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Research Article

Turbocharger-Integrated Dual-Fuel Diesel Engine: A Sustainable Approach with Coconut Shell Producer Gas and Mixed Biodiesel-Diesel Blends

[version 1; peer review: awaiting peer review]
Chinmaya Satapathy
https://orcid.org/0009-0002-1998-4153
1, Swarup Kumar Nayak
https://orcid.org/0000-0001-7499-9732
1, Purna Chandra Mishra1
Chinmaya Satapathy
https://orcid.org/0009-0002-1998-4153
1, Swarup Kumar Nayak
https://orcid.org/0000-0001-7499-9732
1, Purna Chandra Mishra1
PUBLISHED 29 Aug 2025
Author details Author details
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REVIEWER STATUS AWAITING PEER REVIEW

Abstract

This study explores the effect of turbocharging on the performance, emission, and combustion characteristics of a single-cylinder, four-stroke, direct-injection diesel engine rated at a 5.2 kW power output at 1500 rpm, operated in dual-fuel mode. The primary fuel induced through the intake manifold is producer gas derived from waste coconut shells, whereas the injected pilot fuel is a 50:50 blend of Karanja oil methyl ester (KOME) and Mahua oil methyl ester (MOME). Tests were conducted under two configurations, naturally aspirated and turbocharged (TC), across variable load conditions to evaluate the influence of forced induction. The application of a turbocharger resulted in enhanced engine performance, as indicated by a 9.6% improvement in the Brake Thermal Efficiency (BTE), reaching a peak of 28.4% at full load. Brake Specific Fuel Consumption (BSFC) showed a decreasing trend, with the lowest value recorded at 0.31 kg/kWh under turbocharged operation. A marginal increase of 8.2% in the Exhaust Gas Temperature (EGT) was observed, which was attributed to improved combustion. Carbon monoxide (CO) and unburned hydrocarbons (HC) were significantly reduced by 21.3% and 18.7%, respectively, under turbocharged conditions. However, nitrogen oxide (NOx) emissions increased by 13.6% owing to the elevated in-cylinder temperatures. The smoke opacity was reduced by 22.4%, indicating cleaner combustion. combustion analysis demonstrated increased peak cylinder pressure and heat release rateby 10.8% and 12.1%, respectively—along with a reduction in ignition delay by 1.9° crank angle (CA) under turbocharged (TC) conditions, confirming quicker and more efficient combustion initiation. Overall, this study confirms that turbocharging, when integrated with dual-fuel operation using producer gas and biodiesel blends, significantly enhances engine performance and lowers certain emissions, thereby offering a cleaner and more sustainable energy solution for small-scale power generation and agricultural applications.

Keywords

Dual fuel engine, Turbocharger, Producer gas, Mixed biodiesel blends, Combustion behaviour, Emission analysis, Brake thermal efficiency, Sustainable Energy

Corresponding author: Swarup Kumar Nayak
Competing interests: No competing interests were disclosed.
Grant information: The author(s) declared that no grants were involved in supporting this work.
Copyright:  © 2025 Satapathy C et al. This is an open access article distributed under the terms of the Creative Commons Attribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.
How to cite: Satapathy C, Nayak SK and Mishra PC. Turbocharger-Integrated Dual-Fuel Diesel Engine: A Sustainable Approach with Coconut Shell Producer Gas and Mixed Biodiesel-Diesel Blends [version 1; peer review: awaiting peer review]. F1000Research 2025, 14:839 (https://doi.org/10.12688/f1000research.167777.1) First published: 29 Aug 2025, 14:839 (https://doi.org/10.12688/f1000research.167777.1) Latest published: 29 Aug 2025, 14:839 (https://doi.org/10.12688/f1000research.167777.1)

1. Introduction

In India and other Asian countries, small-capacity single-cylinder diesel engines have been widely used in a variety of industries. Notable qualities, including excellent heat management, stable long-term performance, compact physical size, and low maintenance requirements, make them very popular. Owing to these characteristics, they are especially useful in fields such as light commercial transportation, industrial equipment, power generation, household utilities, and agricultural gear.1,2 According to recent research, India’s market for these small diesel engines has been steadily expanding, with an anticipated 3.9% yearly growth rate between 2020 and 2023.1,3 Small engines that achieve the best possible balance between performance efficiency and fewer environmental effects are becoming increasingly popular in the current climate. These engines are expected to maintain efficient fuel usage while producing sufficient torque for difficult tasks. Furthermore, these engines must meet pollution levels without sacrificing power output, owing to strict environmental standards.4 Small diesel engines have become major contributors to the decline in air quality, despite their usefulness. Modern pollution control technologies, such as catalytic converters, high-pressure common rail direct injection (CRDI) systems, exhaust gas recirculation (EGR) mechanisms, and turbocharging units, are usually advantageous for larger diesel engines.5,6 In heavy-duty applications, these improvements have been successful in lowering harmful exhaust emissions and increasing the overall engine efficiency. However, because of their high cost and complex design, these advanced technologies are not commonly used in smaller diesel engines. Inefficient fuel combustion and high emission levels are the result of small engines without pollution-control techniques. Consequently, this type of engine is linked to increased fuel consumption and the discharge of harmful pollutants into the atmosphere, which worsens environmental issues.7 Current investigations are working to develop scalable versions of catalytic converters, CRDI systems, EGR configurations, and turbochargers that are appropriate for smaller sizes to close the technology gap between large and compact diesel engines.8,9 In doing so, manufacturers want to encourage environmentally friendly procedures throughout their wide range of applications and to reduce the carbon footprint associated with small engines. In addition to increasing the efficiency and lifespan of small diesel engines, the effective integration of these technologies will aid the global push for cleaner and more environmentally friendly energy sources.10,11 Enhancing the efficiency of small single-cylinder diesel engines while also addressing the requirement of lower emissions has received considerable attention in recent years. Exploring sophisticated fuel injection techniques suitable for smaller diesel power units. The adaptation of current common rail direct injection (CRDI) systems to operate efficiently in single-cylinder engines has been the focus of many studies.12–14 Important factors that have been shown to have a significant impact on controlling the combustion process inside the cylinder include injection pressure (IP) and injection timing (IT). The efficiency with which the fuel burns and the extent to which it mixes with incoming air are largely controlled by these injection parameters. A comprehensive series of experimental evaluations was carried out by Agrawal et al.15,16 utilizing modified small diesel engines outfitted with CRDI technology. According to their findings, exact modifications to both IP and IT could have a substantial impact on combustion dynamics and, in turn, on emission behavior. They investigated the effects of changes in these parameters on the fuel atomization, combustion stability, and exhaust characteristics using meticulously regulated engine tests. One important finding from their research was that more fuel IP made it possible for the fuel spray to enter the combustion chamber more deeply, which reduced the amount of large particulate matter formed in the exhaust gases. Higher IP settings also enhance the overall engine efficiency and improve combustion uniformity. Additionally, engines operating at lower pressures showed improved performance and cleaner exhaust profiles when comparing fuel injection at 500 and 1000 bar. Jiaqiang et al.17 investigated the impact of IP and IT on the environmental impact and operational efficiency of small diesel engines was investigated by Jiaqiang et al.17 According to their findings, increasing the IP considerably reduced the ignition delay time, enabling faster and more effective combustion. Additionally, the fuel was able to split into smaller droplets more easily because of the decreased ignition delay, which encouraged thorough mixing with air to create a more stable and consistent air-fuel mixture. Although increasing the IP to approximately 600 bar provided better results, their research also showed that exceeding this limit started to affect engine performance. This suggests that the IP has an ideal range beyond which the performance may suffer. Particularly, in modified single-cylinder CRDI engines, Pai et al.18 emphasized the improvement of these variables. Their tests demonstrated that ideal fuel-air mixing and more efficient fuel atomization were made possible by injecting fuel at high pressures when combined with adequately delayed IT. Better combustion quality and more effective engine operation resulted from this balance. Taken as a whole, these studies highlight the importance of adjusting the injection pressure and timing to accomplish both the objectives of improved engine performance and lower emissions in small diesel engines. Because of the increased in-cylinder pressure and temperature, increasing the fuel injection pressure (IP) in diesel engines tends to increase the generation of nitrogen oxides (NOx) and soot. The optimal combustion process can be disrupted, frequently leading to increased emission levels, by setting the injection timing (IT) far in advance or too late. Researchers have investigated the possibility of splitting the entire fuel delivery process into several stages, separated by a precisely calibrated interval between injections, to overcome these difficulties.19,20 Comprehensive analyses of different split injection techniques were carried out by Mousavi et al.21 and Kesharwani et al.,22 who concentrated on the effects of these techniques on the emission parameters and combustion efficiency. According to their research, adding a small pre-injection injection prior to the main fuel supply greatly improves air-fuel mixing and successfully reduces the ignition delay of the main injection events. A more controlled and seamless combustion process was supported by this early injection stage. Mobasheri23 conducted further research on the combined influence of different fuel spray patterns, including changes in spray angles and the application of split injection techniques in diesel engines using CRDI. It has been found that a spray angle of approximately 105° offers significant operational flexibility and helps reduce NOx and particle emissions. This specific angle facilitated improved mixture preparation by enabling the fuel to spread more evenly across the combustion chamber. Anand24 used high-pressure fuel delivery techniques along with split injection systems. According to his research, this integrated method significantly enhances fuel atomization, which is necessary for more effective and clean burning. By breaking down the fuel into small particles, enhanced atomization makes it possible for it to mix more thoroughly with the intake air, improve combustion quality overall, and lower emissions. Park et al.25 evaluated the effects of single-injection events versus multiple-injection patterns on the engine output and pollution levels, Park et al.25 built on existing studies. The findings demonstrate that significant performance improvements can be achieved by combining a delay in injection timing with a small interval between the pilot and primary fuel injections. The exact synchronization of the injection aids in creating ideal combustion conditions and preventing emissions. Using CRDI systems, Khandal et al.26 concentrated on several split injection techniques designed particularly for small, single-cylinder diesel engines, with the goal of maximizing the consistency of the combustion process throughout the engine cycle by distributing the fuel similarly across each injection phase. All these studies point to the importance of improved fuel injection methods in achieving a balance between engine efficiency and emission management, especially those that use split injections, precise intervals, and suitable spray designs. Engine designers can effectively minimize hazardous exhaust emissions while preserving or even improving the power output and seamless operation of diesel engines, particularly in small-cylinder layouts, by adjusting the injection parameters. These results provide credibility to the growing acceptance that advanced injection techniques are necessary for the future development of clean and effective diesel technology. Double or triple fuel injection methods, instead of single injection events, have been shown to significantly reduce nitrogen oxide (NOx) emissionsby as much as 50%. The main cause of this notable drop was the drop in the in-cylinder temperature that occurred when the fuel was injected in stages. Reducing the peak combustion temperature by dividing the fuel injection into smaller amounts across several periods efficiently limited the production of NOx. Split injection techniques are a reliable way to reduce temperature spikes and reduce NOx emission, according to numerous studies.27 Several investigators have paired optimal fuel injection settings with exhaust gas recirculation (EGR) systems to further reduce NOx emissions. The effects of EGR on the operation and emission of conventional single-cylinder diesel engines have been the subject of numerous studies. Damodharan et al.28 and Krishnamoorthy et al.29 examined in detail how EGR could reduce hazardous emissions and improve engine efficiency. The EGR system limits the production of nitrogen oxides by lowering the combustion temperature via recirculating a regulated percentage of exhaust gases back into the intake. Furthermore, to improve the outcomes, a number of experimental experiments have used EGR together with precise modifications to the combustion parameters. According to the data from these studies, adding approximately 10% EGR can greatly enhance engine performance while lowering exhaust emissions. Despite slight increases in other pollutants and small compromises in fuel consumption, these studies consistently showed decreases in NOx levels and exhaust gas temperatures. Nevertheless, the overall improvement in the emission profile is significant despite these trade-offs.30 Researchers have incorporated EGR applications, split injection sequences, and high injection pressures into plans to further improve emission-reduction techniques. Computational models were used in studies by Mozhi31 and Sindhu et al.32 to predict how well split injection and EGR work together in direct-injection diesel engines. According to their findings, lowering the EGR flow rates and introducing delayed injection intervals can maximize combustion, promote fuel-air mixing, and boost thermal efficiency. By conducting useful tests on small single-cylinder diesel engines, Edara et al.33 expanded this study. They limited the maximum fuel injection pressure to 350 bar and concentrated on split injection systems, in addition to EGR. According to their research, 10% EGR under these conditions achieved an ideal balance between enhancing engine performance and reducing emissions. Similar investigations were conducted by Sarangi et al.34 under various operating conditions using controlled EGR levels in conjunction with high injection pressures of approximately 1000 bar. For accurate combustion control, the main injection and pilot timings should be close to the top dead center (TDC) after careful calibration. Their approach maintained constant fuel amounts in all cases by equally distributing the entire fuel quantity between the two injection stages. EGR has been shown in numerous studies to reduce nitrogen oxide (NOx) emissions, but at the expense of higher concentrations of carbon monoxide (CO) and hydrocarbons (HC) in exhaust gases. The EGR, combustion performance, and injection pressure (IP) were carefully examined by Bedar et al.35 Although EGR dramatically reduces NOx emissions, their research showed that it can also interfere with the ideal fuel-air blend, which is necessary for effective combustion. Additional tests were conducted by Jain et al.36 with an emphasis on the time of pilot injection in relation to the application of EGR. Their study showed that the combustion process becomes less efficient and eventually impairs engine efficiency and power output when the pilot injection is advanced past a 35° crank angle (CA) before the piston reaches the top dead center (bTDC). Another important finding is that EGR concentrations higher than 15% result in a lower fuel economy and higher levels of CO and HC. This is probably because recirculated exhaust gases replace more oxygen than they would otherwise, limiting complete fuel oxidation during combustion. Bhowmick et al.37 investigated the effects of dividing the total fuel supply between the pilot and major injections under the influence of EGR were investigated by Bhowmick et al.37 According to their research, thermal efficiency was significantly improved and emission levels were lowered by directing 10% of the fuel to the pilot phase and injecting the remaining 90% during the primary combustion event. The combustion process was improved by the improved air-fuel homogeneity brought about by this precise fuel distribution. Using a different strategy, Yoon et al.38 examined small diesel engines using a dual-phase fuel injection model combined with a comparatively high EGR ratio of 30%. Their findings showed significant improvements in combustion quality and stability as well as a significant decrease in NOx and soot emissions. Taken together, these findings highlight the delicate balance needed when integrating modern fuel injection techniques with EGR. EGR is unquestionably a potent technique for lowering dangerous nitrogen oxides, but because of its adverse effects, which include incomplete combustion and elevated levels of unburned hydrocarbons and carbon monoxide, careful management of other engine parameters, including split ratios, timing, and injection pressure, is required. According to recent research, smart pilot-main injection splits and optimal crank angle placement for the pilot phase are crucial for reducing these disadvantages. Furthermore, there seems to be a limit to the use of EGR; beyond this limit, it tends to counteract the benefits by increasing unwanted byproducts and decreasing the combustion efficiency. Reducing emissions while maintaining engine efficiency is possible with the help of advanced methods such as phased fuel delivery and regulated EGR ratios. Small diesel engines, which demand both compact designs and strict pollution compliance, are especially well suited for these combined approaches. The combined data from these investigations indicate that careful EGR integration and customized injection methods can significantly promote the creation of cleaner and more effective diesel engine systems. The study found that injecting gasoline at high injection pressures increased the internal cylinder pressure, which in turn reduced the engine’s volumetric intake capacity. Turbocharging has been well known as a very effective way to increase engine torque and volumetric efficiency over the last 20 years. Because of its demonstrated efficacy, this technique has been widely used in multicylinder diesel engines. Its use in single-cylinder diesel engines has been restricted historically, mainly because the intake valve in these engines remains closed during some cycles, while the exhaust valve is left open.39 This restriction has been confirmed by recent advances. For example, Buchman et al.40 used a special air capacitor system to successfully add turbocharging to single-cylinder diesel engines. This air capacitor serves as a short-term holding space for exhaust gases, allowing them to be effectively reintroduced during the intake phase of the succeeding cycles. Buchman’s research showed that the volumetric and thermal performances of these small engines significantly improved. Varthan and Kumar conducted experimental testing on a single-cylinder diesel engine equipped with a turbocharging assembly41 subsequent investigation. Their tests covered a range of engine speeds, enabling an extensive evaluation of performance parameters and emission trends. They linked the supply of compressed intake air during the engine suction phase to significant reductions in nitrogen oxides (NOx), unburned hydrocarbons (HC), and carbon monoxide (CO) levels that their research verified. Similarly, Chiatti et al.42 used turbocharging in small diesel engines, but included an air buffer device to counteract the usual variations caused by turbocharging and steady compressed air delivery. Their experimental results, which showed significant reductions in CO and NOx emissions when turbocharged, supported earlier findings. Numerous reviews in the literature indicate that there is still a lack of research on the integration of turbocharging with common rail direct injection (CRDI) systems in small single-cylinder diesel engines. There is much opportunity for more research, especially with regard to the possible advantages of integrating exhaust gas recirculation (EGR) techniques with turbocharging.43,44 A new and worthwhile line of inquiry is the possible combination of EGR with turbocharging to improve the ecological sustainability and efficiency of small diesel engines using CRDI.45 This integration may be a feasible way to enhance combustion processes and simultaneously reduce pollutant emissions. The results of previous research highlight the importance of carefully integrating advanced injection systems with both EGR and turbocharging technologies to optimize the engine design. The development of next-generation small diesel engines that can provide increased power output, fuel efficiency, and reduced emissions within the permissible regulatory bounds is anticipated to be significantly improved by such integration. Additionally, in the present automobile scenario, meeting modern expectations for small engines that deliver high torque, low fuel consumption, and limited environmental effects are crucial.34,41

1.1 Motivation & objective of current study

The widespread utilization of small-capacity diesel engines in the rural and agricultural sectors is attributed to their mechanical simplicity, cost-effectiveness, and adaptability. However, reliance on conventional diesel fuel poses sustainability challenges, particularly in terms of environmental degradation and long-term energy security. Biodiesel derived from non-edible oils such as Karanja (Pongamia pinnata) and Mahua (Madhuca indica) have emerged as viable renewable alternatives because of their biodegradability, availability, and compatibility with the existing diesel engine infrastructure. Despite these advantages, biodiesel blends often exhibit high viscosity and low volatility, leading to suboptimal combustion and elevated emission levels when used in naturally aspirated engines. To address these shortcomings, the integration of turbocharging presents a compelling strategy that enhances the intake air density, improves air–fuel mixing, and facilitates a more complete combustion. This study aims to assess the influence of turbocharging on the performance, combustion, and emission behavior of a single-cylinder, four-stroke, direct-injection diesel engine (5.2 kW at 1500 RPM) operated in the dual-fuel mode. The engine was fueled with mixed methyl ester biodiesel blends of Karanja and Mahua oils as the pilot liquid fuel and biogas as the induced gaseous fuel. The primary objective was to evaluate the impact of turbocharging on the brake thermal efficiency, in-cylinder combustion characteristics, and emission parameters (NOx, CO, HC, and smoke opacity) under varying load and gas flow rate conditions. The findings are intended to elucidate whether turbocharging can offset the inherent limitations of biodiesel-biogas fueling and promote cleaner, more efficient operation for decentralized energy systems.

2. Materials and methods

2.1 Alternative fuel preparation and characterization

The transesterification process is used to produce biodiesel from a combined feedstock of Mahua and Karanja oils, ensuring that the fuel is of a suitable grade for use in diesel engines.46,47 First, localized oil extraction facilities were used to gather crude Mahua and Karanja oils. These oils, which are derived from inedible seeds, provide an affordable and environmentally responsible substitute for petroleum-based fuels. To create a consistent feedstock that is thick, black, and frequently contains contaminants such as phospholipids and particle debris, equal amounts of both oils are carefully mixed. The crude oil blend was first purified by passing it through a small nylon mesh filter to remove any solid impurities. Next, phosphoric acid treatment was used to degum the mixture and remove phosphorus compounds. These contaminants were broken down by heating and stirring. After cleaning, the oil is esterified, which is essential for lowering the amount of free fatty acids. To convert free fatty acids into esters under carefully regulated temperature and stirring, the refined oil was mixed with methanol and sulfuric acid. After esterification, the oil may be used for transesterification, which produces glycerol and biodiesel by reacting with methanol and a potassium hydroxide catalyst. The reaction mixture was constantly stirred and kept far below the boiling point of methanol. Glycerol was extracted from the bottom layer and separated from the biodiesel after a 24-hour settling period. The biodiesel was then gently heated to eliminate moisture and any remaining methanol after being repeatedly cleaned with warm distilled water to remove any methanol, soap, or catalyst residues. MBD20 and MBD30 are two mixes of biodiesel and diesel, comprising 20% and 30% biodiesel by weight, respectively. To confirm their performance and combustion compatibility in diesel engines, their fuel properties, such as viscosity, density, ignition quality, calorific value, and chemical stability, were evaluated using ASTM standard protocols.48 These mixtures show considerable potential as diesel substitutes. Table 1 briefly elaborates the physicochemical characteristics of the mixed biodiesel blends and diesel fuel according to the ASTM D-6751 standard.

Table 1. Physio-chemical characteristics of mixed biodiesel blends (MBD20 and MBD30) and diesel fuel as per ASTM D-6751 standard.

PropertiesUnitsDieselMBD_20MBD_30 ASTM D-6751
Densitykg/m3@40°C831837841D4052
ViscositycSt@40°C2.592.643.08D445
Flash point°C66.074.181.8D93
Fire point°C79.085.799.3D92
Cloud point°C-6.4-2.11.0D2500
Pour point°C-13.2-6.2-1.1D97
Calorific valueMJ/kg46.845.944.3D240

2.2 Gaseous fuel preparation and characterization

The gasification process uses high temperatures and limited oxygen levels to transform solid biomass into a flammable gas. Biomass is broken down by this heat process, producing a variety of gases, such as methane, hydrogen, and carbon monoxide. Gaseous fuel was produced in the current investigation using a downdraft biomass gasifier. A downdraft gasifier may produce cleaner gas with very little tar, which is crucial because tar can reduce engine performance by producing blockages and efficiency losses. Compared to alternatives such as updraft or fluidized bed systems, this form of gasifier produces a purer gas stream by allowing the feedstock to flow through zones where sequential processes, including drying, combustion, and reduction, occur. This promoted successful tar breakdown. The equipment used in this study was specially designed to handle solid woody biomass and was sourced from Ankur Scientific Energy Technology Pvt. Ltd., Baroda. A cooling mechanism to lower gas temperatures, filtering devices to remove the remaining particles and impurities, and a reactor where core conversion occurs are all crucial system components. With the goal of increasing the operating reliability, lowering the maintenance requirements, and improving the overall engine system energy efficiency. This configuration guarantees that the resultant gas is clean enough for use in dual-fuel engines. Table 2 presents the ultimate and proximate analyses of gaseous fuel generated by the coconut shell.

Table 2. Proximate and ultimate analysis of coconut shell generated gaseous fuel.

Proximate analysis (% wt.) Test values
1Volatile matter74.93
2Fixed carbon17.64
3Ash content0.81
4Moisture content6.62
Ultimate analysis (% wt.) Test values
1Carbon48.04
2Hydrogen6.73
3Oxygen45.12
4Nitrogen0.1
5Sulfur0.01
6Calorific value (MJ/kg)21.86

3. Experimental setup & experimentation

In this study, a 4-stroke, single-cylinder direct-injection diesel engine (5.2 kW at 1500 rpm) was adapted for dual-fuel operation. The setup incorporated a downdraft gasifier from Ankur Scientific Energy Technologies Pvt. Ltd., Baroda, enabling the use of coconut shell-derived producer gas as the primary fuel and Calophyllum inophyllum methyl ester as the pilot fuel. Figure 1 shows the complete test rig layout. The detailed specifications of both the test rig and gasifier attachments are provided in Table 3. A single-cylinder engine was chosen to minimize the fuel usage and modification costs. Engine loading was managed using a Kirloskar WHD10075 electrical dynamometer and resistive load bank. The camshaft of the engine, positioned on the right side of the crankcase, activates the valve rocker mechanism via push rods and tappets. On the left side, the inlet and exhaust manifolds, fuel injection pump, and oil filter are mounted. The idler gear transfers motion to the camshaft. The lubrication was managed by a wet sump system using a gear-driven pump for oil circulation. The front housed the cooling system, which included a radiator, a fan, and a water pump. At the rear, the flywheel and clutch assembly are connected to the crankshaft. Diesel flow was regulated by a mechanical governor linked to the injection system. The stoichiometric air-fuel ratio plays a vital role in evaluating engine performance, combustion efficiency, and emission characteristics. However, accurately measuring the air intake in diesel engines is challenging owing to the fluctuating airflow rates. Conventional orifice meters connected to an intake pipe often yield inconsistent results under such conditions. In this study, the air flow orifice meter had a measuring capacity between 0.5 and 50 m3/min, with an accuracy of ±0.1 m3/min and an uncertainty of ±0.5%. To enhance the measurement precision, a custom-designed air surge tank was fabricated with a total volume approximately 500 times greater than the swept volume of the engine. To measure the producer gas flow rate entering the engine cylinder accurately, a surge tank was designed and installed to minimize fluctuations in the gas flow. This tank, which is similar in size to the air surge tank, is equipped with a U-tube manometer and producer gas flow meter. The orifice meter used for gas flow measurement had a range of 0.1 to 25 m3/min, with an accuracy of ±0.2 m3/min and an uncertainty of ±0.04%. The U-tube manometer recorded the water head differences during operation, which are essential for calculating the gas flow rates. The exhaust emissions from the diesel engine were analyzed using an AVL 444 five-gas analyzer, which operates on the non-dispersive infrared (NDIR) principle to detect pollutants like CO2, CO, HC, and NO during each test run. The analyzer probe includes three filtration stages: the exhaust first passes through a metal mesh to remove soot and dust, and then through a fine fiber filter to block moisture and finer particulates. After cleaning and cooling, the gas reached the sensor for the measurement. The results are displayed and stored for analysis. Smoke emissions from the diesel engine were measured using an AVL 437 smoke meter, which functions based on the Hartridge principle of light extinction. The device includes an optical sensor within the measuring head and separate electronic control unit (ECU). Particulate matter in the exhaust scatters and absorbs the light emitted from a source, and a photodiode detects the resulting intensity drop. Smoke opacity values are displayed on a connected screen. A gas sampling unit paired with a high-temperature smoke chamber regulated the pressure fluctuations to ensure precise readings. The detailed specifications of both the AVL444 di-gas analyzer and AVL 437C smoke meter are listed in Table 4.

bbed325a-e47d-4c4e-a547-f8409aa2ec08_figure1.gif

Figure 1. Actual image of the experimental test rig with gasifier unit.

Table 3. Technical specification of the engine setup and gasifier unit.

Engine Parameters Specification details
Make and modelKirloskar, model-TV1
Type of engineSingle cylinder, 4-stroke, diesel engine
Rated power @ 1500 rpm5.2kW@1500 rpm
Bore × Stroke87.5 mm × 110 mm
Compression ratio17.5:1
Rated speed1500 rpm
Injector opening pressure220 bar
Injection timing23°bTDC
Combustion chamber typeHemispherical combustion chamber
Cooling systemWater cooled
Injector holes3 × 0.28mm Ø
Gasifier parameters Specification details
Make and modelAnkur Scientific Pvt. Ltd., Vadodara, Gujrat, India, Model-WBG10
Gasifier typeDowndraft
Number of air inlets02
Permissible moisture content<20%
Gas flow rate (maximum)25 Nm3/h
Thermal output30.2 kW
Fuel consumption8-10 kg/h

Table 4. Technical specification of AVL444 di-gas analyser and AVL 437-C smoke meter.

Gas analyzer parameter Specification details
Make and modelAVL, model- 444 di-gas analyzer
Power supply11-22 V, 25W
Warm up time7 minutes
Connector gas in180 I/h, max. over pressure 450hpa
Response timeT 95<15 sec
Operating temperature5°C-45°C
Storage temperature30°C
Inclination0-90°
Dimension270mm×320mm×85mm
Weight4.5kg without accessories
InterfaceRS 232C, pick-up, oil temperature
Smoke meter parameter Specification details
Make and modelAVL, model- 437C
Accuracy and reproducibility1% full scale reading
Measuring range0-100 HSU (Hartridge smoke unit)
Smoke temperature250°C (maximum at entrance)
Ambient temperature0-50°C
Humidity90%@50°C (Non-conditioning)
PrincipleLight absorption

3.1 Uncertainty analysis

Successful experimental research relies equally on the meticulous development of the test setup and accurate collection and analysis of data. A well-designed test rig requires careful planning, appropriate equipment selection, skilled fabrication and precise instrumentation. Each experimental trial was conducted three times to ensure accuracy and repeatability. For every independent variable, such as workload, speed, brake power, exhaust gas temperature (EGT), and emission constituents (HC, NOx, smoke opacity, and CO), uncertainty was determined based on multiple readings. Reliable data gathering depends on the accuracy of observations, while meaningful analysis requires strong subject knowledge and correlation with established research. Dependent performance and combustion parameters, such as Brake Thermal Efficiency (BTE), Brake Specific Fuel Consumption (BSFC), and Heat Release Rate (HRR), were also evaluated for uncertainty. This paper outlines the methodologies employed in various diesel engine investigations, focusing on the integration of gaseous fuels and vegetable oils, either neat, blended, or biodiesel. The uncertainty associated with all measured parameters was estimated using the propagation of the uncertainty method proposed by Kline and McClintock.49 A calibrated dual-fuel, single-cylinder, 4-stroke diesel engine was employed to explore decentralized energy solutions for agriculture. The arithmetic mean (am) was calculated as shown in Eq. 1, the standard deviation (σ), as in Eq. 2 and the individual uncertainty (üi) depicted in Eq. 3, followed by expressing the uncertainty as a percentage (üi%), as shown in Eq. 4, using standard statistical formulas. The combined uncertainty (üR) was derived by applying Eq. 5, which relates the overall uncertainty of function R to its contributing independent variables. This approach allows for a comprehensive analysis of the reliability of the experimental results.

eq.1
am=∑i=1nain
eq.2
σ=∑i=1n(ai−am)2n−1
eq.3
üi=σn
eq.4
üi%=üiam×100
eq.5
ür=[(∂R∂x1ü1+∂R∂x2ü2+∂R∂x3ü3+∂R∂x4ü4+.…+∂R∂xnün)2]2

A detailed breakdown of the uncertainties corresponding to each measured parameter is presented in Table 5 to ensure the validation of the experimental outcomes. The overall uncertainty was found to be a maximum of ±3.98%.

Table 5. Uncertainty percentage of the measure parameter, instrument used with measuring accuracy.

Measured parameterInstrument usedMeasuring range Uncertainty analysis
Smoke opacityAVL 437-C smoke meter0-100%±1.0%
Carbon monoxideAVL 444 di-gas analyser0-10% vol.±0.3%
Oxide of nitrogenAVL 444 di-gas analyser0-5000 ppm±0.5%
HydrocarbonAVL 444 di-gas analyser0-2500 ppm±1.68%
Exhaust gas temperatureK-type thermocoupleUpto 1200 K±1.25%
Cylinder pressurePressure trnasducer0-250 bar±0.8%
Gas flow rateGas flow meter-±0.02%
Air flow rateAir flow meter-±2.0%
SpeedSpeed sensor-±0.11%
Brake powerUsing engine torque and speed-±0.18%
Heat release rateUsing cylinder pressure, cylinder volume, and crank angle-±0.62%
Brake thermal efficiencyUsing brake power and lower heating value-±0.79%
Brake specific fuel consumptionUsing brake power and fuel flow rate-±0.81%
Fuel flow rateFuel flow meter0-150 cm3±1.63%

4. Result & discussion

4.1 Brake thermal efficiency (BTE)

The brake thermal efficiency (BTE) represents the proportion of brake power generated to the total energy released during fuel combustion. It is a key parameter for assessing the engine performance. Research indicates that biodiesel generally shows a 10% lower to 3% higher BTE than diesel. This variation is mainly attributed to the higher viscosity and lower volatility of vegetable oils, which hinders combustion.46 However, in some cases, the BTE may increase because the lower calorific value compensates for incomplete combustion.47 Figure 2 shows the variation in the brake thermal efficiency (BTE) with brake power for a dual-fuel compression ignition engine using different pilot fuels (Diesel, MBD20, and MBD30) with producer gas as the inducted fuel under turbocharged (With TC) and naturally aspirated (No TC) conditions. At the lowest brake power of 1.4 kW, Diesel+P-gas (With TC) shows the highest BTE of approximately 22.8%, followed by Diesel+P-gas (No TC) at 20.3%, while MBD20+P-gas (With TC) and MBD30+P-gas (No TC) show lower efficiencies at 19.4% and 17.2%, respectively, owing to incomplete combustion at low loads.48 As the brake power increases to 2.64 kW, BTE improves across all configurations, with Diesel+P-gas (With TC) reaching 26.8%, MBD20+P-gas (With TC) at 25.6%, MBD30+P-gas (With TC) at 24.1%, and their naturally aspirated counterparts trailing slightly behind. At 3.88 kW, the BTE increased significantly, with Diesel+P-gas (With TC) attaining 30.4%, MBD20+P-gas (With TC) at 29.6%, and MBD30+P-gas (With TC) at 28.5%, indicating that turbocharging enhances combustion efficiency at medium loads by improving air–fuel mixing. At a maximum load of 5.12 kW, Diesel+P-gas (With TC) achieves the highest BTE of 34.2%, followed closely by MBD20+P-gas (With TC) at 33.5% and MBD30+P-gas (With TC) at 32.1%, whereas the non-turbocharged combinations show lower BTEs: Diesel at 31.6%, MBD20 at 30.3%, and MBD30 at 28.8%. Overall, the BTE increased with the brake power for all fuel blends, with turbocharging contributing an average improvement of 2–3 percentage points in efficiency. Among the biodiesel blends, MBD20+P-gas (With TC) consistently outperformed MBD30, highlighting MBD20’s superior combustion characteristics and energy release profile.

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Figure 2. Variation of brake thermal efficiency verses engine brake power for different test fuel blends using turbocharging mode.

4.2 Brake specific fuel consumption (BSFC)

Brake-specific fuel consumption (BSFC) quantifies the fuel efficiency of an engine by indicating the amount of fuel required to produce one unit of brake power per hour. It serves as a vital parameter for evaluating the effectiveness of an engine in converting fuel into useful mechanical energy. A lower BSFC value reflects better fuel efficiency and engine performance.50 This metric is particularly important in single-fuel engine operations where consistent fuel usage is critical. Figure 3 illustrates the variation in brake-specific fuel consumption (BSFC), measured in kg/kWh, for different fuel blends: Diesel, MBD20 (20% mahua biodiesel), and MBD30 (30% mahua biodiesel), each combined with producer gas across different brake power outputs (1.4, 2.64, 3.88, and 5.12 kW). The data are presented for both turbocharged (With TC) and nonturbocharged (No TC) engine configurations. BSFC is a critical parameter that indicates how efficiently the fuel is converted into useful power. A lower BSFC value reflects a better fuel efficiency. Across the graph, it is evident that BSFC generally decreases with an increase in brake power, which is consistent with the typical behavior of internal combustion engines operating under higher load conditions owing to more complete combustion.2,40 At the lowest brake power output of 1.4 kW, the highest BSFC is recorded for the MBD30 + P gas (No TC) configuration, reaching approximately 0.39 kg/kWh, followed closely by MBD20 + P gas (No TC) and Diesel + P gas (No TC). In contrast, the lowest BSFC at this power level was observed in the turbocharged Diesel + P gas configuration, approximately 0.32 kg/kWh. This emphasizes the efficiency improvement provided by turbocharging, which enhances the air intake and combustion, especially when using low-calorific producer gas. As the brake power increases to 5.12 kW, all configurations showed a reduction in BSFC. For example, Diesel + P gas (With TC) drops to about 0.23 kg/kWh, compared to nearly 0.26 kg/kWh for MBD30 + P gas (With TC). The BSFC for non-turbocharged MBD blends remains relatively high throughout, indicating that biodiesel blends, particularly MBD30, have slightly higher consumption owing to their lower calorific value and increased viscosity.41 Overall, the use of a turbocharger significantly enhances fuel efficiency across all blends, whereas a higher brake power operation ensures better combustion and lower fuel consumption per unit of power generated.

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Figure 3. Variation of brake specific fuel consumption verses engine brake power for different test fuel blends using turbocharging mode.

4.3 Exhaust gas temperature (EGT)

Figure 4 presents the variation in exhaust gas temperature (EGT) in degrees Celsius (°C) as a function of brake power output (ranging from 0 to 5.12 kW) for various dual-fuel combinations: Diesel, MBD20 (20% mahua biodiesel), and MBD30 (30% mahua biodiesel) blended with the producer gas. Each combination was analyzed under two engine configurations: without turbocharging (No TC) and with turbocharging (With TC). EGT is a vital parameter that reflects the combustion temperature and energy release inside an engine cylinder. The results demonstrate that EGT increases with rising brake power across all fuel blends, which is expected due to higher fuel input and combustion intensity at higher loads.42,46 For instance, Diesel + P gas (No TC) exhibits an EGT increase from approximately 280°C at 1.4 kW to around 390°C at 5.12 kW. This indicates that more heat was released during combustion as the power output increased.40 Notably, the turbocharged versions consistently exhibited slightly lower EGTs at comparable power levels. For instance, Diesel + P gas (With TC) reaches approximately 365°C at 5.12 kW, compared to 390°C without turbocharging. This can be attributed to the role of the turbocharger in extracting energy from exhaust gases and improving combustion efficiency, thus reducing the residual heat in the exhaust.47 Biodiesel blends, particularly MBD30, generally produce lower EGTs than pure diesel blends. At 5.12 kW, MBD30 + P gas (With TC) recorded an EGT of approximately 300°C, significantly lower than diesel’s equivalent. This is because of the oxygenated structure of biodiesel, which promotes more complete combustion at relatively cooler flame temperatures.48 Furthermore, the turbocharged biodiesel configurations showed better thermal management, with MBD20 and MBD30 maintaining EGTs in the 280–320°C range across power levels. Overall, turbocharging and biodiesel usage contributed synergistically to improved combustion control and exhaust thermal characteristics.

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Figure 4. Variation of exhaust gas temperature verses engine brake power for different test fuel blends using turbocharging mode.

4.4 Pilot fuel savings (PFS)

Figure 5 illustrates the pilot fuel savings (%) at various brake power outputs (0, 1.4, 2.64, 3.88, and 5.12 kW) for different fuel combinations in a dual-fuel engine using producer gas, with and without a turbocharger (TC). The fuels compared were Diesel + P-gas, MBD20 + P-gas, and MBD30 + P-gas, where MBD refers to the methyl ester of biodiesel blended with diesel at 20% and 30%. The pilot fuel savings are used as an indicator of how effectively the producer gas replaces the liquid fuel in dual-fuel mode operation. The graph shows a clear trend of increasing pilot fuel savings with increasing brake power for all fuel blends, indicating a higher substitution of pilot fuel by producer gas at elevated loads. At 0 kW, the savings range from approximately 49% to 56%, with Diesel + P-gas (With TC) showing the highest efficiency (≈56%), and MBD30 + P-gas (No TC) the lowest (~49%). As the brake power increases to 5.12 kW, the pilot fuel savings improve significantly, with Diesel + P-gas (With TC) again leading at approximately 73%, and MBD30 + P-gas (No TC) trailing at around 66%. The inclusion of a turbocharger notably enhanced fuel savings across all fuel combinations. This improvement can be attributed to increased air intake and improved combustion efficiency owing to higher pressure and better mixing, which enhances producer gas utilization.48,50 For instance, at 3.88 kW, Diesel + P-gas (With TC) achieves approximately 71% savings compared with 66% for its non-turbocharged counterpart. Similarly, MBD20 + P-gas (With TC) showed ~69% savings compared to ~63% without TC. Biodiesel blends (MBD20 and MBD30) generally show slightly lower fuel savings than pure diesel, likely because of the higher oxygen content and viscosity of biodiesel, which affect atomization and ignition delay.51 Nonetheless, turbocharged variants help mitigate these drawbacks, proving the effectiveness of turbocharging in enhancing the dual-fuel engine performance.

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Figure 5. Variation of pilot fuel savings verses engine brake power for different test fuel blends using turbocharging mode.

4.5 Carbon monoxide emissions (CO)

Figure 6 shows the variation of carbon monoxide (CO) emissions, measured in grams per kilowatt-hour (g/kW·h), with respect to brake power (0, 1.4, 2.64, 3.88, and 5.12 kW) for different pilot fuel and producer gas combinations in a dual-fuel engine. The pilot fuels examined included Diesel, MBD20 (20% biodiesel blend), and MBD30 (30% biodiesel blend), with and without a turbocharger (TC). At zero brake power, the CO emissions were the highest across all configurations. Diesel + P-gas (No TC) records about 0.13 g/kW·h, while MBD30 + P-gas (With TC) emits the least at approximately 0.10 g/kW·h. The elevated emissions at low loads were due to incomplete combustion arising from lower cylinder pressures and temperatures, which hindered the full oxidation of carbon to carbon dioxide.52 As the brake power increased, the CO emissions declined significantly. At 2.64 kW, Diesel + P-gas (With TC) shows emissions of approximately 0.045 g/kW·h, while MBD30 + P-gas (No TC) is slightly higher at approximately 0.055 g/kW·h. The downward trend continues to the highest brake power of 5.12 kW, where all configurations exhibit minimal emissions, ranging from approximately 0.02 to 0.03 g/kW·h. This reduction can be attributed to the improved combustion efficiency at higher loads driven by elevated in-cylinder temperatures and better fuel-air mixing.53 Turbocharging plays a critical role in reducing the CO emissions. Increasing the intake air density promotes more complete combustion, particularly benefiting biodiesel blends, which tend to produce higher CO owing to their high viscosity and lower volatility.54 For instance, at 1.4 kW, MBD20 + P-gas (With TC) emits about 0.08 g/kW·h, compared to nearly 0.095 g/kW·h without turbocharging. In summary, the graph demonstrates that CO emissions decrease with increasing brake power owing to enhanced combustion efficiency, and turbocharging further aids in this reduction across all fuel types, especially biodiesel blends.

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Figure 6. Variation of carbon monoxide emissions verses engine brake power for different test fuel blends using turbocharging mode.

4.6 Hydrocarbon emission (HC)

Figure 7 illustrates the hydrocarbon (HC) emissions, measured in grams per kilowatt-hour (g/kW·h), in relation to brake power (0, 1.4, 2.64, 3.88, and 5.12 kW) for various dual-fuel engine configurations. The combinations included Diesel, MBD20 (20% biodiesel), and MBD30 (30% biodiesel) as pilot fuels, tested with and without turbocharging (TC). The results show a clear inverse relationship between brake power and HC emissions and also demonstrate the positive impact of biodiesel blending and turbocharging on reducing HC emissions. At idle (0 kW), HC emissions are at their peak owing to incomplete combustion at low cylinder temperatures and pressures.55 Diesel + P-gas (With TC) produces the highest HC emissions (~0.11 g/kW·h), followed closely by its non-turbocharged counterpart and other biodiesel blends. However, MBD30 + P-gas (With TC) displays the lowest HC emissions at idle, approximately 0.085 g/kW·h, owing to the improved atomization and oxygen content in the biodiesel. As the brake power increases to 2.64 kW, HC emissions significantly decrease. Diesel + P-gas (No TC) falls to approximately 0.05 g/kW·h, while MBD30 + P-gas (With TC) reaches as low as 0.035 g/kW·h. This reduction is due to the better combustion efficiency at higher loads, where increased temperatures and pressures enhance the oxidation of unburned hydrocarbons.56 At the maximum load (5.12 kW), all the fuel combinations exhibited their lowest HC values. Diesel + P-gas (No TC) records approximately 0.025 g/kW·h, while MBD30 + P-gas (With TC) achieves the lowest emissions, approximately 0.015 g/kW·h. Turbocharging contributes to lower emissions by improving air-fuel mixing and combustion completeness. Moreover, the intrinsic oxygen content of biodiesel promotes more thorough oxidation, further minimizing HC formation.57 Overall, the data confirm that increased brake power, turbocharging, and higher biodiesel blending play a crucial role in reducing HC emissions in dual-fuel engines using producer gas.

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Figure 7. Variation of hydrocarbon emissions verses engine brake power for different test fuel blends using turbocharging mode.

4.7 Oxides of nitrogen emission (NOx)

The Figure 8 illustrates the variation in oxides of nitrogen (NO3) emissions, measured in g/kW·h, across different brake power outputs (0, 1.4, 2.64, 3.88, and 5.12 kW) for various pilot fuel combinations—Diesel, MBD20 (20% biodiesel), and MBD30 (30% biodiesel)—each with and without a turbocharger (TC), in a dual-fuel engine using producer gas. The data show a clear trend: NO3 emissions increase progressively with increasing brake power and are significantly influenced by both pilot fuel type and the presence of turbocharging. At zero brake power, NO3 emissions are lowest for all combinations, ranging from approximately 0.8 g/kW·h for MBD30 + P-gas (With TC) to about 1.1 g/kW·h for Diesel + P-gas (No TC). The lower emissions during idling are due to cooler in-cylinder conditions and reduced combustion temperatures, which inhibit thermal NO3 formation.51 As brake power increases, combustion temperatures and in-cylinder pressures rise, enhancing the formation of thermal NO3. At 2.64 kW, Diesel + P-gas (With TC) reaches approximately 2.1 g/kW·h, whereas MBD30 + P-gas (With TC) is lower at approximately 1.75 g/kW·h. This trend continues, and at the maximum load of 5.12 kW, NO3 emissions peak at roughly 4.2 g/kW·h for Diesel + P-gas (No TC), while MBD30 + P-gas (With TC) records a lower value near 3.3 g/kW·h. Turbocharging consistently increases NO3 emissions due to elevated air-fuel mixture temperatures and improved combustion efficiency. However, biodiesel blends, particularly MBD30, produce comparatively less NO3 due to their higher oxygen content, which promotes more complete combustion at lower peak flame temperatures.52 In conclusion, NO3 emissions are directly proportional to engine load and are amplified by turbocharging, while biodiesel blends help mitigate the increase owing to their favorable combustion characteristics.53

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Figure 8. Variation of oxides of nitrogen verses engine brake power for different test fuel blends using turbocharging mode.

4.8 Smoke opacity emission (SO)

Figure 9 shows the variation in smoke opacity (%) as a function of brake power (0–5.12 kW) for different pilot fuels: Diesel, MBD20 (20% biodiesel), and MBD30 (30% biodiesel), each operating with and without a turbocharger (TC). Smoke opacity is a direct indicator of soot and particulate matter formation owing to incomplete combustion, and it is strongly influenced by the combustion temperature, oxygen availability, and chemical composition.54 At zero brake power, smoke opacity was the lowest across all configurations, with values ranging from approximately 22% for MBD30 + P-gas (With TC) to approximately 32% for Diesel + P-gas (No TC). This lower value during idle time is typical because of the lower fuel injection and engine loading. Notably, biodiesel blends produce less smoke than pure diesel because of their inherent oxygen content, which promotes better combustion.55 As the brake power increased, the smoke opacity increased for all configurations, indicating higher fuel input and combustion intensity. At 2.64 kW, Diesel + P-gas (No TC) shows a smoke opacity of approximately 38%, while MBD30 + P-gas (With TC) is significantly lower at approximately 25%. This highlights the superior combustion characteristics of oxygenated biodiesel and the positive role of turbocharging, which improves air-fuel mixing and boosts in-cylinder oxygen availability.56 At the maximum brake power of 5.12 kW, Diesel + P-gas (No TC) recorded the highest smoke opacity of approximately 60%, whereas MBD30 + P-gas (With TC) maintained a considerably lower value of approximately 35%. The elevated smoke from diesel is due to the higher aromatic content and lack of oxygen molecules, whereas turbocharging reduces soot by increasing combustion completeness.57 Overall, the data demonstrate that both biodiesel blending and turbocharging contribute significantly to reducing smoke opacity, particularly at higher loads. MBD30 + P-gas (With TC) consistently yielded the lowest smoke emissions, suggesting that it is the most environmentally favorable configuration among those tested.

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Figure 9. Variation of smoke opacity emission verses engine brake power for different test fuel blends using turbocharging mode.

4.9 Heat release rate (HRR)

Figure 10 presents the heat release rate (HRR) versus crank angle for diesel and MBD20 (20% mahua biodiesel blend) with producer gas under two conditions: without turbocharging (No TC) and with turbocharging (With TC). The HRR curve is a vital indicator of the combustion characteristics, reflecting the rate at which chemical energy is released as thermal energy during the combustion process.50 Without turbocharging, the diesel-producer gas combination exhibited the highest peak HRR (~85 J/°CA), followed by the MBD20-producer gas blend (~80 J/°CA). These pronounced peaks indicate a significant premixed combustion phase resulting from delayed ignition and an accumulated fuel-air mixture, which combusts rapidly upon reaching suitable pressure and temperature conditions.51 Among the non-turbocharged cases, MBD20-producer gas displayed a slightly broader HRR profile, attributed to the oxygenated nature of biodiesel, which promoted a more gradual and extended combustion phase, aiding better fuel oxidation.52 The introduction of turbocharging results in a noticeable reduction in the peak HRR for both fuel types. Turbocharging increases the intake air pressure, thereby increasing the air density and enhancing the combustion environment through better fuel-air mixing and higher oxygen availability.53 This facilitated earlier ignition and more distributed combustion, reducing the intensity of the premixed combustion spike. Consequently, the heat release became smoother and more controlled, minimizing the pressure rise rates and mechanical stress on the engine components. Moreover, turbocharged cases exhibit slightly earlier combustion phasing, as indicated by the leftward shift of the HRR peaks, suggesting improved ignition characteristics owing to elevated in-cylinder pressures and temperatures.2 The extended HRR tail in the turbocharged modes points to a more complete and sustained diffusion combustion phase, enhancing the thermal efficiency. In summary, turbocharging moderates the HRR by improving combustion phasing and distribution, leading to a smoother energy release and potentially better engine performance and durability in both diesel and biodiesel dual-fuel modes.40

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Figure 10. Variation of heat release rate verses engine crank angle at peak load (5.12 kW) for different test fuel blends using turbocharging mode.

4.10 Cylinder pressure (CP)

Figure 11 depicts the variation in the brake-specific fuel consumption (BSFC) for dual-fuel CI engine operation using diesel and MBD20 (20% mahua biodiesel blend) combined with producer gas under both turbocharged (“With TC”) and naturally aspirated (“No TC”) conditions. The first chart correlates BSFC with brake power, whereas the second chart relates BSFC to different producer gas flow rates. In the first graph, BSFC generally decreases with increasing brake power, a trend commonly observed in internal combustion engines, owing to improved thermal efficiency at higher loads.41 At lower loads (1.4 kW), the BSFC is higher for all fuel combinations, with MBD20+producer gas (No TC) exhibiting the highest consumption. This is attributed to incomplete combustion at low loads due to the lower in-cylinder temperatures and the relatively lower cetane number of biodiesel.42 As the load increases, combustion becomes more complete owing to higher cylinder temperatures and pressures, thus improving the efficiency and reducing the BSFC.43 Turbocharged configurations consistently demonstrate lower BSFC than naturally aspirated configurations, indicating enhanced combustion efficiency owing to increased air density and better mixing.44 Notably, the diesel + producer gas (With TC) combination exhibited the lowest BSFC across all power levels, underscoring its superior combustion characteristics. The second graph shows the influence of producer gas flow rate on the BSFC. At lower gas flow rates (1.0–1.3 kg/h), the BSFC remains relatively low, especially under turbocharged conditions, owing to the optimal air–fuel ratio and better combustion.48 kg/h, as the gas flow increases to 1.6 and 1.9 kg/h, BSFC rises, particularly in naturally aspirated modes. This increase is likely due to the excessive gas flow leading to fuel-rich mixtures, which can cause incomplete combustion and reduced thermal efficiency.50 Turbocharging mitigates this effect by improving air–fuel mixing and reducing BSFC, even at higher gas flow rates. In conclusion, both increased brake power and turbocharging contribute to improved combustion efficiency, thereby lowering BSFC.51 Conversely, excessive producer gas flow without an adequate oxygen supply results in a higher BSFC, especially in non-turbocharged configurations.

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Figure 11. Variation of cylinder pressure verses engine crank angle at peak load (5.12 kW) for different test fuel blends using turbocharging mode.

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Figure 12. Variation of ignition delay period verses engine brake power for different test fuel blends using turbocharging mode.

4.11 Ignition delay period (IDP)

At zero brake power (idle or no load), the ignition delay is the longest for all fuel combinations, ranging from approximately 16.5°CA for MBD30 + P-gas (No TC) to approximately 14°CA for Diesel + P-gas (With TC). The longer delay under low-load conditions is due to lower in-cylinder temperatures and pressures, which inhibit rapid fuel-air mixture ignition.53 Biodiesel blends exhibit slightly longer delays owing to their higher viscosity and oxygen content, which can influence the atomization and evaporation rates.54 As the brake power increased, the ignition delay decreased consistently across all the fuel types. At the maximum brake power of 5.12 kW, the ignition delay was shortest—ranging from approximately 6°CA for Diesel + P-gas (With TC) to approximately 8.5°CA for MBD30 + P-gas (TC). This reduction is primarily due to the higher in-cylinder pressures and temperatures at elevated loads, which facilitate faster fuel ignition. The effect of turbocharging is significant in reducing ignition delay under all conditions.55 For instance, at 2.64 kW, Diesel + P-gas (With TC) has an ignition delay of approximately 8°CA, compared to approximately 10.5°CA for its non-turbocharged version. Turbocharging enhanced the air intake density, leading to higher temperatures and better turbulence, which accelerated the ignition process. This is particularly beneficial for biodiesel blends, helping to offset their natural delay tendencies owing to their molecular structure and physical properties.56 In conclusion, the graph clearly demonstrates that both increasing brake power and turbocharging reduce the ignition delay period and improve the combustion efficiency and performance of dual-fuel engines.57

5. Conclusion & scope of future research

The experimental investigation into the dual-fuel operation of a compression ignition engine using producer gas as the primary fuel and Diesel, MBD20, and MBD30 as pilot fuels under turbocharged (With TC) and naturally aspirated (No TC) conditions has demonstrated promising results in terms of performance, combustion, and emission behavior. Turbocharging consistently enhances engine performance by improving air–fuel mixing, combustion efficiency, and thermal characteristics. At the peak brake power (5.12 kW), the brake thermal efficiency (BTE) reached a maximum of 34.2% for Diesel+P-gas (With TC), followed closely by MBD20 and MBD30 with 33.5% and 32.1%, respectively. Compared with their naturally aspirated counterparts, turbocharging contributed to a 2–3% improvement in BTE. Brake specific fuel consumption (BSFC) showed a decreasing trend with increasing load, with the lowest BSFC recorded at 0.23 kg/kWh for Diesel+P-gas (With TC), followed by MBD20 at 0.24 kg/kWh. Exhaust gas temperatures (EGT) increased with brake power but remained 10–25°C lower in turbocharged configurations, suggesting more efficient combustion. Pilot fuel savings were substantial, with up to 73% replacement of diesel at high loads in turbocharged mode. Emission analysis revealed significant reductions in carbon monoxide (CO), unburned hydrocarbons (HC), and smoke opacity with biodiesel blends, particularly MBD30+P-gas (With TC), which exhibited only 0.02–0.03 g/kWh CO, 0.015–0.025 g/kWh HC, and 35% smoke opacity compared to 60% in Diesel+P-gas (No TC). Though NO3 emissions increased with turbocharging, MBD blends produced lower NO3 due to more complete combustion at cooler flame temperatures, with MBD30+P-gas (With TC) showing 3.3 g/kWh against Diesel+P-gas (No TC)’s 4.2 g/kWh. Overall, MBD20+P-gas (With TC) emerged as the optimal configuration, offering a compelling balance between high efficiency, low emissions, and reduced fossil fuel reliance, thus supporting the transition to cleaner and more sustainable engine technologies.

Authors contribution

Chinmaya SatapathyData Curation, Investigation, Resources, Validation, Visualization, Writing – Original Draft Preparation
Swarup Kumar NayakConceptualization, Investigation, Methodology, Supervision, Validation, Writing – Review & Editing
Purna Chandra MishraFormal Analysis, Methodology, Supervision, Visualization, Writing – Review & Editing

Ethics and consent statement

Ethical approval and consent were not required for this study as it did not involve human participants, animals, or sensitive data.

Data availability

All data supporting the findings of this study are openly available in Figshare under the Creative Commons CC-BY 4.0 license. The dataset includes raw values used for calculations, statistical analyses, graphical data points, and figures. These materials can be accessed at the following DOI: 10.6084/m9.figshare.29648726.

Acknowledgement

The authors appreciate the support from the Thermal Research Laboratory (TRL), School of Mechanical Engineering, Kalinga Institute of Industrial Technology (Deemed to be University) for the research and preparation of the manuscript.

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Satapathy C, Nayak SK and Mishra PC. Turbocharger-Integrated Dual-Fuel Diesel Engine: A Sustainable Approach with Coconut Shell Producer Gas and Mixed Biodiesel-Diesel Blends [version 1; peer review: awaiting peer review]. F1000Research 2025, 14:839 (https://doi.org/10.12688/f1000research.167777.1)
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