Numerical and experimental analysis on the effects of turbocharged compressed bio-methane-fueled automotive spark-ignition engine

This work focuses on the numerical and experimental analysis of turbocharger selection and boost pressure effects on a CNG-fueled spark-ignition engine. Because of this, investigations are carried out on the influence of downsized compression ratio of 10.5:1 at different boost pressures and compared with a naturally aspirated compression ratio of 12.5:1. In order to perform the experimentation, a twin-cylinder, port fuel-injected, CNG engine with 15.5 kW at 3400 rpm is modified to utilize compressed bio-methane as fuel under 100% throttle condition. A simulation is performed to study the compressor impeller for T1 and T2 turbochargers using the ANSYS turbomachinery tool. Results indicate that the circumferential velocity of T1 is higher than of T2 at all boost pressures. Subsequently, experimentation is performed using T1 and T2 at three different boost pressure levels in a compression ratio of 10.5:1 at 1.1, 1.3, and 1.5 bar. T2 developed a maximum boost pressure of 1.1 bar compared to T1. T1 is chosen for further experimentations. At 1.3 bar of boost pressure, a rise in brake power was recorded by 19.3% compared to 12.5:1 under the naturally aspirated mode. Consequently, there is a reduction in fuel consumption by 10.1%, and hydrocarbon, carbon monoxide, and carbon dioxide emission levels reduce by 25%, 8.2%, and 4.9%, respectively. Therefore, turbocharging at a lower compression ratio exhibits better performance and reduces emissions compared to a higher compression ratio under naturally aspirated mode.


Introduction
In the developing countries with a higher rural population, there is a motive to improve energy security (Muhumuza et al. 2018), to reduce greenhouse emissions concerning climate change (Huppmann et al. 2019), and to meet the norms of the United Nations sustainable development goals (Vu et al. 2015). Petroleum-based fuels have a greater effect on the climate which can lead to greenhouse emissions; therefore, minimizing carbon in the atmosphere has caused dependability on renewable energy resources (Ray 2019). Concerning these factors, the usage of dedicated alternative energy resources such as bio-methane has been realized for internal combustion engines. Further, the utilization of waste recovery management and the production of cleaner energy are advantageous also. The anaerobic decomposition of biomass yields biogas which, volume-wise, consists of 60% methane (CH 4 ) and 40% carbon dioxide (CO 2 ), and by upgrading the biogas, the possibility of obtaining biomethane can be explored. The bio-dependable resources such as food waste and cattle manure waste hold a good agreement for biogas production (Abbas et al. 2020). Also, in the up-gradation process, several advanced methods have been implemented such as thermochemical recovery through exhaust gas reformation (Lau et al. 2012), membrane method (Molino et al. 2013), and concurrent elimination (Tippayawong and Thanompongchart 2010) of the contaminants such as CO 2 and H 2 S packaged through media column reactors (Chandra et al. 2011). Various sources of bio-recyclable materials to transform to biogenic methane, a refined fuel through the anaerobic digestion process, and different advanced methods of CO 2 capture techniques are available (Heesterman 2019).
The following literature represents the engine experimentation using bio-methane as fuel which describes the merits and demerits toward performance. The performance of the engine is estimated using biogas and bio-methane as fuels. A well-to-wheel review is performed to determine the potential of natural gas and bio-methane in terms of GHG (greenhouse gas) emission reductions; besides, the outcomes were compared with gasoline (Bordelanne et al. 2011). Methane is the primary surrogate of bio-methane neglecting the 2% CO 2 and all fossil fuels (Leiker 1972), with a lower carbon to hydrogen ratio. Methane (CH 4 ) possesses a higher autoignition temperature, higher flame velocity, and higher antiknock property (Weaver 1989;Turns 1996) than other fossil fuels; therefore, this favors an increase in compression ratio. Table 1 shows the properties of different fuels.
The investigation and comparison of methane with gasoline show non-uniformities in the air-fuel mixture that surrounds the spark plug, and this evolves to randomness in motion of the mixture at the vicinity of the spark source during ignition (Amann 1985). Wong (1977) conducted investigations on the performance characteristics of methane and gasoline and observed a drop in brake power and fuel consumption for methane compared to gasoline. Arroyo et al. (2013) examined and compared reductions in brake mean effective pressure for methane and gasoline fuels. The emissions such as HC, CO 2 , CO, and NO exhibit reduced levels for methane fuel compared to gasoline for various engine speeds at wide-open throttle and 50% opening. (Thurnheer et al. 2009) compared combustion characteristics for the engine operation with gasoline at stoichiometric air-fuel mixture, at 2000 rpm speed and brake mean effective pressure of 2 bar and showed that the combustion duration of methane was shorter than that of gasoline.
A review of previous literature, in terms of engine performance and combustion parameters, enhancement of the bio-methane fuel composition becomes necessary for determining optimum fueling for the spark-ignition engine. As upgraded biogas fuel results in the formation of compressed bio-methane gas, the composition of biogas upon engine performance must be investigated. To determine engine operational parameters similar to bio-methane experimentation, biogas is used as fuel and engine experimentations are conducted to improvise the engine performance, emission, and combustion characteristics. (Porpatham et al. 2008) studied the effect of reduced concentrations of CO 2 in biogas fuel to ascertain the performance, emissions, and combustion levels. Engine performance increased significantly and hydrocarbon emissions registered a reduction while operating the engine at lower levels of CO 2 -composed biogas. At lower CO 2 levels, the lean operation limit showed an enhancement. On the other side, the biogas demonstrated lower heating value and laminar flame velocity as compared to natural gas because of the lower content of methane and the presence of CO 2 results in a reduction in power output and lean operation limit (Chandra et al. 2011;Cardona and Amell 2013;Hinton and Stone 2014;Raju 2001).
Technical updates, such as the addition of hydrogen, enhancement of compression ratio, and turbocharging the engine, have also proven to be useful in maximizing the overall engine performance for the biogas-fueled engine. Jung et al. (2015) studied biogas-fueled turbocharged engine performance and oxides of nitrogen emissions were studied at different biogas compositions, namely 50% methane and 50% CO 2 and 70% methane and 30% CO 2 . To obtain a brake power of 22.5 kW, the CH 4 :CO 2 ratio of 50:50 requires a boosting of 1.14 bar pressure at a relative air-fuel ratio of 1.3. The boosting is more than the range of 1.0-1.5 for the relative air-fuel ratio of 1.7, 1.9. For CH 4 :CO 2 = 70:30, the reference brake power of 24.8 kW was attained and the boost pressure levels of 1.13, 1.27, and 1.44 bar at the relative air-fuel ratios 1.3, 1.5, and 1.7 respectively. The brake power increases with an increase in boost pressure for both cases at the relative air-fuel ratio of 1.1, thereby denoting the stoichiometric mode. Several studies conducted for investigating full-load performance with a turbocharged engine showed a significant rise in engine torque which was double compared to fully open wastegate mode. There is an increase in the incylinder pressures at turbocharged mode, reduced unburnt HC, and CO emissions (D'Ambrosio et al. 2006), and a better power to weight ratio (Einewall 1997). This has led the automotive manufacturer to introduce downsizing concepts that optimize a suitable compression ratio necessary for turbocharging which offers enhanced charge cooling, knockfree operation, and effective scavenging of the residual gases to improve low-end torque (Shivapuji and Dasappa 2014;Padmavathi and Nandhakumar 2010). Selection of the turbocharger required for the engine becomes essential and an extensive study of the compressor and a turbine map must be done based on the simulated and experimental results of  (Gumus 2011;Qi et al. 2016;Porpatham et al. 2013;Koonaphapdeelert et al.  the turbomachinery performance. The turbocharger housing unit's design affects the airflow rate at the compressor inlet and exhaust backpressure development and flow rate over the turbine. The turbocharger's compressor housing size determines the magnitude of airflow rate, which is required at wide ranges of the engine speed (Mataczynski et al. 2016). The critical element of achieving turbomachine efficiency depends on scroll design. The two main design parameters of the scroll that affect the flow properties are the cross-sectional area (A) and the centroid radius (r) from the rotational center of the turbomachinery axis (Niemi and Laurén 2002).
To select the right turbocharger, the airflow rate consumption is estimated based on the engine power requirements, and then, it is identified on the compressor performance map to determine the range of pressure ratio concerning corrected mass flow rate (for engine experiments actual airflow rate is considered), lines of turbocharger rotational speed, efficiency islands, surge and choke boundaries (Mohan et al. 2019). The matching process can be further analyzed based on the turbocharger compressor, turbine, and engine performance, such as intake pressure and temperature, in-cylinder pressure, indicated work (Shivapuji and Dasappa 2014). Methods such as the first turbocharger equation are also employed to determine the turbocharger selection process which correlates the performance of the compressor, turbine, and engine (Nguyen-Schäfer 2012). The present work aims to develop a low displacement automotive SI engine capable of operating at a lower compression ratio using CBM as fuel with improved performance and reduced emissions by incorporating a suitable turbocharger. The specific objectives of this work are to improve the performance through intake charge and combustion system modifications. To improve the concentration of methane in biogas, calcium oxide was used to remove the CO 2 . Further, the naturally aspirated engine was modified to operate the engine using a turbocharger at a lower compression ratio. The numerical study was carried out to compare the performance of the two turbocharger compressor impeller at different boost pressures. The best turbocharger was identified based on the pressure ratio and compressor efficiency. To compare the numerical and experimental study, the experimental study was carried out using two turbochargers at different boost pressures at a reduced compression ratio. The experimental study was also carried out at a higher compression ratio under naturally aspirated mode. The performance, emission, and combustion characteristics were compared at higher and lower compression ratios.
The novelty of this work is in the application of turbocharging the naturally aspirated engine and study its effect on different boost pressures using compressed bio-methane (CBM) as fuel for a small displacement engine having less than one-liter capacity. The power utilized for driving the turbine is obtained from the waste energy that is available through the engine exhaust gas. The experimental results suggest that turbocharging at a lower compression ratio with an optimized boost pressure results in superior performance compared to engines operating at a higher compression ratio under naturally aspirated mode. This yields a higher power to weight ratio and hence an increase in mileage and reduction in CO 2 emissions.

Experimental setup
The experiments are conducted on a TATA Ace CNG twincylinder engine, and a naturally aspirated engine was converted to a turbocharged engine. Table 2 depicts the engine technical specifications. Figure 1a demonstrates the experimental setup layout. By modifying the thickness of the cylinder head gasket, the compression ratio is revised. The "Appendix" gives the calculation for the compression ratio. The piston crown has a flat geometry, and the bowl configuration is provided at the upper region of the combustion chamber. This provides adequate swirl and tumble motion of the fuel mixture. The details of the shape and dimensions are shown in Fig. 1b.
The inline twin-cylinder has top dead center (TDC) offset to each other by 180 degrees, which facilitates different valve timings. The firing interval is short between each cycle compared to a multi-cylinder engine. The valve timing configuration aims at attaining maximum volumetric efficiency and scavenging of exhaust gases, with better flow of the medium through the split design of the intake and exhaust port minimizing shocks.
The engine load is varied using the eddy current dynamometer (Make: Dynalec) and is coupled to a starter motor for motoring or cranking. The engine is modified to a turbocharged version from the naturally aspirated system suitable for investigating various gaseous fuels. The smaller version turbocharger is used for turbocharging analysis. An electric wastegate actuator is used to control the exhaust gas mass flow rate to the turbine. A K-type thermocouple measures the exhaust gas temperature at the exhaust manifold after the turbine outlet. Woodward electronic control unit could alter the spark timings. The port fuel injectors are replaced by LPG/CNG gas injectors having working pressure up to 10 bar. The fuel flow rate is determined by the thermal mass flowmeter (Make: FOX). The airflow is determined using an airflow meter (Make: FMG). The intake line consists of an air drum for damping the airflow. An intercooler is placed after the compressor outlet for cooling the high-temperature compressed air. The in-cylinder pressure is evaluated by a piezoelectric pressure sensor (Make: KISTLER) with a resolution of up to 0.1 crank angle degree. The manifold absolute pressure and temperature sensor (Make: BOSCH) are mounted before and after the compressor for measuring the intake manifold temperature and pressure. A 0.1 resolution crank angle encoder is coupled to crankshaft and synchronizes with a data acquisition system and acquires data for 100 combustion cycles. The equivalence ratio is calculated by employing an oxygen sensor, and a manual feedback loop is used to control the injection system. A five-gas analyzer was employed to measure the exhaust emissions (Make: Horiba) which primarily measures the HC, CO, NO, and O 2 levels.
The bio-methane production is from the fine refinement of the biogas (Koonaphapdeelert et al. 2019). The biogas is generated in a floating dome reactor where anaerobic digestion of food waste and cow dung takes place. The food waste used for the generation of biogas consists primarily of cooked rice along with vegetable peelings and wheat products. The main objective of the utilization of biogas in the SI engine is to study the performance and emissions. A separate CO 2 scrubber unit is utilized for biogas up-gradation which separates CO 2 from the biogas composition at desired ranges. Calcium oxide is mixed with water to obtain calcium hydroxide, which is stored in the scrubber unit, and subsequently, biogas is bubbled through it. The reaction of calcium hydroxide with the CO 2 present in the biogas leads to calcium carbonate formation.
The scrubbing process was done repeatedly till the CO 2 concentrations is attained to the desired levels. The concentrations of methane and CO 2 are derived from a non-dispersive infrared analyzer. The composition of biogas attained is 84% of methane, 13% of CO 2 , and some minor traces of CO and other contaminants. The gas is then taken as a lowpressure inlet from the CO 2 scrubber and is fed as the inlet to the high-pressure booster pump, which stores the highpressure horizontal gas tank's bio-methane gas through the multivalve arrangements. The injection pressure is regulated to 2 bar and delivered to the port fuel injector common rail unit.
Initially, experimentations are conducted with a naturally aspirated system at a compression ratio of 10.5:1, and then, turbocharging was performed at 1.1, 1.3, and 1.5 bar boost pressures, respectively. Then, the experiment is conducted at naturally aspirated mode at a compression ratio of 12.5:1. Operating of the engine done at speed at full-throttle conditions ranging from 900 to 3400 rpm and fueled with compressed bio-methane. The outcome of this research can be realized in commercial vehicles for transportation purposes.

Simulation of the compressor impeller
Initially, the simulation process is carried out in ANSYS Turbomachinery software corresponding to different pressure ratios for T1 and T2 concerning turbocharged engine operating boundary conditions. The compressor parameters are analyzed and compared to select the suitable turbocharger. Prediction of the compressor aerodynamic performance for T1 and T2 is initially studied by conducting a steady-state numerical simulation in ANSYS turbomachinery software.

Determination of compressor impeller parameters
The objective of the simulation experiments is to evaluate and compare various compressor performance parameters between T1 and T2 at the respective boost pressures. Preliminary calculations are done based on meeting the targeted engine horsepower along with the study conducted upon the compressor variables using the compressor map. The actual mass flow rate and pressure ratio of air are determined from the relations (Garrett 2019) as given in Eqs. (1), (2), and (3), considering the non-adiabatic conditions of the turbocharger. In the simulation's initial process, the design details and boundary conditions of the compressor are specified to generate the compressor impeller model.
where MAP req = manifold absolute pressure (kPa) required to meet the horsepower target, W a = airflow rate actual (kg/s), R = gas constant (639.6), which is 287 J/kgK, T m = intake manifold temperature (K), Rankine to Kelvin scale, V e = volumetric efficiency, N = engine speed (rpm), V d = displacement volume (m 3 ) where Π c = pressure ratio, p 2c = compressor discharge pressure (kPa), p 1c = compressor inlet pressure (kPa), MAP req = manifold absolute pressure (kPa) required to meet the horsepower target, Δp loss = Pressure losses due to the air filter, intake manifold, and intercooler (kPa), p amb = ambient pressure (kPa).
The compression work is generally a polytrophic process accompanied by friction and losses in the compressor, which leads to entropy generation. The compressor efficiency is analyzed based on the isentropic conditions and in it no heat transfer takes place. However, during the turbocharger's actual engine operating conditions, the compressor is required to compensate for the losses through additional work. The total-to-total compressor isentropic efficiency is determined (Nguyen-Schäfer 2012) using Eq. (4) represented below, which denotes the energy conversion from the kinetic energy at the impeller to the pressure head at the diffuser.
w h e r e c = Compressor is entropic efficiency, p 2c = Pressure at the compressor outlet (kPa) , p 1c = Pressure at the compressor inlet (kPa) , T 2c = Temperature at the compressor outlet (K) , − 1 T 1c = Temperature at the compressor inlet (K) , k = Ratio of specific heat capacity.
The following relation describes the circumferential velocity for a compressor impeller. As per the Euler turbine Eq. (5), since the enthalpy ( h t ) of the air medium is higher, at the impeller exit results in higher circumferential velocity ( C u2 ) which is higher for T1 compared to T2. This can be determined using the following Eq. (5): where h t = specific enthalpy (kJ/kg), h t2 = specific enthalpy (kJ/kg), = turbocharger speed (r/s), r 2 = radii (m), C u2 = circumferential velocity (m/s).

Geometric modeling of the compressor impeller
The geometric dimensions of the compressor impeller show similar configurations for T1 and T2 based on the inducer diameter which assembles to the shroud of the casing.
The key difference is that the compressor scroll casing is smaller for T1 compared to T2, and in some cases, there are also variations in trim ratios. The photographic view of various compressor impeller components such as diffuser and impeller is shown in Fig. 2a, c.
In the pre-processing mode, the overall turbocharger operating ranges and design geometry details are given as input in the Vista CCD, and the compressor impeller model is generated. The model is imported in ANSYS Turbogrid tool, and the mesh analysis is performed. The two-dimensional hexahedral mesh with a total node of 304,530 and a total element of 274,355 is generated. The mesh view of the compressor impeller is shown in Fig. 2b. The computational fluid dynamics (CFD) analysis is performed for the compressor impeller. The overall pressure ratio of 1.5, speed of 175,000 rpm, and the mass flow rate of 0.03 kg/s are derived from the calibrated values of the compressor maps for T1 and T2. The average inlet temperature is considered between 301 and 312 K, and the atmospheric pressure of 101,325 Pa is considered. The compressor's inducer and hub diameter showed similarity for both T1 and T2 and are 25 mm and 9 mm, respectively. The compressor has five main radial compressor blades along with splitter blades. The only variations accounted for in the simulation were the air mass flow rate for T1 and T2 based on different engine boost pressures at its corresponding turbocharger speed. Considering the compressor scroll domain, the actual simulation input data are given in terms of analytically calculated inlet airflow rate, which differs individually for T1 and T2, respectively.

Simulation run parameters
The nature of the study is an analysis of a steady-state fluid medium flow through the compressor impeller, which involves energy transfer from kinetic energy to static pressure. The fluid model used for the simulation was turbulence shear stress transport. The solver controls used are high-resolution first order, and convergence is set to a residual target of 0.0001.

Mesh independence study
The mesh independence study for T1 is illustrated in Fig. 2g, denoting pressure ratio with actual airflow at three different ranges of mesh elements. A mesh independent study is performed and comparisons are made by meshing the generated model using coarse, medium, and fine elements having a total count of 92,483, 169,572, and 431,576, respectively, from which the convergence of solution can be determined. The mesh independence study for T2 also showed similar results. Table 3 tabulates the instruments employed for the experimentation purpose. The measuring instrument's uncertainty is evaluated by the Gaussian distribution method with an error range of ± 2r (Moffat 1988). The accuracy and uncertainty of the instruments used for measurement are listed.

Estimation of laminar flame speed
The numerical simulation is carried out in the Chemkin Pro software package to analyze the flame speed for the CNG and compressed bio-methane. Initially, input files related to methane combustion are used for the simulation studies. The boundary conditions are given for the unburnt gas such as pressure and temperature. The mesh details are provided for the flame model; a total number of 5000 grids are deployed for meshing for the flame length of 0.0155 m. The entire combustion analysis is done in a stoichiometric air-fuel ratio. The species given as input is methane, and the oxidizers are oxygen and nitrogen. The combusted product is given as carbon dioxide, water, and nitrogen. Then, the solver is defined as having a relative tolerance of 0.0001. The results of the laminar flame speed obtained for CNG and compressed bio-methane are 34.8 cm/s and 34.4 cm/s, respectively. Table 1 depicts the obtained results.

Numerical analysis of compressor impeller
With reference to the compressor map, the targeted boost pressure necessary for the twin-cylinder spark-ignition engine is to be optimized. Based on the speed lines, T1 is set to three speed ranges: 58%, 80%, and 100% of the overall compressor speed of 175,000 rpm, and attains a maximum pressure ratio of 1.5. In contrast, for T2, the maximum possible would be within the 1.1 pressure ratio that is obtained at a speed of 57% of the overall speed. Design points are created from the boundary conditions, and the iteration process is set to simulate the compressor performance. Figure 2d, e shows that the numerical and simulation data are superimposed on the original manufacturer compressor maps (Mataczynski et al. 2016) for T1 and T2. The results indicate that T1 attained a maximum of 1.5 pressure ratio which has 1.1 (1.1953, 1.1858, 1.1950, 1.2034, 1.1904, 1.1727), 1.3, and 1.5 bar range of pressure ratio, whereas T2 attained at 1.1 (1.18480, 1.18070, 1.18350, 1.18850, 1.18890, 1.1743) pressure ratio values is lesser compared to T1. At a 1.1 pressure ratio, the peak compressor efficiency is higher for T1 as compared to T2 at the higher airflow rate. The higher airflow rate is superior for T2 as shown in the variation in slope between T1 and T2 efficiency due to flow instabilities and thermal losses.
In the experiment at regions of lower air mass flow rate, at the minimum engine speed of 900 rpm the turbo lag occurs and the trend ascends to a constant range due to an increase in turbine power, and hence, constant turbocharger rotational speed is kept through the wastegate actuator. In the simulation, the starting region of the air mass flow rate ignores this turbo lag because the given input turbocharger rotational speed is the same for all points. Figure 2f shows the compressor efficiency with actual airflow rate for T1 and T2 for simulation and experimentation conditions. Experimental results show that at the greatest brake thermal efficiency region (at 0.012 kg/s) the airflow rate, the compressor efficiency increased for T1 by 2% compared to T2, and also at higher airflow the values were higher for T1. In the simulation at regions of higher airflow rate, the compressor efficiency of T1 surpasses T2 because of the higher circumferential velocity of T1. In the simulation experiment, the speed of the impeller is the only input parameter that decides the performance between T1 (101,500 rpm) and T2 (100,000 rpm) at a 1.1 pressure ratio.
In the case of the scroll geometry, large stall occurrence leads to a less peak pressure ratio for T2. The factor impacting the compressor efficiency and stability of the flow depends on compressor volute design that varies in size individually for both T1 and T2 and significantly influences the axisymmetric flow pattern from the compressor inlet to the diffuser region. Studies claim that there are significant influences upon flow distortions with respect to variation in the A/r ratio. It is revealed that the reduction in the slope of the A/r ratio of the volute resulted in minimized flow distortions. Similarly, this also has resulted in a higher total pressure ratio and peak compressor efficiency (Sun et al. 2019). The volute (scroll) of T1 reveals a smaller A/r ratio that accommodates smaller volumes of air and develops a higher-pressure head than T2, which typically shows a larger A/r ratio. Therefore, the lower mass flow rate of air T2 scroll has an accelerated flow, resulting in reduced total momentum having a final uniform velocity less than T1 at the diffuser region. Distortions also influence the additional variations in flow due to bent pipes during real-time engine experimentation. The compressor efficiency becomes inversely proportional to the pressure ratio. The simulation shows that at the 1.3 pressure ratio, the compressor efficiency for T1 is 2.6% higher than 1.5. The energy produced per stage triggers a rise in temperature with an increase in pressure during the compression process. At a greater pressure ratio, the temperature ratio increases, leading to heat losses through the compressor scroll. Research investigation states that heat transfer occurs from the turbine region to the compressor region, causing an increase in compressor outlet temperature and this leads to compressor efficiency deterioration (Rautenberg and Malobabic 1984;Shaaban 2006). The incylinder temperature and exhaust gas temperature reduce with an increase in boost pressure due to lean combustion, although the compressor outlet temperature increases due to heat transfer taking place from turbine work (total enthalpy controlled by a wastegate, mass flow rate led to turbine increases with an increase in turbocharger speed), bearing friction, and compression work. The overall heat transfer occurring from the exhaust gas is based on turbine work, frictional work, and compressor work (Burke et al. 2015).
The numerical results of the blade-to-blade view flow distributions comparing the circumferential velocity components for T2 at 1.1, T1 at 1.1,1.3, and 1.5 pressure ratios are depicted in Fig. 3a- T1 attains a maximum of 1.2% higher circumferential velocity compared to T2 at a 1.1 pressure ratio. This is also denoted in Eq. (5) which describes the circumferential velocity parameter of the compressor impeller. The specific enthalpies (kJ/kg) are denoted by ht available at inlet and outlet of the impeller, radii (m) at the impeller outlet, the circumferential velocity of air given as (m/s) at the circumferential direction of the impeller outlet, and the turbocharger speed is denoted by (r/s). The higher enthalpy causes more work transfer on the air medium by the compressor's impeller blades, leading to the apparent increase in pressure distribution that can be seen from T1 at the impeller exit (Zhao et al. 2020). Figure 3e-h shows the numerical results of meridional pressure view distributions for T2 at 1.1, T1 at 1.1,1.3, and 1.5 bar pressure ratio. The static pressure rise at the inducer tip denotes the pressure rise of the impeller for T1 by 1% compared to T2 at a 1.1 pressure ratio. At 1.3 bar and 1.5 bar pressure, the ratio T1 is higher by 15% and 30.8%, respectively. The pressure rises primarily due to higher circumferential velocity at the impeller exit for T1 and the minimum recirculation of flow that tends to lower the pressure generation. This is more common in T2, due to the flow rate of air happening along with the larger scroll. In contrast, the flow in T1 results in developing higher pressure where the compressor scroll matches engine volume. The compressor efficiency deviation is due to flow acceleration instabilities, flow at sonic velocity, and entropy generation.
The turbocharged engine at lower engine speed is known to have higher levels of residual gas and tends to stall or surge the turbocharger compressor. In the experimental approach at a lower engine speed of 900 rpm, the operating regions for T1 seem closer to surge margin within stable limits at all boost pressures. This parameter also shows improvements toward low-end engine torque. From the numerical simulation, more significance of surge is observed while appearing closer and beyond the margin because of the constant speed input given at each mass flow rate.
At the region of 1.1 bar pressure ratio, T1 develops higher boost pressure much earlier from lower airflow rate (lower engine speed) and lasts throughout compared to T2. At higher airflow rate (higher engine speed) regions, the T2 compressor leads to a choke margin due to the higher generation of volume flow. This is also evident from the actual turbocharger engine experiments performed with T2. Even at fully closed wastegate conditions, the maximum boost pressure yielded is only 1.1 bar when the engine achieves the speed of 3,400 rpm. Simultaneously, T1 gives a better transient response and greater boost pressure development up to 1.5 bar than T2.
In order to optimize the boost pressure for T1, the stable operation for T1 is at a 1.3 bar pressure ratio, which has 80% of the turbocharger's overall speed and is equivalent to a 140,000 rpm speed line. At this operating region, T1 has a safe surge at the region of lower airflow rate. It nears the efficiency island region and away from the choke region at a higher airflow rate, suggesting an improvement in the higher-pressure ratio. Overall, the scroll design had a significant effect on compressor performance characteristics. Overall, we can conclude that T1 shows better compressor performance than T2 and is suitable for smaller engines. Also, a T1 turbocharger is preferable to a 1.3 pressure ratio, which is necessary for optimal engine performance.

Testing of turbocharger on the engine
The engine experimentations at full-throttle conditions are performed at naturally aspirated mode with a port fuel injection of CBM fuel that forms the base readings. Subsequently, turbochargers T1 and T2 are tested on a twin-cylinder SI engine, to determine the range of the boost pressure. Table 4 shows the details of the turbochargers T1 and T2.
The variations of boost pressure with engine speed are illustrated in Fig. 4a. The engine speed ranges from 900 to 3400 rpm. The T1 turbocharger develops a boost pressure much earlier with a partial closure of the wastegate actuator than T2. The T1 attains a boost pressure of 1.1 bar at full closure of wastegate actuator position (1%) at lower engine speed originating from 1200 rpm engine speed and has an overall superior performance than T2. Further, the maximum boost pressure achieved during the experimentation is 1.5 bar, much below the knocking range. The boost pressure is kept constant using the wastegate actuator at all engine speed. An increase in boost pressure leads to higher engine torque and lower brake-specific energy consumption and emissions. The increase in boost pressure under turbocharged conditions leads to leaner operation and improves engine torque and minimizes the emissions (Iyer et al. 2011). In comparison, T2 has a slower response even at fully closed wastegate conditions toward boost pressure development and attains a maximum of 1.1 bar pressure at a maximum speed of 3400 rpm. As aforementioned, the smaller A/r ratio of the compressor scroll for the T1 turbocharger results in greater development of boost pressure than of volume flow rate.
For the naturally aspirated engine, the intake airflow becomes an essential factor that denotes the volumetric efficiency and increasing the brake mean effective pressure. Brake mean effective pressure is related to volumetric efficiency, and turbocharging causes an increase in air density and this increases the brake mean effective pressure  (Heywood and Welling 2009). The variation of volumetric efficiency as a function of engine speed is shown in Fig. 4b. At all speeds of the engine, the volumetric efficiency for T1 is higher in comparison to T2 at 1.1 bar boost pressure. The smaller compressor scroll and turbine housing of T1 needs less energy of turbine and impeller work to generate the boost pressure than mass flow rate of air as compared to T2. At maximum engine torque and brake thermal efficiency region of 1800 rpm engine speed, the volumetric efficiency increases for T1 at 1.1, 1.3, and 1.5 bar boost pressure by 16.7%, 23.5%, and 40% and for T2 at 1.1 bar by 5% compared to the naturally aspirated mode at 10.5:1 compression ratio. Turbocharging induced improved volumetric efficiency increases the brake mean effective pressure as compared to the naturally aspirated mode. Figure 4c demonstrates the variations of the wastegate actuator opening with engine speed. At lower boost pressures, the opening of the wastegate is minimum and increases with the rise in engine speed. At higher boost pressures beyond 1.3 bar, the opening of the wastegate actuator discharges more exhaust gas at regions of higher engine speed. For developing a boost pressure of 1.1 bar, T1 shows more wastegate opening compared to T2 which is due to the smaller volume of the turbine housing of T1. The indicated mean effective pressure increases with boost pressure because of better fuel conversion efficiency. In some cases, due to flow instability, the gas work transfer to the piston, which must be available to the driveshaft, is not completely performed. The work not utilized, also known as friction work, also affects the development of boost pressure specifically while turbocharging through all engine factors. The pumping work influences the flow of the working gases at the intake and exhaust manifold. For T1, it becomes necessary to optimize the boost pressures among three pressure levels to maximize flow stability. With the increase in engine speed toward attaining maximum brake power, the flow gets choked at the valve seat and achieves the sonic speed. Here, it becomes more significant in the exhaust manifold of T1 where the pumping work becomes dependent on the exhaust volume, including the port volume, manifold, and turbine volute.
Being a smaller turbine volute for T1, boost pressure rise beyond 1.3 bar causes an increase in pumping mean effective pressure and the flow tends to choke when boosting attains 1.5 bar pressure. Since the engine has a two-cylinder with a smaller exhaust volume the firing interval is larger compared to a multi-cylinder engine. Therefore, the exhaust volume increases during the blowdown phase of the exhaust stoke and this leads to higher pumping mean effective pressure (Kellermayr et al. 2019). Among all the boost pressures, 1.5 bar exhibits higher frictional mean effective pressure, signifying the total frictional work. The significant losses are in pumping work on the fluid medium since the exhaust pressure is relatively higher than the intake. This shows that T1 at 1.3 bar boost pressure is optimal for the overall performance of the engine. Overall, the results show that T1 performance is better than T2.

Engine performance, emission, and combustion parameters
The experimentation describes the comparative study between turbocharging using T1 at a reduced compression ratio of 10.5:1 and a naturally aspirated compression ratio of 12.5:1 optimized compression ratio for CNG fuel, as per the literature for the spark-ignition engine fueled with compressed bio-methane with a composition of CH 4 :CO 2 = 84:13. Discussion is made for various parameters associated with engine performance, emission, and combustion characteristics for both compression ratios under turbocharged and naturally aspirated mode. Figure 5a illustrates the changes in engine torques with speed at different compression ratios. The increase in boost pressure leads to an increase in engine torque at a compression ratio of 10.5:1. At boost pressure of 1.1, 1.3, and 1.5 bars, the resultant torques obtained are 47.3, 52.3, and 57.3 Nm, respectively. This is more than that of the naturally aspirated compression ratio of 12.5:1 which has a maximum engine torque of 44.3 Nm. Allowing more air in the engine as a result of increasing the boost pressure results in increased levels of engine torque. The experimental analysis states that the effect of increasing the boost pressure with full closure of wastegate generated an increase in torque, better fuel conversion efficiency, and reduced carbon dioxide equivalent emission (D'Ambrosio et al. 2006). The minimization of mechanical and thermal loading at a higher compression ratio of 12.5:1 is feasible through turbocharging at a compression ratio of 10.5:1. It is noteworthy that higher the boost pressure beyond 1.5 bar leads to engine knocking. The advantage of a rise in boost pressure leads to an enhanced in-cylinder pressure closer to TDC. It is known that gasoline engine is limited to higher range boosting beyond 1.7:1 pressure ratio due to knock influenced combustion. An increase in brake mean effective pressure can be achieved by the intercooler, retarding the ignition timing and reducing the compression ratio (Watson 1988). Figure 4d illustrates the variation of knock ratio with combustion cycle for T1 at different boost pressures. The values tend to spike up to 3 with 1.5 bar boost pressure due to secondary spontaneous combustion occurring, especially at higher engine speed regions. Therefore, it results in knocking of the engine due to an increase in boost pressure. Hence, the wastegate must be set to the optimal level of boosting to minimize mechanical vibrations due to torque fluctuations. Higher boosting and abnormality in combustion lead to unstable combustion, which causes an increase in the COV of IMEP. Figure 5b depicts the variation observed in brake power with engine speed. The effect of increasing boost pressure leads to enhanced brake power. Brake power rises by 19.3% at a compression ratio of 10.5:1.
The formation of a homogenous mixture and higher charge density with a rise in the boost pressure led to the increase of brake power. The downsizing effect was realized at a compression ratio of 10.5:1 since the brake power obtained on the onset of turbocharging is more than or equivalent to the naturally aspirated compression ratio of 12.5:1. A minimum of 1.1 bar boost pressure is necessary to attain more than the reference brake power at a naturally aspirated compression ratio of 12.5:1.
The reduced compression ratio of 10.5:1 at 1.3 bar boost pressure has brake thermal efficiency of 26.3%, whereas at 12.5:1 it was 24.8% under naturally aspirated mode. This is also seen from the Otto cycle, where the increase in boost pressure corresponds to an increase in compression ratio (Jung et al. 2015). Figure 5d shows the variation of brakespecific energy consumption with speed. The increase in boost pressure reduces the energy consumption from 14.5 to 13.7 MJ/kWh at the above-mentioned operating conditions. However, it is observed that at 1.5 bar boost pressure the energy consumption is lower. By considering the safety aspects of the engine, 1.3 bar boost pressure is chosen. This signifies the maximum energy conversion efficiency through the utilization of fuel energy available because heat energy and the exhaust gas recovery through turbocharging lead to lesser energy consumption for turbocharging at a compression ratio of 10.5:1.
The variation of HC emission with speed is shown in Fig. 5e. There was a significant decrease in HC emission while increasing the boost pressure because the lower carbon to hydrogen ratio of CBM and more admittance of air altogether resulted in minimizing the HC emission. The greatest reduction in HC emission is recorded at a compression ratio of 10.5:1 with 1.5 bar boost pressure (60% reduction than naturally aspirated compression ratio of 12.5:1). At the maximum brake thermal efficiency region, the optimized boost pressure of 1.3 bar shows reduced levels of HC emission by 25% compared to the naturally aspirated compression ratio of 12.5:1. The effect of engine performance with increase in the boost pressure raises the brake thermal efficiency and reduces HC and CO emission with the onset of an increase in oxides of nitrogen emission (Singh and Sandhu 2021). The variation of CO emission as a function of engine speed is shown in Fig. 5f. The effect of boosting leads to increased air levels in the mixture; therefore, reduced CO emission levels result in better oxidation of CO molecules to CO 2 during the combustion process. At the maximum brake thermal efficiency region, the CO emission lowered by 8.2% while boosting at a compression ratio of 10.5:1 with 1.3 bar boost pressure compared with a naturally aspirated compression ratio of 12.5:1.
The variation of CO 2 emission regarding engine speed is depicted in Fig. 5g. The levels of CO 2 emission may increase with the increase in boost pressure when compared to naturally aspirated mode. This indicates the occurrence of complete combustion due to the greater presence of air necessary for complete combustion. This signifies a reduction in greenhouse emissions and their effect on the environment.
The variation of NO emission is illustrated in Fig. 6a. The nitrogen oxide (NO) levels substantially increase with boosting pressure due to better combustion. The decrease in NO level at a compression ratio of 10.5:1 with 1.5 bar boost pressure is due to retarded spark timing to avoid knocking at higher boost pressure. The retarded spark timing leads to lower peak gas temperature. Also, a finite amount of heat transfer occurs due to higher in-cylinder pressure and temperature, resulting from the adiabatic flame temperatures at a 1.3 bar boost. Hence, the NO emission is higher compared to 1.5 bar boosting. It is also known that the heat transfer tends to be peak at stoichiometric conditions and decreases as it attains either lean or rich mixture (Heywood 2011). This also corresponds to the drop of in-cylinder gas temperature with an increase in boost pressure, as the combustion becomes lean and the premixed flame temperature becomes reduced with excess air (Singh and Sandhu 2021;Tanin et al. 1999); therefore at higher boosting of 1.5 bar, there is a reduction in NO emission.
The in-cylinder pressure variations over crank angle for different test conditions at 1800 rpm are illustrated in Fig. 6b. The in-cylinder pressure analysis gives useful data about the cycle-to-cycle variation for an average of 100 cycles. The in-cylinder pressure was analyzed for the compression ratio of 10.5:1, and increasing the boost pressure led to higher levels of in-cylinder pressure. This was also more when compared with a naturally aspirated compression ratio of 12.5:1. The higher levels of in-cylinder pressure are due to higher turbulence effects accompanied by faster flame speed. The increase in boost pressure forms a rise in peak pressure occurring closer to the top dead center compared to naturally aspirated mode, and the ignition timing was retarded to avoid knock. The presence of more air in the mixture due to higher boost pressure results in an accelerated chemical reaction which increases the turbulence and causes higher levels of in-cylinder pressure (Mathur and Sharma 2014). The heat release rate as shown in Fig. 6c is higher, while the boost pressure was increased. Figure 6d shows the COV of IMEP with respect to engine speed. The COV of indicated mean effective pressure Fig. 5 Variations of torque (a), brake power (b), brake thermal efficiency (c), brake-specific energy consumption (d), hydrocarbon emission (e), carbon monoxide emission (f), and carbon dioxide emission (g) with engine speed ◂ appears to be more at lower engine speed and higher boost pressures. At lower engine speed, the mean piston speed is lower and reduced exhaust gas velocities exiting the exhaust port cause a turbo lag which results in incomplete combustion. At higher boost pressure, the combustion instabilities increase due to more turbulence. Also, during the blowdown phase, the residual gas transfer occurs on account of the choked flow of the exhaust gas, resulting in variation in combustion and raising the indicated work.

Conclusions
An automotive spark-ignition engine fueled with compressed bio-methane experimented at a compression ratio of 10.5:1 and 12.5:1 at turbocharged and naturally aspirated mode, respectively. Two turbochargers T1 and T2 were selected and compared with different boost pressures. Based on the numerical study of T1 and T2, T1 attained a higher-pressure ratio with a smaller A/r ratio of compressor scroll. The experimental results reveal that an increase in charge density using T1 turbocharger at a compression ratio of 10.5:1 with 1.3 bar boost pressure increases brake power by 19.3% compared to the naturally aspirated compression ratio of 12.5:1. Also, the brake thermal efficiency increased from 24.8 to 26.3% and reduced energy consumption from 14.5 MJ/kWh to 13.7 MJ/kWh. The reduced levels of HC, CO, and CO2 emissions by 25%, 8.2%, and 4.9% increase NO emission. The novelty of this research work is to reduce the compression ratio using the turbocharging technique.