0

Flow measurements using combustion image

velocimetry in diesel engines

HENRIK W. R. DEMBINSKI

Licentiate thesis

Department of Machine Design Royal Institute of Technology

SE-100 44 Stockholm

TRITA MMK 2012:03 ISSN 1400-1179

ISRN/KTH/MMK/R-12/03-SE

1

TRITA MMK 2012:03

ISSN 1400-1179 ISRN/KTH/MMK/R-12/03-SE

Flow measurements using combustion image velocimetry in diesel engines

Henrik W. R. Dembinski

Licentiate thesis

Academic thesis, which with the approval of Kungliga Tekniska Hgskolan, will be presented

for public review in fulfilment of the requirements for a Licentiate of Engineering in Machine

Design. The public review is held at Kungliga Tekniska Hgskolan, Brinellvgen in 83, room B319 Gladan, 1th of March at 10:00.

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Abstract This work shows the in-cylinder airflow, and its effects on combustion and emissions, in a

modern, heavy-duty diesel engine. The in-cylinder airflow is examined experimentally in an

optical engine and the flow field inside the cylinder is quantified and shown during

combustion, crank angle resolved. Cross-correlation on combustion pictures, with its natural

light from black body radiation, has been done to calculate the vector field during the

injection and after-oxidation period. In this work, this technique is called combustion image

velocimetry (CIV). The quantified in-cylinder flow is compared with simulated data, calculated

using the GT-POWER 1-D simulation tool, and combined with single-cylinder emission

measurements at various in-cylinder airflows. The airflow in the single cylinder, characterised

by swirl, tumble and turbulent intensity, was varied by using an active valve train (AVT),

which allowed change in airflow during the engines operation. The same operation points

were examined in the single-cylinder engine, optical engine and simulated in GT-POWER.

This work has shown that the in-cylinder airflow has a great impact on emissions and

combustion in diesel engines, even at injection pressures up to 2,500 bar, with or without

EGR and load up to 20-bar IMEP. Swirl is the strongest player to reduce soot emissions.

Tumble has been shown to affect soot emissions negatively in combination with swirl.

Tumble seems to offset the swirl centre and the offset is observed also after combustion in

the optical engine tests. Injection pressure affects the swirl at late crank angle degrees

during the after-oxidation part of the combustion. Higher injection pressure gives a higher

measured swirl. This increase is thought to be created by the fuel spray flow interaction. The

angular velocity in the centre of the piston bowl is significantly higher compared with the

velocity in the outer region of the bowl. Higher injection pressure gives larger difference of

the angular velocity.

Calculated swirl number from the CIV technique has also been compared with other

calculation methods, GT-POWER and CFD-based method. The result from the CIV

technique are in line with the other methods. CFD-based calculations, according to [62], has

the best fit to the CIV method. The GT-POWER calculations shows the same trend at low

swirl number, but at high swirl number the two methods differs significantly.

3

Acknowledgements This thesis presents the results of the project Transient Air Management, commissioned by

Scania CV AB, the Royal Institute of Technology (KTH) and the Swedish energy agency. I

want to thank my supervisor, Prof. Hans-Erik ngstrm, my colleague at KTH and Scania.

My family, for the support during this period of hard work.

4

List of publications I. 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

Operation. SAE paper: 2011-01-1830. Henrik W. R. Dembinski, Hans-Erik ngstrm.

II. Optical study of swirl during combustion in a CI engine with different injection

pressures and swirl ratios compared with calculations. SAE paper: 2012-01-0682.

Henrik W. R. Dembinski, Hans-Erik ngstrm.

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

Submitted to: International Journal of Engine Research IJER-12-0006. Henrik W. R.

Dembinski, Hans-Erik ngstrm.

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Contents Abstract ................................................................................................................................. 2

Acknowledgements ............................................................................................................... 3

List of publications ................................................................................................................. 4

Introduction ........................................................................................................................... 8

The basic principle of the diesel engine ............................................................................. 8

Historical view ...............................................................................................................10

Diesel engine combustion .............................................................................................12

Basic turbulent flows .....................................................................................................17

Swirl and tumble flow ....................................................................................................20

Squish flow ....................................................................................................................22

Emissions from diesel engines ......................................................................................24

Engine transients ...........................................................................................................27

Project motivation .............................................................................................................29

Method..............................................................................................................................29

Methodology .........................................................................................................................30

Test equipment .................................................................................................................31

Single-cylinder engine with an AVT system ...................................................................32

Optical engine ...............................................................................................................33

Steady-state flow rig ......................................................................................................34

Cylinder head design .....................................................................................................36

Theoretical models and calculation methods ........................................................................37

Calculation of SN and TN ..............................................................................................37

Velocity measurement techniques ....................................................................................39

Optical evaluation method on combustion pictures ...........................................................40

Calculation of a CAD-resolved vector field.....................................................................41

Calculation of a CAD-resolved SN .................................................................................42

Calculation of a mean vector field ..................................................................................45

Calculation of angular velocity profile inside the piston bowl ..........................................45

Results .................................................................................................................................47

Swirl and tumble effects on smoke emissions ...................................................................47

Flow pattern during after-oxidation and its affect on smoke emissions ..............................50

Swirl and tumble effects on NOx and CO emissions ......................................................52

Lambda and SN dependency on emissions ......................................................................53

Airflow effects on combustion ...............................................................................................53

Ignition delay.....................................................................................................................55

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In-cylinder airflow pattern .....................................................................................................57

Piston bowl and flame interaction ......................................................................................57

Flow pattern and swirl at late CAD ....................................................................................59

The injection effect on flow pattern and swirl .....................................................................62

Comparison of CIV method and other calculation methods for SN ....................................68

Conclusions ..........................................................................................................................69

Future work ..........................................................................................................................70

References ...........................................................................................................................71

Appendix 1: CO and NOx emissions ....................................................................................75

Appendix 2: Ignition delay ....................................................................................................77

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8

Introduction The compression-ignited engine, named the diesel engine, after its inventor Rudolf Diesel

(born in Paris 18 March 1858 and died 29 September 1913), has been the main power

source used in heavy-duty vehicles for a long period of time. The reason is mainly

economical, as the diesel engine is more efficient, and historically cheaper fuel together with

higher durability compared with its competitor, the otto engine (also named after its inventor,

Nicolaus Otto, born in Holzhausen 14 June 1832 and died 26 January 1891).

Over the decades, since the diesel engine was patented in 1894, many improvements to the

engine have been carried out. However, there are still many challenges left to deal with.

Emission legislation is increasingly strict, together with higher demands on engine efficiency.

This equation is not very easy to solve. Higher engine efficiency does not automatically mean

decreased emissions; rather the opposite is true with todays engines. A better

understanding of the processes in the engine is one key to solving the emission/efficiency

equation. This work is just one small piece of a giant puzzle that many scientists around the

world are trying to piece together.

The basic principle of the diesel engine

The basic principle of the compression ignition (CI) engine is to increase the temperature of the air in the cylinder, typically in the order of 1,000 K, well over the self-ignition temperature of the fuel, and then inject the fuel. In this way, the combustion can be controlled in a better way compared with the premixed spark ignited (SI) engine. The temperature increase is caused by the increase in the cylinder pressure during compression. In a pressure volume (PV) diagram, see Figure 1, the operation of a CI engine can be observed. The PV diagram comes from the original patent by Rudolf Diesel and shows, in principle, the operation of the engine. When the pressure and the temperature increases, due to the change in cylinder volume (from 1 to 2 in diagram), the fuel is added to the compressed air in the cylinder [1]. The fuel injection maintain until 3 in the original patent. The maximum pressure starts to drop under the expansion and the cylinder pressure does not increase more than the compression pressure. In a modern engine, the PV diagram looks like Figure 2. The cylinder pressure greatly increases during combustion. In the original patent, the aim was to also use water injection. According to the patent, the exhaust temperature can be reduced to below the inlet temperature(!) of the engine by carrying out mechanical work on the crankshaft. The engine should not need any external cooling either. Neither of these two statements has been proven to work in reality, but the diesel process does work and powers many different applications.

Figure 1. Pressure and volume diagram from Rudolf Diesels patent [1]. Pressure is shown on the y-axis

and volume is shown on the x-axis.

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However, when the fuel is injected into the hot air in the cylinder, the relatively cold fuel starts

to increase in temperature and starts to evaporate. Energy is taken from the hot air in the

cylinder and the fuel starts to decompose into other chemical substances, for example

Hydroxide (OH). OH plays an important role during the ignition delay period (and also in the

after-oxidation period). Ignition delay period is the time between when the fuel has started to

be injected until when exothermal reaction is observed (typically on the cylinder pressure

trace). OH is a reactive substance that, at a certain critical mass, ignites the fuel. The ignition

delay in CI engines is often explained by the Arrhenius correlation, which takes the in-

cylinder average temperature and pressure, together with the activation energy of the fuel,

into account to calculate the ignition delay [2]. However, as shown later in this work, other

things, such as airflow in the cylinder, affect the ignition delay. After the first premixed

combustion phase the diffusion combustion part takes over. The fuel is continuously injected,

mixed, transformed and combusted in a non-premixed diffusion flame. Temperature and

pressure rapidly increase in the cylinder. After some crank angle degree (CAD) the fuel

injection stops and, shortly afterwards, so does the diffusion flame period. The after-oxidation

combustion period takes then over, where the mix of air/fuel residuals that are in the cylinder,

and proceeding the combustion of unburned products. This process is maintained under the

expansion until the in-cylinder temperature decreases to a level where the exothermal

reactions stop, due to the chemical timescale being too long. In other words, the temperature

is too low for the after-oxidation to happen due to a decrease in cylinder pressure caused by

volumetric expansion (piston movement). When the volume reaches the highest value, the

exhaust valve opens and the piston lowers the volume again, resulting in the exhaust being

evacuated out from the cylinder. When the piston reaches the top dead centre (TDC), with

the lowest cylinder volume, the exhaust valve closes and the inlet valve opens. Fresh air is

sucked into the cylinder, typically at a slightly lower pressure compared with the exhaust

stroke. The cycle starts over again when the piston reaches the bottom dead centre (BDC)

and the inlet valve closes. In Figure 2, two load cases are plotted in the PV diagram, medium

and high load. As seen, the overall pressure is higher in the high-load case compared with

the medium-load case. In the high-load case, the inlet and exhaust pressure are increased

due to turbocharging. In this way, the air density is increased in the engine, more fuel can be

added and the load on the engine can be increased significantly. The drawback with

turbocharging is that the exhaust back pressure increases, which is caused by the exhaust

turbine that drives the turbo compressor. This means that the residual gases left in the

cylinder increase with the back pressure and reduce the volumetric efficiency. But the

increased inlet pressure more than compensates with increased trapped air mass in the

cylinder.

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Figure 2. PV diagram showing a four-stroke diesel engine operating on medium and full load.

Historical view

In the original patent from Rudolf Diesel, he describes that solid, liquid and gaseous fuels

can be used in his invention. The cooling of the engine was not required due to the controlled

combustion. He compared his patent with other internal engine patents (i.e. Ottos patent).

According to Diesel, the fast explosion that appears in premixed engines increases the heat

and pressure to extremely high levels, which leads to the problem of lubricating and cooling

the engine. Thanks to his more controlled way of introducing the fuel after the compression

and by introducing it over a controlled duration, he claimed he could have greater control

over the combustion velocity. With this control, the fuel is combusted with the same velocity

as it provides mechanical work on the flywheel. In this way, no extra heat is produced that

needs to be cooled away. In the patent, an insulating jacket may be provided outside the

cylinder liner. Rudolf Diesels original engine, which did not have a cooling function, was not

a good idea. He missed the fact that the heat transfer from the in-cylinder air to the engine

also acts during compression. The combustion also produces a lot of heat that dissipates into

the engine components and cannot be isolated. But still today some research has been done

on isolating components in the cylinder [3]. In Diesels patent, he also showed that water

injection could be used in the compression stroke to further increase the efficiency. He even

believed that the exhaust gases should, by introducing water during the compression stroke,

be cooler than the surrounding air. However, it has not been possible to show this in the

reality.

The first fuel injection system that Diesel showed in the first patent was based on coal

powder. In Figure 3, the mechanism is shown at four different positions, where the

0,0 0,1 0,2 0,3 0,4 0,5 0,6 0,7 0,8 0,9 1,0Volume

PC

YL1 [

bar]

0

20

40

60

80

100

120

140

160

180

200

Medium load, IMEP 11 bar

High Load, IMEP 20 bar

0,0 0,1 0,2 0,3 0,4 0,5 0,6 0,7 0,8 0,9 1,0Volume

PC

YL1 [

bar]

0

1

2

3

4

5

6

7

8

9

10

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combustion chamber is placed under the device and the coal is fed from above, at points a,

b, c and d. In the first position, the coal powder is fed into the notch. The notch is then

rotated down to position b, where the powder is introduced into the compressed air in the

cylinder (TDC combustion). The powder is pressurised and starts to combust when it comes

into contact with the compressed air. By controlling the amount of powder that is in contact

with air (the small throat area), the combustion velocity is thereby also controlled.

Figure 3 Rudolf Diesels patented injection system for coal powder.

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In the patent, a fuel system for liquid fuels was also described, where the fuel is introduced

into the cylinder by a higher pressure than the highest possible cylinder pressure. For this, he

aimed to use some kind of feed pump (not shown in this patent) that could pressurise the fuel

up to those levels. In his next patent [4], he developed the injection system further, especially

for liquid and gaseous fuels. He understood that mixing in a non-premixed combustion

process was important. Different layouts of the combustion chamber with different injection

systems are shown in this patent. The pre-chamber combustion geometry is introduced as

well as the first version of todays combustion bowl geometry for direct injected diesel

engines, Figure 4.

Figure 4 The first version of a piston bowl (K in Fig. 3) and the first pre-chamber diesel (Fig. 5) [4].

Today, liquid fuel is typically used in the diesel process, but natural gas can also be used in a

diesel engine. For example, the Westport dual fuel system [5], where a small pilot diesel

injection is injected (and combusted) and, shortly after, the main gas injection starts to be

injected.

Diesel engine combustion

Diesel combustion is a sort of non-premixed combustion where the fuel is injected, typically

in liquid phase directly in the preheated combustion air. The combustion process is divided

into three different stages, the premixed- (ignition delay), diffusion- and after-oxidation

period. The classical paper from Dec [6] shows how the process proceeds and where the

different emissions are created. This is a simplified conceptual model, used to simplify the

discussion about diesel combustion.

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Premixed combustion period

When the fuel is injected with high velocity into the hot air (which has a lower velocity),

droplet break-up occurs. The velocity difference increases the evaporation rate, and thereby

also the heat and air exchange around the droplet [7]. When enough fuel has evaporated

and mixed with oxygen at the right temperature, and the reaction time is long enough,

ignition of the fuel can occur.

The ignition delay in a diesel engine is commonly explained by the Arrhenius correlation [2].

In the correlation, the ignition delay is based on the chemical reaction time needed for the

fuel to evaporate, decompose and excite the activation energy ( AE ). The environmental

conditions that effect the reaction times is temperature (T) and pressure (p). A and n are

constants depending on fuel and, to some extent, the injection and airflow characteristics [2].

TR

EpA Anid exp (1)

where R is the gas constant.

The ignition of fuel is a complex chemical reaction with many different reaction steps.

Detailed reaction models for diesel-like fuels can be found in [8]. A model that takes the

turbulent mixing and history during self-ignition into account can be found in [9].

When ignition occurs, the premixed combustion phase starts. When the premixed fuel is

consumed, the combustion passes over to a diffusion flame period where the rate of heat

release (HR) is governed by the amount of fuel that is injected per unit of time. The

combustion flame starts to look something like the image in Figure 5.

Diffusion combustion period

The injected fuel core is expanding in the area when it propagates from the injector. Hot air is

mixed with the fuel and the fuel starts to evaporate and decompose. First, the fuel is

combusted at low with soot production as a result, the grey area in Figure 5. Air mixes into

the flame and the combustion propagates. The yellow area is typically what is easy to

observe when lots of black body radiation creates a bright yellow light. Most of the thermal

nitrogen oxide (NOx) is created on the surface of the flame, marked in green, where the

temperature is high and oxygen is located.

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Figure 5 Decs classical combustion plume [6].

In Figure 6, particle image velocimetry (PIV) measurements on the air entrainment in non-

burning spray, the instantaneous velocity field are showed. The spray area increases

downstream from the spray when air is mixed up with the fuel spray. Local vortices are

created around the spray and increase the air-mixing rate.

Figure 6 Fuel spray in a spray chamber that shows air entrainment into the spray visualised by PIV

measurements [10].

When the Dec model is compared with pictures from an optical engine, Figure 7, it can

clearly be seen that the geometry and airflow in the engine strongly affect the combustion.

When only one flame is injected it is easy to see how it splits into two halves when it reaches

the edge of the bowl. The swirl direction is marked in the picture and shows that the flame

15

that is on the leeward side of the spray moves in the direction of the swirl. The flame on the

opposite side travels some distance before it is moved back again in the direction of the

swirl. When eight sprays are introduced in the combustion chamber, Figure 7, the reflected

flames contribute and are reflected back to the bowl centre. If this geometry is done correctly,

the reflected flames form a recirculation zone in the bowl and improve the mixing of residual

gases, heat and unburned oxygen. This is of paramount importance for diesel combustion.

Figure 7 One combustion flame compared with eight combustion flames at 1,500-bar injection pressure.

The arrows show the swirl direction and the reflected combustion flames.

After-oxidation period

The after-oxidation period is of great importance to reduce smoke emissions and much (in

the order of 2545%) of the total HR is released during the after-oxidation part of the

combustion. At the end of injection (EOI), when the injection needle starts to close, the fuel

flow through the injection nozzle decreases with the fuel velocity. The fuel penetration is

thereby limited and the fuel spray creates less turbulence. The mixing with air is restricted

with higher soot production as a result, this is shown in Figure 8. When the spray velocity

decreases, the bright soot illumination is also seen in the centre of the bowl. When the

injection has ended, the bright illumination increases in the middle of the bowl. This bright

section in the middle of the bowl has a high soot content that needs to be oxidised. This

means that a fast needle closure is beneficial to reduce the late soot production.

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Figure 8. Combustion at 500 bar injection pressure, swirl number 6.3, end of injection period.

The soot and unburned residuals have time to be reduced during the after-oxidation period.

Increased airflow is beneficial during this period. The after-oxidation period is maintained as

long as the in-cylinder temperature is high enough. During the expansion stroke, the volume

increases, and thereby decreases the pressure and temperature in the cylinder. The time

window required to reduce the emissions can be affected by the injection timing. A later

injection gives a shorter time window for after-oxidation, with higher soot emissions as a

result.

This description of diesel combustion is valid for normal diesel combustion. Other types of

combustion modes that can be applied in a diesel engine are, for example, low-temperature

combustion modes. With high amounts of exhaust gas recirculation (EGR) and long ignition

delay, all of the fuel is injected into the cylinder typically before combustion. Therefore, the

fuel has a longer mixing time [11], and the rich fuel zones are thereby avoided with

decreased soot and NOx formations. Examples of these types of combustion modes can be

seen in [12], [13], [14], [15], [16], [17], [18] and [19]. One drawback with this type of

combustion is that it is best suited for low-load operation and is very difficult to control in a

fast engine transient. This is basically the reason why this is not used in this work, which

deals with fast engine transients and fairly high load.

17

Basic turbulent flows

Most of the flow in nature and engineering applications is turbulent. Turbulent flow is

especially the dominant flow in combustion engines. Turbulence can be observed all around

us. When we stir our drink or tap on a glass of water we can easily observe turbulence. The

cumulus clouds are turbulent, the boundary layer around a sailboat is turbulent and so on.

Without turbulence many things would not work. For example, the combustion in an SI

premixed engine would be too slow, with the result that the maximum possible engine speed

would be limited to a couple of revolutions per minute. Turbulence is characterised by its

irregular random nature with circular motions (eddies) in all three dimensions, see Figure 9. It

behaves randomly in time and space and always occurs at high Reynolds numbers (Re). Re

is defined by,

dudu

Re (2)

where is the dynamic viscosity, is the density, d is the pipe-flow diameter or a

characteristic length for the problem, is the free-stream velocity and is the kinematic

viscosity.

Figure 9. This picture is a side view of the large eddies in a turbulent boundary layer. Laser-induced

fluorescence is used to capture the quasi-periodic coherent structures. The flow is from left to right. Note

the reflection from the polished surface [40].

Turbulence cannot maintain itself; it depends on its environment to obtain energy. Turbulent

flows are generally shear flows. If the energy supply is shut off, the turbulence quickly

dissipates and is transformed to heat. In a diesel engine, where lots of both large eddy and

small eddy turbulence is created during inlet stroke, only the large flows can survive for a

longer time in the cylinder, and thereby has a chance to effect combustion, when the inlet

valves are closed (energy supply). The mean small eddy turbulence lifetime is much smaller

than the time for induction and compression [41].

The transition from laminar to turbulent flow does not have an exact critical Re number,

where the flow is either one or the other and the transition is poorly understood. The scientist

Osborne Reynolds (born 1842 and died 1912), who first experimented on high Re numbers

in pipe flow, showed a critical Re (where the flow becomes turbulent) at 13,000. His

experiments were repeated in the 1970s, in Manchester, where Reynolds original

experimental apparatus still exists. The critical Re number was found to be much less than

18

13,000. The reason can be found in the transition region, where the flow is very sensitive to

external disturbances, as the flow is going from laminar to turbulent.

The circular motions in the turbulent layer increase the mixing of the reactants and products

significantly, and thereby also combustion velocity. The oxidation, fuel pieces, radicals and

heat from the combustion are mixed and exposed to each other more rapidly compared with

laminar combustion. A wide range of length scales exists in turbulent flow, from the biggest

dimensions of the flow field to diffusive actions of molecular viscosity. The smallest scales,

called Kolmogorovs microscale, have relatively small timescales, which makes them quite

statistically independent of the big and relatively slow vortices. The large eddies lose most of

their kinetic energy with one turnover time to the smaller-length scales (called dissipation).

This means that turbulence is a strongly damped non-linear stochastic system [42].

A schematic sketch of a turbulent boundary layer near a wall is seen in Figure 10. From the

stochastic boundary layer profile, a time-average thickness profile can be plotted that

describes the mean thickness of the boundary layer, )(x . The instantaneous velocity, u ,

can be divided into two velocity components, mean (U or u ) and a fluctuating part (u ),

Uuu (also called Reynolds decomposition). In Figure 10, u is the fluctuating velocity

in x-direction, v in y-direction and w in z-direction (not shown in this picture).

At the wall, all the velocity components are zero, and outside the boundary layer the velocity

is the same as the free-stream velocity U . A time-average velocity profile for the turbulent

layer can be plotted that describes the behaviour of the mean velocity. Compared with the laminar case, the turbulent-velocity profile has a higher flow velocity near to the wall. The difference is explained by the fact that the transverse transport (transport in the y-direction) of momentum and vorticity in laminar flow is driven by the viscous shear stress in the fluid. In the turbulent case, the transverse transport is driven by convection, by the fluctuating turbulent eddies vortices (or, in other words, by the turbulence itself of the circular motions). This gives a higher velocity, momentum and friction near the wall in the turbulent case. The drag coefficient, Cd, is thereby also higher compared with the laminar case.

Figure 10. Turbulent boundary layer with time-average thickness, turbulent- and laminar-velocity profile

sketches.

The Reynolds decomposition:

iii uUu (3)

u

v y

x

U

19

The flow in engines in not stationary. But to simplify calculations in engine flow, the velocity

components can be defined as stationary for a small period of time (or for a small CAD

window). The mean velocity for stationary flows is defined as [42]:

Tt

ti

Ti dtu

TU

0

0

1lim (4)

and the mean value of the fluctuating part (velocity) is zero by definition; and can be written

as:

Tt

tii

Ti dtUu

Tu

0

0

01

lim (5)

In turbulent flows, the convection dominates over the molecular diffusion. Turbulence has

fast shifts in pressure and velocity. To describe the motion of the gas, two basic equations

are required: the mass conservation and momentum equation. First, we have the mass

conservation equation. If we think about a control volume, the mass in the control volume (

cvM ) is the mass transported into the volume minus the mass transported out of the volume:

in out

cv mmt

M (6)

The mass can also be described as velocity and density, which gives the mass conservation

or the continuity equation [43]:

0

U

t

(7)

where

3

1i ix (8)

The momentum equation (NavierStokes equation) is based on Newtons second law and

relates the fluid particle acceleration to the surface forces and body forces. The Navier

Stokes equation is defined as [43]:

Up

Dt

DU 2 1

(9)

The NavierStokes equation, together with the continuity equation, describes the

conservation of mass, momentum and energy in a flow field. With the restriction of an

incompressible flow field, there is no need for the energy equation to describe the flow. In

compressible flow, like supersonic flow, or when heat transfer is involved, the energy

equation cannot be ignored.

When a velocity difference exists between the fluid particles, initiated by fluid motion, shear

stress is generated. When a particle with a mass is transported in a flow with different

velocity, compared with the particle, the shear stress occurs and wants to accelerate or slow

down the particle. The total mean shear stress for two dimensional flow, as showed in Figure

10, can be written as:

vu

y

U

(10)

20

Where vu is the turbulent stress (or Reynolds stress) and the other component is the

viscous stress. Reynolds stress is an internal stress that acts on mean turbulence flow. The

viscous stress is acting on the particle by the fluid viscosity.

When the turbulence dissipates, the large eddies lose their kinetic energy first to smaller-

length scales and then to heat. The rate of conversion of turbulence into heat by molecular

viscosity is called the dissipation rate, . At the smallest turbulent vortices, it is still the fluid

viscosity that dissipates kinetic energy to heat.

Vorticity is a fluid parcel with a tendency to rotate around itself. If the vorticity is zero, the

fluid parcel can still move in a curve, but it does not rotate around its own axis. Vorticity has

the dimension of frequency (sec^-1).

zyx eeeu

y

u

x

v

x

w

z

u

z

v

y

w (11)

Swirl and tumble flow

The airflows inside the cylinder are commonly characterised by swirl, tumble and turbulent

intensity. Swirl and tumble are large-scale vortices that can exist inside the cylinder, one at a

time or in combination. The vortices are created during the inlet stroke, when the piston

moves down and inlet air passes over the inlet ports. Depending on the inlet port design,

swirl and/or tumble vortices are created and conserved (or dissipate slowly) in the cylinder

when the inlet ports are closed. Unlike small-scale vortices, these big-scale vortices do not

dissipate so fast and survive a sufficient time and can thereby effect the combustion and

after-oxidation. Friction against cylinder walls and in the airflow by itself makes the swirl flow

to slowly dissipate, it is just a matter of time. Swirl is the angular velocity around the cylinder

centre axis and tumble is the perpendicular velocity to the cylinder axis. Both variables can

be normalised towards the engine crankshaft velocity and the dimensionless numbers, swirl

number (SN) and tumble number (TN), can be defined as:

Engine

SwirlSN

(12)

Engine

TumbleTN

(13)

where

Swirl Air angular velocity around cylinder centre axis.

Tumble Air angular velocity perpendicular to the cylinder axis.

Engine Engine angular crankshaft velocity.

When both vortices exist, they are combined to create one big resulting vortex. Typically,

swirl is used in DI diesel engines and tumble in SI engines. In diesel engines, the after-

oxidation part of the combustion is of great importance. To reduce soot emissions, swirl has

the opportunity to survive the compression and the combustion, and thereby affect the

combustion in the latter part of the engine cycle. With a variable valve train it is possible to

control SN and TN. Variable valve trains are in production in light-duty engines [20], [21], but

21

this is not yet common in heavy-duty diesels, although research and development is in

progress [22]. Port designs on diesel engines have historically been very important [23], [24].

SN has been an important factor in good combustion and low smoke at moderate injection

pressures. Recently, research on the injection system using higher injection pressure and

EGR has shown significantly decreased emissions. Although SN has a demonstrable effect

on emissions and combustion [25], it has not changed appreciably until the introduction of

todays high-injection pressures. Today, some manufacturers have nearly no swirl

(quiescent) in their combustion systems.

During the engine cycle, from that when swirl is created and the inlet valves are closed, the

swirl changes in velocity during compression and combustion. During compression, the swirl

velocity increase at late CAD when air flow is forced into the piston bowl. This spin-up is

created when the conserved momentum in the airflow is forced into smaller radius. When the

piston moves down again, the opposite happens. The flow is also slowing down due to the

friction against the cylinder walls. But if the swirl flow is assumed to be constant for a small

time window, the swirl vortex conservation can be observed if the continuity and momentum

equation, mentioned earlier, is rewritten to polar-cylindrical coordinates for the mean flow.

The details can be seen in [44]. The tangential equation from the momentum equation

describes the conservation of mean flow angular momentum in the flow. If a fluid is moved

outside its normal track around the cylinder axis (increased radius), where it maintains its

angular velocity in the swirl, it will have an angular momentum loss compared with the

surrounding fluid particles. The centripetal forces acting on the fluid element are lower than

the opposite net pressure force caused by the mean radial pressure gradient, with the result

that the fluid is forced into its normal equilibrium orbit. This demonstrates why swirling flow in

a cylinder can be stable and does not dissipate into smaller-scale turbulence as fast, as with

other types of turbulent flows, when the energy source is closed. With this above-mentioned

statement, it is easy to understand why swirling flows in an engine are often modelled as a

solid-body rotation. Even if PIV measurements show that a perfect solid-body rotation does

not exist [45], [46], this can be a decent assumption in modelling.

Tumble is of paramount importance for an SI engine to increase combustion velocity. The

tumble vortex is transformed to small-scale turbulence around TDC, due to the geometric

change of the combustion chamber during compression. Fuel and air are premixed before

combustion, and a spark plug ignites the mixture. The flame front propagates through the

premixed air/fuel. First, the flame propagation is laminar (around 0.3 m/s), but later it

becomes turbulent (in the order of 1080 m/s depending on for example engine speed), as

turbulence exists in the cylinder [2]. At TDC, the tumble vortex is transformed to small-scale

turbulence, which greatly influences the flame velocity. Without the turbulence, the flame

velocity should remain around laminar combustion velocities with a low flame speed, and

thereby unable to combust all of the mixture in the cylinder before the expansion stroke has

been finished. Turbulence has many characterising elements, in this work turbulent intensity

is used to characterise the small scale vortices. The turbulent intensity is normalized with the

mean piston speed, and defined as:

pV

uNTI (14)

22

m/spiston theofity Mean veloc

m/sintesity Turbulent

pV

u

Turbulence is also of great importance for diesel combustion. Compared with an SI engine,

which has a long time to mix air and fuel, the diesel process is a non-premixed combustion.

The mixing is done while the combustion takes place (in the diffusion phase of the

combustion). The spray creates lots of turbulent flows, which increase the mixing process.

Higher injection pressure means increased turbulent kinetic energy.

Squish flow

A squish region is the flat area on the piston that is located on the outer radius. Typically, this

area is designed to have a small clearance to the cylinder head, and in a diesel engine there

are several reasons for this [47]. First of all, as much of the air trapped in the cylinder as

possible should be located in the piston bowl cavity at TDC to contribute in the combustion

when the fuel is injected. For a given compression ratio (higher compression ratio gives

higher efficiency to a certain level), the heat transfer at TDC is high, and by decreasing the

volume outside the bowl, the heat transfer is also lowered. During the late stage of the

compression, right before TDC, the volume that is above the squish area (the squish volume)

rapidly decreases. The air trapped in this volume is forced towards the centre of the cylinder

and creates a strong flow. This squish flow can thereby strongly affect the combustion,

depending on when the fuel injection starts. The flow affects the swirl in the bowl and,

depending on how strong the swirl flow is, the squish flow contributes in different ways.

According to [44] and [47], in a light-duty engine at moderate SN, the squish flow flows into

the bowl horizontally and creates a rotating vortex in the bowl, as illustrated in Figure 11. At

high SN, the squish flow is deflected from the horizontal track to follow the bowl geometry

down into the piston cavity. This results in a change in the direction of the created vortex in

the bowl.

23

Figure 11. Squish flow interaction with swirl flow in the piston bowl [44], numerical simulation of flow

velocity at different SN. Flow velocity (Sp) is expressed in mean piston speed.

Cycle-to-cycle variations

All engines have cycle-to-cycle variations. It is normally coefficient of variation (CoVIMEP) of

indicated mean effective pressure (IMEP) that is the measuring variable. Definition:

avg

N

i

avgi

IMEP

IMEPIMEPN

1

2

IMEP

1

CoV (15)

where IMEPavg is the average of all measured cycles (N).

A rule of thumb is to keep CoVIMEP bellow 3% on engines in vehicles to maintain a good

drivability. When CoVIMEP rises above this value, problems with efficiency and emissions

typically arise as well. The main reason to cycle variations is based on the fluctuating nature

of turbulent airflows inside the cylinder. Typically, CoVIMEP is higher for SI engines compared

with CI engines. The main reason is that the combustion velocity in an SI engine is

dependent on the in-cylinder turbulence. Small changes in the turbulence give a larger

change in the combustion propagation velocity in a SI engine, because of the premixed

combustion. Specially the ignition event and the first laminar flame propagation is effected

due to stratification in the cylinder. In the CI engine, the fuelling is induced more or less

during the combustion and the HR rate is thereby mainly controlled by how fast the fuel

enters the combustion chamber. Still, the cycle-to-cycle variations of the airflow also

influence the combustion for CI engines. In [26], an investigation of the cycle-to-cycle

variations on SN was done using PIV measurements for two different SN settings. The

engine tested was a 4-stroke SI engine with a 2-valve cross-flow cylinder head and the swirl

was controlled either with a shrouded poppet valve with SN 6.0 and TN 2.0 (as determined in

a steady-state flow bench) and a standard poppet valve with SN 0.7 and TN 2.5. At TDC, the

24

measured (with PIV) SN for the shrouded valve decreased to a mean of SN 5.2, and for the

std valve it increased to a mean of SN 0.9. It was found that the low swirl case has more

cycle-to-cycle variations in SN compared with the high-SN case. A stronger swirl vortex

seems to be beneficial to have small-cycle variations in SN.

Emissions from diesel engines

The emission legislation for heavy-duty diesel engines has been strongly restricted since

1992, when the Euro I emission standards were set until the coming levels, Euro VI that is

seen in Table 1. The most challenging emissions from a diesel engine are particulate matter

(PM) and NOx. These emissions have decreased greatly to todays hard-to-measure levels.

The reason for lowering the emissions is justified by the fact of how they affect humans and

environment. Both soot and NOx irritate the respiratory organs [30]. The harmfulness of NOx

was found by tyre manufacturers when ground-level ozone caused rubber tyres to fracture.

The NOx in combination with HC from combustion was found to be the reason for both smog

and ozon. PM consist of Polycyclic aromatic hydrocarbons (PAH), solid coal and condensed

hydrocarbons. If the particles are small enough, they pass the cilia and go into the lungs.

Exposure to ultra-fine particles may cause cardiovascular diseases [28]. PAH are considered

to be highly carcinogenic. Carbon monoxide (CO) interferes with haemoglobin, which impairs

the ability to pick up and transport oxygen. CO has 300 times higher affinity compared with

the affinity of oxygen. This means that the haemoglobin takes up the CO instead of oxygen.

Table 1. Emission standards for HD diesel engines [27].

Emissions formation in a diesel engine

NOx emissions can be formed by four different mechanisms during combustion with air: the

Zeldovich mechanism, the Fenimore mechanism, the nitrous oxide- (N2O) intermediate

mechanism and the NNH mechanism [7]. The Zeldovich (or thermal) mechanism is the

dominating mechanism at high-temperature combustion over a wide range of and is the

mechanism that is most used to explain NOx formation from diesel engine combustion. The

Fenimore mechanism is particularly important in rich combustion. Fenimore (also called

prompt NO) discovered that NO was rapidly produced in a laminar premixed flame zone a

long time before thermal NOx had time to form. The hydrocarbon radicals react with nitrogen

and form amines or cyano compounds. These compounds are then converted to

25

intermediate compounds that, after a while, form NO. The N2O-intermediate mechanism has

an important role at very lean, low-temperature combustion processes (typically gas turbines

in lean operating conditions). Those three mechanisms can contribute in premixed and

diffusion flames. The NNH mechanism is a relatively new discovered reaction pathway, and it

seems to be particularly important in hydrogen combustion and methane combustion. NOx

can also be created if the fuel by itself contains nitrogen, which typical diesel fuel have only

in very low concentrations.

But as earlier mentioned, the Zeldovich mechanism is the dominating mechanism in diesel

combustion. As it is called today, the extended Zeldovich [7] version has three reactions:

O+N2 NO+N

N+O2NO+O

N+OHNO+H

The first reaction has a very high activation energy due to strong triple bond in the N2

molecule. The Zeldovich mechanism needs four things to form NOx: heat, time, oxygen and

nitrogen. If some of these components are reduced or removed, the NOx formation

decreases. The Fenimore mechanism is linked to the combustion chemistry of hydrocarbons.

HC are usually low from conventional diesel combustion. Some of the fuel that is injected is

formed into different hydrocarbons that are not present in the fuel. For example, it is found

small amounts of methane, formaldehyde and aromatics in the exhaust. The influence on the

human body has been discussed earlier, so the understanding of how these emissions are

created is of paramount importance. PAHs are formed under fuel-rich conditions, and they

are important precursors in soot formation. Another source of HC comes from injection

needle sack volume [2]. When the injection has ended, some fuel is left in the needle sack

volume that is ventilated to the combustion chamber. During the expansion stroke, when the

cylinder pressure decreases, the evaporating fuel that is left in the sack is evacuated into the

cylinder. When the pressure and temperature decrease during the expansion, the HC is not

combusted before the exhaust valve opens and evacuates the HC emissions. In Figure 12, a

bright soot cloud is created during cylinder pressure drop at some CAD after EOI. At later

CAD, when pressure and temperature decreases even more, the HC in the sack leaks out

without forming soot around the injector tip. By reducing the sack volume, this emission can

be reduced.

26

Figure 12. Optical engine picture of the combustion at the after-oxidation phase of combustion. The blue

circle indicates an injection sack leak after the injection has ended.

PM occurs in incomplete, cold and/or rich combustion. PM is kinetically controlled. PM starts

to grow when PAH molecules are conglomerated and start to form a particle. The particle

then starts to grow by addition of acetylene and coagulation. In Figure 13, the different steps

of PM formation are shown. The medicines used to reduce the soot emissions are the three

ts of combustion: time, temperature and turbulence [32]. This means, just increase the

temperature in the combustion and wait for a longer time with a good mixing, and the soot is

no longer a problem. This can be done with earlier start of injection (SOI). But, as always,

there is no free lunch, and NOx starts to form. The reason for this is the longer residual time

and higher cylinder temperature, which increases thermal NOx formation. This is the

classical NOx-soot trade-off problem that diesel engines struggle with. By reducing one of

the emissions, the other emission starts to increase.

27

Figure 13. Schematic soot formation in combustion [32].

CO comes from incomplete and cold combustion in diesel engines. Increased EGR, late

combustion facing, lack of oxygen and, as shown later in this work, too high swirl can

increase the CO emissions. HC and CO can easily be treated in an oxidation catalyst

mounted after the engine. The after-treatment system for NOx and soot is more complex and

expensive. If the engine-out emissions can be reduced, the after-treatment system can be

made simpler or ignored. This is not an easy goal.

Engine transients

Fast engine load build up, called transient, is a challenge to keep the exhaust emissions at

legal levels. In Figure 14, a transient load build up is plotted in forms of IMEP and inlet

plenum pressure at 1,000 rpm for a turbo diesel engine. The requested load, the blue line, is

the desired value and the red line shows the actual load on the engine and these values

differ. The problem is that the air supply from the turbocharger unit, the green line, needs

some time to build up the boost pressure. The amount of fuel needs to be restricted during

the boost pressure build up before it is at a normal operating level. If the fuel mass is not

restricted, the in-cylinder will go under a critical level and the engine will start to produce

high smoke levels. In the engine, electronic control unit (ECU) functionality to detect and

control low operation is constantly developed. Different models are implemented in the

ECU to predict the oxygen trapped in the cylinder, allowing the right amount of fuel to be

added for each cycle during the transient. As shown in examples of control system models

[33], [34], [35], [36], the cannot pass below a critical value, typically 1.251.30 for an

engine without a diesel particle filter, before smoke emissions increase rapidly. Besides a

disappointed driver, who wants engine power as fast as possible, there is also a longer time

period where the engine operates at unfavourable conditions. As long as the air is restricted,

an EGR engine cannot use EGR (as the turbo pressure is too low). The in-cylinder mean

28

temperature increases due to less gas mass being kept in the cylinder, with higher NOx

emissions as a result. The engine efficiency is also negatively affected when the SOI needs

to be retarded to maintain NOx levels at acceptable levels.

If the combustion system can maintain low emissions of PM and NOx at a lower , the

available exhaust energy and engine torque will be increased. This results in a faster build up

of boost pressure. Also, it can be traded against lower emissions during the transient.

Figure 14. Turbo diesel engine transient from 3-bar IMEP to full load, in this case, 23-bar IMEP.

Requested load, engine-out load and inlet pressure plotted versus time.

In the following, the tests that have been performed in this work and the attached

papers/journal are based on this engine transient. Load points are chosen from this line and

repeated in a single-cylinder engine and an optical engine with the same boundary

conditions. The only difference is that the respective load point is repeated at steady-state

conditions in the laboratory engines where the transient phenomena can be observed in

detail. The advantage is that relatively fast load build up can be examined and changes in;

for example, in-cylinder flow and injection parameters can be applied and studied in a

controlled way.

1

1.3

1.6

1.9

2.2

2.5

0

5

10

15

20

25

-1 0 1 2 3 4 5

Inle

t p

ress

ure

[bar

]

IMEP

[b

ar]

Time [s]

Req. Load IMEP [bar]

IMEP engine out [bar]

Inlet pressure [bar]

Load 1

Load 3

Load 2

Load 4

29

Project motivation

The main project focus is on how airflow in the cylinder affects combustion and the emissions

from a diffusion flame diesel engine at transient operation. As explained above, emissions

from an engine transient become increasingly important. During a transient, the emissions

increase. By creating a better understanding of the processes, the combustion can be

controlled in a better way during the transient. Ignition delay, diffusion combustion and after-

oxidation period are the main stages in diesel combustion. The after-oxidation period has

been shown by other work [37], [38] to be of paramount importance for engine-out emissions.

During after-oxidation, the mixing process is of paramount importance. Optical engine

measurements on flow structure during the after-oxidation part of combustion can give

important knowledge on the flow structure that affects the emissions from the engine.

The main objectives of this work are to:

1. Demonstrate the potential emission reduction when variable swirl and tumble is

applied to the combustion system at high-injection pressure.

2. Quantify the airflows in the cylinder before and during combustion. This will be done

with 1-D simulations and experimentally.

3. Look into the after-oxidation part of combustion and understand the effects on this

when injection pressure and swirl are changed.

Method

Fast load increase means low operation with less remaining oxygen for the after-oxidation

of the fuel. The demands on the mixing process in the cylinder are therefore higher. The

method for this work is to first measure what is happening during an engine transient,

examine where the critical points are and repeat them in a controlled way with different in-

cylinder airflow. The repetition is done in a single-cylinder engine with AVT, which is opened

up for variable in-cylinder airflow. Simulation tool GT-POWER, together with constant flow-rig

measurements, are used to quantify the airflow before combustion. To create an

understanding on how the flow field influences the combustion, optical engine measurements

are taken using a high-speed camera. The flow inside the cylinder during combustion is

quantified with optical calculations in a cross-correlation program. Simulated, measured and

processed data on flow quantities, emissions and combustion data is combined to examine

the airflow effect on diesel combustion during transient.

30

Methodology The experimental equipment and simulation tools used in this work can be seen in Figure 15.

Data from three different engines was used: a six-cylinder turbocharged production engine

1), a single-cylinder engine 6) and an optical engine 9). The six-cylinder engine was a Scania

DC13 Euro 5, equipped with EGR. The engine was used to measure the transient behaviour

and the data provided 2) was used to set up a single-cylinder and optical-engine test series.

The single-cylinder engine used in the tests was equipped with a Lotus AVT system 7). With

this system, the valves were hydraulically controlled, enabling the valve profiles 4) to be

shifted during operation. With different valve profiles, the airflow in the cylinder can be

affected over a wide range. The airflow was quantified as SN, TN and normalised turbulent

intensity (NTI). The 1-D simulation program, GT-POWER, 5) was used to calculate these

quantities. The two tested cylinder heads were measured in a constant flow rig 3). Swirl and

tumble at valve lifts from 1 to 15 mm, with increments of 1 mm, were measured for each

valve individually and with the two valves together. Valve profiles 4) were created in

MATLAB. These were then used in the Lotus AVT system and in GT-POWER. In this way,

the SN, TN and NTI were calculated and then used to plot, for example, emissions measured

in the single-cylinder engine, as a function of airflow in the cylinder 8). The tested load points

in the single-cylinder engine were then repeated in the optical engine 9). The combustion

pictures 10) was captured with a high-speed Phantom camera. The pictures were then

evaluated using LaVision DaVis 7.2 PIV software 11). Velocity vector fields were calculated

12) and CAD resolved during the injection and after-oxidation part of the combustion. The

results from single-cylinder tests were put together with the velocity vector figures calculated

from the high-speed combustion film. Data evaluation was then done from this set of data.

31

Figure 15. Overview of used experimental equipment and calculation tools.

Test equipment

The three different test engines used in this work have the same, or nearly the same,

combustion system layout, presented in Table 2 and Table 3. Some deviation was, of course,

inevitable, due to the different engine constructions, as seen in Table 3. The engine layout

was based on a Scania DL Euro 5 combustion system equipped with a common rail XPI

injection system capable of injection pressures up to 2,500 bar. This is a swirl-supported

combustion system with a re-entry combustion bowl design that does not need any extra

after-treatment system for Euro V emission legislation when EGR is used. It has a 4-valve

cylinder head with a centrally placed 8-hole injector.

80

70

60

50

40

30

20

10

0

-10

-20

-30

-40

-50

-60

-70

-4

-2

0

2

4

6

8

14

710

13 Angle on cylinder head

Tum

ble

val

ue

Valve lift [mm]

One valve tumble 6-84-6

2-4

0-2

-2-0

-4--2

300 350 400 450 500 5500

2

4

6

8

10

12

14

CAD

Valv

e lift

[mm

]

15 mm std

10 mm std

5 mm std

15 mm step

10 mm step

5 mm step

1

1,3

1,6

1,9

2,2

2,5

0

5

10

15

20

25

-1 0 1 2 3 4 5

Inle

t p

ress

ure

[bar

]

IMEP

[b

ar]

Time [s]

Req. Load IMEP [bar]

IMEP engine out [bar]

Inlet pressure [bar]

Load 1

Load 3

Load 2

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

Flow -

bench

GT-POWER

Single

cyl

Lotus

AVT

PIV DaVis

Emissions data

combined with

flow data

Combine

data

300 350 400 450 500 5500

2

4

6

8

10

12

14

CAD

Valv

e lift

[mm

]

15 mm std

10 mm std

5 mm std

15 mm step

10 mm step

5 mm step

1

2

3 4

5 6 7

8

9

1011 12

32

Table 2. Engine specifications.

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

Common test engine data

16

0.187 / 0.163

The difference between the tested engines can be seen in Table 3. The fist obvious

difference is that bore and stroke differs slightly between the engines. The reason is that

127/154 mm is the old engine configuration and 130/160 mm is the new configuration. The

optical engine was updated with the new bore but not the new stroke. The combustion

system (injection system, bowl geometry, cylinder head design and so on) is the same for all

test engines.

Table 3. Test engine individual specifications.

Engine type Optical engine Scania single cylinder Scania DL Euro 5

engine

Bore/stroke [mm] 130/154 127/154 130/160

Connecting rod [mm] 255 255 255

Valve system Camshaft Active valve train Camshaft

Single-cylinder engine with an AVT system

The single-cylinder engine was equipped with a Lotus AVT system that is a fully hydraulic

system. One hydraulic cylinder was coupled to each valve in the cylinder head, as seen in

Figure 16. Hydraulic oil, with a pressure of approximately 200 bar, was supplied to each side

of the piston inside the hydraulic cylinder, making the piston move. The oil flow was

controlled by a servo valve that directs the oil to one side of the piston at a time. In this way,

the engine valves are controlled with a good level of accuracy.

33

Figure 16. The Lotus AVT valve actuator with inlet valve (a) and the actuators on the cylinder head (b)

Different valve profiles can be implemented in the AVT system. They are created in MATLAB

and then uploaded to the system. During engine operation it is then possible to change the

valve profiles. Some examples of valve profiles used in this work are seen in Figure 17. The

system enables SN variations between 0.4 to 6.7 and TN from 0.5 to 4.0 with the used

cylinder head configuration.

Figure 17. The tested valve profiles. The dark-blue profile is the standard valve profile used in a standard

engine.

Optical engine

The optical engine layout can be seen in Figure 18, with the two 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. The camera, Phantom v7.3, is installed next to the engine and the

combustion light is transferred to the camera by a mirror mounted inside the piston

extension. The engine is capable of running with cylinder pressures up to 160 bar. A titanium

clamping ring was mounted above the piston glass to fix it. This restricted the field of view to

a diameter of 80 mm, compared with the total cylinder bore of 130 mm. Two different shapes

300 350 400 450 500 5500

2

4

6

8

10

12

14

CAD

Valv

e lift

[mm

]

15 mm std

10 mm std

5 mm std

15 mm step

10 mm step

5 mm step

b

a

34

on the piston bowl glass were tested, bowl-shaped piston and flat piston bowl. A schematic

spray path is plotted, Figure 18, in the two piston bowls. The optical engine had a normal

camshaft valve mechanism. To change the in-cylinder airflow, two different cylinder heads

were used. To further extend the possible airflow in the cylinder of the optical engine, one of

the inlet ports was also blocked (or not).

Figure 18. Principal layout of the optical engine on the left. On the right, the two tested piston bowl shapes

with plotted schematic spray.

Due to the long piston extension, the effective compression ratio is lower than the

geometrical compression ratio and decreases with cylinder pressure. At 160 bar, the

distance between cylinder head and squish area on the piston increased by 1.5 mm

compared with atmospheric pressure. To compensate for the lower compression ratio, the

boost pressure and inlet temperature was increased so the motoring cylinder pressure at

TDC in the optical engine was equal to the single-cylinder engine. The was slightly higher

in the optical engine compared with the single-cylinder engine. The increase in inlet

temperature was also done to compensate the increased ignition delay in the optical engine,

since only one combustion event was performed during the measurement.

Steady-state flow rig

A honeycomb type steady-state flow rig was used to generate swirl and tumble data for the

GT-POWER simulations. In Figure 19, a sketch of the flow rig can be seen with the most

important parts marked with a number. 1) the tested cylinder head is mounted on a cylinder

liner 2) with an inner diameter equal to the engine bore. A honeycomb torque meter 3) is

mounted in the stagnation chamber 4), which measures the torque that is created by the

swirling flow from the cylinder head. The pressure difference between the stagnation

chamber and the atmosphere is created by the fan 8) and monitored by 5) and kept constant

for all measurements at 25 mbar. To measure the airflow that is passing the cylinder head, a

roots blower is mounted in 6), which works as an airflow meter. By monitoring the pressure

35

difference before and after the roots blower 7) and then adjusting the rotational speed so the

pressure drop is zero across it, the airflow is proportional to the speed.

Figure 19 Sketch of the steady-state flow rig with the most important parts marked in the figure [39].

To measure the tumble flow in the same rig, the cylinder liner was replaced with a 90

cylinder liner, as seen in Figure 20. The tumble that is created in this way is transformed to a

swirling motion that can be measured in the same way as when swirl is measured. When

both swirl and tumble exist in the airflow, the cylinder head needs to be rotated on the 90

liner to find the highest torque measurement for a certain valve lift height (as marked in the

figure).

Figure 20. Cylinder liner with 90 bend to transfer the tumble motion from the cylinder head to a swirl-

like motion which can be measured on the honeycomb torque meter [39].

The swirl and tumble changes with lift height. Maximum tumble was measured at different

angles on the cylinder head depending on lift height. The cylinder head angle therefore

needs to be adjusted to maintain tangential flow. The cylinder head was measured every 5

degrees for valve lifts 115 mm, with increments of 1 mm. Figure 21 shows the measured

tumble for one valve, operating in the straight port seen in Figure 22. The highest mean

0

(-)

(+)

36

tumble during a standard lift valve profile was chosen for the data set marked 70 and used

in the following calculations.

Figure 21. Tumble for the straight port with valve lift on the y-axis, cylinder head angle on the x-axis and

tumble on the z-axis.

Cylinder head design

The cylinder head design used in the test is seen in Figure 22 a. To increase the SN, extra

masking of the valve seats was done. This is shown in Figure 22 b see the red circles. The

extra masking was used in the single-cylinder engine and optical engine. The six-cylinder

engine does not have this extra masking. The optical engine, however, does not have the

AVT system to vary the airflow in the cylinder. In this case, one cylinder head with masking

and one without were used to vary the airflow. The SN increased from 2.1 to 3.4 when this

masking was applied. By blocking one inlet port (the lower blue arrow in Figure 22), the SN

increased from 3.4 to 6.3. In this way, four different airflow characteristics were applied in the

optical engine. For the single-cylinder engine, only the cylinder head with the extra masking

was used. By changing the valve profiles, the SN could be varied between 0.4 to 6.7 and

tumble between 0.5 to 4 with the same cylinder head.

80

70

60

50

40

30

20

10 0

-10

-20

-30

-40

-50

-60

-70

-3

-2

-1

0

1

2

3

4

5

6

7

1 4

7 10

13 Angle on cylinder head

Tu

mb

le v

alu

e

Valve lift [mm]

One valve tumble 6-7 5-6 4-5 3-4 2-3 1-2 0-1 -1-0 -2--1 -3--2

37

Figure 22 . Orientation of the valves and the inlet ports for the tested cylinder heads. The inlet valve seats

have extra masking for the high-SN head, marked in picture b, to increase the swirl.

Theoretical models and calculation methods

Calculation of SN and TN

Swirl in CI engines is usually defined as the rotational flow eddy around the cylinder centre

axis at BDC. The flow is assumed to be a solid-body rotation around the geometrical cylinder

centre [48]. Two methods, the engine-simulation program GT-POWER and the Thien

method, have been studied and compared. In both cases, data for the calculations are

measurements from a constant flow ring (Figure 19), where torque is measured as a function

of valve lift. The two methods are shown here.

Thien

Thien [49] uses the movement of the piston to estimate the air volume passing the inlet

valve, assuming incompressible flow over the valve during the entire inlet stroke. The dz/d

term in Eq 16 is the volume change depending on CAD. Eq 16 is the swirl coefficient

definition for Thien. Tumble is calculated in the same way with flow data for the perpendicular

contribution inserted in the equation instead.

d

d

dzNSNT

2)(

)( (16)

2

22)(q

MsN

(17)

Where: s = stroke [m], M= measured torque [Nm], 2 = air density [kg/m3] and q = volume

flow [m3].

1-D simulation tool

As with the Thien method, GT-POWER assumes solid-body rotation in the cylinder and uses

flow-rig data to calculate SN. The flow-rig data for respective valve lift is put into Eq 18,

where the torque, M, is normalised to the air mass flow, m , the isentropic airflow (through

the valve), isV and cylinder bore, B. The created vector, which depends on valve lift, is used

by the program to estimate the SN and TN. The program calculates the air mass flow

passing into the cylinder, depending on the pulsating pressure/flow in the inlet system and

a b

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the valve lift. With the calculated airflow, the momentum contribution from the airflow moving

into the cylinder is added to the rotating air mass in the cylinder. Swirl and tumble, which are

calculated using GT-POWER, are called SNgt and TNgt in this work.

2/BVm

MC

is

S

(18)

The test-cell single-cylinder engine was built up in the 1-D program with the right geometries

for inlet, exhaust-system and cylinder parameters. Measured engine data was compared with

simulated engine data to tune the model so it behaves like a real engine. More on how this

can be done is shown in [50]. The built-in flow model was used to estimate in-cylinder

airflow and heat transfer, and the combustion model CombDIJET was used. The flow

model uses a simplified cylinder geometry and a k- model to estimate, for example,

normalised turbulent intensity. More information about how GT-POWER handles in-cylinder

flow can be found in [51], [52] and [53]. The tumble algorithm in GT-POWER is quite similar

to the calculation method for swirl, and this is shown in detail in [51].

The NTI is a global mean parameter for the entire cylinder. The values shown in this work

are at the start of the cycle, 100 CAD before top dead centre ( BTDC), calculated using the

GT-POWER built-in k--based model.

The SN, TN and NTI output from GT-POWER is shown in Figure 23. The results represent

two valves operating at 15 mm lift and standard valve profile for the entire engine cycle.

During the intake stroke, the swirl and tumble build up. Their values depend on valve lift. At

the inlet valve closure (IVC), the tumble and swirling motions are conserved in the cylinder.

The airflow speed accelerates during compression due to the geometric change. The

turbulent intensity is high during the inlet stroke when the air flows into the cylinder at high

velocity. The turbulent intensity drops during compression and at TDC it starts to increase

again when the tumble motion is transformed into turbulence with a smaller-length scale.

This behaviour has also been observed in large eddy simulations (LES) in [54]. At TDC, the

fuel injection creates high turbulent intensity in the cylinder, as shown in Figure 23. This

should only be seen as a rough estimate.

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Figure 23. The SN, TN and NTI plotted for the entire engine cycle.

The outer gas exchange system GT-POWER is handled by using a 1-D approach. In a pipe

that contains a pulsating flow, the velocity in the axial direction in the pipe is significantly

higher than the velocity and flow in the cross-sectional plane. By just modelling the flow in

the axial direction, the continuity, momentum and energy equation can be greatly simplified

by just solving the equations for the x-axis. This is the fundamental approach of 1-D

simulation. GT-POWER handles the flow in the different objects with a space-discredited

form of the continuity, momentum and energy equation. For more information about 1-D

simulation in GT-POWER, I recommend that you read [55].

Velocity measurement techniques

Hot wire measurements are a common measurement technique to use when a gas flow is

being examined. It is possible to measure both the mean velocity and the turbulent velocity

wherever the probe is placed. By heating a thin wire, typically made of platinum-coated

tungsten, and placing it in the gas stream (turbulent or laminar), the energy that is needed to

maintain a constant temperature is proportional to the gas mass flow that passes the wire. By

choosing a thin wire, the probes response time can be decreased so both the fluctuating part

(u) and the mean part (U) of the turbulence velocity can be measured [41].

Hot wire measurements around TDC have turnout to give large errors in respect of

turbulence intensity levels around TDC. The indicated values have been a factor two or

greater than those obtained with, for example, Laser Doppler Anemometry (LDA) [41]. LDA is

estimated as a more accurate measurement method.

NTI from

injection

40

Optical evaluation method on combustion pictures

In the optical engine described earlier, high-speed films are captured with a Phantom high-

speed colour camera [56]. In this work, the total resolution of the pictures from the camera

was 256 x 256 pixels, with a colour depth of 14 bit. The time delay between every picture, t,

was set to 28 s, which means 0.17 CAD at 1,000 rpm. The same load points that have been

tested in the single-cylinder engine were repeated in the optical engine. Two sets of film were

captured for every test point, one with an external light source (4 x 70W Xenon lights) and

one with natural combustion light. With the external light source the spray and early injection

phase could be captured on film. When the combustion starts, the illumination from the

combustion is many times larger than the external lights. Another shutter time and diaphragm

scale needs to be selected to avoid overexposure of the pictures. The captured pictures from

the combustion were then evaluated with PIV software, DaVis 7.2 [57].

It is difficult to obtain normal PIV measurements during diesel engine combustion. The

seeding particles that are introduced in the inlet air can be combusted or destroyed by the

high-temperature combustion. The laser sheet that is introduced into the cylinder can easily

be drowned in the bright combustion light that contains a broad spectrum of wavelengths.

The flow in the cylinder is three-dimensional and turbulent. To catch the turbulent

phenomena, the PIV measurements need to be carried out both vertically and horizontally in

different positions. With cycle-to-cycle variations, observed in [41], numerous laser sheets

need to be introduced at the same time. This makes it very complex to capture double-

exposure pictures on each laser sheet when all of the sheets are introduced into the

combustion chamber. Instead, the clouds of glowing soot particles that are created during

combustion can be directly traced in the PIV software, which I have called combustion image

velocimetry (CIV) in this work. By comparing two pictures at a time, the glowing particles are

traced using cross-correlation. The light from the glowing particles is the tracer in the PIV

software, 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. Off course, this method makes it difficult to calculate the total velocity vector when the

z-axis is not included, but it gives a good picture of the flow in the 2-D plane close to the

piston. This means that information about the fluctuating velocity components can be hard to

capture, but the mean velocity part can easy observed, which this work is concentrated

around. Another problem with this method is to determine the height where the flow is

observed. During combustion, the traced flame is near the piston glass. The illumination from

the flame is strong, so the traced particles are estimated to be near the piston bowl

geometry. During after-oxidation, and especially at late CAD, the illumination decreases and

light from particles in the centre of the combustion chamber are the easiest particles to trace.

The reason is that the temperature in the centre of the combustion chamber is still high, so

the soot particles can still illuminate. Near the cylinder head and the glass in the piston bowl,

the temperature is much lower, which means that the soot cannot illuminate as strongly as

the particles in the middle of the combustion chamber. In this way, the soot particles that is

after-oxidising is traced and their flow orbit can be examined. Basically, the flow that

contribute to the after-oxidation mechanism in the diesel combustion.

Two different piston bowls were used in this work, one with a flat piston bowl and one with a

bowl-shaped piston. The flat piston gives no distortion to the picture due to the refractive

index, but the bowl-shaped piston does, as seen in Figure 24. The glass is placed on graph

paper with the combustion chamber side attached to the paper. As seen in the outer region

41

of the glass, the picture is distorted. The distortion becomes less visible when the distance

from the glass decreases. The problem is obvious, as, depending on where the observed

object is placed in the combustion chamber, the distortion becomes different. Close to the

glass, no distortion is observed. The injected spray has the same angle as the glass and is

thereby located near the glass at TDC. In this case, no compensation is needed. During the

after-oxidation part of the combustion, when the piston is moving down, the distortion may be

a problem. The distortion also changes with the moving piston, but the traced particles are

assumed to be placed near the glass, even at the after-oxidation period, and thereby no

compensation is applied.

Figure 24 Bowl -shaped piston photographed upside down on graph paper.

Calculation of a CAD-resolved vector field

By comparing two pictures for different CAD with cross-correlation, a picture series with the

vector field and a resolution of 0.17 CAD between the pictures can be created. The

information is then used to calculate the mean SN from the pictures CAD resolved.

As will be shown later, the results for velocity vectors are quite stable, indicating that enough

traceable information, from the glowing soot particles, remain in sight between the pictures.

In Figure 25, the principle of the picture evaluation can be seen. Every picture was divided

into 16 x 16 (or less for diffusion flame figures) pixel integration windows in which a mean

velocity vector

22yx VVV (19)

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

picture evaluation, the last picture was evaluated with the next picture. A film with velocity

vectors was thereby created. To reduce the error reading from the pictures, the DaVis built-in

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

pictures. The missing or erroneous vectors mainly came from limited or missing structure,

mainly caused from lack of glowing soot particles, in the area that can be tracked by DaVis.

In Figure 25, 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

42

movement and the traceable particles are well captured by the PIV program. In the figure, a

colour scale indicates which mean velocity every box has; arrows indicate the direction of

the flow and velocity; and the black dot indicates missing or erroneous data in the box, and

this is not included in further calculations.

Figure 25. Evaluated combustion pictures at 25.9 and 26.1 ATDC together with the calculated vector

field.

Calculation of a CAD-resolved SN

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

calculations were made. To calculate the SN