Numerical study on ignition amelioration of a hydrogen-enriched Wankel engine under lean-burn condition

Applied Energy 255 (2019) 113800

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Numerical study on ignition amelioration of a hydrogen-enriched Wankel engine under lean-burn condition

T

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Cheng Shia, Changwei Jib, , Yunshan Gea, Shuofeng Wangb, Jianhui Baoc, Jinxin Yangb a

School of Mechanical Engineering, Beijing Institute of Technology, Beijing 100081, PR China College of Environmental and Energy Engineering, Beijing Laboratory of New Energy Vehicles and Key Lab of Regional Air Pollution Control, Beijing University of Technology, Beijing 100124, PR China c School of Microelectronics, Dalian University of Technology, Dalian 116024, PR China b

H I GH L IG H T S

simulation model of the hydrogen-enriched Wankel engine was built and validated. • ASingle-spark ignition strategies consist of split ignition, high-energy ignition. • Twin-spark ignition strategies cover asynchronous ignition and energy allocation. • The preferable arrangement of the twin spark plug was T/L = 1.5. • Optimal combustion was gained as the leading plug was ignited earlier and stronger. •

A R T I C LE I N FO

A B S T R A C T

Keywords: Wankel engine Lean-burn Hydrogen-enriched Ignition amelioration Twin-spark ignition Discharge energy

For the hydrogen-enriched spark-ignited Wankel engine, the optimization of ignition strategy is conducive to improve combustion performance and specifically effective to lessen the unburned region due to the elongated rotor chamber. In this paper, the role of the number of the ignition source, twin-spark plug location, asynchronous ignition, and energy allocation in improving lean combustion was investigated through the threedimensional computational fluid dynamics model coupling with kinetic mechanisms. The model was validated by experiment, and good agreements between measured and predicted combustion pressure and the heat release rate was obtained. Results showed that the improvements of engine combustion were limited by single-spark ignition strategies, and the twin-spark ignition configuration was capable of enhancing combustion efficiency drastically. The arrangement of the twin-spark plug determined the space for flame development, and it was favorable for the trailing plug to stand a greater offset from the minor axis of the engine. An earlier leading-spark ignition enabled flame propagation faster and occurred quenching rapidly, which contributed to higher pressureoutput and better heat-release. The higher energy of leading-spark ignition made the mixture consumption faster, combustion pressure higher, and combustion duration shorter. The optimum strategy on combustion was expressed as follows: the location of trailing-spark plug is offset from the minor axis by 20.7 mm; the spark timing and discharge energy of leading-spark plug is 325°EA and 0.03 J, respectively; and those of trailing-spark plug is 335°EA and 0.01 J. It was recommended that the leading-spark ignition was set earlier and stronger for practical operations.

1. Introduction

efficiency of the internal combustion engine is of great interest [3]. The two most common engine prototypes are rotary-piston engine and reciprocating-piston engine. Rotary-piston and reciprocating-piston are two different types of mechanisms, which have been conducted by engine designers and manufacturers to provide modern society with a viable but efficient power unit [4]. The reciprocating engine plays a dominant role on internal combustion engines as a result of the

The ever-increasing prominence of the energy dilemma and environmental pollution, has given rise to the search for solutions such as exploiting renewable energy or improving conversion efficiency [1]. For the former, hydrogen, extracted from multiple renewable routes, has wide prospect for application [2]. For the latter, enhancing

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Corresponding author. E-mail address: [email protected] (C. Ji).

https://doi.org/10.1016/j.apenergy.2019.113800 Received 3 June 2019; Received in revised form 18 August 2019; Accepted 27 August 2019 0306-2619/ © 2019 Elsevier Ltd. All rights reserved.

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Nomenclature BDC CFD CO EA EA10 EA50 IHRF L-plug MAP

MPRR PRF SI SSI TDC TKE TSI T-plug WE rpm Φ αH2 η

bottom dead center computational fluid dynamics carbon monoxide eccentric angle the eccentric angle whose accumulated heat amount at 10% of total heat amount the eccentric angle whose accumulated heat amount at 50% of total heat amount integrated heat release fraction leading-spark plug manifolds absolute pressure

maximum pressure rising rate primary reference fuel spark ignition single spark-plug ignition top dead center turbulent kinetic energy twin spark-plug ignition trailing-spark plug Wankel engines revolutions per minute equivalence ratio the volume of enriched hydrogen combustion efficiency

high-efficiency gaseous fuel, hydrogen has become an ideal substitute for non-renewable energy sources owing to its superior property and unlimited source [22]. Growing concerns with hydrogen combustion have made conventional engine fuels to be reformed by designers [23]. The characteristics of hydrogen combustion are distinct from that of gasoline in three major aspects: burning velocity, quenching distance, and flammability limit [24]. Inside the engine, the faster burning velocity of hydrogen implies that the heat release can lead to higher combustion temperature, which increase cooling losses [25]. Besides, the shorter quenching distance of hydrogen means that the flame propagates closer towards the inner wall of the cylinder [26]. To overcome these problems, WE designers have benefited from the ability of hydrogen to burn under lean conditions and introduced it into the basic fuel without framework modification [27]. The hydrogen-enriched WE are therefore suitably operating with leaner mixtures. On the one hand, lean-burn increases the amount of operating gas heated through more air suction at a constant mass of fuel charge. On the other hand, the burning speed of lean mixtures is lower, which inhibits the excessive heat release thus decreases the peak temperature in rotor chambers [28]. However, in the negative aspect, less heat release induces more exhaust losses. In this connection, the developers in SI WE have devoted to the improvement of mass burning rate under lean-burn conditions. Amelioration of ignition strategy is one of the main studied objects due to the initial flame growth, which is more significant compared with subsequent combustion phasing, especially for WE. To enhance the ignition level of SI engines under lean-burn conditions, varying strategies have been adopted in WE, such as multipoint ignition, the arrangement of the spark plug, ignition timing, discharge energy, etc. The overarching focus of related studies lies in the number and location of ignition points. An earlier modeling work by Ohzeki et al. [29] concluded that two ignition points proved to be superior to single sparkplug ignition (SSI) configuration, and a favorable position of the spark plug is beneficial for enhancing engine performance. Abraham and Bracco [30] performed a three-dimensional computational fluid

simplicity of the in-cylinder sealing method [5]. However, numerous issues on the piston-crank method are still under dispute, which restricts the improvement of this mechanism efficiency [6]. Analogously, the airtightness and durability of the rotary mechanism have limited Wankel engines (WE) to be an economical and practical power source [7]. Owing to the development of innovative technologies, the attractiveness of WE is giving a comeback in the minds of many researchers even though it had been faded out for a period of time, such as a compact design, lightweight architecture, favorable power-to-weight ratio, multi-fuel capability, easy maintenance, and reduced mechanical vibrations [8]. Recently, utilizing WE as an alternative power plant can be generally categorized into three applications: (1) Driven by the electrification tendency of automobile powertrains [9], there has been great interest in the viable use of WE as range extenders for battery electric or hybrid vehicles [10]. (2) WE are becoming increasingly popular with regard to their use as an auxiliary power supply for unmanned aerial vehicles or gliders [11]. (3) Various compressors and pumps based on the operating principle of WE must be mentioned [12]. However, these special applications have not managed the leap into production. In addition to the sealing problem, the underlying reason for this lies in the poor fuel economy and unwished noxious emissions [13]. In recent decades, an agreement has reached on the in-cylinder combustion that satisfies a term of “rapid combustion” criterion, namely the fuel-burning progress taking place more rapidly which thereby results in the efficiency enhancement directly [14]. The prominent merits of rapid combustion compared with normal burning-rate combustion are depicted as follows: Firstly, it produces a more powerful and repeatable burning, and thus permits operating with leaner mixtures without deteriorating engine operation or combustion stability [15]. Secondly, rapid combustion combined with leaner mixtures has a prospect to obtain greater pollutants control within the combustion chamber accompanied with certain improvements in fuel economy due to the reduced heat loss, decreased pumping work and declined exhaust temperature [16]. The third attractive feature lies in the increased resistance to knock whilst adopting rapid combustion, which enables the fuel economy associated with higher compression ratios to be performed [17]. Learning from existing literature, there are various technical options for accomplishing fast burning, and therein the optimizations of fuel property and ignition strategy should be noticed in terms of the spark ignition (SI) WE. Altering fuel components is of great importance to increase combustion rate via accelerating the flame spread and enlarging the frontal area of the flame. Moreover, fuels used in engines should take the energy crisis and environmental concern into consideration as well [18]. In this regard, a variety of alternative fuels have been successfully implemented for SI WE as an alone fuel or additional fuel, such as hydrogen [12], natural gas [19], methane [20], and alcohol fuels [21]. Thereinto, the relevant studies concerning the assessment of hydrogen additive in WE have special importance. Acting as an eco-friendly and

Table 1 Engine technical specifications.

2

Parameter

Value

Manufacturer Engine type Fuel supplement Generating radius Eccentricity Rotor width Compression ratio Displacement Power delivery Intake opening Intake closing Exhaust opening Exhaust closing

Hongling Electromechanical Co. Ltd. Single rotor, side-ported, air-cooled Port-fuel injection 69 mm 11 mm 40 mm 8:1 0.16 L 3.8 kW@4200 rpm −465°EA, −105°EA, 255°EA −209°EA, 151°EA, 511°EA 208°EA, 568°EA, 928°EA 250°EA, 610°EA, 970°EA

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emissions. As a consequence, multipoint ignition viably serves as a preeminent configuration, and it is better characterized by a reasonable arrangement to obtain rapid combustion for SI WE. In the aspect of ignition timing, some scholars have engaged in this concept by experiments and simulations, and only a few of them are based on multipoint ignition mode in WE. Spreitzer et al. [33] developed a CFD model for predicting combustion and flow phenomena through WE with two plugs. In their viewpoints, the change of spark timing was crucial to influence the burning duration and heat release of the engine. Fan et al. [34] illustrated that an increase of mass burning rate was originated from the increment of the tumble effect time as twospark points ignited earlier. As mentioned in Ref. [35], researchers experimentally studied the engine performance of gasoline-butanol WE, and the cycle-to-cycle variations were initially weakened and then worsened when the ignition advance was increased. The simulation was also conducted by Shi et al. [36], and it was revealed that the better flame propagation could be achieved through the pre-ignition of the trailing spark plug while igniting both spark plugs simultaneously, which would make a counter effect. To sum up, optimizing the multipoint ignition timing has a great potential to improve burning initiation and engine performance of SI WE. Another effective way to decrease the inflammation time for successful lean-burn is to generate robust flame kernels by spark discharge energy [37]. To promote rapid combustion and extend the lean-stability limit of SI engines, much effort to develop ignition systems has centered on ameliorating ignition energy in reciprocating engines. Modeling work by Yossefi et al. [38] found that compared with fuel composition, spark energy exhibited a larger influence on initial combustion, with shorter ignition lag time. As for experiments about spark energy, Srivastava et al. [39] evaluated the effect of high-energy ignition on the combustion of methane. In their studies, the improved burning stability, higher heat release and less fuel consumption were observed at higher spark energy. Jung et al. [40] analyzed the effects of spark discharge energy and in-cylinder turbulence level in a SI engine, and their results showed that the combination of these elements made it possible to operate at ultra-lean limit. Chen et al. [41] demonstrated that high spark energy tangibly reduces the cyclic variation, and combustion stability and duration could be optimized to acquire high thermal efficiency and engine performance consequently. In view of these researches, limited implementation is only based on reciprocating SI engines in terms of ignition energy, and the influence of this factor on WE running under lean-burn conditions remains to be observed. In summary, both hydrogen enrichment and ignition optimization are effective approaches to realize rapid combustion and high mass burning rate for WE. Despite the substantially positive influence, the detailed and in-depth mechanism for combustion enhancement of hydrogen-enriched WE coupling with varying ignition strategies is not fully understood, particularly under lean-burn conditions. Importantly, as literature survey implies, there seem to be limited opportunities concentrating on the synergistic effect of different ignition strategies, and perhaps no coherent investigation simultaneously involving in the arrangement, time difference and discharge energy of twin spark-plug ignition (TSI) compared with SSI configuration. Hereby, for purpose of filling the gap in this field, the objective of this research is to clarify the role of spark plug location, asynchronous ignition and discharge energy in combustion variations of port-fuel injection WE fueled with gasolinehydrogen blends, with addressing the lean-burn process by the numerical model. In current work, the simulation was conducted in threedimensional CFD models of port-fuel injection WE. After the calculation tool was verified with experimental results, synchronization acquisition through the flow phenomena and characteristic parameters were carried out, and varying ignition strategies were employed. The present paper was organized elaborately as follows: first, numerical approach and model validation was introduced with detailed descriptions of the simulated model and experimental apparatus. Then, the comparative analysis of SSI and TSI on combustion performance, followed by effects

Fig. 1. Schematic diagram of the operated engine.

Fig. 2. Meshing model of the computational domain.

Fig. 3. Geometry of the simulated combustion chamber.

Table 2 Summary of boundary conditions. Region

Type

Value

Air intake Exhaust outlet Rotor Stator Spark plug Spark electrode Inlet port Outlet port

Inflow Outflow Moving wall Fixed wall Fixed wall Fixed wall Fixed wall Fixed wall

0.035 MPa/293 K 0.1 MPa/700 K 550 K 550 K 750 K 850 K 293 K 293 K

dynamics (CFD) of the two-ignition source WE fueled with nature gas, and found that the shorter flame propagation and lower hydrocarbon emissions are achieved by two spark points per chamber. Jaber et al. [31] estimated the possible use of the placement of spark plugs in a Mazda rotary engine, and their study indicated that the increment of spark plug numbers gave rise to strengthened combustion in rotor chambers. Based on a peripheral-ported WE, Zambalov et al. [32] assessed hydrogen combustion in WE using laser ignition, and the numerical simulation had shown the dual ignition mode was a very promising option because of its high engine efficiency and low carbon 3

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(a) Schematic

(b) Photograph Fig. 4. Schematic and photograph of experimental apparatus.

2. Numerical approach

Table 3 Summary of measurement sensitivities. Parameter

Instrument

Manufacturer

Uncertainty

Engine speed Torque Cylinder pressure Gasoline mass flow rate Hydrogen volumetric flow rate Air volumetric flow rate Air-to-fuel ratio

CAC6 CAC6 6117BCD17 FC2210 D07-19BM 20N060 MEXA-730λ

Power link Power link Kistler Power link Seven star Tociel Horiba

≤ ± 1 rpm ≤ ± 0.4% F.S. ≤ ± 0.3 bar ≤0.8% F.S ≤ ± 0.02 L/min ≤ ± 0.1 L/min ≤ ± 0.007

2.1. Engine geometry and computational domain In this study, the model follows the geometry of the Hongling Electromechanical Co. Ltd. designed Z160F rotary engine, and the technical specifications of the engine as listed in Table 1. Fig. 1 plots the schematic diagram of the prototype. The heart of the engine consists of an eccentric shaft and a rotor. Such two parts are the only moving components of this engine concept. The rotor and trochoidal stator divide the housing into three independent combustion chambers. Three chambers change simultaneously in volume along with the eccentric shaft and rotor rotate, thus creating intake, compression, power and exhaust strokes. Each rotor rotation equals three revolutions of the eccentric shaft because of the kinematics. The eccentric angle (EA) is counted from the spark-ignited top dead center (TDC), as remarked in

of TSI location, asynchronous ignition and discharge energy on the early flame development and combustion characteristics were discussed step-by-step to realize preferable ignition strategy. And the gist of conclusions is finally drawn. 4

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chambers since it is acceptable for predicting the performance of target model with minimum computational cost actually. To resolve spark ignition by generating a fine grid around the spark plug, the minimum grid scale of 0.125 mm through adequate mesh refinement is specified that applies to the entire ignition phasing in the simulation of the realistic SI WE. As depicted in Fig. 2, the spacious space compensates for sectors such as dented or protruded inlet and outlet ports, possibly fuel injectors and other structural details, which does not taken into account on the description of an engine built in the domain. Fig. 3 shows the geometry of the simulated combustion chamber. The total number of the grid is being around 30,000–90,000 per chamber. 2.2. Mathematical model and boundary condition In the course of this work, the RNG k-ε model is chosen as the turbulence model, which has been validated by the PIV experiment [34] and numerical simulation on flow phenomena of WE [32]. The Wall-function heat transfer model is adopted to deal with wall heat loss and boundary layers [43]. To capture the right ignition response at the spark plug gap, a spark-energy deposition model is used to describe virtual spark as a source term and explicitly compute the generation of the initial kernel with the calculation solver, detailed of such model can be found in the literature [44]. As a selected combustion model, the SAGE model offers the possibility to assess the premixed combustion and diffusion combustion in detail [33]. The chemical kinetic model is based on the skeletal mechanism of primary reference fuel (PRF), which performs more accurate with 41 components and 124 reaction steps [45]. The Multi-zone function is implemented for facilitating the solution of chemical kinetics. A surrogate fuel of 92% (v/v) iso-octane and 8% (v/v) n-heptane is served as the representative of the commercial gasoline (#92 Sinopec, China-V). The presetting of boundary conditions imposed for the current simulation is listed in Table 2. Furthermore, the following assumptions of this model are made to feature a dedicated simulation environment. Firstly, the space of engine-built is enclosed completely by computational domain. Secondly, the running process is fully transient. Thirdly, it is supposed that the considered WE refers to the “perfect mixing” scavenging concept in which air, PRF and hydrogen are instantly mixed before they are aspirated into the combustion chamber and that the exhaust leaving the chamber has the same species as the in-charge in the intake port. Finally, the effects of the heat transfer with the environment and residual exhausts are neglected. To save computational time, Chamber III sector (Fig. 1) is chosen as a research target. The simulation range is from −105°EA (the intake port connects to Chamber III) to 568°EA (the exhaust port connects to Chamber III) while TDC is 360°EA. During all subsequent calculations,

Fig. 5. Model validation for pressure and heat release rate vs. EA. Table 4 Comparison of SSI and TSI strategy schemes. Case

A0 A1 A2 A3

L-plug (original)

T-plug

Timing

Energy

Timing

Energy

335°EA 325/335°EA 335°EA 335°EA

0.02 J 0.02/0.02 J 0.04 J 0.02 J

– – – 335°EA

– – – 0.02 J

Fig. 1. From TDC to bottom dead center (BDC), the rotor rotates 90°EA, the eccentric shaft is three times faster by 270°EA. A rotor flank reaches after three cycles of the eccentric shaft its original position. Referring to Fig. 1, the rotating moment of Chamber I is at TDC (i.e. 0°EA). In this work, a three-dimensional geometric model of the port-fuel injection SI WE is established by CATIA code. A rapid and steady grid generation can be achieved through automatic grid technology based on the orthogonal grid (i.e. cubes) using CONVERGE code, and this grid generation arithmetic is insensitive to the wall-surface quality. Moreover, it is competent in permitting the grid scale to change with run-time transiently and handling moving boundaries properly, which is necessary for the modeling of the complex motion of WE. Before the modeling research, the accuracy of the computational model has been calibrated and verified [42]. And well fitness can be observed from the grid scale of 2 mm with adequate mesh refinement for combustion

Fig. 6. Contours of flame propagation and streamlines distribution at varying EAs. 5

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(a) Pressure

(b) Tmax and

Fig. 7. Comparisons of combustion characteristics for SSI and TSI.

the engine is run at 4500 rpm, the equivalence ratio (Φ) is fixed at 0.8, and the volume of enriched hydrogen (αH2) is adjusted to 3%. The Φ and αH2 are severally expressed in Eqs. (1) and (2) [13]:

Table 5 Coordinates of different T-plug arrangements. Case

T/L

T-plug coordinates

A3

1

(13.8, −63.76, 0)

B1

0.5

(6.9, −63.16, 0)

Model (at TDC)

Φ=

VH 2 ρH 2 AFst , H 2 + mPRF AFst , PRF Vair ρair

αH 2 = B2

1.5

(20.7, −64.12, 0)

B3

2

(27.6, −63.91, 0)

VH 2 × 100% VH 2 + Vair

(1)

(2)

where VH2 and Vair represent the volume of hydrogen and air at the standard atmospheric pressure when the temperature is 300 K, respectively. ρH2 and ρair denote the densities of hydrogen and air at the standard atmospheric pressure when the temperature is 300 K. AFst,H2 and AFst,PRF indicate the stoichiometric air-to-fuel ratios of hydrogen and PRF, severally. mPRF is the mass flow rate of PRF. The combustion efficiency (η) mentioned in the current work is calculated by Eq. (3) [46]:

Fig. 8. Contours of temperature distribution at varying EAs. 6

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in the engine system are implemented in two aspects. From one side, remove the carburetor and magnetoelectric ignition system, and replace them with gasoline injector and electronically controlled ignition module, respectively. From the other side, to realize enriched hydrogen function and synchronized with gasoline injection, a hydrogen injector is mounted on the intake port, and the self-developed hybrid electronic control unit governs the spark module and port-injection strategy. More details regarding experimental setup and information about data measurement procedure are mentioned in Ref. [49]. In the experiment, αH2 was controlled through altering the injection duration by the hybrid electronic injection system. All the tests were run on the lean combustion condition of the excess air coefficient of 1.25 (Φ = 0.8), which was determined by operating the gasoline injection pulse width based on the wideband lambda sensor. For each testing point, the engine was imposed to run for some minutes until the excess air coefficient, exhaust gas temperature, as well as the carbon dioxide concentration, had reached their steady-state values, and then the data were recorded repeatedly. The resolutions and sensitivities of the measurement instruments are depicted in Table 3. Theoretically, the more of the cycles are, the better the statistics about pressure data are, and accordingly the accuracy of using 200 cycles can realize pressure acquisition. Besides, the adoptive gasoline is mixed with 2% engine oil in volume to lubricate the sealing elements due to the prototype engine without independent lubricating system. The mean pressure and instantaneous heat release curves of experiment and computation are shown in Fig. 5. The simulation is carried out for the same ambient and operating conditions that are observed under the experiment. The discrepancy (in percent) of pressure at maximum pressure course is less than 0.2% when αH2 is 3%. Regarding the instantaneous heat release rate, it is observed that the simulative phase lag match measured values within 1°EA. Undoubtedly, the modeling results perform an agreement with experimental data, and therefore the simulation model and subsequent numerical study are credible.

Fig. 9. Contours of velocity magnitude at spark moment.

4. Results and discussions 4.1. Comparative analysis of SSI and TSI configurations In this section, the combustion characteristics comparison between SSI and TSI configuration is carried out when αH2 is 3%. For SSI, based on the ignition parameter of the original engine (A0), this section designs two different strategies of SSI, among which A1 is a split ignition mode, while A2 is designed with higher discharge energy. For TSI (A3), L-plug and the trailing-spark plug (T-plug) are mounted symmetrically with respect to the minor axis, and the ignition parameter of T-plug is consistent with that of L-plug. It must be stated here that, as is known that ignition energy has a greater influence on combustion, so the total ignition energy (0.04 J) remains unchanged for all subsequent schemes to concentrate on desired research. The specified parameters of SSI and TSI are listed in Table 4. Fig. 6 describes the influence of variations of the ignition strategy on the flame growth and streamlines spatial distribution at 340°EA, 350°EA, TDC, and 380°EA. The stretched flame surface is characterized by the maximum magnitude of OH radical concentration. Around 5°EA after the spark (340°EA), there is one or two kernels induced by the spark when the spark ignition happened for all the schemes, while the flame kernel of the split ignition has a bigger size than that of the high discharge energy. This is because of the earlier occurrence of spark ignition. With the effect of the mainstream field in the rotor chamber, it can be clearly observed that the flame front of varying cases is similar and propagates in the same rotating direction of the rotor. At the moment of TDC, the twin kernels for A3 begin to connect together to form an extensive flame distribution, and the combustion front occupies the trailing section of the rotor chamber quickly enough (380°EA). It should be noticed that there is a competition between the unidirectional field

Fig. 10. Sum of OH, O and H radical peak values.

∑ x i QHV ·i ⎫ × 100% η = ⎧1 − ⎨ [mPRF /(mPRF + mair + mH 2 )] QHV ·PRF ⎬ ⎭ ⎩

(3)

the terms xi and QHV∙i is severally the mass fractions and lower heating values of incomplete combustion products, namely hydrogen, unburned hydrocarbon and carbon monoxide [47]. mair and mH2 denote the mass of air and hydrogen. QHV∙PRF is the lower heating value of PRF, i.e. isooctane and n-heptane [48]. 3. Model validation The confirmation of the developed approach is carried out by comparison of predicted engine performance with available experimental data for the prototype engine. The schematic and photograph of the experimental apparatus are illustrated in Fig. 4. The modifications 7

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(a) Pressure

(b) MPRR

Fig. 11. Comparisons of combustion characteristics for varying T-plug location. Table 6 Comparison of TSI timing schemes. Case

B2 C1 C2 C3 C4

Spark timing L-plug

T-plug

335°EA 325°EA 335°EA 345°EA 335°EA

335°EA 335°EA 325°EA 335°EA 345°EA

and the flame backward-spread in the chamber. The flame front of SSI cases could hardly extend in the reverse direction of the rotor rotation, and accordingly, there is an amount of unburned mixture in the trailing side of the chamber compared with TSI case. The combustion pressure curves of A1, A2, and A3 as a function of EA are depicted in Fig. 7(a). Compared with A0 (Fig. 5), the pressure traces for other cases are shifted to the upwards clearly while the peak positions occur earlier. A3 has higher overall pressure and steepest pressure gradient, relative to those in the other cases. Due to the existence of twin-spark plug configuration, the initial growth of combustion is facilitated as demonstrated in Fig. 6. Thereafter the advanced burning phasing with ignition source increase is conducive to thermal efficiency, and higher combustion pressure is exerted consequently.

Fig. 13. Contours of TKE distribution at varying spark moments.

Fig. 12. Contours of CH2O distribution at varying EAs. 8

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Regional distributions of temperature across the rotor chamber for different T-plug arrangements at 350°EA, TDC, 370°EA, and 390°EA are shown in Fig. 8. The distance between L-plug and T-plug determines the space for flame development. In the initial combustion stage (at 350°EA and TDC), the high-temperature region across the chamber decreases with the increment of the T/L value. This is primarily caused by the increase in the velocity magnitude around T-plug at spark moment (Fig. 9), which contributes to combustion initiation and yields a tremendous amount of heat. However, the contrary pattern is witnessed at subsequent 370°EA and 390°EA such that high concentration region of temperature for B2 and B3 covers the larger section than that for A3 and B1. This is due to the fact that the combustion sources for A3 and B1 are connected together at the early moment (TDC). After two burning points connected to each other, the rising rate of the combustion area begins to slow down [32], thereby decreasing high-temperature regions for A3 and B1. It is interesting to note that compared with B3, the temperature spread of B2 is faster and crashes upon the leading edge of the rotor chamber rapidly. This is mainly because of the discrepancy of the reaction rate and combustion intensity, and the concentration of highly active radicals can be treated as a reasonable indicator to reflect this phenomenon [50]. In Fig. 10, the sum of peak mass fractions of OH, O and H radical for different four schemes are given by stacked column, respectively. Evidently, the sum value of B2 presents higher than that of the other schemes. This indicates more hydrogen and radicals participate in some significant branching reactions in high-temperature oxidation [13], such as OH + H2 ⇔ H + H2O, H + O2 ⇔ OH + O, and therefore promotes temperature extension and combustion progress. That fact is presented even clearer by considering the pressure curves in Figs. 7(a) and 11(a). During the early stage of the combustion process, B2 has lower combustion pressure, analogous to the trend in Fig. 8, which confirms the analysis above. However, as the combustion continues, the peak pressure of B2 becomes higher and the corresponding position occurs earlier. Compared with B1, the peak pressure of B2 is increased by 9.6%, while the corresponding position is advanced by 3°EA. This clear trend is also appreciated in Fig. 11(b), exhibiting the maximum pressure rising rate (MPRR), which could characterize the combustion intensity and performance of WE [51]. It can be seen that different T-plug locations have an influence on the MPRR. Referring to Fig. 10, the combustion rate is enhanced and heat release is increased by optimizing TSI arrangement, this is the main reason for the increase in MPRR with a proper T-plug location. As shown in Fig. 11(b), the positions of MPRRs for different four cases are all after TDC. The MPRR of B2 presents the highest, and the other cases all decrease somewhat comparing with B2. This means when T-plug is arranged at T/L = 1.5, the engine performs a better combustion performance.

Fig. 14. Variations of HO2 concentration as a function of EA.

Moreover, it can also be seen from Fig. 7(b) that, from A0 to A3, the maximum temperature in the rotor chamber increases successively, which implies a higher thermal atmosphere in the stator thus promote the oxidation of incomplete combustion products, and contributes to the increment of the combustion efficiency as expected in Fig. 7(b). Therefore, as a result of the comparative analysis of these ignition strategies, the schemes of split ignition and high discharge energy divulge rather limited impact, though these two elements have a positive influence on improving combustion efficiency. On the other side, the substantial improvement in combustion resulting from extensive flame distribution by introducing TSI is realized in this section. With that in mind, the combined effect of different ignition strategies on combustion amelioration of TSI WE will be presented for further analysis. 4.2. Influence of TSI locations on combustion For the convenience of analysis, the definition of the coordinate system has been given in Fig. 1. The coordinate of L-plug (original source) is (−13.8, −63.76, 0). With the arrangement for L-plug remains constant, different location placement of T-plug is displayed in Table 5. The dimensionless positions of T-plug are represented by T/L, where T and L are the distance of the minor axis to the ignition point of T-plug and to the ignition point of L-plug, respectively. All of the spark timing is fixed at 335°EA, and the ignition mode is synchronous for TSI in this section.

(a) Pressure

(b) EA10 and EA50

Fig. 15. Comparisons of combustion characteristics for synchronous and asynchronous ignitions. 9

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Fig. 16. Contours of hydrogen distribution at varying EAs.

is advanced (or delayed), T-plug timing remains the same, and the opposite is true when T-plug timing is advanced (or delayed), as listed in Table 6. Fig. 12 displays the CH2O concentration distributions for different ignition strategies associated with B2, C1, C2, C3 and C4 with EA evolution from 350°EA to 380°EA. The cut on Fig. 12 runs through the median cross-section of the engine, perpendicular to the eccentric shaft. The edge of CH2O concentration is employed to represent the flame front surface [34]. It can be clearly observed that the flame surface of five schemes presents quite different during the combustion initiation, though the flame kernels of L-plug and T-plug have formed and their propagation direction is identical to the same the direction of rotation of the rotor. Amongst, the flame surface of C1 and C2 are larger, and the initial flame growth of T-plug is faster than that of L-plug by contrasting C1 to C2. This phenomenon can be explained by the following reasons. Firstly, the initial-spark timings of C1 and C2 are relatively earlier, which results in an increase of the kernel extension time. Secondly, the difference in the turbulence field is the underlying cause of how the spark timing affects the combustion initiation, early ignition produces more intense turbulence at spark moment than late ignition (Fig. 13), which facilitates the flame propagation. Finally, as shown in Fig. 13, the surrounding turbulence kinetic energy (TKE) amplitude of T-plug is higher than that of L-plug regardless of 325°EA or 335°EA, which contributes to the flame of C2 has a broad stretched surface. In addition, with the proceeding of combustion, C1 yields faster flame spread and is quenched in the leading edge of the rotor chamber at 380°EA. Since CH2O is mainly consumed via chain-branching reactions of CH2O + OH ⇔ HCO + H2O and HCO + O2 = HO2 + CO [52], it can be observed from Fig. 14 that a higher mass fraction of HO2 is produced under C1 condition, and thereby leads to more rapid combustion. Fig. 15(a) gives the pressure variation of Ci over EA. Taking Fig. 11(a) together with Fig. 15(a), it can be noticed that compared with synchronous ignition, an earlier spark timing of L-plug or T-plug provides the higher pressure output and faster flame growth. As expected in Fig. 15(b), both EA10 (the eccentric angle whose accumulated heat amount at 10% of total heat amount) and EA50 (the eccentric angle whose accumulated heat amount at 50% of total heat amount) have the tendency of being advanced with respect to the schemes of C1 and C2. Specifically, for C1 and C2, even though EA10 of C1 is somewhat delayed, unexpectedly EA50 of C1 is advanced. When L-plug is ignited earlier, the combustion is more efficient. The duration of an oxidation reaction in the succeeding combustion is likely to be shortened because of higher temperature in the stator. To sum up, the introduction of the pre-ignition of L-plug leaves a positive effect on combustion performance of WE.

Fig. 17. Contours of temperature distribution at 340°EA.

Fig. 18. Variations of reactants as a function of EA.

4.3. Influence of asynchronous ignition on combustion Based on the preferable TSI arrangement, the effect of time difference of TSI on the combustion process is further investigated by contrast cases in this part of the simulation. As mentioned above, the timings of TSI plugs is synchronously set to be 335°EA for the base case (B2). L-plug and T-plug are ignited at 325°EA (or 345°EA) which are indicated as C1 and C2 (or C3 and C4), respectively. When L-plug timing

4.4. Influence of discharge energy on combustion It has been demonstrated that asynchronous ignition is prone to promote flame propagation and combustion rate under C1 condition. To 10

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(a) Pressure

(b) IHRF

Fig. 19. Comparisons of combustion characteristics for synchronous and asynchronous ignitions.

Importantly, it is specifically concentrated on the effects of the number of the ignition source, twin-spark plug arrangement, asynchronous ignition, and discharge energy allocation. Based on the thorough analysis, the conclusions achieved in current work may be summarized as follows:

further ameliorate the ignition and combustion of WE, the influence of discharge energy for TSI on combustion is studied herein. For C1, the discharge energy of L-plug and T-plug is allocated equally with 0.02 J. For D1, the discharge energy of L-plug is 0.01 J while that of T-plug is 0.03 J. The contrary manner is allocated for D2, namely the discharge energy of L-plug is 0.03 J while that of T-plug is 0.01 J. Besides, TSI is adjusted to the same spark timings, which is the same as that of C1. Fig. 16 shows H2 concentration across the cross-section area of the rotor chamber with EA variation after 345°EA onwards at different values of the TSI energy. It can be seen that the evolution of combustion process presents an apparent difference of three ignition strategies. The discharge energy of L-plug and T-plug defines the early flame kernel (Fig. 17), and thus accelerates hydrogen consumption with an increment of combustion velocity, which is due to higher discharge energy making larger flame surface [53]. It is worth noting that burning of the D2 mixture is more rapid than other cases in the leading part of the rotor chamber, and the locations of the right-propagated front are closer to the trailing edge at 375°EA. This is because the duration between the spark timing and blending moment of the TSI source is increased, the flame kernel of T-plug has more time for propagating towards the trailing side of the rotor chamber, which contributes to enlarge the combusted area. It can be deduced that a faster flame extension of T-plug has a negative effect on the combustion quality. As quantitated in Fig. 18, the consumption of the reactants increases with the reduced ignition energy of T-plug, and the initiation of combustion is advanced slightly. It is also verified in Fig. 16 that the proportion of unburned mixture in corner regions is less under D2 condition. The comparisons of combustion pressure and integrated heat release fraction (IHRF) of C1, D1, and D2 are displayed in Figs. 15(a) and 19. Owing to the higher rate of burning out of the mixture, it can be found that the rate of pressure increase of D2 is greater than that of C1 and D1. The peak pressure increases with an increment of L-plug energy. It is widely accepted that IHRF directly characterizes the speediness of the combustion process. As shown in Fig. 19(b), the initiation and termination of the heat release is advanced with the increase of L-plug energy. D2 has steepest IHRF gradient compared to that of others. D2 is capable of creating fast fuel consumption and high overall pressure, and accordingly generates more heat and shortens combustion durations.

(1) For single-spark ignition, despite split ignition and high discharge energy have a positive effect on facilitating combustion process, these two factors divulge rather limited. Additionally, the introduction of twin-spark ignition presents a substantial improvement in the flame spread and combustion efficiency because of the extensive burning distribution. (2) The growth rate of combustion zone is not constant, this change is involved in the shift of the flame into high-velocity regions and the connection of the burning source within the rotor chamber. The twin plug location determines the space for flame development. When the location of the leading plug is locked, the burning intensity is strengthened, combustion pressure increases and maximum pressure rising rate presents higher as the trailing plug is arranged at T/L = 1.5. (3) An earlier ignition of the leading plug or trailing plug enables the flame propagates faster, and the mixture is consumed more quickly. The discrepancy in the turbulent kinetic energy amplitude is responsible for the influence of the timing difference for twin-spark ignition. Only when the turbulence is strong will the combustion initiation be accelerated. As the leading plug is ignited earlier, the flame yields faster speed and occurs quenching rapidly, which contributes to higher pressure output and better heat release. (4) Under the condition of the same total ignition energy, the reactants consumption rate increases with an increment of the leading plug discharge energy. Compared with other schemes of energy allocation, lower discharge energy of the trailing plug makes combustion pressure higher and combustion duration shorter as a whole. It can be concluded that faster flame extension of the trailing plug has a negative effect on the combustion quality. (5) Considering the combustion characteristics, D2 is the optimum ignition strategy under this computational condition. Namely, with the placement for the leading plug remains constant, the location of the trailing plug is offset from the minor axis by 20.7 mm; the spark timing and discharge energy of L-plug are 325°EA and 0.03 J, respectively; and those of the trailing plug are 335°EA and 0.01 J correspondingly. (6) With respect to the engineering application, the dual-spark plug is a superior configuration for Wankel engines, especially for lean operation. Leaving a longer distance between the twin spark plug has

5. Conclusions In the present study, twelve schemes of the ignition strategy were implemented with establishing a three-dimensional computational fluid dynamics model to simulate the kernel formation, flame propagation and combustion process of hydrogen-enriched Wankel engines. 11

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the potential to increase combustion efficiency. Moreover, it is highly recommended that the ignition of the leading plug is set earlier and stronger than the ignition of the trailing plug. To realize ideal lean-burn performance in real applications, the synergistic effects of suitable plug arrangement, asynchronous ignition mode and inhomogeneous energy allocation is an effective and practical way.

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