Research on Fuel Efficiency and Emissions of Converted Diesel Engine with Conventional Fuel Injection System for Operation on Natural Gas

Abstract

This paper presents a study on the energy efficiency and emissions of a converted high-revolution bore 79.5 mm/stroke 95 mm engine with a conventional fuel injection system for operation with dual fuel feed: diesel (D) and natural gas (NG). The part of NG energy increase in the dual fuel is related to a significant deterioration in energy efficiency (ηi), particularly when engine operation is in low load modes and was determined to be below 40% of maximum continuous rating. The effectiveness of the D injection timing optimisation was established in high engine load modes within the range of a co-combustion ratio of NG ≤ 0.4: with an increase in ηi, compared to D, the emissions of NOx+ HC decreased by 15% to 25%, while those of CO2 decreased by 8% to16%; the six-fold CO emission increase, up to 6 g/kWh, was unregulated. By referencing the indicated process characteristics of the established NG phase elongation in the expansion stroke, the combustion time increase as well as the associated decrease in the cylinder excess air ratio (α) are possible reasons for the increase in the incomplete combustion product emission.

Full text

energies
Article
Research on Fuel Efficiency and Emissions of
Converted Diesel Engine with Conventional Fuel
Injection System for Operation on Natural Gas
Sergejus Lebedevas 1, Saugirdas Pukalskas 2, Vygintas Daukšys 1, Alfredas Rimkus 2,
Mindaugas Melaika 2and Linas Jonika 1, *
1Department of Marine Engineering, Faculty of Marine Technologies and Natural Sciences, Klaipeda
University, Bijunu Str. 17, LT-91225 Klaipeda, Lithuania; sergejus.lebedevas@ku.lt (S.L.);
vygintasdauksys@gmail.com (V.D.)
2Department of Automobile Engineering, Transport Engineering Faculty, Vilnius Gediminas Technical
University, J. Basanaviˇciaus Str. 28, LT-03224 Vilnius, Lithuania; saugirdas.pukalskas@vgtu.lt (S.P.);
alfredas.rimkus@vgtu.lt (A.R.); mindaugas.melaika@gmail.com (M.M.)
*Correspondence: linas.jonika@apc.ku.lt
Received: 20 May 2019; Accepted: 19 June 2019; Published: 23 June 2019


Abstract:
This paper presents a study on the energy efficiency and emissions of a converted
high-revolution bore 79.5 mm/stroke 95 mm engine with a conventional fuel injection system for
operation with dual fuel feed: diesel (D) and natural gas (NG). The part of NG energy increase
in the dual fuel is related to a significant deterioration in energy efficiency
(ηi)
, particularly when
engine operation is in low load modes and was determined to be below 40% of maximum continuous
rating. The effectiveness of the D injection timing optimisation was established in high engine load
modes within the range of a co-combustion ratio of NG
≤
0.4: with an increase in
ηi
, compared to D,
the emissions of
NOx+
HC decreased by 15%
to
25%, while those of
CO2
decreased by 8% to16%;
the six-fold
CO
emission increase, up to 6 g/kWh, was unregulated. By referencing the indicated
process characteristics of the established NG phase elongation in the expansion stroke, the combustion
time increase as well as the associated decrease in the cylinder excess air ratio (
α
) are possible reasons
for the increase in the incomplete combustion product emission.
Keywords:
compression ignition engine; conventional fuel injection system; natural gas; energy and
emission indicators; fuel injection phase
1. Introduction
When comparing autonomous heat engines, modern compression ignition engines (CIEs) are
characterised by the highest energy efficiency [
1
]. This is one of the most important operational
indicators, and has a significant influence on the extensive use of these engines in transportation,
households, and non-road machines. According to greenhouse gas emission statistics data,
CIEs constitute the following parts of the total transport energy balance: 72.1% of the road vehicle
sector, 13.6% of marine transportation, and 0.5% of rail transport [
2
]. Although numerous CIE processes
involve heterogeneous fuel mixture combustion across a large temperature inconsistency (at local
temperatures up to 2500 to 3000 K and higher, with an air–fuel equivalent ratio up to 2 to 4 units),
a relatively high level of harmful component emission is generated in exhaust gas [
1
,
3
,
4
]. Among the
harmful exhaust components, the most dangerous to people’s health and fauna are nitrous oxides
and particulate matter (PM), which mainly consist of soot [
1
]. Soot particles of 10
µ
m and less may
cause airway diseases such as mesothelioma [
3
]. CIEs that burn fuel containing sulphur can release
sulphuric oxides that are the main cause of acid rain and can cause damage flora [5].
Energies 2019,12, 2413; doi:10.3390/en12122413 www.mdpi.com/journal/energies

Energies 2019,12, 2413 2 of 32
Therefore, research on the reduction of harmful emissions into the atmosphere as well as CIE
standardisation has become prevalent during recent decades (for example, EN590, EPA, 2012). One of
the most widespread technologies is the conversion of CIE for dual-fuel operation and the use
of petroleum-derived fuel or either compressed natural gas (CNG) or liquefied natural gas [
6
–
11
].
The introduction of NG into the cylinder can be achieved with either high-pressure injection or in
combination with liquid fuel spray or carburation with air into air inlet manifold. Owing to the inferior
NG auto-ignition properties, the fuel mixture may be ignited with electrical discharge, as with Otto
engines [
12
–
14
], or by using a pilot fuel portion or high reaction fuel (HRF) [
15
]. The second method
for igniting NG is used most frequently, owing to the easy technological realisation of converting CIE
to work with NG, without any need for substantial changes in the CIE structure, while the engine can
continue to operate on liquid fuel only [16].
These types of engines, also known as dual-fuel engines, can provide significant improvements
in efficiency and emissions [
16
–
25
]. For example, in research on a dual-fuel cargo truck engine [
26
]
a decrease in
NOx
of up to six times and decrease in
CO
of up to 83% were observed compared
to diesel only operation. The MARPOL 73/79 VI annex standard Tier III norms were achieved in a
Wärtsillä company average revolution 20DF ship engine when the engine operated with NG fuel
feed, without using secondary emission reduction technologies (such as selective catalytic reduction
technology) [
24
]. Compared to petroleum-based fuel, at least on a theoretical basis, the NG chemical
composition contains a carbon/hydrogen ratio that is more favourable for decreasing
CO2
emissions
responsible for the greenhouse effect by a quarter.
However, serious drawbacks exist when converting a functioning CIE for dual-fuel operation.
When using NG greenhouse gas
CO2
emissions are reduced because NG, when compared to diesel,
has a smaller proportion of carbon atoms, and also the calorific value of NG is 15–25% higher. Moreover,
even when there is an increase in HC emissions from using NG, the overall impact on the greenhouse
effect is low, when compared to using diesel only [
26
]. Also, operation during low load modes results
in instability, leading to a reduction in efficiency indicators, and during nominal power load operation
and with a large gas fraction of 90% to 95% in the fuel balance, this may result in knocking [
24
,
27
].
To solve these problems, motor methods are generally used: adjustment of the compression ratio
and combustion chamber optimisation [
10
,
28
,
29
], air vortex movement increase in the cylinder [
28
],
and exhaust gas recirculation (EGR) technology [
24
,
26
], among others. Research results attest to the
conclusion that depending on the operating conditions, the use of NG can increase [
24
] as well as reduce
instances of engine knocking phenomena [
27
]. For example, in numerical research [
24
] combustion
pressure high frequency pulsation manifests at early and late (by retarding injection timing more than
40 CAD BTDC) pilot fuel injection timing. While at the same time at injection timing from 25 to 40
CAD BTDC pulsations were observed to be minimal. Authors of this publication used a CIE with a
conventional fuel injection system. NG had a positive effect and led to a reduction of high frequency
pressure amplitude.
HRF characteristic optimisation is a prevalent method, which includes the pilot fuel portion
phase, pressure, injection law change, and multiphase injection [
30
]. The appeal of this method is
related to its relative simplicity, which is relevant for operational CIE conversion, with regulation
flexibility across a wide speed and load range, and importantly, enables improvements in engine energy
efficiency and emission parameters. The majority of research in this area has been based on complex
experimental investigation and internal cylinder mathematical modelling, with respect to the injection
phase influence on the operation mixture combustion dynamics. For example, in papers [
24
,
25
,
30
–
32
],
the physical mechanism and factors leading to the diesel and gas–environment air engine cylinder
chemical kinetics as well as process cylinder dynamics were investigated.
In [
30
], a one-cylinder (bore 137.2 mm/stroke 105.1 mm) engine with a common rail (CR) fuel
injection accumulative system was investigated, and two separate operational mixture combustion
physical mechanisms were established, depending on the HRF. The research was conducted on a 25%
partial load and partial part gas phase, with an energy value of co-combustion ratio of NG (CCR NG)

Energies 2019,12, 2413 3 of 32
of approximately 75%. During the late diesel injection time (DIT), which does not exceed 30
◦
CA
before top dead centre (BTDC), auto-ignition and combustion dynamics were observed, specific to
the CIE cycle: OH radicals were grouped closely with HRF torch, initiating intensive combustion in
accordance with the kinetic mechanism, with a further transition to flame diffusion transfer into the
peripheral combustion bowl zones. Therefore, a relatively small increase in the induction period can
be observed by advancing the DIT by 4 to 5
◦
CA BTDC, and the combustion process is postponed into
earlier crankshaft angles; the maximum pressure
Pmax
and maximum combustion temperature
Tmax
increase, which leads to superior
NOx
emission and efficiency (
ηe
). The investigated DIT range of 30 to
50
◦
CA BTDC operational mixture thermodynamics becomes insufficient for rapid HRF combustion.
The double increase in the induction period surpasses the HRF increase, bringing the combustion
process back to TDC. The active auto-ignition OH centres include a significant part of the combustion
chamber (CC) volume, and the fuel gas–air mixture combustion becomes single phase. The local
temperature field in the CC is equalised, thereby reducing the
NO2
emissions. Moreover, the reduction
not encompassed by the peripheral combustion zones in the CC also leads to a decrease in the emissions
of the incomplete combustion products CO and PM.
Analogous results were found in the DIT range between 50
◦
CA BTDC and 5
◦
CA after top
dead centre (ATDC) in a high-revolution (bore 85 mm/stroke 90 mm) engine with a common rail
system [
1
]. Furthermore, the influence of the DIT on the energy efficiency and emissions was established.
The authors of the paper have expanded their research by changing the HRF injection pressure in
the range of 500 to 1000 bar, optimising the DIT speed, and separating the HRF into component
parts, in combination with pressure optimisation. It was established that HRF injection during the
early phases is an effective method for improving dual-fuel engine energy efficiency and emissions.
A maritime purpose single-cylinder average rpm CIE, Wärtsillä 20DF (bore 200 mm/stroke 280 mm),
with a CR system, was also studied [24].
This research was conducted to establish the HRF distribution and structure, as well as influence
of the thermodynamic properties in a compression process on the HRF auto-ignition and fuel mixture
combustion process. The experimental research was conducted with medium, nominal engine loads at
constant revolutions; the HRF injection pressure was adjusted within the range of 1300 to 2100 bar;
the DIT ranged from 15 to 50
◦
CA BTDC. It was established that, in the DIT range of up to 30
◦
CA
BTDC, the HRF combustion reaction potency, as assessed by the localised HRF equivalence ratio,
was high, which is in accordance with the research results. A predominant
β
value range of 0.2 to
0.8 determines the comparatively short ignition delay of 10 to 17
◦
CA BTDC and intense combustion.
Later, the DIT is accompanied by a noticeable
β
=0.2 to 0.8 part reduction, causing a homogenous
operational mixture combustion close to the TDC.
Combustion obtains “soft” characteristics, equalising the localised temperature fields and reducing
emissions. Moreover, it was established that the fuel mixture thermodynamic parameter changes in
the cylinder (by advancing the inlet valve closing phase) are no less effective than the increased HRF
ignition delay period. It is important to note that, during this event, the DIT range is significantly
reduced: in the conducted experiment from 45
◦
CA BTDC up to 32
◦
CA BTDC, emissions were reduced
to or below the MARPOL 73/78 Tier III standard established values. Similar results were obtained
in an experiment with a six-cylinder engine (Bore 11.2 mm/Stroke 13.2 mm) [
33
] with a CR system,
by advancing the DIT up to 32
◦
CA BTDC (CCR NG 90%), and low load mode emissions were reduced.
As opposed to the experimental results from [
23
,
27
,
28
,
30
,
34
], in which engine parameter deterioration
resulted in an increase in hydrocarbon emissions, a reduction in the energy efficiency was observed.
A short overview of the research demonstrates the significance of improving modern engine
energy efficiency and emission parameters, changing their operation to NG. However, the separate
CIE category conversion to NG operation faces several difficulties. Firstly, problems include models
with traditional mechanical fuel injection systems, which exhibit the characteristic of a limited DIT
range change and a substantially lower diesel fuel injection pressure compared to the CR system.
However, numerous operational CIEs exist with traditional mechanical fuel injection systems. Therefore,

Energies 2019,12, 2413 4 of 32
the objective of reducing environmental pollution is inseparable from modernising current generation
CIEs to work with NG.
Moreover, it should be noted that most research has concentrated on separate speed and load
engine operation mode experiments. However, when a fixed CCR NG gas component exists, there is a
lack of data regarding the fuel mixture composition in wide and rational engine work mode ranges,
as well as rational dual-fuel distribution in real-life operational range modes. Work that evaluates
diesel reading changes under operational conditions has passive properties as a rule, without the
realisation of experiments [
4
,
26
,
35
]. This is because of operational health and safety regulations for
parameter changes, which limit fast-acting processes.
This paper presents the results from a team of Klaipeda University and Vilnius Gediminas Technical
University scientists. The research object was a high-revolution four-stroke engine with a conventional
fuel injection system, converted for operation with dual-fuel feed diesel and compressed NG.
The research objectives were as follows:
•
Rational estimation of the dual D–NG fuel composition, justifying solutions in accordance with
energy efficiency, emissions, and reliability values, while the engine is operating in a wide range
of modes.
•
The influence of fuel injection timing on the characteristics of the conventional injection fuel
system engine.
•
The evaluation of rational directions ofa converted dual-fuel CIE operation process, with the purpose
of establishing a higher level of energy efficiency and environmentally friendly effectiveness.
2. Experimental Methodology
2.1. General Description of Dual-Fuel Engine
Direct injection high revolution four cylinder turbocharged CIE tests were performed at the
Internal Combustion Engines Laboratory of the Automobile Transport Department, Faculty of Transport
Engineering, Vilnius Gediminas Technical University. A turbocharged 1.9 litre engine with an
electronically controlled BOSCH VP37 distribution-type fuel pump and turbocharger was used for the
tests. The EGR system was disabled during the tests. Diesel injection timing was controlled using
Pulse-Width-Modulation (PWM), by forming electronic control signals for the fuel pump (Figure 1).
The main CIE parameters are listed in Table 1.
Table 1. Engine specification.
Displacement (L) 1.896
Bore ×stroke (mm) 79.5 ×95.5
Power (kW)/speed (rpm) 66/4000
Torque (Nm)/speed (rpm) 180/2000–2500
Cooling type Water cooling
Fuel supply system Direct injection
Cylinders 4 in line
Compression ratio 19.5:1
Aspiration Turbocharge
2.2. Test Bench
The scheme of the laboratory equipment is illustrated in Figure 1. A KI-5543 engine brake stand
was used for the load M
B
and crankshaft speed determination. The torque measurement error was
±
1.23 Nm. The hourly fuel consumption
Bf
was measured by SK-5000 electronic scales and a stopwatch,
and the accuracy of the Bfdetermination was 0.5%.
The NG fuel was measured by a Coriolis-type mass flow meter. The fuel flow meter was a
RHEONIK RHM 015 (see Figure 1pos. 25), connected into the high-pressure fuel supply system before

Energies 2019,12, 2413 5 of 32
the gas reducer, which reduced the gas to a pressure of 1.5 bar. The flow meter measuring range was
0.004 to 0.6 kg/min with a high measurement accuracy of ±0.10%.
Energies 2019, 10, x FOR PEER REVIEW 5 of 33
Figure 1. Scheme of the engine testing equipment: 1: engine; 2: high-pressure fuel pump; 3:
turbocharger; 4: EGR valve; 5: air cooler; 6: connecting shaft; 7: engine load stand; 8: engine torque
and rotational speed recording equipment; 9: fuel injection timing sensor; 10: cylinder pressure
sensor; 11: exhaust gas temperature meter; 12: intake air temperature meter; 13: intake air pressure
meter; 14: air mass meter; 15: exhaust gas analyser; 16: opacity analyser; 17: cylinder pressure
recording equipment; 18: fuel injection timing control equipment (PWM); 19: fuel injection timing
recording equipment; 20: crankshaft position sensor; 21: fuel tank; 22: fuel consumption measuring
equipment; 23: CNG tank; 24: pressure regulation valve; 25: gas flow meter; 26: pressure reducer; 27:
ECU; 28: gas metering valve; 29: gas injectors; 30: air and gas mixer; 31: computer.
2.3. Exhaust Gas Emission Measurement Equipment
The pollutants in the exhaust gas were measured using several gas analysers: AVL DiCom 4000
(AVL, Austria) (Table 2) and HORIBA PG-250 (HORIBA, Japan)(Table 3), TESTO 350 Maritime
(TESTO, Indonesia) (for CO, CO2, HC, and NOx) (Table 4), and AVL DiCom 4000 and MDO-2 LON
(MAHA, Germany) (for absorption coefficient K-value) (Table 5). The HORIBA PG-250, TESTO
350M, and MDO-2LON were used as measurement value control units to ensure the accuracy of the
measurement results. The main equipment for the exhaust gas analysis was the AVL DiGas 4000/AVL
DiCom 4000 (AVL, Austria).
Table 2. Measurement range and resolution of AVL DiCom 4000 gas analyser.
Parameter Measurement Range Measurement Accuracy
Nitrous oxides (NOx) 0–5000 ppm (vol.) 1 ppm
Hydrocarbons (HC) 0–20,000 ppm (vol.) 1 ppm
Carbon monoxide (CO) 0–10% (vol.) 0.01% (vol.)
Carbon dioxide (CO2) 0–20% (vol.) 0.1% (vol.)
Oxygen (O2) 0–25% (vol.) 0.01% (vol.)
Absorption (K-value) 0–99.99 m−1 0.01 m−1
Lub. oil temperature 0–150 °C 1 °C
HORIBA PG-250 exhaust gas analyser measurement range and resolution.
Table 3. Measurement range and resolution of HORIBA PG-250 analyser.
Figure 1.
Scheme of the engine testing equipment: 1: engine; 2: high-pressure fuel pump; 3: turbocharger;
4: EGR valve; 5: air cooler; 6: connecting shaft; 7: engine load stand; 8: engine torque and rotational
speed recording equipment; 9: fuel injection timing sensor; 10: cylinder pressure sensor; 11: exhaust
gas temperature meter; 12: intake air temperature meter; 13: intake air pressure meter; 14: air mass
meter; 15: exhaust gas analyser; 16: opacity analyser; 17: cylinder pressure recording equipment;
18: fuel injection timing control equipment (PWM); 19: fuel injection timing recording equipment;
20: crankshaft position sensor; 21: fuel tank; 22: fuel consumption measuring equipment; 23: CNG
tank; 24: pressure regulation valve; 25: gas flow meter; 26: pressure reducer; 27: ECU; 28: gas metering
valve; 29: gas injectors; 30: air and gas mixer; 31: computer.
2.3. Exhaust Gas Emission Measurement Equipment
The pollutants in the exhaust gas were measured using several gas analysers: AVL DiCom
4000 (AVL, Austria) (Table 2) and HORIBA PG-250 (HORIBA, Japan)(Table 3), TESTO 350 Maritime
(TESTO, Indonesia) (for CO, CO
2
, HC, and NO
x
) (Table 4), and AVL DiCom 4000 and MDO-2 LON
(MAHA, Germany) (for absorption coefficient K-value) (Table 5). The HORIBA PG-250, TESTO
350M, and MDO-2LON were used as measurement value control units to ensure the accuracy of the
measurement results. The main equipment for the exhaust gas analysis was the AVL DiGas 4000/AVL
DiCom 4000 (AVL, Austria).
Table 2. Measurement range and resolution of AVL DiCom 4000 gas analyser.
Parameter Measurement Range Measurement Accuracy
Nitrous oxides (NOx) 0–5000 ppm (vol.) 1 ppm
Hydrocarbons (HC) 0–20,000 ppm (vol.) 1 ppm
Carbon monoxide (CO) 0–10% (vol.) 0.01% (vol.)
Carbon dioxide (CO2) 0–20% (vol.) 0.1% (vol.)
Oxygen (O2) 0–25% (vol.) 0.01% (vol.)
Absorption (K-value) 0–99.99 m−10.01 m−1
Lub. oil temperature 0–150 ◦C 1 ◦C
HORIBA PG-250 exhaust gas analyser measurement range and resolution.
This equipment contains a gas preparation block, which cools down the gases and thus removes
condensate (H
2
O) surplus that may accumulate during combustion. During analysis, the quality of
the outlet gases is established. The complex exhaust gas analyser TESTO 350 Maritime measurement
range and resolution are given in Table 4.

Energies 2019,12, 2413 6 of 32
Table 3. Measurement range and resolution of HORIBA PG-250 analyser.
Emission Component Measurement Range Measurement Accuracy
CO 0–5000 ppm ±1% F.S.
CO20–20 vol.% ±1% F.S.
SO20–3000 ppm ±1 ppm F.S.
NOx0–2500 ppm ±1 ppm F.S.
O20–25% ±1 ppm F.S.
Table 4. Measurement range and resolution of TESTO 350M analyser.
Emission Component Measurement Range Measurement Accuracy
CO 0–5000 ppm ±5% F.S.
CO20–50% ±0.3 vol.% +1% F.S.
NOx0–5000 ppm ±5 ppm (0–99 ppm); ±5% F.S.
(+100–+500 ppm)
SO20–5000 ppm ±5% F.S.
O20–25% ±0.2%
Exhaust gas opacity analyser MDO-2 LON measurement range and accuracy.
Table 5. Measurement range and resolution of MDO-2 LON analyser.
Measurement Parameter Measurement Range Measurement Accuracy
Opacity 0–100% ±2% F.S.
Absorption (K-value) 0–99.99 m−1±2% F.S.
The in-cylinder pressure (
Pcyl
) was recorded by an AVL GH13P piezo-sensor (sensitivity 16 pC/bar,
linearity of FSO
≤ ±
0,3%), which was integrated into the preheating plug and recorded using an
AVL DiTEST DPM 800 amplifier (input range 6000 pC, signal ratio 1 mV/pC, overall error complete
temperature range 1%) and LabView Real-Time equipment. The intake air mass flow meter was
measured by a BOSCH HFM 5 with an accuracy of 2%. The intake manifold pressure was measured
with a Delta OHM HD 2304.0 pressure gauge. A TP704-2BAI sensor device with an error of
±
0.0002
MPa was mounted ahead of the intake manifold. The exhaust and intake gases temperature meter
K-type thermocouple (accuracy ±1.5◦C) was used.
2.4. Fuel Specification
Two fuel types were used during the experiment: liquid and gas (Table 6). During the dual-fuel
mode, standard diesel fuel (EN 590) and standard compressed NG (ISO 6976:1995) were used.
Table 6. Fuel properties.
Fuel type Natural gas Diesel
Density (kg/m3)0.74 829.0
Cetane number - 49
HU(MJ/kg) 51.7 42.8
Viscosity (cSt 40 ◦C) - 1.485
H/C ratio - 1.907
Component (% vol.) Methane: 91.97 Carbon: 86.0
Ethane: 5.75 Hydrogen: 13.6
Propane: 1.30 Oxygen: 0.4
Butane: 0.281
Nitrogen: 0.562
Carbon dioxide: 0.0
2.5. Basis for Numerical Research
The CNIDI (Central Diesel Research Institute, St. Petersburg, Russia) mathematical model from
the TEPLM software package has been used in order to conduct analysis of experimentally obtained

Energies 2019,12, 2413 7 of 32
indicated process diagrams [
36
], which used a closed thermodynamic cycle energy balance model,
evaluating the heat transfer through the cylinder walls. Initial values for the software calculations
included the dual-fuel cycle portion (
qc
) and lower heating value
HL
, while the chemical composition
(C, H, O) was established using the following formula:
qc=qcD·HLD+qcNG ·HLNG
HL.
where,
qc: Overall fuel consumption per cycle, g/cycle;
qcD: Diesel fuel consumption per cycle, g/cycle;
qcNG : NG fuel consumption per cycle, g/cycle;
HLD: Lower heating value of diesel fuel, MJ/kg;
HLNG : Lower heating value of NG, MJ/kg.
HL, the lower heating value of the fuel (MJ/kg) was calculated by the Mendeleev equation [36]:
HL=337.5·C+1025·H−108.3·O,
where,
C=CD·(100 −CCR NG)+CNG ·CCR NG,
H=HD·(100 −CCR NG)+HNG ·CCR NG,
O=OD·(100 −CCR NG)+ONG ·CCR NG,
CCR NG =qcN G ·HLNG
qcNG ·HLNG +qcD·HLD
·100%.
where,
CCR NG: Co-combustion ratio of natural gas, %
CD: Element carbon composition in diesel fuel
CNG: Element carbon composition in NG.
Energy efficiency parameters:
The indicated efficiency (ηi) could be established using the following formula:
ηi=3.6·Pe
HLD·Gf D +HLNG ·Gf NG .
where,
HLD and HLNG
: lover heat values of diesel fuel and NG, respectively, MJ/kg;
BfD and B fNG
: diesel
fuel and NG consumption, respectively, kg/h.
The brake thermal efficiency could be established using the classical expression
ηe=ηi·ηm
,
where
ηm
is the mechanical resistance coefficient established using the combined mechanical engine
losses Pm, and Pmcould be determined from the experimental indicator diagrams.
2.6. Experiment Execution Plan
The experimental engine efficiency and emission research was conducted with a wide range
of loads
(Pme)
and rpm
(n)
, as well as various HRF injection timing angles
ϕinj
(see Table 7).
In every mode, characterised by different combinations
(Pme(D)
,
n
,
ϕinj)
, the engine parameters were
measured using diesel only (D), and dual D and NG fuel: D60/NG40, D40/NG60, and D20/NG80
(here, the numbers following “D” and “NG” correspond to the diesel and NG percentage parts of the
total energy balance). The engine load modes were named in the following manner:
Pme =
5.97
bar
,
high load mode (HLM);
Pme =
3.98
bar
, medium load mode (MLD); and
Pme =
1.98
bar
, low load
mode (LLM).

Energies 2019,12, 2413 8 of 32
Table 7. Experiment execution plan.
n=2500 min−1D/GD n=2000 min−1D/GD n=1500 min−1D/GD
ϕinj D D60/G40 D40/G60 D20/G80 D D60/G40 D40/G60 D20/G80 D D60/G40 D40/G60 D20/G80
Pme =0.597 MPa
−1 X X X X
−7 X X X X
−13 X X X X
Pme =0.398 MPa
−1 X X X X X X X X
−7 X X X X X X X X
−13 X X X X X X X X
Pme =0.198 MPa
−1 X X X X X X X X
−7 X X X X X X X X
−13 X X X X X X X X
Note: The engine parameters were investigated in the following ranges:
ϕinj =
1; 4; 7; 10; 13
◦CA BTDC
; the provided results are in the range:
ϕinj =
1; 7; 13
◦CA BTDC
.
X: investigated parameters.

Energies 2019,12, 2413 9 of 32
For every load mode energy efficiency value, the emission and indicated parameters were
registered at least three times, with further average calculations and rough error removal using
statistical methods (MAthWorks—Matlab, Microsoft Excel). The indicator diagram pressure data
arrays were established as the averages of 100 registered indicator diagrams.
3. Research Results and Discussion
The conversion of a CIE for operation with NG fuel feed encompasses the evaluation of energy
efficiency, emissions, and reliability values (cylinder used, piston group detailed mechanical load ratio
criteria, and maximum cylinder pressure Pmax).
3.1. Changes in Energy Efficiency
The effective
ηe
and indicated
ηi
efficiency were used as the energy efficiency parameters. Changes
in the parameter
ηe
when increasing CCR NG in different load and rpm modes (
Pme
), and the injection
timing ϕinj range results are provided as well (see Figure 2).
Energies 2019, 12, 2413 9 of 33
(a)
(b)
Figure 2. Cont.

Energies 2019,12, 2413 10 of 32
Energies 2019, 10, x FOR PEER REVIEW 10 of 33
(c)
(d)
Energies 2019, 10, x FOR PEER REVIEW 11 of 33
(e)
Figure 2. Influence of CCR NG and 𝜑 (𝜙) on engine 𝜂 at (a) HLM, 𝑛 = 2000 min; (b) MLM,
n = 2000 min; (c) LLM, 𝑛 = 2000 min; (d) MLM, 𝑛 = 2500 min; (e) LLM, 𝑛 = 1500 min.
The experimental data are denoted by dots in the graph.
It should be noted that the obtained dependencies 𝜂=𝑓(𝑃
,𝜑, 𝐶𝐶𝑅 𝑁𝐺) are qualitatively
and quantitatively identical to those of an engine operating in the investigated revolution range of
𝑛 = 1500 min, 𝑛 = 2000 min, 𝑛 = 2500 min (see Figure 2).
Based on this change in the CCR, the engine parameters were specified as n = 2000 min−1.
The load characteristics at 𝑛 = 1500 min and 𝑛 = 2500 min data were limited, only
disclosing their specific features. However, the CCR NG quantitative influence did differ strongly for
different load modes.
Operating in a mode close to the nominal engine load 𝑃 = 5.98 𝑏𝑎𝑟 CCR NG, the increase in
the influence on the 𝜂 parameter was minimal: a 0.8% to 1.8% 𝜂 decrease for every CCR NG
increase of 10%. The range top limit values were obtained with relatively low process dynamics, at
𝜑 = 1 to 4 °CA BTDC; lower values are specific to processes with high work process dynamics at
𝜑 = 10 to 13°CA BTDC (see Figure 3).
Figure 2.
Influence of CCR NG and
ϕinj φin j
on engine
ηe
at (
a
) HLM,
n=
2000
min−1
; (
b
) MLM,
n=
2000
min−1
; (
c
) LLM,
n=
2000
min−1
; (
d
) MLM,
n=
2500
min−1
; (
e
) LLM,
n=
1500
min−1
.
The experimental data are denoted by dots in the graph.

Energies 2019,12, 2413 11 of 32
It should be noted that the obtained dependencies
ηe=f(Pme
,
ϕinj
,
CCR NG)
are qualitatively
and quantitatively identical to those of an engine operating in the investigated revolution range of
n=1500 min−1,n=2000 min−1,n=2500 min−1(see Figure 2).
Based on this change in the CCR, the engine parameters were specified as n=2000 min−1.
The load characteristics at
n=
1500
min−1
and
n=
2500
min−1
data were limited, only disclosing
their specific features. However, the CCR NG quantitative influence did differ strongly for different
load modes.
Operating in a mode close to the nominal engine load
Pme =
5.98
bar
CCR NG, the increase
in the influence on the
ηe
parameter was minimal: a 0.8% to 1.8%
ηe
decrease for every CCR NG
increase of 10%. The range top limit values were obtained with relatively low process dynamics,
at
ϕinj =
1
to
4
◦
CA BTDC; lower values are specific to processes with high work process dynamics at
ϕinj =10 to 13◦CA BTDC (see Figure 3).
Energies 2019, 10, x FOR PEER REVIEW 12 of 33
Figure 3. Influence of CCR NG portion increase on diesel engine 𝜂.
The analysed 𝜑 change was evaluated as one of the most technologically simple measures,
capable of improving the engine parameters for dual-fuel operation. Moreover, an experiment on the
changes in 𝜑 within the range of 1 to 13 °CA BTDC was conducted to expand the obtained
results for engines with different dynamic characteristics (𝑃, pressure increase for maximum
speed (𝑑𝑃 𝑑𝜑
⁄), average speed (𝑑𝑃 𝑑𝜑
⁄), and pressure increase for (λ)).
The MLM 𝑃 =3.98 bar parameter 𝜂 decreased for every CCR NG increase of 10%, making
up 2.5% to 3.5%, and for the LLM 𝑃 =1.99 bar from 4.7% to 6.0%. It is possible to adjust the
parameter 𝜑 to improve 𝜂 for operation with NG in different load modes, and 𝑃 differs
significantly. When the engine is operating in the HLM 𝜑 advancement (∆𝜙) to 3 °CA BTDC,
compensating for the negative effect of an increase in the CCR increase on 𝜂 (Figure 3).
Overall, it is rational that for every CCR NG increase of 20%, there should be an advancement
of 𝜑 (∆𝜙) by 3 °CA BTDC. The obtained results are specific to the investigated 𝜑 range,
which means that it is applicable for engine models with different operational process dynamics: in
the investigated object, the pressure increase 𝑃 𝑃
⁄ was changed (where 𝑃 is the final
compression pressure) within the range from 0.9 to 1.8.
In the MDL mode, the influence of the 𝑃 =3.98 bar parameter 𝜑 on 𝜂 decreased. In
order to compensate for the 𝜂 decrease with the engine operating on NG, it is necessary to advance
𝜑 (∆𝜙) to 3 to 6 °CA for every 20% increase in CCR NG: the lower limit values are consistent
with a low dynamics process. For the LLM 𝑃 = 1.99 bar , the injection timing 𝜑 must be
advanced ∆𝜙 to 9 to 12 °CA or even more.
The adjustment of 𝜑 for improving 𝜂 is inseparable from the necessity to control the
operation process dynamic parameters, including the maximum cycle pressure, in order to avoid
mechanical overloading of the engine. The overall tendency of the investigated engine load 𝑃 is
Figure 3. Influence of CCR NG portion increase on diesel engine ηe.
The analysed
ϕinj
change was evaluated as one of the most technologically simple measures,
capable of improving the engine parameters for dual-fuel operation. Moreover, an experiment on the
changes in
ϕinj
within the range of 1
to
13
◦
CA BTDC was conducted to expand the obtained results for
engines with different dynamic characteristics (
Pmax
, pressure increase for maximum speed
(dP/dϕ)max
,
average speed (dP/dϕ)mid, and pressure increase for (λ)).
The MLM
Pme =
3.98
bar
parameter
ηe
decreased for every CCR NG increase of 10%, making
up 2.5%
to
3.5%, and for the LLM
Pme =
1.99
bar
from 4.7%
to
6.0%. It is possible to adjust the
parameter
ϕinj
to improve
ηe
for operation with NG in different load modes, and
Pme
differs significantly.
When the engine is operating in the HLM
ϕinj
advancement
∆φinj
to 3
◦
CA BTDC, compensating for
the negative effect of an increase in the CCR increase on ηe(Figure 3).

Energies 2019,12, 2413 12 of 32
Overall, it is rational that for every CCR NG increase of 20%, there should be an advancement of
ϕinj ∆φinj
by 3
◦
CA BTDC. The obtained results are specific to the investigated
ϕinj
range, which means
that it is applicable for engine models with different operational process dynamics: in the investigated
object, the pressure increase
Pmax/Pc
was changed (where
Pc
is the final compression pressure) within
the range from 0.9 to 1.8.
In the MDL mode, the influence of the
Pme =
3.98
bar
parameter
ϕinj
on
ηe
decreased. In order
to compensate for the
ηe
decrease with the engine operating on NG, it is necessary to advance
ϕinj
∆φinj
to 3 to 6
◦
CA for every 20% increase in CCR NG: the lower limit values are consistent with a
low dynamics process. For the LLM
Pme =
1.99
bar
, the injection timing
ϕinj
must be advanced
∆φinj
to 9 to 12 ◦CA or even more.
The adjustment of
ϕinj
for improving
ηe
is inseparable from the necessity to control the operation
process dynamic parameters, including the maximum cycle pressure, in order to avoid mechanical
overloading of the engine. The overall tendency of the investigated engine load
Pmax
is that a decrease
in
Pmax
occurs with an increase in CCR NG (see Figure 4). It is known that the size of
Pmax
is determined
by the heat release
QPmax
at the maximum pressure phase
ϕPmax
[
37
–
40
]. The size is mainly formed
by the heat released during the first kinetic phase, which is influenced by the auto-ignition delay
period
ϕi
. A decrease in the amount of fuel injected through
ϕi
or in the period of injection
ϕi
ensures
a decrease in
Pmax
[
41
–
43
]. The
Pmax
dependence on the CCR NG portion variable is nonlinear: at a
CCR NG increase of up to 0.4,
Pmax
changed only slightly; however, when the CCR NG increased to a
larger portion than 0.4 and up to 0.8, the decrease in Pmax reached 10 to 15 bar (see Figure 4).
Energies 2019, 12, 2413 13 of 33
that a decrease in 𝑃 occurs with an increase in CCR NG (see Figure 4). It is known that the size of
𝑃 is determined by the heat release 𝑄 at the maximum pressure phase 𝜑 [37–40]. The
size is mainly formed by the heat released during the first kinetic phase, which is influenced by the
auto-ignition delay period 𝜑. A decrease in the amount of fuel injected through 𝜑 or in the period
of injection 𝜑 ensures a decrease in 𝑃 [41–43]. The 𝑃 dependence on the CCR NG portion
variable is nonlinear: at a CCR NG increase of up to 0.4, 𝑃 changed only slightly; however, when
the CCR NG increased to a larger portion than 0.4 and up to 0.8, the decrease in 𝑃 reached 10 to 15
bar (see Figure 4).
(a)
(b)
Figure 4. Cont.

Energies 2019,12, 2413 13 of 32
Energies 2019, 12, 2413 14 of 33
(c)
Figure 4. Injection timing 𝜑

influence on maximum cycle pressure 𝑃

(𝑛 = 2000 min

): (a)
HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
Therefore, it is appropriate to increase the CCR NG and compensate for the 𝜂 losses with an
increase in 𝜑. Thus, in order to evaluate the changes in 𝑃 for CIE conversion for operation
wit h NG fuel fe ed it is nec essary to u se differe nt por tions of NG , such a s CCR NG 0 to 0.2 and 0 to 0.4.
Figure 5 provides the engine indicator diagrams with the speed of heat release characteristics 𝑑𝑄/𝑑𝜑.
(a) (b)
(c) d)
Figure 5. Cont.
Figure 4.
Injection timing
ϕinj
influence on maximum cycle pressure
Pmaxn=2000 min−1
: (
a
) HLM;
(b) MLM; (c) LLM. The dots in the graph represent the experimental data.
Therefore, it is appropriate to increase the CCR NG and compensate for the
ηe
losses with an
increase in
ϕinj
. Thus, in order to evaluate the changes in
Pmax
for CIE conversion for operation with
NG fuel feed it is necessary to use different portions of NG, such as CCR NG 0
to
0.2 and 0
to
0.4.
Figure 5provides the engine indicator diagrams with the speed of heat release characteristics
dQ/dϕ
.
Energies 2019, 12, 2413 14 of 33
(c)
Figure 4. Injection timing 𝜑

influence on maximum cycle pressure 𝑃

󰇛𝑛  2000 min

󰇜: (a)
HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
Therefore, it is appropriate to increase the CCR NG and compensate for the 𝜂 losses with an
increase in 𝜑. Thus, in order to evaluate the changes in 𝑃
 for CIE conversion for operation
with NG fuel feed it is necessary to use different portions of NG, such as CCR NG 0 to 0.2 and 0 to 0.4.
Figure 5 provides the engine indicator diagrams with the speed of heat release characteristics 𝑑𝑄/𝑑𝜑.
(a) (b)
(c) (d)
Figure 5. Cont.

Energies 2019,12, 2413 14 of 32
Energies 2019, 10, x FOR PEER REVIEW 15 of 33
(e) (f)
Figure 5. Combustion process indicator diagrams and heat release dynamics 𝑑𝑄 𝑑𝜑
⁄ for
different 𝜑

𝜑

= −1 and − 13 °CA, (𝑛 = 2000 min

): (a) and (b) correspond to LLM; (c) and
(d) correspond to MLM; (e) and (f) correspond to HLM.
For every engine load mode, the diesel auto-ignition phase for 𝜑 −𝑖𝑑𝑒𝑚 did not change in
the CCR NG portion range from 0 to 0.8. The ignition delay period 𝜑 also did not change for 𝜑=
81°CA in all investigated 𝜑. Overall, the decrease in 𝑃 can be linked to the changes in the
heat release dynamics when the CCR NG is increased. A decrease in the diesel portion decreases the
heat release in the heterogeneous combustion kinetics phase, and increases the total energy balance
part of the NG volume specific combustion part [18,23,44]. The combustion process moves towards
an expansion stroke, and therefore the decrease in 𝑃 occurs in parallel with the deterioration in
𝜂.
The advancement of the injection timing 𝜑 moves the combustion process towards TDC,
compensating for the changes in 𝜂 relating to the CCR NG increase. However, the
𝜑 advancement in engines with conventional fuel injection systems exhibits limited capabilities in
the aforementioned range. One method for improving the 𝜂 characteristics is accelerating the
combustion process in the second (main) phase, which forms the energy efficiency indicators of the
cycle. In order to achieve this, the cylinder air load is increased; moreover, the fuel portion vortex
movement degree is increased by using different technological methods, such as incorporating fuel
additives [42,43].
In the HLM, in which the CCR NG constitutes 0.2, to compensate for the deterioration in 𝜂,
𝜑 is advanced to 3°CA BTDC, and the maximum cycle pressure is increased to 7 to 10 bar (see
Figure 4). In the MLM, the 𝜑 advancement to 6°CA BTDC leads to an increase in 𝑃 of up to
15 to 25 bar: the result is that 𝑃 is at a larger pressure than that obtained by using diesel only
(CCR NG = 0).
The LLM 𝑃 can reach 20 to 25 bar. Although 𝑃 is lower during the medium and low load
modes compared to the modes that are closer to the nominal engine load, and engine operation in
the HLM only occurs for a limited time, a meaningful increase can significantly decrease the engine
reliability. Therefore, it is not considered as rational for practical use. A CCR NG increase for the LLM
up to 0.4, together with a 𝜑 advancement to 6 °CA, can lead to an increase in 𝑃 to 15 to 25 bar.
During the MLM, the 𝑃 increase can reach up to 30 bar and more. It should be noted that
changes in 𝑃 are established when the engine is operating with diesel fuel feed in a stationary
state of 𝜑 =1 to 4 °CA BTDC, or low process dynamics. When the system is in a higher state of
𝜑 = 7 to 13 °CA BTDC or in a more dynamic DE operation state, the advancement of 𝜑 becomes
irrational: for a small 𝜂 restoration effect, a large increase in 𝑃 is necessary.
Overall, it should be stated that application of the 𝜑 advancement to compensate for the
decrease in energy efficiency when the engine is operating in the dual-fuel mode is limited during
the HLM and portions of the CCR NG from 0 to 0.4.
The exchange of diesel for NG and the use of NG in different load modes are essentially
expanded by applying the 𝜑 advancement method, when there is also a decrease of approximately
Figure 5.
Combustion process indicator diagrams and heat release dynamics
dQ/dϕ
for different
ϕinj ϕin j =−1 and −13 ◦CA
,
n=2000 min−1
: (
a
) and (
b
) correspond to LLM; (
c
) and (
d
) correspond
to MLM; (e) and (f) correspond to HLM.
For every engine load mode, the diesel auto-ignition phase for
ϕinj −idem
did not change in the
CCR NG portion range from 0 to 0.8. The ignition delay period
ϕi
also did not change for
ϕi=
8
±
1
◦CA
in all investigated
ϕinj
. Overall, the decrease in
Pmax
can be linked to the changes in the heat release
dynamics when the CCR NG is increased. A decrease in the diesel portion decreases the heat release in
the heterogeneous combustion kinetics phase, and increases the total energy balance part of the NG
volume specific combustion part [
18
,
23
,
44
]. The combustion process moves towards an expansion
stroke, and therefore the decrease in Pmax occurs in parallel with the deterioration in ηe.
The advancement of the injection timing
ϕinj
moves the combustion process towards TDC,
compensating for the changes in
ηe
relating to the CCR NG increase. However, the
ϕinj
advancement
in engines with conventional fuel injection systems exhibits limited capabilities in the aforementioned
range. One method for improving the
ηe
characteristics is accelerating the combustion process in the
second (main) phase, which forms the energy efficiency indicators of the cycle. In order to achieve this,
the cylinder air load is increased; moreover, the fuel portion vortex movement degree is increased by
using different technological methods, such as incorporating fuel additives [42,43].
In the HLM, in which the CCR NG constitutes 0.2, to compensate for the deterioration in
ηe
,
ϕinj
is
advanced to 3
◦
CA BTDC, and the maximum cycle pressure is increased to 7
to
10 bar (see Figure 4).
In the MLM, the
ϕinj
advancement to 6
◦
CA BTDC leads to an increase in
Pmax of
up to 15
to
25 bar: the
result is that Pmax is at a larger pressure than that obtained by using diesel only (CCR NG =0).
The LLM
Pmax
can reach 20
to
25 bar. Although
Pmax
is lower during the medium and low load
modes compared to the modes that are closer to the nominal engine load, and engine operation in
the HLM only occurs for a limited time, a meaningful increase can significantly decrease the engine
reliability. Therefore, it is not considered as rational for practical use. A CCR NG increase for the LLM
up to 0.4, together with a ϕinj advancement to 6 ◦CA, can lead to an increase in Pmax to 15 to 25 bar.
During the MLM, the
Pmax
increase can reach up to 30 bar and more. It should be noted that
changes in
Pmax
are established when the engine is operating with diesel fuel feed in a stationary
state of
ϕinj =
1
to
4
◦CA BTDC
, or low process dynamics. When the system is in a higher state of
ϕinj =
7
to
13
◦CA BTDC
or in a more dynamic DE operation state, the advancement of
ϕinj
becomes
irrational: for a small ηerestoration effect, a large increase in Pmax is necessary.
Overall, it should be stated that application of the
ϕinj
advancement to compensate for the decrease
in energy efficiency when the engine is operating in the dual-fuel mode is limited during the HLM and
portions of the CCR NG from 0 to 0.4.
The exchange of diesel for NG and the use of NG in different load modes are essentially expanded
by applying the
ϕinj
advancement method, when there is also a decrease of approximately 3%. The LLM
CCR NG is expanded to 40%, while the MLM possible CCR NG application reaches 20% when decreases

Energies 2019,12, 2413 15 of 32
of 34% and 6% occur, respectively. An increase occurs when operating with diesel fuel equal to 7 bar,
which remains unchanged.
3.2. Emission
The engine emission evaluation is directly related to the engine purpose. Vehicle engines
are regulated by international standards throughout the entire range of harmful emissions:
NOx
,
CO
,
CnHm
,
and SOx
, as well as PM [
45
–
48
]. Moreover, marine-purpose DEs are regulated
by MARPOL 73/78 convention Annex VI [
49
] for
NOx
and
SOx
emissions. However, as opposed to
DEs for other purposes, marine power plants are also regulated by the decision of the International
Maritime Organization to limit greenhouse gas
CO2
emissions [
50
]. Therefore, it is correct for the
evaluation of converted dual-fuel engine emissions to be conducted in accordance with their purpose.
3.2.1. Nitrous oxides
The majority of research has pointed out that CIE conversion for operation with NG fuel feed
fundamentally decreases
NOx
emissions, owing to the equalisation of the combustion temperature
field in the cylinder and decrease in high-temperature zones [
45
–
47
,
50
]. The results obtained from this
research agree with this tendency. The effects of the
NOxe
mission decrease with an increase in the
CCR are nonlinear – the maximum effect in the investigated options was detected in the CCR NG >
40% range. The NOxmainly decreased in the LLM Pme =1.99 bar.
Moreover, depending on the stationary
ϕinj
value while operating on diesel only, for every CCR
NG increase of 10%, the
NOx
emissions decreased by 7% to 3% in the HLM, and 9% to 10% in the
MLM and LLM (earlier values of
ϕinj
exhibited lower changes in
NOx
emissions). With this manner
of exchanging diesel for NG, the
NOx
decrease constitutes approximately 65% in the HLM, and a
decrease in NOxof 90% to 95% can be reached in the MLM and LLM (see Figure 6).
However, the CCR NG increase is followed by a decrease in
ηe
of up to 20% in the MLM and 45%
in the LLM. It is obvious that the changes in
ηe
and
NOx
must be coordinated, for example, by applying
the ϕinj advancement method (see Figure 6).
Energies 2019, 10, x FOR PEER REVIEW 16 of 33
3%. The LLM CCR NG is expanded to 40%, while the MLM possible CCR NG application reaches
20% when decreases of 34% and 6% occur, respectively. An increase occurs when operating with
diesel fuel equal to 7 bar, which remains unchanged.
3.2. Emission
The engine emission evaluation is directly related to the engine purpose. Vehicle engines are
regulated by international standards throughout the entire range of harmful emissions:
NO,CO,CH,and SO, as well as PM [45–47]. Moreover, marine-purpose DEs are regulated by
MARPOL 73/78 convention Annex VI [49] for NO and SO emissions. However, as opposed to DEs
for other purposes, marine power plants are also regulated by the decision of the International
Maritime Organization to limit greenhouse gas CO emissions [50]. Therefore, it is correct for the
evaluation of converted dual-fuel engine emissions to be conducted in accordance with their purpose.
3.2.1. Nitrous oxides
The majority of research has pointed out that CIE conversion for operation with NG fuel feed
fundamentally decreases NO emissions, owing to the equalisation of the combustion temperature
field in the cylinder and decrease in high-temperature zones [45–47, 50]. The results obtained from
this research agree with this tendency. The effects of the NO 𝑒mission decrease with an increase in
the CCR are nonlinear – the maximum effect in the investigated options was detected in the CCR NG
> 40% range. The 𝑁𝑂 mainly decreased in the LLM 𝑃 =1.99 bar.
Moreover, depending on the stationary 𝜑 value while operating on diesel only, for every CCR
NG increase of 10%, the NO emissions decreased by 7% to 3% in the HLM, and 9% to 10% in the
MLM and LLM (earlier values of 𝜑 exhibited lower changes in NO emissions). With this manner
of exchanging diesel for NG, the NO decrease constitutes approximately 65% in the HLM, and a
decrease in NO of 90% to 95% can be reached in the MLM and LLM (see Figure 6).
However, the CCR NG increase is followed by a decrease in 𝜂 of up to 20% in the MLM and
45% in the LLM. It is obvious that the changes in 𝜂 and NO must be coordinated, for example, by
applying the 𝜑 advancement method (see Figure 6).
(a)
Figure 6. Cont.

Energies 2019,12, 2413 16 of 32
Energies 2019, 10, x FOR PEER REVIEW 17 of 33
(b)
(c)
Figure 6. Influence of dual-fuel CCR NG portion and 𝜑 on NO emissions (𝑛 = 2000 min): (a)
HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
A CCR NG increase to 0.2 in the HLM was implemented practically without a significant
decrease in the energy efficiency; therefore, even without the use of 𝜑 advancement, the
NO emissions decreased. When increasing the CCR NG portion up to 0.4, it is rational to use a
limited 𝜑 advancement solution, which will ensure a certain amount of 𝜂 and NO reduction.
For example, in the HLM, a𝜑 advancement to 3 °CA BTDC instead of 6°CA BTDC ensured a
reduction in 𝜂 and NO of 3% and 0.9 g/kWh, respectively. Without changing 𝜑 , 𝜂 was
reduced by 6.5%, and the NO emissions were reduced by 1.2 to 1.7 g/kWh or 35% compared to the
engine operating on diesel only.
In the MLM, a partial deterioration of the 𝜂 compensation with a change in 𝜑 ensured a
NO decrease of 0.7 to 1.1 g/kWh or approximately 15%. In the LLM, exchanging 0.2 of diesel with
NG, without 𝜑 adjustment, a reduction in NO emissions of 2 to 3.3 g/kWh or 50% was achieved.
3.2.2. Carbon monoxide
Figure 6.
Influence of dual-fuel CCR NG portion and
ϕinj
on
NOx
emissions (
n=
2000
min−1)
:
(a) HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
ACCR NG increase to 0.2 in the HLM was implemented practically without a significant decrease
in the energy efficiency; therefore, even without the use of
ϕinj
advancement, the
NOx
emissions
decreased. When increasing the CCR NG portion up to 0.4, it is rational to use a limited
ϕinj
advancement
solution, which will ensure a certain amount of
ηe
and
NOx
reduction. For example, in the HLM,
a
ϕinj
advancement to 3
◦
CA BTDC instead of 6
◦
CA BTDC ensured a reduction in
ηe
and
NOx
of 3%
and 0.9 g/kWh, respectively. Without changing
ϕinj
,
ηe
was reduced by 6.5%, and the
NOx
emissions
were reduced by 1.2 to 1.7 g/kWh or 35% compared to the engine operating on diesel only.
In the MLM, a partial deterioration of the
ηe
compensation with a change in
ϕinj
ensured a
NOx
decrease of 0.7
to
1.1 g/kWh or approximately 15%. In the LLM, exchanging 0.2 of diesel with NG,
without ϕinj adjustment, a reduction in NOxemissions of 2 to 3.3 g/kWh or 50% was achieved.

Energies 2019,12, 2413 17 of 32
3.2.2. Carbon monoxide
According to various sources, the conversion of a CIE for dual-fuel operation is related to a
significant increase in harmful partial combustion components, such as
CO
and HC emissions [
45
–
48
].
During the experiment, it was determined that when a portion of CCR NG of 0.8 was reached, the CO
emissions increased 8 to 30 times in the HLM from the base level of 0.5 g/kWh, 20 to 30 times in the
MLM from approximately 1 g/kWh, and 10 to 20 times in the LLM from approximately 2 to 6 g/kWh,
during different operation dynamics (see Figure 7).
Energies 2019, 10, x FOR PEER REVIEW 18 of 33
According to various sources, the conversion of a CIE for dual-fuel operation is related to a
significant increase in harmful partial combustion components, such as CO and HC emissions [45–
48]. During the experiment, it was determined that when a portion of CCR NG of 0.8 was reached,
the CO emissions increased 8 to 30 times in the HLM from the base level of 0.5 g/kWh, 20 to 30 times
in the MLM from approximately 1 g/kWh, and 10 to 20 times in the LLM from approximately 2 to 6
g/kWh, during different operation dynamics (see Figure 7).
(a)
(b)
Figure 7. Cont.

Energies 2019,12, 2413 18 of 32
Energies 2019, 10, x FOR PEER REVIEW 19 of 33
(c)
Figure 7. Influence of dual-fuel CCR NG portion and 𝜑 on CO emissions (𝑛 = 2000 min): (a)
HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
In the investigated range, 𝜑 advancement becomes an ineffective measure for CO emission
control. Overall, for CCR NG 0.2 and 0.4 portions, an increase in CO emissions could be detected in
the HLM of 3 to 5 g/kWh and 5.5 to 10 g/kWh, and in the MLM, with CCR NG 0.2, an increase
occurred for 7 to 10 g/kWh.
3.2.3. Hydrocarbons
No less intensive is the HC increase when the engine is operating on NG (see Figure 8).
(a)
Figure 7.
Influence of dual-fuel CCR NG portion and
ϕinj
on
CO
emissions (
n=
2000
min−1)
: (
a
) HLM;
(b) MLM; (c) LLM. The dots in the graph represent the experimental data.
In the investigated range,
ϕinj
advancement becomes an ineffective measure for
CO
emission
control. Overall, for CCR NG 0.2 and 0.4 portions, an increase in
CO
emissions could be detected in the
HLM of 3 to 5 g/kWh and 5.5 to 10 g/kWh, and in the MLM, with CCR NG 0.2, an increase occurred for
7 to 10 g/kWh.
3.2.3. Hydrocarbons
No less intensive is the HC increase when the engine is operating on NG (see Figure 8).
Energies 2019, 10, x FOR PEER REVIEW 19 of 33
(c)
Figure 7. Influence of dual-fuel CCR NG portion and 𝜑 on CO emissions (𝑛 = 2000 min): (a)
HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
In the investigated range, 𝜑 advancement becomes an ineffective measure for CO emission
control. Overall, for CCR NG 0.2 and 0.4 portions, an increase in CO emissions could be detected in
the HLM of 3 to 5 g/kWh and 5.5 to 10 g/kWh, and in the MLM, with CCR NG 0.2, an increase
occurred for 7 to 10 g/kWh.
3.2.3. Hydrocarbons
No less intensive is the HC increase when the engine is operating on NG (see Figure 8).
(a)
Figure 8. Cont.

Energies 2019,12, 2413 19 of 32
Energies 2019, 10, x FOR PEER REVIEW 20 of 33
(b)
(c)
Figure 8. Influence of dual-fuel CCR NG portion and 𝜑 on HC emissions (𝑛 = 2000 min): (a)
HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
The HC emissions increased when the engine load was increased, and the influence of the 𝜑
advancement was not significant. In the HLM, for every CCR NG portion increase of 0.1, the HC
emission increase constituted approximately 0.1 g/kWh; when the CCR NG was in the range of 0.6
to 0.8, the HC emissions increased by 0.15 to 0.4 g/kWh. In the MLM, for an analogous CCR NG
portion, the HC emissions increase constituted 0.4 to 0.6 g/kWh and 0.7 to 1.5 g/kWh, respectively;
for the LLM, the values were 1 to 1.5 g/kWh and 2.5 to 3.0 g/kWh.
3.2.4. Carbon Dioxide
Greenhouse gas CO emissions from internal combustion engines are dependent on two factors:
the carbon C in the fuel chemical composition and fuel consumption [45,46,51]. Compared to diesel,
the NG chemical composition has a C part of 75% versus 85% to 86%. Thus, engine conversion for
operating on NG at the same mass portion results in a decrease in fuel consumption and therefore
reduced CO emissions.
Figure 8.
Influence of dual-fuel CCR NG portion and
ϕinj
on HC emissions (
n=
2000
min−1)
: (
a
) HLM;
(b) MLM; (c) LLM. The dots in the graph represent the experimental data.
The HC emissions increased when the engine load was increased, and the influence of the
ϕinj
advancement was not significant. In the HLM, for every CCR NG portion increase of 0.1, the HC
emission increase constituted approximately 0.1 g/kWh; when the CCR NG was in the range of 0.6
to 0.8, the HC emissions increased by 0.15 to 0.4 g/kWh. In the MLM, for an analogous CCR NG portion,
the HC emissions increase constituted 0.4 to 0.6 g/kWh and 0.7 to 1.5 g/kWh, respectively; for the LLM,
the values were 1 to 1.5 g/kWh and 2.5 to 3.0 g/kWh.
3.2.4. Carbon Dioxide
Greenhouse gas
CO2
emissions from internal combustion engines are dependent on two factors:
the carbon
C
in the fuel chemical composition and fuel consumption [
45
,
46
,
51
]. Compared to diesel,
the NG chemical composition has a
C
part of 75% versus 85% to 86%. Thus, engine conversion for

Energies 2019,12, 2413 20 of 32
operating on NG at the same mass portion results in a decrease in fuel consumption and therefore
reduced CO2emissions.
During the experiment, a decrease in
CO2
was only achieved for the HLM
Pme =
5.98
bar
in the
entire CCR NG portion range from 0 to 0.8 (see Figure 9).
Moreover, the
ϕinj
advancement reduced the
CO2
emissions in the CCR NG portion range from
0 to 0.8: when
ϕinj =−
1
◦
CA BTDC, the
CO2
emissions decreased by 9%, for
ϕinj
=
−
13
◦
CA BTDC,
the decrease could reach 16%. For the MLM and LLM, the increased
CO2
emissions led to a deterioration
in energy efficiency. Overall, in the mode of
Pme =
3.98
bar
for a range up to a CCR NG portion of 0.6,
the increase in
CO2
emissions practically remained unchanged, and with a further increase in the CCR
NG up to 0.6 to 0.8 in the LLM, the CO2increase totalled 30% to 45%.
The overall effects of the CIE conversion for dual-fuel operation on the energy efficiency and
emissions are provided in Table 8.
Energies 2019, 10, x FOR PEER REVIEW 21 of 33
During the experiment, a decrease in CO was only achieved for the HLM 𝑃 =5.98 𝑏𝑎𝑟 in
the entire CCR NG portion range from 0 to 0.8 (see Figure 9).
(a)
(b)
Figure 9. Cont.

Energies 2019,12, 2413 21 of 32
Energies 2019, 10, x FOR PEER REVIEW 22 of 33
(c)
Figure 9. Influence of dual-fuel CCR NG portion and 𝜑 on CO emissions (𝑛 = 2000 min): (a)
HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
Moreover, the 𝜑 advancement reduced the CO emissions in the CCR NG portion range
from 0 to 0.8: when 𝜑 = −1 °CA BTDC, the CO emissions decreased by 9%, for 𝜑 = −13°CA
BTDC, the decrease could reach 16%. For the MLM and LLM, the increased CO emissions led to a
deterioration in energy efficiency. Overall, in the mode of 𝑃 =3.98 bar for a range up to a CCR
NG portion of 0.6, the increase in CO emissions practically remained unchanged, and with a further
increase in the CCR NG up to 0.6 to 0.8 in the LLM, the CO increase totalled 30% to 45%.
The overall effects of the CIE conversion for dual-fuel operation on the energy efficiency and
emissions are provided in Table 8.
The evaluation encompassed different 𝜂 compensation possibilities, from 𝜑 advancement
from full 𝜂 establishment when the engine was operating on diesel to 𝜑 =𝑖𝑑𝑒𝑚.
Table 8. Influence of investigated engine (79.5/95.5) conversion to operation with NG fuel feed on
energy efficiency and emission parameters.
Operating mode: HLM 𝑷𝒎𝒆 = 𝟓. 𝟗𝟖 𝐛𝐚𝐫
Changes in engine parameters when CCR NG increases from 0 to 0.2
Stationary
𝜑,°CA
Advanced to
𝜑,°CA ∆𝜼𝒆 ∆𝑃,
bar
∆NO,
g/kWh
∆CO,
g/kWh
∆HC,
g/kWh
∆CO,
% ∆CO,
g/kWh
−1 −4 0 +10 −0.2 +4 +0.22 −9 −47
−1 −1 −3% +5 −1.2 +5 +0.25 −5.5 −25
−4 −7 ~0.7% +7 +0.2 +3 +0.21 −9 −34
−4 −4 −3.5% 0 −1.7 +4 +0.22 −6.5 −16
Operating mode: HLM 𝑷𝒎𝒆 = 𝟓. 𝟗𝟖 𝐛𝐚𝐫
Changes in engine parameters when CCR NG increases from 0 to 0.4
Stationary 𝜑,
°CA
Advanced to 𝜑,
°CA ∆𝜂 ∆𝑃,
bar
∆NO,
g/kWh
∆CO,
g/kWh
∆HC,
g/kWh
∆CO,
% ∆CO,
g/kWh
−1 −4 −2.9% +9 −0.9 +6 +0.45 −15 −63
−1 −1 −6.5% +4 −1.7 +10 +0.5 −9 −49
−4 −10 −1.3% +23 +2.6 +5,5 +0.42 −16 −72
−4 −7 −3.5% +7 −0.6 +6 +0.45 −15 −60
−4 −4 −6.4% 0 −1.2 +8 +0.45 −12 −32
Figure 9.
Influence of dual-fuel CCR NG portion and
ϕinj
on
CO2
emissions (
n=
2000
min−1)
:
(a) HLM; (b) MLM; (c) LLM. The dots in the graph represent the experimental data.
Table 8.
Influence of investigated engine (79.5/95.5) conversion to operation with NG fuel feed on
energy efficiency and emission parameters.
Operating mode: HLM Pme =5.98 bar
Changes in engine parameters when CCR NG increases from 0 to 0.2
Stationary
ϕinj
,
◦CA
Advanced to
ϕinj ,◦CA ∆ηe
∆Pz,
bar
∆NOx,
g/kWh
∆CO,
g/kWh
∆HC,
g/kWh
∆CO2,
%
∆CO2,
g/kWh
−1−4 0 +10 −0.2 +4+0.22 −9−47
−1−1−3% +5−1.2 +5+0.25 −5.5 −25
−4−7 ~0.7% +7+0.2 +3+0.21 −9−34
−4−4−3.5% 0 −1.7 +4+0.22 −6.5 −16
Operating mode: HLM Pme =5.98 bar
Changes in engine parameters when CCR NG increases from 0 to 0.4
Stationary
ϕinj ,◦CA
Advanced to
ϕinj ,◦CA ∆ηe
∆Pz,
bar
∆NOx,
g/kWh
∆CO,
g/kWh
∆HC,
g/kWh
∆CO2,
%
∆CO2,
g/kWh
−1−4−2.9% +9−0.9 +6+0.45 −15 −63
−1−1−6.5% +4−1.7 +10 +0.5 −9−49
−4−10 −1.3% +23 +2.6 +5,5 +0.42 −16 −72
−4−7−3.5% +7−0.6 +6+0.45 −15 −60
−4−4−6.4% 0 −1.2 +8+0.45 −12 −32
Operating mode: MLM Pme =3.98 bar
Changes in engine parameters when CCR NG increases from 0 to 0.4
Stationary
ϕinj ,◦CA
Advanced
to ϕinj ,◦CA ∆ηe
∆Pz,
bar
∆NOx,
g/kWh
∆CO,
g/kWh
∆HC,
g/kWh
∆CO2,
%
∆CO2,
g/kWh
−1−7+1.5% +24 +1.1 +9.0 +0.36 −6.2 −50
−1−4∼0% +9−0.7 +9.7 +0.43 −3.9 −32
−1−1−6.5% +1−1.7 +10 +0.50 2.7 +24
−4−10 0 +13 +2.1 +7.0 +0.30 −7.3 −56
−4−7−1.5% +10 −1.1 +9.0 +0.35 −3.8 −30
−4−4 -4% 0 −3.0 +9.9 +0.43 −1.5 −12

Energies 2019,12, 2413 22 of 32
The evaluation encompassed different
ηe
compensation possibilities, from
ϕinj
advancement from
full ηeestablishment when the engine was operating on diesel to ϕinj =idem.
When diesel is exchanged for NG (CCR NG 0.2), it is more rational to use partial
ϕinj
advancement
(with the purpose of re-establishing a proper
ηe
). In the
Pme =
1.99
bar
mode, where a decrease in
ηe
of
3% to 3.5% was observed, the increase in
Pmax
did not exceed 5 bar. A significant decrease in
NOx
was
achieved at 1.2 g/kWh, the HC increase did not exceed 0.2 g/kWh, and the
CO2
emissions decreased.
For the MLM with Pme =3.98 bar, the ηedeterioration also did not exceed 3% to 3.5%, Pmax increased
by 7 to 9 bar, and CO2decreased by 15%.
ACCR NG portion increase to 0.4 it is more appropriate for use in partial
ϕinj
advancement.
The engine parameter changes are not significantly different from those of the CCR NG 0.2 values.
An effect of a decrease of approximately 15% in
CO2
emissions detected compared to operation on
diesel only.
However, the HC emissions increase is not critical, because a valid standard regulates the total
NOx+
HC emissions [
52
]. The sum of the
NOx
and HC emissions in the investigated engine conversion
to NG was decreased, because the decrease in the absolute
NOx
exceeded the increase in HC emissions
(see Table 8). As a result, the only increase was in the CO emissions. Very few road transport engine
CO emissions are regulated by secondary harmful emission control technologies (oxidation-type
neutralisations) [
45
–
47
,
51
,
52
]. Moreover, marine-purpose power plant CO emissions are not regulated.
However, DE conversion for dual-fuel operation purposes exhibits complex improvements in energy
efficiency and emission values for conducting research and identifying rational solutions.
3.3. Engine Parameter Improvement Field Evaluation
With reference to previously conducted research on HRF, an injection timing advancement in the
range of
−
1 to
−
13
◦
CA BTDC to improve converted CIE operation into dual-fuel is not very effective,
because it is strongly related to an intensive increase in Pmax and NOx.
The search for rational engine parameters has mainly focused on improving the energy
efficiency—particularly the operation indicated process efficiency
ηi
. Based on internal combustion
engine classical theory rules [
39
,
40
,
52
,
53
], changes in
ηi
during experiments and improvement reserves
are analysed in coordination with heat release dynamics in the engine cylinder, as well as the main
operation process parameters (air excess ratio
α
, dynamic parameters such as
Pmax
, and auto-ignition
delay period ϕi, among others).
As established in Section 3.2, the increase in the CCR NG under
ϕinj =idem
conditions influences
the combustion process movement towards expansion stroke, and increases the combustion process
period—one of the main
ηi
influencing factors (see Figure 10). The graph in Figure 10 contains integral
heat release characteristic data in the relative form of
X=f(ϕ)
, and indicates a partially meaningful
decrease about
X
during the expansion stroke, which is even more evident when there a larger portion
of CCR NG and lower engine load exist.
Larger changes in
X=f(ϕ)
are characteristic for low dynamic engine cycles (
ϕinj =
1 to 4 ◦CA BTDC). For an earlier ϕin j, the differences among the X=f(ϕ)characteristics decrease.
For change evaluation of the quantitative heat release characteristics, the relationship with the
ηi
uses 50% of the released heat phase
CA50
in practice is commonly applied to engine cycle process energy
efficiency evaluation [
30
,
33
]. A direct relationship exists between the CCR NG and
CA50
parameter
interaction: in the HLM, when
CA50
increased from 0 to 0.8, an increase of 4
◦CA
occurred; in the MLM,
an increase of 12
◦CA
occurred; and in the LLM, an increase of 42
◦CA
occurred. When increasing the
engine cycle dynamics (
ϕinj =
13
◦CA BTDC)
, the
CA50
exchange of CCR NG in equal parts decreased
intensely; for example, up to 6 ◦CA and 14 ◦CA for the MLM and LLM, respectively.

Energies 2019,12, 2413 23 of 32
Energies 2019, 10, x FOR PEER REVIEW 24 of 33
(a) (b)
(c)
Figure 10. Dual-fuel engine CCR NG fuel portion influence on heat release characteristics: (a) HLM;
(b) MLM; (c) LLM.
Larger changes in 𝑋=𝑓(𝜑) are characteristic for low dynamic engine cycles ( 𝜑 =
1 to 4 °CA BTDC). For an earlier 𝜑, the differences among the 𝑋=𝑓(𝜑) characteristics decrease.
For change evaluation of the quantitative heat release characteristics, the relationship with the
𝜂 uses 50% of the released heat phase CA in practice is commonly applied to engine cycle process
energy efficiency evaluation [30,33]. A direct relationship exists between the CCR NG and
CA parameter interaction: in the HLM, when CA increased from 0 to 0.8, an increase of
4 °CA occurred; in the MLM, an increase of 12 °CA occurred; and in the LLM, an increase of 42 °CA
occurred. When increasing the engine cycle dynamics (𝜑 = 13 °CA BTDC), the CA exchange of
CCR NG in equal parts decreased intensely; for example, up to 6 °CA and 14 °CA for the MLM and
LLM, respectively.
Based on the postulate that the ignition delay has a significant influence (𝜑, based on evaluation
of the I.I. Vibe heat release model) on 𝜂 [55,56], it is appropriate to analyse the relationship between
𝜑 and the engine cycle parameters. For this purpose, Woschi’s mathematical modelling I.I. Vibe
combustion period 𝜑 analytical dependency on the engine cycle implementation parameters was
used [57,58]:
𝜑

=𝜑

󰇡


󰇢

󰇡



󰇢

.
Figure 10.
Dual-fuel engine CCR NG fuel portion influence on heat release characteristics: (
a
) HLM;
(b) MLM; (c) LLM.
Based on the postulate that the ignition delay has a significant influence (
ϕz
, based on evaluation of
the I.I. Vibe heat release model) on
ηi
[
54
–
56
], it is appropriate to analyse the relationship between
ϕz
and
the engine cycle parameters. For this purpose, Woschi’s mathematical modelling I.I. Vibe combustion
period ϕzanalytical dependency on the engine cycle implementation parameters was used [57,58]:
ϕz=ϕz0 n
n0!mα0
αk
.
Here, the “0” index is assigned to the respective parameter values, and is also known as the
“basis” engine cycle mode, for which it is usual to accept the nominal power mode;
n
indicates the
engine revolutions;
α
is the excess air ratio; and
m
and
k
are engine constants over a wide range of fuel
types [
56
,
57
]. It should be noted that the
ϕz
dependency was established based on a large scope of
diesel engine experimental data.
In accordance with the
ϕz
expression during
n=idem
, the variable (
α
) remains the main
influencing factor on the
ϕz
period, as well as when evaluating the relationship between
ηi
and
ϕz
,
and the energy efficiency in the engine cycle parameter
ηi
[
58
,
59
]. A change in the CCR NG fuel portion
in different load modes attests to the following (see Figure 11):

Energies 2019,12, 2413 24 of 32
•
An increase in the amount of CCR NG has an inverse effect on the
α
value, which can be explained
by a partial exchange of air with NG, as well as NG having a larger stoichiometric coefficient
compared to diesel—17 kg air/kg NG versus 14.6 kg air/kg diesel when the engine is operating in
different modes (when the engine is operating in different modes, the CCR NG/air inlet pressure
Pkremains constant).
•
The largest effect from the CCR NG fuel portion on
α
was detected in the LLM, and decreased
significantly when ϕinj was advanced.
Energies 2019, 10, x FOR PEER REVIEW 25 of 33
Here, the “0” index is assigned to the respective parameter values, and is also known as the
“basis” engine cycle mode, for which it is usual to accept the nominal power mode; 𝑛 indicates the
engine revolutions; 𝛼 is the excess air ratio; and 𝑚 and 𝑘 are engine constants over a wide range
of fuel types [56, 57]. It should be noted that the 𝜑 dependency was established based on a large
scope of diesel engine experimental data.
In accordance with the 𝜑 expression during 𝑛=𝑖𝑑𝑒𝑚, the variable (𝛼) remains the main
influencing factor on the 𝜑 period, as well as when evaluating the relationship between 𝜂 and 𝜑,
and the energy efficiency in the engine cycle parameter 𝜂 [58,59]. A change in the CCR NG fuel
portion in different load modes attests to the following (see Figure 11):
• An increase in the amount of CCR NG has an inverse effect on the 𝛼 value, which can be
explained by a partial exchange of air with NG, as well as NG having a larger stoichiometric
coefficient compared to diesel—17 kg air/kg NG versus 14.6 kg air/kg diesel when the engine is
operating in different modes (when the engine is operating in different modes, the CCR NG/air
inlet pressure 𝑃 remains constant).
• The largest effect from the CCR NG fuel portion on 𝛼 was detected in the LLM, and decreased
significantly when 𝜑 was advanced.
(a)
(b)
Figure 11. Cont.

Energies 2019,12, 2413 25 of 32
Energies 2019, 10, x FOR PEER REVIEW 26 of 33
(c)
Figure 11. Influence of dual-fuel CCR NG portion and 𝜑 on 𝛼 (𝑛 = 2000 rpm): (a) HLM; (b) MLM;
(c) LLM. The dots in the graph represent the experimental data.
The changes in the parameter (𝛼) are attuned with the changes in the parameter (𝜂) arising
from the same factors of CCR NG and 𝜑 (see Figure 12 compared to Figure 11).
(a)
Figure 11.
Influence of dual-fuel CCR NG portion and
ϕinj
on
α
(
n=
2000
rpm)
: (
a
) HLM; (
b
) MLM;
(c) LLM. The dots in the graph represent the experimental data.
The changes in the parameter (
α
) are attuned with the changes in the parameter
(ηi)
arising from
the same factors of CCR NG and ϕinj (see Figure 12 compared to Figure 11).
On this basis, the composed graphical
ηi
and
α
dependencies (see Figure 13a) confirm that
α
has a
meaningful influence on the engine cycle
ηi
. The investigated engine load modes of
Pme −idem
exhibited
similar dependencies of
ηi=f(α)
, independent of the CCR NG and
ϕinj
values. The determinant
coefficient R2=0.8 to 0.994 attests to the strong correlations of the ηiand αparameters.
Energies 2019, 10, x FOR PEER REVIEW 26 of 33
(c)
Figure 11. Influence of dual-fuel CCR NG portion and 𝜑 on 𝛼 (𝑛 = 2000 rpm): (a) HLM; (b) MLM;
(c) LLM. The dots in the graph represent the experimental data.
The changes in the parameter (𝛼) are attuned with the changes in the parameter (𝜂) arising
from the same factors of CCR NG and 𝜑 (see Figure 12 compared to Figure 11).
(a)
Figure 12. Cont.

Energies 2019,12, 2413 26 of 32
Energies 2019, 10, x FOR PEER REVIEW 27 of 33
(b)
(c)
Figure 12. Influence of dual-fuel CCR NG portion and 𝜑 on 𝛼 (𝑛 = 2000 min): (a) HLM; (b)
MLM; (c) LLM. The dots in the graph represent the experimental data.
On this basis, the composed graphical 𝜂 and α dependencies (see Figure 13a) confirm that α
has a meaningful influence on the engine cycle 𝜂. The investigated engine load modes of 𝑃 −𝑖𝑑𝑒𝑚
exhibited similar dependencies of 𝜂=𝑓(𝛼), independent of the CCR NG and 𝜑 values. The
determinant coefficient 𝑅 = 0.8 to 0.994 attests to the strong correlations of the 𝜂 and α parameters.
The ratio change of the dependency of parameters 𝜂 and 𝛼 comes close to a functional with
𝑅≈1.0 (see Figure 13b).
Figure 12.
Influence of dual-fuel CCR NG portion and
ϕinj
on
α
(
n=
2000
min−1)
: (
a
) HLM; (
b
) MLM;
(c) LLM. The dots in the graph represent the experimental data.
The ratio change of the dependency of parameters
ηi
and
α
comes close to a functional with
R2≈1.0 (see Figure 13b).
In a practical sense, the increase in the parameter (
α
) for a converted DE does not result in many
technological difficulties. The increase in the supply of air to the ICE engine cycle can be ensured
by changing the air compression equipment to that with a higher capacity or modifying the current
compressor [52].

Energies 2019,12, 2413 27 of 32
Energies 2019, 10, x FOR PEER REVIEW 28 of 33
(a) (b)
Figure 13. (a) Relationship between 𝜂

and α during dual-fuel engine operation with different loads
of CCR NG and 𝜑

(CCR NG = 0 to 0,8; 𝜑

= 1 to13°CA BTDC). (b). Relationship between 𝜂

 and
𝛼 with dual-fuel engine operating on different loads of CCR NG and 𝜑

.
In a practical sense, the increase in the parameter (𝛼) for a converted DE does not result in many
technological difficulties. The increase in the supply of air to the ICE engine cycle can be ensured by
changing the air compression equipment to that with a higher capacity or modifying the current
compressor [52].
The established relationship between 𝜂 and α attests to the fact that α is not the only variable
that forms the 𝜂 value. The engine cycle energy efficiency is also determined by other factors,
because it is demonstrated that an increase in the engine load, at the same values of α, leads to an
increase in 𝜂 (see Figure 13a). Additional factors influencing 𝜂 may be other engine cycle
dynamics parameters: 𝑃 and the change in pressure λ=𝑃
 𝑃
⁄ (here, 𝑃 is the pressure at
TDC) [58]. An increase in the load leads to an increase in the engine cycle dynamic parameters. It also
likely that an increase in the fuel macro- and micro-turbulence influences the combustion intensity
positively [40,58]. During large load modes, an increase in the air inlet pressure and the respective
air–NG mixture movement speed through the air inlet valve accelerates the fuel combustion main
diffusion phase, resulting in an increase in 𝜂 [59].
Micro- and macro-turbulence accelerate the active radical OH and local combustion centre
diffusion into the cylinder periphery, ensuring an entire volume combustion process [40,60].
An analogous energy efficiency increase result was obtained in the works of [28,33], where a
non-traditional CIE cycle HRF injection timing angle advancement of up to 50 °CA BTDC was used.
As noted previously, the energy efficiency increase of a converted CIE for dual-fuel operation also
manifests itself as a method for emission control when operating on NG. An increase in the air mass
during the engine cycle leads to a decrease in the PM, CO, and HC emissions [19,61,62,63]. As well
as the temperature field being equalised in the cylinder during combustion, there is also a decrease
in the number of combustion centre concentrations in the cylinder, and as a result, a decrease in the
NO emissions [27,33,63]. An important improvement effect on 𝜂 is the greenhouse gas CO
emission decrease, as observed in the LLM and MLM, owing to the smaller portion of carbon in the
chemical composition of NG of 0.75 compared to diesel with 0.86.
4. Conclusion
The energy efficiency and emissions of a converted (Bore 79.5 mm/Stroke 95 mm) CIE for
operation on dual D and NG fuel were established through experimental research:
Figure 13.
(
a
) Relationship between
ηi
and
α
during dual-fuel engine operation with different loads of
CCR NG and
ϕinj
(CCR NG =0 to 0, 8;
ϕinj =
1
to
13
◦CA BTDC
). (
b
) Relationship between
ηi
and
α
with dual-fuel engine operating on different loads of CCR NG and ϕinj .
The established relationship between
ηi
and
α
attests to the fact that
α
is not the only variable that
forms the
ηi
value. The engine cycle energy efficiency is also determined by other factors, because it is
demonstrated that an increase in the engine load, at the same values of
α
, leads to an increase in
ηi
(see
Figure 13a). Additional factors influencing
ηi
may be other engine cycle dynamics parameters:
Pmax
and the change in pressure
λ=Pmax/Pc
(here,
Pc
is the pressure at TDC) [
58
]. An increase in the
load leads to an increase in the engine cycle dynamic parameters. It also likely that an increase in
the fuel macro- and micro-turbulence influences the combustion intensity positively [
40
,
58
]. During
large load modes, an increase in the air inlet pressure and the respective air–NG mixture movement
speed through the air inlet valve accelerates the fuel combustion main diffusion phase, resulting in an
increase in ηi[59].
Micro- and macro-turbulence accelerate the active radical OH and local combustion centre
diffusion into the cylinder periphery, ensuring an entire volume combustion process [40,60].
An analogous energy efficiency increase result was obtained in the works of [
28
,
33
], where a
non-traditional CIE cycle HRF injection timing angle advancement of up to 50
◦
CA BTDC was used.
As noted previously, the energy efficiency increase of a converted CIE for dual-fuel operation also
manifests itself as a method for emission control when operating on NG. An increase in the air mass
during the engine cycle leads to a decrease in the PM, CO, and HC emissions [
19
,
61
–
63
]. As well
as the temperature field being equalised in the cylinder during combustion, there is also a decrease
in the number of combustion centre concentrations in the cylinder, and as a result, a decrease in the
NOx
emissions [
27
,
33
,
63
]. An important improvement effect on
ηi
is the greenhouse gas
CO2
emission
decrease, as observed in the LLM and MLM, owing to the smaller portion of carbon in the chemical
composition of NG of 0.75 compared to diesel with 0.86.
4. Conclusions
The energy efficiency and emissions of a converted (Bore 79.5 mm/Stroke 95 mm) CIE for operation
on dual D and NG fuel were established through experimental research:
•
A deterioration in the energy efficiency was established when increasing the NG portion in the
energy balance up to 80% or the CCR NG up to 0.8: in the HLM, 7 to 15%; in the MLM, 15 to 38%,
and in the LLM, 40 to 45%.

Energies 2019,12, 2413 28 of 32
•
An increase occurred in the incomplete combustion products in exhaust gases: the HC could
exceed 30 times in the HLM and MLM and up to 80 times in LLM with 0.1 g/kWh while operating
on D; the
CO
increased up to 30 times from 0.5 g/kWh; although the emissions of the most harmful
pollutant, NOx, decreased approximately 4 to 5 times.
•
The pilot D portion HRF injection timing DIT optimisation with respect to energy efficiency could
be practically effective only when it did not exceed a CCR NG of 0.4 when the engine was operating
in the HLM (
Pme ∼
6
bar)
: by advancing DIT by 3 to 6
◦CA
, the
NOx
+HC emissions decreased
by 10 to 15%; in the MLM and LLM, DIT optimisation did not have a positive effect.
The indicated process parameter analysis in the investigated CCR NG =0 to 0,8 and DIT =–1 to
–13
◦CA BTDC
range attests to the fact that conversion of a CIE for NG operation does not undermine
the CIE characteristic interdependent correlated cycle indicated efficiency and characteristic parameters
(such as
α
and
Pmax
). On this basis, one of the main reasons for the deterioration in
ηi
when the
engine is converted for NG operation was the NG combustion phase increase in the expansion stroke,
and therewith, the associated cylinder air excess deterioration (evaluated by the air excess ratio α).
The research results can be used for the evaluation of similar experiments, in which converted
CIEs with conventional fuel injection systems for operation with dual-fuel feed without significant
changes to the engine design were used.
Author Contributions:
Conceptualization, S.L. and S.P.; methodology, S.L., V.D. and A.R.; software, M.M and
V.D.; formal analysis, V.D.; validation, A.R. and M.M.; writing – original draft preparation, S.L., L.J. and V.D.;
writing – review and editing, S.L. and L.J.; supervision, S.L. and S.P.; project administration, S.L.
Funding: This research received no external funding.
Conflicts of Interest: The authors declare no conflicts of interest.
Nomenclature
kOptical absorption coefficient (m−1)
nRotational speed of the crankshaft (min−1)
TgExhaust gas temperature (K)
TKAir temperature after compressor (K)
PeBrake power (kW)
αExcess air coefficient
βFuel excess coefficient
εCompression ratio
ϕinj φinjHigh reaction fuel injection time (◦CA)
HULower heating value (kJ/kg)
ηmMechanical efficiency coefficient
ηiIndicated thermal efficiency
ηeBrake thermal efficiency
Pmax Maximal cylindrical pressure (bar)
PkAir pressure after compressor (bar)
λ=Pz
PcCylinder pressure increase rate
X=f(ϕ)Relative heat release ratio (◦CA)
dQ
dϕHeat release rate (kJ/◦CA)
Pme Brake mean effective pressure (bar)
Pmi Indicated mean effective pressure (bar)
Pcyl Pressure in cylinder (bar)
CA50 Half of heat released during cycle (◦CA)
(dP/dϕ)max Maximum pressure increase rate in cylinder (bar/◦CA)
(dP/dϕ)mid.Average pressure increase rate in cylinder (bar/◦CA)

Energies 2019,12, 2413 29 of 32
Abbreviations
CIE Compression ignition engine
CR Common rail fuel injection
CCR NG Co-combustion ratio of natural gas
CC Combustion chamber
DIT Diesel fuel injection timing
HRR Heat release rate
HRF High reaction fuel
HLM High load mode
MLM Medium load mode
LLM Low load mode
ULLM Ultra-low load mode
BTDC Before top dead centre
ATDC After top dead centre
CA Crank angle
CNG Compressed natural gas
LNG Liquefied natural gas
CO Carbon monoxide
HC Hydrocarbon
NG Natural gas
NOxNitrous oxide
CO2Carbon dioxide
References
1.
Chen, Z.; Yao, C.; Wang, Q.; Han, G.; Dou, Z.; Wei, H.; Wang, B.; Liu, M.; Wu, T. Study of Cylinder-to-Cylinder
Variation in a Diesel Engine Fueled with Diesel/Methanol Dual Fuel. Fuel 2016,170, 67–76. [CrossRef]
2.
Greenhouse Gas Emissions from Transport. Available online: https://www.eea.europa.eu/data-and-maps/
indicators/transport-emissions-of-greenhouse-gases/transport-emissions-of-greenhouse-gases- 11 (accessed
on 30 May 2019).
3.
Kumar, S.; Cho, J.; Park, J.; Moon, I. Advances in Diesel–Alcohol Blends and Their Effects on The Performance
and Emissions of Diesel Engines. Renew. Sustain. Energy Rev. 2013,22, 46–72. [CrossRef]
4.
Cheenkachorn, K.; Poompipatpong, C.; Ho, C. Performance and Emissions of a Heavy-Duty Diesel Engine
Fuelled With Diesel and LNG (Liquid Natural Gas). Energy 2013,53, 52–57. [CrossRef]
5.
Frei, J.K.; Orenic, C.; Smith, N. Effects of Acid Rain on Epiphytic Orchid Growth. Stud. Environ. Sci.
1984
,25,
271–285. [CrossRef]
6.
Pedrozo, V.; May, I.; Dalla Nora, M.; Cairns, A.; Zhao, H. Experimental Analysis of Ethanol Dual-Fuel
Combustion in a Heavy-Duty Diesel Engine: An Optimisation at Low Load. Appl. Energy
2016
,165, 166–182.
[CrossRef]
7.
Nithyanandan, K.; Lin, Y.; Donahue, R.; Meng, X.; Zhang, J.; Lee, C. Characterization of Soot From Diesel-CNG
Dual-Fuel Combustion In A CI Engine. Fuel 2016,184, 145–152. [CrossRef]
8.
Liu, J.; Zhang, X.; Wang, T.; Zhang, J.; Wang, H. Experimental and Numerical Study of The Pollution
Formation in a Diesel/CNG Dual Fuel Engine. Fuel 2015,159, 418–429. [CrossRef]
9.
Arteconi, A.; Brandoni, C.; Evangelista, D.; Polonara, F. Life-Cycle Greenhouse Gas Analysis of LNG As A
Heavy Vehicle Fuel in Europe. Appl. Energy 2010,87, 2005–2013. [CrossRef]
10.
Banapurmath, N.; Budzianowski, W.; Basavarajappa, Y.; Hosmath, R.; Yaliwal, V.; Tewari, P. Effects
of Compression Ratio, Swirl Augmentation Techniques and Ethanol Addition on the Combustion Of
CNG–Biodiesel In A Dual-Fuel Engine. Int. J. Sustain. Eng. 2013,7, 55–70. [CrossRef]
11.
Abagnale, C.; Cameretti, M.; De Simio, L.; Gambino, M.; Iannaccone, S.; Tuccillo, R. Numerical Simulation
and Experimental Test of Dual Fuel Operated Diesel Engines. Appl. Therm. Eng.
2014
,65, 403–417. [CrossRef]
12.
Athenstaedt, G. Entwiklung Stationarer Gasmotoren Seit Dem Inkrafttreten Der TA-Luft; Springer: Wiesbaden,
Germany, 1993; p. 4.
13. Van Basshuysen, R.; Schäfer, F. Handbuch Verbrennungsmotor; Springer: Berlin, Germany, 2007; p. 1032.

Energies 2019,12, 2413 30 of 32
14.
Dietrich, W. Die Gemischbildung Bei Gas—Und Dieselmotoren Sowie Ihr Einfluss Auf Die
Schadstoffemissionen—Rückblick Und Ausblick Teil 1. Mtz-Mot. Z. 1999,60, 28–38. [CrossRef]
15. Woschni, G. Verbrennungsmotoren; TU Munchen: Munchen, Germany, 1988; p. 303.
16.
Boretti, A. Numerical Study of The Substitutional Diesel Fuel Energy in a Dual Fuel Diesel-LPG Engine with
Two Direct Injectors Per Cylinder. Fuel Process. Technol. 2017,161, 41–51. [CrossRef]
17.
Wang, B.; Li, T.; Ge, L.; Ogawa, H. Optimization of Combustion Chamber Geometry for Natural Gas Engines
with Diesel Micro-Pilot-Induced Ignition. Energy Convers. Manag. 2016,122, 552–563. [CrossRef]
18.
Nithyanandan, K.; Zhang, J.; Li, Y.; Meng, X.; Donahue, R.; Lee, C.; Dou, H. Diesel-Like Efficiency Using
Compressed Natural Gas/Diesel Dual-Fuel Combustion. J. Energy Resour. Technol.
2016
,138, 052201.
[CrossRef]
19.
Carlucci, A.; Laforgia, D.; Saracino, R. Combustion Development and Exhaust Emissions of a Dual-Fuel DI
Diesel Engine With Variable in-Cylinder Bulk Flow and Methane Supply Strategies. In Proceedings of the
ASME 2009 Internal Combustion Engine Division Fall Technical Conference, Lucerne, Switzerland, 27–30
September 2009.
20.
Carlucci, A.; Laforgia, D.; Saracino, R. Effects of In-Cylinder Bulk Flow and Methane Supply Strategies on
Charge Stratification, Combustion and Emissions of a Dual-Fuel DI Diesel Engine; SAE Technical Paper Series;
SAE International: Detroit, MI, USA, 2009. [CrossRef]
21.
Mustafi, N.; Raine, R.; Verhelst, S. Combustion and Emissions Characteristics of a Dual Fuel Engine Operated
on Alternative Gaseous Fuels. Fuel 2013,109, 669–678. [CrossRef]
22.
Banapurmath, N.; Tewari, P.; Yaliwal, V.; Kambalimath, S.; Basavarajappa, Y. Combustion Characteristics of
A 4-Stroke CI Engine Operated on Honge Oil, Neem and Rice Bran Oils When Directly Injected and Dual
Fuelled with Producer Gas Induction. Renew. Energy 2009,34, 1877–1884. [CrossRef]
23.
Ramadhas, A.; Jayaraj, S.; Muraleedharan, C. Dual Fuel Mode Operation in Diesel Engines Using Renewable
Fuels: Rubber Seed Oil and Coir-Pith Producer Gas. Renew. Energy 2008,33, 2077–2083. [CrossRef]
24.
Garc
í
a Valladolid, P.; Tunestål, P.; Monsalve-Serrano, J.; Garc
í
a, A.; Hyvönen, J. Impact of Diesel Pilot
Distribution on The Ignition Process of a Dual Fuel Medium Speed Marine Engine. Energy Convers. Manag.
2017,149, 192–205. [CrossRef]
25.
Carlucci, A.; Ficarella, A.; Laforgia, D.; Strafella, L. Improvement of Dual-Fuel Biodiesel-Producer Gas Engine
Performance Acting on Biodiesel Injection Parameters and Strategy. Fuel 2017,209, 754–768. [CrossRef]
26.
Zhang, C.; Song, J. Experimental Study of Co-Combustion Ratio on Fuel Consumption and Emissions of
NG–Diesel Dual-Fuel Heavy-Duty Engine Equipped with a Common Rail Injection System. J. Energy Inst.
2016,89, 578–585. [CrossRef]
27.
Papagiannakis, R.; Rakopoulos, C.; Hountalas, D.; Rakopoulos, D. Emission Characteristics of High Speed,
Dual Fuel, Compression Ignition Engine Operating in a Wide Range of Natural Gas/Diesel Fuel Proportions.
Fuel 2010,89, 1397–1406. [CrossRef]
28.
Carlucci, A.; Ficarella, A.; Laforgia, D. Potentialities of A Common Rail Injection System for The Control of
Dual Fuel Biodiesel-Producer Gas Combustion and Emissions. J. Energy Eng.
2014
,140, A4014011. [CrossRef]
29.
Carlucci, A.; Colangelo, G.; Ficarella, A.; Laforgia, D.; Strafella, L. Improvements in Dual-Fuel
Biodiesel-Producer Gas Combustion at Low Loads Through Pilot Injection Splitting. J. Energy Eng.
2015
,141,
C4014006. [CrossRef]
30.
Singh, R.; Singh, S.; Pathak, B. Investigations on Operation of CI Engine Using Producer Gas and Rice Bran
Oil in Mixed Fuel Mode. Renew. Energy 2007,32, 1565–1580. [CrossRef]
31.
Yousefi, A.; Birouk, M.; Guo, H. An Experimental and Numerical Study of The Effect of Diesel Injection
Timing on Natural Gas/Diesel Dual-Fuel Combustion at Low Load. Fuel 2017,203, 642–657. [CrossRef]
32.
Maghbouli, A.; Saray, R.; Shafee, S.; Ghafouri, J. Numerical Study of Combustion and Emission Characteristics
of Dual-Fuel Engines Using 3D-CFD Models Coupled With Chemical Kinetics. Fuel
2013
,106, 98–105.
[CrossRef]
33.
Maurya, R.; Mishra, P. Parametric Investigation on Combustion and Emissions Characteristics of a Dual Fuel
(Natural Gas Port Injection and Diesel Pilot Injection) Engine Using 0-D SRM and 3D CFD Approach. Fuel
2017,210, 900–913. [CrossRef]
34.
Zhang, C.; Zhou, A.; Shen, Y.; Li, Y.; Shi, Q. Effects of Combustion Duration Characteristic on the Brake
Thermal Efficiency and Nox Emission of a Turbocharged Diesel Engine Fueled with Diesel-LNG Dual-Fuel.
Appl. Therm. Eng. 2017,127, 312–318. [CrossRef]

Energies 2019,12, 2413 31 of 32
35.
Li, J.; Wu, B.; Mao, G. Research on The Performance and Emission Characteristics of the LNG-Diesel Marine
Engine. J. Nat. Gas. Sci. Eng. 2015,27, 945–954. [CrossRef]
36.
Ivanchenko, N. Visokii Nadduv Dizelei; Mashinnastraenie Leningradskaya Oddelenye: Leningrad, Russia,
1983; p. 198.
37.
Lebedevas, S.; Lebedeva, G. Mathematical Model of Combined Parametrical Analysis of in Indicator Process
and Thermal Loading on the Diesel Engine Piston. Transport 2004,19, 108–118. [CrossRef]
38. Mollenhauer, K.; Tschöke, H. Handbook of Diesel Engines; Springer: Berlin, Germany, 2010.
39. Schwarz, C.; Merker, G.; Stiesch, G.; Otto, F. Simulating Combustion; Springer: Berlin, Germany, 2005.
40.
Guzzella, L.; Onder, C. Introduction to Modeling and Control. of Internal Combustion Engine Systems; Springer:
Berlin, Germany, 2010.
41. Heywood, J. Internal Combustion Engine Fundamentals; McGraw-Hill: New York, NY, USA, 1988.
42. Lieuwen, T.; Yang, V.; Yetter, R. Synthesis Gas Combustion; CRC Press: Boca Raton, FL, USA, 2010; p. 384.
43.
Boehman, A.; Corre, O. Combustion of Syngas in Internal Combustion Engines. Combust. Sci. Technol.
2008
,
180, 1193–1206. [CrossRef]
44.
Wang, T.; Zhang, X.; Zhang, J.; Hou, X. Numerical Analysis of The Influence of The Fuel Injection Timing
and Ignition Position in a Direct-Injection Natural Gas Engine. Energy Convers. Manag.
2017
,149, 748–759.
[CrossRef]
45.
Daisho, Y.; Yaeo, T.; Koseki, T.; Saito, T.; Kihara, R.; Quiros, E. Combustion and Exhaust Emissions in a
Direct-Injection Diesel Engine Dual-Fueled with Natural Gas; SAE Technical Paper Series; SAE International:
Detroit, MI, USA, 1995.
46.
Dishy, A.; You, T.; Iwashiro, Y.; Nakayama, S.; Kihara, R.; Saito, T. Controlling Combustion and Exhaust Emissions
in A Direct-Injection Diesel Engine Dual-Fueled with Natural Gas; SAE Technical Paper Series; SAE International:
Detroit, MI, USA, 1995.
47.
Faghani, E.; Kheirkhah, P.; Mabson, C.; McTaggart-Cowan, G.; Kirchen, P.; Rogak, S. Effect of Injection Strategies
on Emissions From a Pilot-Ignited Direct-Injection Natural-Gas Engine-Part II: Slightly Premixed Combustion; SAE
Technical Paper Series; SAE Internaniotal: Detroit, MI, USA, 2017.
48.
Air Pollution. Available online: http://www.imo.org/en/OurWork/Environment/PollutionPrevention/
AirPollution/Pages/Air-Pollution.aspx (accessed on 3 April 2019).
49.
GHG Emissions. Available online: http://www.imo.org/en/OurWork/Environment/PollutionPrevention/
AirPollution/Pages/GHG-Emissions.aspx (accessed on 3 April 2019).
50.
Abdelghaffar, W. Performance and Emissions of a Diesel Engine Converted to Dual Diesel-CNG Fuelling.
Eur. J. Sci. Res. 2011,56, 279–293.
51.
Emission Standards: Europe: Nonroad Engines. Available online: https://dieselnet.com/standards/eu/
nonroad.php#vessel (accessed on 3 April 2019).
52.
Krasovskij, I. Izdatel’stvo: Mashinostroenie (Vse Knigi Izdatel’stva) Mesto Izdanija; Leningrad God: Leningrad,
Russia, 1983.
53.
Kruggel, O. Untersuchungen Zur Stickoxidminderung an Schnelllaufenden Großdieselmotoren; VEB Verlag Technik:
Berlin, Germany, 1989; pp. 29–36.
54.
Wiebe, I. Brennverlauf Und Kreisprozesse Vonverbrennungsmotoren; VEB Verlag Technik: Berlin, Germany, 1970.
55. Vibe, I. Novoe O Rabochem Cikle Dvigatelej; Moskva-Sverdlovsk: Moscow, Russia, 1962.
56.
Woschni, G. Eine Methode Zur Vorausberechnung Der Änderung Des. Brenverlaufs Mittelschnellaufender
Dieselmotoren Bei Geanderten Betriebsbedigungen; Springer: Wiesbaden, Germany, 1973; pp. 106–110.
57.
Woschni, G. Die Berechnung Der Wandverluste Und Der Thermischen Belastung Von Dieselmotoren; Springer:
Wiesbaden, Germany, 1970.
58.
Kavtaradze, R. Teorija Porshnevyh Dvigatelej. Special’nye Glavy: Uchebnik Dlja Vuzov; MGTU of N. E. Bauman:
Moscow, Russia, 2008; p. 720.
59.
Lebedev, S.; Lebedeva, G.; Matievskij, D.; Reshetov, V. Formirovanie Konstruktivnogo Rjada Porshnej Dlja Tipaža
Vysokooborotnyh Forsirovannyh Dizelej; Akademija Transporta RF: Barnaul, Russia, 2003; p. 89.
60.
Lebedev, S.; Matievskij, D. Analiz Indikatornogo KPD I Harakteristiki Teplovydelenija Dizelej Tiporazmera
CN16,518,5 Pri Ih Forsirovanii Do Pme =2,0 MPa; Vestnik Altajskogo Tehnicheskogo Universiteta: Barnaul,
Russia, 2000; pp. 103–107.
61.
Zhang, Q.; Li, M.; Shao, S. Combustion Process and Emissions of A Heavy-Duty Engine Fueled With Directly
Injected Natural Gas And Pilot Diesel. Appl. Energy 2015,157, 217–228. [CrossRef]

Energies 2019,12, 2413 32 of 32
62.
Mittal, M.; Donahue, R.; Winnie, P.; Gillette, A. Exhaust Emissions Characteristics of a Multi-Cylinder 18.1-L
Diesel Engine Converted to Fueled With Natural Gas and Diesel Pilot. J. Energy Inst.
2015
,88, 275–283.
[CrossRef]
63.
Taniguchi, S.; Masubuchi, M.; Kitano, K.; Mogi, K. Feasibility Study of Exhaust Emissions in a Natural Gas Diesel
Dual Fuel (DDF) Engine; SAE Technical Paper Series; SAE International: Detroit, MI, USA, 2012.
©
2019 by the authors. Licensee MDPI, Basel, Switzerland. This article is an open access
article distributed under the terms and conditions of the Creative Commons Attribution
(CC BY) license (http://creativecommons.org/licenses/by/4.0/).