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Effects of oxygen enriched air on the operation andperformance of a diesel-biogas dual fuel engine
Karen Cacua*, Andres Amell, Francisco Cadavid
Grupo de Ciencia y Tecnologıa del Gas y Uso Racional de la Energıa, Faculty of Engineering, University of Antioquia,
Calle 67 No 63-108 Medellın, Colombia
a r t i c l e i n f o
Article history:
Received 28 December 2010
Received in revised form
15 May 2012
Accepted 1 June 2012
Available online 30 June 2012
Keywords:
Dual fuel engine
Oxygen enriched air
Biogas
Ignition delay
Thermal efficiency
* Corresponding author. Tel.: þ57 4 219 85 45E-mail addresses: karenpaolacacua@hotm
0961-9534/$ e see front matter ª 2012 Elsevhttp://dx.doi.org/10.1016/j.biombioe.2012.06.
a b s t r a c t
The effect of oxygen enriched air was tested for a diesel-biogas dual fuel engine. The
operation and performance characteristics, such as thermal efficiency, pollutant emissions
and combustion parameters were determined. Experiments have been carried out with
a stationary compression ignition (CI) engine coupled with a generator in dual mode using
a typical biogas composition of 60 vol. %CH4 and 40% vol. %CO2. For every engine load
evaluated, the oxygen concentration in the intake air engine was varied from 21% to 27% O2
v/v. Ignition delay time and methane emissions were decreased when using oxygen
enriched air with respect to normal air (21%O2), whereas the thermal efficiency was
increased.
ª 2012 Elsevier Ltd. All rights reserved.
1. Introduction There are some problems associated with the performance
Biogas is an alternative energy source and it is produced from
anaerobic fermentation of organic material. It is a low heating
value fuel and methane and carbon dioxide are its main
components. Rural households use biogas for cooking, while
farms normally go for heat, shaft power and electricity gener-
ation. Because of its high octane number, biogas is suitable for
engines with a relatively high compression ratio to maximize
thermal efficiency and may be applied to conventional
compression ignition engines with minor modifications [1].
A dual fuel engine is a diesel engine operatingwith gaseous
fuels whilemaintaining some liquid fuel injection as source of
ignition. The main objective of dual engines is to reduce the
use of fossil fuels and maximize their substitution with
alternative fuels, in attention to economic and environmental
reasons [2].
; fax: þ57 4 211 90 28.ail.com, karen.cacua@ud
ier Ltd. All rights reserved003
of dual engines. At light load, the dual engine tends to exhibit
lower fuel utilization, low thermal efficiency, higher pollutant
emissions and long ignition delay. This is due to low ignit-
ability of gaseous fuels like methane and the dilution with
CO2. On the other hand, operation at light load is associated
with a greater degree of cyclic variations in performance
parameters, such as peak cylinder pressure and ignition delay.
The principal cause of this behavior is the low flame propa-
gation velocity from the pilot fuel ignition into the lean
gaseous fuel mixture [3].
Ignition delay is a critical parameter to control the perfor-
mance and emissions of internal combustion engines. In dual
engines, ignition delay increases due to a reduction in the
partial pressure of oxygen in the intake air, reduced reaction
activity because of the inert in the fuel, and changes in the
effective temperature during compression. A long and
ea.edu.co (K. Cacua)..
Table 1 e Test engine characteristics.
Type Lister Petter TR2, DI, four stroke,
two cylinders, naturally aspirated,
air cooled
Displacement 1.55 � 10�3 m3
Bore � stroke 0.098 � 0.101 m
Compression ratio 15.5:1
Rated power 20 kW at 3000 rpm
Maximum torque 76 Nm at 1800 rpm
Inlet valve open 36� BTDC (Before
Top Dead Center)
Exhaust valve close 32� BTDC
b i om a s s an d b i o e n e r g y 4 5 ( 2 0 1 2 ) 1 5 9e1 6 7160
variable ignition delay time is undesirable as it leads to an
increase in the premixed part of combustion in a heat release
diagram. This produces the reduction in engine efficiency,
increase in exhaust emissions, and damage in mechanical
parts [3e7].
Many research efforts have been done to provide effective
solutions for further improvement of dual operation at light
load. Such solutions include changes in the initial charge
temperature and pressure, quantity and quality of liquid fuels,
air-fuel ratio, and injection characteristics of liquid fuels such
as increasing injection pressures. All these measures cause
increases in thermal efficiency and decreases in pollutants
emissions and ignition delay times [3,8e11].
Another possible and simpler solution to improve the
operation of dual diesel engines at light load is the increase of
oxygen concentration in the air intake up to 30%O2 by volume.
This has shown successful results in spark ignition and
compression ignition engines such as increases in power
density, thermal efficiency and decreases in pollutants emis-
sions [12e15]. All of these effects are due to increase in
burning velocities. However, NOx emissions were increased
due to higher temperatures inside the combustion chamber
[16e18]. Other researchers have conducted experiments with
oxygen concentrations below 24% O2 by volume. For oxygen
concentrations up to 23%, the particulate matter decreased as
well as the ignition delay, and NOx emissions were in
a permissible range [12,19,20]. Additionally, the combination
of water-diesel emulsions and oxygen enrichment (24%
molar) was assessed. An increase of 10% in effective efficiency
decreases in particulate matter and a low increase in NOx
emissions were obtained at full load [19].
In spark ignition engines with oxygen enriched air and
gasoline as fuel, several researches found a decrease in carbon
monoxide and hydrocarbon emissions, and an increase in the
effective power in whole range operation [14,21e23]. Maxwell,
Setty, Jones and Narayan [23] worked with oxygen concen-
trations in air of 23%and 25% O2 v/v in a spark ignition engine
fueled with gasoline and natural gas. An Increase in effective
power by 5e17%, a decrease in carbonmonoxide emissions by
25e32%, as well as a decrease in hydrocarbons by 30e40%
were obtained.
There are several processes currently available for
producing oxygen. The most efficient at large scale is cryo-
genic air separation. Pressure-swing absorption is used at
medium to small scale. The most common method to deliver
air enriched with oxygen to engines is selective permeation
through nonporous, polymeric membranes. However, their
costs are higher and their operation requires high pressures,
resulting in additional cost of the engine tests [21,24]. Because
of the recent progress in nonporous polymeric membrane
methods to enhance oxygen in air and the reduction in their
production costs, this solution will be technically and
economically feasible in few years [13e15,21,25e31].
It is expected that the use of oxygen enriched air in
a biogas-diesel dual engine attenuates the effect of CO2 in
decreasing the laminar burning velocity, adiabatic flame
temperature and ignition delay time of methane [32]. More-
over, an increase in thermal efficiency and a decrease in
pollutants emissions are expected. However, information
about this issue for dual diesel-biogas engines is limited in
literature. Therefore, in this study the results of the effect of
varying the concentration of oxygen from 21 to 27% O2 by
volume in a dual diesel-biogas engine are discussed.
2. Experimental methodology
2.1. Experimental setup
The experimental tests were performing at a region placed at
1500 m over the sea level (Medellın e Colombia). The local
environmental conditions were 298 K and 85.3 kPa. A
stationary CI engine was coupled with a generator to run at
maximum torque speed (1800 rpm). Table 1 shows the tech-
nical engine characteristics [33].
The experimental work started with preliminary investi-
gation of the engine running on neat diesel fuel, in order to
determine its performance characteristics. Electric power
outputs at 40%, 50%, 70% and 100% of full load were obtained.
Engine loadswere set from3 to 10 kWwith a variable electrical
resistance bank connected to the electricity generator.
In the dual fuel mode, Colombian commercial diesel was
used and biogas was simulated with a typical composition of
60%CH4and40%CO2onavolumetric basis. TheflowrateofCO2
and CH4 were both measured with hot wire sensors (Omega;
FM5400). Theflow rate ofdiesel fuelwasmeasuredwithCoriolis
sensor (Siemens; SITRANS 2100 DI). Table 2 summarizes
important properties of the fuels used in the experiments.
The load of the engine was fixed and the biogas was
injected into the air intake manifold using a “Tee” mixer at
a point that ensures homogenous mixture. Parameters such
as power produced by the engine, engine speed, fuel
consumption, air flow, temperatures and emission charac-
teristics were measured.
Oxygen (99.9% O2) from gas bottles was injected in counter
flow via a tube with three orifices into the air intake manifold.
The air composition in a volumetric basiswasmeasured at the
duct using a gas analyzer with the paramagnetic technique
(MAIHAK S710). An orifice meter and a U-tube manometer
were used to measure the air consumption of the engine.
The in-cylinder pressure was measured using a piezo-
electric pressure transducer (Kistler; model: 6125B) which
was flush mounted in the cylinder head. The air intake
pressure was measured with a piezoresistive pressure
Table 2 e Fuel properties.
Property Diesel Biogas Natural gas Guajira
API gravity at 60 �F 31.9 e e
Low heating value (MJ/kg) 43 23.73 48.77
Cetane number 44 e e
Viscosity at (m2/s) 4.66 � 10�6 e e
Cloud point (K) 274 e e
Simplified chemical composition C10.8H18.7 60% CH4;
40% CO2
(by volume)
97.76% CH4; 0.38%
C2H6; 0.2% C3H8; 1.29% N2;
0.37% CO2 (by volume)
Stoichiometric air-fuel ratio (AFR) 14.32 6.05 e
Lower Woobe index (kJ/Nm3) e 22.176 52.344
Methane number e 160 e
b i om a s s a n d b i o e n e r g y 4 5 ( 2 0 1 2 ) 1 5 9e1 6 7 161
transducer (Kistler; Model: 4005A) mounted in the air intake
manifold. Its output signal was connected to a charge
amplifiers mounted in a signal conditioner (Kistler; model:
2853A). To measure the crank angle position, a precision
shaft encoder (Kistler; model: 2614A) was coupled with the
engine crank-shaft. The in-cylinder pressure-crank angle
history data acquisition was performed using a program
based on LabView. In cylinder pressure-crank angle history of
600 consecutive cycles was recorded for each test conditions.
The ignition delay and the heat release rate were calculated
using the mean measured cylinder pressure diagram and the
crank angle signal flow rate signals of fuels and air, as well as
pressure signal in the cylinder and crank angle were recorder
on a personal computer via a National Instrument� acquisi-
tion card [34].
The exhaust emissions were measured using a non
dispersive infrared sensor for CO2, CH4 and CO and para-
magnetic sensor for O2 in a gas analyzer (MAIHAK 610). K-type
fine thermocouples were used for measuring the mean
temperatures of the exhaust gas, cooling air inlet, and engine
lubricating oil. In Fig. 1 a schematic diagram of the experi-
mental setup is shown.
2.1.1. Uncertainty analysis of the experimental dataIn the experimental stage, three operating modes were
chosen: Diesel mode, dual mode and dual mode with air
enriched with oxygen. The accuracy of the measurements
was estimated determining the coefficient of variance (COV)
for each measured parameter. COV of parameters was calcu-
lated using Equation (1):
COVðxÞ ¼ sx,100% (1)
where x ¼ Pni¼1xi=n and (s) standard deviation
s ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiXni¼1
ðxi � xÞ=ðn� 1Þs
(2)
Table 4 show the COV for the measured quantities and its
values were below 20 percent. This confirms that the
measurements were repeatable.
2.2. Experimental procedure
The maximum substitution levels in dual fuel mode were
found for each load. In the first stage of the experimental
phase, the maximum value of biogas that could be used in the
engine without shutting down or severe deterioration of
combustion stability was determined monitoring the voltage
curve vs. encoder pulses. Fig. 2 shows an unstable operation of
the engine due to excessive presence of biogas. The start of
combustion occurred after of TDC (Top Dead Center) and the
cyclic variation in the maximum pressure angle was high,
showing motored conditions in some cycles. The maximum
substitution level was found decreasing the biogas concen-
tration in the mixer inlet to re-establish the combustion
stability such as shown in Fig. 3. Furthermore, when oxygen
enriched air was used, stability was improve and the substi-
tution level was increased such as shown in Fig. 4.
The substitution level, Z, was determined using the diesel
mass flow rate in diesel mode _mD (kg/s) and dual mode _mp (kg/
s), according to Equation (2):
Z ¼ _mD � _mp
_mD,100½%� (3)
In the second experimental stage, three levels of oxygen
(22%, 25%, 27% v/v) were injected in the air intake at 40%, 50%,
and 70% of full load. This was done at conditions where the
maximum substitution level in dual fuel mode with atmo-
spheric air had been obtained.
The experimental factorial design employed to recollect
and analyze the data is shown in Table 3. Experiments were
replicated two times according to the procedure described by
Montgomery, 2004 [35]. The engine performance was evalu-
ated by comparing thermal efficiency, CO and CH4 emissions,
cylinder pressure traces, ignition delay and total heat release
rate.
The thermal efficiency was defined as the ratio of the
electric power output hE to the energy contribution of biogas
and diesel, as follows in Equation (3):
hE ¼ NE
_mBLHVB þ _mDLHVB,100½%� (4)
where _mB [kg/s] is the biogas mass flow rate, _mD [kg/s] is the
diesel mass flow rate, LHVB [kJ/kg] is the biogas low heating
value and LHVB [kJ/kg] is the diesel low heating value.
The ignition delay,qR, was defined as the crank angle
difference between the start of the diesel injection into de
combustion chamber and the start of combustion, as follows
in Equation (4):
qR ¼ qI � qINY½�CA� (5)
Fig. 1 e Schematic diagram of the experimental setup.
b i om a s s an d b i o e n e r g y 4 5 ( 2 0 1 2 ) 1 5 9e1 6 7162
where qI and qINY [�CA BTDC] are the crank angle at ignition
and at injection respectively. The start of the diesel fuel igni-
tion was estimated when a change in the slope of pressure
injection-crank angle position diagram occurred. To deter-
mine the crank angle at the start of combustion, the first
derivative of the in-cylinder pressure related to the crank
Table 3 e Experimental design for recollecting andanalysis data.
Factor Level description Level designation
Oxygen content
in air (%) by
volume
1 21
2 22
3 25
4 27
Load (%) 1 40
2 50
3 70
Engine speed (rpm) 1 1800
angle (dP/dq) was used. The slope in (dP/dq) vs. Crank angle
diagram changes its concavity when combustion starts.
3. Results and discussion
In the first experimental stage, the level of maximum substi-
tution with atmospheric air (21% O2) was determined and the
dual fuel engine performance was evaluated. Fig. 5 shows the
substitution percentage with oxygen enriched air for all loads.
Table 4 e Coefficient of variance (COV) for the measuredquantities.
Measured quantity COV (%)
Exhaust gas temperature 8.9
Methane emissions 13.9
CO emissions 17.3
Diesel flow 2.3
3000 3200 3400 3600 3800 4000 4200 4400 4600 4800 50000
1
2
3
4
5
6
Encoder pulses
Volta
ge [V
]
Start of combustion
Maximum cylinderpressure
Motored conditions
Fig. 2 e Unstable operation dual diesel-biogas engine 40%
load with 66% substitution level.
3000 3200 3400 3600 3800 4000 4200 44000
1
2
3
4
5
6
7
8
Encoder pulses
Volta
ge [V
]Fig. 4 e Stable operation dual diesel-biogas engine 40%
load with 62.4% substitution level and air enriched with
oxygen (27%O2).
b i om a s s a n d b i o e n e r g y 4 5 ( 2 0 1 2 ) 1 5 9e1 6 7 163
A higher oxygen concentration allows increasing the substi-
tution level at 50% and 70% loads due to stable operation of the
dual engine and a more uniform combustion, whereas there
are not major differences at 40% of full load.
Fig. 6 shows the ignition delay time at 40%, 50% and 70% of
full load at several levels of oxygen enrichment. The ignition
delay was lower for all the enrichment levels due to a higher
amount of oxygen available for the combustion process. This
allows the acceleration in preignition reactions of diesel. The
addition of just 1% of oxygen to the atmospheric air (21% O2)
reduced the ignition delay time. This produces better
3000 3200 3400 3600 3800 4000 4200 44000
1
2
3
4
5
6
7
Encoder pulses
Volta
ge [V
]
Fig. 3 e Stable operation dual diesel-biogas engine 40%
load with 62% substitution level.
performance characteristics of the dual engine and general
improvement in the combustion process.
For the whole load range, the thermal efficiency was
increased up to 28% in 40% of full load and 27% O2 in the air.
This is shown in Fig. 7. The improvements on the thermal
efficiency are due to an increase in the activity of the partial
oxidation reactions by improving propagation of flame fronts
from diesel and the increase of the overall mixture tempera-
tures. Both phenomena allow an increase in the reactivity of
40 50 60 7050
60
70
80
Load [%]
Subs
titut
ion
leve
l [%
]
21%O222%O225%O227%O2
Fig. 5 e Variation of the substitution level with engine load
and oxygen enriched air.
40 50 60 7012
13
14
15
16
17
18
Load [%]
Igni
tion
Del
ay [º
CA]
21%O222%O225%O227%O2
Fig. 6 e Ignition delay time at 40%, 50% and 70% loads with
engine load and oxygen enriched air.
−100 −50 0 50 1000
10
20
30
40
50
60
70
Crank Angle [ºCA]
Cyl
inde
r Pre
ssur
e [b
ar]
21%O222%O225%O227%O2
Fig. 8 e Cylinder pressure traces to crank angle position at
40% of full load for each level of enrichment.
b i om a s s an d b i o e n e r g y 4 5 ( 2 0 1 2 ) 1 5 9e1 6 7164
both fuels because of higher flame propagation velocities and
the decrease in ignition delays times of the diesel fuel [30,36].
Figs. 8, 9 and 10, show cylinder pressure traces corre-
sponding to 40%, 50% and 70% of full load for all oxygen
enrichment levels. The peak pressure was higher for 25% and
27% oxygen enrichment compared to the case of atmospheric
air for all loads. This is due to an increase in the reactivity of
both fuels during the premixed combustion stage and
a decrease in the ignition delay time. For 22%O2 the difference
40 50 60 709
10
11
12
13
14
15
16
17
18
19
20
Load [%]
Ther
mal
effi
cien
cy [%
]
21%O222%O225%O227%O2
Fig. 7 e Variation of thermal efficiency with engine load
and oxygen enriched air.
in the peak pressure was not significant regarding to atmo-
spheric air, although the ignition delay time decreased.
Figs. 11, 12 and 13, show the total heat release (dQ/dq)
related to the crank angle at 40%, 50% and 70% of full load for
all the enrichment levels. The peak value of the heat release
rate during premixed combustion of the diesel and biogas is
not differentiable in the experimental heat release rate
diagram. This is due to a long delay time,where a large portion
of pilot fuel is mixed with air, producing a fast energy release
and a high value peak of total premixed combustion.
−100 −50 0 50 1000
10
20
30
40
50
60
70
Crank Angle [ºCA]
Cyl
inde
r Pre
ssur
e [b
ar]
21%O222%O225%O227%O2
Fig. 9 e Cylinder pressure traces to crank angle position at
50% of full load for each level of enrichment.
−100 −50 0 50 1000
10
20
30
40
50
60
70
Crank Angle [ºCA]
Cyl
inde
r Pre
ssur
e [b
ar]
21%O222%O225%O227%O2
Fig. 10 e Cylinder pressure traces to crank angle position at
70% of full load for each level of enrichment.
−20 0 20 40
0
10
20
30
40
50
60
70
80
90
100
Crank Angle [ºCA]
dQ/d
Thet
a [ºC
A]
21%O222%O225%O227%O2
Fig. 12 e Total heat release traces related to crank angle
position at 50% load for the each level of oxygen
enrichment.
b i om a s s a n d b i o e n e r g y 4 5 ( 2 0 1 2 ) 1 5 9e1 6 7 165
At 40% and 50% of full load, the total heat release (dQ/dq)
showed earlier premixed combustion and lower diffusion
combustion on the expansion stroke for all the oxygen
enrichment levels, which indicates more energy being
released due to decreases in ignition delay and more efficient
combustion.
The main pollutants of the exhaust gas of dual diesel-
biogas engines are methane and carbon monoxide. When
the dual engine operates at light load, a significant amount of
the methane and products of the preignition and partial
combustion remain at the exhaust stage. This is because of
−20 0 20 40
0
10
20
30
40
50
60
70
80
90
100
Crank Angle [ºCA]
dQ
/dTh
eta
[J/ºC
A]
21%O222%O225%O227%O2
Fig. 11 e Total heat release traces related to crank angle
position at 40% load for the each level of oxygen
enrichment.
the flame fronts propagation from various ignition centers do
not extend to all regions of the cylinder [2]. However changes
were presented in the extension of the flammability interval
of mixture air-biogas when oxygen air enriched was utilized.
This improves the propagation of the flame fronts originated
from diesel with a faster heat release.
Fig. 14 shows carbon monoxide (CO) emissions variations
related to load engine for oxygen enrichment levels. At 40% of
full load and 25% O2, carbon monoxide decrease by 19.5%
−20 0 20 40
0
10
20
30
40
50
60
70
80
90
100
Crank Angle [ºCA]
dQ/d
Thet
a [J
/ºCA]
21%O222%O225%O227%O2
Fig. 13 e Total heat release traces related to crank angle
position at 70% load for the each level of oxygen
enrichment.
40 45 50 55 60 65 700.11
0.115
0.12
0.125
0.13
0.135
0.14
0.145
0.15
0.155
0.16
Load [%]
CO
[%]
21%O222%O225%O227%O2
Fig. 14 e Variation of carbon monoxide (CO) related to
engine load with enrichment oxygen.
b i om a s s an d b i o e n e r g y 4 5 ( 2 0 1 2 ) 1 5 9e1 6 7166
regarding to atmospheric air (21% O2). This is due to decreases
in fuel/air equivalence ratio and increases in preignition
reactions of biogas. At 50% of full load, carbon monoxide
emissions increase up to 11% for 22% oxygen and up to 7.5%
for 25% oxygen due to an increase in partial oxidation of
biogas and higher substitution levels regarding to atmo-
spheric air (21% O2).
Methane emissions were decrease up 35% for 27% O2 for all
loads such as shown in Fig. 15. The most noticeable change
was presented at 50% of full load and 25%O2, where a decrease
of 38% in methane emissions was reached. However, the
methane emissions were increased for all enrichments
40 50 70
0.6
0.7
0.8
0.9
1
Load [%]
CH
4 [%
]
21%O2
22%O2
25%O2
27%O2
Fig. 15 e Variation of methane emissions (CH4) related to
engine load with enrichment oxygen.
percentages. This could be due to increases in substitution
level with oxygen enrichment. A deep study is necessary to
explain this effect.
4. Conclusions
In thiswork, an experimental studywas developed to evaluate
the effects of air enriched with oxygen on a stationary dual
fuel engine performance using biogas as primary fuel. The
results showed the following:
� Small additions of O2 to intake combustion air improve
combustion stability in a biogas-diesel engine. The addi-
tional O2 helps to attenuate negative effects of CO2 in the
combustion such as decreases in overall gas-air mixture
temperature and low burning velocities of biogas regarding
to methane.
� Oxygen enrichment is a viable technique for dual diesel-
biogas engine at light loads due to improvement in impor-
tant characteristics of performance such as thermal effi-
ciency, decreases in the ignition delay, high burning rates,
as well as decreases in methane emissions.
Acknowledgments
The authors gratefully acknowledge the financial support of
“COLCIENCIAS” to the project “Optimizacion de motores
duales diesel-biogas en el piso termico Colombiano” and the
sostenibility program of Vicerrectoria de Investigacion of
University of Antioquia.
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