International journal of automotive technology, tập 10, số 2, 2009

Copyright © 2009 KSAE
1229−9138/2009/045−01

International Journal of Automotive Technology, Vol. 10, No. 2, pp. 131−139 (2009)

DOI 10.1007/s12239−009−0016−2

COMBUSTION AND EMISSION CHARACTERISTICS OF BD20
REFORMED BY ULTRASONIC ENERGY FOR DIFFERENT INJECTION
DELAY AND EGR RATE IN A DIESEL ENGINE
1)

2)

S. Y. IM , D. S. CHOI

3)*

and J. I. RYU

Graduate School of Mechanical Engineering, BK21 Mechatronics Group, Chungnam National University,
Daejeon 305-764, Korea
Division of Mechanical & Automotive Engineering, Kongju National University, Chungnam 330-717, Korea
Department of Mechanical Engineering, BK21 Mechatronics Group, Chungnam National University,
Daejeon 305-764, Korea
1)

2)

3)

(Received 25 April 2007; Revised 11 September 2008)

ABSTRACT−The purpose of this study is to understand the operational characteristics of a diesel engine that uses BD20

reformed by ultrasonic energy irradiation. In particular we study the effects of tuning injection delay and EGR rate. BD
containing about 10% oxygen has attracted attention due to soaring crude oil prices and environmental pollution. This oxygen
decreases soot by promoting combustion, but it also increases NOx. To solve this problem, injection timing may be delayed
or an EGR system may be applied. These adjustments normally lower engine power and increase exhaust emission but, in
using fuel reformed by ultrasonic energy irradiation (which is changed physically and chemically to promote combustion), we
may hope to circumvent this problem. To control the duration of the ultrasonic energy irradiation, the capacity of the chamber
in an ultrasonic energy fuel supply system was tested at 550cc and 1100cc capacities. As for the results of the experiment, we
could identify the optimum EGR rate by investigating the engine performance and the characteristics of exhaust emissions
according to the injection timing and the EGR rate while ultrasonically irradiated BD20 was fed to a commercial diesel engine.
With UBD20 (at an injection timing of BTDC 16o), the optimum EGR rate, giving satisfactory engine performance and
exhaust emissions characteristics, was in the range of 15~20%.

KEY WORDS : BD (biodiesel fuel), BD20 (diesel 80% + biodiesel fuel 20%), EGR (exhaust gas recirculation), Injection
delay, UBD (ultrasonic energy irradiated biodiesel fuel), Ultrasonic energy irradiation

compounds other than NOx are known increase with an
increase of the EGR rate. There are also a few problems to
solve in order for it to be put into practice universally,
including contamination of the engine induction system by
increase of the break specific fuel consumption (BSFC)
and soot, the abrasion and corrosion of parts in the engine,
and the blazing fire of lubricating oil (Lue et al., 2001;
Ham and Chun, 2002).
When fuel is irradiated by ultrasonic energy, the fuel
undergoes physical and chemical changes induced by the

energy of the ultrasonic irradiation. For example, the
aromatic constituents become constituents of the fatty
group, and the isoparaffins turn into normal paraffins. This
leads to an increase in the cetane number and heating
value, while decreasing viscosity, surface tension, and the
spray diameter (SMD). In this way, fuel quality is reformed,
physically and chemically, promoting combustion and decreasing soot (Choi, 1996; Jeong et al., 1991; Lee and Ryu,
2003; Song, 2005).
There is a trade-off however, as NOx reduction leads to
the increase of soot, while soot reduction leads to the

1. INTRODUCTION

Due to soaring crude oil prices and environmental problems, interest in biodiesel fuel (hereafter called BD) has
grown. BD is similar to diesel fuel and can be applied to
commercial diesel engines without any special modification.
Although BD contains about 10% oxygen and thereby
decreases soot by promoting combustion, this active combustion leads to an increase in NOx emissions due to the
increase in combustion temperature (Agarwal et al., 2006;
Ryu and Oh, 2003; Baik, 2006; Bae et al., 2002; Oh et al.,
2002).
In particular, BD20, which is a blend of diesel fuel and
BD in a volume ratio of 8:2, contains a substantial quantity
of O , making it particularly effective in reducing soot
(Baik, 2006; Bae et al., 2002; Ryu and Oh, 2003).
To overcome this problem, the EGR (Exhaust Gas Recirculation) system has been widely put into practical use.
The EGR system is inexpensive to install, but emission of
2

*Corresponding author. e-mail:

131


132

S. Y. IM, D. S. CHOI and J. I. RYU

increase of NOx. Hence, it is very difficult to reduce both
quantities simultaneously. The fuel injection timing is
markedly delayed to reduce NOx, because this leads to a
raise in the temperature of combustion. The temperature at
the intake also increases when using hot EGR. Alternatively, cooled EGR may be used to display a constant effect in
the range over mid-load. The EGR method has been widely
applied to small diesel vehicles but the soot in the exhaust
gas contains sulfur oxides that enter the engine, causing
enhanced wear on the piston and cylinders. The life of the
engine oil is also badly affected. Hence it is essential to
designate the driving range with a large amount of NOx in
order to maintain an appropriate EGR rate (Uchida, 1993).
Therefore, this study aims to decide the optimum injection timing and the EGR rate to obtain a simultaneous
reduction of soot and NOx to within minimum values,
without deteriorating engine performance using BD20
reformed by ultrasonic energy irradiation.

2. EXPERIMENTAL SYSTEM AND METHODS
2.1. Ultrasonic Energy Fuel Supply System
Figure 1 provides a schematic diagram for the ultrasonic
energy fuel supply system used for ultrasonic irradiation.

The specifications of the transducer used in this experiment

are presented in Table 1.
A bolted Langevin transducer (BLT) was used for the
ultrasonic energy fuel supply system. Its structure is such
that when an ultrasonic oscillator (50W) functions at AC
220V, ultrasonic vibrations are produced at 28 kHz and are
transmitted through the horn and into the fuel. To maximize the ultrasonic energy irradiation, fuel is supplied to
the lower part of the chamber of the ultrasonic energy fuel
supply system and flows out through the upper part. An air
vent was installed in the back of the upper reflector of the
chamber so that the air bubbles generated by the ultrasonic
cavitations could escape (Dale Ensminger, 1988). The ultrasonic energy fuel supply system was connected between a
fuel flowmeter and an injection pump so that the reformed
(irradiated) fuel could be provided to the engine.
2.2. Experimental System and Method
The schematic diagram and photograph of the experimental apparatus is presented in Figures 2 and Figure 3.
The apparatus consists of a dynamometer, a test engine,
measurement instruments, and a data acquisition system.
The specifications of the test engine are shown in Table 2.
This experiment was performed according to the KSRISO

Figure 2. Schematic diagram of experimental apparatus.
Figure 1. Schematic diagram of an ultrasonic energy fuel
supply system.
Table 1. Specifications of the bolted Langevin Transducer
(BLT).
Items
Specifications
Frequency (kHz)
28±0.4
Admittance (mΩ)

35±7
Thermal quality (Qm)
1,000
Capacitance (pF)
3000±15%
Maximum velocity (cm/s)
60
Maximum altitude (µm)
9.6
Dia. & Thickness (mm)
φ 29.5 & 5

Figure 3. Photograph of experimental apparatus.


COMBUSTION AND EMISSION CHARACTERISTICS OF BD20 REFORMED BY ULTRASONIC ENERGY
Table 2. Specifications of test engine.
Items
Specifications
Engine type
Water cooled 4 stroke cycle
4 cylinder diesel engine
Combustion chamber type Swirl combustion chamber
Valve mechanism
O.H.C
(Over Head Cam shaft)
Injection pump
Distributor type
Bore×Stroke
91.1×95 mm

Total displacement
2,476cc
Max. power
77ps/4,200 rpm
Max. torque
15.5 kgf·m/2,000 rpm
Fuel ignition timing
BTDC 16o
Coolant temperature
80±2 oC
Ignition order
1-3-4-2
Table 3. Properties of test fuels.
Biodiesel Fuel
Items
BD20 BD100
o
48
174
Flash point (PM, C)
Pour point (oC)
−17.5
−2.5
Sulfer (Wt. %)
0.018
0.011
Specific gravity (15/4oC) 0.8317 0.8815
Cetane number
50.5
57.5

Kinematic viscosity
2.614
4.255
(40oC, cSt)
Oxygen content (%)
−
11.03

Diesel
Fuel
44
−17.5
0.022
0.8211
51.8
2.350
0

2534 testing standard (KSRISO, 2003). The engine load is
adjusted after fixing the engine speed. Also, the injection
timing was set for BTDC 11o and 16o, with the engine load
set to 25%, 50%, 75% and 100% according to a maximum
torque for each engine speed. The engine speed was varied
from 1000 rpm to 3500 rpm in 500 rpm intervals.
Table 3 shows properties of the test fuels used for the
experiment. BD is made out of soybean oil and in our tests
was blended with 20% (by volumetric ratio) commercial
diesel fuel.
In order to investigate the effects of the duration of the
ultrasonic energy irradiation, the capacity of the chamber

for the ultrasonic energy fuel supply system was set to
either 500cc or 1100cc. If the capacity of the chamber is
too small, the influence of the ultrasonic energy irradiation
is not significant, and if the capacity of the chamber is too
large, the temperature of the supplied fuel becomes high
enough to negatively affect the engine. The amount of
recirculating exhaust gas is controlled by the EGR valve
after measuring the intake air volume (m3/h). The EGR rate
was computed using the following formula (1):
V 0 – V-a × 100
EGR rate ( % ) ---------------(1)
V0

133

Table 4. Specifications of measuring equipment.
Items
Specifications
Co. Model HE-130.
Dynamometer Hwanwoong
Absorption torque: 35 kgf·m
Pressure
Kistler Co. type 6052B
transducer
Piezoelectric pressure transducer
Charge
Kistler Co. type 5011B
amplifier
Oscilloscope
Tectronic Co. type DS360

A/D-D/A
National Ins. Co. type
converter
PCI-6024E
Load cell
Jungwoo Co. type JW-U2SB
AND Co. type HF-2000GD
Fuel flow meter Capacity: 2100g,
Resolution: 0.01g
Kwang precision Co. type Gold 707
Smoke meter Su
Measurable range: 0~100%±2% F.S
HORIBA KOREA Ltd. MEXA Chemical method
NOx analyzer 554JKNOX
Measurable range: 0~5,000 ppm
±20 ppm
BOSCH ETT 008.55
Non-Dispersive Infrared Method
CO analyzer
Measurable range :
0.00~10.00% vol ±0.06% vol
Here, V0 is intake air volume (m3/h) with no EGR, while Va
is that when EGR is taken into account.
Since the temperature of the EGR gas changes in accordance with each engine load, it was controlled here to be
24oC by using a cooling circulation system. A filter was
used to remove particles in the recirculated exhaust gas.
The temperature of the cooling water was maintained at
a constant 80±2oC, regardless of the testing condition.
After the completion of each experiment, the fuel filter was
replaced, and the fuel supply system was examined for

need of repair to ensure that previous experiments would
not affect subsequent ones.
The list of equipment used for testing engine performance, combustion characteristics, and exhaust emissions
is presented in Table 4.

3. CHARACTERISTICS OF INJECTION DELAY
Figure 4 shows the relationship between the engine power
and the engine speed for the test fuels (commercial diesel,
BD20) according to the amount of ultrasonic energy irradiation. Here, the chamber capacities were 550cc and 1100cc
at fuel injection timings of BTDC 11o and BTDC 16o.
For the BTDC 16o case, the engine power of commercial
diesel fuel was 0~3% higher than that for BD20. With
regard to total engine speed, UBD20 was 0~2% higher than


134

S. Y. IM, D. S. CHOI and J. I. RYU

diesel with a chamber capacity of 550cc, and 1~6% higher
than diesel with a chamber capacity of 1100cc. In other
words, the engine power was enhanced the most when
using irradiated fuel. For the BTDC 11o case, when the fuel
injection timing was delayed and ultrasonic irradiation was
applied, the engine power was enhanced or remained
almost the same within a 2~3% margin as compared to
BD20 over the entire range of speeds. The fuel injection
timing was delayed to ensure optimum fuel injection timing for NOx reduction. Although this delay has the effect of
reducing engine power, the fuel reformation effect due to
ultrasonic irradiation caused a compensating increase in

power. Hence, the overall engine power observed for
UBD20 was similar to that for BD20. In fact, the engine
power characteristics observed for BD20, UBD20 and
diesel were similar throughout the whole range of engine
speeds used here. With a fuel injection timing of BTDC
16o, the 1100cc chamber configuration showed a particular
tendency toward power increase. This implies that the
heating value of BD is lower than that of diesel, but that
fuel reformation, achieved by ultrasonic energy irradiation
lead to increased thermal efficiency.
Figure 5 shows the maximum combustion pressure for
cases with and without ultrasonic irradiation with fuel
injection timings of BTDC 16o and BTDC 11o. At BTDC
16o, the maximum combustion pressure of UBD20 increased by 5% using the 550cc chamber, and by approximately
2~6% with an 1100cc chamber.
In addition, the UBD20 maximum combustion pressure
was found to be larger in the 550cc case by approximately
2~3% over that for the 1100cc at BTDC 11o (the injection
timing was delayed compared to that used for BD20).
This is because the fuel reformed by the ultrasonic
irradiation promotes combustion via an improvement in
ignition. This is indicative of promoted evaporation of fuel
and reduced Sauter mean diameter (SMD) of droplets as
compared to the case of BD20.
As stated above, the optimum fuel injection timing for

Figure 4. Comparison of engine power under varying
engine speed at engine load 75%.

Figure 5. Comparison of Pmax under varying load at

engine speed of 2,000 rpm.
reduction of NOx in UBD20 is more delayed than that for
BD20, and thus the maximum combustion pressure is
lower. However, as presented in Figure 5, increased
duration of irradiation leads to somewhat of an increase in
the maximum combustion pressure. Accordingly, to ensure
optimum fuel injection timing for UBD20, the maximum
combustion pressure was reduced slightly by delaying the
fuel injection timing. However, providing enough duration
of ultrasonic energy irradiation contributed to an effective
fuel reformation, which improved combustion. The improvement seems to be due to the fact that the maximum
combustion pressure was similar to that for BD20 at BTDC
16o.
Figure 6 compares the combustion pressure, the heat
release rate and the mass burning rate of UBD20 with
1100cc chamber and BTDC 11o timing, diesel and BD20 at
the standard optimum timing, BTDC 16o. The engine load
was held at 75%, and the engine speed at 2,000 rpm.
When UBD20 was supplied to the engine, its combustion started slightly later than that for BD20, although it
was found to have similar characteristics. essentially, when
ultrasonic energy irradiation was performed on BD20 (used
at BTDC 11o), its ignition delay and the combustion
duration were shortened due to acceleration of combustion.
It is thought because the 10% oxygen content of BD
actively promotes combustion. It is also thought that the
ultrasonic energy reforms fuel quality physically and chemically leading to an increase in heating value and ignition
quality, while decreasing the viscosity, the surface tension
and the spray diameter (SMD) size.
Figure 7 shows the brake specific fuel consumption
(BSFC) of the engine according to the duration of ultrasonic energy irradiation (as implied by the chamber capacity) at fuel injection timings of BTDC 11o and BTDC 16o.

At BTDC 16o, the BSFC of commercial diesel fuel was
1~2% higher than that of BD20. Over the whole range of
engine speed, the BSFC of UBD20 was 1~2% lower with a
550cc chamber, and 1~3% lower with a 1100cc chamber.
Thus, the BSFC was most enhanced for the irradiated


COMBUSTION AND EMISSION CHARACTERISTICS OF BD20 REFORMED BY ULTRASONIC ENERGY

135

Figure 7. Comparison of BSFC under varying engine
speeds at an engine load of 75%.

Figure 8. Comparison of NOx under varying engine speeds
at an engine load of 75%.
Figure 6. Comparison of cylinder pressure, heat release rate
and mass burning rate at an engine speed of 2,000 rpm and
load of 75%.
fuel. With delayed timing, UBD20 showed a BSFC that
was better than the commercial diesel by 1~2% over the
whole range of engine speeds.
As mentioned above, it was anticipated that BSFC
would be deteriorated by delaying the fuel injection timing
in order to obtain optimal NOx reduction for UBD20. In
reality, for fuel reformed by sufficient ultrasonic energy
irradiation, the overall BSFC was enhanced. Furthermore,
BSFC gradually decreased with the amount of irradiation
of BD, but not when such a procedure was applied to
commercial diesel. It is believed that the oxygen contained

in BD promotes combustion, which then leads to enhanced
combustion efficiency.
Figure 8 shows NOx characteristics according to engine
speed and the duration of ultrasonic energy irradiation
given to the test fuels (commercial diesel, BD20, where
duration is implied by chamber capacity).

Figure 9. Comparison of soot concentration under varying
engine speeds at an engine load of 75%.
The engine load was 75% and the fuel injection timing
was set to BTDC 11o or BTDC 16o. At BTDC 16o, NOx
emission from commercial diesel fuel was 3~5% less than
that of BD20 over the whole range of engine speeds.
UBD20 NOx was 4~26% higher than BD20 with a 550cc
chamber, and 25~42% higher with an 1100cc chamber.


136

S. Y. IM, D. S. CHOI and J. I. RYU

As mentioned in the discussion of engine performance,
the ultrasonic energy irradiation reformed the fuel quality
physically and chemically, which promoted combustion.
This leads to a pressure increase in the combustion chamber, and a resultant increase in combustion temperature.
Higher temperature leads to increased NOx emissions as
compared to BD20.
NOx tends to generally increase with BD content,
because oxygen contained in BD promotes combustion,
thereby raising the combustion chamber temperature. When

the fuel was irradiated, NOx was remarkably increased as a
result of the reformed fuel.
Figure 9 shows the soot characteristics according to
engine speed depending on the duration of irradiation
(implied by chamber capacity). The engine load was 75%
with fuel injection timings of BTDC 11o and BTDC 16o.
The soot of commercial diesel fuel was 13~60% higher
than that from BD20 over the entire speed range. For
UBD20, it was 13~33% lower with a 550cc chamber and
40~67% lower an 1100cc chamber. Notably, when the fuel
injection timing was delayed to BTDC 11o, together with
the ultrasonic energy irradiation, the soot level was 20~
60% lower than that for BD20 over the entire speed range.
Although the combustion duration was not enough (due to
the delayed timing), the soot was reduced because of the
physical and chemical properties of irradiated fuel.

Figure 11. Comparison of engine power under varying
EGR Rates at engine load of 75% (2000 rpm).

Figure 10 shows the maximum combustion pressure according to the EGR rate and irradiation at timings of BTDC 16o
and BTDC11o for each of the test fuels (commercial diesel
fuel and BD20). For UBD20 at BTDC 16o, the decrease of
the maximum pressure with the EGR rate showed a
minimum of 1.4% at an EGR of 20%, while showing a
dramatic decrease of 5% at an EGR rate of over 30%.
It is thought that, at an EGR rate of over 30%, the
maximum combustion pressure quickly decreases because
the concentration of oxygen becomes insufficient for

combustion.
Figure 11 shows the relationship between the engine
power and the EGR rate for the test fuels (commercial
diesel, BD20) according to irradiation when the engine
speed is 2000 rpm and engine load is 75% with fuel
injection timings of BTDC 11o and BTDC 16o.
For BTDC 16o, the engine power of commercial diesel
fuel was 4~11% higher than that for BD20. For UBD20,
the engine power was 6~11% higher with an 1100cc chamber, and 4~5% higher with a 550cc chamber. For BTDC
11o, the engine power was increased by 6~8%. When the
EGR was applied, the engine power of BD20 tended to
decrease more than that of the commercial diesel fuel. With
increased irradiation, the engine power increased, similar
to the non-EGR case.
Figure 12 shows the relationship between the BSFC and
the EGR rate of test fuels (commercial diesel, BD20)
according to irradiation with an engine speed of 2000 rpm
and engine load of 75% and timings of BTDC 11o and
BTDC 16o.
At BTDC 16o, depending on the EGR rate, the BSFC of
commercial diesel fuel was larger by 2~4% than that of
BD20. The BSFC for UBD20 was about 3~5% lower than

Figure 10. Comparison of Pmax under varying EGR Rate
at engine speed of 2,000 rpm.

Figure 12. Comparison of BSFC under varying the EGR
Rates at engine load of 75% (2000 rpm).

4. CHARACTERISTICS RELATED TO

CHANGES OF EGR RATE


COMBUSTION AND EMISSION CHARACTERISTICS OF BD20 REFORMED BY ULTRASONIC ENERGY

Figure 13. Comparison of NOx under varying the EGR
Rates at engine load 75% (2000 rpm).
that for BD10 while using an 1100cc chamber, and about
2% lower while using a 550cc chamber. At BTDC 11~ the
BSFC for UBD20 was about 3% lower.
Varying the fuel injection timing at an EGR rate of 10%
does not make a significant difference in the BSFC under
identical engine speed and load. However, the range of
fluctuation tended to expand with increasing EGR rate.
Figure 13 shows the relationship of NOx according to
the EGR rate and irradiation for the test fuels (commercial
diesel, BD20) at an engine speed of 2000 rpm and load of
75% with timings of BTDC 11o and BTDC 16o.
When the fuel injection timing was set to BTDC16o, the
NOx from commercial diesel fuel was 2~8% lower than
that of BD20. UBD20 NOx was 5~8% higher than BD20
with a 550cc chamber, and 8~20% higher with an 1100cc
chamber.
At BTDC 11o, NOx was decreased by 5~8% compared
to BD20 over the whole range of engine speed. Against this
backdrop, it was reassuring that reduction of NOx was
greatly affected by the fuel injection timing.
NOx was dramatically decreased in accordance with the
increase of the EGR rate, especially in the 40% range
where the reduction was most notable. When EGR was

applied to the engine, some of the air inhaled into the
combustion cylinder was replaced with inert exhaust gas.
Because of this combustion temperature was lowered and
combustion was delayed, eventually leading to a remarkable reduction of NOx. NOx emissions were, therefore,
rapidly reduced with an increase of the EGR rate. In
addition, the NOx reduction rate was found to increase
with increasing engine load, but it could be predicted that
exhaust emissions would be remarkably increased due to
the decrease of oxygen.
Figure 14 shows the relationship between soot and the
EGR rate of the test fuels (commercial diesel, BD20)
according to irradiation at an engine speed of 2000 rpm,
and load of 75% with timings of BTDC 11o and BTDC 16o.

137

Figure 14. Comparison of soot under varying EGR Rates at
an engine load 75% (2000 rpm).
When the fuel injection timing was delayed together
with application of ultrasonic energy irradiation, soot was
decreased by 34~43% over the entire range when compared to the soot from commercial diesel fuel. Although
the combustion duration was insufficient due to the delayed
fuel injection timing, the soot was reduced because of
various factors related to the ultrasonic energy that promote
combustion. In other words, soot increased with an increase
in EGR rate, but the case of mixed diesel/BD emitted less
soot than with diesel alone.
When the EGR rate was more than 20%, soot was
increased over the whole range. The recirculated exhaust
gas reduced the amount of oxygen in the intake air that was

inhaled into the combustion chamber, disturbing the combustion process.
In addition, when the EGR was applied to BD20, soot
was reduced by a larger amount than was seen with diesel
fuel alone. This is because oxygen contained in BD promoted oxidation of the fuel. It is thought that oxygen
contained in the fuel promoted combustion by prompting
chemical reactions with hydrocarbons. Exceptional soot
reduction was achieved over the whole range of the EGR
rate due to the irradiation. This is because the ultrasonic
fuel reformation, together with biodiesel fuel elements,
contributed to combustion promotion.
Figures 15 and 16 show the correlation of the BSFC,
NOx, engine power and soot with cooled EGR rate at an
engine speed of 2000 rpm and load of 75% for BD20 at
BTDC 16o, and for ultrasonically reformed BD20 at BTDC
11o.
The BSFC and soot for both BD20 and the reformed fuel
(UBD20) increased in accordance with the increase of the
EGR rate, while engine power and NOx tended to gradually decrease.
Soot increased with the increase in the EGR rate, but it
was emitted relatively less with BD20 than with commercial diesel fuel. Soot was also emitted relatively less from
the irradiated fuel, UBD20.
However, with an EGR rate of more than 20%, the soot


138

S. Y. IM, D. S. CHOI and J. I. RYU

considerations simultaneously is in the range of 10~20%.
For UBD20 at BTDC 16o and EGR rate larger than 20%,

soot levels were similar or a little more than in the case of
the commercial diesel fuel or BD20. If the EGR rate was
less than 5%, NOx levels were similar to the case of
commercial diesel fuel.
Therefore, when UBD20 was used at the fuel injection
timing of BTDC 16o, the optimum EGR rate should be
15~20% to reduce both soot and NOx. Against this
backdrop, the optimum EGR rate should be considered in
the range that does not deteriorate characteristics of the
engine performance and minimizes exhaust emissions.
5. CONCLUSION

Figure 15. Comparison of NOx vs. BSFC under varying
the EGR Rates at engine load of 75% (2000 rpm).

Figure 16. Comparison of Power vs. Soot under varying
the EGR Rates at engine load of 75% (2000 rpm).
rapidly increased. This is because the recirculated exhaust
gas reduces the amount of oxygen sucked into the combustion chamber, resulting in insufficient oxygen for combustion.
In addition, NOx was dramatically reduced in accordance with the increase in EGR rate. As described previously, this feature of the EGR is due to lower combustion
temperature and slackened combustion speed.
Therefore, for UBD20 at BTDC 16o, the optimum EGR
rate to satisfy the BSFC, NOx, engine power and soot

We studied the use of fuel reformed by ultrasonic energy
irradiation in diesel engines in the context of optimizing
performance and emissions. In particular, we studied the
tuning of fuel injection timing and EGR rate to obtain the
following results:
(1) The maximum combustion pressure of the chamber

increased by up to 6% with an engine speed of 2,000
rpm and load of 75% upon irradiating BD20. When
UBD20 was used with injection further delayed by 5o,
the pressure increased by up to 3%.
(2) For the reduction of NOx from biodiesel fuel (an
oxygenated fuel with high NO), the optimum injection
timing of the fuel reformed by ultrasonic energy irradiation should be delayed compared to that of commonly
used diesel fuel. As regards BSFC, the results were
found to improve as the ultrasonic energy irradiation
duration became longer (the chamber capacity larger).
(3) NOx emission from UBD20 was 42% higher than
BD20 over the whole range of this experiment, while
soot was a maximum of 67% lower.
(4) When BD20 was used with a fuel injection timing of
BTDC 16o, the optimum EGR rate providing satisfactory BSFC and engine power was in the range of
10~20%.
(5) When BD20 was used with the fuel injection timing of
BTDC 16o, the optimum EGR rate for reduction of soot
and NOx was in the range of 15~20%.
When the fuel reformed by the ultrasonic energy irradiation
was applied to the diesel engine, the optimum EGR rate
was identified to be 15~20% to reduce NOx and to
promote the BSFC. Also, when the fuel injection timing
was delayed and the duration of the ultrasonic energy
irradiation was prolonged, the reduction effect increased.
For reformed fuel and delayed timing (UBD20 with BTDC
11o), engine power and BSFC were enhanced, while soot
and NOx were both reduced as compared to BD20 levels.
We find that in order to enhance engine performance and
reduce exhaust emissions it is essential to precisely control

the injection timing and the duration of ultrasonic energy
irradiation and to consider the optimum EGR rate.


COMBUSTION AND EMISSION CHARACTERISTICS OF BD20 REFORMED BY ULTRASONIC ENERGY

REFERENCES

139

Engineers , , 65−71.
KSRISO 2534 (2003). Load Vehicle - Engine Test Method Gross Shaft Power, Korean Agency for Technology and
Standards. Gyeonggi. Korea.
Lee, B. H. and Ryu, J. I. (2003). A study on relationship
between fuel characteristics and combustion characteristics of reformed diesel fuels by ultrasonic energy irradiation(II)-Relationship between chemical structure and
cetane number-. Trans. Korean Society of Automotive
Engineers , , 64−71.
Lue, Y. F., Yeh, Y. Y. and Wu, C. H. (2001). The emission
characteristics of a small D. I. diesel engine using biodiesel blended fuels. J. Environmental Science and Health
, , 845−859.
Oh, Y. T., Choi, S. H. and Kim, S. W. (2002). A study on
characteristics of rise bran oil as an alternative fuel in a
diesel engine(I), Trans. Korean Society of Automotive
Engineers , , 15−22.
Ryu, K. H. and Oh, Y. T. (2003). A study on the usability of
biodiesel fuel derived from rice bran oil as an alternative
fuel for IDI diesel engine. KSME Int. J. , , 310−317.
Song, Y. S. (2005). The Characteristics of Performance
13 6

Agarwal, D., Sinha, S. and Agarwal, A. K. (2006). Experimental investigation of control of NOx emissions in
biodiesel-fueled compression ignition engine. Renewable
Energy , , 2356−2369.
Bae, M. W., Ryu, C. S., Kim, H. M., Tsuchiya, K. and
Mochimaru, Y. (2002). Effect of intake mixture temperature on performance and exhaust emissions in diesel
engines with scrubber EGR system. Proc. (Int. Session)
31 14

68th Annual Conf. Japan Institution of Marine Engineering, 51−58.

Baik, D. S. (2006). Combined effects of BD20, low sulfur
diesel fuel and diesel oxidation catalyst in a HD diesel
engine. Int. J. Automotive Technology , , 653−658.
Choi, D. S. (1996). A Study on the Automization Charac7 6

teristics and the Engine Performance Characteristics of
the Ultrasonic-Energy-Added Diesel Fuel. Ph.D. Disser-

tation. Chungnam University. Daejeon. Korea.
Dale, E. (1988). Ultrasonics. MARCEL DEKKER. 2nd
edn. New York.
Ham, Y. Y. and Chun, K. M. (2002). Parametric study for
reducing NO and soot emissions in a DI diesel engine by
using engine cycle simulation. Trans. Korean Society of
Automotive Engineers , , 35−44.
Jeong, M. J., Cho, G. S. and Ryu, J. I. (1991). The performance imorovement of the diesel engine by the
ultrasonic vibrations. Trans. Korean Society of Automotive
10 5

11 1

36 5

10 2

17 2

and Exhaust in Diesel Engine of Ultrasonic Irradiation
Bio-diesel Blended Fuel. Ph.D. Dissertation. Chungnam

University. Daejeon. Korea.
Uchida, N. (1993). Combined effects of EGR and supercharging on diesel combustion and emissions. SAE Paper
No. 930601.


International Journal of Automotive Technology, Vol. 10, No. 2, pp. 141−149 (2009)

DOI 10.1007/s12239−009−0017−1

Copyright © 2009 KSAE
1229−9138/2009/045−02

STRATEGIES FOR IMPROVING THE MODE TRANSITION IN A
SEQUENTIAL PARALLEL TURBOCHARGED AUTOMOTIVE DIESEL
ENGINE
J. GALINDO, H. CLIMENT, C. GUARDIOLA and J. DOMÉNECH
*

CMT Motores Térmicos, Universidad Politécnica de Valencia, Camino de Vera s/n, Valencia 46022, Spain
(Received 1 August 2007; Revised 2 September 2008)

ABSTRACT−Parallel sequential turbocharging systems are able to operate in different modes, which are defined according

to the turbochargers that simultaneously boost the engine, and are controlled by means of specific valves. In order to cover the
full engine operating range, a smooth transition between turbocharging operating modes must be ensured. However, important
disturbances affect both boost and exhaust pressure when shifting the operation mode, thus causing non-negligible torque
oscillations. This paper presents different methods for smoothing such undesirable effects during mode transition. Strategies
covering optimal synchronization of the control valves, control of the valves’ position, and correction of the injected fuel
during the transition are analysed. A fully instrumented passenger car engine is used for illustrating the different torque
smoothing methods, and experimental results for transitions during both steady operation and engine accelerations are shown.

KEY WORDS : Sequential turbocharging, Diesel engine, Transition, Boost control

NOMEMCLATURE

The problem of the turbo-lag is of great importance
when an acceleration is demanded and exhaust gas energy
availability is low, which occurs at low engine speed and
load. In this case, the driver’s torque demand produces a
step increase in the fuel supplied; however, the required air
to ensure a correct combustion process is not supplied by
the compressor as fast as it should be. This situation results
in an increase in the fuel-to-air ratio, and hence, characteristic ‘black smoke’ appears. In order to limit smoke
emissions, current diesel engines limit the injected fuel
during the accelerations on the basis of minimum air-tofuel ratio criteria, which also negatively affects engine performance during load transients.
Current demands on diesel engine performance require
the solving of problems related to the turbo-lag. Downsizing is a general trend that leads to engines that have the
same power output with lower displacement. The downsized engine has lower emissions and consumption but the
low end torque is worsened. As a result, some new technologies have been developed, such as the variable geometry
turbines (VGT) (Watson and Janota 1982; Arnold et al.,

2002), which are able to adapt the effective area of the
stator. Another method to improve the performance in
transient conditions is to employ a smaller turbocharger
since: a) the turbine is better designed for working with low
exhaust gases mass flow, and b) it accelerates faster because
of its lower inertia. Both methods allow faster transient
engine response. However, small size turbochargers limit
the maximum engine speed and load. In order to overcome
this problem, sequential operation of turbochargers arrang-

1T
: engine operation mode with a single turbocharger
2T
: engine operation mode with two turbochargers
C1
: compressor of TC1
C2
: compressor of TC2
TC1 : turbocharger number 1
TC2 : turbocharger number 2
T1
: turbine of TC1
T2
: turbine of TC2
VC2 : control valve of C2
VGT : variable geometry turbine
Vrecirc: valve of the recirculation circuit
VT2 : control valve of T2
WG : waste-gate

1. INTRODUCTION
Turbocharging is a widely used method that improves
internal combustion engine performance. However, turbocharged engines cause a delay during load transients (Watson
and Janota, 1982).
The causes of the delay are mechanical, thermo- and
fluid dynamic. The first involves the inertia of the turbocharger, the second includes the processes of mass and
energy transfer between the exhaust valve and the turbine,
and between the compressor and the cylinders (Benajes et
al., 2002; Payri et al., 2002).
*Corresponding author. e-mail:
141


142

J. GALINDO, H. CLIMENT, C. GUARDIOLA and J. DOMÉNECH

ed in parallel can be used.
Sequential parallel turbocharging systems are able to
improve diesel engine transient response, since low inertia
turbochargers allow faster accelerations (Arnold
.,
2002). This turbocharging system has been used in the past
for marine Diesel (Ren ., 1998; Benvenuto and Campora,
2002; Kench and Klotz, 2002) and gasoline automotive
engines (Tashima
., 1991, 1998). The present work is
focused on the development of a passenger-car diesel
engine. In contrast to marine applications, engine speed in
land vehicles is subject to more variation, and thus, the

dynamic torque response is a key factor. In contrast to
conventional turbocharging systems, parallel sequential
systems are more complex and include non-standard elements. Additional details on the advantages and disadvantages of this kind of turbocharging system compared to
serial turbocharging (Serrano
, 2008) and others is
presented by Portalier . (2006) and Galindo . (2009).
Most research on diesel engines has been centered on
marine applications where the engine running conditions
are fairly constant; very few investigations have analyzed
the load transient that appears during the sequence from
one to two turbochargers in operation and vice versa. This
process will be referred to hereafter as a transition.
This paper presents experimental work related to a
sequential parallel double turbocharged diesel engine. Since
the engine is designed for automotive applications, important variations on both engine speed and load are expected.
Thus, the main objective of the paper is to illustrate the
main phenomena causing torque oscillations during the
transition, and to present several strategies in order to
reduce them.
et

al

et al

et al

et al.

et al

et al

2. TURBOCHARGING SYSTEM DESCRIPTION
In this section the sequential parallel turbocharging system
is described. Parallel sequential turbocharging systems comprise two independent turbochargers (TC1 and TC2). One
of them (TC1) is always in operation, while the second one
(TC2) operates only in the high speed region of the engine
map.
TC1 is a basic charger and operates throughout the
complete operational range. However, it is smaller than the
equivalent turbocharger used in a conventional system, and
is therefore lighter and has lower inertia. These characteristics allow TC1 to accelerate faster in lower ranges of
speed and load where TC2 is not operating. In addition, and
due to the smaller compressor size, higher boost pressures
are admissible at low engine speeds, as the surge limit of
TC1’s compressor is less restrictive than those of conventional compressors.
TC2 is even smaller than TC1 and operates at high
engine speed and load range, where both turbochargers are
working simultaneously. The activation of TC2 is possible
by means of auxiliary air valves in the engine.
The system is completed by additional valves that allow

managing the transition between the two basic operation
modes (1T and 2T), and controlling the boost pressure. A
schematic of the system is shown in Figure 1, where the top
plot shows valve positions in 1T operation mode and the
bottom plot shows valve positions in 2T operation mode. In
both cases a waste-gate (WG) is used for controlling boost
pressure. A short description of the functionality of each

valve is provided below:
(1) A waste-gate (WG) is used for boost pressure control.
Although variable geometry turbines (VGT) could also
be used, this possibility was not considered due to
overall cost restrictions of the turbocharging system.
WG-based boost control in the parallel sequential turbocharging system is similar to conventional turbocharged engines. Hence, only WG influence on the transition
will be highlighted.
(2) A control valve at turbine 2 (VT2) is used to feed
turbine 2 (T2) when necessary. This valve activates the
transition between 1T and 2T modes. In addition, small
openings of the VT2 are used to prepare the transition
from 1T mode to 2T mode.
(3) A valve placed downstream of compressor 2 (VC2) is
used to prevent internal air leaks from the intake manifold to C2 when this compressor is not operating. Both
VC2 and VT2 determine the turbocharging system
configuration. The engine runs with two turbochargers
in parallel if both valves are open.
(4) As previously mentioned, TC2 must never be completely stopped: lubrication and a minimum thermal level
must be ensured in order to avoid breaking TC2 when
shifting from 1T to 2T operation. A minimum flow
through C2 is ensured by slightly opening VT2, and
using a recirculation circuit with a control valve (Vrecirc).

Figure 1. System schematic and valve configuration for the
two main operation modes (top: 1T; bottom: 2T).


STRATEGIES FOR IMPROVING THE MODE TRANSITION IN A SEQUENTIAL PARALLEL TURBOCHARGED 143
This valve must be open in 1T mode and closed in 2T
mode to avoid air leaks in the intake manifold. Beyond

ensuring lubrication and minimum temperature, the
recirculation circuit helps prepare the transition form
1T to 2T via the progressive acceleration of TC2.
(5) Finally, the engine is also fitted with an EGR circuit
with an EGR control valve. As this system only opens
in the 1T operation zone, no real interaction occurs with
the parallel sequential turbocharging system. For clarity
reasons, all diagrams presented have been simplified
by excluding the EGR system.
2.1. Operating Range Limits
Important limitations in the operation ranges of each of the
turbocharging modes (1T and 2T) exist, hence it is not
possible to run the engine in a specific turbocharger configuration for the entire operation range. A short theoretical
approach based on engine volumetric efficiency is now
discussed.
Volumetric efficiency is defined as follows:
m· a
η V = -----------------(1)
zV ρ niD i

where m· a is the engine air mass flow, is the number of
cylinders, D is the cylinder displacement, ρi is the intake
manifold air density, is the engine speed, and equals 0.5
in a four-stroke engine. Equation (1) can be conveniently
transformed into:
z

V

n

i

pi - = ---------------------------RTi - m· a
-------

(2)

patm p atm zV D ni ηV

where the first term is equal to the compression ratio in the
compressor if the pressure losses in the air filter and in the
charge air cooler are neglected.
Equation (2) expresses the linear relationship between
the air mass flow and the compression ratio for a constant
engine speed if it is assumed that the temperature at the
charge air cooler outlet and the volumetric efficiency remain
constant. The equation is valid when no EGR is performed
and in steady operation, where there are no mass storage
effects (i.e. variation of manifold pressure).
Equation (2) is used to plot three engine characteristic
lines for three different engine speeds in a typical compressor chart and is shown in Figure 2. At low engine
speed, the compression ratio is limited by either the compressor surge zone (Galindo
, 2006) or the exhaust
energy availability. At high engine speed, the compression
ratio, and hence the intake pressure, is limited by the
compressor over-speed zone.
The design steps in conventional turbocharging systems
consists in selecting a single turbocompressor to cover the
full engine operation range. Big turbochargers are needed

to cover the high speed region, and surge is usually a
limiting factor in the low speed range.
A smaller turbocharger can be selected in sequential
parallel systems. This turbocharger is better adapted to the

Figure 2. Engine operation points plotted on a compressor
chart.
low speed region, but over-speed is a problem for high
boost pressures and engine speeds.
When the two turbochargers of the sequential system
operate simultaneously in the 2T mode, equation (2)
becomes:

pi
RT i
·
·
------p - = ---------------------------p zV ni η- ( m a C + ma C
atm

atm

D

V

,

1

,

2

)

(3)

This means that the equivalent operation point is shifted
to the left on the compressor chart as the total flow is
supplied by the two turbochargers. In this case, the overspeed limit is not reached, but surges are possible at low
engine speeds and high boost pressures. In addition, low
exhaust gas energy availability in this region makes 2T
operation impossible.
These two limits are represented on the engine map in
Figure 3. In addition to these two restrictions, some mild
criteria for the operating mode selection can be added: at
low engine load and speed, the engine transient response is
faster if running at 1T mode, so the 2T mode is not
recommended, and there is also a zone at medium load and
high engine speed that corresponds to highway driving
conditions, where operation in 1T mode would result in

et al.

Figure 3. Engine map illustrating operation mode restrictions.


144

J. GALINDO, H. CLIMENT, C. GUARDIOLA and J. DOMÉNECH

Figure 4. Scheme of the experimental setup. Main information fluxes are marked as gray arrows, while main
control actions are marked as black arrows.
higher fuel consumption.

3. EXPERIMENTAL SET-UP
The engine tested was a 4-stroke, 2.2 liter, four cylinder
diesel engine with two turbochargers arranged in parallel.
The test bench configuration is shown in Figure 4. The
engine is controlled by an ECU that is externally calibrated
by means of an ETAS ES1000 system. In addition, a signal
bypass is used to include the sequential parallel turbocharging management system: an independent module for
controlling the four listed valves was programmed and
integrated into the system using ASCET software and an
ES1130 card. A bypass to the ECU fueling strategy was
also available, which enabled fuel injected quantity corrections during the transition.
An eddy current dynamometer was used to load the
engine and perform transient tests. ECU and ES1130 card
variables were acquired using INCA software. In addition,
intake and exhaust manifold pressure and temperature,
engine speed and torque, and turbocharger speeds were
acquired with a PUMA system. Instantaneous opacity measurements were performed with an AVL439 opacimeter.

4. TRANSITION AND TORQUE OSCILLATIONS
Although the sequential parallel turbocharging system in
diesel engines presents several advantages as previously
described, there is a significant drawback: the transition
between both operation modes becomes critical. Torque
oscillations during the transition are of little importance in

marine applications, where these types of engines have
been largely employed, since the engines run in nearly
steady conditions with few transitions. However, the problem is notable in automotive engines due to the higher
frequency of transitions (Galindo
, 2009) and the
potential negative effects of torque variations on the vehicle
driveability.
et al.

Several dynamic processes occur when the transition
takes place, including variations in both intake and exhaust
manifold pressure, acceleration of the turbochargers, and
thermal transients. All these phenomena interact with the
combustion process and the engine pumping work, thus
resulting in a non-steady evolution of engine torque during
the transition.
Engine efficiency can vary depending on the operation
mode and during the transition. The difference in pumping
losses and the variation in combustion efficiency, which
strongly depends on the in-cylinder pressure at the end of
the intake process, are primary factors affecting engine
efficiency. Thus, different steady torque values are obtained when operating with the same fuel mass in 1T and 2T
operation modes, and oscillations appear in the torque
profile.
In addition, the sensed air mass flow during the transition can undergo important oscillations, which can trigger
the injection fuel limitation process to control smoke. In
order to avoid the interaction of this strategy with torque
oscillation due exclusively to the transition process, the
smoke control strategy was deactivated during the transition. There were two reasons underlying this decision:
(1) The smoke control strategy is based on the measurement of a hot wire anemometer placed just downstream

of the air filter. During the transition from 2T to 1T
operation, the pressure of the portion of intake manifold between C2 and VC2 decreases. This rapid variation in the mass stored in the manifold causes an
evident variation in the sensed air mass flow, while the
admitted air mass flow is less affected. Consequently,
the smoke control methodology overestimates the
required fuelling restriction and causes an additional
torque drop.
(2) The opacity level during the transition becomes unimportant when compared with the beginning of a load
transient. Smoke emission during the transition is a
short-term smoke peak that could presumably be dealt
with via other emission control techniques like particulate filters. This decision can be made during the
final calibration.
Hence, none of the experimental results in the present
paper consider the fuelling limitation due to the smoke
control strategy. However, opacity measurements are
provided in order to check the transition impact on smoke
emissions.
Additionally, valve management during the transition is
critical, and slight variations in the timing of the closing
and opening process of different involved valves can result
in magnification of the torque oscillation. Controlling this
becomes a critical issue for this type of engine (Cantemir,
2001).
To achieve optimal valve synchronisation requires an
iterative optimization process. In this case a pilot study was
performed by means of simulation (Galindo
, 2007),
with experimental tests used later for fine tuning and
et al.

STRATEGIES FOR IMPROVING THE MODE TRANSITION IN A SEQUENTIAL PARALLEL TURBOCHARGED 145
validation. The following paragraphs detail the transition
from1T to 2T and from 2T to 1T.
4.1. 1T to 2T Transition
When passing from 1T to 2T, C2 must be accelerated and
the part of the intake system between C2 and VC2 must be
pressurized. Hence, additional energy is needed to perform
these actions.
A first step when performing a transition is to close the
waste gate to capture all available exhaust gas energy. The
next step is opening VT2 to enable exhaust gases to flow
through turbine 2 and accelerate the turbocharger. All these
steps are done with VC2 closed and Vrecirc open in order
to avoid surge and reverse flow through the compressor.
Once the C2 outlet pressure is close to the intake manifold
pressure, VC2 opens and Vrecirc closes.
The top plots in Figure 5 depict the torque evolution
during the transition from mode 1T to 2T; the bottom plots
illustrate the valve opening sequence applied. The lefthand plots show the transition at 3250 rpm and 10 mg of
fuel per stroke, and the right hand plots show a similar
transition at the same speed and 30 mg of fuel per stroke. It
is notable that the engine torque under steady conditions in
1T and 2T modes is not the same, since these engine tests
were performed with constant injected fuel.
To illustrate the sensibility of the system to WG actuation timing, two tests are shown in the figure: one using
optimal timing, and the other increasing the WG closing by
400 ms. When the WG closes too early there is an increase
in the exhaust manifold pressure which increases the
pumping losses and reduces the engine torque (Luján

.,
2007). It is possible to achieve a synchronization between
the WG and the VT2 in order to obtain a smooth torque
evolution. This synchronization is more important at low
rather than at high engine loads. The peak value of the
torque deterioration is 2 Nm for the low load case and 4
Nm for the high load case, but the relative variation is 8%
et al

Figure 6. 2T to 1T mode transitions with different valve
timing. Two transitions are shown per plot: optimal (dashed)
and delaying the WG opening 200 ms (solid). See Figure 5
caption for details.
and 1.5%, respectively.
Vrecirc and VC2 timings are not shown as they were not
important to the transition: recirculation is used for preserving C2 from surge, and has no direct effect on the
torque evolution. On the other hand, VC2 pneumatic
actuation system does not enough force to open the valve in
the case of reverse pressure gradient. Thus, once the
control action of VC2 is set to ‘open’, VC2 acts as a semipassive valve, which automatically opens when pressure at
C2 outlet is about equal to the intake manifold pressure.
4.2. 2T to 1T Transition
Although less critical, valve synchronization during the
transition from the 2T operating mode to 1T is also
relevant. In this case, the WG must open at the same time
VT2 closes. A delay in opening the waste gate generates a
higher torque drop, due to the increase in backpressure and
therefore, the pumping losses increase.
Figure 6 illustrates transitions from the 2T to 1T mode at
3250 rpm and 10 and 30 mg/stroke. In this case, the delay

between the two tests performed at the same engine
condition is 200 ms. However, the sensitivity to WG timing
is not as high as in the 1T to 2T transition.
It is noticeable that the torque dynamics during the 2T to
1T transition are faster than in the 1T to 2T case, and a
steady value is reached in less than 1 s.

5. TRANSITION IMPROVEMENT STRATEGIES

Figure 5. 1T to 2T mode transitions with different valve
timing. Two transitions are shown per plot: optimal (dashed)
and closing the WG 400 ms early (solid). Left: 3250 rpm
and 10 mg/stroke. Right: 3250 and 30 mg/stroke. Top: torque
evolution. Bottom: valves’ position.

Basic transitions shown in the previous section have, even
in the optimal synchronization case, two undesirable effects
on the torque: the first is that the different modes have
important differences in steady operation torque, and the
second is oscillations in torque, especially in the 1T to 2T
transition.
This section is focused on different strategies for smoothing the torque evolution. For steady operation torque difference, it is necessary to correct the injected fuel mass


146

J. GALINDO, H. CLIMENT, C. GUARDIOLA and J. DOMÉNECH

depending on the operation mode. For oscillations, several
possibilities are explored to improving the short-time torque

oscillations during the transition: adding corrections to the
instantaneous injected fuel mass, and using other strategies
in the VT2 opening profile. The latter of these involves
slightly opening VT2 for transition preparation (it will be
referred as VT2 pre-lift) and slow actuation of VT2.
As the fuel-to-air ratio is modified, opacity measurements are provided in order to check the effect of the
strategy on the overall emissions.
Finally, with the aim of covering different driving conditions, transitions have been tested both in steady operation
(with fixed engine speed and injected fuel) and in engine
accelerations simulating road conditions.
5.1. Injected Fuel Mass Corrections
The injected fuel mass correction strategy consists in modifying the fuel injected during the transition. Therefore, the
methodology of this strategy is to counteract the positive
torque oscillation peaks with negative pulses of fuel and
vice versa.
The system is able to modify the nominal fuel by adding
3 fuel pulses that are completely independent and configurable as illustrated in Figure 7. In addition, a steady correction value is added to compensate for the difference in
torque between 1T and 2T, which can be positive or
negative depending on the operating conditions.
5.1.1. Engine steady conditions
Several amplitude and timing configurations for the injected fuel pulses were tested before achieving reasonable
performance. In all cases, high accuracy was needed in the
correction timing, since injecting a positive fuel mass pulse
during a positive torque peak would result in an even more
problematic situation.
Figure 8 illustrates the results of a transition with and
without the fuel correction strategy performed in steady
conditions, 2750 rpm and 200 Nm. The transition was from
the 1T to 2T mode. The top left graph shows that torque
evolution and an oscillation of 15 Nm (7.5%) is observed if

no fuel correction is applied. The fuel correction evolution
during the transition is plotted in the bottom left graph. The

Figure 7. Fuel correction profile. Labels indicate configurable parameters for the initial fuel pulse.

Figure 8. 1T to 2T mode transitions at 2750 rpm and 200
Nm with (solid) and without (dashed) the fuel correction
strategy. Top left: torque. Top right: fuel-to-air relative ratio.
Bottom left: injected fuel. Bottom right: opacity.
fuel-to-air ratio and the exhaust gas opacity are shown in
the right-hand plots.
In the previous case, since the fuel-to-air ratio is low
enough, the impact of this strategy is very small. However,
when the engine approaches full load, the situation becomes
critical, since the engine operates near stochiometric conditions.
Figure 9 illustrates a transition performed at full load,
2500 rpm and 365 Nm. A variation of 40 Nm (11%) is
observed without the fuel correction strategy. The fuel
evolution that smoothes the engine torque is also represented, where 15 mg/stroke additional fuel is necessary to
counterbalance the engine torque reduction. However, this
strategy has serious consequences at high load, where an
opacity of 90% and a fuel-to-air ratio higher than 1.1
appear because of the increase in fuel consumption.
The 2T to 1T mode transitions are again depicted at
partial load (2750 rpm and 200 Nm) in Figure 10, and at
full load (2500 rpm and 365 Nm) in Figure 11. The transition at partial load leads to an engine torque deterioration
of 15 Nm, which is similar to the value obtained for the

Figure 9. 1T to 2T mode transitions at 2500 rpm and full
load with (solid) and without (dashed) the fuel correction

strategy. See Figure 8 caption for details.


STRATEGIES FOR IMPROVING THE MODE TRANSITION IN A SEQUENTIAL PARALLEL TURBOCHARGED 147

Figure 10. 2T to 1T mode transitions at 2750 rpm and 200
Nm with (solid) and without (dashed) the fuel correction
strategy. See Figure 8 caption for details.

Figure 11. 2T to 1T mode transitions at 2500 rpm and full
load with (solid) and without (dashed) the fuel correction
strategy. See Figure 8 caption for details.
opposite mode transition. Moreover, this backwards transition at full load results in a torque drop of 20 Nm, which is
comparably lower than what was observed for the 1T to 2T
mode transition.
Since these transitions do not present strong oscillations
in the engine torque, the required fuel corrections in order
to smooth the engine operation achieve maximum values
of 3 and 5 mg/stroke at partial and full loads, respectively.
Therefore, when using the correction fuel strategy, the fuelto-air ratio and exhaust gas opacity give slightly higher,
though still acceptable, values.
5.1.2. Engine acceleration conditions
A transition during engine acceleration is quite different
from the steady operation equivalent: since the WG is
closed during the acceleration process, the engine torque
evolution only slightly increases when VT2 is activated
due to the reduction of the exhaust manifold pressure and
hence the pumping losses. Although the torque oscillations
in these conditions are not as relevant as in steady transitions, the fuel correction strategy has been also tested in
order to completely eliminate the torque peak.

Figure 12. 1T to 2T mode transition during an engine
acceleration. Results are shown with (solid) and without
(dashed) the fuel correction strategy. Top left: torque. Top
right: detail of torque evolution during the transition. Bottom
left: injected fuel. Bottom right: opacity during the transition.
Figure 12 illustrates a transition under an engine acceleration. The upper graph shows the torque curves with and
without the fuel correction strategy. The graph on the right
shows in detail the torque peak generated during the transition, where 10 Nm of torque oscillation (4%) is detected
when no fuel correction strategy is applied. Since the goal
is to reduce the engine torque, only reductions in the
injected fuel pulses are applied, as depicted in the left-hand
bottom plot. This prevents negative consequences from
arising due to increases in exhaust gas opacity.
5.2. VT2 Pre-lift Actuation
The strategy is intended to improve 1T to 2T transition by
means of an initial acceleration of TC2: in a first step, VT2
does not open completely, but just enough to allow the
passage of the exhaust gases through T2. The purpose of
this action is to accelerate TC2 and pressurize the ducts
downstream of C2 and upstream of VC2. Once TC2 is
launched, the second step is to completely open VT2. The
VT2 pre-lift actuation was tested as a fuel correction strategy
at both steady and transient operation.
5.2.1. Engine steady conditions
Figure 13 right-hand plot replicates the conditions of
Figure 8 (2750 rpm and 200 Nm), and the left-hand plot
those of Figure 9 (2500 rpm and 365 Nm, full load). By
using VT2 pre-lift in steady conditions, engine torque
oscillations are reduced, although slight fluctuations still

appear. They correspond precisely to the two-step opening
profile of VT2, as shown in the bottom plots.
At partial load, the maximum oscillation is reduced from
16 Nm at normal VT2 actuation, to 8 Nm (from 8% to 4%).
At full load the reduction is significant, from 40 Nm to 10
Nm (or from 11% to 3%). The duration of the VT2 pre-lift
period must increase with engine load to properly accelerate TC2.


148

J. GALINDO, H. CLIMENT, C. GUARDIOLA and J. DOMÉNECH

Figure 13. Effect of VT2 pre-lift actuation strategy (solid)
during a 1T to 2T transition at 2500 rpm and 365 Nm (left)
and at 2750 and 200 Nm (right). Dashed lines show
nominal transitions. Top: torque. Bottom: VT2 position.

Figure 15. Effect of the VT2 slow actuation strategy (solid)
during a 1T to 2T (left), and a 2T to 1T (right) transition.
Dashed lines show nominal transitions. Top: torque. Bottom:
VT2 position.

5.2.2. Engine acceleration conditions
As previously noted, engine torque oscillations in full-load
accelerations are not as important as they are in steady
operation. Nevertheless, some improvement is possible with
pre-lift actuation. Figure14 replicates the conditions of the
results of Figure 12. Engine torque differences at the early
stages of the acceleration process are due to slight modifications in the smoke limiting strategy between both tests.

The figure shows the VT2 strategy in the left-bottom
plot, which completely smoothes torque evolution during
the transition.

A restriction in the vacuum system of the pneumatic
actuator was used to modify the valve response time. Thus,
the effective diameter of the duct that links the actuator
chamber of the VT2 valve with the rest of the vacuum
control circuit was reduced, increasing the response time.
Results for both steady and transient engine conditions
are shown.

5.3. VT2 Slow Actuation
The last strategy evaluated to reduce the torque oscillations
during the transition consists in modifying the VT2 actuation speed.
For 1T to 2T transitions, the main purpose of this
strategy is to avoid the sudden aperture of VT2, which
causes an abrupt drop of exhaust manifold pressure and
increases engine torque.

Figure 14. 1T to 2T mode transition during an engine
acceleration. Results are shown using the VT2 pre-lift
strategy (solid) and using nominal valve actuation (dashed).
See Figure 12 caption for details.

5.3.1. Steady engine conditions
The left-hand plots in Figure 15 illustrate the transition
from 1T to 2T mode at high engine load, 2500 rpm and 300
Nm. The standard VT2 aperture takes about 200 ms, while
the duration exceeded 1 s using the slow VT2 actuation.

The top plot shows a reduction in the amplitude of the
torque oscillation from 30 Nm to 20 Nm (from 10% to
6.7%), whereas the duration of the oscillation is doubled.
The transition from 2T to 1T mode at 2500 rpm and 300
Nm is represented in the right-hand plots of Figure 15. In
this case, no significant improvements are obtained. Since
the VT2 is set in series with T2, VT2 effective area has
important effects on the flow when it is comparatively
lower than the effective area of the turbine. For the particular design considered, this only occurs when the valve is
nearly closed. This justifies the effect on the 1T to 2T mode
transition, since the valve opens slowly at the beginning of
the transition and has little influence on the 2T to 1T mode
transition, where it closes at the later stages.
5.3.2. Engine acceleration conditions
The VT2 slow actuation was also examined in engine
acceleration conditions and the results are illustrated in the
Figure 16.
The results show that the small torque peak is eliminated. However, the exhaust gas energy bypassed through
VT2 affects the boost pressure and also the combustion
efficiency, yielding a loss in torque during the transition
process. Thus, this strategy is not suitable for air-deficient
full-load accelerations.


STRATEGIES FOR IMPROVING THE MODE TRANSITION IN A SEQUENTIAL PARALLEL TURBOCHARGED 149
Automobile Engineering, , 431–441.
Benvenuto, G. and Campora, U. (2002). Dynamic simulation of a high-performance sequentially turbocharged
marine diesel engine. Int. J. Engine Research , , 115−
125.
Cantemir, C.-G. (2001). Twin turbo strategy operation. SAE

Paper No. 2001-01-0666.
Galindo, J., Serrano, J. R., Guardiola, C. and Cervello, C.
(2006). Surge limit definition in a specific test bench for
the characterization of automotive turbochargers. Experimental Thermal and Fluid Science, , 449–462.
Galindo, J., Luján, J. M., Climent, H. and Guardiola, C.
(2007). Turbocharging system design of a sequentially
turbocharged diesel engine by means of a wave action
model. SAE Paper No. 2007-01-1564.
Galindo, J., Climent, H., Guardiola, C., Tiseira, A. and
Portalier, J. (2009). Assessment of a sequentially turbocharged diesel engine on real-life driving cycles. Int. J.
Vehicle Design, .
Kench, J. M. and Klotz, H. (2002). Model-based sequential
turbocharging optimization for series 8000 M70/M90
engines. SAE Paper No. 2002-01-0378.
Luján, J. M., Climent, H., Guardiola, C. and García-Ortiz,
J. V. (2007). A comparison of different algorithms for
boost pressure control in a heavy-duty turbocharged
diesel engine. Prc. Instn. Mech. Engrs. Part D, J. Automobile Engineering, , 629–640.
Payri, F., Benajes, J., Galindo, J. and Serrano, J. R. (2002).
Modelling of turbocharged diesel engines in transient
operation. Part 2: Wave action models for calculating the
transient operation in a high speed direct injection engine.
216

3

3

30

Figure 16. 1T to 2T mode transition during an engine
acceleration. Results are shown using VT2 slow actuation
strategy (solid) and with nominal valve actuation (dashed).
See Figure 12 caption for details.

49

6. CONCLUSIONS

Transition phenomena and optimized valve management of
a sequential parallel turbocharged engine have been shown.
In addition, different strategies for improving the mode
transition were also presented and experimentally compared. The main conclusions of this comparative study are:
(1) A fuel correction strategy allows the complete elimination of torque oscillations during the transition for
both steady and transient cases. However high levels of
exhaust gas opacity are obtained when approaching full
load conditions, thus limiting the application range of
this strategy. The fuel correction strategy also compensates for the steady state torque difference between the
two operation modes.
(2) The VT2 pre-lift strategy smoothes the torque oscillation in both steady and transition engine conditions
without significant drawbacks. However, small oscillations are still present.
(3) The VT2 slow actuation strategy shows good results in
steady conditions, however it is not recommendable
under transient engine conditions.
These strategies for smoothing the mode transition can
be combined. In the final version of the engine, the VT2
pre-lift strategy was combined with the fuel correction
strategy to provide smooth transitions in both steady and
transient conditions (Galindo et al., 2009).

221

Prc. Instn. Mech. Engrs. Part D, J. Automobile Engineering, 216, 479–493.

Portalier, J., Blanc, J. C., Garnier, F., Hoffmann, N., Schorn,
N., Kindl, H., Galindo, J., Jeckel, D., Uhl, P. and Laissus,
J.-J. (2006). Twin turbo boosting system design for the
new generation of PSA 2,2 liter HDI diesel engines.
Proc. Thiesel Conf. 2006, 589−607.
Ren, Z., Campbell, T. and Yang, J. (1998). Theoretical and
experimental study on the performance of a sequentially
turbocharged diesel engine. 6th Int. Conf. Turbocharging
and Air Management Systems, C554/010/98.
Serrano, J. R., Arnau, F. J., Dolz, V., Tiseira, A., Lejeune,
M. and Auffret, N. (2008). Analysis of the capabilities of
a two-stage turbocharging system to fulfil the us2007
anti-pollution directive for heavy duty diesel engines.
Int. J. Automotive Technology , , 277–288.
Tashima, S., Tadokoro, T., Okimoto, H. and Niwa, Y. (1991).
Development of sequential twin turbo system for rotary
engine. SAE Paper No. 910624.
Tashima, S., Okimoto, H., Fujimoto, Y. and Nakao, M.
(1998). Sequential twin turbocharged rotary engine of
the latest RX-7. SAE Paper No. 941030.
Watson, N. and Janota, M. S. (1982). Turbocharging the
Internal Combustion Engine. MacMillan Publishers Ltd..
Houndmills.
9 3

REFERENCES

Arnold, S., Groskreutz, M., Shahed, S. M. and Slupski, K.
(2002). Advanced variable geometry turbocharger for
diesel engine applications. SAE Paper No. 2002-01-0161.
Benajes, J., Luján, J. M., Bermúdez, V. and Serrano, J. R.
(2002). Modelling of turbocharged diesel engines in
transient operation Part 1: Insight into the relevant
physical phenomena. Prc. Instn. Mech. Engrs. Part D, J.


Copyright © 2009 KSAE
1229−9138/2009/045−03

International Journal of Automotive Technology, Vol. 10, No. 2, pp. 151−160 (2009)

DOI 10.1007/s12239−009−0018−0

EXPERIMENTAL STUDIES ON LOW PRESSURE SEMI-DIRECT FUEL
INJECTION IN A TWO STROKE SPARK IGNITION ENGINE
M. LOGANATHAN and A. RAMESH
1)

1)

2)*

Department of Mechanical Engineering, Annamalai University, Annamalai Nagar 608002, India
Internal Combustion Engines Laboratory, Department of Mechanical Engineering,
Indian Institute of Technology Madras, Chennai 600036, India
2)

(Received 13 August 2007; Revised 6 October 2008)

ABSTRACT−In this work a two-stroke scooter engine was modified to work with semi-direct injection of gasoline at a

pressure of 8 bar from an injector in the cylinder barrel pointed toward the cylinder head. The influence of injection timing,
injection pressure, spark plug location and air-fuel ratio, on performance, emissions and combustion characteristics has been
investigated. In addition, a comparison has been made with manifold injection of gasoline on the same engine at a given speed
and various outputs. A significant reduction in HC emissions and fuel consumption with no adverse effects on NOx emissions
and combustion stability was observed. A small drop in power and increase in CO emission were observed disadvantages of
the new injection system. Injection timing was found to be the most important factor and a balance between reduction in shortcircuited fuel by late injection, and time for mixture preparation by advancing the injection, was found to be essential.

KEY WORDS : Two stroke engines, Semi-direct injection, Emission control, Fuel Injection

1. INTRODUCTION

tion with a three-way catalytic converter) performs adequate
emission control. Such a technique will not yield significant results in a two-stroke engine, however, as the main
problem of short-circuiting of the fuel will still be present.
Injection of the fuel directly into the combustion chamber
is a desirable modification as it presents the possibility of
completely eliminating short-circuiting. Yamagishi et al.
(1972) developed a mechanical fuel injection system for a
350 cc two-stroke SI (Spark Ignition) engine wherein the
injector could be located in different places, for example,
the cylinder bore, head and transfer port (the port leading
the fresh charge from the crank case to the cylinder)
sequentially. In this work, injection at the cylinder head
was found to give the best results. An air assisted injection
system was developed for small outboard engines by

Leighton et al. (1994). This system used a small air compressor and a combination of air and fuel injectors mounted
on the cylinder head. The resultant highly-stratified fuel air
mixture reduced the specific fuel consumption by 40% and
exhaust emissions by 60%.
Pierre and Stephane (1996) developed an air assisted
fuel injection device called the IAPAC (Injection assistee
Air Comprime) system. Here, the crankcase compressed
air is used to push fuel that has been pre-injected into a
cavity into the engine. The engine uses a mechanically
operated valve located on the cylinder head to inject the
fuel. Emmenthal et al. (1985) and Schechter and Levin
(1991) have also developed air-forced injection systems.
Gentilli et al. (1994) developed a pumpless air-assisted

The demand for small capacity engines with high power to
weight ratio and low emissions is well known. Though the
four-stroke engine has replaced its two-stroke counterpart
in many applications, two-stroke engines have been the
primary choice for vehicles like mopeds, small scooters,
snowmobiles and hand held power tools. However, these
engines are also known for their high emissions of HC and
CO and fuel consumption. These defects are due to shortcircuiting of the fuel supply during the scavenging phase
and to the dilution of the fresh charge by residual exhaust
gases. The trapped residual charge is used to lower NOx
emissions. If the two-stroke design can be improved to
reduce the short-circuiting of fuel its advantages may make
it more attractive for use in applications in which fourstroke engines are the traditional choice. Despite the deficiencies inherent to current two-stroke designs, there are
still several two-stroke engine designs being manufactured
and used every year in applications such as those mentioned above. There are also many old engines still in service.
Hence, it will be desirable to devise methods that improve

the performance of current designs by simple modifications
and components that may allow retrofitting of existing
systems.
Injection of fuel into the manifold of four-stroke engines
has become a highly-perfected technique that (in conjunc*Corresponding author. e-mail:
151


152

M. LOGANATHAN and A. RAMESH

injection system that uses crank case compressed air and a
rotary injection valve. Cobb (2001) developed an injection
system for a small two-stroke engine that injects a mixture
of air and fuel using the energy of a compression wave
generated by cylinder gases. Syverten
(1996) found
that the injection spray type, spark plug location, injection
timing and fuel–air mixing are all very influential. A wide
spray produces a well-mixed cloud of fuel in the vicinity of
the spark plug to improve combustion. A narrow spray
produces a stratified air and fuel mixture near the spark
plug, resulting in unpredictable combustion. Johnson and
Wong (1998) have performed a related experimental investigation using a 1100cc 3 cylinder two stroke SI engine. The
fuel injectors used in the system are accumulator type fuel
injectors with high-speed solenoid valves. The engine is
capable of operating with very short injection durations
(less than 800 microseconds) with delivery a very low
amount of fue (22 mg). The fuel injector was located in the

cylinder head and the injection pressure was 140 bar. The
performance of the engine revealed that the HC emissions
decreased by 76% while the fuel consumption decreased
by 42%.
For two-stroke engines that must be run at high speeds,
achieving stable operation with direct injection is known to
be a difficult task. Stability is typically accomplished using
costly equipment, either to inject fuel at high pressures, or
to provide an air assisted injection system. Such technologies have proven difficult to control and so expensive
that the cost advantage of the two-stroke engine design is
lost. It is thus desirable to use injection equipment typically
used on four stroke SI engines to reduce the short-circuiting of fuel in two stroke engines (thereby capitalizing on
the low cost of widely used technology). These injectors,
however, cannot be used for direct injection as they cannot
withstand high temperatures and pressures. The pump used
also is not capable of developing the pressures necessary
for direct injection.
Semi-direct injection has already been investigated as a
means to reduce short circuiting by several researchers.
Douglas and Blair (1982) reported that injection of fuel in a
two-stroke engine at a pressure of 0.28 MPa from the
cylinder bore can lead to about a 30% reduction in fuel
consumption and a 60% reduction in exhaust emissions.
Vieilldent (1978) investigated a cylinder bore injection
system and indicated that it could reduce emissions and
specific fuel consumption while manifold injection resulted
in performance figures similar to that of a carbureted system.
Grasas-alsina
. (1986) investigated the performance
of low pressure fuel injection (0.3 MPa) at the transfer port

of a two stroke engine and concluded that that the fuel
consumption was reduced by 10% to 30% compared to a
carbureted engine. The injection timing for lowest fuel
consumption was 125 CA aTDC (after Top Dead Center)
and 130 CA aTDC at an engine speed of 3000 rpm. The
power at wide-open throttle decreased. Motoyama
.
(2002) investigated cylinder wall injection in two-stroke
et al.

et al

o

o

et al

engines for using a 430 cc, single cylinder engine. They
found that an injector position as close to TDC as possible
on the rear side of the transfer port is suitable and that the
entire amount of fuel should be injected toward the top of
the piston surface. The optimal timing was found to be
55 CA bTDC (before Top Dead Center) at a fuel injection
pressure of 0.6 MPa. A reduction in HC emission ranging
from 40% to 90% and of fuel consumption by 15% to 35%
was obtained.
o

2. PRESENT WORK

This work aims at bringing out the potential of the cylinder
barrel (semi-direct) injection method in a two-stroke engine
by means of a comparison to a well-tuned manifold injection system. A low-cost semi-direct injection system, based
on a conventional manifold gasoline injector used in four
stroke spark ignition engines, was developed for a small
two-stroke scooter engine. The injector was located on the
cylinder barrel opposite the exhaust port and was directed
toward the cylinder head. A small positive-displacement
gasoline pump, capable of developing a pressure of up to 8
bar, was used. The injector was positioned so as to obtain
stable combustion. Detailed tests were conducted to study
the influence of injection pressure, injection timing, spark
plug position and engine output at a constant speed of 3000
rpm. The spark timing was maintained at the optimum
level. Performance, emissions and combustion parameters
have been obtained. Earlier results obtained on the same
engine (by the authors) under manifold injection of gasoline are used for comparison.

3. EXPERIMENTAL SETUP
A single cylinder two-stroke spark ignition scooter engine
whose specifications are shown in Table 1 was connected
to an eddy current dynamometer. Fuel flow was measured
on a mass basis. Air flow was measured by a turbine flow
meter. Atmospheric pressure and temperature were used to
convert the volume flow rate of air to a mass flow rate. The
air-fuel ratio that is reported is the overall value as
measured at the inlet of the engine. A lubricating oil supply
system consisting of a plunger pump driven by a variable
speed motor was used.
The lubricating oil was injected directly into the air

stream leading to the crankcase. Since it was found that, at
certain operating conditions, the fuel touched some portion
of the cylinder wall, a small amount of lubricant was also
added to the fuel to avoid piston seizure. A potentiometer
was used to determine the throttle position.
An infrared gas analyzer (HORIBA, MEXA 554J) was
used to measure the exhaust hydrocarbon (n-hexane equivalent), carbon monoxide and carbon dioxide concentrations. The concentration of the NO in the exhaust was
detected using a chemiluminescence analyzer. A flush
mounted piezoelectric pressure transducer was used to


EXPERIMENTAL STUDIES ON LOW PRESSURE SEMI-DIRECT FUEL INJECTION

153

Table 1. Specifications of the engine.
Make
Bajaj chetak
Number of cylinders
One
Bore x stroke
57×57 mm
Compression ratio
8.8:1
Displacement volume
145.5×10−6m3
Rated power output
4.8 kW @ 5500 rpm
Ignition timing
22obTDC

Exhaust port opening (EPO)
72obBDC
Exhaust port closing (EPC)
72oaBDC
Transfer port opening (TPO) 52obBDC
Transfer port closing (TPC)
52oaBDC
Inlet port opening (IPO)
50obTDC
Inlet port closing (IPC)
50oaTDC
obtain cylinder pressure signals. The signals were captured
on a high-speed data acquisition system and then analysed
for computing heat release rate. Heat release rate was
calculated from the cylinder gas pressure history (Lanzafame
and Messina, 2003). The schematic of the test setup is
shown in Figure1. An electronic circuit was used to vary
the spark timing from 40o before TDC to 15oCA bTDC to
set the optimal value.
The semi-direct fuel injection system is comprised of a
gasoline injector, a fuel pump, a pressure regulator and
associated electronic circuits for controlling the injection
timing and duration. A positive-displacement pump was
used to supply gasoline at a pressure of 5 to 8 bar. The

Figure 2. Photographic view of the injector hole and the
spray (seen from the cylinder head side).

Figure 1. Experimental setup.
1. Engine, 2. Dynamometer, 3. Dynamometer controller, 4.

Data acquisition system, 5. Pressure transducer, 6. DC
sensor, 7. Ice path, 8. Exhaust gas analyzer, 9. Engine
speed indicator, 10. Temperature indicator, 11. Fuel
injector, 12. Pressure regulator, 13. Pressure gauge, 14.
Fuel pump, 15. Weighing machine, 16. Fuel tank, 17.
Throttle body, 18. Throttle position sensor, 19. Surge tank,
20. Air flow meter, 21. Electronic control system, 22. 12V
Battery, 23. Lubrication oil tank, 24. Lubrication oil pump,
25. Motor, 26. Air compressor

injector used was of the solenoid operated single hole type
with a pintle that gave a conical spray angle of 27o. A hole
was drilled in the cylinder barrel at an angle of 45o to the
cylinder axis, opposite the exhaust port, to locate the fuel
injector. This angle was selected so that the fuel would
reach the spark plug without interfering with the cylinder
wall.
A photograph of the injector and fuel spray is seen in
Figure 2. The injector holder was welded to the barrel and
care was taken to see that there was no deformation of
cylindrical barrel surface. The level of the injector above
the level of the exhaust port had to be carefully decided. If
the injector is not well above the level of the exhaust port
there will be little flexibility in the injection timing. On the
other hand, if the hole is much above the exhaust port, the
injector will be exposed to high pressure and high temperature exhaust gases during the expansion stroke. A bench
test on the injector used in this work indicated that it could
tolerate a back pressure of about 8 bar without any leak.
Based on the cylinder pressure crank angle curve obtained during normal engine operation, the level of the hole
was finally decided so that it would be completely closed at

82oCA bTDC. A specially developed monostable multivib-


154

M. LOGANATHAN and A. RAMESH

rator-based electronic circuit was used to generate pulses at
the correct timing and for the correct duration to energize
the injector. The output of the crankshaft position sensor
was conditioned and used to trigger this circuit. The range
of injection timings was 72oCA b BDC (before Bottom
Dead Center) to 11oCA bBDC and the injection duration
ranged between 1.5 ms and 6 ms. The amount of fuel
injected for a given injection duration depends on the
supply pressure of the fuel and on the downstream pressure
of the injector. The downstream pressure of the injector
rises if the injection extends into the region where the
exhaust port is closed.

4. RESULTS AND DISCUSSION
The engine was first tested with the semi-direct injection
system at an injection pressure of 6 bar (0.6 Mpa). The
injection pressure was continuously monitored and maintained. The effect of injection timing was initially determined at a representative throttle setting of 25% at a constant speed of 3000 rpm. Injection timings in the range of
72oCA bBDC to 11oCA bBDC were tried. It may be noted
that the port opening times are 72oCA bBDC for the
exhaust port and 52oCA bBDC for the transfer port. In
addition, the injector opening time was found out to be
1 ms. In all of the figures the start of injection actually
refers to the start of the pulse given to the injector, which

means that the injector needs a further (1 ms) 18o at 3000
rpm to open. The results obtained with the semi-direct
injection system have been compared with the results
obtained with a manifold injection system under similar
operating conditions. The manifold injection results were
obtained on the same engine and have been discussed in an
earlier paper by the authors(Loganathan
., 2006).

Figure 3. Variation of brake power with injection timing.

et al

4.1. Effect of Injection Timing, Injection Pressure and Spark
Plug Position
Changing the width of the pulse applied to the injector
varies the air-fuel ratio. As mentioned earlier the speed was
held constant at 3000 rpm and the throttle was kept at 25%.
The effects of power output and brake thermal efficiency
are seen in Figures 3 and 4. Very early injection will lead to
greater short-circuiting losses. Late injection may result in
insufficient time for good mixture preparation. We find
that, in the case of semi-direct injection, the injection timings of 72o and 49oCA bBDC are suitable. Later injection
timings lead to lower power and thermal efficiency. This is
probably predominantly because ocaused by having insufficient time to produce a combustible mixture. The injection
duration, with respect to the amount of fuel injected per
cycle, is seen in Figure 5 with data for various injection
timings. The curves within the dotted line indicate the
injection duration in milliseconds (ms) while while those
outside give the injection duration in crank angle degrees.

For the most suitable injection timings, 72o and 49oCA
bBDC, the injection duration extends from about 45o to 55o

Figure 4. Variation of brake thermal efficiency with injection timing.
crank angle. In the case of the retarded injection timings the
injection extends up to about 70oCA because of lower
thermal efficiency and the need to inject large quantities of
fuel. In all cases the injection stops before both the transfer
and exhaust ports close. Further, retardation of the injection
timing leads to irregular operation as the injection occurs,
even after the injection port closes. As the injection timing
is advanced, peak power is generally shifted to leaner
mixtures (Figure 3). Injection timings earlier than 72oCA
bBDC were not tried because the exhaust port would only
open at this angle. Too early injection (72oCA bBDC)
results in a good power probably due to the formation of a
homogeneous mixture. Though the injection timing of 49o
CA bBDC results in reduced power as compared to 72oCA
bBDC when the air-fuel ratio reduces, it leads to the
highest brake thermal efficiency on account of a reduction
in short-circuiting. It also leads to higher power with lean
mixtures. The peak thermal efficiency at 49oCA bBDC is


EXPERIMENTAL STUDIES ON LOW PRESSURE SEMI-DIRECT FUEL INJECTION

155

Figure 5. Variation of fuel flow rate with Injection duration.

Figure 7. Variation of CO emission with Injection duration.

26.8% at an air air-fuel ratio of 20.5:1, while the manifold
injection system gives a 25% efficiency at an air-fuel ratio
of 17:1 (Figure 4). Thus there is a significant improvement
in the thermal efficiency as compared to the manifold
injection method, which is mainly due to reduced shortcircuiting losses. Since the short-circuited charge is mostly
air, the trapped air-fuel ratios in the case of semi-direct
injection are expected to be lower than the overall air-fuel
ratios indicated in the graphs. The air-fuel ratios indicated
in the graphs are those obtained from the overall air and
fuel flow rates.
Emissions of HC and CO are seen in Figures 6 and 7.
There is a significant decrease in the HC level with semidirect injection due to the reduction in short circuiting. In a
two-stroke engine, HC emissions will depend on how much
fuel is short-circuited and also on the extent of combustion
of the trapped fuel. Hence, in the case of retarded injection
timings (even though a better trapping of the injected fuel
can be expected) the HC levels are higher due to poor

combustion because of insufficient time available for mixture formation.
Here again we see that the injection timing of 49oCA
bBDC is the best and is followed by 29oCA bBDC. The
minimum HC with manifold injection is at 1544 ppm with
an air-fuel ratio of 16.8:1 whereas it is only 600 ppm at an
air-fuel ratio of 19:1 with semi-direct injection. In the case
of an injection timing of 72oCA bBDC, the HC levels are
higher than those at 49oCA bBDC due to increased shortcrcuiting, even though the differences in peak thermal
efficiencies are not significant and power output is higher.
For any given air-fuel ratio, the CO level with semi-direct

injection is higher than with manifold injection because the
trapped air and fuel mixture is richer. With lean mixtures,
where almost all of the trapped fuel can be expected to
burn, we see that the CO level increases as we retard the
injection timing ie., as more fuel is trapped and the trapped
mixture becomes richer (Figure 7).
The NO levels seen in Figure 8 are lowest for the

Figure 6. Variation of HC emission with Injection duration.

Figure 8. Variation of NO emission with Injection duration.


156

M. LOGANATHAN and A. RAMESH

manifold injection method where all of the air and fuel are
premixed and the charge temperature is lowest. This leads
to attaining lower peak temperatures in this case.
The peak also occurs at a slightly leaner mixture than
stoichiometric in this case. With an injection timing of
72oCA bBDC, where the power is the maximum, the NO
level is also the highest. With an injection timing of 49oCA
bBDC, NO levels are also significantly higher than those
from the manifold injection technique. Also, in the case of
semi-direct injection (49o and 72oCA bBDC), the peak NO
level occurs when the overall mixture is lean. This is
because the trapped mixture is richer.
The exhaust gas temperature, as seen in Figure 9, is

highest at any given air-fuel ratio for the case with the most
retarded injection timing. Exhaust temperatures are lowest
under the manifold injection technique where the shortcircuiting of fuel will be the highest. As we retard the
injection timing, the amount of short-circuited fuel will be
reduced and this is probably the reason for the increase in
the exhaust gas temperature (which is actually the temperature of the exhaust, short-circuited air and fuel). The mass
of fuel injected per cycle with different injection timings is
shown in Figure 5.
In all cases the mass of fuel injected at any air-fuel ratio
is higher with the semi-direct injection system. This means
that the mass of air inducted is higher. In the case of the
manifold injection system, this is because the vaporized
fuel displaces some of the air. We find that the difference is
reduced as we move to leaner mixtures. Based on the
thermal efficiency and HC emissions, an injection timing
of 49oCA bBDC was selected as the most setting. The heat
release rates at different injection timings (at air-fuel ratios
that correspond to about an output of 2.3 kW) are compared in Figure 10. We see that the 72o and 49o timings
result in the highest heat release rates. This also means that
combustion is completed in a very short time, leading to
good thermal efficiency. Even though the heat release rate
is slightly higher with an injection timing of 72oCA bBDC

Figure 9. Variation of Exhaust gas temperature with Injection duration.

Figure 10. Variation of heat release rate with injection timing.

Figure 11. Variation of COV of IMEP with injection timing.
the short-circuiting losses are expected to be lower with
49oCA bBDC, leading to better efficiency. The COV of

IMEP with different timings (seen in Figure 11) shows that
the values are comparable for 72oCA and 49oCA The low
values indicate that the engine is quite stable at both of
these injection timings.
In subsequent experiments the injection pressure was
increased to study the influence of fuel atomization. Pressures over 8 bar could not be tried due to the limitations
posed by the pump and injector. These tests were also done
at 25% throttle, 3000 rpm and at an injection timing of
49oCA bBDC. The influence of injection pressure on
power, thermal efficiency and emissions is seen in Figures
(12)~(15). In between 6 and 8 bar, there is very little effect
seen in these parameters (there is a small decrease in the
HC and CO emissions). Hence, in all subsequent experiments, the injection pressure was held at 8 bar. Three spark
plug positions were tried (as shown in Figure 16). The air
movement in the cylinder will depend on the angle at
which the fuel spray is injected and this will affect the final