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Article

Effects of Hydrotreated Vegetable Oil and Diesel Blends on Combustion, Energy Performance, and Emissions of a Compression Ignition Engine Under EGR-Controlled Operation

by
Alfredas Rimkus
1,
Justas Žaglinskis
2 and
Saugirdas Pukalskas
1,*
1
Department of Automobile Engineering, Faculty of Transport Engineering, Vilnius Gediminas Technical University, Plytinės Str. 25, 10105 Vilnius, Lithuania
2
Department of Marine Engineering, Lithuanian Maritime Academy, Vilnius Gediminas Technical University, I. Kanto 7, 92123 Klaipeda, Lithuania
*
Author to whom correspondence should be addressed.
J. Mar. Sci. Eng. 2026, 14(7), 665; https://doi.org/10.3390/jmse14070665
Submission received: 13 March 2026 / Revised: 28 March 2026 / Accepted: 31 March 2026 / Published: 31 March 2026
(This article belongs to the Section Marine Ecology)
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Abstract

The decarbonization of marine transport requires the wider use of alternative low-carbon fuels that can be applied in existing compression ignition (CI) engines without major modifications. Hydrotreated vegetable oil (HVO) is considered a promising renewable drop-in fuel due to its favorable physicochemical properties and high cetane number. This study investigates the influence of neat HVO and its blends with conventional diesel fuel on the combustion characteristics, energy, and emission indicators of a CI engine operating under different load conditions and exhaust gas recirculation (EGR) ratios. Experimental tests were carried out on a four-cylinder CI engine at constant speed and variable load using diesel fuel (D100), HVO100, and their blends (D80_HVO20 and D50_HVO50). In-cylinder pressure measurements and combustion analysis were performed using AVL instrumentation and AVL BOOST software. The results show that increasing the HVO fraction slightly advances combustion phasing and increases maximum in-cylinder pressure by approximately 4–5%. The use of HVO was found to reduce brake-specific fuel consumption by up to 3.4% and increase brake thermal efficiency by about 1.9%, although volumetric fuel consumption increases due to the lower fuel density. In addition, higher HVO content significantly reduces smoke opacity by up to 42% and decreases CO2 emissions by 4.7–6.3%, while the influence on NOx emissions depends on the applied EGR strategy. The results indicate that HVO and its blends can be effectively applied in CI engines; however, optimal performance and emission characteristics require appropriate calibration of EGR rate and fuel injection timing.

1. Introduction

In April 2018, the International Maritime Organization (IMO) submitted its contribution to the United Nations Framework Convention on Climate Change (UNFCCC) Talanoa Dialogue, outlining its Initial Strategy for the reduction in greenhouse gas (GHG) emissions from ships. The submission, issued from IMO Headquarters in London, United Kingdom, emphasized the shipping sector’s commitment to supporting global climate objectives under the Paris Agreement [1].
The document detailed the IMO’s main ambitions for reducing emissions from international shipping. These include a target to reduce the carbon intensity of international shipping by at least 40% by 2030, with further efforts aimed at achieving a 70% reduction by 2050, compared with 2008 levels. In addition, the strategy aims to reduce total annual GHG emissions from international shipping by at least 50% by 2050, while pursuing their complete phase-out within this century [1].
The IMO strategy highlights that the development and deployment of alternative zero-carbon and fossil-free fuels are essential for achieving the established emission reduction targets. Alongside other GHG mitigation measures, these approaches span short-, medium-, and long-term pathways, forming an integrated and coherent framework for policy implementation.
The role of these fuels in the GHG reduction process is significant, as the use of alternative low-carbon, carbon-free, and fossil-free fuels with favorable performance characteristics can reduce not only GHG emissions but also other harmful air pollutants from marine engines [2,3,4,5]. Advanced alternative fuels can significantly influence emission toxicity indicators [6,7,8,9,10], the reduction of which typically requires the application of costly and complex technologies such as scrubbers, selective catalytic reduction (SCR), selective non-catalytic reduction (SNCR), exhaust gas recirculation (EGR), water mist, and other emission control strategies [11,12].
In maritime practice, various fuels are increasingly used as alternatives to conventional marine distillates and heavy/residual fuels. Marine power plants commonly utilize natural gas, methanol, ammonia, hydrogen, and biodiesel [13,14,15] as alternative fuels. However, with the exception of biodiesel, these options are typically implemented in newly built ships or require complex and capital-intensive retrofit operations [16,17]. In contrast, the use of biodiesel is considerably less demanding, as it does not require fundamental changes to engine operating principles, unlike ammonia, natural gas, methanol, or hydrogen. Consequently, the integration of biodiesel technologies in both new and existing ships is more straightforward and economically feasible [18,19]. Furthermore, advanced biodiesels (hydrotreated vegetable oil and biomass-to-liquid fuels, etc.) differ from first-generation biodiesel (fatty acid methyl esters) in that they can be used in marine power plants with minimal or no operational limitations, including under varying climatic conditions [20,21].
HVO is classified as an advanced biofuel that exhibits better operational and storage properties than fatty acid methyl esters (FAME) [22]. Due to its similar or superior physicochemical properties, paraffinic HVO can replace conventional diesel fuel used in both on-road and off-road vehicles, as well as petroleum distillates used in marine transport [23,24]. HVO contains no sulfur compounds or aromatic hydrocarbons and is characterized by a relatively low average molecular weight, resulting in a high cetane number ranging from 70 to 95, compared to 40–55 for fossil diesel [22]. The chemical properties responsible for the high cetane number also contribute to improved low-temperature performance and a “softer” heat release profile, avoiding a pronounced premixed combustion peak [25]. These combined effects lead to a more stable ignition process and smoother pressure rise, making the combustion more controllable; moreover, the shorter ignition delay associated with high-cetane HVO can, under certain operating conditions, help limit peak NOx formation by reducing the intensity of the premixed combustion phase [26,27].
According to sources [28,29,30], HVO can reduce carbon dioxide emissions by 60–95% compared to fossil diesel when evaluated over the entire fuel life cycle. In terms of engine-out emissions, pollutant reductions are more moderate: carbon dioxide (CO2) by approximately 2–4%, nitrogen oxides (NOx) by 2.5–25%, and particulate matter (PM) by 50–80% [31,32,33,34,35]. These results indicate that HVO is a promising fuel in the context of the IMO’s short-, medium-, and long-term GHG reduction strategies and the transition toward alternative zero-carbon and fossil-free fuels [1].
In the maritime sector, interest in HVO has been steadily increasing due to its suitability as a drop-in replacement for conventional marine distillates and its compatibility with existing propulsion and auxiliary engines without requiring hardware modifications. Recent assessments have demonstrated that HVO can provide tangible benefits under typical operating profiles of marine diesel generators, which often run at constant speed and variable load, while ensuring stable ignition behavior, improved combustion controllability, and reliable performance across diverse climatic regions [22,23,24]. Owing to its paraffinic composition and excellent cold-flow characteristics, HVO remains operational in low-temperature marine environments where traditional marine gas oils may require cold-flow additives or pre-heating systems [22]. Furthermore, studies report that its inherently low aromatic content and cleaner oxidation pathways substantially reduce particulate matter and smoke opacity—an advantage of particular relevance for vessels operating in ports, coastal areas, and emission control zones [31,32,33,34,36]. In addition, several marine engine manufacturers and certification bodies have already approved HVO100 for selected propulsion and auxiliary engines, highlighting its practicality as an immediately deployable measure for reducing local pollutant emissions and supporting compliance with IMO decarbonization objectives [23,24,36]. Taken together, these findings position HVO as a technically mature transitional fuel capable of delivering near-term environmental benefits while the maritime sector gradually develops the infrastructure required for future zero-carbon energy carriers.
Comprehensive investigations into the use of HVO indicate that, despite consistent benefits in terms of particulate matter reduction and combustion performance, its influence on NOx emissions remains uncertain, as highlighted by varying and sometimes contradictory findings depending on engine configuration and operating conditions [37,38].
To comprehensively evaluate the influence of HVO on engine performance parameters, this study investigates neat HVO and its blends with diesel fuel (20% and 50% v/v HVO) at a constant engine speed under varying load conditions of a compression ignition engine. This approach was selected to replicate the operating conditions of marine auxiliary diesel engines (generator sets) and diesel-electric propulsion systems operating under variable load profiles [39,40].
The objective of this study was to evaluate the influence of engine load and EGR valve position on energy–economic indicators (brake specific fuel consumption (BSFC) and brake thermal efficiency (BTE)), emission characteristics (CO2, NOx, and smoke opacity), and combustion parameters (mass burned fraction (MBF), in-cylinder pressure, start of injection (SOI), and rate of heat release (ROHR)) when using HVO and its blends with diesel fuel. The analysis of these parameters and their interrelationships enables a comprehensive assessment of the performance similarity between HVO-based fuels and conventional diesel fuel [41,42].

2. Materials and Methods

2.1. Experimental Engine and Test Bench

Experimental investigations were carried out using a four-cylinder compression ignition engine equipped with a turbocharger and an electronically controlled fuel injection system. The engine is fitted with a BOSCH VP37 (Robert Bosch GmbH, Stuttgart-Feuerbach, Germany) distributor-type high-pressure fuel pump, with the start of injection controlled by the engine control unit (ECU). Fuel is directly injected into the combustion chamber using a single-injection strategy.
The main technical parameters of the tested engine are listed in Table 1. The engine was installed on a laboratory test bench and coupled to a dynamometer for controlling engine load and speed. Engine torque and speed were recorded using dedicated measurement equipment integrated into the dynamometer system.
Fuel consumption was determined using a gravimetric method. The fuel tank was placed on electronic scales, and the consumed fuel mass was measured over a defined time interval. The intake air mass flow rate was measured using a BOSCH HFM5 (Robert Bosch GmbH, Stuttgart-Feuerbach, Germany) air flow meter. Intake manifold pressure was monitored using a precision pressure sensor installed in the intake manifold.
The concentrations of exhaust gas components (CO, CO2, HC, and NOx) and smoke opacity were measured using an AVL DiCom 4000 (AVL DiTEST, Graz, Austria) exhaust gas analyzer.
The experimental test bench and measurement equipment are shown in Figure 1.

2.2. In-Cylinder Pressure Measurement

In-cylinder pressure was measured using a piezoelectric pressure sensor AVL GH13P (AVL DiTEST, Graz, Austria) integrated into the glow plug location. The sensor sensitivity is approximately 15.8 pC/bar, and its linearity error does not exceed ±0.3% of full scale.
The signal from the pressure sensor was amplified using an AVL DiTEST DPM 800 (AVL DiTEST, Graz, Austria) charge amplifier and recorded as a function of crank angle. The crankshaft position was determined using a rotational position sensor mounted on the crankshaft. The pressure data were acquired over multiple engine cycles and averaged to reduce cycle-to-cycle variations.
The recorded pressure traces were subsequently used to determine combustion characteristics such as the rate of heat release and the pressure rise rate.
Jmse 14 00665 g001
Figure 1. The scheme of engine testing equipment: 1—tested CI engine; 2—engine load (brake); 3—external air mass flow meter; 4—turbocharger; 5—exhaust gas analyzer; 6—opacimeter; 7—air pressure sensor; 8—exhaust gas temperature sensor; 9—exhaust gas recirculation system; 10—intake air temperature sensor; 11—high pressure fuel pump; 12—fuel meter; 13—fuel tank; 14—engine ECU; 15—in-cylinder pressure recording equipment; 16—fuel injection timing sensor; 17—engine torque and speed meter; 18—fuel injector with timing sensor; 19—in-cylinder pressure sensor; 20—crankshaft position sensor; 21—EGR control after disconnecting the ECU.
Figure 1. The scheme of engine testing equipment: 1—tested CI engine; 2—engine load (brake); 3—external air mass flow meter; 4—turbocharger; 5—exhaust gas analyzer; 6—opacimeter; 7—air pressure sensor; 8—exhaust gas temperature sensor; 9—exhaust gas recirculation system; 10—intake air temperature sensor; 11—high pressure fuel pump; 12—fuel meter; 13—fuel tank; 14—engine ECU; 15—in-cylinder pressure recording equipment; 16—fuel injection timing sensor; 17—engine torque and speed meter; 18—fuel injector with timing sensor; 19—in-cylinder pressure sensor; 20—crankshaft position sensor; 21—EGR control after disconnecting the ECU.
Jmse 14 00665 g001

2.3. Fuels and Test Conditions

The experimental study was performed using conventional diesel fuel and its blends with HVO. Fuel blends were prepared by mixing diesel and HVO in the required volumetric proportions prior to testing (Table 2).
Engine experiments were conducted under steady-state operating conditions. Measurements were performed at selected engine loads while maintaining constant engine speed. The tested operating regimes were selected to represent typical operating conditions of marine auxiliary engines. During the first test phase, the EGR ratio was adjusted between 0 and 0.3 using a modulated control signal, and all tested fuels were examined over this range. During the second phase of the test, the EGR ratio was controlled by the engine ECU; under these conditions, the recirculation rate was determined by engine load and fuel composition. This broad experimental approach was adopted to assess the effect of exhaust gas recirculation over a wide range, especially at low EGR levels between 0 and 0.20. Investigating such low EGR rates, i.e., below those typically commanded by the ECU, is particularly important for fuels with a high HVO content, because HVO substantially lowers NOx emissions. This creates the possibility of reducing the EGR ratio while achieving an additional decrease in smoke emissions.
During the experiments, the engine coolant temperature was maintained at approximately 85–90 °C and the lubricating oil temperature was kept near 100 °C to ensure stable operating conditions. All measurements were repeated six times, and the average values were used for further analysis.

2.4. Combustion Analysis

Combustion process analysis was performed using experimental in-cylinder pressure data. The heat release rate and pressure rise rate were determined using AVL BOOST (Version 2021 R2, AVL, Graz, Austria) software with the BURN sub-module. This tool calculates combustion parameters based on measured pressure traces and thermodynamic relationships describing heat transfer and combustion processes inside the cylinder.
The obtained combustion parameters were used to evaluate the influence of different fuel compositions on the combustion process in the compression ignition engine.

3. Results and Discussion

3.1. Combustion Indicators

During tests with D_100, HVO_100, and their mixtures, with EGR and SOI managed by the engine ECU, it was found that the ECU applies different EGR-valve opening and SOI settings depending on the fuel composition [36,43,44,45]. These control-induced variations are important when interpreting combustion behavior, as well as the engine’s energy and environmental performance. At low load (BMEP = 0.2 MPa), replacing D_100 with HVO_100 reduced the EGR ratio from 0.41 to 0.36 (12%) (Figure 2a). As the load increased to BMEP = 0.8 MPa, the ECU decreased the EGR ratio for D_100 to 0.16 to avoid a pronounced rise in smoke. Under the same condition, the EGR ratio for HVO_100 decreased to 0.09, which is 43% lower than for D_100.
For different fuels and their mixtures, the ECU also does not maintain another key engine control parameter, the start of fuel injection (SOI), at a constant value. With EGR off at BMEP = 0.2 MPa, the ECU maintained SOI = 2.0 CAD BTDC for D_100, whereas for HVO_100 it set SOI = 2.5 CAD BTDC (Figure 2b). This is attributable to the higher volumetric consumption of HVO_100 compared with D_100 due to the lower fuel density, even though the lower heating value of HVO is higher than that of diesel [38,46]. When the load was increased to BMEP = 0.8 MPa, SOI advanced in both cases as the injected fuel quantity increased, reaching 3.7 and 5.2 CAD BTDC for D_100 and HVO_100, respectively. With EGR on, the ECU further advanced injection timing to compensate for slower combustion [47] and the slightly higher fuel consumption at low load. For D_100, SOI was 2.3 CAD BTDC, while for HVO_100 it was 2.9 CAD BTDC. At BMEP = 0.8 MPa, the SOI differences between fuels were governed not only by the different volumetric fuel consumption but also by the different EGR ratios. With EGR on, SOI advanced to 4.1 and 5.7 CAD BTDC for D_100 and HVO_100, respectively. The mixtures D80_HVO20 and D50_HVO50 showed analogous SOI trends, scaling with the relative proportions of D_100 and HVO_100.
To investigate the influence of low EGR ratios on energy and emissions performance for different fuel compositions in more detail, tests were conducted at BMEP = 0.4 MPa with the EGR ratio controlled independently of the ECU by means of a simulated signal driving the vacuum EGR control valve. The results in Figure 3a show that, as the EGR ratio increases from 0 to 0.3 at BMEP = 0.4 MPa, the air excess ratio decreases progressively from 2.8 to 1.75. Only a minor effect of fuel composition on the air excess ratio was observed when EGR = 0, with D_100 exhibiting a slightly higher air excess ratio than HVO_100. As the EGR ratio increases, mixtures with a higher HVO content require more air mass for stoichiometric combustion because of the higher hydrogen content. However, this increased air requirement is partially offset by the lower fuel mass consumption associated with the higher lower heating value of HVO_100. The air excess ratio is also affected by a reduction in turbocharger boost pressure, which decreases the volumetric efficiency.
With EGR off and D_100 operation, increasing BMEP from 0.2 to 0.8 MPa reduced the air excess ratio from 3.6 to 1.65 due to the higher cycle fuel quantity (Figure 3b). For diesel blended with HVO (20% and 50%) and for neat HVO_100, the air excess ratio remained close to these values, at approximately 3.65 and 1.6 across the same load range. With EGR on, where the EGR ratio and SOI are controlled by the engine ECU as described above, the ECU’s EGR strategy markedly alters the air excess ratio. Owing to the higher EGR ratio with D_100 (Figure 1), the lowest air excess ratios were obtained, namely 2.05 and 1.25 at BMEP = 0.2 and 0.8 MPa, respectively. As the HVO share increases, the ECU reduces the EGR ratio in response to higher volumetric fuel consumption, and for HVO_100, the air excess ratio becomes higher, reaching 2.33 and 1.44, respectively. Overall, fuel-composition-driven changes in EGR ratio, air excess ratio, and SOI substantially influence the combustion process and the associated energy and emissions indicators [43,48,49].
At one operating point within the engine test map, a combustion-indicator analysis was performed to evaluate the combustion process of the fuels and their mixtures investigated in this study. Combustion indicators were determined using the AVL BOOST Burn subroutine under medium engine speed (n = 2000 rpm) and a medium load of 120 Nm (BMEP = 0.8 MPa), with EGR off. This operating condition is representative of marine ship generator/auxiliary engine duty, i.e., a medium-speed diesel engine operating regime. The start of injection (SOI) was controlled by the engine ECU, and an SOI advance from 2.9 CAD BTDC to 4.2 CAD BTDC was observed as the HVO concentration in the fuel increased from 0% to 100%. This can be attributed to the increased volumetric fuel consumption caused by the lower fuel density [44], which, in turn, affects the ECU control strategy and results in an earlier SOI.
Differences in the physicochemical fuel properties, particularly the higher cetane number of HVO and the resulting shorter ignition delay (ID), together with the advanced SOI (1.3 CAD) relative to neat diesel, affect the combustion indicators. Figure 4a shows that the in-cylinder pressure traces are very similar; however, as the HVO share increases, a moderate rise in maximum cylinder pressure (pmax) and a slight shift in combustion phasing are observed. When moving from D_100 to HVO_100 at the same operating point, pmax increases by approximately 4–5%, and its occurrence advances by up to 0.8 CAD. This trend is consistent with the expected behavior of HVO as a high-cetane paraffinic fuel: a slightly earlier SOI and shorter ID advance the heat release, which can increase the pressure during the early expansion phase. Published studies likewise indicate that HVO often leads to earlier combustion phasing and similar or slightly higher pmax values, depending on the engine ECU calibration. When using the other investigated mixtures (D80_HVO20 and D50_HVO50), the in-cylinder pressure peak is attained at intermediate values between those for D_100 and HVO_100.
The rate of heat release (ROHR) profile for HVO_100 rises earlier (Figure 4b), indicating an earlier start of combustion (SOC), which can be ascribed to the advanced SOI and the shorter ignition delay (ID). During the premixed combustion phase, the peak ROHR shows only minor variation across the fuels, remaining at about 49 J/CAD and occurring at approximately 14 CAD ATDC. For HVO_100, the diffusion and late-combustion phases also end sooner. This trend accords with the literature: owing to its paraffinic composition and low aromatic content, HVO typically promotes more favorable oxidation and suppresses late combustion. Nevertheless, the resulting ROHR shape is highly sensitive to the injection strategy and to the influence of fuel density and viscosity on injection dynamics and the delivered fuel quantity [44].
The combustion evolution of the tested fuels is further elucidated by the mass burned fraction (MBF). The MBF traces indicate that, as the HVO share increases, combustion occurs slightly earlier (the MBF curve shifts to the left) and, over certain intervals, proceeds marginally faster (Figure 5a). HVO_100 exhibits a higher MBF in the early interval (approximately 5–25 CAD), indicating an earlier combustion onset and an earlier CA10. Specifically, CA10 advances from 7.2 CAD (D_100) to 6.4 CAD (HVO_100). CA50 advances for HVO50_D50 and HVO_100 by 0.60 and 1.00 CAD, respectively. In the later phase (approximately 30–50 CAD), HVO50_D50 reaches a slightly higher final MBF and approaches 1.0 more rapidly, which can be interpreted as a slightly more complete combustion process. The faster burn observed at 50% HVO is attributed to the reduced fuel viscosity, which improves spray atomization, and to the higher hydrogen content of HVO, which together enhance combustion quality and burn rate. Increasing the HVO share to 100% does not further shorten the combustion duration because the lower density increases the injected fuel volume and tends to extend the injection duration. Accordingly, CA10–CA90 decreases by approximately 2.0% (HVO20_D80), 9.0% (HVO50_D50), and 3.8% (HVO_100). The relatively small changes in combustion indicators suggest that HVO can be used as a drop-in fuel without engine recalibration; however, a shift in combustion phasing and changes in heat-release characteristics are commonly observed, particularly due to differences in cetane number and physical properties [50].
The combustion-indicator traces show that increasing the HVO fraction slightly advances combustion phasing and increases both the in-cylinder pressure and temperature. However, the calculated in-cylinder temperature for HVO_100 shows a more pronounced rise only at the onset of the premixed combustion phase (Figure 5b), whereas the highest peak temperature is obtained with D_100 (1971 K). By contrast, for HVO_100, the peak temperature reaches 1955 K and occurs 2 CAD earlier. Under real ship operating conditions, the use of different diesel–HVO mixtures may require revisiting SOI settings and re-evaluating EGR strategies to preserve the targeted combustion–emissions trade-off [23].
These results indicate that the observed shift in combustion phasing was caused not only by the physicochemical properties of HVO but also by the fuel-dependent ECU response, particularly the simultaneous variation in EGR rate and SOI. Therefore, under stock control conditions, the combustion behavior should be interpreted as the combined result of fuel properties and closed-loop engine management rather than as a pure fuel effect.

3.2. Energy Performance Indicators

When assessing the engine energy performance, the brake specific fuel consumption (BSFC), brake specific fuel volume consumption (BSFC_V), and brake thermal efficiency (BTE) are evaluated [51,52]. At BMEP = 0.4 MPa, increasing the EGR ratio from 0 to 0.25 leads to an approximately 4% increase in BSFC (Figure 6a), which is attributable to degraded combustion quality and slower heat release due to CO2 dilution from recirculated exhaust gases, together with the reduced air excess ratio (Figure 3). With EGR off, increasing the HVO share from 0% to 100% reduces BSFC from 263 g/kWh to 255 g/kWh (3.4%), and a comparable relative reduction is maintained at BMEP = 0.4 MPa as the EGR ratio is increased to 0.25.
Across the load range from BMEP = 0.2 MPa to 0.8 MPa, with the EGR ratio regulated by the engine ECU, BSFC decreases with increasing load, whereas EGR increases BSFC. For HVO_100 with EGR on, the fuel consumption becomes comparable to, or lower than, that of D_100 with EGR off (Figure 6b). For D_100 with EGR off, BSFC values of 338 g/kWh and 236 g/kWh were measured at BMEP = 0.2 MPa and 0.8 MPa, respectively. For HVO_100 with EGR on, BSFC reaches 334 g/kWh and 234 g/kWh at the same loads. The close agreement between D_100 (EGR off) and HVO_100 (EGR on) is explained by the lower EGR ratio applied by the ECU with HVO_100 (Figure 2a), together with faster (Figure 5a) and thermally more effective combustion when operating on HVO_100. Since the ECU applies a higher EGR ratio for D_100 than for HVO_100 (Figure 2a), at BMEP = 0.8 MPa, the BSFC for D_100 increases to 251 g/kWh, which is 6.8% higher than under EGR-off operation. For HVO_100 at BMEP = 0.8 MPa, BSFC increases from 228 g/kWh with EGR off to 234 g/kWh (2.6%) with EGR on.
From a user perspective, volumetric fuel consumption is often more relevant than mass-based consumption. BSFC_V is strongly influenced by fuel density, which for HVO_100 is 6% lower than for D_100 [53]. Figure 7a shows that at BMEP = 0.4 MPa with EGR = 0, increasing the HVO share to 20% (D80_HVO20) yields a BSFC_V of 317 mL/kWh, which is close to the value obtained with HVO_100. Increasing the HVO share to 50% and 100% increases BSFC_V by approximately 1% and 3.8%, respectively. As the EGR ratio increases, BSFC_V rises in line with the increase in BSFC; however, at a fixed EGR level, the relative differences in volumetric consumption between fuels remain consistent with those noted above.
With EGR off at BMEP = 0.8 MPa, BSFC_V for D_100 decreases to 285 mL/kWh, while HVO_100 reaches 292 mL/kWh (2.5% higher) (Figure 7b). The narrowing difference in BSFC_V between D_100 and HVO_100 with increasing BMEP indicates that HVO utilization becomes more effective at higher loads. With ECU-controlled EGR at BMEP = 0.8 MPa, the volumetric consumption for D_100 is 303 mL/kWh, whereas for HVO_100 it is 299 mL/kWh (1.3% lower). Under EGR-on operation, the lowest BSFC_V of 294 mL/kWh (3.0% lower than D_100) is achieved with the D50_HVO50 mixture.
The thermal efficiency of an engine depends on fuel mass consumption, the lower heating value, and the delivered brake power. At BMEP = 0.4 MPa with EGR off, BTE is 0.319 for D_100; with HVO shares of 20%, 50%, and 100%, BTE increases to 0.323, 0.324, and 0.325, corresponding to improvements of 1.3%, 1.6%, and 1.9%, respectively (Figure 8a). When the EGR ratio is increased to 0.2, the reduced combustion efficiency lowers BTE for D_100 to 0.312 (2.2%), whereas with HVO-containing fuels, BTE increases with the HVO fraction, following the same trend as noted above.
At BMEP = 0.8 MPa with EGR off, BTE increases to 0.356 for D_100 and to 0.362 (1.7%) for HVO_100 (Figure 8b). At BMEP = 0.8 MPa with EGR on, the BTE of D_100 decreases to 0.335 (5.9%), while the BTE of HVO_100 decreases to 0.354, remaining comparable to the efficiency of D_100 with EGR off. The BTE of the D50_HVO50 mixture also remains high (0.351). This results, on the one hand, from more efficient combustion with higher HVO content and, on the other hand, from the ECU response to increased volumetric fuel consumption, namely reducing the EGR ratio and advancing SOI, which confirms the engine control calibration requirement [54].

3.3. Environmental Performance Indicators

The brake-specific carbon dioxide emissions [g/kWh] depend on fuel mass consumption, fuel composition, particularly the carbon content, and combustion completeness [53,55,56]. At BMEP = 0.4 MPa with EGR off, CO2 emissions amount to 775 g/kWh for D_100, whereas for H100 they reach 735 g/kWh, which is 5.2% lower (Figure 9a). This reduction is associated with a decrease in the fuel C/H ratio from 6.8 to 5.6 (23%) and a 3.0% decrease in BSFC (Figure 6a). When the EGR ratio is increased to 0.2, CO2 emissions for all fuels and mixtures rise proportionally by about 6.0%.
Over the load range from BMEP = 0.2 MPa to 0.8 MPa with EGR off, the reduction in CO2 emissions for HVO_100 relative to D_100 varies between 4.7% and 6.3% (Figure 9b). The lowest CO2 emissions are obtained at BMEP = 0.8 MPa, where D_100 emits 716 g/kWh and H100 emits 678 g/kWh. When the engine is operating in EGR-on mode, more pronounced differences in CO2 emissions are observed for different fuels at different loads, as the engine ECU sets the EGR ratio individually for different fuels and at different loads (Figure 2a). With D_100, the ECU sets the highest EGR ratio due to the shortest injection duration, and the corresponding increase in CO2 emissions is the greatest. As the HVO fraction increases, the longer injection duration prompts the ECU to reduce the EGR ratio. Consequently, through the combined effects of the reduced EGR ratio and the fuel properties, the CO2 emissions of HVO and the D50_HVO50 mixture approach, or even fall below, those of D_100 under EGR-off operation.
Nitrogen oxide emissions depend on the EGR ratio, but the physical and chemical properties of the fuel, the air excess ratio, and the SOI also have a significant impact [53,56]. The analysis of these parameters is provided in Section 2.1. At BMEP = 0.4 MPa with EGR off, NOx emissions are 3.8 g/kWh for D_100 and 3.4 g/kWh for H100 (Figure 10a), corresponding to a 10% reduction. The lower NOx emissions with HVO_100 are primarily attributed to a lower in-cylinder temperature during combustion. On the one hand, the larger cyclic injected volume with HVO_100 leads the ECU to advance SOI (Figure 2b), which tends to increase the combustion temperature because heat release occurs closer to TDC. On the other hand, several factors act to reduce the combustion temperature: the higher cetane number shortens the ignition delay; the lower HVO_100 density slows the rate of energy delivery to the cylinder and reduces the intensity of the rate of heat release; and the higher hydrogen content in the fuel increases the stoichiometric air requirement, thereby lowering the air excess ratio. When the EGR ratio is increased to 0.2, combustion dilution reduces NOx emissions for D_100 by 43% to 2.15 g/kWh, while for HVO_100 the emission level decreases to 1.95 g/kWh, remaining 9% lower.
When emissions are analyzed over the BMEP range from 0.2 MPa to 0.8 MPa with EGR off, NOx emissions increase with load, and the effect of HVO varies with operating point (Figure 10b). At BMEP = 0.2 MPa, NOx emissions are 3.7 g/kWh for D_100 and 3.0 g/kWh for HVO_100, representing a 19% reduction. When the load is increased fourfold to BMEP = 0.8 MPa, D_100 NOx emissions rise to 4.2 g/kWh (13%), whereas with HVO_100 they decrease to 3.9 g/kWh, which corresponds to a 7% reduction for HVO_100 compared to D_100. This indicates that, at higher loads, HVO exerts a weaker influence on the combustion process and the in-cylinder temperature.
With EGR on, NOx emissions for D_100 are 1.5 g/kWh at BMEP = 0.2 MPa and increase to 2.3 g/kWh at BMEP = 0.8 MPa, which corresponds to reductions of 59% and 45% relative to EGR off. When the HVO additive is introduced under EGR-on operation, NOx does not decrease as might be expected; instead, it increases because the ECU reduces the EGR ratio and advances SOI. When using HVO without recalibration, NOx may not decrease and may, in some modes, even increase [44]. The largest NOx increases with HVO_100 under EGR-on operation occur at BMEP = 0.8 MPa, where emissions reach 3.1 g/kWh, which is 34% higher than for D_100.
In all cases, increasing the HVO fraction in diesel mixtures reduces smoke opacity and soot [48,53]. At BMEP = 0.4 MPa with EGR off, smoke opacity is 0.097 m−1 for D_100 and 0.070 m−1 for HVO_100, which is 28% lower (Figure 11a). When the EGR ratio is increased to 0.2, smoke opacity increases by approximately a factor of three due to the reduced oxygen availability in the cylinder and combustion suppression, reaching 0.29 m−1 for D_100 and 0.22 m−1 for HVO_100. For HVO_100, smoke remains 24% lower due to several factors, including lower carbon content, a simpler molecular structure, lower viscosity, and improved spray quality, among others.
As the load increases to BMEP = 0.8 MPa with EGR off, smoke opacity rises because of the decreasing air excess ratio, reaching 0.197 m−1 for D_100 and 0.113 m−1 for H100 (Figure 11b). The difference between D_100 and HVO_100, therefore, increases to 42%. An even larger smoke difference is observed under EGR-on operation. The smoke peak occurs at BMEP = 0.6 MPa, where D_100 reaches 1.13 m−1, and HVO reaches 0.51 m−1. At this operating point, D_100 smoke exceeds HVO_100 by 55% because the ECU maintains an EGR ratio of 0.25 for D_100 and 0.19 for HVO_100 (Figure 1). The combined NOx and smoke results confirm that HVO widens the potential operating window for cleaner combustion, since smoke is reduced substantially even at elevated load. However, when the ECU responds by lowering EGR and advancing SOI, part of this benefit is shifted from particulate reduction towards higher combustion intensity, which may limit the expected decrease in NOx. This shows that when using fuel mixtures, engine control calibration is required [44,57].
When the load is increased to BMEP = 0.8 MPa, smoke decreases due to the lower EGR ratio and the higher combustion temperature; however, the difference between D_100 and HVO_100 increases further to 63%. It can also be observed that the smoke opacity of the D50_HVO50 mixture is close to that of HVO_100, reflecting the pronounced influence of HVO on the fuel properties and environmental pollution.
In the context of marine auxiliary engines operating at nearly constant speed and variable load, these findings suggest that diesel–HVO blends can be introduced without major hardware changes; however, the full benefit of HVO can only be realized when EGR and injection settings are adjusted to the specific blend composition.

4. Conclusions

The experimental results presented and discussed above allow a comprehensive assessment of the influence of HVO and its blends with diesel fuel on the combustion process, energy performance, and emission characteristics of a compression ignition engine operating under different load conditions and EGR ratios. The main findings of this study are summarized as follows.
  • Increasing the HVO share in diesel mixtures changes key fuel properties, including density and ignition characteristics, which in turn alter combustion development and the resulting energy and emissions indicators.
  • The higher heating value of HVO reduces fuel mass consumption by up to 3.4%, while improved combustion efficiency increases brake thermal efficiency by as much as 1.9% when neat HVO is used. At the same time, the approximately 6% lower density of HVO increases volumetric fuel consumption by up to 3.8%, raises the injected fuel volume, and prolongs the injection duration. For this reason, the EGR ratio at a given operating point was reduced by 12–43%, indicating that the ECU response varied with fuel properties rather than remaining identical for all tested fuels.
  • In general, the addition of HVO reduced smoke opacity by 24–42% and decreased CO2 emissions by 4.7–5.2%, mainly due to lower carbon intensity (C/H ratio) and reduced brake-specific fuel consumption. However, the greater injected fuel volume can lead to an ECU-driven reduction in EGR and an advance in SOI, which may partly offset the anticipated reduction in NOx emissions. In EGR-on modes, high-HVO fuels may exhibit NOx levels close to those of D100, because of the combined effect of lower EGR dilution and earlier combustion phasing.
  • The results indicate that replacing conventional fuels with various mixtures without recalibration can shift the combustion and emissions trade-off, especially when maintaining unadjusted EGR and injection control. Achieving the best overall performance requires fuel-specific optimization of EGR and SOI control maps when changing the diesel–HVO composition.
  • These outcomes are particularly relevant for medium-speed marine diesel generator and auxiliary engines, where maintaining the desired efficiency, NOx, and smoke trade-off requires fuel-specific optimization of EGR and SOI control maps when the diesel–HVO composition is changed.

Author Contributions

Conceptualization, A.R. and J.Ž.; methodology, S.P.; software, A.R.; validation, A.R., J.Ž. and S.P.; formal analysis, A.R.; investigation, A.R., J.Ž. and S.P.; resources, J.Ž.; data curation, A.R.; writing—original draft preparation, A.R. and J.Ž.; writing—review and editing, S.P.; visualization, A.R.; supervision, J.Ž. All authors have read and agreed to the published version of the manuscript.

Funding

This research received no external funding.

Institutional Review Board Statement

Not applicable.

Informed Consent Statement

Not applicable.

Data Availability Statement

Data is contained within the article. The original contributions presented in this study are included in the article. Further inquiries can be directed to the corresponding author.

Acknowledgments

A part of the results of the research described in this article were obtained using the engine simulation tool AVL BOOST, acquired by signing a Cooperation Agreement between AVL Advanced Simulation Technologies and the faculty of the Transport Engineering of Vilnius Gediminas Technical University.

Conflicts of Interest

The authors declare no conflicts of interest.

Abbreviations

The following abbreviations are used in this manuscript:
ATDCAfter top dead center
BMEPBrake mean effective pressure
BSFCBrake specific fuel consumption
BTDCBefore top dead center
BTEBrake thermal efficiency
CADCrank angle degree
COCarbon monoxide
CO2Carbon dioxide
D_100Pure diesel fuel
D50_HVO5050/50 (v/v) diesel–HVO blend
D80_HVO2080/20 (v/v) diesel–HVO blend
ECUElectronic control unit
EGRExhaust gases recirculation
FAMEFatty acid methyl esters
GHGGreenhouse gas
HCHydrocarbons
HVOHydrotreated vegetable oil
HVO_100Pure hydrotreated vegetable oil
IDIgnition delay
IMOInternational Maritime Organization
MBFMass burned fraction
NOxNitrogen oxides
NSCRNon-selective catalytic reduction
pmaxMaximum cylinder pressure
ROHRRate of heat release
SCRSelective catalytic reduction
SOIStart of fuel injection

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Figure 2. Engine control parameters at different engine loads and fuel compositions: (a) EGR ratio at different engine loads and for different fuel compositions; (b) Start of fuel injection at different engine loads and for different fuel compositions.
Figure 2. Engine control parameters at different engine loads and fuel compositions: (a) EGR ratio at different engine loads and for different fuel compositions; (b) Start of fuel injection at different engine loads and for different fuel compositions.
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Figure 3. Variation in air excess ratio at different EGR ratios, engine loads, and fuel compositions: (a) air excess ratio at different EGR ratios for different fuel compositions; (b) air excess ratio at different engine loads and for different fuel compositions.
Figure 3. Variation in air excess ratio at different EGR ratios, engine loads, and fuel compositions: (a) air excess ratio at different EGR ratios for different fuel compositions; (b) air excess ratio at different engine loads and for different fuel compositions.
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Figure 4. Combustion parameters for different fuel compositions: (a) in-cylinder pressure traces for different fuel compositions; (b) rate of heat release depending on fuel composition.
Figure 4. Combustion parameters for different fuel compositions: (a) in-cylinder pressure traces for different fuel compositions; (b) rate of heat release depending on fuel composition.
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Figure 5. Combustion process parameters for different fuel compositions: (a) mass burned fraction depending on fuel composition; (b) in-cylinder temperature, depending on the fuel composition.
Figure 5. Combustion process parameters for different fuel compositions: (a) mass burned fraction depending on fuel composition; (b) in-cylinder temperature, depending on the fuel composition.
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Figure 6. Variation in brake specific fuel consumption under different EGR ratios, engine loads, and fuel compositions: (a) brake specific fuel mass consumption (BSFC) at different EGR ratios for different fuel compositions; (b) brake specific fuel mass consumption at different engine loads and for different fuel compositions.
Figure 6. Variation in brake specific fuel consumption under different EGR ratios, engine loads, and fuel compositions: (a) brake specific fuel mass consumption (BSFC) at different EGR ratios for different fuel compositions; (b) brake specific fuel mass consumption at different engine loads and for different fuel compositions.
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Figure 7. Variation in brake specific fuel volume consumption under different EGR ratios, engine loads, and fuel compositions: (a) brake specific fuel volume consumption (BSFC_V) at different EGR ratios for different fuel compositions; (b) brake specific fuel volume consumption (BSFC_V) at different engine loads and for different fuel compositions.
Figure 7. Variation in brake specific fuel volume consumption under different EGR ratios, engine loads, and fuel compositions: (a) brake specific fuel volume consumption (BSFC_V) at different EGR ratios for different fuel compositions; (b) brake specific fuel volume consumption (BSFC_V) at different engine loads and for different fuel compositions.
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Figure 8. Variation in engine brake thermal efficiency under different EGR ratios, engine loads, and fuel compositions: (a) engine brake thermal efficiency (BTE) at different EGR ratios for different fuel compositions; (b) engine brake thermal efficiency (BTE) at different engine loads and for different fuel compositions.
Figure 8. Variation in engine brake thermal efficiency under different EGR ratios, engine loads, and fuel compositions: (a) engine brake thermal efficiency (BTE) at different EGR ratios for different fuel compositions; (b) engine brake thermal efficiency (BTE) at different engine loads and for different fuel compositions.
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Figure 9. Variation of carbon dioxide emissions under different EGR ratios, engine loads, and fuel compositions: (a) carbon dioxide (CO2) emissions at different EGR ratios for different fuel compositions; (b) carbon dioxide (CO2) emissions at different engine loads and for different fuel compositions.
Figure 9. Variation of carbon dioxide emissions under different EGR ratios, engine loads, and fuel compositions: (a) carbon dioxide (CO2) emissions at different EGR ratios for different fuel compositions; (b) carbon dioxide (CO2) emissions at different engine loads and for different fuel compositions.
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Figure 10. Variation of nitrogen oxides emissions under different EGR ratios, engine loads, and fuel compositions: (a) nitrogen oxides (NOx) emissions at different EGR ratios for different fuel compositions; (b) nitrogen oxides (NOx) emissions at different engine loads and for different fuel compositions.
Figure 10. Variation of nitrogen oxides emissions under different EGR ratios, engine loads, and fuel compositions: (a) nitrogen oxides (NOx) emissions at different EGR ratios for different fuel compositions; (b) nitrogen oxides (NOx) emissions at different engine loads and for different fuel compositions.
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Figure 11. Variation in smoke opacity under different EGR ratios, engine loads, and fuel compositions: (a) smoke opacity at different EGR ratios for different fuel compositions; (b) smoke opacity at different engine loads and for different fuel compositions.
Figure 11. Variation in smoke opacity under different EGR ratios, engine loads, and fuel compositions: (a) smoke opacity at different EGR ratios for different fuel compositions; (b) smoke opacity at different engine loads and for different fuel compositions.
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Table 1. Technical specifications of the tested compression ignition engine.
Table 1. Technical specifications of the tested compression ignition engine.
ParameterSpecification
Engine displacement1896 cm3
Cylinders/valve train4/OHC
Compression ratio19.5:1
Rated power66 kW at 4000 rpm
Maximum torque180 Nm at 2000–2500 rpm
Cylinder bore79.5 mm
Piston stroke95.5 mm
Fuel injection systemDirect injection (single-stage)
Injection pump typeAxial-piston distributor-type pump
Injector nozzleHole-type
Nozzle–holder assemblyTwo-spring design
Nozzle opening pressure190–200 bar
Table 2. Main properties of the tested fuels and fuel blends.
Table 2. Main properties of the tested fuels and fuel blends.
PropertiesD_100D80_HVO20D50_HVO50HVO_100
Diesel/HVO content, % (v/v)100/080/2050/500/100
Stoichiometric air to fuel ratio, kg air/1 kg fuel14.5014.6114.7915.10
C/H ratio6.806.576.225.60
Density at 15 °C, g/mL830.0820.5805.8780.0
Lower heating value (LHV), MJ/kg42.8242.9743.2143.63
Cetane number50.854.761.877.4
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MDPI and ACS Style

Rimkus, A.; Žaglinskis, J.; Pukalskas, S. Effects of Hydrotreated Vegetable Oil and Diesel Blends on Combustion, Energy Performance, and Emissions of a Compression Ignition Engine Under EGR-Controlled Operation. J. Mar. Sci. Eng. 2026, 14, 665. https://doi.org/10.3390/jmse14070665

AMA Style

Rimkus A, Žaglinskis J, Pukalskas S. Effects of Hydrotreated Vegetable Oil and Diesel Blends on Combustion, Energy Performance, and Emissions of a Compression Ignition Engine Under EGR-Controlled Operation. Journal of Marine Science and Engineering. 2026; 14(7):665. https://doi.org/10.3390/jmse14070665

Chicago/Turabian Style

Rimkus, Alfredas, Justas Žaglinskis, and Saugirdas Pukalskas. 2026. "Effects of Hydrotreated Vegetable Oil and Diesel Blends on Combustion, Energy Performance, and Emissions of a Compression Ignition Engine Under EGR-Controlled Operation" Journal of Marine Science and Engineering 14, no. 7: 665. https://doi.org/10.3390/jmse14070665

APA Style

Rimkus, A., Žaglinskis, J., & Pukalskas, S. (2026). Effects of Hydrotreated Vegetable Oil and Diesel Blends on Combustion, Energy Performance, and Emissions of a Compression Ignition Engine Under EGR-Controlled Operation. Journal of Marine Science and Engineering, 14(7), 665. https://doi.org/10.3390/jmse14070665

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