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RESEARCH ARTICLE
Dependence of combustion dynamics in a gasoline engineupon the in-cylinder flow field, determined by high-speed PIV
M. Buschbeck • N. Bittner •
T. Halfmann • S. Arndt
Received: 16 May 2012 / Revised: 9 August 2012 / Accepted: 31 August 2012 / Published online: 26 September 2012
� Springer-Verlag 2012
Abstract We apply time-resolved high-speed particle
image velocimetry (PIV) in an optically accessible gasoline
engine to determine the effect of the in-cylinder flow field
upon combustion dynamics. Our PIV setup involves solid
particles as tracer, which enables also measurements at
firing top dead center and during the combustion process
itself. We analyze the flow field for the entire intake and
compression phase, as well as the decay of a prominent
large-scale tumble structure in the flow field. The data
indicate significant cycle-to-cycle flow field variations,
characterized by detection of kinetic energy and tumble
center. Measurements in fired engine operation demon-
strate the influence of the flow field on combustion
dynamics. At stoichiometric operation, we find that varia-
tions in the kinetic energy of the flow field are a major
cause of cycle-to-cycle variations. From simultaneous
imaging of the combustion flame and PIV at lean opera-
tion, we find that the velocity distribution in the flow field
induces a macroscopic motion of the flame kernel—which
significantly effects the combustion process.
1 Introduction
Spark ignition internal combustion engines show random
variations of the in-cylinder pressure from one cycle to the
next. This phenomenon is known as cycle-to-cycle varia-
tions. For an overview of the literature, see Ozdor et al.
(1994) and references therein. Cycle-to-cycle variations
occur in direct injection as well as in port fuel injection
engines. Moreover, the cyclic variability is especially
dominant at light loads (particularly at idle) or high dilu-
tion due to lean mixture or exhaust gas recirculation
(Heywood 1988).
The variations induce losses in terms of thermodynamic
efficiency and hence affect the specific fuel consumption.
As an example, fast burning cycles tend to knock and
therefore limit the engine compression ratio. One of the
main challenges in engine development is to gain a fun-
damental understanding of the origin of cycle-to-cycle
variations. Usually, the cyclic variability is attributed to
random fluctuations in equivalence ratio and in-cylinder
flow field. Therefore, the fluid mechanics of internal
combustion engines attracted scientific interest already for
many decades (Arcoumanis and Whitelaw 1987). Particle
image velocimetry (PIV) serves as a powerful tool to
determine such fluid mechanics. The flow under investi-
gation is seeded with small particles, which follow the
motion of the flow. Light pulses with sufficiently short time
delay illuminate the particles to generate images on a
camera. The displacement of the particles during the time
interval between the light pulses enables determination of
the velocity of the flow.
One of the first applications of PIV to an internal
combustion engine was carried out by Reuss et al. (1989).
Further PIV studies investigated the cyclic variation of the
in-cylinder flow field (Reuss 2000; Li et al. 2001). Recent
developments in camera and laser technology allow the
crank angle resolved acquisition of PIV images (Towers
and Towers 2004; Muller et al. 2010). This high-speed PIV
yields information on the evolution of the flow field for
M. Buschbeck (&) � N. Bittner � S. Arndt
Robert Bosch GmbH, Corporate Research,
Robert-Bosch-Platz 1, 70839 Gerlingen-Schillerhohe, Germany
e-mail: [email protected]
T. Halfmann
Institute of Applied Physics, TU Darmstadt,
Hochschulstr. 6, 64289 Darmstadt, Germany
e-mail: [email protected]
123
Exp Fluids (2012) 53:1701–1712
DOI 10.1007/s00348-012-1384-3
individual engine cycles. In most PIV studies, the engine
was operated under motored conditions, but there are also
some recent documentations on PIV in fired combustion
engines (Fajardo et al. 2006; Gajdeczko and Bracco 1999;
Fajardo and Sick 2007). In particular, Fajardo and Sick
(2007) performed high-speed UV PIV in stratified opera-
tion of a fired direct-injection transparent engine. The
authors analyzed the flow field just before fuel injection
and up to the early stages of flame propagation.
PIV may be implemented with different seeding parti-
cles or tracers (for an overview, see, for example, Reeder
et al. (2010)). Many studies applied liquid droplets as
tracers, for example, Muller et al. (2010) and Fajardo et al.
(2006) used oil droplets with a size of *1 lm. As an
advantage, liquid tracers are not abrasive in the engine. As
a drawback, liquid droplets evaporate at larger temperature.
This makes measurements at firing top dead center and
during the combustion process very difficult. Application
of solid particles avoids the latter problem. Towers and
Towers (2004) seeded the flow with micro balloons, which
exhibit a large light scatter efficiency compared to oil
droplets. However, these particles show an abrasive
behavior during engine operation and reduce the barrel
lifetime significantly. Also another typical solid tracer TiO2
shows an abrasive behavior. Solid lubricants, for example,
boron nitride and graphite, overcome this problem. They
are non-abrasive and suitable for in-cylinder applications
(Neubert et al. 2011).
In our work, presented in the following sections, we
apply graphite particles as tracers in high-speed PIV. This
enables PIV measurements during all phases of the engine
cycle in motored operation, as well as outside the flame
area in fired operation. Our aim is to analyze the evolution
and cycle-to-cycle variations of the in-cylinder flow field.
In particular, we perform measurements in fired operation
and investigate the effect of the flow field upon the com-
bustion process.
2 Experimental setup
Figure 1 depicts the experimental setup of the high-speed
PIV system applied to the optically accessible gasoline
engine. The single-cylinder engine includes 4 valves, an
82 mm bore and an 86 mm stroke, yielding a displacement
volume of 454 cm3. The compression ratio is 9.6. For the
experiments discussed in the following, the engine speed is
set to 1000 rpm, and the inlet manifold pressure is
450 mbar. The engine is equipped with a horizontally
dividable twin inlet manifold. By closing the lower half of
the inlet manifold, we induce a distinct tumble motion. The
engine offers optical access to the combustion chamber by
a cylindrical quartz glass ring below the cylinder head, two
pent roof windows, located at the front and back side of the
engine, as well as a quartz glass disk inserted into the
piston. Operation of the engine is possible either at mo-
tored or fired condition. For fired operation, we replace the
spark plug (located between the exhaust valves) by a
Nd:YAG laser ignition system operating at 1064 nm, pro-
viding laser pulses with nanosecond duration. The laser is
focused by a lens with a focal length of 15 mm and
propagates through an entrance window into the combus-
tion chamber. The laser-induced plasma is generated
approximately 6 mm behind the entrance window. There-
fore, the ignition location is identical to the spark plug
ignition. We note that the ignition location is highly
reproducible from cycle to cycle. The lifetime of the laser-
induced plasma is orders of magnitude shorter compared to
the spark plug ignition. Thus, the plasma is not signifi-
cantly deflected by the flow field. Moreover, the ignition
location is optically accessible and no shadowing effects
from the spark plug mass electrode perturb the measure-
ments. In order to collect data under realistic engine con-
ditions, we use ambient air and a reference fuel (CEC-RF
08-A-85) as cylinder charge. The fuel injector is located at
the inlet manifold 80 cm upstream of the intake valves.
Due to the rather long path toward the combustion cham-
ber, this setup yields an almost perfectly homogeneous
mixture—avoiding effects of an inhomogeneous fuel dis-
tribution on the combustion process. We monitor the
pressure variation in the engine during individual com-
bustion cycles by a pressure detector (Kistler, 6043Asp).
The high-speed PIV system consists of a frequency-
doubled, double-cavity Nd:YAG laser (Edgewave, IS6II-
DE) operating at 532 nm. The laser pulse duration is of the
order of 10 ns, with a maximum repetition rate of 10 kHz.
The pulse-to-pulse separation of the two cavities can be
matched to the engine speed between 15–60 ls. We focus
Fig. 1 Experimental setup of high-speed PIV in an optically
accessible engine. For better visibility in the two-dimensional sketch,
the camera is drawn parallel to the plane of the light sheet—though
indeed it is oriented perpendicular to the light sheet
1702 Exp Fluids (2012) 53:1701–1712
123
the laser beam by a cylindrical concave (f1 = -20 mm)
and a cylindrical convex (f2 = 750 mm) lens to generate a
light sheet with a thickness of 1 mm and width of 60 mm
in the combustion chamber. The light sheet is located in the
center of the combustion chamber (i.e., between the two
intake valves) and oriented perpendicular to the axis of the
tumble motion.
The graphite tracer particles (Thielmann, colloidal
graphite 43019/a) are provided by a powder disperser
(Palas, RBG1000 D) in the air flow at the inlet manifold.
The particles exhibit a mean diameter of approximately
3 lm and survive temperatures up to 900�C. Therefore, the
particles do not survive the flame itself. But outside the
flame and in motored conditions, we do not observe any
burn of the particles. Due to (graphite) particle deposition
on the inner cylinder surfaces, the combustion chamber is
cleaned after every second engine run.
The light scattered by the tracer particles is detected on a
high-speed CMOS camera (Photron, SA1.1), oriented
perpendicular to the laser beam direction. An optical
bandpass filter with center wavelength at 532 nm ± 10 nm
serves to suppress background radiation. The frame rate
of the camera is 6 kHz at a spatial resolution of
960 9 704 px. At an engine revolution speed of 1000 rpm,
the temporal resolution of the camera corresponds to an
image taken every 2� of the crank angle.
To process the PIV image data, we use a commercially
available software (LaVision, DaVis8). As an additional step
of pre-processing, we subtract a sliding background (spatial
average of the vicinity) with a length scale of 150 px from the
PIV raw images and carry out an intensity normalization. We
apply a multi-pass cross-correlation algorithm to obtain
velocity fields, resulting in final interrogation windows of
32 9 32 px with 50 % overlap. To obtain clearer data, we
remove inconsistent vectors (i.e., differing substantially
from the median of the adjacent four vectors). In this case, we
also check the next best (second, third or fourth) correlations
between the images. If one of these correlations leads to a
better agreement with the local flow field, we use the
velocities from this correlation peak instead. The velocity
field finally obtained after processing has a size of
60 9 42 mm with a resolution of 1.4 9 1.4 mm.
3 Results and discussion
For a first impression of the in-cylinder flow dynamics,
Fig. 2 shows a sequence of ensemble-averaged velocity
fields under motored conditions. We recorded the flow field
for the intake and compression stroke from 360� before top
dead center (BTDC) until firing top dead center (TDC),
with a resolution of 2� CA for 20 consecutive cycles. We
record an air flow into the cylinder, when the intake valves
open around 330� BTDC. At 284� BTDC, the flow reaches
largest velocities of up to 50 m/s. Since the lower half of
the inlet manifold is closed, the flow is guided above the
intake valves, leading to a strong tumble inside the com-
bustion chamber. Between 238� BTDC and 192� BTDC,
the piston moves further downwards, and the tumble center
moves below the field of view. The intake valves close now
and the velocity of the flow reduces to values around 20
m/s. At 146� BTDC, the intake valves are fully closed and
the piston moves upwards again. The flow field shows a
large tumble structure, which spans over the entire com-
bustion chamber and exhibits a tumble center slightly
below the field of view. In the later compression phase (i.e.,
from 100� BTDC to 54� BTDC), the tumble moves
upwards and is squeezed vertically, while the magnitude of
the velocity field stays rather constant around 15 m/s.
Finally at 8� BTDC (i.e., typical ignition timing), the
tumble structure of the flow field is still visible.
A common way to quantify the flow field utilizes the
kinetic energy per unit mass Ekin within the light-sheet
plane (Druault et al. 2005):
Ekin ¼Xnx
i¼1
Xny
j¼1
1
2
u21ðxi; yjÞ þ u2
2ðxi; yjÞnxny
:
with the numbers nx and ny of vectors in x and y-directions,
the velocities u1(xi, yj) and u2(xi, yj) in x and y-directions at
the point (xi, yj). We note that we consider only the
velocity components within the light-sheet plane and
neglect the out-of-plane velocity. Since we induce a dis-
tinct tumble motion, the out-of-plane velocity (i.e., swirl
motion) is much lower than the velocity within the plane.
This was also shown by Muller (2012) in an identical
optical engine by means of stereoscopic PIV. Therefore,
the contribution of the out-of-plane velocity to the kinetic
energy can be neglected.
Figure 3 shows the kinetic energy in the flow field vs.
the crank angle for data from Fig. 2. At the beginning of
the intake stoke (360� BTDC), the kinetic energy is close
to zero—as the flow field is at rest. When the intake valves
open (i.e., from 350� BTDC to 290� BTDC), the kinetic
energy increases to a maximal value of approximately 250
m2/s2. A specific ‘‘staircase’’ shape is visible between
330� BTDC and 310� BTDC. By looking at the vector
fields, it is noticeable that the tumble structure is formed
directly after the ‘‘staircase’’ (i.e., 310� BTDC). From
260� BTDC on the kinetic energy decreases until
160� BTDC, when it reaches values around 50 m2/s2.
During this time, the intake valves close. From 160� BTDC
until shortly before firing TDC (i.e., at 20� BTDC), the
energy stays on a constant level around 50 m2/s2. At the
end of the compression stroke, the kinetic energy decreases
Exp Fluids (2012) 53:1701–1712 1703
123
x position [mm]
-30 -20 -10 0 10 20 30-30 -20 -10 0 10 20 30
45
30
15
0
45
30
15
0
45
30
15
0
45
30
15
0
y po
sitio
n [m
m]
y po
sitio
n [m
m]
y po
sitio
n [m
m]
y po
sitio
n [m
m]
x position [mm]
45
30
15
0
velo
city
[m/s
]
15
10
5
0
velo
city
[m/s
]
15
10
5
0
velo
city
[m/s
]
15
10
5
0
velo
city
[m/s
]
30
20
10
0
30
20
10
0
15
10
5
0
15
10
5
0
330° BTDC
8° BTDC54° BTDC
100° BTDC146° BTDC
238° BTDC 192° BTDC
284° BTDC
piston
piston
piston
1704 Exp Fluids (2012) 53:1701–1712
123
slightly. Since the combustion process takes place around
TDC, this particular timing is of high interest for com-
bustion diagnostics. Therefore, we will analyze now the
flow field around TDC in more detail.
Figure 4 shows the evolution of the ensemble-averaged
flow field around firing TDC for motored engine operation.
We measure the flow field of 36 individual engine cycles
from 100� BTDC to 100� after TDC (ATDC) in intervals
of 2� CA. At the beginning of the sequence (i.e., at
20� BTDC), a strong tumble motion with velocities up to
15 m/s shows up. The tumble structure is still visible at the
end of compression (TDC). However, at this time,
the velocity decreases to 10 m/s. During the expansion, the
tumble motion decays completely. While at 10� ATDC, a
weak tumble structure is still visible, it disappears com-
pletely at 30� ATDC. We also note that the flow field slows
down significantly toward velocities of only 4 m/s at 30�ATDC.
The decay of the tumble structure can be quantified by
the kinetic energy and the vorticity x. The latter is defined
via the rotation of the velocity field x = r 9 u (Adrian
et al. 2000). Since we record the flow field in x and y-
directions, our analysis yields the z component of x. We
also average the vorticity over the entire vector field. Fig-
ure 5 shows the kinetic energy and the vorticity of the flow
field for the PIV data depicted in Fig. 4. From the data, the
decay of the tumble flow structure becomes very obvious.
Between 40� BTDC and 20� BTDC, the kinetic energy
decreases only slightly. From 20� BTDC, the kinetic energy
drops significantly to values around 20 m2/s2 at TDC. This
corresponds to 1/3 of the energy remaining in the flow field
compared to the compression phase. We assume that the
‘‘lost’’ energy dissipates into small-scale dynamics in the
flow field, which enhance turbulent flame propagation
(Arcoumanis and Whitelaw 1987; Towers and Towers
2004). The resolution of PIV is not sufficient to resolve such
small-scale dynamics. Moreover, the kinetic energy might
also move outside the light-sheet plane or to the z-compo-
nent of the velocity field. Around 20� ATDC, the kinetic
energy reaches values close to zero, concluding the decay of
the tumble structure. The evolution of the vorticity shows a
similar behavior (see Fig. 5b). The vorticity stays on a
constant level until 20� BTDC and drops afterward. At
TDC, it exhibits a value of x = 600 s-1, which corre-
sponds to more than half the value in the compression
phase. From TDC onwards, the vorticity decreases further,
until at 20� BTDC it reaches values close to zero. Our
analysis of the kinetic energy and vorticity confirms the first
impressions of the velocity field sequence (see Fig. 4). The
tumble structure, generated during the intake stroke, starts
to dissipate at 20� BTDC and is completely vanished at
20� ATDC. At TDC and typical ignition timings (i.e.,
5–15� BTDC), the large-scale tumble structure is still
present in the flow field.
3.1 Cycle-to-cycle fluctuations
The previous results dealt with ensemble-averaged data of
the flow field. However, the flow field also exhibits strong
variations from cycle to cycle. To analyze this effect, we
perform a further PIV measurement and increase the
number of recorded cycles to 75. In return, we shorten the
recording range to 100�BTDC—TDC. The recording
interval is kept constant at 2� CA. Figure 6 shows the
evolution of the flow field of three individual engine
cycles. We randomly choose the cycles (14, 25, 47) from
the set of data and depict the flow field in these cycles at
90�, 60� and 30� BTDC. All three cycles show a distinct
tumble structure. However, the single cycles vary strongly
in the specific shape of the tumble flow and the location of
the tumble center. In cycle 14, at 90� BTDC, the tumble
center and the center of the combustion chamber coincide.
During compression, the tumble shifts upwards toward the
pent roof. At 30� BTDC, the tumble shows a symmetric
shape centered to the combustion chamber. In cycle 25, at
90� BTDC, the tumble center is located in the lower left
corner of the combustion chamber. While the piston moves
upwards, the tumble moves toward the center. Finally, the
tumble is vertically squeezed and an elliptical shape of the
tumble occurs at 30� BTDC. In cycle 47, at 90� BTDC,
the tumble center is slightly left-shifted with respect to the
cylinder axis. During the compression phase, the center
moves to the upper right corner of the cylinder. At 30�
0
100
200
300
400
° BTDC
Eki
n [m2/s
2]
0501001502002503003500
2
4
6
8
10
valv
e lif
t [m
m]
Ekin
valve lift
Fig. 3 Kinetic energy in the flow field vs. crank angle for intake and
compression stroke. The solid, blue line depicts ensemble-averaged
kinetic energy (20 cycles). The gray background indicates the range
of the standard deviation. The dashed, black line shows the lift of the
intake valves
Fig. 2 Temporal evolution of the ensemble-averaged flow field
during the intake and compression stroke, recorded by high-speed
PIV. To indicate the geometry, the figure also shows the regions of
the piston and intake valve. The velocity is color-coded (see legends
on the right side of the images). We note that the velocity scale
changes in the sequence from image to image
b
Exp Fluids (2012) 53:1701–1712 1705
123
BTDC, the center is located at the right side of the com-
bustion chamber. Beyond cycle-to-cycle variations of the
tumble structure, we also monitor variations in the mag-
nitudes of the velocity vectors. While cycle 14 indicates
maximal velocities in the range of 23 m/s at 30� BTDC, in
cycle 25, the maximal velocities are lower (i.e., around 18
m/s) at the same time.
In the following, we quantify the cycle-to-cycle varia-
tions of the flow field using the kinetic energy and the
spatial location of the tumble center. Figure 7 shows the
kinetic energy of the flow field for 75 individual engine
cycles (red circles) and the ensemble average (blue line) for
the same data set as depicted in Fig. 6. Since the three
cycles in Fig. 6 are chosen randomly, they show no
extreme behavior in kinetic energy and are rather close to
the ensemble-average, which reproduces the trend of the
data already illustrated in Fig. 5. The averaged kinetic
energy stays on a constant level (60 m2/s2) between 60�BTDC and 20� BTDC. Afterward, it decreases to values of
20 m2/s2 at TDC. The spread in kinetic energy of indi-
vidual cycles is indeed large. At 60� BTDC, the spread
ranges from 40 m2/s2 to 80 m2/s2, corresponding to a
standard deviation of ±14 % with regard to the ensemble
average. During the compression, the relative fluctuations
even increase. At TDC, the standard deviation is ±25 %.
The kinetic energy of individual cycles ranges from 7 m2/s2
up to 36 m2/s2.
We focus our attention now to the determination of the
tumble center—which can be used to quantify the cycle-to-
cycle variations of the large-scale tumble structure. We
determine the tumble center (as proposed by Jackson et al.
(1997)) by the position (x, y) in the flow field with a
maximal value of the rotation of the normalized velocity
field bx � r� ujuj.
Figure 8 depicts the position of the tumble center in the
engine for the data set of individual cycles, as already
introduced above. The tumble motion of the ensemble-
averaged flow field is shown as a black line, recorded from
100� BTDC to TDC. Additionally, we plot the tumble
center location of our 75 individual engine cycles for three
different crank angle positions (i.e., 90� (blue triangles),
60� (red squares) and 30� BTDC (green circles)). During
the period from 100� BTDC to 20� BTDC, the tumble
center of the ensemble-average moves diagonally from the
bottom left to upper right part of the engine. At 20�BTDC, the center moves slightly in upper left direction
toward the middle of the combustion chamber. The tumble
centers of individual cycles spread significantly around the
15
10
5
0
9
6
3
0
4
2
0
4
2
0
velo
city
[m/s
]
6
4
2
0
velo
city
[m/s
]
12
8
4
0
velo
city
[m/s
]
45
35
25
y po
sitio
n [m
m]
45
35
25
y po
sitio
n [m
m]
45
35
25
y po
sitio
n [m
m]
x position [mm]-20 -10 0 10 20 -20 -10 0 10 20
x position [mm]
30° ATDC20° ATDC
10° ATDCTDC
10° BTDC20° BTDC
Fig. 4 Temporal evolution of the ensemble-averaged flow field around firing TDC, recorded by high-speed PIV
1706 Exp Fluids (2012) 53:1701–1712
123
ensemble average. Inspecting the path of individual cycles
(not shown in the figure), it is noticeable that for most of
the cycles, the tumble centers proceed in parallel to each
other. However, also a crossing of the individual path can
be observed sometimes. At 90� and 60� BTDC the tumble
location of individual cycles is quite uniformly distributed
around the ensemble-average position. The scatter region
has a diameter of around 20 mm, corresponding to a
quarter of the combustion chamber diameter. At 30�BTDC, the distribution becomes elongated in horizontal
direction, yielding a displacement range of 34 mm hori-
zontally and 9 mm vertically. Thus, both the kinetic energy
in the flow field and the large-scale flow field structure (i.e.,
the tumble) vary substantially from cycle to cycle.
3.2 Fired engine operation
To analyze the effect of the cycle-to-cycle variations in the
flow field upon the combustion process, we operate the
engine now in fired condition. In the compression phase,
we record the flow field by high-speed PIV until ignition.
After ignition, we evaluate the combustion properties by
thermodynamic pressure analysis. The time when 50 % of
the mass fraction is burned (MFB50) is used as a parameter
to describe the time-scale of the combustion process. We
performed the measurement for 75 consecutive combustion
cycles with stoichiometric mixture (k = 1). The mean
effective pressure is IMEP = 2.7 bar, which corresponds to
lower partial load. The ignition timing is set to 6� BTDC
in order to achieve an optimal average MFB50 of
12.6� ATDC. To identify key parameters in our data set,
we consider the 10 cycles with lowest, middlemost and
highest MFB50 and average the flow fields from the cycles
in the sub-sets.
Figure 9 shows the kinetic energy of the flow field for
the averaged cycles with lowest, middlemost and highest
MFB50. In addition, the plot shows the kinetic energy
averaged over the complete ensemble (blue line). The
kinetic energy of the cycles with middlemost MFB50
(12.5�ATDC) coincides with the ensemble average. The
slow cycles (i.e., late MFB50 (15.1� ATDC)) hold for most
of the time a significantly lower kinetic energy compared to
the ensemble average. Beyond 20� BTDC, the energy
difference decreases, so that at ignition timing (6� BTDC),
the slowest cycles have almost the same energy as the
ensemble average. The fast burning cycles (i.e., early
MFB50 (10.7� ATDC)) hold the highest kinetic energy in
the flow field. Also here, at the end of the compression
phase, the difference in kinetic energy decreases toward the
value of the ensemble average. In summary, cycle-to-cycle
variations of the flow field significantly change the
dynamics of the combustion process. Fast burning cycles
show a considerably higher energy in the compression
phase. We assume that the energy is dissipated at the end of
compression toward small-scale dynamics in the flow field
(which are not resolvable by the PIV measurements). The
small-scale structures lead to enhanced flame propagation
velocities and therefore a faster combustion.
In order to confirm the effect of the flow field upon the
combustion process for individual cycles, we average the
kinetic energy of individual cycles in the time frame of
60�–20� BTDC and plot it vs. the corresponding MFB50
(Fig. 10). In addition, the figure shows a linear fit to the
data and the coefficient of determination. Figure 10a shows
the data for stoichiometric operation (k = 1). We observe a
distinct trend toward faster combustion (i.e., lower
MFB50) with increasing kinetic energy of the flow field.
The coefficient of determination is R2 = 0.59, which
confirms the effect of the flow field upon the combustion
dynamics. Thus, cycle-to-cycle variations of the kinetic
energy in the flow field are clearly a source of fluctuations
in the combustion process at stoichiometric operation.
Figure 10b shows the data for lean operation (k = 1.5).
The manifold pressure is kept constant at 450 mbar. Due to
the lean combustion, the mean effective pressure reduces to
IMEP = 2 bar. Moreover, the flame propagation is slower.
To compensate this, the ignition timing is shifted to
18� BTDC in order to maintain the optimal mean MFB50
0
20
40
60
80E
kin [m
2/s
2]
TDC
(a)
−40−2002040
0
400
800
1200
ω [1
/s]
° BTDC
(b)
Fig. 5 Evolution of the kinetic energy a and vorticity b around firing
top dead center. The solid, blue line depicts ensemble-averaged
kinetic energy. The gray background indicates the range of the
standard deviation
Exp Fluids (2012) 53:1701–1712 1707
123
of 11.9� ATDC. In this case, we observe a slight trend
toward faster combustion (i.e., lower MFB50) for cycles
with higher kinetic energy in the flow field. The scatter of
01020304050600
20
40
60
80
° BTDC
Eki
n [m2/s
2]
Fig. 7 Evolution of the kinetic energy for individual combustion
cycles (red circles). The solid, blue line depicts ensemble-averaged
kinetic energy. The gray background indicates the range of the
standard deviation
cycle 14 cycle 25 cycle 4715
10
5
0
velo
city
[m/s
]
20
15
10
5
0
velo
city
[m/s
]
x position [mm]x position [mm]x position [mm]
y po
sitio
n [m
m]
45
30
15
0
y po
sitio
n [m
m]
y po
sitio
n [m
m]
90° BTDC 90° BTDC 90° BTDC
60° BTDC 60° BTDC 60° BTDC
30° BTDC 30° BTDC 30° BTDC
-30 -20 -10 0 10 20 30 -30 -20 -10 0 10 20 30 -30 -20 -10 0 10 20 30
45
30
15
0
45
30
15
0
20
15
10
5
0
velo
city
[m/s
]
piston piston piston
Fig. 6 Flow fields of three individual combustion cycles (labeled with numbers 14, 25, 47) from a set of 75 cycles, at timings 90�, 60� and
30� BTDC, recorded by high-speed PIV
−30 −20 −10 0 10 20 300
10
20
30
40
x position [mm]
y po
sitio
n [m
m]
ensemble average90° BTDC60° BTDC30° BTDC
Fig. 8 Motion of the tumble center in an ensemble-average flow field
(black line). Position of the tumble center for individual cycles at 90�(blue triangles), 60� (red squares) and 30� BTDC (green circles).
Data determined from a set of 75 individual cycles, recorded by high-
speed PIV
1708 Exp Fluids (2012) 53:1701–1712
123
the data around the regression line is quite large, and the
coefficient of determination is lower (R2 = 0.26) compared
to the case of stoichiometric operation. Thus, we suspect
that the kinetic energy in the flow field has a smaller effect
on the combustion dynamics at lean operation. To deter-
mine the source of the larger fluctuations in lean operation,
we imaged the combustion flame simultaneously with PIV.
3.3 Simultaneous PIV and flame imaging
To image the flame propagation in parallel to PIV, we
slightly modified our experimental setup and replaced the
optical bandpass filter by a long-pass filter with cutoff
wavelength at 515 nm (Schott, OG515). The filter still
protects the camera from the bright flash of the ignition
plasma (i.e., mainly in the ultraviolet spectral region), but
transmits enough radiation from the flame to detect a
distinct flame contour. We reduce the field of view of the
camera to 32 9 16 mm and the spatial resolution to
704 9 368 px. This permits an increase of the frame rate to
16 kHz, corresponding to one image of the flow field taken
every 0.8� crank angle. The frame rate is sufficiently fast to
resolve the dynamics of the flame propagation. We record
the flow field and flame propagation in the time interval
from ignition to 20� ATDC for 75 individual combustion
cycles. For the calculation of the flow field, we mask the
flame area by application of an upper intensity threshold.
Figure 11 shows the evolution of the flow field in three
individual combustion cycles (18, 51, 58), which represent
examples for a slow, medium and fast combustion. The
measurement is performed at lean operation (k = 1.5). The
ignition timing is set to 22� BTDC and the mean MFB50 is
11.6� ATDC. In all cases, a clear vortex structure is vis-
ible. However, the local flow field (recorded at a timing of
20� BTDC) near the flame kernel varies strongly from
cycle to cycle. As the combustion process proceeds, the
flame kernel follows the motion of the in-cylinder flow
field. In cycle 18, the flow field pushes the flame kernel
leftwards to the chamber wall. Under these conditions, the
flame propagation is retarded due to heat losses at the wall
(i.e., quenching) and the fact that the flame front cannot
propagate isotropically in all directions. The characteristics
of this delayed combustion are visible in the pressure trace
and the late MFB50 of 24.0� ATDC. In cycle 51, the flame
kernel moves to the middle of the combustion chamber and
is afterward carried upwards by the tumble structure. At
this point, the flame can propagate isotropically in all
directions. Thus, we get a more regular flame propagation.
The MFB50 of 11.1� ATDC is close to the ensemble
average of 11.6� ATDC. Cycle 58 shows a flow field
pointing downwards at the ignition location. Therefore, the
flow field pushes the flame kernel straight toward the pis-
ton. On the piston surface, we get a stagnation point and the
flame kernel splits into two section. One part of the flame
kernel moves to the left and the second one to the right.
102030405020
30
40
50
60
70
80
90
° BTDC
Eki
n [m2/s
2]
ensemble averagelowest MFB50middlemost MFB50highest MFB50
Fig. 9 Kinetic energy for fired operation (k = 1). Combustion cycles
with lowest (green squares), middlemost (red triangles) and highest
(black diamonds) MFB50, as well as ensemble-averaged kinetic
energy (solid, blue line)
30 40 50 60 70 80 90 100
8
10
12
14
16
18
MF
B50
[° A
TD
C]
Ekin
[m2/s2] Ekin
[m2/s2]
(a)
30 40 50 60 70 80
5
10
15
20
25
MF
B50
[° A
TD
C]
(b)R2 = 0.59 R2 = 0.26
Fig. 10 MFB50 determined from pressure traces versus kinetic
energy of the flow field (averaged between 60–20� BTDC) deter-
mined by PIV, for a stoichiometric operation (k = 1) and b lean
operation (k = 1.5). The blue circles depict the data of individual
cycles. The dashed, black line shows a linear fit
Exp Fluids (2012) 53:1701–1712 1709
123
Afterward, both parts propagate independently from each
other. Due to this specific situation of two flame sections
(i.e., a large flame area), the combustion process becomes
very fast. This is also visible from thermodynamic pressure
analysis, which yields an early MFB50 of 8.9� ATDC.
However, we note that when using a glass piston as in the
experiments discussed here, the heat losses at the piston are
lower than with a standard metal piston. Higher heat losses
might retard the flame propagation, which would attenuate
the effect of the larger flame area.
Figure 12 shows the flame propagation and flow field for
stoichiometric operation. Also in this figure, three individual
combustion cycles (42, 54, 12), representing a slow, medium
and fast combustion, are shown. The ignition timing is set to
7� BTDC, which results in a mean MFB50 of 12.5� BTDC.
Due to the faster combustion at stoichiometric operation, the
time interval between the images is shortened. During the
combustion process, the influence of the macroscopic flow
field is much lower compared to lean operation (compare
Fig. 11). In none of the cycles, a significant motion of the
flame kernel can be observed. We only notice a minor
deformation of the flame by the flow field. In cycle 12 (fast
combustion, MFB50 = 9.1� BTDC), the flame is slightly
stretched by the rotation of the tumble vortex. In cycle 54
(medium combustion, MFB50 = 12.1� BTDC), this effect is
even weaken and in cycle 12 (slow combustion,
MFB50 = 15.5� BTDC), no deformation of the flame ker-
nel occurs. The deformation of the flame may lead to a faster
combustion due to larger flame area. For stoichiometric
operation, this effect seems to be secondary, since the kinetic
energy is the major source for cycle-to-cycle fluctuations in
the combustion process—as shown in Fig. 10a. Anyhow, the
macroscopic influence of the flow field might explain the
slight scatter of the data in Fig. 10a.
We conclude that for stoichiometric operation, the
macroscopic motion of the flame kernel is of minor
importance. Since the flame propagation is faster, the time
slot for macroscopic variations in the flame kernel is sig-
nificantly shorter. Moreover, the ignition timing is closer to
TDC, when the kinetic energy (and therefore also the
velocities) of the flow field is lower (see Fig. 5a). For lean
operation, the velocity distribution in the flow field clearly
determines macroscopic motion of the flame kernel. This
can either lead to advantages for flame propagation (e.g., a
20
15
10
5
0
velo
city
[m/s
]
45
40
35
30y po
sitio
n [m
m]
x position [mm]x position [mm]x position [mm]
y po
sitio
n [m
m]
y po
sitio
n [m
m]
y po
sitio
n [m
m]
20° BTDC CDTB°02CDTB°02
CDTB°61CDTB°61CDTB°61
12° BTDC CDTB°21CDTB°21
8° BTDC CDTB°8CDTB°8
cycle 18 “slow” cycle 51 “medium” cycle 58 “fast”
-20 -10 0 10 -20 -10 0 10 -20 -10 0 10
45
40
35
30
45
40
35
30
45
40
35
30
Fig. 11 Simultaneous imaging of flame propagation and flow fields of three individual combustion cycles (labeled with numbers 18, 51, 58)
from a set of 75 cycles, at timings 20�, 16�, 12�, and 8� BTDC, recorded by high-speed PIV, at lean operation (k = 1.5)
1710 Exp Fluids (2012) 53:1701–1712
123
large flame area due to fragmentation) or to disadvantages
(e.g., flame quenching near the walls). This effect seems to
be more important than the influence of the kinetic energy
upon the combustion process—and hence may explain the
large scatter of the data in Fig. 10b.
4 Conclusions
We investigated the effect of the in-cylinder flow field upon
combustion dynamics in an optically accessible gasoline
engine by high-speed PIV with solid particles (i.e., graphite)
as tracer. The solid tracer particles permit measurements at
firing top dead center and during the combustion process. In
particular, we monitored the in-cylinder flow field (i.e., the
velocity distribution) for the intake and compression strokes
of the engine. As a prominent feature in the flow field, we
observed and studied the dynamics of a large-scale tumble
structure—which is still present at typical ignition timings.
Moreover, we observed large cycle-to-cycle fluctuations of
the flow field, which we quantify by variations in the kinetic
energy and the location of the tumble center. We conducted
measurements in fired operation to determine the effect of
such fluctuations upon the combustion dynamics. As a small
extension of the PIV setup, we simultaneously imaged the
motion of the combustion flame. At stoichiometric operation
of the engine, cycle-to-cycle variations of the combustion
process are mainly due to variations in the kinetic energy of
the flow field. The macroscopic motion of the combustion
flame by the flow field is of minor importance at stoichi-
ometric operation—as the flame propagation is faster and the
time for macroscopic variations of the flame is shorter. At
lean operation, the flow field induces a considerable, mac-
roscopic motion of the flame kernel—which significantly
effects the combustion process. Thus, at lean operation, it is
mainly the large-scale flow structure, which leads to cycle-
to-cycle variations in the combustion process—while vari-
ations of the kinetic energy play a minor role in this case. In
summary, high-speed PIV of the in-cylinder flow field serves
as a valuable tool to determine and quantify sources for
cycle-to-cycle fluctuations in the combustion process of a
gasoline engine.
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