The effects of injection pressure on swirl and flow pattern in diesel combustion

Abstract Single cylinder and optical engine measurements were performed for different in-cylinder

flow conditions. Swirl and tumble were varied, and emission data, together with optical

evaluation of in-cylinder flow field during the injection event and during the after-oxidation

event, were measured. Particle image velocimetry (PIV) software was used to evaluate the

combustion pictures and to calculate the flow field in the cylinder with crank angle

resolution during combustion. The glowing soot particles from the combustion were used as a

tracer. Single cylinder tests with an active valve train were used, which gives a variation in

swirl number, between 0.4 to 6.7, and tumble, between 0.5 to 4.0.

The main conclusion from this work is that the injection pressure strongly

affects the flow field in the cylinder, both before and after the oxidation period. The big

swirling vortex survives the combustion. At different injection pressures the angular velocity

at different positions in the bowl differs, and the difference increases with injection pressure.

Introduction The importance of spray, mixture formation and after oxidation in the diesel combustion

process is well established. Many diesel engines have a tangential swirl motion to improve

combustion. The swirl flow has been proven to reduce particulate matter (PM) emissions

from the engine and much research on swirl flow has been carried out for a long period of

time [7]. Lately, the main focus of the research has been on injection systems with extremely

high injection pressures [8], and, in many cases, with a quiescent combustion chamber (with

no, or nearly no, swirl in the cylinder). The higher injection pressure improves droplet break-

up air/fuel mixing in the spray [1] and increases turbulent intensity in the combustion

chamber. This is of paramount importance, especially during an engine transient, when the

combustion system needs to handle low air/fuel ratio (λ) conditions without producing

extremely high PM, especially for engines without diesel particulate filters (or other PM-

reducing after-treatment). With swirl, the PM can be greatly reduced [2] during a transient,

even when high injection pressures are used.

The two main flow structures in the cylinder are swirl and tumble. Swirl

number (SN) and tumble number (TN) are defined at bottom dead centre(BDC) (calculated

with GT-POWER, SNgt and TNgt) as:

Engine

SwirlgtSN

(1)

Engine

TumblegtTN

(2)

where

Swirl Air angular rotational velocity around the cylinder centre axis.

Tumble Air angular rotational velocity perpendicular to the cylinder axis.

Engine Engine angular rotational crankshaft velocity.

To create an understanding of how the in-cylinder flow behaves, PIV measurements are

commonly used in both constant flow rigs and motored engines. In [5], PIV measurements

were done from inlet stroke until close to firing TDC and compared with CFD calculations on

the same engine geometry as used in this paper.

In [2], tests with up to 2,000-bar injection pressure at λ 1.2 and a load of 10-bar

indicated mean effective pressure (IMEP) was performed. The emissions were greatly

affected when SN and TN were changed. To examine why the emissions were so greatly

affected at different flow structures, optical engine tests were performed in [3]. PIV software

was used to calculate the flow characteristics from the pictures taken in the optical engine

after the injection had ended. The SNgt was varied from 2.3 to 6.3 and the TNgt from 1.0 to

2.2. The results from [3] show that the injection pressure also affects the flow structure in

the cylinder during the after-oxidation part of the combustion. With higher injection pressure

the angular velocity in the inner bowl region is significantly higher than the outer bowl

region. To further understand why the emissions were affected and the angular velocity

profile was so greatly affected, complementary optical measurements and single-cylinder

engine tests were carried out. The injection event and the after-oxidation event were

examined from the optical pictures, using the same technique as shown in [3]. The optical

measurements were then combined with single-cylinder engine tests with emission results

for different SNgt and TNgt. The aim of this work is to show how the in-cylinder flow pattern

affects engine emissions.

Test setup The test points of interest were chosen from a measured engine load transient at 1,000 rpm.

The load transient, from 3- to 23-bar IMEP, was performed on a six-cylinder engine with a

similar combustion system as in the optical engine. The three points of interest, seen in

Table 1, were investigated further in both single-cylinder engine, with an active valve train

(Lotus AVT) and in an optical engine with the same boundary conditions as in the six-

cylinder engine, but they were tested in steady-state conditions. The workflow and used test

engines/equipment can be seen in Figure 1.

Load point 1 is the same as tested in [2] and corresponds to a maximum natural

aspirated engine load, ~10-bar IMEP, with injection pressures from 500 to 1,500 bar, with

500-bar increments, and λ = 1.25. The fuel mass was kept constant by changing the injection

duration for the different injection pressures. The crank angle for maximum cylinder

pressure (APmax) was kept constant at 11° ATDC by adjusting the start of injection (SOI).

The SNgt was varied from 0.4 to 6.7 by using the AVT system in the single-cylinder engine. In

the optical engine the SN and TN was varied by using two different cylinder heads and by

blocking (or not) one inlet port. To maintain constant λ, the boost pressure was adjusted

slightly to maintain the same air mass flow into the engine when the inlet geometry was

varied.

Load point 2 was performed at 2,500-bar injection pressure and a higher load,

~20-bar IMEP, and λ 1.1. The load represents when the before the boost pressure reaches

steady-state conditions on a normal turbocharged DI engine. All load points in this work

were performed at an engine speed of 1,000 rpm.

Figure 1 Overview of the equipment used.

Calculations of swirl, tumble and turbulent intensity were performed in GT-POWER. GT-

POWER data for airflow calculations was measured in a constant flow-rig, where swirl and

tumble were measured as a function of valve lift. For more details, see [2].

Table 1 Selected load cases with boundary conditions. SNgt and TNgt are calculated using GT-POWER.

The optical- and single-cylinder engines used in this study are based on a Scania engine

geometry and injection system capable of 2,500-bar injection pressure, see Table 2. In Figure

2, the optical engine can be seen with the different tested piston bowls. On the original

piston, a piston extension is mounted that leads to the optical piston and the liner that it is

fitted into. A high-speed colour camera, Phantom v7.3, is installed next to the engine and the

combustion light is transferred to the camera by a mirror mounted in the piston extension.

For more information on the optical engine, see [4].

Two different-shaped piston bowls were tested in this study: one with a flat

glass and one with the original bowl-shaped glass, as tested in the single-cylinder engine.

The glass is mounted in a titanium piston that transfers the combustion pressure on the

glass to the piston extension. The titanium piston restricts the vision field to a diameter of

80mm. To be able to vary the in-cylinder airflow, two different cylinder heads were used in

this study (named low- and high-SN head). Each head was tested with one or two operating

inlet ports, called v1 and v2, respectively, in this report. When using one port, the SNgt and

TNgt increased compared with the results that were achieved when both ports were

activated.

Load 1 a,b Load 1c Load 2

Rail press. [bar] 500, 1000 1500 2500

SNgt (GT-POWER) at BDC 2 - 6.3 0.4 - 6.7 0.4 - 6.7

TNgt (GT-POWER) at BDC 1 - 2.2 0.5 - 4.0 0.5 - 4.0

SOI [° ATDC] -11°, -6° -4° -2°

IMEP [bar] 10 10 20

Flow -

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13 Angle on cylinder head

Tum

ble

val

ue

Valve lift [mm]

One valve tumble 6-8

4-6

2-4

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GT-POWER

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Lotus

AVT

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Inlet pressure [bar]

Load 1

Load 3

Load 2

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PIV DaVis

Emissions

data combined

with flow data

Swirl

Tum

ble

Injection pressure 1000 bar, load 1b

0 1 2 3 4 5 6 7

0.5

1

1.5

2

2.5

3

3.5

4 Smoke [FSN]

one valve

two valves

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

Combine

data

Figure 2 Principal layout of the optical engine, on the left. On the right, the two tested piston-bowl shapes with plotted schematic spray.

Table 2 Engine specifications for optical- and single-cylinder engine.

Optical evaluation method on combustion pictures

The pictures from the combustion were evaluated using PIV software, DaVis 7.2, where

velocity vector fields were extracted from the moving combustion cloud. By comparing two

pictures at a time, the glowing particles are traced using cross-correlation. The light from the

glowing particles is what the PIV software traces, and no form of extra seeding or laser layer

is used. In this way, the movement in the x-y plane can be traced with information from not

just one thin layer, but from a line of sight. This is a advantage and a disadvantage: it is hard

to determine which height the traced particles are actually on, but it gives more complete

information on flows that are not just acting on one thin layer. This method cannot evaluate

the total velocity vector, since the z-axis is not included. As will be shown later, the results for

velocity vectors are quite stable, indicating that enough glowing soot particles remain in

sight between the pictures. The total resolution of the pictures from the camera is 256x256

pixels, with a colour depth of 14 bit. The time step between every picture, Δt, was set to 28

µs, which means 0.17 CAD at 1,000 rpm. In Figure 3, the principle of the picture evaluation

can be seen. Every picture was divided into 16x16 (or less for diffusion-flame figures) pixels

integration windows in which a mean velocity vector

22yx VVV

(6)

was calculated between two pictures (at t and t + Δt) for every integration window. For the

next picture evaluation another picture pair were evaluated. A film with velocity vectors was

therefore created. To reduce the error reading from the pictures, a DaVis built-in median

filter and sliding average filter were used to fill up missing or erroneous vectors in the

Test engine 4-stroke Scania D12

Bore/stroke [mm] 130/154

Connecting rod [mm] 255

Compression ratio 17.3:1

No. of valves 4

Injection system Scania common rail XPI

Injector holes 8

Spray angle [deg]

(° between cyl.head and spray)

Injector hole diameter

(inner/outer) [mm]

Max Injection pressure [bar] 2500

16

0.187 / 0.163

pictures. The missing or erroneous vectors mainly came from limited numbers of discernible

particles in the area that could be tracked by DaVis. In Figure 3, the result of two evaluated

pictures taken with 0.17 CAD between the exposures can be seen. The vector field, as seen, is

what the observer can see of the movement, and the traceable particles are well captured by

the PIV program.

To reduce the influence of potential pressure fluctuations, which might give

velocity vectors that are pointed in the direction of the pressure wave, an average vector field

was created from 20 pictures. The time interval between 10 frames represents one period of

the observed pressure fluctuation, and by using the average result, this error can be

neglected when pictures with velocity vectors are plotted. The pressure oscillations are

created by the premixed combustion, and at long ignition delay this can give large pressure

oscillations that disturb the data evaluation. In an optical engine with single combustions,

the ignition delay is longer than for a continuously firing engine, since no residual gases

remain in the cylinder and the surfaces in the combustion chamber have lower temperatures,

which increases the ignition delay, see [2].

From the PIV software, the velocity vectors were exported to MATLAB, where

further calculations were made. To calculate the SN from the velocity vectors some

assumptions were needed. The tangential part from the geometrical centre, seen in Figure 3,

of every velocity vector was estimated to give a contribution to the total rotational velocity in

the cylinder. The tangential part of the velocity vector is calculated with:

i

iz

r

ree

(7)

where

(0,0,1) axiscylinder thealongr Unit vectoze

z)y,(x, coordinate vector radial ir

The angular velocity is then calculated:

i

i

ir

eu

(8)

where

i angular velocity [rad/s]

iu Velocity vector (ux, uy, uz)

, where uz was set to zero, due to no z-axis information (2-D camera shots).

Angular velocities above 6,000 rad/s and below -1,000 rad/s were assumed to be unrealistic

and excluded from the SN calculations. From the angular velocities in the cylinder, the mean

SN can be calculated by:

nSN

engine

n

i

i

1

(9)

where vectorsspeedangular ofnumber n

When the bowl-shaped piston is used, SN is also calculated with respect to the evaluated

volume in the respective integration window:

totengine

n

i

ii

V

V

SNw

1

(10)

iV = Volume in evaluated integration window

n

i

itot VV1 (11)

Figure 3 Evaluated combustion pictures at 25.9° and 26.1° ATDC together with the calculated vector fields. In the vector field picture, the coordinate system is shown for SN calculation.

Results In Figure 4, three heat releases (HR) were plotted for load point 2 (20-bar IMEP) at different

SN. The HR was measured in a single cylinder metal engine with an active valve train, the

same engine as in [2]. This gave the opportunity to vary the SN in a wider range, by

implementing different valve profiles and strategies than were possible in the optical engine.

The same cylinder head was then used in the optical engine, but the combustion pictures

were only taken for SNgt 3.4 (standard valve profiles) in this case. The combustion picture

arrows mark on the green HR line where the pictures were taken. In this case, the piston

bowl-shaped glass was used together with a high-SN cylinder head. The injection event is

marked at the bottom of the figure. It is clear to see that the HR was affected by the different

SN at an injection pressure of 2,500 bar. In the first picture, where the combustion has just

started, yellow light can be seen on the leeward side of the spray. This picture has an

external light source (4 x 70w Xenon bulbs) outside the cylinder, to illuminate the sprays, but

all other pictures have only natural combustion light. In picture 2, all sprays have developed

a combustion flame that is slightly orientated to the leeward side of the sprays. In the next

two pictures the combustion has the highest intensity, according to the HR. Between the two

pictures, a figure with velocity profiles is shown, this is combustion picture 4, which has been

evaluated using the DaVis software. The spray creates lots of turbulent flow in the

combustion chamber, Figure 4. When the injection stops, the turbulence created from the

injection just survives some CAD before the swirl flow in the cylinder again dominates the

flow structure, and a mainly circular rotation can be seen on the pictures. The glowing soot

particles that are moving with the swirling vortex are bright and easy to follow even at this

late CAD.

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Pic 1: At 25.90° ATDC

Pic 2: At 26.07° ATDC

iu

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Figure 4 HR and combustion pictures at load 20-bar IMEP, 2,500-bar injection pressure. The HR are plotted for three different SN, GT-POWER definition, SNgt. The pictures are from the SNgt 3.4 case. The red arrows mark where on the HR it represents.

Emissions dependency on swirl and tumble

Single cylinder tests were performed at load point 1c (10-bar IMEP), with 1,500-bar injection

pressure, and λ was kept constant at 1.25 for the different valve profiles. Swirl and tumble

was varied and plotted together with soot emissions in Figure 5 (top picture). An optical

engine test was then performed for the same load point. The after-oxidation period was then

evaluated using DaVis software at 16.5 to 19.9 CAD, with injection stops at approximately 9

CAD, and plotted together with the emissions. The arrows on the emission plot mark where

the respective image corresponds to. In the top-right-hand corner of the vector figures, the

mean calculated SN for the picture is shown. To vary the SN and TN in the optical engine,

two different cylinder heads were used: low- and high-SN head. Used together with blocking

(or not blocking), one port gives four different SN and TN flows. This is because the optical

engine has a regular camshaft and the different valve profiles were not implemented in this

engine compared with the single-cylinder engine equipped with the AVT system. Calculations

of SNgt and TNgt were performed in GT-POWER (BDC for SN and TN). The low-SN head was

tested with a flat piston bowl and the high-SN head with a shaped piston bowl. The first

trend that we can see in Figure 5 is that higher SN radically decreases the soot emissions. If

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Lambda 1.1

Shaped piston

High SN-Head

Injection period

higher tumble is introduced the soot increases. For example, SNgt 2.5 is measured at two

different TNgt, where the high tumble case has nearly 50% more soot. Therefore, the SN

needs to be increased even more to lower the soot emissions if high TN is used. In [5] and [3]

it has been shown that high tumble disturbs the swirl and offsets the swirl centre. If the

different sprays, that are injected symmetrically, are exposed to asymmetric swirl flow, the

soot production is increased in those sprays that have an unfavourable airflow. The picture

with the lowest SNgt shows the most unstructured flow field, 7 CAD after injection, in Figure

5. The soot formation is at the highest level in this case (compared with the other showed

vector figures). When one port is deactivated, the next picture (mean SN 4.107), shows a

much more uniform flow structure. Also, the soot emissions decrease and the velocity

gradients increase with higher SN. The next two vector figures show uniform flow fields, but

the highest SN case also has the highest velocity gradients, and in this case the tumble is

higher, which creates the offset in the swirl centre. The high SN gives low soot emissions. To

quantify all the velocity vectors, the angular velocity (in the bottom-left diagram) and mean

velocity (on the right) are plotted against the radial position in the bowl, at 18 CAD ATDC.

The vectors that are at a certain radius are summed up and a mean value is calculated

(which is shown on the graphs). The first difference that can be observed between the curves

is that the velocity in the outer region of the bowl radius is more or less constant for each

piston bowl geometry, but there is a big difference between the geometries. Higher velocity is

shown in the shaped piston bowl case. The reason for this is that more of the piston bowl

volume is placed in the outer region of the bowl in the shaped piston case compared with the

flat piston, which has the same volume over the observed area. Compensations in SN

calculations for the shaped piston bowl are made later in this work. In the velocity profiles, it

is clear to see that when the SN is increased by port deactivation in the two-cylinder head

configurations, both the angular velocity and vector velocity increase in the middle of the

piston bowl. The velocity in the outer bowl region remains nearly the same for the same

piston configuration.

Figure 5 Emissions at different SNgt (named swirl in graph) and TNgt after oxidation velocity vector pictures, angular velocity and velocity at different piston radius plotted for load 1c, 1,500-bar injection pressure. All velocity figures are plotted at 18 CAD. The red boxes on the emission graph indicate where the two optical engine tests are performed with flat piston and low-SN head. Those tests are not repeated exactly in the single-cylinder engine, as in the bowl-shaped piston case and high-SN head.

Swirl

Tum

ble

Injection pressure 1500 bar, load 1c

0 1 2 3 4 5 6 7

0.5

1

1.5

2

2.5

3

3.5

4 Smoke [FSN]

one valve

two valves

0

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Om

ega [

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2v 1500bar,Flat pist,Low SN-head,SN=2.6682

1v 1500bar,Flat pist,Low SN-head,SN=4.1073

2v 1500bar,Bowl pist,High SN-head,SN=5.1262

1v 1500bar,Bowl pist,High SN-head,SN=5.7798

0 5 10 15 20 25 30 35 40 450

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Velo

city [

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low-SN-head

Shaped piston,

high-SN-head

Diffusion flame dependency of swirl

In Figure 6, the diffusion flame is traced with the PIV software for two different SN. By

deactivating one port, the SNgt increases from 2.0 to 3.5 (top, low SN and bottom picture pair

high SN). The marked area in the right picture corresponds to the evaluated area shown in

the figure on the left. The evaluation is done over 2 CAD, so the flow observed is a mean

value. The stagnation point, where the spray hits the flat piston glass, can be clearly seen

(marked with a red arrow). The combustion flame is then divided and the recirculation zone

is created around the combustion flame. The global swirl vortex is moving in a clockwise

direction. The combustion flame is affected by this, with longer velocity vectors on the

leeward side of the flame indicating a higher velocity in this direction (maximum observed

velocity of 44 m/s). Effects from the nearest combustion flame on the windward side of the

evaluated spray can be seen. Velocity vectors that are pointing in the direction of the swirl

tell us that mass transport from the nearest spray interfere with the evaluated spray.

When the SN is increased, shown in Figure 6 (bottom picture), the velocity

vectors on the windward side decrease and more flow from the adjacent combustion flame on

the windward side is clearly observed. More flow can be detected in the inner region of the

bowl, which indicates that more fuel is in the centre region of the bowl. The highest observed

velocity is still on the leeward side of the evaluated combustion flame, but the maximum

velocity is decreased to 37 m/s (white arrow in the picture) compared with 44 m/s in the two-

port case (low SN). With one port (increased SN), the penetration depth is decreased. The

strong side wind that is acting on the spray forces the injected fuel to bend away from its

initial direction. More fuel is therefore left in the inner region of the bowl and the

spray/combustion flame that hits the piston wall has less kinetic energy. The observed

velocity vectors that have been reflected in the piston are lower compared with the low-SN

case. Because of this, deflected velocity vectors are higher for the low-SN case.

Figure 6 Low-SN head and flat piston with both inlet ports activated (SNgt 2 in GT-POWER) in the top pictures and one inlet port activated (SNgt 3.5) in the bottom pictures at 8 CAD ATDC. Load 1c (10-bar IMEP) 1,500-bar injection pressure.

The flames in the bowl shaped piston are shown at relative low-injection pressure, 500 bar,

in Figure 7. The spray stagnation point is harder to see, see the red arrow. However, the flow

structure behaves differently compared with the flat piston. The schematic differences in

spray behaviour are shown in Figure 2. The flame hits the glass in the flat piston case and is

divided into two observable flames. In the shaped piston bowl case, the flame is also divided

in a similar way, but the flame follows the glass shape in a “twisting-like motion”, see Figure

7. The global swirl vortex forces the combustion cloud to rotate in the same clockwise

direction as the swirl by itself. The two forces, from the spray and from the global swirl flow,

together with the round shape of the outer piston bowl, force the combustion flame flow to

twist, as observed in the figure. This gives good mixing during the injection event in the

outer bowl region.

Emission measurements show that this particular injection pressure and

engine speed, 3.4 SNgt (two port), gives a lower amount of smoke emissions compared with

6.3 SNgt (one port). When the SN increases, shown in Figure 7 (bottom pictures), the

observed twisted flow in the outer bowl region is not as well structured as in the two-port

case. The flow on the leeward side increases and the flow is more tangential in the outer

piston bowl, with just a little twisted flow left. The stagnation point for the spray cannot be

seen. The mixing during the injection event is probably not as favourable compared with the

two-port case for this particular injection pressure and engine speed. The highest observed

vector velocity for the one-port case is 28 m/s, which is observed on the leeward side of the

spray. This is slightly higher compare to the two port case, as seen in Figure 7, with 25 m/s

on the leeward side. The reason why we observe the opposite result in this case, compared

with the flat piston bowl, is that the swirling flow velocity is higher and the geometry is

different. The injection pressure is also lower compared with the flat piston case. The black

ring seen in the right-hand picture is the sharp bowl radius, where the con of the middle part

of the bowl meets the circle formed “ditch” in the outer part of the bowl, see bowl geometry in

Figure 2.

Figure 7 Bowl-shaped glass piston with high-SN head, from 4.0° ATDC to 9.5° ATDC, and 500-bar injection pressure.

Flow pattern dependency from injection

At increased injection pressure, the SN can be increased further for smoke reduction, as

observed in [2]. In Figure 8, three different injection pressures are shown, 500 bar, 1,000 bar

and 1,500 bar. The observed flow field during the injection period indicates an increased

vector velocity. Higher injection pressure gives increased kinetic energy in the spray and,

therefore, a higher velocity. The following pictures show testing with a high-SN head, two-

port operation and shaped piston bowl. Observe the different velocity scales between spray

pictures and after-oxidation pictures. The calculated SNw that is printed in the top right-

hand corner of the pictures is the weighted SN calculated by using the velocity vectors of

each respective figure. The three tested cases have the same boundary conditions before

injection, load case 1. The calculated SNw is higher in 1,000-bar and 1,500-bar cases

compared with the low-pressure case. A big difference is observed in the after-oxidation part.

The low injection pressure case has significantly lower SNw, 3.6 compared with 4.4 and 4.6

for the higher injection pressures. The difference in velocity gradients increases with

injection pressure. In the high-injection pressure cases (1,000 and 1,500 bar), the earlier

discussed twisting motion in the outer part of the piston bowl (Figure 7) can be easily

observed. These cases also have lower smoke emissions compared with the low-pressure case.

In all of the cases, the eight combustion flames, caused by the eight injection nozzles holes,

can be observed in the pictures. In the high-injection pressure case, the recirculation zones

and the interaction between the sprays can be clearly observed, marked by a red arrow in the

figure. The interactions between the sprays creates a recirculation zone, and number of

sprays (distance between the sprays) influence the lift-off length [9] and thereby the air

entrainment in the spray.

The high velocities, which are created at higher injection pressures, increase

the turbulence in the cylinder, but, as seen in the after-oxidation picture on the right, the

small vortexes are also quickly reduced when the injection event is stopped, nearly

independent of injection pressure. The big-scale swirl vortex is the dominating flow just a few

CAD after injection. A small-scale turbulence vortex has a small timescale and is quickly

transformed into smaller scales, and in the end to heat, when the vortex length scale is the

same as the viscous length scale [6]. Even if the small-scale turbulence is transformed, a

difference in velocity is still observed in the after-oxidation figure (on the right-hand side).

The difference in velocity also increases with injection pressure. The flow can be expected to

exist in the z-plane due to the difference in velocity seen in the 2-D plane in the figures.

To understand how much the injection pressure affects the velocity gradients that are

observed in the earlier figures, points are plotted which show the sum of angular velocity

corresponding to the respective radial position in the piston bowl. This is shown in Figure 9

for injection pressures 500 to 1,500-bar and load case 1, 10-bar IMEP. Angular velocity is

plotted in the two top graphs, on the left, at 16.5 to 19.9 CAD, and on the right, at 26.5 to

29.9 CAD. The bottom picture shows the weighted SN plotted against engine CAD for the

three different injection pressures and a mark where the two pictures above are represented

on the graph.

The first observation is that the angular velocity in the inner bowl region is

much higher compared with the outer region and strongly dependent on injection pressure.

In the outer bowl region, marked with a green dotted ellipse, it seems to have the same

angular velocity in all cases. As seen in Figure 5, with varying SN and piston geometry, the

velocity in the outer bowl region seems to be affected by bowl geometry and not as much SN

and injection pressure, as seen here in Figure 9. The middle region of the bowl differs

significantly in angular velocity at different injection pressures. With higher angular velocity

gradients the vorticity increases and thereby the mixing increases, which is beneficial during

the after-oxidation part of the combustion.

10 CAD later (top right-hand figure), the angular velocity gradients decrease in

all cases, but the cases with highest gradients decrease the most. The reason for this is due

to the fact that the high gradients drop from a higher level, which gives higher flow losses.

The angular velocity decreases in the centre region of the combustion chamber and increases

slightly in the outer piston bowl region. The shear stress through the velocity profile,

together with the turbulence eddies, had time to equalise at the different radii due to

dissipation.

The SNw is affected by the injection pressure, seen at the bottom of Figure 8.

The 500-bar injection pressure case has a much lower SNw compared with 1,000- and 1,500-

bar cases. The injection ends at 9 CAD and the signal, especially for 1,500 bar, is quite noisy.

Before 9 CAD it can be hard to see the trend as after 30 CAD in the 1,500-bar case, where

the glowing soot particles start to be hard to follow – (better after-oxidation at higher

injection pressures). It is also observed that the SNw seems to be accelerated at higher

injection pressures after the injection has ended.

Figure 8 Injection event plotted from 4.0° ATDC to 9.0° ATDC in the left-hand pictures. The after-oxidation period in the right-hand pictures is a mean value between 16.5-19.9° ATDC. Injection ends at 9.0° ATDC. Observe different velocity scales between the injection and after-oxidation event.

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Inj. 1000bar

Inj. 500bar

Figure 9 Angular velocity profile versus radial position in the shaped piston bowl at three different injection pressures in the figures at the top. SNw is plotted against CAD in the bottom figure. The position where the

mean velocity profiles are calculated is marked on the graph. The mean value is calculated over 3.4 CAD, from 16.5 to 19.9 CAD for the first case and 26.5 to 29.9 CAD for the second case.

When the load increases to 20-bar IMEP together with higher injection pressure, in this case

2,500 bar, even higher velocities are observed in the outer part of the piston bowl during the

injection event, see Figure 10.

How can we explain the difference in angular velocity at different radii when

the injection pressure is changed? Here is one suggestion to an answer: when fuel is injected,

it is affected by the clockwise swirling airflow and the spray starts to curve, see Figure 10.

When the flame reaches the outer bowl region, gas is transported back to the middle by the

piston bowl geometry and the force of the injection. Some of the flame and gas is reflected

with an offset compared with where the fuel is injected from the injector holes. The swirling

flow influences the reflected flame/gas and the offset increases. When the spray reaches the

inner bowl region, the other seven sprays, which also are reflected and have an offset,

cooperate and speed up the velocity of the swirling gas in the middle of the bowl. Higher

injection pressure gives a higher velocity of the reflected spray, shown in this work. The

offset of the spray, together with higher spray velocity, contributes to increase the SN in the

middle of the combustion chamber, as shown schematically in Figure 10. The top pictures in

the figure show EOI when the injection stops and the swirling flow starts to strongly affect

the flames in the middle of the bowl. The increased swirl velocity can be easily observed

when the flame in the last picture is not as straight as the earlier picture, when the injection

is still occurring.

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SN

w [

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1000bar inj.p

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Figure 10 Injection pressure, 2,500 bar, with schematic flow lines (from injection) in the bottom picture. The top pictures show EOI where the injection stops and the swirl flow “overtakes” the earlier strong injection.

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CONCLUSIONS

This work has shown an optical investigation of in-cylinder flow during the injection event

and after-oxidation period together with emission and heat release comparison at different

in-cylinder flows.

1. Injection pressure has a strong influence on the degree of irregularity on the swirl

vortex in the after-oxidation part of the combustion. Higher injection pressure means

increased kinetic energy and greater differences in angular velocity in the combustion

chamber. This increases the vorticity and the mixing of residuals during the after-

oxidation part of the combustion.

2. There is an observed difference in flow behaviour when a flat piston bowl is used

compared with a shaped piston bowl.

3. The inner bowl region velocity is affected by SN and injection pressure. The outer

bowl region velocity is affected by piston bowl geometry.

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Off Region of a Heavy Duty Diesel Engine Using Raman Spectroscopy”, SAE 2009-01-1357.

2. Dembinski, Henrik and Ångström, Hans-Erik. “An Experimental Study of the Influence of Variable In-

Cylinder Flow, Caused by Active Valve Train, on Combustion and Emissions in a Diesel Engine at Low λ

Operation”, SAE 2011-01-1830.

3. Dembinski, Henrik and Ångström, Hans-Erik. “Optical study of swirl during combustion in a CI engine

with different injection pressures and swirl ratios compared with calculations”, SAE 2012-01-0682.

4. Lindström, Mikael and Ångström, Hans-Erik. “A Study of In-Cylinder Fuel Spray Formation and its

Influence on Exhaust Emissions Using an Optical Diesel Engine”, SAE. 2010-01-1498.

5. Nordgren, Henrik; Hildingsson, Leif; Johansson, Bengt; Dahlén, Lars; Konstanzer, Dennis. “Comparison Between In-Cylinder PIV Measurements, CFD Simulations and Steady-Flow Impulse Torque

Swirl Meter Measurements”, SAE paper 2003-01-3147.

6. Pope, Stephen B. “Turbulent flows”, Cambridge University Press, 2000, ISBN 978-0-521-59125-6.

7. D. E. Winterbone; J. H. Horlock and Rowland S. Benson, “The thermodynamics and gas dynamics of

internal-combustion engines”, Oxford: Clarendon Press, 1982, ISBN 0-19-856210-1.

8. Mahr, B. “Future and Potential of Diesel Injection Systems” THIESEL 2002 Conference on Thermo- and

Fluid-Dynamic Processes in Diesel Engines

9. Polonowski, Christopher; Mueller, Charles and et al. “An Experimental Investigation of Low-Soot and

Soot-Free Combustion Strategies in a Heavy-Duty, Single-Cylinder, Direct-Injection, Optical Diesel

Engine”, SAE. 2011-01-1812