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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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