Energy Conversion and Management 123 (2016) 209–217

Contents lists available at ScienceDirect

Energy Conversion and Management

journal homepage: www.elsevier.com/locate/enconman

Effect of the Miller cycle on the performance of turbocharged hydrogen internal combustion engines ⇑ Qing-he Luo, Bai-gang Sun

School of Mechanical Engineering, Beijing Institute of Technology, Beijing 100081, China article info abstract

Article history: Hydrogen is a promising energy carrier, and the port fuel injection (PFI) is a fuel-flexible, durable, and Received 11 January 2016 relatively cheap method of energy conversion. However, the contradiction of increasing the power den- Received in revised form 16 May 2016 sity and controlling NOx emissions limits the wide application of PFI hydrogen internal combustion Accepted 12 June 2016 engines. To address this issue, two typical thermodynamic cycles—the Miller and Otto cycles—are studied Available online 20 June 2016 based on the calculation model proposed in this study. The thermodynamic cycle analyses of the two

cycles are compared and results show that the thermal efficiency of the Miller cycle (gMiller) is higher than Keywords: g , when the multiplied result of the inlet pressure and Miller cycle coefficient (d c ) is larger than Miller cycle Otto M M that of the Otto cycle (i.e., the value of the inlet pressure ratio multiplied by the Miller cycle coefficient Hydrogen internal combustion engine Numerical calculation is larger than the value of the inlet pressure ratio of the Otto cycle). The results also show that the intake valve closure (IVC) of the Miller cycle is limited by the inlet pressure and valve lift. The two factors show the boundaries of the Miller cycle in increasing the power density of the turbocharged PFI hydrogen engine. The ways of lean burn + Otto cycle (LO), stoichiometric equivalence ratio burn + EGR + Otto cycle (SEO) and Miller cycle in turbocharged hydrogen engine are compared, the results show that the Miller cycle has the highest power density and the lowest BSFC among the three methods at an engine speed of 2800 rpm and NOx emissions below 100 ppm. The brake power of the Miller cycle increases by 37.7%

higher than that of the LO and 26.3% higher than that of SEO, when cM is 0.7. The BSFC of the Miller cycle decreases by 16% lower than that of the LO and 22% lower than that of SEO. However, the advantage of the Miller cycle decreases with an increase in engine speed. These findings can be used as guidelines in developing turbocharged PFI hydrogen engines with the Miller cycle and indicate the boundaries for the development of new hydrogen engines. Ó 2016 Elsevier Ltd. All rights reserved.

1. Introduction be achieved without substantial modifications to traditional internal combustion engines [2–4].

Given the seriousness of the energy problem and the lack of Two fuel supply methods can be implemented for H2ICE: the stringent emission regulations, the search for a new type of clean port fuel injection hydrogen internal combustion engine (PFI- renewable energies has become the most important topic in the H2ICE) and direct injection hydrogen internal combustion engine energy sector. As an energy carrier, hydrogen is a promising means (DI-H2ICE). DI-H2ICE is considered an effective means of increasing of achieving clean combustion given that its major combustion power density and limiting NOx emissions [5]. The direct injection product is H2O when NOx emissions are well controlled [1]. Hydro- strategy can have a higher in-cylinder trapped air mass at the start gen can generally be used in two ways: through direct combustion of fuel injection after IVC. However, this method is difficult to and through hydrogen-fueled cells. However, hydrogen fuel cells apply because of the limited lubrication and cooling of the injector are limited by technical bottleneck and high price, such as proton and the large cycle variation. The PFI strategy does not result in exchange membrane, carbon plate, and the high price of the metals such problems, and most of the parts of the gasoline engines can for catalysis. The direct combustion of hydrogen via hydrogen- be used in PFI-H2ICE, which makes the PFI-H2ICE system cheaper fueled internal combustion engines (H2ICE) can easily and cheaply and more reliable than the DI-H2ICE system [6,7]. However, the contradiction of the power density and control-

ling NOx emissions limits the PFI-H2ICE, because hydrogen is the ⇑ Corresponding author. gas fuel that takes up the cylinder volume, leading to a decrease E-mail address: [email protected] (B.-g. Sun). in the volume coefficient [8]. To solve this contradiction, the http://dx.doi.org/10.1016/j.enconman.2016.06.039 0196-8904/Ó 2016 Elsevier Ltd. All rights reserved. 210 Q.-h. Luo, B.-g. Sun / Energy Conversion and Management 123 (2016) 209–217

supercharged PFI-H2ICE has been studied by many researchers. 2. Materials and method Verhelst et al. [9] results showed that maintaining the supercharg- ing stoichiometric mixtures coupled with exhaust gas recirculation 2.1. Test equipment and process (EGR) is a better method of controlling NOx emissions than the method using lean mixtures. Ghent University developed a super- A correctly set-up engine and test cell with an appropriate data- charging H2ICE (GM 7.4 L V8) [10], which can increase maximum acquisition system were built to acquire precise and reliable infor- torque output by 60% compared to naturally aspirated H2ICE. mation. Numerous tests were conducted on a 2.3 L four-cylinder White et al. [11] presented a review of H2ICEs on boosted H2ICEs hydrogen engine with turbocharger. and showed that such engines are a compelling alternative to fuel The test cell included a CW250 eddy current dynamometer, cells when used in a series hybrid vehicle featuring very low NOx which was used for energy absorption and engine speed regula- emissions without any post-treatment at low equivalence ratios. tion. The dynamometer had a capacity of 250 kW and a maximum Berckmüller et al. [12] worked on a supercharged stoichiometric rated speed of 8000 rpm and was attached to an external blower, 3 equivalence ratio H2ICE and showed that the peak power output which volume flow rate is 500 m /h. The engine speed could be of a four-stroke single cylinder engine is about one third higher controlled at the desired level, and the torque was varied. The than that of a naturally aspirated gasoline engine. Natkin et al. speed, engine oil temperature, coolant temperature, and intake [13] reported work on a supercharged 2.3 L PFI engine with a stan- air temperature were recorded automatically from the dynamome- dard (air-to-air) intercooler and a fixed equivalence ratio of 2. Their ter control console. The cylinder pressure was measured using a results showed that the torque decreases by 28% compared to the Kislter 6117B pressure sensor, and the crank angle position was base (naturally aspirated) gasoline engine. In the recently intro- specified by the crank angle encoder, which type is Kislter 2613B duced Ford E-450 H2ICE vehicles, a 6.8 L supercharged PFI engine (which can measure 0–20,000 r/min and the accuracy can reach was used at a low equivalence ratio (not specified), eliminating to 0.1 °CA). The cylinder pressure and the corresponding crank the need for post-treatment [14]. From the above, supercharging angle were captured through a high-speed data acquisition system can increase the power density under acceptable NOx limits. How- (Kibox combustion analyzer). The output from these measure- ever, supercharging consumes the useful work of the output, and ments was the diagrams of pressure versus crank angle. The inlet the EGR cooler system is difficult to achieve because of the com- air flow was measured by the Hot-film Air Mass Flow Meter, which bustion products, including a huge amount of water vapor. The type is ToCeiL20N. The hydrogen mass flow was measured by the combustion products lead to a very high export temperature in Coriolis Mass Flow Meter, which type is CMF025 made by EMER- the EGR cooler, along with an increase in the inlet temperature. SON. The specifications were listed in Table 1. In the 1940s, Miller proposed a different Otto cycle with Before conducting the test, the engine was warmed up to ensure unequal compression and expansion stroke called the Miller cycle. that it reached the operating temperatures and that it stabilized. The most interesting feature of the Miller cycle is the shorter effec- Tests were conducted after running the engine until it reached a tive compression stroke and longer expansion stroke, which can steady state at an oil temperature of 90 °C and cooling water tem- increase the thermal efficiency [15]. Li et al. [16] reported the perature of 80 °C. The data was then recorded after running the effects of early intake valve closure (EIVC) and later intake valve engine with hydrogen fuel. All the tests were run at the MBT to closure (LIVC) on the fuel economy of a boosted DI gasoline pro- obtain comparable data. The MBT tests were done through modify- duction engine reformed with a geometric CR of 12.0 are experi- ing the ignition timing at different engine speed and load. For mentally compared at low and high loads. They found that the example, to get the MBT at 2800 r/min and the wide open throttle, LIVC could improve the brake specific fuel consumption (BSFC) the engine speed should be kept at 2800 r/min and the throttle was up to 4.7% than the CR of 9.3 at high load, but not ELVC. At the wide open, then modifying the hydrogen mass flow to keep the low load operation, combined with CR12.0, LIVC and EIVC improve equivalence ratio at constant (such as 0.55). After these steps, the the fuel economy by 6.8% and 7.4%, respectively, compared to the ignition timing was modified internal of 1 °CA or bigger values production engine. Hou [17] comparison of performances of air through the electronic system. Record the test data until finding standard Atkinson and Otto cycles with heat transfer considera- the MBT. Repeat above steps to get the other test data at different tions including combustion constants, compression ratio and speeds, equivalence ratios and throttle open angles. intake air temperature. They found that an Atkinson cycle (one All measurements of physical quantities have some degree of kind of Miller cycle) has a greater work output and a higher ther- uncertainty, owing to various sources. Therefore, uncertainty anal- mal efficiency than the Otto cycle at the same operating condition. ysis is desirable, to confirm the precision of the tests. The uncer- On the other hand, the technology of PFI-H2ICEs has been very tainty for the experimental results is determined according to mature [18] and the way of PFI can ensure the uniformity of mix- the principle of root-mean square method, to get the magnitude ture to decrease the cycle variations. Thus, this study proposes of the error given by Gaussian distribution, as follows: using the Miller cycle in PFI-HICEs with turbocharger to solve the &’ = 2 2 2 1 2 contradiction of the power density and controlling NOx emissions @R @R @R DR ¼ Dx1 þ Dx2 þþ Dxn ð1Þ as the limitation of PFI-H2ICE. To study the effect of the Miller @x1 @x2 @xn cycle, a calculation model is developed based on experimental data on a turbocharged hydrogen engine.

Table 1 Average uncertainty of the main measured parameters.

Variable Device Accuracy Parameters 2800 rpm, WOTa Error (%) Engine speed FC2000 ±1 r/min Brake power 28.1668 0.0574 0.20 Crank angle Kislter 2613B ±0.02° ITEb 32% 0.07% 0.23 In-cylinder pressure Kislter 6117B ±0.4% ARa ISFCb 93.7450 0.2127 0.23 Air mass flow rate ToCeiL20N ±1%FSa Volumetric efficiency 83.27% 0.77% 0.92 Hydrogen mass flow rate CMF025 ±0.1%FSa Equivalence ratio 0.5170 0.0026 0.51

a FS: Full scale, AR: All range, WOT: wide open throttle. b ITE: Indicated Thermal Efficiency, ISFC: Indicated Specific Fuel Consumption. Q.-h. Luo, B.-g. Sun / Energy Conversion and Management 123 (2016) 209–217 211

Fig. 1. Schematic of the hydrogen internal combustion engine with turbocharger. where DR is the uncertainty in the computed result, R is a given BDUR is user-entered combustion duration (10–90%); WEXP is function of the computed results, x1, x2, xn are the independent user-entered exponent in Wiebe function. measured variables, Dx1, Dx2, Dxn are the corresponding uncertainty In the current work, Wiebe 2 zone combustion model is used. values of the independent measured variables [19]. The inputs for the Wiebe 2 Zone combustion model are same as Table 1 summarizes the average uncertainties of the measured the Wiebe function. However instead of one mass averaged tem- parameters based on the specification of the instruments and perature, two temperatures (burned and unburned zones) are cal- experimental error analysis. The error analysis was performed by culated [20]. One more advantage of using this model is that this considering error rates in the measurement range of devices model also predicts the knocking and NOx characteristics of the (according to their calibration values) used in the experimental engine, provided the actual heat release is described properly by studies. the Wiebe function specified. In the previous work, the parameters for combustion model can be set using the test data, which has a 2.2. Calculation model and validation good correlation with the calculation model. Thus, this paper more emphasis is given on the test data. The gas dynamic approach was used to develop the calculation model. Fig. 1 shows the diagram of the model of the four-cylinder 2.2.2. Details of the engine friction model four-stroke PFI-H2ICE with turbocharger. Table 2 lists the real For calculation of engine friction, the Chen-Flynn Fiction model engine specifications. The analysis of the effect of the Miller cycle is used. The equation used to calculate friction is given below: on the power density of the PFI-H ICE with turbocharger was per- 2 Xncyl hi formed using this model. ¼ þ ð Þþ ð Þ þ ð Þ2 ð Þ FMEP Acf Bcf Pmax Ccf Sfact i Ccf Sfact i 3 The calculation model mainly included four sub-models: geom- i¼1 etry, combustion, heat transfer, and friction models. The geometry Sfact ¼ RPM stoke=2 ð4Þ model included the main engine specifications, as shown in Table 2.

The SI Wiebe, Woschni heat transfer and Chen-Flynn Friction mod- where Acf, Bcf, Ccf, Qcf need user input according to parameters of the els were employed for the combustion, heat transfer and friction calculation engine (the value of FMEP is modified using the test data processes, respectively. for this study); Pmax is maximum cylinder pressure; RPM is cycle- average engine speed; stroke is cylinder stroke. 2.2.1. Details of the combustion model The Wiebe function is often used to approximate the actual heat 2.2.3. Details of the heat transfer model release characteristic of the engine [20]. The cumulative mass frac- The heat transfer model used for the calculation is Woschini tion burned as a function of crank angle is given by the following: model which is available in the Boost software and this model cal- culates the heat transfer rates to the piston, cylinder head, and ¼ : ð ð = ÞðWEXPþ1Þ ð Þ W 1 0 exp AWI h BDUR 2 liner [21]. Wall temperatures are provided as the input to the heat transfer model and the wall temperatures of piston, cylinder head, where AWI is internally calculated parameter to allow BDUR to liner at TDC and liner at BDC are assumed as 585 K, 585 K, 585 K cover the range of 10–90%; h is degrees past start of combustion; and 408 K respectively [22].

Table 2 2.2.4. Details of the NOx model Engine characteristics. For accurate treatment of NOx kinetics, which is strongly tem- Engine type Hydrogen internal combustion engine perature dependent, the non-homogeneity of the temperature field Bore 90 mm within the combustion chamber must be taken into account. Thus, Stroke 90 mm the NOx emissions sub-model requires specification of the 2-zone Cylinders 4 combustion thermodynamics. All the NOx is assumed to be in the Compression ratio 9.3 form of NO during the prompt formation phase as well as the ther- Crank radius 45 mm mal phase described below by the extended Zeldovich mechanisms Connecting rod radius 153 mm Exhaust valve open 53.64°BBDC (Before Bottom dead center) of NOx formation [20]. The concentration of NO versus time is Exhaust valve close 44.36°ATDC (After Top dead center) solved using an open system in which the above elementary reac- Inlet valve open 23.3°BTDC (Before Top dead center) tions are used with those rate constants reported by Heywood [21]. Inlet valve close 58.7°ABDC (After Bottom dead center) For the first reaction equation, the rate constant, R1, is given by: 212 Q.-h. Luo, B.-g. Sun / Energy Conversion and Management 123 (2016) 209–217

AERC1 ðTa Þ R1 ¼ A ARC1 e T ð5Þ the cylinder enters the turbine, which drives the turbine to work. The work is used to compress the inlet air in the compressor. The For the second and third reaction equations, the rate constant, compression pressure of engine inlet-air is less than 2 bar in the R2/3, is given by: experimental studies. ð = Þ Ta T Before using the calculation model to study the Miller cycle, its R2=3 ¼ A e ð6Þ precision must be validated using the test data. Fig. 2 compares the where A is the pre-exponential constant; ARC1 is user-enter cylinder pressure and air mass of the calculated data with those of pre = exponent multiplier (the value is set to 1.5); Ta is activation the test data. temperature for the reaction; AERC1 is user-enter exponent (the The calculated and experimental results tend to agree with each value is set to 1); T is burned-zone temperature. other. Given an average deviation around 10%, a discrepancy is The calculation is terminated when the temperature in the assumed between the calculated and the experimental results. burned zone reaches a low enough level so that the kinetics The effects of blow-by, leakage, and cycle-to-cycle variations par- become inactive and total NO no longer changes. tially account for the discrepancies [18,23]; the uncertainties of the measured parameters also contribute to the discrepancies. 2.2.5. Details of the knock model The engine performance of PFI-H2ICE with turbocharger can there- The knock model available in the WAVE calculates the mini- fore be predicted through the present calculation model. Further mum octane number required for the engine operation free of estimation and optimization can be performed using this model. knock [20]. The induction time (ignition delay) in seconds is calculated at every time step using the following equation: 2.3. Method of achieving the Miller cycle in turbocharging hydrogen 3:4107 engines ON : 3800=A s ¼ 0:01869=A P 1 7exp T ð7Þ p 100 T To prevent the PFI-H2ICEs from backfiring, EIVC was chosen to where A俄 is user-entered pre-exponential multiplier; ON is user- achieve the Miller cycle. The timing of IVC and the value of the entered fuel octane number (130 for hydrogen); P is cylinder pres- valve lift have an unknown relationship [24]. An earlier timing of IVC generally leads to a smaller valve lift. If the valve lift is con- sure in atm; AT is user-entered activation temperature multiplier; T is unburned gas temperature, K. stant, a shorter duration of the intake stroke increases the velocity, In general, this induction time continually decreases as com- acceleration, and contact stress of the valve lift, which causes unre- bustion progresses and the unburned zone temperature rises. The liability or even damages in the valve seat. Thus, achieving the end-gas auto-ignites (knocks) if the induction time is less than Miller cycle requires not only changing IVC but also changing the the flame arrival time. When knock occurs, a spontaneous mass valve lift with IVC. The main factors of determining the valve lift burning rate due to knock is determined and fed back to the cylin- are the velocity and acceleration of the valve lift, which are gov- der, leading to rapid rise in cylinder pressure and temperature. The erned by a valve lift coefficient. In this study, the valve lift curves in-cylinder heat transfer coefficient is also increased during knock. were obtained by changing the IVC and valve lift coefficients at The model assumes that auto-ignition occurs when: given coefficients of the Miller cycle, as shown in Fig. 3. Z The relationship between the IVC and Miller cycle coefficient ti ds ¼ ð Þ can be calculated using the following formula: s 1 8 t0 qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 3 1 2 2 where s is induction time, defined above; t0 is start of end-gas com- V ¼ pr ð1 cos uÞþ 1 1 k sin u ð9Þ x k pression; ti is time of auto-ignition. ¼ð þ Þ=ð þ ÞðÞ The developed model has four main parts: the intake system, cM V x V 0 V h V 0 10 engine crank train, turbocharger system, and exhaust system.

The air cleaner is connected to the air, as the turbine is. The com- where Vx is the volume when the piston movement to the crank- pressor, turbine, and inlet cooler comprise the turbocharger sys- shaft angle is u; V0 is the clearance volume; Vh is the cylinder effec- tem, which can compress the air from the air cleaner. The tive volume; r is the crank radius; l is the connecting rod radius; and compressed air enters the engine cylinder. The exhaust gas from k ¼ r=l is the ratio of the crank radius and connecting rod radius.

Fig. 2. Comparison of the air mass (a) of different throttle-openings and cylinder pressure (b). Q.-h. Luo, B.-g. Sun / Energy Conversion and Management 123 (2016) 209–217 213

0 (2) 2 -2 and 1-2 are the adiabatic compression processes: 0 j p2 V 2 ¼ 0 ð12Þ p V 2 2 p V j 1 ¼ 2 ð13Þ p2 V 1 (3) The heat addition occurs in 2-3 and is isochoric: ¼ ð Þ¼ ð ÞðÞ q23 CV T3 T2 V 2 p3 p2 14

(4) 3-4 is the adiabatic expansion process: p V j 3 ¼ 4 ð15Þ p4 V 3 (5) 4-5 is the isochoric expansion process: ¼ ð Þ¼ ð ÞðÞ q45 CV T4 T5 V 5 p4 p5 16 (6) 5-6-70-10 and 5-6-7-1 are the pumping work:

0 0 ¼ 0 0 þ 0 Fig. 3. Changes in the relationship between the valves lift and crank angle at w567 1 q7 1 q67 q65 q15 different cM. ¼ 0 ð 0 0 Þþ 0 ð 0 Þ ð Þ p7 V 1 V 7 V 7 p7 p6 p6 V 5 V 6 ð ÞðÞ 3. Results and discussion V 1 p1 p5 17

¼ þ 3.1. Thermodynamic cycle analysis of the Miller cycle in turbocharged w5671 q71 q67 q65 q15 engines ¼ ð Þþ ð Þ ð Þ ð Þ p7 V 1 V 7 V 7 p7 p6 p6 V 5 V 6 V 1 p1 p5 ð18Þ Fig. 4(a) compares the thermodynamic cycle of the Miller cycle 0 0 and that of the Otto cycle in turbocharging engines. The p-V figure 0 where p1, p2, p1, p2, p3, p4, p5, p6, p7 and p7 are the pressure and vol- shows that the Miller cycle consists of 10-20-2-3-4-5-6-7-10 and the 0 0 0 0 ume at states 1, 2, 1 ,2,3,4,5,6,7and7 respectively; V 1, V 2, V 1, 0 Otto cycle 1-2-3-4-5-6-7-1. The Miller and Otto cycles consist of 0 V , V 3, V 4, V 5, V 6, V 7 and V 7 are the volume at the states, respec- the following process: 2 tively; T2, T3, T4, and T5 are the absolute temperatures at states 2,

3, 4 and 5, respectively; j is the specific heat ratio (CP/CV), CP and (1) 10-20 is the adiabatic expansion process: CP are the constant pressure and constant volume specific heats,

0 j respectively; q23 and q45 are heat of the isochoric process for 2-3 p0 V 1 2 0 0 ¼ 0 ð11Þ and 4-5, respectively; w567 1 and w5 6 7 1 represent the pump- 0 0 0 p2 V 1 ing work of 5-6-7 -1 and 5-6-7-1, respectively.

Fig. 4. The thermodynamic cycle of comparison between Miller and Otto cycle. 214 Q.-h. Luo, B.-g. Sun / Energy Conversion and Management 123 (2016) 209–217

To make the formula concise, the following definitions are 3.2. Relationship between the IVC and inlet pressure of the Miller cycle given: The occurrence of abnormal combustion phenomena, such as

The turbocharger pressure ratio is given by backfiring, is a crucial limitation of PFI-H2ICEs. The above theoret- ical thermodynamic analysis shows that the larger the product of 0 p7 p7 the inlet pressure and Miller cycle coefficient, the larger the pump- dM ¼ ; dO ¼ ð19Þ p6 p6 ing work, and the higher the thermal efficiency. However, the pro- duct of the inlet pressure and Miller cycle coefficient is limited by The compression ratio is given by the abnormal combustion phenomena, particularly the knock, V e ¼ 1 ð20Þ because the pressure at the end of the compression stroke cannot V 2 exceed the limit (i.e., 3.5 MPa for the engine in this study). The The pressure rise ratio is given by Miller cycle coefficient and the IVC have a one-to-one relationship, p3 as discussed below. Fig. 5 shows the changes in the pressure at the kp ¼ ð21Þ p2 end of the compression stroke with the IVC and inlet pressure. To avoid the knock [25], 3.5 MPa is set as the limitation in this study. Eqs. (2)–(13) show that the thermal efficiency of the Miller cycle Fig. 5 shows that the smaller the IVC, the higher the inlet pres- can be deduced as follows: sure for the Miller cycle when the pressure at the end of the com- q q þ q 0 0 þ q 0 q q 1 d c d pression stroke is held constant. When the values of IVC and the g ¼ 23 45 7 1 67 65 15 ¼ 1 þ M M O Miller k1 k1ð Þ limit of the pressure remain the same, the inlet pressure at a high q23 e e kp 1 engine speed is larger than that at a low speed. The first reason for ð22Þ this phenomenon is that the calculated result of the pressure at the where dM and dO are the turbocharger pressure ratio; e is compres- end of the compression stroke is obtained when the mixture in the sion ratio; kp is pressure rise ratio; gMiller is the thermal efficient of cylinder consists of hydrogen and air rather than air only. The Miller cycle. higher the speed is, the smaller the compression polytropic coeffi- A comparison of the thermal efficiency of the Miller cycle and cient [26]. Thus, a higher inlet pressure is required to achieve the that of the Otto cycle is shown in Fig. 4(b). When the value of dM- same compression pressure because of the smaller compression cM is equal to dO, Eq. (22) is the thermal efficiency of the Otto cycle. polytropic coefficient. The second reason is that the duration of The thermodynamic cycle analyses of the two cycles are compared the mixture charging process shortens, while IVC remains the same and results show that the thermal efficiency of the Miller cycle at both speeds. To reach the same pressure at the end of compres-

(gMiller) is higher than gOtto, when the multiplied result of the inlet sion, a higher inlet pressure can provide the same amount of in- pressure and Miller cycle coefficient (dMcM) is larger than that of cylinder trapped mixture mass. However, the decrease in IVC is the Otto cycle (i.e., the value of the inlet pressure ratio multiplied limited. On the one hand, the smaller the IVC is (the higher the by the Miller cycle coefficient is larger than the value of the inlet inlet pressure), the more difficulties for the turbocharger system. pressure ratio of the Otto cycle). The proportions of increase are On the other hand, the smaller the IVC is, the shorter the intake determined by the difference between the values of dMcM and dO. stroke duration, and the smaller the valve lift. This result is caused This result theoretically proves the benefit of using the Miller cycle by a decrease in the inlet air flow mass in turn caused by pressure in turbocharging engines, which provides the basis for the follow- blocking. Fig. 6(a) shows the changes in brake power with changes ing analysis. in the IVC. The power curve rises at the beginning and falls at the Fig. 4(c) shows the effect of j on the thermal efficient changing high point. As the IVC increases, the flow intersection area of the with compression ratio. The value of j of adiabatic cycle can be set intake valve increases, which is the limitation of the in-cylinder to 1.4 as the average value for the theory analysis. However, this charging mass. Meanwhile, the inlet pressure decreases and value for the polotropic cycle is different for the different kinds becomes the controlling factor of the in-cylinder charging mass of type engine. From review [23], the variation value of j for gaso- as the IVC increases. As the IVC increase, this establishes the trend line engine is smaller than the hydrogen one, because of high of the indicated power. Thus, the IVC of the Miller cycle is limited hydrogen in-cylinder pressure and gas temperature. Thus, the by two factors, as shown in Fig. 6(b). The lines on the left is the hydrogen average value of j is also bigger than the gasoline one. smallest one limited by the turbocharger system and lift valve, Fig. 4(c) shows that the benefit of Miller cycle used for hydrogen the other one is limited by the insufficient of power. The two lines engine is higher than the gasoline engines. can limit the boundary of the inlet pressure of the Miller cycle in

Fig. 5. Changes of pressure at the end of compression with the IVC at different inlet pressures. Q.-h. Luo, B.-g. Sun / Energy Conversion and Management 123 (2016) 209–217 215

Fig. 6. Limitation caused by the power, turbocharger system, and the valve lift.

Fig. 7. Comparison of the broke power (a), BSFC (b), and temperature at the end of the compression stroke (c) of the three technical methods. 216 Q.-h. Luo, B.-g. Sun / Energy Conversion and Management 123 (2016) 209–217

Fig. 8. Three ways of brake power (a) and BSFC (b) change with engine speed. turbocharging engines. The design values of the Miller cycle coeffi- in the cylinder reduces the temperature at the cylinder, which in cient and inlet pressure can be obtained through the method turn decreases NOx emissions. The combined result consists of described in the real engines as the demanding power density. In an increase in the output power when the NOx emissions are lim- this study, 0.7was selected as the Miller cycle coefficients. ited below 100 ppm. However, the exhaust gas takes up the vol- ume, which means that ability of SEO in increasing power is 3.3. Comparison of three technical methods of turbocharging hydrogen smaller than that of the Miller cycle. In the thermodynamic cycle, engines the higher average temperature, the better the thermal efficiency of the cycle. Thus, the BSFC of SEO is the highest among the three To compare the performance and fuel economy of the Miller methods. The logic behind this finding is similar to the case of NOx cycle, lean burn (/ = 0.55 [6,22]) + Otto cycle (LO), and stoichio- emissions below 100 ppm when the NOx emissions are limited metric equivalence ratio burn + EGR + Otto cycle (SEO) in tur- below 200 ppm. bocharging hydrogen engines, 100 ppm was selected as the From Fig. 8(a), a comparison of brake power changing with threshold value for NOx emissions [8]. The values of EGR for SEO engine speed shows that the advantage of the Miller cycle in were adjusted by the values of NOx produced in the exhaust sys- increasing the power density decreases with an increase in the tem. Fig. 7 shows the results for the brake power, BSFC, and tem- engine speed. At 5200 r/min, the brake power of Miller cycle just perature at the end of the compression stroke of the three higher as 2.8% (100 ppm) and 7.2% (200 ppm) than the way of methods at 2800 and 5200 rpm. To validate the regular, the results SEO. The first reason for this phenomenon is the larger heat trans- of the three methods are also presented in Fig. 7, when the NOx fer that occurs at a high engine speed [27]. The second reason is emissions are lower than 200 ppm. Ma, which increases with an increase of engine speed. Substantial Fig. 7 shows that the method of the Miller cycle is the best experimental experience shows that the larger the Ma, the less the among the three because it has not only the largest brake power air intake mass. Thus, the advantage of the Miller cycle decreases at output but also the lowest BSFC. The brake power of the Miller a high engine speed. Fig. 8(b) shows that the BSFC of Miller cycle is cycle is higher as 37.7% and 26.3% than those of LO and SEO, the smallest changing with engine speed. The comparison of the results for the three methods show the respectively, when the Miller cycle coefficient (cM) is 0.7. The BSFC of the Miller cycle is lower as 16% and 22% than that of LO and SEO, potential of the Miller cycle to solve the contradiction of the power respectively. These results are obtained at 2800 rpm and emission density and controlling NOx limitation for PFI-H2ICEs with tur- of NOx limited below 100 ppm. The reasons for these findings can bocharger, particularly at a lower engine speed. When the NOx emissions are held at a constant value, the Miller cycle can increase be explained by two aspects of the followings. First, gMiller is larger the power density and decrease BSFC. than gOtto when the product of the turbocharger pressure ratio of the Miller cycle (dM) and cM is bigger than the turbocharger pres- sure ratio of the Otto cycle (dO) based on the analysis of the ther- 4. Conclusions modynamic cycle. With a similar pressure during the exhaust stroke, the higher intake pressure decreases the pumping loss. The Miller cycle not only shortens the compression, lengthens Therefore, reducing the pumping loss of the Miller cycle should the expansion stroke, and improves the thermal efficiency of PFI- be one of the most important contributors to the improvement HICEs but also significantly increases the power density using the of the fuel economy. Second, the temperature at the end of the turbocharging system. The test using Miller cycle with turbocharg- compression stroke of the Miller cycle is the smallest among the ing hydrogen engine should be done basic on the theory and calcu- three methods (about 17.5%–2800 r/min and 19.9%–5200 r/min lation results in the future. The above analysis leads to the smaller than other ways). This finding can imply that the NOx following conclusions: emissions of the Miller cycle are less than those of the Otto cycle, as suggested in Fig. 7(c). When the NOx emissions are limited 1. In the thermodynamic cycle analysis, gMiller is larger than gOtto, below 100 ppm, the equivalence ratio of the Miller cycle can be when dMcM is larger than dO. The additional pumping work in set larger, which in turn increases the output power. The heat the Miller cycle derives from the EIVC and turbocharging pres- transfer of the Miller cycle is smaller than that of the Otto cycle, sure ratio, which is the basic theory underlying the preceding which can be explained through the Woschni model [27]. analysis. The potential of the Miller cycle to increase the power The brake power of LO is the smallest, but its BSFC is better than density is determined by the turbocharging pressure ratio.2. To that of SEO. The method of SEO is at the stoichiometric equivalence avoid knock, 3.5 MPa is selected as the limitation of pressure at ratio burn, which can increase the output power. The exhaust gas the end of compression stoke. The IVC and inlet pressure of the Q.-h. Luo, B.-g. Sun / Energy Conversion and Management 123 (2016) 209–217 217

Miller cycle have bending relationship, which is limited by the [6] Dennis Peter A, et al. Performance of a port fuel injected, spark ignition engine pressure. The IVC of the Miller cycle is limited by the valve lift, optimised for hydrogen fuel. SAE technical paper no. 2012-01-0654 2012. [7] Rahman MM, Kamil Mohammed, Bakar Rosli A. Engine performance and and the inlet pressure is limited by the turbocharging system. optimum injection timing for 4-cylinder direct injection hydrogen fueled The two factors limit the Miller cycle coefficient, which deter- engine. Simul Model Pract Theory 2011;19:734–51. mines the boundary of the Miller cycle in increasing the power [8] Nakagawa Kenji, Yamane Kimitaka, Ohira Tetsuya. Potential of large output power, high thermal efficiency, near-zero NOx emission, supercharged, lean- density of hydrogen engines. burn, hydrogen-fuelled, direct injection engines. Energy Proc 2012;29:455–62. 2. Comparing three technical methods of turbocharging hydrogen [9] Verhelst S. Recent progress in the use of hydrogen as a fuel for internal engines, the Miller cycle way can get the higher brake power, as combustion engines. Int J Hydrogen Energy 2014;39(2):1071–85. [10] White CM, Steeper RR, Lutz AE. The hydrogen-fueled internal combustion 37.7% and 26.3% than those of LO and SEO, and lower BSFC, as engine: a technical review. Int J Hydrogen Energy 2006;31:1292–305. 16% and 22% than that of LO and SEO, respectively at 2800 r/ [11] Berckmüller M, Rottengruber H, Eder A, Brehm N, Elsässer G, Müller-Alander min. However, the benefits decrease with engine speed, espe- G, et al. Potentials of a charged SI-hydrogen engine. SAE technical paper no. 2003-01-3210 2003. cially the brake power only 2.8% higher comparing with SEO [12] Natkin RJ, Tang X, Boyer B, Oltmans B, Denlinger A, Heffel JW. Hydrogen IC at 5200 r/min. engine boosting performance and NOx study. SAE technical paper no. 2003- 3. When the emission of NOx is limited below 200 ppm, the brake 01-0631 2003. power of Miller cycle is higher as 34.9% and 25.8% than those of [13] Lapetz J, Natkin R, Zanardelli V. The design, development, validation and delivery of the Ford H2ICE E-450 shuttle bus. In: Proceedings 1st international LO and SEO, which is smaller than the values at 100 ppm. The symposium on hydrogen internal combustion engines, September 28–29, results show that the turbocharging Miller cycle is the best Graz, Austria. method of increasing power density and controlling NOx emis- [14] Gottschalk W, Lezius U, Mathusall L. Investigations on the potential of a variable Miller cycle for SI knock control. SAE technical paper no. 2013-01- sions in the limited boundary, particularly at a low engine 1122 2013. speed and lower emission of NOx. [15] Caton JA. The thermodynamic characteristics of high efficiency, internal combustion engines. Energy Convers Manage 2012;58:84–93. [16] Li Tie, Gao Yi, Wang Jiasheng, Chen Ziqian. The Miller cycle effects on improvement of fuel economy in a highly boosted, high compression ratio, Acknowledgement direct-injection gasoline engine: EIVC vs. LIVC. Energy Convers Manage 2014;79:59–65. [17] Hou Shuhn-Shyurng. Comparison of performances of air standard Atkinson The financial support of the National Natural Science Founda- and Otto cycles with heat transfer considerations. Energy Convers Manage tion of China (NSFC 51276019), the National High-Tech Research 2007;48:1683–90. and Development Plan (863 2006AA11A1B6) is greatly acknowl- [18] Sun Bai-gang, Zhang Dong-sheng, Liu Fu-shui. Cycle variations in a hydrogen internal combustion engine. Int J Hydrogen Energy 2013;38:3778–83. edged. And the authors would like to express their sincere grati- [19] Gürbüz Habib, Akçay Ismail Hakkı, Buran Dinçer. An investigation on effect of tude to Jing Wei of Beijing Institute of Technology for providing in-cylinder swirl flow on performance, combustion and cyclic variations in language assistance and proofreading. hydrogen fuelled spark ignition engine. J Energy Inst 2014;87(1):1–10. [20] Ricardo WAVE Theory Manual, version 2009. [21] Ricardo WAVE User Guide, version 2009. References [22] Das, et al. An experimental study of turbocharged hydrogen fuelled internal combustion engine. SAE technical paper no. 2015-26-0051 2015. [1] Verhelst S, Wallner T. Hydrogen-fueled internal combustion engines. Prog [23] Gürbüz Habib. Parametrical investigation of heat transfer with fast response Energy Combust Sci 2009;35(6):490–527. thermocouple in SI engine. J Energy Eng 1943:4016014. http://dx.doi.org/ [2] Verhelst S, Sheppard CGW. Multi-zone thermodynamic modeling of spark- 10.1061/(ASCE)EY.-7897.0000350. ignition engine combustion – an overview. Energy Convers Manage 2009;50 [24] Liu FuShui, Sun BaiGang, et al. Development of performance and combustion (5):1326–35. system of Atkinson cycle internal combustion engine. Sci China Technol Sci [3] Chintala V, Subramanian KA. Hydrogen energy share improvement along with 2014;57(3):471–9. NOx (oxides of nitrogen) emission reduction in a hydrogen dual-fuel [25] Soylu S. Prediction of knock limited operating conditions of a natural gas compression ignition engine using water injection. Energy Convers Manage engine. Energy Convers Manage 2005;46(1):121–38. 2014;83:249–59. [26] Luo Qing-he, Sun Bai-gang. The characteristic of polytropic coefficient of [4] Verhelst S, Maesschalck P, Rombaut N, Sierens R. Increasing the power output compression stroke in hydrogen internal combustion engine. Int J Hydrogen of hydrogen internal combustion engines by means of supercharging and Energy 2014;39:13787–92. exhaust gas recirculation. Int J Hydrogen Energy 2009;34:4406–12. [27] Woschni G. A universally applicable equation for the instantaneous heat [5] Lee Jongtai, Lee Kwangju, Lee Jonggoo, Anh Byunghoh. High power transfer coefficient in the internal combustion engine. SAE technical paper no. performance with zero NOx emission in a hydrogen-fueled spark ignition 670931 1967. engine by valve timing and lean boosting. Fuel 2014;128:381–9.