Influence of Heavy Fuel Properties on Spray Atomization for Marine

2009-01-1858
Influence of Heavy Fuel Properties on Spray Atomization
for Marine Diesel Engine Applications
Nikolaos Kyriakides, Christos Chryssakis, Lambros Kaiktsis
National Technical University of Athens (NTUA)
Dept. of Naval Architecture & Marine Engineering
Copyright © 2009 SAE International
ABSTRACT
In the present work, a model with the thermophysical
properties of Heavy Fuel Oil, typically used in marine
diesel engines, has been developed and implemented
into the KIVA CFD code. The effect of fuel properties on
spray atomization is investigated by performing
simulations in a constant-volume high-pressure
chamber, using the E-TAB and the USB breakup
models. Two different nozzle sizes, representative of
medium- and low-speed marine diesel engines, have
been considered. The simulations have been performed
for two values of chamber pressure, corresponding to
operation at partial and full engine load. The results
indicate that, in comparison to a diesel spray, the Heavy
Fuel spray is characterized by comparable values of
penetration length, and larger droplet sizes. These
findings are correlated to experimental results from the
literature.
INTRODUCTION
Current research efforts on optimizing diesel engine
operation attempt in minimizing pollutant emissions
without sacrificing fuel economy. To this end,
Computational Fluid Dynamics (CFD) simulations are a
valuable tool. The computed levels of pollutant
concentrations and engine output are heavily dependent
on a proper description of spray atomization processes.
Currently used physical models of spray atomization are
based on conditions representative of small automotive
engines. In large marine diesel engines, the nondimensional parameters affecting the spray dynamics
differ substantially, due to both the larger size of
injectors and the use of Heavy (or Residual) Fuel Oil
(HFO). Roughly 2/3 of all merchant ships are operated
with HFO. Commonly, the composition of HFO varies
substantially, introducing many uncertainties in
modeling. Reported values of the kinematic viscosity of
HFO at 50oC can range between 50 cSt and 800 cSt,
compared to approximately 3 cSt for automotive diesel
fuel. Furthermore, surface tension values are
substantially higher for HFO (by order 20%). It is
expected that the differences in fuel properties affect the
dynamics of spray atomization, droplet evaporation and
fuel-air mixing in marine diesel engines. Thus, validation
and further development of the existing models is
necessary for modeling spray dynamics in these large
engines.
With reference to medium-speed marine engines, a first
experimental study performed by Fink et al. [1] suggests
that, in comparison to automotive diesel sprays, HFO
sprays exhibit small differences in terms of tip
penetration, with the deviations in droplet sizes being
significant. For large (low-speed) marine diesel engine
conditions, a constant-volume combustion chamber has
been developed [2], and experiments are currently
underway. In 2006, Goldsworthy [3] presented a
simplified model for heavy residual fuel, with constant
viscosity, density and latent heat of evaporation, in
which the surface tension was defined as a function of
temperature. The model was used in CFD simulations
for two representative fuels, and good agreement was
reported between measured and computed data for
ignition delay, burning rate, and spatial form of the spray
and flame. More recently, Struckmeier et al. [4] adopted
Goldsworthy’s approach, and elaborated it by adding
multi-component droplet evaporation. This was achieved
The Engineering Meetings Board has approved this paper for publication. It has successfully completed SAE’s peer review process under the supervision of the
session organizer. This process requires a minimum of three (3) reviews by industry experts.
All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted, in any form or by any means, electronic,
mechanical, photocopying, recording, or otherwise, without the prior written permission of SAE.
ISSN 0148-7191
Positions and opinions advanced in this paper are those of the author(s) and not necessarily those of SAE. The author is solely responsible for the content of
the paper.
SAE Customer Service: Tel:
877-606-7323 (inside USA and Canada)
Tel:
724-776-4970 (outside USA)
Fax:
724-776-0790
Email: [email protected]
SAE Web Address:
http://www.sae.org
Printed in USA
by applying different saturation pressures for light and
heavy fuel components. The comparison of spray
simulations with experimental measurements showed
good agreement in terms of spray penetration,
evaporation, and flame lift-off length.
Motivated by similar studies in automotive engines,
recent attempts in optimizing pollutant emissions and
engine performance of marine diesel engines exploit
advanced injection strategies with pilot injection [5-6].
Typically, CFD studies as these of [5-6] rely on diesel oil
properties for modeling spray dynamics. Thus, future
studies should take into account the effects of marine
fuel properties. To this end, a new fuel model is
developed and tested in the present study. The new
model accounts for the thermophysical properties of a
representative heavy marine fuel. The model is
implemented into a KIVA-based CFD code, and
evaluated for a prototype fuel spray in a constant volume
combustion chamber. Results for HFO are compared to
computational results for a conventional diesel fuel, as
well as with the experimental measurements of [1], in
terms of spray breakup behavior.
SPRAY MODELING
The developed HFO model has been implemented into a
modified version of the KIVA code [7-8]. In order to study
the effects of fuel properties on spray atomization, two
different spray breakup models have been utilized. The
breakup models under consideration are a modified
version of the Enhanced - Taylor Analogy Breakup (ETAB) model [9], and the Unified Spray Breakup (USB)
model [10]. The selection of the two spray models is
motivated by the facts that the E-TAB model has been
extensively tested for diesel spray simulations in marine
engine applications [5, 6, 9, 11], while the USB model
has been recently developed and validated under a wide
range of conditions, yielding promising results for
conditions applicable to typical diesel sprays. In the
following, basic features of the two models are outlined.
E-TAB MODEL
The E-TAB model reflects a cascade of droplet
breakups, in which the breakup condition is determined
by the Taylor droplet oscillator dynamics. The droplet
size is reduced in a continuous manner, until the product
droplets reach a stable condition. The model maintains
the droplet deformation dynamics of the TAB model [12].
According to this approach, the droplet distortion is
described by a forced damped harmonic oscillator, in
which the forcing term corresponds to the aerodynamic
droplet-gas interaction, the restoring force is due to
surface tension, while damping is attributed to the
liquid’s viscosity.
Breakup occurs when the normalized (with respect to
the initial radius) droplet distortion exceeds the critical
value of 1. The rate of droplet creation is
d
m(t ) = −3K br m(t )
dt
where m(t) is the mean mass of a droplet’s product
distribution, and Kbr a constant that depends on the
breakup mechanism [9]. This correspondingly leads to
an exponential relation between the product and parent
droplet radius, r and α, respectively:
r
α
= e − K br t .
Droplets are initialized with a “negative” deformation
velocity in order to avoid the almost immediate breakup
of highly unstable initial ligaments, and to extend their
lifetime to levels comparable with experimentally
observed jet breakup lengths [9].
USB MODEL
In this model, the spray breakup has been divided into
three
distinct
sub-processes,
namely,
primary
atomization, drop deformation due to aerodynamic drag,
and secondary atomization [10].
The primary atomization is modeled based on the Huh et
al. approach [13]. Here, the effects of both infinitesimal
wave growth on the jet surface and jet turbulence
(including cavitation dynamics) are considered. Initial
perturbations on the jet surface are induced by turbulent
fluctuations of the flow inside the nozzle. This approach
overcomes the inherent difficulty of wave growth models,
where the (exponential) growth of disturbances becomes
zero at zero perturbation amplitude. The primary
atomization model is based on two main assumptions:
1. The integral length scale of turbulence is the
dominant length scale of primary atomization.
2. The time scale of the atomization is a weighted
sum of a turbulent and a wave growth time
scale.
The drop deformation and secondary atomization model
builds on top of already existing drops, generated by
primary atomization. The secondary atomization has
been further divided into four breakup regimes, based on
experimental observations reported in the literature [14].
In accordance with the findings of [14], determination of
the appropriate secondary atomization regime depends
only on the Weber number of the droplets, defined as:
We =
ρ GU 2 d o
,
σ
where ρG is the density of the ambient gas, U the
droplet velocity, do the droplet initial diameter upon its
creation, and σ the surface tension. For low Weber
numbers (less than 12), atomization does not occur, and
only droplet deformation takes place. For higher values
of Weber number, the following regimes are identified in
[14]:
•
•
•
•
to 953.7 kg/m3, compared to 848 kg/m3 for the standard
diesel fuel models in KIVA. It is noted that the liquid
density is a decreasing function of temperature, which
will be addressed in future development of the model.
Bag breakup, 12<We<20
Multimode breakup, 20<We<80
Shear breakup, 80<We<800
Catastrophic breakup, 800<We
In [10], the breakup times and resulting droplet sizes are
estimated from experimentally verified correlations for
each breakup regime. Once a droplet’s secondary
breakup has been completed, further disintegration
(tertiary breakup) is not possible [15], and droplets reach
a stable condition.
In experiments with isolated droplets, Faeth and
coworkers [14-15] observed that, for sufficiently high
values of the liquid’s viscosity, the limits of breakup
regimes are affected. The effect of the liquid’s viscosity
is accounted for in the Ohnesorge number, defined as:
Oh =
µL
,
(ρ L d oσ )1/ 2
where µL, ρL are the dynamic viscosity and density of the
droplet, respectively. As the viscosity increases, the
value of Weber number required for the onset of
breakup increases. As a consequence, the transitions
between the four above-mentioned regimes occur at
higher Weber numbers; thus, breakup of the liquid
droplets takes place at a slower pace. This is illustrated
in Figure 1, where a map of the breakup regimes is
presented, based on the findings of [14], as compiled by
Chryssakis and Assanis [10]. On this map, the areas of
typical gasoline and diesel sprays for automotive
applications are identified. As illustrated in Figure 1, the
dependence of breakup on Ohnesorge number is
important only for values higher than order 1.
Five other thermophysical properties have been
accounted for, in the HFO model. These are the liquid’s
dynamic viscosity, surface tension, vapor pressure,
latent heat of evaporation, and specific enthalpy.
Calculation of the five above-mentioned properties
depends on molecular weight and critical pressure and
temperature of an equivalent fuel. These values have
been determined by considering the specific gravity of
HFO (ratio of fuel density to water density) taken at 288
K, and its distillation curve, according to the ASTMD1160 standard (Dr. G. Bellos, private communication).
The corresponding values for molecular weight and
critical properties are determined by discretizing the
distillation curve into small segments, and assuming one
hydrocarbon for each segment. Finally, the fuel is
treated as a mixture of these components, and its
properties are calculated. Based on these calculations,
the Molecular Weight of HFO has been set to 463.
The critical temperature of HFO has been approximated
by using the correlation [16]:
Tc = 189.8 + 450.6 ⋅ S + Tb ⋅ (0.4244 + 0.1174 ⋅ S ) +
+
14410 − 100688 ⋅ S
Tb
where Tb is the normal boiling point of the fuel and S the
standard specific gravity, here set to 0.9537. The normal
boiling point can be estimated from the relationship [16]:
(1.8 ⋅ T )
1 .3
Kw =
Breakup Regimes
100000
Catastrophic
Weber
1000
Shear
100
10
Multimode
Gasoline
Bag
Osc. Deform.
1
0.001
0.01
0.1
Ohnesorge
S
,
where Kw is the Watson characteristic factor, and can be
taken equal to 11.7 for HFO (Dr. G. Bellos, private
communication). Solving for Tcr, the value of 1033 K is
obtained for the critical temperature of HFO.
Diesel
10000
b
1
10
Figure 1: Map of secondary atomization regimes as
functions of Ohnesorge and Weber numbers, in which
the areas representative of automotive gasoline and
diesel sprays are identified [10].
HEAVY FUEL OIL MODEL
The fuel properties used in the present work are
representative of an average residual fuel, typically used
in large marine diesel engines. The liquid density is
assumed to be constant. Here, the density has been set
In Figures 2 and 3, the temperature dependence of the
dynamic viscosity and surface tension is presented, both
for automotive diesel and for HFO. It is evident that,
depending on the temperature, the dynamic viscosity is
1-2 orders of magnitude higher for HFO, while the
surface tension is 20%-30% higher for HFO.
while heavy components evaporate at a slower rate.
This effect should be taken into account in future
development of the present model.
Diesel
10000
HFO
1000
100
1.00E+00
10
1.00E-01
Vapor Pressure [bar]
Dynamic Viscosity [cP]
100000
1
0.1
250
300
350
400
450
Temperature [K]
Figure 2: Dynamic viscosity for diesel fuel and HFO, as
a function of temperature.
1.00E-02
1.00E-03
1.00E-04
1.00E-05
Diesel
1.00E-06
HFO
1.00E-07
250
300
350
400
450
Temperature [K]
0.035
Figure 4: Vapor pressure for diesel fuel and HFO, as a
function of temperature.
HFO
0.03
0.02
0.015
250
300
350
400
450
Temperature [K]
Figure 3: Surface tension for diesel fuel and HFO, as a
function of temperature.
As deduced from the previous discussion on atomization
regimes, the increased liquid viscosity of HFO will raise
significantly the Ohnesorge number; this effect will, to
some extent, be counteracted by the increased liquid
density and surface tension of HFO. On the other hand,
the increased values of surface tension will lead to lower
Weber numbers, resulting in slightly slower atomization.
However, it should be noted here that marine fuels are
typically injected at higher temperatures (of the order of
80oC), in order to reduce their viscosity, and enable the
flow in the injection system lines. Thus, as deduced from
Figure 3, the use of HFO has a minor effect on the
surface tension values. In the following, an injection
temperature of 353 K will be considered for HFO.
In Figure 4, the vapor pressure curves for diesel and
HFO are presented. It is clear that HFO has 2-3 orders
of magnitude lower vapor pressure, resulting in lower
evaporation rates.
Heat of Evaporation [kcal/kg]
180.0
0.025
160.0
140.0
120.0
Diesel
100.0
HFO
80.0
60.0
250
300
350
400
450
Temperature [K]
Figure 5: Heat of evaporation for diesel fuel and HFO,
as a function of temperature.
The specific enthalpy of the heavy fuel is shown in
Figure 6, and compared with the specific enthalpy of
automotive diesel fuel. The thermal conductivity of HFO
has been maintained the same as for the automotive
diesel fuel.
-400.0
Specific Enthalpy [kcal/kg
Surface Tension [N/m]
Diesel
-420.0
Diesel
-440.0
HFO
-460.0
-480.0
-500.0
-520.0
-540.0
250
300
350
400
450
Temperature [K]
In Figure 5, the heat of evaporation curves for diesel and
HFO are presented. Only one value was available for
HFO (at 273 K); as a simplifying assumption, the slope
of this curve has been maintained the same as for the
diesel fuel. It should also be noted that the utilized heat
of evaporation is an average value. In reality, the fuel
consists of many components, resulting in a gradual
evaporation process: light components evaporate first,
Figure 6: Specific enthalpy for diesel fuel and HFO, as a
function of temperature.
Finally, the effect of the thermophysical properties of
HFO on engine operation is demonstrated by presenting
simulation results for a large two-stroke marine diesel
engine.
RESULTS AND DISCUSSION
TEST CASES
The problem setup consists of a constant-volume
chamber with dimensions 20cm×10cm×10cm (axial and
two vertical directions). The chamber is filled with N2, the
pressure is maintained at 100 bar (10 MPa) or 30 bar (3
MPa), and the gas and wall temperatures at 400 K. The
temperature has been kept low so that only the liquid
atomization of the fuel jets is investigated, by eliminating
the effects of evaporation. The two different gas
pressures correspond to full and partial load conditions
in a marine diesel engine. Two different nozzle
diameters are tested, one representative of large lowspeed two-stroke marine diesel engines, and one of
medium-speed marine engines, as shown in Table 1.
Table 1: Nozzle characteristics.
Diameter
[mm]
Nozzle A
Large, LowSpeed Engine
0.9
Nozzle B
MediumSpeed Engine
0.37
14
Inj. Velocity [m/s]
Nozzle A
12
Diesel
p=30 bar
HFO
10
8
6
Nozzle B
4
Nozzle A
2
0
0.00
0.50
1.00
1.50
2.00
Time [ms]
10
p=100 bar
Diesel
8
Tip Penetration [cm]
600
400
An interesting observation is that for both pressures
investigated, initially, the penetration of diesel fuel is
slightly higher for Nozzle A, while the opposite trend is
observed for Nozzle B. This indicates that the nozzle
size and injection velocity (Weber number) affect the
primary atomization mechanism. These effects need to
be quantified experimentally.
16
The injection velocity profile for Nozzle A (Figure 7) is
representative of a typical velocity profile of a two-stroke
engine operating at full load, and has been measured by
Wärtsilä Switzerland (Dr. G. Weisser, private
communication). For Nozzle B, a simplified, constant
injection rate has been used, with an injection velocity of
300 m/s, representative of injection in medium-speed
marine engines [11]. In all cases the behavior of diesel
and HFO fuels is investigated. Diesel fuel is simulated
using C14H30, similarly to many typical engine
simulations.
500
The effect of fuel properties on tip penetration is shown
in Figure 8, for gas pressures of 30 and 100 bar, based
on the USB model predictions. In both cases, the tip
penetration is initially lower for Nozzle A (d=0.9 mm),
since the injection velocity increases from zero to
approximately 500 m/s, but settles at significantly higher
values at around 2 ms after the start of injection. (In
Nozzle B, the velocity is held constant at 300 m/s).
Tip Penetration [cm]
Application
INFLUENCE OF FUEL PROPERTIES ON SPRAY
ATOMIZATION
HFO
Nozzle A
6
4
Nozzle B
2
300
Nozzle B
200
0
0.00
100
0.50
1.00
1.50
2.00
Time [ms]
0
0
5
10
15
20
25
Figure 8: Tip penetration vs. time, for diesel fuel and
HFO, as predicted with the USB model.
Time [ms]
Figure 7: Injection velocity histories for Nozzles A and
B.
A grid sensitivity analysis was performed by the authors
in [17], and showed that both the E-TAB and USB
models are not grid sensitive for computational cell sizes
smaller than 2 mm in the direction of the spray axis. In
the present work, a resolution of 1.5 mm in the axial
direction has been implemented.
In Figure 9, the spray tip penetration lengths, as
calculated from the E-TAB model, are presented, for
Nozzle B (d=0.37 mm). The tip penetration prediction is
significantly higher than the predictions of the USB
model. The large discrepancies illustrate the need for
experimental data, for validating both models under
conditions relevant to marine diesel engines. In this
context, the models can be optimized, see [18] for spray
model adaptation for small nozzles. Interestingly, the ETAB model predicts consistently longer tip penetration
lengths for the HFO, compared to diesel fuel. This
behavior should be attributed to the larger droplet sizes
predicted by the E-TAB model, as discussed
subsequently.
20
Tip Penetration [cm]
Diesel
p=30 bar
HFO
15
10
p=100 bar
5
0
0.00
the control volume centered at 4 cm from the nozzle exit.
For the diesel fuel, the SMD values are smaller but
comparable to the predictions of the USB model.
However, for HFO, the predicted SMD values are much
higher, close to 200 µm. It is very likely that this is the
reason for the consistently longer tip penetration lengths
predicted with the E-TAB model, for HFO (larger
droplets face smaller aerodynamic drag per unit
volume). It is expected that the addition of a catastrophic
breakup mechanism, in the context recently
implemented by Tanner [19], as well as appropriate
adaptation of the model, will lead to more realistic
results.
500
0.50
1.00
1.50
2.00
Diesel
400
Figure 9: Tip penetration vs. time in the case of Nozzle
B, for diesel fuel and HFO, as predicted with the E-TAB
model.
Further insight into the breakup process can be gained
by examining the droplet size predictions. These are
presented in the form of the Sauter Mean Diameter
(SMD) at a control volume centered 4 cm downstream of
the nozzle exit. Specifically, the control volume has a
height of 6 mm, and all droplets found therein have been
used to calculate the SMD value. In Figure 10, the SMD
predictions of the USB model are presented for Nozzle B
(d=0.37 mm), for both gas pressures of 100 and 30 bar.
Under both pressures, the predicted SMD is significantly
higher in the HFO case, as compared to diesel fuel. This
is in agreement with the experimental measurements of
Fink et al. [1], who measured higher SMD values for
HFO, compared to pure diesel. For comparable
conditions (the gas pressure was maintained at 14 bar in
their experiments), they measured HFO droplet sizes of
approximately 50 µm and diesel fuel droplet sizes of 3035 µm. For HFO, the increase in SMD values is
attributed to the increased surface tension and viscosity
values, compared to diesel fuel.
200
SMD [µm]
Time [ms]
HFO
300
200
100
0
0
0.5
1
1.5
2
Time [ms]
Figure 11: Sauter Mean Diameter vs. time, for diesel
fuel and HFO, as predicted with the E-TAB model, for
Nozzle B, for gas pressure of 30 bar.
In order to understand the effects of the liquid’s viscosity
and surface tension, the computational predictions can
be presented in the form of Weber number vs.
Ohnesorge number plots, combined with the secondary
atomization regimes, in the context of Figure 1. In Figure
12, the secondary breakup regimes for Nozzles A and B
are presented, as predicted with the USB model. Each
point on the maps represents one computational parcel,
corresponding to a number of identical fuel droplets. The
Weber and Ohnesorge numbers for each droplet are
calculated based on its diameter and velocity, at the
onset of secondary breakup. (The limiting curves on the
maps separate the different breakup regimes, see also
Figure 1).
Diesel
SMD [µm]
150
HFO
p=100 bar
100
50
p=30 bar
0
0
0.5
1
1.5
2
Time [ms]
Figure 10: Sauter Mean Diameter vs. time, for diesel
fuel and HFO, as predicted with the USB model, for
Nozzle B, for gas pressures of 100 bar and 30 bar.
The SMD predictions of the E-TAB model are presented
in Figure 11, for Nozzle B and gas pressure of 30 bar, at
The clouds of droplets on the left hand side of the maps
in Figure 12, with Ohnesorge numbers close to 0.1,
correspond to the diesel fuel, while the clouds with
Ohnesorge numbers close to 1.0 correspond to HFO.
While the range of Weber numbers covered is
comparable between the two fuels, the Ohnesorge
number is roughly an order of magnitude higher for
HFO. The difference in Ohnesorge number is due to the
viscosity of HFO, which is significantly higher, even
when the fuel is preheated. Furthermore, it can be
observed that, for Nozzle A (d=0.9 mm), there are
droplets with Weber numbers higher than 10,000, as
opposed to Nozzle B, where 10,000 seems to be an
upper limit. This is a result of both larger droplet sizes
and higher droplet velocities for Nozzle A. It is noted
that, in all cases, the catastrophic atomization
mechanism is the dominant one, followed by the shear
atomization mechanism, see also Figure 1.
10000
Weber [ - ]
It is also observed in Figure 12 that, for Nozzle A, a
small number of droplets is characterized by very low
Weber numbers. This is due to the initial slope of the
velocity profile, corresponding to the opening of the
needle in the injector nozzle (Figure 7). As for Nozzle B,
a constant injection velocity of 300 m/s was assumed,
this effect is absent. It is noted here that droplets with
low Weber numbers will break up at a slower rate, also
leading to low evaporation rates. This may also result in
combustion of liquid fuel, which, in engine operation, is
likely to cause increased local soot formation rates.
100
1
0.01
1.00
10.00
Nozzle A, p=100 bar
Weber [ - ]
10000
Diesel
HFO
1000
100
10
1
0.01
0.10
1.00
10.00
Ohnesorge [ - ]
Nozzle B, p=30 bar
100000
Diesel
HFO
10000
Weber [ - ]
0.58
2.416
Common Rail
0.10
Ohnesorge [ - ]
1000
100
10
-96
1
0.01
120
0.10
1.00
10.00
Ohnesorge [ - ]
105
A full load operating condition, with a single fuel injection
starting at 2o CA aTDC has been considered. This
operating condition has been studied in previous work
by the authors [5, 6], in which the computational results
were validated by comparing against experimental data.
The present simulations have been performed by using
the same initial fuel temperature (320 K) for diesel
(approximated by C14H30) and HFO, for direct
comparison. However, under realistic engine operating
conditions, higher fuel temperatures (of the order of 350
K) are used for HFO. The effect of fuel properties on
evaporation is demonstrated in Figure 13, where
isosurfaces of fuel-air equivalence ratio φ=2.0 are shown
for both diesel and HFO. These areas include the
regions φ>2.0. It is evident that the size of the fuel-rich
regions is significantly smaller for HFO, demonstrating
slow evaporation rates due to poor spray atomization.
This leads to low burning rates and potentially to
incomplete combustion, associated with lower
temperatures, thus further contributing to low
evaporation rates.
Nozzle B, p=100 bar
100000
10000
Weber [ - ]
Bore Diameter [m]
Stroke [m]
Injection System
Exhaust Valve Closing, EVC
[o aTDC]
Exhaust Valve Opening, EVO
[o aTDC]
Engine Speed [RPM]
HFO
10
APPLICATION OF THE HFO MODEL IN A MARINE
ENGINE
Table 2: Main engine characteristics
Diesel
1000
100000
A preliminary calculation in a large-bore marine diesel
engine has been performed, in order to demonstrate the
effects of fuel properties on fuel evaporation and
combustion. The engine geometry corresponds to the
RT-flex58T-B Sulzer engine. Each cylinder has three
injectors, located symmetrically on the periphery of the
cylinder head; each injector has five orifices, to enhance
dispersion of the liquid fuel into the cylinder. In general,
the injection is in a co-swirl direction. The main engine
characteristics are given in Table 2.
Nozzle A, p=30 bar
100000
Diesel
HFO
1000
100
10
1
0.01
0.10
1.00
10.00
Ohnesorge [ - ]
Figure 12: USB model predictions: secondary breakup
regimes for diesel fuel and HFO, for Nozzles A and B,
under gas pressures of 100 bar and 30 bar.
The effect of HFO on combustion is illustrated in Figure
14, where the calculated cylinder pressure traces for
diesel and HFO are presented. Due to the lower
evaporation rates, combustion is substantially weaker for
HFO.
Diesel
for both fuels, the catastrophic breakup mechanism is
the most relevant one, followed by the shear breakup
mechanism. Finally, first simulations of combustion in a
large two-stroke marine diesel engine have been
performed with the HFO model.
HFO
Further work on HFO should include more detailed
evaporation studies, as well as adaptation of the spray
breakup models against experimental data.
ACKNOWLEDGMENTS
We would like to thank Dr. G. Bellos of Motor Oil Hellas
(currently with KBC Process Technology Ltd.,
Netherlands) for providing the thermophysical properties
data for Heavy Fuel Oil, and for helpful discussions. We
would also like to thank Mr. C. Pantazis of NTUA for
contributing to the engine simulations, and Dr. G.
Weisser of Wärtsilä Switzerland for helpful discussions.
The second author acknowledges the financial support
by a Marie-Curie International Reintegration Grant,
Agreement No. 207232.
REFERENCES
Figure 13: Isosurfaces of fuel-air equivalence ratio
φ=2.0, at different time instants for diesel and HFO.
Cylinder Pressure [bar]
160
140
120
100
80
60
HFO
40
Diesel
20
0
-30
-20
-10
0
10
20
30
Crank Angle [deg. aTDC]
Figure 14: Computed cylinder pressure traces for diesel
and HFO.
CONCLUSIONS
A physical model for Heavy Fuel Oil, typically used in
large marine diesel engines, has been developed. The
main properties that affect spray atomization are the fuel
viscosity and surface tension. The model has been
implemented in the CFD code KIVA, and evaluated with
simulations in a constant-volume chamber. Computational results for two nozzle sizes have been
presented, using two different spray atomization models.
Two different gas pressures have been considered. In
agreement with the experimental data of [1], the present
results illustrate that tip penetration lengths for HFO and
diesel are comparable, while droplet sizes are
significantly larger for HFO. It has been established that,
1. Fink, C., Buchholz, B., Niendorf, M., Hard, Harndorf,
H., Injection Spray Analyses from Medium Speed
Engines Using Marine Fuels, presented at ILASSEurope 2008, Lake Como, Italy, September 2008
2. Herrmann, K., Schulz, R., Weisser, G., Development
of a Reference Experiment for Large Diesel Engine
Combustion System Optimization, presented at
CIMAC Congress 2007, Vienna, May 2007
3. Goldsworthy, L., (2006). Computational Fluid
Dynamics Modelling of Residual Fuel Oil
Combustion in the Context of Marine Diesel
Engines, Int. J. Engine Research, Vol. 7, pp. 181199
4. Struckmeier, D., Goldsworthy, L., Tajima, H.,
Kawauchi, S., Tsuru, D., Modelling of Ignition and
Combustion Characteristics of Heavy Residual Fuel
Oil on Two Component Approximation, presented at
7th International Colloquium Fuels, Stuttgart,
Germany, January 2009
5. Kontoulis, P., Chryssakis, C., Kaiktsis, L., Evaluation
of Pilot Injections in a Large Two-Stroke Marine
Diesel Engine, Using CFD and T-φ Mapping,
presented at COMODIA 2008, Sapporo, Japan, July
2008
6. Andreadis, P., Chryssakis, C., Kaiktsis, L.,
Optimization of Injection characteristics in a Large
Marine
Diesel
Engine
Using
Evolutionary
Algorithms, presented at SAE World Congress,
2009-01-1448, Detroit, MI, April 2009
7. Amsden, A.A., KIVA-3: A KIVA Program with BlockStructured Mesh for Complex Geometries, Los
Alamos National Laboratory LA-12503-MS, 1993
8. Amsden A.A., KIVA-3V: A Block-Structured KIVA
Program for Engines with Vertical or Canted Valves,
9.
10.
11.
12.
13.
14.
Los Alamos National Laboratory LA-13313-MS, July
1997
Tanner, F.X., Weisser, G., Simulation of Liquid Jet
Atomization for Fuel Sprays by Means of a Cascade
Drop Breakup Model, presented at the SAE World
Congress, 980808, Detroit, MI, 1998
Chryssakis, C., Assanis, D.N. (2008). A Unified Fuel
Spray Breakup Model for Internal Combustion
Engine Applications, Atomization and Sprays, Vol.
18, No. 5, pp. 375-426
Larmi, M., Rantanen, P., Tiainen J., Kiijarvi, J.,
Tanner, F.X., Stalsberg-Zarling, K., Simulation of
Non-Evaporating Diesel Sprays and Verification with
Experimental Data, presented at SAE World
Congress 2002-01-0946, Detroit, MI, 2002
O’Rourke, P.J., Amsden, A.A., The TAB method for
numerical calculation of spray droplet breakup, SAE
Technical Paper 872089, 1987
Huh, K.Y. Lee, E. Koo, J.-Y., Diesel Spray
Atomization Model Considering Nozzle Exit
Turbulence Conditions, Atomization and Sprays, vol.
8, pp. 453–469, 1998
Faeth, G.M., Hsiang, L.-P., Wu, P.-K. (1995).
Structure and Breakup Properties of Sprays,
International Journal of Multiphase Flow, Vol. 21,
Suppl. pp. 99-127
15. Hsiang, L.-P., Faeth, G.M., (1993). Drop Properties
after Secondary Breakup, Int. J. Multiphase Flow,
Vol. 19, No. 5, pp. 721-735
16. Wauquier, J.-P., Crude Oil Petroleum Products
Process Flowsheets, p. 97, Editions Technip, IFP
Publications, 1995
17. Chryssakis, C., Kaiktsis, L., Evaluation of Fuel Spray
Atomization Models for Conditions Applicable to
Marine Engine Applications, presented at ILASSEurope-2008, Lake Como, Italy, September 2008
18. Pizza, G., Wright, Y.M., Weisser, G., Boulouchos, K.
(2005). Evaporating and Non-Evaporating Diesel
Spray Simulation: Comparison Between the ETAB
and Wave Breakup Model, Int. J. Vehicle Design,
Vol. 45, Nos. 1/2, pp. 80-99
19. Tanner, F.X. (2004). Development and Validation of
a Cascade Atomization and Drop Breakup Model for
High-Velocity Dense Sprays, Atomization and
Sprays, Vol. 14, No. 3
CONTACT
Mr. Nikolaos Kyriakides: [email protected]
Dr. Christos Chryssakis: [email protected]
Prof. Lambros Kaiktsis: [email protected]