)

Equivalence Ratio

Reaction Ratio

Figure 8. Temperature, equivalence ratio and reaction ratio distribution. SOFI = -40o aTDC, SOSI = -5o aTDC, MIFI = 52%


Energy & Fuels

100 98

55 96

50

94 92

45

Indicated Thermal

90

Combustion

40

Combustion efficiency

Indicated Thermal efficiency

60

88

-60

-50

-40

-30

-20

SOFI (CAD aTDC)

(a)

0.5

NOx

1.2

Soot

1

0.4

0.8

0.3

0.6

0.2

0.4

0.1

0.2

0

Soot (g/kgf)

0.6

NOx (g/kgf)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

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0

-60

-50

-40

-30

-20

SOFI (CAD aTDC)

(b) Figure 9. SOFI variation (a) Combustion and Indicated thermal efficiency (b) emissions


Page 29 of 33

Pressure (MPa)

20 SI -10 CAD aTDC SI -8 CAD aTDC SI -6 CAD aTDC SI -4 CAD aTDC SI -2 CAD aTDC

15 10 5 0 -30

-20

-10 0 10 Crank Angle (deg)

20

30

15

20

(a) 2000 SI -10 CAD aTDC SI -8 CAD aTDC SI -6 CAD aTDC SI -4 CAD aTDC SI -2 CAD aTDC

HRR (J/CAD)

1500 1000 500 0 -10

-5

0

5

10

Crank Angle (deg)

(b) 10

35

25 20

8 15 10

PPRR (bar/CAD)

30

CA50 (CAD aTDC)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

Energy & Fuels

CA50

5

PPRR

6

0

-10

-8

-6

-4

-2

SOSI (CAD aTDC)

(c) Figure 10. SOSI variation (a) Pressure curves (b) Heat release rate (c) Peak pressure rise rate and CA50


Energy & Fuels

100

55 50

99

45

Indicated Thermal Combustion

40



60

98

-10

-8

-6

-4

-2

SOSI (CAD aTDC)

(a) 0.4

0.1

NOx

0.2

0.05

Soot

0.1 0

Soot (g/kgf)

0.3 NOx (g/kgf)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

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0

-10

-8

-6

-4

-2

SOSI (CAD aTDC)

(b) Figure 11. SOFI variation (a) Combustion and Indicated thermal efficiency (b) emissions


Page 31 of 33

20 MIFI 60%

Pressure (MPa)

MIFI 70%

15

MIFI 80% MIFI 90%

10 5 0 -40

-20


40

(a) 1200 MIFI 60%

HRR (J/CAD)

MIFI 70%

800

MIFI 80% MIFI 90%

400

0 -5

0

5

10

15

20

25

30

Crank Angle (deg)

(b) 20

15

12 10

8 5

4

PPRR (bar/CAD)

20

16 CA50 (CAD aTDC)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

Energy & Fuels

CA50 PPRR

0

0

60

70

80

90

MIFI (%)

(c) Figure 12. MIFI variation (a) Pressure curves (b) Heat release rate (c) Peak pressure rise rate and CA50


Energy & Fuels

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

MIFI = 90% Temperature (K)

Page 32 of 33

MIFI = 60%

Equivalence Ratio

Temperature (K)

Equivalence Ratio

Figure 13. Temperature and equivalence ratio distribution. SOFI = -50o aTDC, SOSI = -10o aTDC, MIFI 90% (left panels) and 60% (right panels)


Page 33 of 33

100 98

55

96

50 94

45

92

Indicated Thermal Combustion

40



60

90

60

70

80

90

MIFI (%)

(a) 0.4

0.1 NOx

0.3

Soot

0.2

0.05

0.1 0

Soot (g/kgf)

NOx (g/kgf)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

Energy & Fuels

0

60

70

80

90

MIFI (%)

(b) Figure 14. MIFI variation (a) Combustion and Indicated thermal efficiency (b) emissions


A Computational Investigation of Double Injection Strategy for Ethanol

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