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Article

Effects of Water Injection Strategies on Oxy-Fuel Combustion Characteristics of a Dual-Injection Spark Ignition Engine

1
School of Computer Science and Technology, University of Bedfordshire, Luton LU1 3JU, UK
2
State Key Laboratory of Engines, Tianjin University, Tianjin 300072, China
3
UMR 9189–CRIStAL—Centre de Recherche en Informatique Signal et Automatique de Lille, CNRS, Centrale Lille, Université de Lille, F-59000 Lille, France
4
Junia, Smart Systems and Energies, Université de Lille, F-59000 Lille, France
5
School of Engineering, University of Lincoln, Lincoln LN6 7TS, UK
*
Authors to whom correspondence should be addressed.
Academic Editors: Jamie W.G. Turner and Constantine D. Rakopoulos
Energies 2021, 14(17), 5287; https://doi.org/10.3390/en14175287
Received: 11 July 2021 / Revised: 17 August 2021 / Accepted: 23 August 2021 / Published: 26 August 2021

Abstract

Currently, global warming has been a serious issue, which is closely related to anthropogenic emission of Greenhouse Gas (GHG) in the atmosphere, particularly Carbon Dioxide (CO2). To help achieve carbon neutrality by decreasing CO2 emissions, Oxy-Fuel Combustion (OFC) technology is becoming a hot topic in recent years. However, few findings have been reported about the implementation of OFC in dual-injection Spark Ignition (SI) engines. This work numerically explores the effects of Water Injection (WI) strategies on OFC characteristics in a practical dual-injection engine, including GDI (only using GDI), P50-G50 (50% PFI and 50% GDI) and PFI (only using PFI). The findings will help build a conceptual and theoretical foundation for the implementation of OFC technology in dual-injection SI engines, as well as exploring a solution to increase engine efficiency. The results show that compared to Conventional Air Combustion (CAC), there is a significant increase in BSFC under OFC. Ignition delay (θF) is significantly prolonged, and the spark timing is obviously advanced. Combustion duration (θC) of PFI is a bit shorter than that of GDI and P50-G50. There is a small benefit to BSFC under a low water-fuel mass ratio (Rwf). However, with the further increase of Rwf from 0.2 to 0.9, there is an increment of 4.29%, 3.6% and 3.77% in BSFC for GDI, P50-G50 and PFI, respectively. As WI timing (tWI) postpones to around −30 °CA under the conditions of Rwf ≥ 0.8, BSFC has a sharp decrease of more than 6 g/kWh, and this decline is more evident under GDI injection strategy. The variation of maximum cylinder pressure (Pmax) and combustion phasing is less affected by WI temperature (TWI) compared to the effects of Rwf or tWI. BSFC just has a small decline with the increase of TWI from 298 K to 368 K regardless of the injection strategy. Consequently, appropriate WI strategies are beneficial to OFC combustion in a dual-injection SI engine, but the benefit in fuel economy is limited.
Keywords: oxy-fuel combustion (OFC); water injection (WI) strategies; dual-injection spark ignition (SI) engine; simulation oxy-fuel combustion (OFC); water injection (WI) strategies; dual-injection spark ignition (SI) engine; simulation

1. Introduction

The global warming crisis has apparently been a severe problem over the last few decades. Furthermore, the 10 warmest years on record have occurred since 1998, with the 4 warmest years occurring since 2014 [1]. Hence, carbon neutrality has become a desperate need to resolve the climate crisis, as it is closely linked to anthropogenic Greenhouse Gas (GHG) emissions, particularly Carbon Dioxide (CO2) [2,3,4]. To help achieve carbon neutrality by decreasing CO2 emissions, a variety of new technologies have been developed for passenger vehicles, such as hybrid electric, plug-in hybrid electric, lithium-ion battery electric, solar-powered, fuel-cell electric, etc. [5,6,7].
However, due to the relatively high cost and low torque output, these technologies are rarely utilised in non-road mobile machineries, such as vessels and boats. In order to realise zero CO2 emissions in conventional Internal Combustion Engine (ICE) fuelled with fossil fuels, Oxy-Fuel Combustion (OFC) technology with Carbon Capture and Storage (CCS) has been an attractive alternative in recent decades. Yaverbaum initially proposed OFC technology, which was afterward widely applied in power stations [8,9,10,11,12]. Equation (1) and Table 1 present the OFC reaction process and the gas physicochemical properties, respectively. The discrepancies of physicochemical properties between CO2 and nitrogen have significant impacts on OFC and Conventional Air Combustion (CAC). The chemical reaction indicates that OFC can take advantage of eliminating nitrogen during the combustion process. The comparison of gas properties shows that CO2 heat capacity is much higher than nitrogen on a mole basis. Besides, both the thermal diffusivity and oxygen diffusion of CO2 are much lower than those of nitrogen.
C x H y + x + y 4 O 2 x CO 2 + y 2 H 2 O
Regarding the study of OFC technology in Spark Ignition (SI) engines, Bilger [15] initially proposed a new approach named Internal Combustion Rankine Cycle (ICRC) in 1999. In recent years, Wu et al. [16,17,18,19,20,21,22,23,24,25] made a significant contribution to ICRC implementation in Port Fuel Injection (PFI) SI engines fuelled with propane under 40% to 70% oxygen volume fraction. Water Injection (WI) has been a proven method to control oxy-fuel combustion process of ICRC engines based on some findings [17,18,19,20,21]. Engine performance could be directly affected by the timing, duration, pressure, mass and temperature of WI. Thermal efficiency and indicated work could be enhanced by an appropriate rate of EGR matches with ignition timing and increased WI temperature. Li et al. [26] initially investigated the impacts of intake charge on OFC in a dual-injection SI engine fuelled with gasoline under GDI, P-G and PFI three injection strategies.
Based on this literature survey about OFC implementation in SI engines, most of them focused on PFI engines fuelled propane. However, few studies have been reported about the implementation of OFC in GDI engines or dual-injection SI engines. However, Gasoline Direct Injection (GDI) technology has become the mainstream of SI engines in academia and industry [27,28,29,30,31,32,33,34,35]. In 2005, an advanced technology named dual-injection was commercially applied to SI engines by Toyota [36]. Furthermore, with the advantages of flexible fuel injection, dual-injection SI engines have been a hot topic for researchers in recent years [37,38,39,40,41,42]. The thermal efficiency, gaseous and particulate emissions have been investigated in this type of engine by using gasoline plus various fuels such as methanol, ethanol and 2,5-dimethylfuran.
Therefore, the investigation of the effects of WI strategies on combustion characteristics of OFC in a dual-injection SI engine fuelled with gasoline was firstly reported in this article. It will provide valuable insight into the effects of WI strategies on OFC characteristics in a practical dual-injection engine, including GDI (only using GDI), P50-G50 (50% PFI and 50% GDI) and PFI (only using PFI). The findings will help build the conceptual and theoretical foundation for the OFC implementation in dual-injection SI engines, as well as exploring a solution to increase engine efficiency. This work belongs to the ‘RIVER’ project, which was proposed to develop a novel non-carbon boat with an environment-friendly powertrain system. The overview of the designed novel ‘RIVER’ system can be depicted in Figure 1, which mainly includes the subsystems of oxygen feeding, intake organisation, CO2 capture and storage. Pure oxygen is mixed with recirculated gas (CO2), then the mixture will enter into engine combustion chambers. Besides, the excess CO2 can be captured in a storage tank to achieve zero carbon emissions.

2. Research Approach and Numerical Method

2.1. Research Approach and Key Parameters

In order to keep this numerical study well organised, the procedure can be summarised in Figure 2. First, the validation of this model is performed based on the experimental data of 2000 rpm-10 bar BMEP under CAC mode. Second, as nitrogen is replaced by CO2, engine operation converts into the OFC mode with nitrogen replaced by CO2, a comparison in performance between CAC and OFC is presented. Finally, the effects of WI mass, WI timing and WI temperature on OFC characteristics are explored and analysed in turn.
In order to increase the efficiency of water utilisation for optimising OFC process, it is usually injected directly into the engine chamber by a separated injector [17,18,19,20,21]. Figure 3 shows the schematic of two representatives.
The system presented in Figure 3a is adopted in this numerical study, which advantage is compact size, simple structure and control measure. A container coupled with a pump is utilised to deliver water to a rail. Afterward, water will be directly injected into combustion chambers with the assistance of an air booster. The booster provides a high WI pressure to enhance water droplets’ breakup, diffusion and vaporisation.
The other strategy in Figure 3b does not include storage for water supply. Instead, the water vapour in the exhaust gas can be condensed and separated, then heated up prior to being recirculated and injected into the combustion chambers. The advantage of this system is that overall thermal efficiency can be improved by the benefit of water vaporisation inside the cylinder chamber and the waste heat recovery from the exhaust gas.
In order to make the parameters more visually, WI mass is parameterised by water-fuel mass ratio ( R w f ) in Equation (2).
R w f = τ w τ f
Here, τ w (mg/cycle) and τ f (mg/cycle) denotes the injection quantity of water and fuel per engine cycle, respectively.
WI timing and WI temperature is abbreviated with t W I and T W I , respectively. Brake Specific Fuel Consumption (BSFC) is introduced to evaluate fuel consumption. φ C A 10 , φ C A 50 and φ C A 90 represents the Crank Angle (CA) where 10%, 50% and 90% of the total heat has been released, respectively. Ignition delay ( θ F ) represents the period between spark timing and φ C A 10 . Combustion duration ( θ C ) denotes the period between φ C A 10 and φ C A 90 .
During the study, engine operation is kept at a typical mid-high load of 2000 revolutions per minute (rpm) and 10 bar Brake Mean Effective Pressure (BMEP). The spark timings are fixed under the conditions of R w f = 0. WI pressure is maintained at 60 bar to ensure the injected water is in the liquid state. The water boiling point at 60 bar is around 549 K, which is much higher than the WI temperature (298 K to 368 K) in this study. Besides, to make the investigation process more explicit, the throttle opening angle, fuel injection timings, intake temperature and Oxygen Mass Fraction (OMF) remain constant with CAC mode throughout this study.

2.2. Engine and Testbed

The model of numerical study is set up based on a practical turbocharged dual-injection SI engine fuelled with gasoline. The engine specifications and testbed overview are presented in Table 2 and Figure 4, respectively. The used fuel of this study is commercial gasoline, which physical and chemical parameters are shown in Table 3. Regarding the data collection and analysis of the test, cylinder pressure is measured and analysed by spark-plug type piezo-electric sensors (AVL-GH13Z), a crank sensor (Kistler 2614CK1), a charge amplifier (Kistler 5018A) and a combustion analyser (AVL 641). The pressure data are recorded and averaged by 200 consecutive cycles to ensure the accuracy of measurements. A direct current electrical dynamometer measures and controls the engine’s speed and torque in real-time. A programmable Electronic Control Unit (ECU) is used to adjust the ignition, fuel injection and other engine operating parameters. The air-fuel ratio is kept constant at stoichiometric ratio (lambda = 1 ± 0.01). The uncertainties of the measured parameters in the test are shown in Table 4 by the root mean square method.

2.3. Model Description

The model used in this study is set up based on GT-Power software, which is an industry-standard simulation software and widely used both industry and academia to predict combustion performance of SI engines [43,44,45,46].
In this model, the engine configuration settings are according to its detailed specifications, such as basic geometry of pipes, engine displacement, cylinder bore, connecting rod length, valves, etc. The heat transfer process in ICE can be classified into thermal conduction, thermal convection, thermal radiation and thermal vaporisation. The heat transfer submodel is set to be ‘Woschni model’, which was put forward to provide an instantaneous, spatially averaged heat transfer coefficient for the evaluation of heat transfer rate to the cylinder walls [47,48,49]. Some key correlations in the ‘Woschni model’ are given by Equations (3)–(5).
Q W = 0 c y c l e i h A i T T w i d φ
h = 110 d 0.2 P 0.8 T 0.53 C 1 c m + w c 0.8
w c = C 2 V S T 1 P 1 V 1 P P 0
Here, Q W is total heat transferred; h is heat transfer coefficient; A i is heat absorbing area of the surfaces; T is in-cylinder mean gas temperature; T w i is mean surface temperature of A i ; φ is CA; d is cylinder bore diameter; P is cylinder pressure; T is in-cylinder mean gas temperature; C 1 is a constant related to airflow velocity coefficient; c m is mean piston speed; w c is gas velocity caused by combustion; C 2 is a constant related to combustion chamber; V S is cylinder volume; P 0 is cylinder pressure when the engine is started. T 1 , P 1 and V 1 is cylinder temperature, pressure and volume at the beginning of compression, respectively.
The ‘SI turbulent flame combustion model’ is chosen and set up for the submodel of combustion and heat transfer. This submodel is commonly used to predict the in-cylinder burn rate, knocking occurrence and emissions for SI engines [43]. Moreover, the flame-wall interactions can be calculated in this submodel according to the settings of cylinder geometry, spark location, head region and piston cup. Some key formulas are presented as Equations (6)–(9).
S L = S L , 0 ( T u T r e f ) α ( p p r e f ) β = ( B m B Ø Ø Ø m 2 ) T u T r e f α p p r e f β f D
f D = 1.0 0.75 λ D E M ( 1.0 1.0 0.75 λ D E M D 7 )
α = 2.4 0.271 Ø 3.51
β = 0.357 + 0.14 Ø 2.77
Here, S L is laminar flame speed; S L , 0 is laminar flame speed under the condition of 298 K and 101.325 kPa; T u is unburned gas temperature; T r e f is 298 K; α is temperature exponent; p is pressure; p r e f is 101.325 kPa; β is pressure exponent; B m is maximum laminar speed; B Ø is laminar speed roll-off value; Ø is in-cylinder equivalence ratio; Ø m is equivalence ratio at maximum speed; f D is dilution effect; D is mass fraction of the residuals in the unburned zone; λ D E M is dilution effect multiplier.

2.4. Model Validation

Figure 5 and Figure 6 show the model validation by comparing cylinder pressure and Heat Release Rate (HRR) between experimental and simulation results at 2000 rpm-10 bar under CAC mode with GDI, P50-G50 and PFI three injection strategies. The time-step sensitivity has also been considered to assure the accuracy of the simulation results further. It can be observed that the cylinder pressure and HRR curves in experimental conditions have a good match with those of simulation. Furthermore, the magnitude and position for the peaks of curves are correctly predicted. Therefore, this model is eligible to make a reasonable prediction in this study.

3. Results and Discussion

3.1. Comparison between CAC and OFC

Figure 7 presents the comparison of HRR between CAC and OFC under 2000 rpm-10 bar. The spark timing is optimised to be the minimum advance for Maximum Brake Torque (MBT) or Knock Limited Spark Advance (KLSA) for each condition.
It can be seen that the peaks of HRR under OFC are generally about 8 degrees advanced than CAC. The peaks of HRR curves decrease from around 110 J/CA of CAC to 90 J/CA of OFC. This is mainly because the engine knocking would occur with the excessively advanced sparking timing at 2000 rpm-10 bar under CAC, and it can be easily avoided since the thermal diffusivity of CO2 is much lower than that of nitrogen.
Figure 8 and Figure 9 shows the comparison of BSFC, spark timing, θ F and θ C between CAC and OFC. It can be seen that BSFC is 317.269 g/kWh, 317.519 g/kWh and 317.524 g/kWh for GDI, P50-G50 and PFI, respectively. There is a significant increase of around 68 g/kWh compared to that of the CAC mode. This is mainly because the thermal diffusivity of CO2 is much lower than nitrogen, resulting in a great difference in flame transmission, which would increase the heat loss under OFC. Even though there are also offset effects caused by the higher thermal conductivity of CO2 than that of nitrogen.
The behaviour can be further explained by the significantly advanced spark timing and prolonged θ F under OFC compared to CAC. It is similar to some previous findings of oxy-fuel burners, which conclude that the ignition delay of a mixture of O2/CO2 is significantly higher than that of an O2/N2 mixture with a similar oxygen fraction [50,51].
Besides, there is no significant difference in θ C between OFC and CAC. In the meantime, the θ C of PFI under OFC is 25.5 degrees, which is a bit shorter than 26.8 degrees of GDI and 26 degrees of P50-G50. This can be attributed to that compared to P50-G50 and GDI, a more homogenous gas-fuel mixture under PFI strategy can help accelerate flame development.

3.2. Effects of WI Mass on OFC Performance

The effects of R w f are discussed in this section, which includes the analysis of BSFC, maximum cylinder pressure ( P m a x ), φ C A 50 , HRR and in-cylinder temperature under three injection strategies (GDI, P50-G50 and PFI). Meanwhile, in order to make the research objective clear and straightforward, t W I and T W I is fixed at −60 °CA and 298 K, respectively.
As shown in Figure 10, BSFC is quite sensitive to R w f . With the increase of R w f from 0 to 0.2, there is a small decrease in BSFC regardless of the injection strategy. The lowest BSFC is each 316.647 g/kWh, 316.497 g/kWh and 316.358 g/kWh for GDI, P50-G50 and PFI, which has a slight saving rate of 0.196%, 0.322% and 0.367% than that of R w f = 0, respectively. This can be attributed to the increased working medium and oxygen atom concentration from injected water.
However, an overall growth trend can be observed in BSFC with the further increase of R w f from 0.2 to 0.9, which is an increment of 4.29%, 3.6% and 3.77% for GDI, P50-G50 and PFI, respectively. Moreover, by further increasing R w f from 0.9 to 2.0, the trend of BSFC would become quite stable. This indicates that an excessive amount of injected water will increase BSFC. However, the deterioration is not serious due to water’s positive effect on combustion by the increased working medium and oxygen concentration.
The variations can be further explained in Figure 11 and Figure 12, which show the effects of R w f on combustion characteristics by P m a x , φ C A 50 , HRR and in-cylinder temperature. It demonstrates that the changing trends in combustion characteristics are similar for GDI, P50-G50 and PFI. With the increase of R w f from 0 to 0.6, P m a x has a reduction of around 4.5 bar and φ C A 50 is delayed by around 2.5 degrees. With further increased R w f , both P m a x and φ C A 50 remain broadly stable with a minor fluctuation. In the meantime, HRR and in-cylinder temperature are also sensitive to R w f . Under all the three injection strategies, as the R w f increases from 0 to 1.0, the peak of HRR is postponed by around 1 degree whilst the peak of in-cylinder temperature reduces about 180 K. Afterwards, HRR and in-cylinder temperature will be less affected by further increasing R w f .

3.3. Effects of WI Timing on OFC Performance

Figure 13 presents the effects of t W I on BSFC under three injection strategies (GDI, P50-G50 and PFI). Meanwhile, T W I is held constant at 298 K in this section.
On the whole, the relatively lower BSFC generally appear with low R w f . The lowest BSFC appears on the operating condition of ‘ t W I = −70 °CA, R w f = 0.1’, ‘ t W I = −70 °CA, R w f = 0.2’ and ‘ t W I = −70 °CA, R w f = 0.2’ for GDI, P50-G50 and PFI, respectively. The corresponding values are each 316.638 g/kWh, 316.383 g/kWh and 316.172 g/kWh, which provides a limited benefit of 0.199%, 0.358% and 0.426% than that of R w f = 0, respectively.
Besides, it can be observed that BSFC is less affected by t W I under the conditions of low R w f . On the conditions of R w f ≤ 0.5, the fluctuation of BSFC is generally kept in a narrow range with the delay of t W I from −100 °CA to −30 °CA under all three injection strategies.
Under the conditions of high R w f ( R w f ≥ 0.8), BSFC is strongly affected by t W I . As t W I postpones from −100 °CA to −30 °CA, BSFC has a sharp decrease of more than 6 g/kWh. Furthermore, this trend becomes more obvious under GDI injection strategy. Because the negative effects on the gas-fuel mixture from the vaporisation process of injected water are a bit stronger under GDI. The variation of BSFC can be attributed to the combustion characteristics, which are presented with P m a x , φ C A 50 , HRR and in-cylinder temperature in Figure 14 and Figure 15.
As shown in Figure 14, P m a x shows a monotonic increase with the delay of t W I from −100 °CA to −30 °CA. The decreased tendency is more apparent under the conditions of R w f = 1.0, particularly with GDI strategy. In the meantime, φ C A 50 generally shows a decreasing trend to around 5 °CA at t W I = −30 °CA, which is also helpful to improve fuel efficiency. Figure 15 shows the effects of t W I on HRR and in-cylinder temperature under GDI with R w f = 1.0. With the postpone of t W I from −100 °CA to −40 °CA, the peak of HRR is advanced by 4.5 °CA, and the peak of in-cylinder temperature is significantly increased by 247 K. With the delay of t W I , the improvement in combustion performance is mainly because the cooling effect is getting weaker by reducing the period of water vaporisation.

3.4. Effects of WI Temperature on OFC Performance

Figure 16 shows the effects of T W I on BSFC under GDI, P50-G50 and PFI injection strategies. In the meantime, t W I is fixed at −60 °CA in this section. It can be seen that BSFC is not sensitive to T W I , it generally has a small decline with the increase of T W I from 298 K to 368 K regardless of the injection strategy. The most significant decline is just around 3.5 g/kWh under R w f = 1.0. Besides, the lowest BSFC are each 316.291 g/kWh, 316.068 g/kWh and 315.884 g/kWh, providing a limited saving rate of 0.308%, 0.457% and 0.516% than that of R w f = 0, respectively.
Figure 17 presents the effects of T W I on P m a x and φ C A 50 under R w f = 0.2, 0.6 and 1.0. The variation of P m a x and φ C A 50 is less affected by T W I compared to the effects of R w f or t W I . By increasing T W I from 298 K to 368 K, P m a x just presents an increase of around 0.5 bar, and whilst φ C A 50 would be advanced by about 0.25 degrees. This is mainly because that by increasing T W I , the suppression of combustion would become slightly weaker during the process of heat absorption by the cooling effect of water. Moreover, increasing T W I would accelerate the vaporisation rate of water, leading to fewer negative effects on combustion [52].
This can also be explained by Figure 18, which is one example to show the effects of T W I on HRR and in-cylinder temperature. It shows that by increasing T W I from 308 K to 368 K, the peak of HRR and in-cylinder temperature have a slight increase of 0.47 J/CA and 19.5 K, respectively.

4. Conclusions

In order to eliminate CO2 emissions from ICE for achieving carbon neutrality and mitigating global warming, the implementation of OFC technology in a dual-injection SI engine fuelled with gasoline was numerically studied. This study systematically explores the effects of WI strategies (mass, timing, temperature) on OFC characteristics in a dual-injection SI engine. The findings offer some important insights to help understand the performance and optimisation of dual-injection SI engine under OFC mode fuelled with gasoline. The main conclusions are summarised as follows.
  • Compared to CAC mode, it can be seen that the peaks of HRR under OFC are generally about 8 degrees advanced than CAC. The peaks of HRR curves decrease from around 110 J/CA of CAC to 90 J/CA of OFC. Furthermore, there is a significant increase of around 68 g/kWh in BSFC under OFC. θ F is significantly prolonged, and the spark timing is obviously advanced. Besides, θ C of PFI is a bit shorter than that of GDI and P50-G50.
  • With the increase of R w f , P m a x and in-cylinder temperature is reduced, while φ C A 50 and the peak of HRR is postponed. There is a small benefit to BSFC under low R w f . However, with the further increase of R w f from 0.2 to 0.9, there is an increment of 4.29%, 3.6% and 3.77% in BSFC for GDI, P50-G50 and PFI, respectively.
  • On the whole, the relatively lower BSFC generally appear with low R w f . The lowest BSFC appears on the operating condition of ‘ t W I = −70 °CA, R w f = 0.1’, ‘ t W I = −70 °CA, R w f = 0.2’ and ‘ t W I = −70 °CA, R w f = 0.2’ for GDI, P50-G50 and PFI, respectively.
  • Under the conditions of R w f ≥ 0.8, BSFC is strongly affected by t W I . As t W I postpones to around −30 °CA, with the improvement in combustion performance, BSFC has a sharp decrease of more than 6 g/kWh, and this decline is more evident under GDI injection strategy.
  • The variation of P m a x and φ C A 50 is less affected by T W I compared to the effects of R w f or t W I . BSFC generally has a small decline with the increase of T W I from 298 K to 368 K regardless of the injection strategy. P m a x just presents an increase of around 0.5 bar, and whilst φ C A 50 would be advanced by about 0.25 degrees.
  • It is feasible to implement appropriate WI strategies to control OFC characteristics and performance of a dual-injection SI engine, but the benefit in fuel economy is limited.
  • The findings of this study are beneficial to establish a conceptual and theoretical foundation for the implementation of OFC technology in dual-injection SI engines, as well as increasing the efficiency of this kind of engine. In future works, the engine efficiency under the OFC mode will be expected to further improve by various optimisation research, such as the effects of different kinds of fuel injection, intake charge components, variable valve actuation, etc.

Author Contributions

X.L.: Methodology, Software, Investigation, Data curation, Writing—original draft. Y.P.: Investigation. D.L.: Project administration. T.A.: Project administration. K.-J.R.: Conceptualization. A.A.: Supervision, Project administration, Funding acquisition. R.M.: Conceptualization. Z.P.: Methodology, Supervision, Writing—review & editing. All authors have read and agreed to the published version of the manuscript.

Funding

This work is funded by European Regional Development Fund (ERDF) via Interreg North-West Europe (Project No. NWE553).

Acknowledgments

This work is financially supported by European Regional Development Fund (ERDF) via Interreg North-West Europe (Project No. NWE553).

Conflicts of Interest

The authors declare no conflict of interest.

Abbreviations

BMEPBrake Mean Effective Pressure (bar)
BSFCBrake Specific Fuel Consumption (g/kWh)
CACrank Angle (degree)
CACConventional Air Combustion
CCSCarbon Capture and Storage
CO2Carbon Dioxide
DIDirect Injection
ECUElectronic Control Unit
ERDFEuropean Regional Development Fund
GDIGasoline Direct Injection
GHGGreenhouse Gas
HRRHeat Release Rate
ICEInternal Combustion Engine
ICRCInternal Combustion Rankine Cycle
KLSAKnock Limited Spark Advance
MBTMaximum Brake Torque
OFCOxy-Fuel Combustion
OMFOxygen Mass Fraction
PFIPort Fuel Injection
P50-G5050% Port Fuel Injection and 50% Gasoline Direct Injection
rpmrevolutions per minute
SISpark Ignition

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Figure 1. Schematic of OFC with CCS in the application of ICE.
Figure 1. Schematic of OFC with CCS in the application of ICE.
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Figure 2. Flow chart of the research approach.
Figure 2. Flow chart of the research approach.
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Figure 3. Schematic of typical WI systems in IC engines.
Figure 3. Schematic of typical WI systems in IC engines.
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Figure 4. Schematic of engine testbed.
Figure 4. Schematic of engine testbed.
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Figure 5. Comparison of cylinder pressure between experimental and simulation results at 2000 rpm-10 bar.
Figure 5. Comparison of cylinder pressure between experimental and simulation results at 2000 rpm-10 bar.
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Figure 6. Comparison of HRR between experimental and simulation results at 2000 rpm-10 bar.
Figure 6. Comparison of HRR between experimental and simulation results at 2000 rpm-10 bar.
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Figure 7. Comparison of HRR between CAC and OFC.
Figure 7. Comparison of HRR between CAC and OFC.
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Figure 8. Comparison of BSFC and spark timing between CAC and OFC.
Figure 8. Comparison of BSFC and spark timing between CAC and OFC.
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Figure 9. Comparison of θ F and θ C between CAC and OFC.
Figure 9. Comparison of θ F and θ C between CAC and OFC.
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Figure 10. Effects of R w f on BSFC.
Figure 10. Effects of R w f on BSFC.
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Figure 11. Effects of R w f on P m a x and φ C A 50 .
Figure 11. Effects of R w f on P m a x and φ C A 50 .
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Figure 12. Effects of R w f on HRR and in-cylinder temperature.
Figure 12. Effects of R w f on HRR and in-cylinder temperature.
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Figure 13. Effects of t W I on BSFC.
Figure 13. Effects of t W I on BSFC.
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Figure 14. Effects of t W I on P m a x and φ C A 50 ( R w f = 0.2, 0.6, 1.0).
Figure 14. Effects of t W I on P m a x and φ C A 50 ( R w f = 0.2, 0.6, 1.0).
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Figure 15. Effects of t W I on HRR and in-cylinder temperature (GDI; R w f = 1.0).
Figure 15. Effects of t W I on HRR and in-cylinder temperature (GDI; R w f = 1.0).
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Figure 16. Effects of T W I on BSFC.
Figure 16. Effects of T W I on BSFC.
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Figure 17. Effects of T W I on P m a x and φ C A 50 ( R w f = 0.2, 0.6, 1.0).
Figure 17. Effects of T W I on P m a x and φ C A 50 ( R w f = 0.2, 0.6, 1.0).
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Figure 18. Effects of T W I on HRR and in-cylinder temperature (GDI; R w f = 1.0).
Figure 18. Effects of T W I on HRR and in-cylinder temperature (GDI; R w f = 1.0).
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Table 1. Physicochemical properties of CO2 and nitrogen at 1000 k and 0.1 MPa [13,14].
Table 1. Physicochemical properties of CO2 and nitrogen at 1000 k and 0.1 MPa [13,14].
PropertyCO2NitrogenRatio (CO2/Nitrogen)
Molecular weight44281.57
Density (kg/m3)0.53620.34131.57
Kinematic viscosity (m2/s)7.69 × 10−51.2 × 10−40.631
Specific heat capacity (kJ/kg K)1.23431.16741.06
Thermal conductivity (W/m K)7.057 × 10−26.599 × 10−21.07
Thermal diffusivity (m2/s)1.1 × 10−41.7 × 10−40.644
Mass diffusivity of O2 (m2/s)9.8 × 10−51.3 × 10−40.778
Prandtl number0.74550.70221.06
Emissivity and absorptivity>0~0-
Table 2. Engine specifications.
Table 2. Engine specifications.
ItemsContent
Engine typeFour-cylinder; Four-stroke
Bore × Stroke (mm)82.5 × 92
Displacement (L)2.0
Injection typeDual-injection (PFI plus GDI)
Intake typeTurbocharged
Compression ratio9.6:1
Rated speed (rpm)5500
Rated power (kW)160
Maximum Torque (N·m)320
Table 3. Fuel properties.
Table 3. Fuel properties.
ItemsContent
Fuel typeGasoline
Chemical formulaC5-C12
Relative molecular mass95–120
Gravimetric carbon content (%)86.4
Gravimetric hydrogen content (%)13.6
Gravimetric oxygen content (%)<0.1
Research octane number95
Density (20 °C) (kg/L)0.73
Dynamic viscosity (20 °C) (mPa·s)0.52
Kinematic viscosity (20 °C) (mm2/s)0.71
Surface tension (20 °C) (N/m)22
Boiling range (°C)30–200
Low heating value (kJ/kg)44,300
Latent heat of vaporisation (kJ/kg)370
Laminar flame speed (20 °C) (m/s)0.33
Stoichiometric air-fuel ratio14.7
Table 4. Uncertainties of measured parameters.
Table 4. Uncertainties of measured parameters.
Measured ParametersUncertainty (%)
Engine speed±0.1
BMEP±0.1
BSFC±0.2
Pressure±0.1
Crank angle±0.1
Lambda±0.3
Coolant temperature±0.4
Intercooler output temperature±0.4
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