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Doctoral Thesis

Hydrothermal spallation drilling experiments in a novel highpressure pilot plant

Author(s): Stathopoulos, Panagiotis

Publication Date: 2013

Permanent Link: https://doi.org/10.3929/ethz-a-009795863

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Diss. ETH No.21163

Hydrothermal spallation drilling

experiments in a novel high pressure pilot

plant

A dissertation submitted to

ETH ZURICH

for the degree of

Doctor of Sciences

presented by

Panagiotis Stathopoulos

Dipl. Eng. National Technical University of Athens

born on March 6, 1982

citizen of

Greece

Accepted on the recommendation of

Prof.Dr., Dr. h.c. PhilippRudolf vonRohr , examiner

Prof.Dr.DimosPoulikakos , co-examiner

Dr.Gunter Siddiqi , co-examiner

Zurich, 2013

c© Panagiotis Stathopoulos, 2013

To see a world in a grain of sand

And a heaven in a wild flower,

Hold infinity in the palm of your hand,

And eternity in an hour.

William Blake (1757-1827), from “Auguries of Innocence”

Acknowledgements

The present thesis would not be possible without the material and im-

material support of many people, whom I would like to thank in these

paragraphs. The project was sponsored from the Swiss Federal Office of

Energy, Swiss Electric Research and the Swiss National Science Founda-

tion. I am obliged to thank Dr. Gunter Siddiqi for his trust in my decisions

and the absolute absence of any pressure from his side. Such a support is

never to be taken for granted. I would also like to thank Prof. Dimos

Poulikakos who agreed to be the co-examiner of the current thesis despite

his full calendar and his many obligations as Department Head.

My most sincere thanks goes to my supervisor Philipp Rudolf von Rohr. I

will be always grateful for the risk he took when he assigned to me one of

the most risky and costly projects of our research group. His insights and

daring propositions throughout the whole duration of the project made me

push my limits to a great extent. The academic freedom I enjoyed while

working with him made me definitely a more competent person and allowed

me to develop my engineering skills.

My earnest thanks goes to Prof. Schmalz and to my colleague Martin

Schuler for their valuable help and patience during the safety analysis of

the plant. The discussions with them offered me alternative points of view

that made the plant a very safe and efficient facility. A special thanks goes

also to Tobi Rothenfluh for the discussions we had on many issues but most

of all on the design of my pressure vessel. You allowed me to learn from

the shortcomings of the plant you were working on and you contributed to

the sound design of my plant.

I

A great thanks goes to Bruno, Peter, Dani and Rene for the helpful discus-

sions and their help by many technical difficulties. I will always remember

the day we all had, while installing the upper cover of my plant. Further, I

would like to thank Markus Mahr and his team at the PSI, who were always

ready to assist me and invested some of their working time to construct my

sensors.

An agreeable and stimulating office environment is also very important, so

I would like to give my thanks to all the LTR members, current and former,

who accompanied me in so many activities. A special thanks goes to Adi

for sharing his office with me throughout our whole PhD and for putting

up with all my bad days, changing moods and controversial ideas. Thanks

to Holger and Nora for their friendship from the beginning of my PhD. I

am really grateful to Karol Prikopsky, Beat Wellig and Markus Weber for

letting me profit from their vast experience and giving me valuable input

during the design of my plant. You all helped me immensely to overcome

my insecurities and fears in building such a big plant.

The amount of work I put in my thesis would by no means be possible with-

out the help of my students. I therefore would like to thank here Pirmin,

Simon, Roland, Basti, Beni, Ruizhi, Stephan, Enzo and my partners in

crime Florian Hofmann and Kaspar Ninck. I thank you all for choosing me

as your tutor and I hope you all learned something from me. The work I

performed with my third partner in crime, Thierry Meier, will continue for

the next three years and I am happy you are going to take over the plant.

Thanks for the courage you gave me and the intensity of our discussions.

Last but not least, I would like to thank my family for supporting me in all

my decisions and providing me with the means to study. I know you made

many sacrifices during all these years and I hope you can be proud of the

person I became.

The greatest thanks of all goes to my wife and partner for life, Xenia. Your

love and understanding during my dead ends and frustrations, helped me

go through all the obstacles I met. I am blessed to have you by my side.

II

Abstract

The current thesis concerns the experimental investigation of a novel ther-

mal drilling technology called hydrothermal spallation drilling. Shallow

spallation drilling applies the impingement of flames on rocks, in order to

break them in small disk-like fragments (spalls), due to the thermal stresses

induced in them. The absence of any contact between the drilling bit and

the rock and the demonstrated higher penetration rates in crystalline rock

drastically reduce the costs for drilling. For the deeper boreholes of geother-

mal power plants and Enhanced Geothermal Systems (EGS), spallation

would require flames in supercritical water (hydrothermal flames). Trans-

ferring spallation drilling to high pressure aqueous environments introduces

several new challenges that have yet to be addressed. The most important

among them are the forced ignition of hydrothermal flames, the entrain-

ment of water in the hot hydrothermal flame-jets and the measurement of

the heat transfer from the flames to the rock.

In order to investigate these issues, a novel high pressure pilot plant with

an installed thermal capacity of 120 kW (which is several times higher than

comparable facilities) has been built and commissioned. The plant and its

vessel operate at temperatures up to 420 ◦C and pressures up to 350 bar

and are able to accommodate rock samples with diameters up to 10 cm.

The combustion chamber of the plant not only accommodates much higher

thermal loads but also demonstrates significant improvements in terms of

combustion stability compared with similar plants.

The approach of the forced ignition problem comprised two phases, starting

with the convective heat transfer measurements in the combustion chamber.

III

In this first phase, mixtures of water, ethanol and nitrogen were used to

model the combustible mixture and heat transfer enhancement, typical for

supercritical fluids, was observed. The following ignition experiments with

mixtures of ethanol, water and oxygen resulted to a reliable ignition and

demonstrated minimum ignition temperatures around the pseudo-critical

point of the mixtures. The successful forced ignition led to a considerable

reduction of the reactants’ temperatures before ignition. This fact not only

increased the operational safety of the facility but, most importantly, shows

the feasibility of down-hole ignition of hydrothermal flames. Indeed, state

of the art down-hole generators can produce the required ignition power

according to our measurements.

To address the entrainment issue, the heat transfer capabilities of six differ-

ent configurations of hydrothermal flame-injection nozzles have been ana-

lyzed through the measurement of their impingement temperature profiles.

An initial increase, followed by a region of constant impingement temper-

atures was observed when increasing the jet power and velocity. Further

increase of the jet power raised the temperatures on the plate again. This

behavior was explained with classical entrainment rules and by consider-

ing the flame-jet properties around its pseudo-critical point. The resulting

insights in the impingement of hydrothermal flames supported the initial

optimization of the drilling tool design.

For the heat transfer measurements in hydrothermal flames two novel heat

flux sensors have been constructed and calibrated. One of them proved to

be suitable for the intended operating conditions, while the design of the

second sensor must be slightly adapted prior to its implementation in the

high pressure plant.

After addressing the above, an optimized version of the combustion cham-

ber nozzles has been successfully implemented in drilling experiments. We

were able to drill small cavities in granite with only 60% of the plant ther-

mal capacity, thus delivering a definite proof of the success of the plant

design and the potential of the concept of hydrothermal spallation drilling.

IV

Zusammenfassung

Die vorliegende Arbeit beschaftigt sich mit der experimentellen Untersu-

chung einer neuartigen Bohrtechnik, des sogenannten “Hydrothermal Spal-

lation Drilling”. Bei Spallation Drilling wird ein Flammenstrahl gegen die

Oberflache eines Gesteins gerichtet, um es durch die entstehenden thermi-

schen Spannungen in scheibenformige Fragmente (“Spalls”) zu zerbrechen.

Dieses beruhrungslose Bohrverfahren erzielt hohe Bohrgeschwindigkeiten

in kristallinen Gesteinsformationen, was zu einer deutlichen Senkung der

Bohrkosten fuhren konnte. Das Ubertragen vom Spallation Drilling Kon-

zept in geothermische Kraftwerke und insbesondere Enhaced Geothermal

Systems (EGS) benotigt den Einsatz von Flammen im uberkritischen Was-

ser (hydrothermale Flammen), was neue technische Herausforderungen mit

sich bringt. Die wichtigsten davon sind die erzwungene Zundung der hydro-

thermalen Flammen, der Wassereintrag in den heissen Flammenstrahl und

die Warmetransportmessung beim Aufprall der Flammen auf das Gestein.

Um diese Themen nachzugehen, wurde eine neue Hochdruck-Pilotanlage

mit einer thermischen Leistung von 120 kW (wesentlich hoher als je zuvor)

entworfen, gebaut und in Betrieb genommen. Sie kann bei Temperaturen

bis 420 ◦C und Drucke bis auf 350 bar betrieben werden und sie bietet die

Moglichkeit Gesteinsproben mit einem maximalen Durchmesser von 10 cm

zu untersuchen. Ihre Brennkammer weist im Vergleich zu bestehenden An-

lagen eine sehr hohe Leistungsdichte und Effizienz auf.

Ausgangspunkt fur die Erforschung der erzwungenen Zundung war die Mes-

sung des Warmeubergangskoeffizienten in der Brennkammer. Als Model fur

das brennbare Gemisch wurden Wasser-Ethanol-Stickstoff Gemische einge-

V

setzt und dabei wurde ein erhohter Warmeubergangskoeffizient festgestellt,

der typisch fur uberkritische Fluide ist. Im Anschluss wurde der Stickstoff

durch Sauerstoff ersetzt und beim pseudo-kritischen Punkt der resultieren-

den Gemische wurde ein Mindestwert der Zundtemperaturen beobachtet.

Durch die deutlich tieferen Vorheiztemperaturen der Reaktanden, die durch

die erzwungenen Zundung erzielt wurden, hat sich die Anlage sicherheits-

technisch verbessert. Ausserdem wurde die Machbarkeit der erzwungenen

Zundung in einem tiefen Bohrloch unter Beweis gestellt, weil die gemes-

senen Zundleistungen auch von bestehenden Bohrlochgeneratoren erreicht

werden konnen.

Um das Problem des Wassereintrags in den hydrothermalen Flammenstrahl

anzugehen, wurden sechs Dusen zur Injektion des Flammenstrahls entwor-

fen und verglichen. Dafur wurden die Temperaturverteilungen auf einer

metallischen, ebenen Plate gemessen, die vom Flammenstrahl senkrecht

angestromt wurde. Wenn die Flammenleistung erhoht wurde, sind diese

Temperaturen zunachst gestiegen, dann haben sie einen konstanten Wert

aufgenommen und schliesslich sind sie weiter gestiegen. Dies lasst sich durch

die Zusammenwirkung des Wassereintrags in den Strahl und der besonderen

Stoffeigenschaften des Flammenstrahls um seinen pseudo-kritischen Punkt

erlautern. Die damit gewonnenen Erkenntnisse im Hinblick auf die Warme-

transportmechanismen fuhrten zu einer Optimierung der Injektionsdusen.

Fur die beabsichtigten Warmetransportmessungen wurden zwei Warme-

flusssensoren konzipiert, gebaut und kalibriert. Einer dieser Sensoren hat

sich als besonders geeignet fur die vorgeschriebenen Bedingungen erwiesen,

wahrend der Andere leichten Anpassungen unterzogen werden muss.

Nach dem Uberwinden der obengenannten Herausforderungen ist die opti-

mierte Injektionsduse in Bohrexperimenten erfolgreich eingesetzt worden.

Bereits ein Betrieb der Anlage bei lediglich 60% der eigentlichen Nenn-

leistung konnte in den Bohrexperimenten kleine Bohrvertiefungen auf der

Oberflache von Granitproben zu Tage fordern. Dies zeigt nicht nur die

erfolgreiche Planung und Realisierung der Anlage sondern auch das hohe

Potential des Bohrkonzepts.

VI

Table of Contents

Acknowledgements II

Abstract IV

Zusamenfassung VI

Nomenclature XVI

1 Introduction and project goals 1

1.1 Energy supply and consumption worldwide and in Europe . 2

1.1.1 Geothermal energy supply and potential . . . . . . . 5

1.2 Cost analysis of geothermal plants . . . . . . . . . . . . . . 7

1.3 Introduction to spallation drilling . . . . . . . . . . . . . . . 9

1.4 Hydrothermal spallation drilling - Goals of the thesis . . . . 14

1.4.1 Thesis structure and goals . . . . . . . . . . . . . . . 16

2 Plant design, construction and commissioning 19

2.1 Process design . . . . . . . . . . . . . . . . . . . . . . . . . 20

2.1.1 Process description . . . . . . . . . . . . . . . . . . . 22

2.1.2 The HAZard and OPerability study . . . . . . . . . 24

2.2 Plant construction . . . . . . . . . . . . . . . . . . . . . . . 29

2.2.1 The fuel sub-system . . . . . . . . . . . . . . . . . . 29

2.2.2 The oxygen sub-system . . . . . . . . . . . . . . . . 31

2.2.3 The CW2 sub-system . . . . . . . . . . . . . . . . . 32

2.2.4 The CW1&3 sub-system . . . . . . . . . . . . . . . . 32

2.2.5 The effluent water sub-system . . . . . . . . . . . . . 35

VII

Table of Contents

2.2.6 Control system . . . . . . . . . . . . . . . . . . . . . 35

2.3 Plant commissioning . . . . . . . . . . . . . . . . . . . . . . 38

2.3.1 Commissioning and temperature control of the heaters 39

2.3.2 Pressure control . . . . . . . . . . . . . . . . . . . . 41

2.3.3 Fuel mixture composition control . . . . . . . . . . . 44

3 Design of the pressure vessel and its combustion chamber 47

3.1 Pressure vessel specifications and geometrical characteristics 48

3.2 Combustion chamber design . . . . . . . . . . . . . . . . . . 52

3.2.1 Combustion chambers used previously in the LTR . 53

3.2.2 Design concept of the combustion chamber . . . . . 56

3.2.3 Bluff body, flame stabilization analysis . . . . . . . . 58

3.2.4 Combustion system construction . . . . . . . . . . . 61

3.3 High pressure auxiliaries . . . . . . . . . . . . . . . . . . . . 64

4 The ignition project 67

4.1 The scientific problem - motivation . . . . . . . . . . . . . . 67

4.1.1 Ignition measurement goals . . . . . . . . . . . . . . 68

4.1.2 The measurement concepts . . . . . . . . . . . . . . 69

4.2 Hot surface ignition state of the art . . . . . . . . . . . . . . 70

4.2.1 Thermal ignition theory . . . . . . . . . . . . . . . . 70

4.2.2 Experimental studies on hot surface ignition . . . . . 72

4.3 Convective heat transfer in supercritical fluids - literature

review . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 74

4.3.1 Physical properties of water and its mixtures with

ethanol, oxygen and nitrogen in the trans-critical region 74

4.3.2 Convective heat transfer phenomena in supercritical

fluids . . . . . . . . . . . . . . . . . . . . . . . . . . 78

4.4 The igniter modules . . . . . . . . . . . . . . . . . . . . . . 81

4.4.1 Ceramic ignition module . . . . . . . . . . . . . . . . 82

4.4.2 Coil igniter . . . . . . . . . . . . . . . . . . . . . . . 85

4.4.3 Electrical connection of the igniters . . . . . . . . . . 87

4.4.4 Data acquisition electronics . . . . . . . . . . . . . . 88

VIII

Table of Contents

4.5 Heat transfer experiments in the combustion chamber . . . 91

4.5.1 Measurement procedure . . . . . . . . . . . . . . . . 91

4.5.2 Heat transfer coefficient results . . . . . . . . . . . . 94

4.5.3 Conclusions of the heat transfer measurements in the

combustion chamber . . . . . . . . . . . . . . . . . . 100

4.6 Ignition experiments . . . . . . . . . . . . . . . . . . . . . . 101


4.6.2 Ignition experiments results . . . . . . . . . . . . . . 103

4.6.3 Conclusions of the ignition experiments . . . . . . . 109

4.7 Flame temperature profile measurements . . . . . . . . . . . 111

4.7.1 Scientific problem, motivation and aims . . . . . . . 111


4.7.3 Experimental results . . . . . . . . . . . . . . . . . . 113

4.7.4 Conclusions of the axial flame temperature measure-

ments . . . . . . . . . . . . . . . . . . . . . . . . . . 117

5 Design of the hydrothermal spallation drilling tool 119

5.1 Heat flux sensor development . . . . . . . . . . . . . . . . . 121

5.1.1 Scientific problem - sensor requirements . . . . . . . 121

5.1.2 Heat flux sensors state of the art . . . . . . . . . . . 121

5.1.3 Choice of the heat flux sensors working principles . . 124

5.1.4 Transverse type anisotropic sensor . . . . . . . . . . 125

5.1.5 Thin film resistance sensor . . . . . . . . . . . . . . 133

5.2 Calibration plant and methodology for heat flux sensors . . 137

5.2.1 Scientific problem - calibration requirements . . . . . 137

5.2.2 Heat flux sensor calibration state of the art . . . . . 138

5.2.3 The calibration concepts . . . . . . . . . . . . . . . . 140

5.2.4 The convection calibration setup . . . . . . . . . . . 142

5.3 Calibration of the sensors . . . . . . . . . . . . . . . . . . . 145

5.3.1 Transverse sensor calibration . . . . . . . . . . . . . 146

5.3.2 Thin film resistance sensor calibration . . . . . . . . 150

5.4 Drilling tool design - Flame impingement temperatures . . . 155

5.4.1 Scientific and technical problem . . . . . . . . . . . . 155

IX

Table of Contents

5.4.2 Literature review and design considerations . . . . . 156

5.4.3 Design of the combustion chamber nozzle . . . . . . 160

5.4.4 Experimental setup . . . . . . . . . . . . . . . . . . . 162


5.4.6 Experimental results and impingement temperature

profiles . . . . . . . . . . . . . . . . . . . . . . . . . 165

5.4.7 Conclusions of the impingement experiments . . . . 174

6 Thesis conclusions 177

6.1 Summary of the technical results . . . . . . . . . . . . . . . 177

6.2 Summary of the scientific results . . . . . . . . . . . . . . . 180

7 Thesis outlook 183

7.1 Sensor project outlook . . . . . . . . . . . . . . . . . . . . . 183

7.1.1 Heat flux measured with optical fibers . . . . . . . . 183

7.1.2 Heat flux sensors made of ceramic thermoelectric oxides185

7.1.3 Optimization of the existing sensors . . . . . . . . . 187

7.2 Drilling tool design outlook . . . . . . . . . . . . . . . . . . 188

7.2.1 Initial drilling experiments . . . . . . . . . . . . . . 189

7.2.2 Future impingement heat transfer experiments . . . 191

7.2.3 Future drilling measurements . . . . . . . . . . . . . 192

Bibliography 195

A The design of the pressure vessel 215

A.1 First concept . . . . . . . . . . . . . . . . . . . . . . . . . . 215

A.2 The final design of the pressure vessel . . . . . . . . . . . . 218

List of publications 221

Curriculum vitae 222

X

Nomenclature

Abbreviations

3WV Three way valve

ASTM American Society for Testing and Materials

ATEX ATmospheres EXplosives

AWG American Wire Gauge

BF Blockage Factor

BRICS Brasil, Rusia, India, China, South Africa

BS Bursting disc

CFD Computational Fluid Dynamics

CSTR Continuous Stirred-Tank Reactor

DI DeIonized (water)

EGS Enhanced Geothermal Systems

EMPA Swiss Federal Laboratories for Materials Science

and Technology

EtOH Ethanol

FP Fuel pump

HAZOP HAZard and OPerability analysis

HPSN Hot Pressed Silicon Nitride

IEA International Energy Agency

LOX Liquid Oxygen

LTR Laboratory of Transport Processes and Reactions

NIST National Institute of Standards and Technology

NTUA National Technical University of Athens

OECD Organization for Economic Co-operation

XI

Table of Contents

and Development

P&I Process and Instrumentation

PID Proportional-Integral-Derivative

PLC Programmable Logic Controller

RTD Resistance Temperature Detector

SCWO Supercritical Water Oxidation

SI Spark ignition

SOD Stand-off distance

WP Water Pump

toe Tonne of oil equivalent(amount energy)

Latin letters

b0,1 Fitting parameter -

cp Isobaric specific heat capacity Jkg−1K−1

d Diameter m

h Convective heat transfer coefficient Wm−1K−1

h Enthalpy Jkg−1

l Length m

m Weibull homogeneity parameter -

mf Mass flow rate of the fuel stream kgh−1

mN2 Mass flow rate of the nitrogen stream NLmin−1

mO2 Mass flow rate of the oxygen stream NLmin−1

nT Refractive index -

p Pressure bar

p Thickness ratio -

q Heat flux Wm−2

tr Average residence time in the combustion

chamber s

ur Local rock penetration velocity ms−1

XII

Table of Contents

vair Air injection velocity ms−1

vf Fuel injection velocity ms−1

vO2 Oxygen injection velocity ms−1

x Thermocouple position m

A Surface Area m2

B Condensation potential -

(Cp)r Rock heat capacity Jkg−1K−1

CL Chip aspect ratio -

Cf Ethanol concentration of the fuel stream % wt.

E Young’s modulus Pa

E Electric field Vm−1

G Mass flow velocity kgs−1m−2

I Current A

KV S Valve flow coefficient m3h−1

L Penetration length m

M Momentum flux Nsm−2s−1

P Combustion chamber thermal power kW

Q Electric power W

R Electric resistance Ω

Rio Electric resistance at 0 ◦C Ω

S Seebeck coefficient VK−1

S Sensitivity of a sensor Vm2W−1

T Temperature ◦CTbulk Fluid bulk temperature ◦CTHX1 Oxygen/nitrogen exit temperature

from its heater ◦CTHX2 Fuel exit temperature from its heater ◦CTs1 Temperature of the covered resistance

of the thin film heat flux sensor ◦CTs2 Temperature of the uncovered resistance ◦C

of the thin film heat flux sensor ◦CTsref Surface temperature of the reference sensor ◦C

XIII

Table of Contents

Tssens Surface temperature of the sensor

to be calibrated ◦CTsurf Surface temperature ◦CΔTsr Rock surface temperature change at

spallation Grad

V Volumetric flow rate m3h−1

V Voltage V

Vjet Jet velocity ms−1

Greek letters

α1,2,3 Fitting parameter -

αr Rock thermal diffusivity m2 sec−1

α Layers angle Deg

α Linear thermal expansion coefficient 10−6K−1

α Temperature coefficient of resistivity K−1

βt Rock thermal expansion coefficient 10−6K−1

δ Thickness m

δ Frank Kamenetskii explosion parameter m

ε Dielectric constant -

η Dynamic viscosity Pas

θ Angle Deg

κ Thermal conductivity Wm−1K−1

λ Oxygen-to-fuel ratio -

ν Poisson’s ratio -

ρ Density kgm−3

σ Electrical conductivity Sm−1

σ0 Weibull rock strength Pa

φ Angle Deg

XIV

Table of Contents

Subscripts

al Aluminum oxide cover layer of the resistances in the thin

film sensor

amb Ambient

av Average

b Bridge

bulk Bulk

c Critical

c Point where condensation is complete

conv Convection

cool Cooling water

crit Critical

cw Cooling water

f Fluid

f Saturated liquid state at the bath pressure

ign Igniter

lead Lead wires

pc Pseudo-critical

r Rock

rad Radiation

ref Reference Sensor

s Saturation

s Conditions at the completion of external expansion

sens Sensor to be calibrated

substrate Aluminum oxide substrate plate for the thin film sensor

trans Transverse

w Wall

0 Injection conditions

∞ Ambient

XV

Table of Contents

XVI

Chapter 1

Introduction and project goals

New factors affecting policy and decision making in the worldwide energy

market have surfaced in the course of the last decade, while others that were

taken for granted have changed. The economic depression of 2008 changed

the concept of an ever growing global energy market, as the power demand

in many countries has stalled or reduced, with the notable exception of

e.g BRICS [1]. On the other hand, the nuclear accident at Fukushima

Daiichi has put the nuclear energy sector under renewed criticism. Many

countries, especially the ones not heavily dependent on nuclear energy, ac-

celerated the decision-making processes in order to stop their nuclear power

plants. Moreover, technologies to unlock unconventional gas resources or

the production of bio-ethanol from food crops sustained great damage of

their public image. In addition to that, the fact that solar and wind re-

sources are highly localized has hindered their further development due to

the strain they impose on electricity grids. In this context, the idea of

a centralized, baseload, renewable energy production emerges as a valu-

able complimentary approach in mitigating their side-effects. Geothermal

energy, being such an energy resource, is one of the alternatives capable

to contribute considerably and change the current state of the renewable

energy sector.

An analysis of the trends and players of the world energy market is not the

focus of this work; rather we aspire to reveal the ground-breaking potential

1

Chapter 1 Introduction and project goals

of a drilling technology called hydrothermal spallation drilling. This tech-

nology may revolutionize geothermal energy production and it may also

have an impact on other resource industries. It will be thus useful to out-

line the current energy market situation and the predictions for its future

in order to fully understand its potential.

1.1 Energy supply and consumption

worldwide and in Europe

The rapid economical and financial rise of the two Asian giants, China and

India, in the last two decades and its side effects in the energy markets can

be clearly seen in Fig.1.1(a). The development of the Asian economies is

closely connected to a rapid increase of their energy consumption (both heat

and electricity). At the same time, the developed regions of the world like

the EU-27 and the USA reduced and stabilized their energy consumption

respectively [2].

The global energy market becomes even more interesting by observing the

evolution from the perspective of utilized fuel types worldwide and in the

EU-27 (see Fig.1.1). From this figure one can deduce the customer and

political choices made in this time period, on energy utilization and supply.

The developing countries founded their development on fossil fuels, whereas

the decisions in the EU-27 countries show exactly the opposite trend. Coal

and oil consumption has been dramatically reduced in these countries, while

renewable energy resources have been successfully introduced to the energy

mix. One cannot fail to notice the reduction in the energy consumption

in this part of the world from the year 2005 on, which can be primarily

attributed to the successful energy saving programs and an increased public

awareness [3]. Another very interesting observation from Fig.1.1(c) is that

the nuclear energy production has been stable throughout the last two

decades, contributing roughly 6% of the total worldwide produced energy.

In fact, approximately 25% of the produced energy in the EU-27 comes

2

1.1 Energy supply and consumption worldwide and in Europe

0

700

1 100

1 500

1 900

2 300

1995 1997 1999 2001 2003 2005 2007 2009

China

US

Middle East

Asia

Russia

EU-27

(a) Global total energy production by region (Mtoe). Source : [2].

0

50

300

250

200

150

400

350

100

1990 19961992 1994 1998 2000 2002 2004 2006 2008 2010

CoalNuclear Oil Renewables

Gas

(b) EU-27 energy production by fuel type(Mtoe). Source : [3].

0500

1 5001 000

2 500

3 5004 000

3 000

2 000

4 500

1996 1998 2000 2002 2004 2006 2008

CoalNuclear Oil Renewables

Gas

(c) Global energy production by fuel type(Mtoe). Source : [2].

Figure 1.1: Energy production by fuel type and region (Mtoe).

3


from nuclear sources, thus presenting a remarkably high dependency on its

ambiguous yet aging nuclear plants [3].

A closer look in the evolution of the electricity produced from renewable

resources is equally revealing. A considerable increase can be observed in

the biomass, wind and solar sectors, where the private sector has been

mainly investing due to policy incentives. In comparison, the evolution

of hydro-plants has stalled, mainly because of the very capital intensive

investments required for its exploitation and the lack of resources that can

be developed. When observed in a global context, one could state that

geothermal electricity supply has been almost constant.

Despite the clear trends provided by the aforementioned figures, it is equally

interesting to study the future projections of power generation made from

the International Energy Agency (IEA) in its most recent world energy

outlook [4]. The report analyzes three scenarios for worldwide policies

and the evolution of the energy supply. The new policies scenario, which

is the benchmark of the study, takes into consideration the global policy

commitments and plans that have been announced by mid-2011 on the

energy supply and its environmental impact. The current policies scenario

works on the assumption that the policies and measures enacted by the

same time will be the only ones implemented until 2035. Finally the 450

scenario (calling for a CO2 concentration in the atmosphere of 450 ppm),

suggests measures which result to a 50% chance of limiting the increase

of the average global temperature by 2 ◦C, compared with pre-industrial

levels.

The outcome of the current policies scenario, shows a change of focus in

China from coal-based electricity to nuclear energy. It also predicts that

despite its fast growth, the modern renewable’s sector will supply less than

a single fossil fuel-based production. The new policies scenario predicts a

worldwide increase of the installed global geothermal electrical power from

11GW to 41GW, and the respective electricity generation from 67TWh

to 271TWh. Even the less optimistic scenario speaks for a tripling of

the installed power and the annually produced energy from geothermal

4

1.1 Energy supply and consumption worldwide and in Europe

01 000

6 0005 0004 0003 000

8 0007 000

2 000

9 000

1995 1997 1999 2001 2003 2005 2007 2009

OilRenewablesCoalNuclear

Gas

(a) Global total electricity generation by fuel (Mtoe).

0

200

400

600

800

1000

1200

1400

1600

1990 19961992 1994 1998 2000 2002 2004 2006 2008 2010

HydroSolarBiomassOceanWindGeothermal

(b) Global electricity generation from renewable resources (TWh).

Figure 1.2: Global electricity generation, by fuel and from renewable re-sources. Sources : [2, 3].

resources in the next twenty years. It is obvious that these numbers pose

a great challenge for an energy sector, which has only roughly doubled its

installed power in the years from 1990 to 2012 (see Fig.1.3).

1.1.1 Geothermal energy supply and potential

Geothermal power has been one of the first renewable energy resources to

be utilized for the production of electricity and heat. However, its develop-

ment along the years has not been as fast as that of wind or solar resources.

5


The geothermal energy association reports a global geothermal electrical

capacity of 11224 MWe for the year 2012 [5]. Apart from the USA, among

the leading countries for geothermal electricity supply are Iceland, El Sal-

vador and Kenya and Indonesia. In comparison, according to the data of

the IEA geothermal road-map of 2011 [6] , the total direct (no heat pumps

considered) heating capacity for geothermal heat accounted 15347 MW-

thermal in 2009, with the leading consumer being China with 46.3 PJ/yr.

Figure 1.3 presents the evolution of the global heating and electrical power

0

10

20

30

40

50

60

1995 2000 2005 2010

HeatElectricity

(a) Electrical and heating capacities (GW).

0

100

200

300

400

1995 2000 2005 2010

HeatElectricity

(b) Produced electricity and heat (PJ/y).

Figure 1.3: Global electricity and heat generation from geothermal re-sources. Sources : [7, 8].

produced from geothermal resources [7, 8]. The heat capacity has increased

6

1.2 Cost analysis of geothermal plants

by a factor of five in the last fifteen years, something that has to be at-

tributed to the great success of geothermal heat pumps. By contrast, only

a moderate increase in the electrical capacity is observed. This nevertheless

led to a considerable increase of the produced electrical energy, due to the

high capacity factor of geothermal power plants.

So far geothermal energy exploitation has focused on locations, where the

tapping of an already existing aquifer is possible in depths around 2-3 km.

The majority of this type of geothermal resources occurs in submarine

regions and despite the efforts to utilize them, no actual technical success

has been so far reported [9]. The German Environmental Agency estimated

the global potential for geothermal electricity and heat production at 45

EJ/yr and 1040 EJ/yr respectively, excluding the hot rock and all the

submarine resources [10].

Although these numbers may seem very high, the potential of the geother-

mal resources in the form of hot rock formations is even higher. As temper-

atures increase with depth, generally at 7-100 ◦C/km, hot rocks are present

nearly everywhere. The heat stored in these rocks can be utilized with the

so-called engineered (or enhanced) geothermal systems (EGS) technology.

This technology, which is still under development and demonstration since

the 1970’s, aims to enhance the existing permeability of rock formations.

It then utilizes the formations as heat exchangers to directly tap their heat

and it could be even used in mature geological regions, where the temper-

ature gradient is as low as 7-15 ◦C/km [11].

1.2 Cost analysis of geothermal plants

Despite the great potential of geothermal energy, its exploitation is hin-

dered by the very high initial risks of virtually all geothermal projects.

Solar and wind power resources offer a high investment security through

their easy and cost effective monitoring over long time periods. By contrast,

geothermal resources need very expensive investments for this procedure.

7


Exploration and test wells are required for the evaluation of a geothermal

resource and this upfront investment has a very high risk, due to the high

probability of not finding a suitable resource. Gehringer et al. [12] divided

the development of geothermal power plants in seven stages, and presented

the share of each one to the total investment cost and risk in Fig.1.4. In

Pre-

Surv

ey

Expl

orat

ion

Test

Dril

ling

Fiel

d de

velo

pmen

tPr

oduc

tion

Plan

ning

Dril

ling

Con

stru

ctio

n

Star

t-up

Ope

ratio

n &

Mai

nten

ance

Risk

CostHigh

Moderate

Low

100%

50%

0

Cum

ulat

ive

inve

stm

ent C

ost

Proj

ect R

isk

Figure 1.4: Investment cost and risk for the various stages of a geothermalpower plant development. Source:[12] (adapted).

most geothermal projects, nearly 15% of the total investment costs have to

be spent without really knowing whether a heat resource will be found or

whether it will supply energy in a commercially viable way. This situation

could be tackled with financial solutions, on which Gehringer et al. [12]

concentrated, by finding ways to finance geothermal projects with the re-

sources of the local and global economy. This perspective offers short term

methodologies and is necessary to keep geothermal projects alive. Unfor-

8

1.3 Introduction to spallation drilling

tunately, it is extremely difficult to achieve the expected growth of the

geothermal sector only with these solutions.

An alternative way to tackle this problem would be to find the key tech-

nological issues acting as obstacles and try to overcome them. As Hance

[13] stresses, drilling wells is the most expensive component of any new

exploration stage. Drilling cost depends on the geological characteristics of

the chosen region and the choice of the developer, whether the exploration

wells will also be used as production wells. Furthermore, 80% of the costs

for the confirmation and testing of the resource are well costs. The actual

site development costs consist almost entirely of well costs and the costs for

planning the surface plant equipment. In summary, some 50% of the total

costs for the whole development of a geothermal plant have to be spent

for different kinds of drilling. More than 50% of these costs have to be in-

vested in a high risk environment with an uncertain outcome. We therefore

argue that more research and development efforts should be dedicated to

technologies aiming at the minimization of drilling costs and at new cost

efficient and more reliable exploration methods.


Generally, most of the known drilling processes impose large forces on the

rock mass and increase the stress field, which exceeds the rock strength and

leads to its brittle failure. Most commonly stresses are induced by combined

shear and impact forces. Rotary drilling technologies use a normal force to

advance the bit into the rock, and rotation to produce the necessary shear

stress and remove its cuttings.

These technologies have been developed in the gas and oil industries and

adapted to the rock types encountered in these applications. Geothermal

power, and specifically EGS, requires drilling to greater depths, higher

temperatures and in hard crystalline rock formations. The combination

of these three requirements poses great challenges for conventional drilling

9


technologies and raises the costs of geothermal wells considerably. The

experience of the recent development projects in Soultz-sous-Forets can

provide a benchmark for the expected performance of the rotary drilling

technology when applied for EGS. The penetration rate of the system used

there ranged from 2-7m/h with an average of approximately 3.5m/h, with

each drill bit having an average life-time of approximately 40 h [14]. This

leads to an average drilling depth of around 140m before each drill bit must

be replaced due to wear. In general, data from the gas and oil industry and

from other completed geothermal projects shows a logarithmic increase of

the total well cost as a function of well depth [15].

Spallation drilling is a promising technology, that could prove to be eco-

nomically advantageous over rotary technologies, when applied for EGS.

It takes advantage of the properties of certain rock types, to spall them

to small disk like fragments, due to thermal stresses induced from the im-

pingement of a flame on them. The absence of contact between the bit

and the rock leads to less wear and a long life expectancy of the drill bits.

Furthermore, higher rates of penetration have been demonstrated with this

technology, especially for some crystalline rock types [16]. These two main

advantages, together with the reduced pipe trip time, are expected to lead

to considerable drilling costs reduction, once this technology becomes com-

mercially available.

Preston and White [17] were the first to perform systematic observations on

the spalling mechanisms of ceramic materials, by introducing clay spheres

abruptly in a high temperature oven. They argued that spalling cannot

be attributed directly to the thermal stresses induced in the specimen. It

is rather a result of the propagation and gathering of existing flaws in the

material due to these stresses (Fig.1.5). Their argument was, that the in-

duced compression stresses are not high enough to break ceramic materials,

which typically demonstrate very high compressive strength. They proved

this argument by testing the method on glass spheres and showing that

their spalling was not possible.

In the years after the work of Preston many companies implemented the

10


σ σ

σ σ

σ σ

σ σ

Figure 1.5: Spallation mechanism and steps. Adapted from [17].

spallation principle as a drilling mechanism. The most important of these

companies and their technologies are presented in a report of Sandia Na-

tional Laboratories [18].

The system of Browning, presented in this report [18], consisted of a simple

gas fuel combustion chamber with a venturi nozzle at its outlet to produce

a hypersonic gas jet. Drilling operations with rates of penetration up to

30m/h and holes 331m deep and 50 cm in diameter were reported in a

separate work of Browning et al.[16]. Their experiments showed that the

heat flux and the way it is transferred to the rock surface must be adapted

to the properties of the rock. Moreover, they stressed that the jet velocity

not only controls the heat transfer coefficient but also provides the neces-

sary kinetic energy to sweep the cuttings away. Unfortunately, their data

connected the rate of penetration with the heat flux at the outlet of the

combustion nozzles and not with that on the rock surface. As a result, it

remains case specific and cannot be transferred to other applications.

The most sophisticated theoretical analysis of the spallation phenomenon

11


and the drilling technology has been presented by Rauenzahn [19] and

Wilkinson [20]. Rauenzahn started from the deterministic model of Pre-

ston and then implemented a statistical model for the brittle failure of

rocks. The model is based on the considerations of Weibull [21] that the

strength of a material is statistically dependent on the distribution of phys-

ical flaws preexisting in its stressed volume. The model uses the Weibull

statistical distribution with two characteristic semi-empirical parameters,

which account for the role of the material properties. The one parameter

models the uniformity of the material and the other its intrinsic strength.

Rauenzahn [19] implemented this failure model to a stress condition re-

sulting from a temperature distribution during drilling. He modeled these

temperature-stress fields and developed two expressions, one for the heat

flux during drilling (eq.1.1) and one for the temperature rise on a rock layer

necessary to induce spallation (eq.1.2).

q = (ρ ·C)r ·(1− ν

βt ·E)

·σ0 ·(2 · 0.693π ·C2

L

) 1m

·(m ·ur

ar

) 3m

·ur (1.1)

ΔTs =

[(q

ρ ·Cp

)3

·((1− ν) ·σ0

βt ·E)m

·(2 · 0.693π ·C2

L

)·(m

ar

)3]m+3

(1.2)

In both equations every parameter is a rock property except the incident

heat flux q. Despite his very simplified heat flux measurements, Rauenzahn

has been able to report that heat flux values of approximately 1MW/m2

and surface temperature changes in the range of 500 ◦C are necessary for

spallation to start in granite. He computed the two Weibull parameters

(m,σ0) experimentally for two rock types and concluded that the model

could predict general trends of the process but it is very sensitive to these

two parameters. Wilkinson [20] refined these measurements and performed

additional simulations. He computed, among other parameters, the Stan-

ton number as the percentage of the heat leaving the combustion chamber

nozzle that reaches the rock surface. His results have shown that this num-

12


ber lies in the region 0.3-1%, which leads to similar values for the spallation

heat flux as these of Rauenzahn.

In the latest research contribution Rodrigues et al. [22] found that the

average penetration rate decreases with time and it eventually levels off at

a value lower than that reported from Browning [18]. Gray [23] predicted

this phenomenon by calculating the temperature field in a rock sample

during the extraction of a spall and by illustrating the temperature change

on the surface. The surface that is exposed after a spall extraction has a

higher temperature than that at the initiation of the process. As a result

the heat flux to it decreases, since the jet temperature and its heat transfer

coefficient stay the same. He concluded that this process leads to a drilling

steady state for every combination of rock type and incident heat flux.

Although spallation drilling showed a great potential, especially in crys-

talline rocks, its implementation has been limited to quarry extractions

and drilling for deep geothermal resources has never materialized. Its ma-

jor drawback was the fact that drilling deeper wells - deeper than 350m -

was connected to boreholes being filled with a water based drilling mud.

The combustion chambers and their drilling jets could not be implemented

in water under high pressure and the system has been considered impossi-

ble to use [24]. As in the late eighties and early nineties the first diffusion

flames have been ignited and continuously operated in water under high

pressure, the initial problem has been overcome giving rise to hydrothermal

spallation drilling. Hydrothermal spallation drilling is a term first used in

the work of Augustine [25] to signify the use of hydrothermal flames in

spallation drilling. This version of spallation drilling is the topic of the

following section.

13


1.4 Hydrothermal spallation drilling - Goals

of the thesis

Water above 22.064MPa and 647K has considerably different physical

properties than when it is at its liquid phase. Its molecules become less

polarized and its dielectric constant falls to values comparable with the

ones of non-polar gases (see Fig.1.6). Both organic substances like a fuel

and non-polar gases like oxygen are soluble in supercritical water, and thus

a flame can be ignited in it. The resulting flames have been initially used

to destroy organic waste in high pressure vessels [26].

Schilling [27] and Steinle [28] were the first to operate diffusion flames

in supercritical water. Although their flames were ignited and operated

batch-like, important data on their self-ignition, their efficiency and on

high pressure explosions has been produced. La Roche [29] and Weber

[30], developed the idea further by igniting and continuously operating

hydrothermal diffusion flames. In water-filled boreholes, three kilometers

0200

400600

8001000

100

80

60

40

20

00 0.2 0.4 0.6 0.8 1

T[°C]

ρ[gcm ]

ε

-3

Figure 1.6: Dielectric constant values of water, adapted from [31].

14

1.4 Hydrothermal spallation drilling - Goals of the thesis

deep, water exceeds its critical pressure and hydrothermal flames can pro-

vide the required heat to spall the rock.

A hydrothermal spallation drilling head consists of a combustion cham-

ber, the high temperature combustion products of which form a hydrother-

mal jet. This flame-jet is directed to the rock surface to induce spallation

through its impingement. Although the concept seems to be straightfor-

ward and easy to implement, the use of hydrothermal flames as free jets

in liquid water introduces new challenges for the technology, which have

remained unsolved in the research so far (see Fig.1.7).

Entrainment• High density differences••

•

•

High temperature differences

→Rapid temperature decayEntrainment control

Heat transfer coefficient• Crucial for spallation performance

• Heat flux sensors developement• Dependency on operation conditions

Rock mechanics - Drilling procedure• Hole size modeling

• Fracture model and drilling velocity predictions

• Interaction flame - rock

Combustion in an aqueous environment

Forced ignition of hydrothermal flamesTemperature and power control

Figure 1.7: Hydrothermal spallation drilling challenges.

The forced ignition of the flames, the jet formation in a aqueous environ-

15


ment and the heat transfer optimization from it to the rock surface, define

the principal aims of the spallation drilling project in the LTR. In paral-

lel with these fluid dynamics and thermodynamics problems, the rock-jet

interaction has to be investigated in depth. New fracture models must be

developed, in order to optimize and adapt the drilling technology to the

requirements of the deep geothermal well drilling.

1.4.1 Thesis structure and goals

The research on hydrothermal spallation drilling is by definition interdis-

ciplinary. It draws on know-how from and contributes innovative results

to a number of various scientific fields, among which combustion and heat

transfer in supercritical fluids and high pressure technologies. Conceptu-

ally, spallation drilling consists of individual process modules, for many of

which no data existed in the kick-off of this project. Some of these modules

are presented on Fig.1.7.

• Forced flame ignition. While the ignition and control of normal

flames at ambient conditions are well known procedures, no forced

ignition has been achieved with hydrothermal flames so far. This

is the first addressed topic of this thesis and the technical solutions

together with the scientific results are presented in chapter 4.

• Water entrainment in the flame-jet. The entrainment of cold,

high-density water to the high-temperature, low-density flame-jet is

crucial for hydrothermal spallation drilling. It contributes consid-

erably to the total heat losses of the flame before its impingement

and it is reducing its length. The physical phenomena connected to

this problem, methods for its characterization and its control and the

respective experimental results are presented in section 5.4.

• Hydrothermal flame impingement heat transfer. The convec-

tive heat transfer coefficient of impinging hydrothermal flames models

the interaction of the fluid with the rock during spallation drilling.

The results of this interaction are temperature and thermal stress

16

1.4 Hydrothermal spallation drilling - Goals of the thesis

distributions in the rock mass. The development and calibration of

heat flux sensors, capable to operate in the vicinity of a hydrothermal

flame, is a significant goal of the thesis and the results on this field

are presented in section 5.1.

• Thermal fracturing of rock. The actual drilling experiments and

the development of the thermal fracturing models conclude the study

of hydrothermal spallation drilling. Although this research module

is not included in the scope of the present thesis, the limited exper-

iments presented in its outlook, form the basis for future investiga-

tions.

Apart from the scientific topics mentioned above, several technical and

engineering aspects of the measurements and the operation of the high

pressure process belong to the long list of obstacles, which had to be over-

come. A spallation drilling pilot plant, the central experimental facility

for the research undertaken, has been built from scratch, together with all

its measurement and control systems. As commercially available heat flux

sensors were unsuitable for the intended application, new ones have been

designed and constructed. Complementary to these sensors a calibration

methodology has been developed and a calibration facility has been built.

The sensors can be now tested and calibrated at conditions simulating the

ones in their intended operation. A whole novel ignition system has been

produced for the ignition of the flames, capable of conducting electricity in

extreme temperatures and pressures.

The structure of this thesis attempts to give a detailed overview of all

the challenges, technical issues and innovative solutions developed over its

course. The difficulty of doing so lies, as stated, in the interdisciplinary

character of the project. Moreover the fact that solutions were developed

as an appropriation and expansion of already existing ideas and concepts

from widely varying fields makes this task even harder. So while the outline

of this thesis follows as closely as possible the steps of designing, building,

commissioning and operating the hydrothermal spallation drilling experi-

mental unit, it is necessary to address some of the peripheral issues, where

17


innovative concepts had to be incorporated. The design decisions and the

specifications of the equipment are justified in the light of recent literature

and the needs of the project.

18

Chapter 2

Plant design, construction and

commissioning

The hydrothermal spallation drilling pilot plant models experimentally the

conditions in a borehole, at depths interesting for hydrothermal spalla-

tion drilling. Its concept is based on existing supercritical water oxidation

plants, but the adaptation of its size and capacity to the drilling process

introduced new challenges to its design, construction and commissioning.

In the following paragraphs, the planning procedure of the plant along with

its detailed safety analysis are summarized. Several details of the built fa-

cility are outlined and the commissioning of parts crucial for the subsequent

measurements is analyzed.

It should be pointed out that the data presented is not comprehensive of the

entire design and construction process. Our criteria for this selection were

on the one hand to convey an understanding of the operating conditions

and on the other hand to document a novel procedure and solutions that

are crucial to further research.

19

Chapter 2 Plant design, construction and commissioning

2.1 Process design

The process of the facility is similar to that of the existing high pressure

plant presented in the work of Weber [30]. The core specifications of the

process, which define the size of the equipment as well as the necessary

safety measures can be summarized as follows:

• The maximum fuel thermal power.

• The upper pressure and temperature limits.

• The fuel and oxidant used for the combustion reaction.

• The disposal of the produced effluent stream.

Successful drilling experiments rely on the availability of thermal power in

the plant. The plant would be poorly designed if the upper limit of its

capacity would be reached without spallation occurring. As a result of the

limited data on hydrothermal spallation drilling, the dimensioning of the

facility exhausted the utility limits provided to it. These limits in terms of

cooling water, compressed air and electricity, dictated a plant fuel thermal

capacity of 120 kW.

The lowest operating pressure of the plant is set by the critical point of

water, resulting in a minimum value of 220 bar. In a similar way, the

maximum operating pressure was chosen according to previous experimen-

tal investigations on hydrothermal flames, while also taking into account

the financial limitations of the project. Hence, an upper pressure limit of

350 bar was chosen.

Safety issues pose stringent limits on the maximum temperature (without

combustion occurring) of any surface of the equipment. Most of the ex-

isting supercritical water oxidation (SCWO) plants operated with reactant

temperatures reaching 550 ◦C [30]. Relying upon our experience, ATEX

specifications and the commercially available equipment, this temperature

was chosen equal to 420 ◦C for our plant.

20

2.1 Process design

Various fuels and fuel mixtures have been used for hydrothermal flames

in the course of the years. La Roche [29] and Weber [30] used mainly

methanol-water and methane-water mixtures. They reported poor mixing

of methane with water close to the critical point of the latter and were

forced to use higher reactant temperatures, to minimize the observed micro-

explosions. Steinle [28] used methane, ethane, methanol and hydrogen, but

his experiments were conducted in high pressure explosion vessels and are

not directly relevant to the current project. For the choice of the fuel in

our case, a list of chemical and physical specifications comprises:

• The fuel should be water-soluble and liquid in atmospheric conditions.

• Avoidance of harmful or toxic substances.

• The shortage of storage space called for a fuel that is easy to store

and low-cost.

• The heating value of the fuel and its mixtures with water should lead

to a manageable size of the fuel pump.

Once an exhaustive list of fuels had been prepared, ethanol was deemed

the optimal solution considering the above criteria. Liquid propane was

considered as the second best fuel. The fuel pump can also operate either,

in case higher fuel power values are necessary.

The choice of the oxidation medium has been easier, because only air,

oxygen (liquid or gas) or hydrogen peroxide were possible. After a short

market assessment, air was excluded because of the resulting compressor

size, whereas liquid oxygen and hydrogen peroxide presented very high

safety hazards. The only choice, leading to realistic equipment sizes and a

reasonably safe operation, was gaseous oxygen.

By all accounts, the disposal of the effluent stream must not cause dan-

gerous situations. In addition to that, the expansion of the effluent stream

from the operating pressure to atmospheric pressure should be carried out

in a technically sound way. To achieve these goals, we limited the outlet

temperature of the effluent stream to 80 ◦C and its disposal took place in

21


the sink of the building.

2.1.1 Process description

The process comprises four sub-systems responsible for the fuel, and oxygen

compression and control, the cooling of the plant and the disposal of its

effluent stream respectively.

In the fuel system, DI water and ethanol are mixed by a three-way valve

(3WV-1, in Fig.2.1) before their inlet to the fuel pump (FP-1). The mixture

flows through a pulsation dampener, and its pressure is measured at the

outlet of the pump (PI-1). The mass flow meter (FMI-1) additionally

records its density and is followed by a non-return valve and a heater (HX-

2). The outlet temperature of the latter is measured and controlled (TIC-2)

and a second non-return valve precedes the fuel injection in the pressure

vessel.

Oxygen is compressed from the storage bottles to a 300 bar reservoir, the

outlet pressure of which defines the inlet pressure for the oxygen line and is

controlled by a front pressure controller (PIC-1, in Fig.2.1). The flow rate

is then set from a flow controller (FC-1), which is followed bya non-return

valve and a manometer. The oxygen flow is finally heated electrically (HX-

1) in the same way as the fuel, and injected in the pressure vessel.

For redundancy, two high pressure pumps (WP-1 & WP2) deliver the cool-

ing water for the plant. WP-1 is fed directly from the DI water network,

and it provides water to the cooling mantle of the pressure vessel. The flow

of this stream is measured with a flow meter (FMI-3) and controlled with

an inverter on the pump. WP-2 is fed from a storage tank (2.5m3) and it

is responsible for the combustion chamber (CW1) and the effluent stream

(CW3) cooling water. The mass flow reaching the plant from this pump

is manually controlled with a by-pass valve (V-7). On the other hand, the

mass flow of the CW1 stream is measured and controlled by a flow meter

(FMI-2) and an needle valve (FC-2) respectively. The two cooling water

22

2.1 Process design

Oxygen

HX

2

NR

V -

10

TIAC

H

RV

-10

OC

-1V

-1N

RV-

1

Fuel

CW1

tank

(2.5

m )

3

SV

-1

PD 1

N2

NR

V -

8

PIC

-1

TIAH–4

CW3 F

2

P&

T -1

BS-8B

S-7

FIC

Fuel

PI -

1

Man

omet

erV

-2N

RV

-2

CW2

HX

1

TIAC

H

PIC

TIAH -1TIAH -2

TIC

-1

TIC

-2

Com

pres

sed

air n

etw

ork

SV

-2

F 3

V-5

V -

6N

RV

–7

TIAH–3

PD

3

PD 2

Drin

king

wat

er

O2

O2

FMI -

1

FMI -

3

NR

V -

6

NR

V-4

NR

V -

9

FIC

PI-5

FP -1

WP

-2

WP

-1

Psw

itch-

1P

I-4

Psw

itch-

2P

I-2

Psw

itch-

3P

I-3

NR

V -

5

NR

V-3

V-1

0

V-1

2

BS

-4BS

-3

BS-5

BS-6

BS

-1BS

-2

SV

-3

V-7

3WV

-1

FC-2

FMI-2

FC-1

DI w

aterD

I wat

er

DI w

ater

Figure 2.1: Plant piping and instrumentation diagram.

23


lines are connected with two valves (V 10 and V 12 in Fig.2.1), which open

when the pressure in the vessel is below 250 bar.

The pressure vessel has four outlets, two for its cooling mantle and two

for the outlet stream of its main chamber. The temperature at each outlet

is measured (TIAH-3&4) and the two streams are mixed after their exit

from the vessel as a first cooling step of the combustion products. The

second cooling step is realized by mixing the resulting stream with the

CW3 steam. A safety valve (SV-3) and a bursting disc (BS-8) provide

overpressure safety to this line and are followed bya filter and a combined

pressure and temperature measurement (P&T-1). The last component of

the plant is the pressure controller (RV-10), which reduces the pressure to

atmospheric.

2.1.2 The HAZard and OPerability study

The complexity of the process and the scale of the plant required a detailed

safety analysis of its operation. Several methodologies have been examined,

and the HAZOP analysis was deemed appropriate for our continuous pro-

cess. An analysis of both the process and the pressure vessel was carried

out, just after the completion of their conceptual design and before starting

the detailed design.

Core of the HAZOP is the careful definition of the functional parameters

of the plant process and its vessel. A thorough application of the method

on each of these parameters identifies the potential deviations from their

targeted values. The method is finalized by formulating strategies either

to avoid deviations or to reduce their impact on the facility [32]. At the

beginning of the analysis a draft sketch of the pressure vessel and an inter-

mediate version of the process P&I Diagram were available (see Fig.2.2).

The safety-relevant parameters of the vessel and the operation of the facil-

ity at full thermal power (and thus the highest risk) were used as functional

specifications for the safety analysis and are presented in Table 2.1.

24

2.1 Process design

12mm ; 300 bar

Oxygen

12m

m ;

300

bar

Gas

-liqu

id s

epar

ator

Rea

ctor

HX

2

NR

V-8

NR

V-5

TIA

CH

TIAC

H

TIC

Pre

ssur

e re

gula

tor

Wat

er p

ump

Fuel

Pum

p

Oxy

gen

Com

pres

sor

12m

m ;

300

bar

3/16

' ; 3

00 b

ar

1’ ;

1bar

V-8

RV

-1

RV

-2V

-5

V-2

V-6 V-3

¼’,2

00ba

r

V-1

NR

V-1

Fuel

CW1

Des

alin

ated

wat

er

tank

(3 m

3)de

salin

ated

wat

er

SV

3

PD

2

PD

1

FC-5

3WV

N2

O2

¼’,2

00ba

r

Par

ticle

Sep

arat

or

TI

12m

m ;

300

bar

NR

V-7

NR

V-6

PIC

Air

inje

ctio

n –

Dill

utio

n an

d co

olin

g

TIC

CW3

L

F 1

F 2

1’ ;

1 ba

r

SV

-6

SV

-5

12m

m ;

300

bar

5/16

' ; 3

00 b

ar

FIC

HX

3FC

-4

FIC

¼’ ;

300

bar

FC-3

FIC

Fuel

3/16

' ; 3

00 b

ar3/

16' ;

300

bar

V-4

NR

V-3

V-7

NR

V-4

3/16

' ; 3

0 ba

r

SV

1FC-2

FIC

NR

V-2

1/8'

; 300

bar

FC-1

FIC

CW2

HX

1

TIA

CH

1'; 1 bar

PIC

1'; 1 bar

TI

1/8'

; 300

bar

12m

m ;

300

bar

TI

12mm ; 300 bar

5/16

'; 30

0 ba

r

TITITI

TIC

TIC

RV

-31'

; 1b

ar

RV

-4

Air

com

pres

sor

Air

pipe

Air

pipe

SV

2

F 3

Figure 2.2: Plant piping and instrumentation diagram at the beginningof the HAZOP safety analysis.

25


Table 2.1: Functional specification for the pressure vessel and the process.

Pressure vessel functional specificationSafe, continuous spalling of rock probes at a pressure of 250 barand at wall temperatures < 400 ◦C, and outflow of the productsin a sub-critical state (300 ◦C).

Process functional specificationContinuous spalling of rock probes in spalls (<10mm) with com-bustion of 20 g/s of a supercritical (50wt% Ethanol/Water) mix-ture (p=250 bar, T=400 ◦C) with oxygen (p=250 bar, T=400 ◦C,λ = 1, 2), with a simultaneous injection of 140 g/s cooling waterat a maximum temperature of T=100 ◦C.

Safety analysis results

The safety analysis not only influenced the design of the plant and the

pressure vessel considerably, but it also produced valuable input for the

evaluation of emergency and normal operational procedures.

The first result was that the likelihood of a potential explosion must be min-

imized, due to the high pressure in which the combustion reaction takes

place. An accumulation of combustible mixture in the volume (approx.

5.8 Liters) of the pressure vessel could cause an explosion with dire con-

sequences for persons and equipment. The most efficient measure against

this hazard is to provide water to the vessel irrespective of the operational

condition of the facility. A second water pump along with its tubing net-

work were introduced in the system to provide redundancy of the cooling

water supply. Moreover, the two water networks were connected on two

points with on-off valves, which are closed during normal operation and

open only during an emergency (i.e a pump outage).

Although the dangers from the combustion reaction are minimized through

these measures, the dangers of overpressure and overheating still remain.

A double safety system, with safety valves operating at 350 bar as a first

protection level and bursting discs at 400 bar as a second, protects the

26

2.1 Process design

plant from overpressure. In the same way, the points of the plant mostly

endangered from overheating were identified and the control system trips

the plant if their temperature exceeds a safety value.

In the case of electricity failure, a safe state of the plant was ensured by

choosing the pressure (RV-10) and CW1 mass flow (FC-2) controllers as

normally-open valves and the oxygen flow controller (FC-1) as a normally-

closed. As a last resort, a connection with the potable water network was

introduced to wash out all the remaining fuel from the system.

In the operational aspects of the plant, the safety analysis resulted in the

start-up and shut-down procedures as well as all the emergency procedures.

It was concluded that no piece of equipment should be heated-up or cooled-

down too fast, and a slow cool-down process was chosen as an emergency

shut-down procedure. The last resort in case of loss of containment in

the tubing is the interruption of the electricity supply to the plant. The

measures taken when an emergency arises are as follows:

• Step 1. Switch 3WV-1 to water. According to the residence time

measurements of section 2.3.3 only water will be present in the fuel

line after six minutes. At the same time both on-off valves (V-10 &

V-12) open and the two water networks are connected to each other.

• Step 2. Switch FC-2 to a fully open position. The maximum cooling

capacity of the combustion chamber is ensured.

• Step 3. Switch the set-point of both heaters to zero. No further

heating occurs.

• Wait step 1. Wait 15 minutes to make sure that no flame is present

in the system.

• Step 4. Shut-down the fuel pump and the oxygen compressor.

• Step 5. Reduce the pressure of the system in steps of 50 bar, and

switch both water pumps off when the pressure controller reaches its

fully open position.

27


Reactor

PD 1

PI -1

FMI -1

FP-1

BS4

HeaterHX-2

NRV – 6

First storey – 2m

NRV – 5

BS3

Manometer

Ethanol

DI-Water line

Tank

3WV-1

TIC - 2

(a) P&I diagram of the fuel network.

(b) The fuel pump and its suction line with thethree-way valve.

(c) Fuel heater (on the rightcorner) and the fuel injectionline.

Figure 2.3: The fuel sub-system as built.

28

2.2 Plant construction


After several revisions and the integration of the HAZOP study, the pro-

cess and instrumentation diagram of the plant took the form presented

on Fig.2.1. A description of the tasks of each network of the plant was

presented in section 2.1.1. In the present section some details on the in-

dividual pieces of equipment will be explained, along with their technical

characteristics.

2.2.1 The fuel sub-system

The fuel sub-system performs for all the functions connected to the fuel

stream. The mixing of water and ethanol, the metering of the total mass

flow of the mixture, its compression and its heating, all take place within

this sub-system.

A high pressure membrane metering pump (see Fig.2.3(b)) was chosen, due

to its flexibility in compressing various fuels and its increased operational

safety. The pump is able to reach pressures up to 500 bar and its mass flow

rate can be controlled manually with a stroke adjuster and automatically

with an inverter. A coriolis mass flow meter measures the flow rate and

the density of the fuel, which are respectively used as measured process

variables for the inverter and the controller of 3WV-1.

The electrical flow heater of the fuel consists of a coiled tube made of

stainless steel that is cast in an aluminum block, which is in turn electrically

heated. The outlet temperature of the heater is measured with a K-Type

thermocouple (TIC-2) and is controlled with a standard PID controller.

This controller sets the output of a stack of triac power switches, which

are responsible for the control of the electrical power of the heater. The

system is ATEX compatible and the temperature of the aluminum block is

monitored from a three-wire RTD. The latter is connected to an electrical

switch that turns the heater off if the temperature of the aluminum block

29


exceeds 440 ◦C.

Reactor

BS-1

HeaterHX-1

NRV – 4

First storey – 2m

OC-1

PIC - 1

O2

BS-2NRV – 3

Manometer

TIC - 1

FC-1

(a) P&I diagram of the oxygen sub-system.

(b) Oxygen compressor &flow controller.

(c) Oxygen heater.

Figure 2.4: The oxygen sub-system as built.

30


2.2.2 The oxygen sub-system

The oxygen sub-system is built on a principle similar to the fuel system

and its P&I diagram is presented on Fig.2.4(a).

High pressure oxygen tubing must comply with standards defining the flow

velocity in it, its materials and its cleanness [33, 34, 35, 36]. Due to the high

pressure and temperature of oxygen in our plant even the most resistant

materials for oxygen service (i.e. MONEL) could not be classified as non-

combustible. As a consequence, we used high quality stainless steel (1.4571)

for the tubes and we focused on the limitation of the flow velocity in them

and their thorough cleaning. After maintenance, the tubing is cleaned

with appropriate solvents, and then purged with a high velocity nitrogen

flow. This nitrogen flow removes all the particles remaining in tubes in

order to minimize the chance of impingement ignition on the walls of the

tubes. Purging is also carried out at the end of every experimental day,

to prevent the accumulation of stagnant, high temperature oxygen in the

system. Finally, all the tubes are electrically grounded to minimize the

effects of static electricity.

The oxygen is provided from a bundle of twelve bottles (50 Liter- 200 bar)

and it is fed to the compressor through a filter. The oxygen compressor

comprises two air driven pressure boosters, each of which has two compres-

sion stages and is able to compress oxygen up to 300 bar from a minimum

feed pressure of 50 bar. The gas is compressed to a high pressure reservoir,

the pressure of which is measured by two pressure switches with a lower

and an upper set-point. Each switch controls one booster and switches it

off and on at the upper and lower pressure set-points respectively. The

oxygen pressure at the inlet of the plant (at the outlet of the compressor)

is controlled with a forward pressure controller (Tescom ER300).

The gas flow rate is controlled from a thermal mass flow controller, oper-

ating between two controlled pressure values (PCI-1 and PI-5 in Fig.2.1).

The flow controller produces the pressure drop that corresponds to the de-

sired mass flow rate and the values of these two pressures. The oxygen

31


heater (HX-1) and its control system are similar to the fuel heater, with

small differences in the material of the coiled tube.

2.2.3 The CW2 sub-system

The CW2 sub-system delivers water to the cooling jacket of the pressure

vessel ( refer to its P&I diagram in Fig.2.5(a)). A high pressure three-head

plunger pump (WP-1) is used for the compression of the water. Its mass

flow rate is measured with a coriolis flow meter (FMI-3) and the signal is

fed to an inverter, which controls the rotational velocity of the pump and

thus its mass flow rate.

The line is connected to the second cooling water line of the plant (see

section 2.2.4) by two valves (V-10 and V-12), the installation points of

which were determined during the HAZOP analysis. The operation of the

pump is monitored from a pressure switch-transducer (PI-4), which opens

the valves V-10 and V-12 in case the pressure in the line falls below 250 bar.

2.2.4 The CW1&3 sub-system

The CW1&3 sub-system delivers the water streams for the combustion

chamber (CW1) and for the cooling of the effluent stream of the plant

(CW3). The injection points are presented in section A.2 and the P&I

diagram of the sub-system can be seen on Fig.2.6(a).

A 2.5m3 DI water storage tank is used to feed the high pressure three-head

plunger pump (WP-2). This water volume is enough to operate the plant

independently for an hour, in the case of an emergency failure of the water

utility system. WP-2 operates always in full capacity and the by-pass valve

(V-7) controls the water flow reaching the plant. The pump operation is

monitored from a pressure switch-transducer (PI-2), in the same way as PI-

4 controls WP-1. Furthermore, a magnetic valve (V-6) connects the facility

to the potable water utility network, for the case of electricity failure (see

32


Reactor

NRV – 10DI-Water

V-12

PD 3

FMI-3WP-1

Connections with CW 1&3 network

SV-2

V-10

P-Switch 1 PI-4

BS - 6Manometer

(a) P&I diagram of the cooling water 2 sub-system.

(b) The high pressure pump of the cool-ing water 2 sub-system.

FMI-3

SV-2

(c) Cooling water 2 sub-system.

Figure 2.5: The CW2 sub-system as built.

33


section 2.1.2).

Reactor

CW1

DI-Water

CW3V-5 V-6 NRV - 7

PD 2

Potable water

NRV - 9

WP-2 Manometer

Connection with CW2 Connection with CW2

SV-1Pswitch 2PI-2

Pswitch 3PI-3

BS-6V-7

V-10

V-12

FC-2FMI - 2

(a) P&I diagram of the cooling water 1&3 sub-system.

(b) The high pressurepump of the cooling wa-ter 1&3 sub-system.

V-10

SV-1

FMI-2 FC-2 V-12

(c) Cooling water 1&3 sub-system.

Figure 2.6: The CW 1&3 sub-system as built.

The flow that reaches the plant is divided in the two respective streams

(CW1 and CW3) at a point next to valve V-10. The flow rate of the CW1

stream is measured with a coriolis mass flow meter (FMI-2), the signal of

which is used to control the opening of the needle valve (FC-2) with a PID

controller. The rest of the aforementioned flow is led automatically to the

CW3 line.

34


2.2.5 The effluent water sub-system

The aim of the effluent water sub-system is to cool the combustion products,

and reduce the pressure of the stream from the working pressure to the

atmospheric.

The temperatures of the streams leaving the pressure vessel of the plant are

measured directly at their exits. After mixing the streams with the CW3

water stream, the temperature and the pressure of the resulting stream

are measured with a combined sensor (P&T -1) just before the pressure

controller (RV-10). This temperature value has a limit alarm of 80 ◦C, sothat no water vapor is produced due to the pressure reduction from 250 bar

to 1 bar. RV-10 is additionally protected from a high pressure metal-mesh

filter with a cut-off of 80 �m.

To achieve pressure reduction and control a needle valve with a triple, com-

bined stem and seat is utilized that reduces the pressure in three consecutive

stages.

2.2.6 Control system

The completion of the mechanical construction of the plant (October 2010)

was followed by the construction of the electrical and the control system. A

programmable logic controller (PLC) with a Profi-bus control network was

chosen. The design and setup of this part of the plant was carried out in

cooperation with a specialized company. The control system architecture

of the plant is presented in Fig.(2.8).

The communication of the user with the control system is realized via a web

server, having a dedicated IP address and a touch panel. The whole system

is built in a control rack, which performs all the control sequences for the

plant and is also responsible for the emergency procedures. Additionally,

an I/O box is located near the pressure vessel, where the local analog

signals are digitized and transported via Profi-bus to the PLC for further

35


RV-10

CW3

Filter P&T 1

PIC

TIAH–3

To Low pressure

TIAH–4

NRV - 8

SV -3

BS-8

(a) P&I diagram of the effluent water sub-system.

(b) The effluent water sub-system.

(c) The pressure controller of the plant.

Figure 2.7: The effluent water sub-system as built.

36


Table 2.2: Technical characteristics of the most important components ofthe plant.

Name Function Technical specifications

FMI Flow meter(Coriolis) Accuracy -Flow 0.15% -Density 0.0005 g/cm3(Repeatability)

Pswitch Pressure switch Set point:250 bar, Accuracy 0.1%SV Safety valve pset=350 barBS Bursting disc pset=400 bar3WV-1 Mixing Valve KV S =0.15m3/h for both streams

FP-1 Fuel Pump Vmax =100 l/h, pmax=550 barHX-2 Fuel heater pmax=350 bar, Tmax = 420 ◦C

Power=65 kW, V=4.3 L

OC-1 Oxygen compressor pmax =310 bar, Vmax=500N l/minFC-1 Flow controller 12-600N l/min ,Accuracy 1%HX-1 Oxygen heater pmax=350 bar,Tmax = 420 ◦C

Power=10 kW, V=0.8 L

WP-1 Water pump pmax =400 bar, Vmax=1.5m3/h

WP-2 Water pump pmax =400 bar, Vmax=3m3/hFC-2 Needle control valve KV S =1.6m3/h,pmax =400 barFilter-2 High pressure filter lcutoff=80 �mRV-10 Needle control valve KV S =0.4m3/h(linear)

validation. Finally, a personal computer can be connected to the web-server

to monitor the experiments and perform online data sampling. Details

regarding the functionality ad operating philosophy of the control system

are presented in the following section concerning the commissioning of the

plat.

37


PC

GUI (Browser)

Ethernet

Profibus DP

Web server

D I/OA I/O

Inverter FP1Inverter

WP1 Control Rack HX2

Control Rack HX1

Con

trol R

ack

Plan

t

Special signals -Heat flux

Ignition etc.

WP2I/O BoxFlow meters

D I/OA I/O

HX1 HX2

PLC

Figure 2.8: Control system architecture.

2.3 Plant commissioning

The commissioning phase aimed at testing all aspects of the plant oper-

ation, with a primary focus on the safety procedures. All the internal

communication systems of the plant were tested and every controller was

tuned. Furthermore, the nominal operation limits of the sub-systems were

verified by operating the plant for 10 hours at these limits. The highest

pressure test was conducted at 300 bar, while the fuel and oxygen heaters

reached maximum temperatures of 420 ◦C and 400 ◦C respectively.

The following sections report the commissioning details of three sub-systems

38


of the facility, which were crucial for the intended experiments.

2.3.1 Commissioning and temperature control of the

heaters

The adequate control of the heaters ensures the reproducibility of experi-

ments and the safe operation of the plant. Moreover, the duration of their

heat-up and cool-down procedures must be subtracted from the 10-12 hours

of a normal experimentation day, because the facility must be supervised

while operating without shift service. Hence, the number of experiments

that can be conducted in one day depends on the duration of these proce-

dures.

Commissioning the heaters included three types of experiments: at a first

stage their maximum operational limits were confirmed. The second phase

consisted of tuning their PID controllers and thirdly the duration of their

heat-up and cool-down procedures was determined. During these experi-

ments DI water and nitrogen were used and the plant operated at 260 bar.

The heaters are controlled from stand-alone racks, each one communicating

with the PLC via digital and analog signals. A PID controller operates

semiconductor power switches, which control the electrical power and the

outlet temperature of the heaters. The results of the controller tuning for

the oxygen heater are presented in Fig.2.9(a).

Fig.2.9(b) presents the temperature difference between the aluminum block

and the outlet temperature of the fuel heater. The block temperature was

measured with one of the two RTDs that are cast in it for its overheating

protection. This temperature difference was approximately 30 ◦C, for low

and medium temperatures and became smaller, the closer the medium was

to its pseudo-critical temperature. The oxygen heater operated with higher

temperature differences, due to the lower heat transfer coefficient in its

tube. This is the reason for its low heating rate, since higher rates could

39


0 20 40 60 80 100 120 140 160 180 2000

50

100

150

200

250

300

350

400

450

Time [min]

T[◦C]

T

HX1Setpoint

(a) Temperature control of the oxygen heater. With no control steps, theheat-up time is two hours. Depending on the flow rate of the gas, the cut-offswitch (Tblock >440 ◦C) is activated at different THX1. Here THX1 =380 ◦C.

0 5 10 15 20 25 30 35 40100

150

200

250

300

350

400

Time [min]

T[◦C]

T

BlockT

HX2

(b) Temperature difference between aluminum block of the fuel heater (HX2)and its medium.

Figure 2.9: Commissioning results of the heaters.

40


not be sustained from the gas side, and a control of the gas temperature

would be harder. A heating circle from room temperature to 400 ◦C takes

two hours for the oxygen and one hour for the fuel heater.

The cool-down experiments, showed that the oxygen heater is practically

impossible to cool in a realistic time period. Hence, only the fuel heater

was cooled down to an outlet temperature of 90 ◦C, limiting the respective

time to approximately 90 minutes. At the end of each day the oxygen line

is purged with nitrogen, to avoid having heated and stagnant oxygen in it

for a long time period.

2.3.2 Pressure control

The pressure in the vessel of the plant is the most important and sensitive

parameter in the facility. All the experiments are conducted in the trans-

critical regions of the involved fluids and small pressure perturbations could

make the reproducibility of the experiments impossible. From an opera-

tional point of view, the oxygen mass flow controller adapts its opening to

the pressure difference between the inlet of the gas line and the pressure

vessel. Since all the oxygen flow controllers are slow to adapt their opening

(in order to avoid ignition of their seat from the medium) the pressure of

the vessel must be as stable as possible. The principal aims of the pres-

sure control were the optimization of its stability and accuracy. In parallel

the erosion of the seat and stem of the pressure control valve had to be

addressed and minimized.

The pressure control loop is realized with a software based PID controller

(programmed in the PLC of the plant), the measured process variable of

which is the pressure in the vessel of the plant. Its output is used as the

set-point for the position controller of the valve RV-10. Initially a single

stage control valve was installed to reduce the pressure from 260 bar to

the atmospheric one. The first tuning of the controller led initially to a

stable and reproducible pressure control. This stability lasted for short

time periods (5-10 minutes) and it was followed bypressure fluctuations

41


0 2 4 6 8 10 12 14 16 18250

255

260

265

270

Time [min]

Pressure

[bar]

PressureSetpoint

(a) Pressure control during the first commissioning stage. The valve has anequal percentage characteristic and KV S =0.63m3/h.

0 2 4 6 8 10 12 14 16 18250

255

260

265

270

Time [min]

Pressure

[bar]

PressureSetpoint±0.7%

(b) Pressure control after optimization. The new valve has a linear char-acteristic, KV S =0.4m3/h and a buffer is implemented in the controllerintegration term.

Figure 2.10: Commissioning experiments for the pressure controller.

42


(see Fig.2.10(a)).

The observed pressure fluctuations were attributed to the controller param-

eters and the valve hardware. Small stochastic changes of the pressure are

to be expected in a plant with three pumps and a gas compressor. In the

present case these changes eventually added up and the output signal of

the PID controller exceeded the input signal resolution (0.3%) of the valve

pneumatic actuator. As a result, the opening of RV-10 changed although

the mean value of the controlled quantity stayed the same. The KV S value

of the valve and its equal percentage KV - characteristic led to the pressure

fluctuations. A validation of these instabilities led to two solutions for the

problem:

• A user defined buffer was implemented in the PID controller. As

long as the control error lies between the two buffer limits, the error

integration stops and the controller output stays constant. When

for example the buffer is 2 and the pressure set point is 260 bar, the

control error is not integrated for pressure values between 258 bar and

262 bar.

• The KV S value of the valve and its characteristic were changed from

KV S =0.63m3/h - equal percentage to KV S =0.4m3/h - linear.

After these changes the control had a satisfactory accuracy of 0.7% and its

stability was remarkably increased, as is presented on Fig.2.10(b).

A solution to the erosion problem of the one-stage valve proved very dif-

ficult. The stem of the valve was eroded to such an extent, that after 30

operating hours the KV S value was changed. The pressure control was not

possible anymore and a replacement of the stem and seat of the valve was

necessary. The problem was solved by replacing the one-stage valve with

one that performs a three stage pressure reduction. The new valve has

a stem with three heads and a triple seat that result to the same control

characteristics with the former one. At the time of writing of this thesis

this valve had already operated for approximately 130 hours without any

marks of erosion.

43


2.3.3 Fuel mixture composition control

In the past, experiments with varying fuel composition were carried out by

preparing several fuel tanks with the different compositions and manually

switching them. Operational and safety reasons make the online control of

the fuel mixture composition very important, because it allows the direct

change from a fuel mixture to water. As a first commissioning step of

this sub-system we calibrated the density as a function of the fuel mixture

composition. The next stage was tuning the density controller, followed

by the measurement of the average residence time in the fuel line. The

aim of the controller tuning was to achieve a high degree of accuracy, while

avoiding problems such as overshooting the set point and unexpected high

ethanol concentrations in the fuel mixture.

For the calibration experiments, four mixtures of ethanol and water were

prepared and fed to the fuel pump. Their density was measured at a vessel

pressure of 260 bar for three different flow rates of the pump, to account

for the compressibility effect. The results of the calibration are presented

on Fig.2.11(a).

The results of the density controller tuning are presented in Fig.2.11(b).

A relatively slow adjustment of the density was chosen in order to avoid

overshooting the set point of the controlled variable. An average of 10

minutes was necessary to reach the target density value, when the plant

operated with mf = 20 kg/h. In contrast, lower adjustment times were

observed when operating at higher fuel mass flow rates (4 minutes when

mf = 40 kg/h).

The average residence time in the fuel line was estimated during the initial

flame experiments, where a step change of the valve position from its con-

trolled value to water was used. In average, 6 minutes were necessary to

purge the approximately 5 Liter fuel line and to put out the flame, when

operating the plant with a fuel mass flow rate of 20 kg/h.

44


0 5 10 15 20 25 30960

970

980

990

1000

1010

1020

Cf [%]

ρ[ kgm

−3]

(a) Calibration of the ethanol-water mixtures density at 260 bar and roomtemperature. The points correspond to the density measurements performedwith FMI-1.

0 1 2 3 4 5 6 7 8 9 10960

970

980

990

1000

1010

1020

Time [min]

ρ[ kgm

−3]

DensitySetpoint

(b) Density control results of ethanol-water mixtures at 260 bar and mf =20 kgh−1.

Figure 2.11: Commissioning experiments for the density controller.

45

Chapter 3

Design of the pressure vessel

and its combustion chamber

The vast majority of the high pressure vessels in the field of supercritical

water oxidation was designed to safely operate hydrothermal flames, and

decompose liquid waste with high efficiency. The vessel of the hydrothermal

spallation drilling plant has the additional task to function as a drilling

device. Hence, its operational framework is considerably expanded and the

design approach must address all the new challenges (Fig.1.7) and safety

issues (see Tab.2.1).

Several details of the vessels built from Weber [30] and Wellig [37] have

been adapted and optimized in the new vessel. The way its load-bearing

walls are protected from high temperatures, its materials and the high

temperature sealing techniques have their origin in the knowledge acquired

in these projects. Beyond these parameters, the drilling procedure sets

further requirements for the geometrical parameters of the vessel.

The following paragraphs summarize the geometrical and operational pa-

rameters of the pressure vessel. The very detailed mechanical stability

calculations, performed on the basis of well established standards, have

been omitted as no significant innovation resulted from them. The chap-

ter concludes with an account of the combustion chamber design and the

positioning devices, which form one functional unit with the vessel.

47

Chapter 3 Design of the pressure vessel and its combustion chamber

3.1 Pressure vessel specifications and

geometrical characteristics

The vessel geometrical specifications comprise its diameter and length,

while its operational and technical specifications include the maximum op-

erational temperature and pressure and its materials. The in-house expe-

rience at the LTR was the foundation of the choices made with regards to

the major operational specifications, while the open literature on spallation

drilling was consulted for the choice of the geometrical characteristics.

Core of the project motivation is the demonstration of the novel drilling

technology, such that its feasibility on a conceptual level is clearly demon-

strated. The size of the boreholes that can be drilled with the facility should

prove this feasibility. Considering the typical borehole size as shown in

Fig.3.1 and evaluating data from other sources [38], the average hole diam-

eter for geothermal applications at depths greater than 2.5 km is between

19 cm and 30 cm. Taking also into account that spallation drilling can be

used for the enlargement of bore holes [39], a scale down factor of 10 was

considered realistic for the demonstration. The aimed hole diameter for

the laboratory experiments is set to approximately 2.5 cm.

The link between the targeted hole diameter and the required rock probe

diameter for the laboratory experiments was found in the existing spallation

drilling literature. During spallation drilling, the thermal confinement of

the heated part of a rock sample is necessary, in order to achieve the desired

thermal stresses in the rock mass. The rock that surrounds the heated

region must be at a much lower temperature, to avoid thermal stress relief.

Rauenzahn [19] and Wilkinson [20] expressed the findings of Preston [17] in

a rule of thumb, which estimates the maximum hole diameter as a function

of the diameter of the rock sample. It states that the directly heated

surface must be below 10% of the total surface of the sample. Hence, a

rock diameter of 9 cm should be enough for the assumption of confinement

to hold, if the desired hole diameter is 2.5 cm.

48

3.1 Pressure vessel specifications and geometrical characteristics

Drill with mudmotor 8 1/2" 2681 -3180 m

12 1/4" hole to 4511 m

1447 m

20" Casing

13 3/8" Casing

574 m

Drill 24"

Drill 17 1/2"

5031 m TVD 5101 m MD

Drill 12 1/4"

Drill 8 1/2"

Drill 12 1/4"

Reopen from 8 1/2"directional to12 1/4"

Figure 3.1: A borehole configuration in the geothermal plant in Soultz-sous-Forets, France. Adapted from [14].

Initial experiments with free hydrothermal flame-jets that were injected in

cold water showed a very strong water entrainment. This effect quenched

the flame-jets and reduced their impingement heat transfer. According to

the literature on jet entrainment [40], the entrainment mass flow rate of

water is proportional to the density and momentum difference of the two

fluids. A way to adapt the density of the water that surrounds the flame-jet

would thus offer the flexibility to tune the flame-jet temperature and its

heat transfer coefficient. In order to achieve this operational flexibility, it

should be possible to operate the vessel space at temperatures as high as

350 ◦C. Such an operating temperature makes the protection of the vessel

walls with a cooling jacket necessary, a CFD model of which resulted to an

inner vessel diameter of 14 cm. Moreover, the diameter of the vessel inner

49


volume was chosen equal to 10 cm, in order to easily accommodate the rock

probes.

Although the length of the vessel may not be significant for its mechanical

stability, it is nevertheless crucial for future drilling experiments, in partic-

ular in respect to measurements on rates of penetration. Considering that

the first drilling tool will not be optimized, the vessel length should allow

the reliable measurement of drilling velocities of approximately 1m/h. All

the same, this length should not make the construction of the vessel too

expensive for a demonstration plant. Accordingly, the length was chosen

equal to 400mm and the resulting total volume is 5.83 Liters.

400

140

100

Figure 3.2: First general drawing of the pressure vessel. The details des-ignated here are presented in Fig.3.3.

50

3.1 Pressure vessel specifications and geometrical characteristics

The specifications of the vessel can be summarized as follows:

• The normal operating pressure lies between 250 - 400 bar. For safety

reasons the highest possible internal pressure is 650 bar.

• The maximum wall temperature during normal operation is 500 ◦C.

• Optical access is provided by two small windows on the upper flange

and two in the main vessel body.

• The inner diameter and the length of the vessel are 140mm and

400mm respectively.

• Silver coated stainless steel rings with a rhomboidal cross section are

used as sealings.

• The high-performance alloy 1.4890 is used as the vessel material.

Some additional process related specifications, reflecting the integration of

the vessel in the pilot plant and best practices from the existing plants are

the following:

• The fuel stream is injected through the upper part of the vessel

head, in the axial direction and its maximum expected temperature

is 450 ◦C.

• The oxygen stream is injected through the same flange, in the radial

direction, with the same maximum expected temperature as the fuel

stream.

• The water stream for the cooling of the main volume of the vessel

(CW1) is injected in the radial direction and its temperature is equal

to 20 ◦C.

• The water for the cooling mantle (CW2) is fed radially, on its upper

side and flows downwards.

Figures 3.2 and 3.3 present a schematic drawing of the pressure vessel,

whereas a detailed account of its design steps and its technical character-

istics is given in Appendix A.

51


Combustion products outlet

CW1 inlet

CW2 outlet

Detail F

Detail E

Detail G

Fuel inlet

Oxygen inlet

Figure 3.3: Details referring to the first general drawing of the pressurevessel (Fig.3.2).

3.2 Combustion chamber design

The combustion chamber of the hydrothermal spallation drilling plant must

function as a combustor and at the same time form the flame jets that

are subsequently used as drilling tools. These two distinct functions were

studied separately in the current thesis. The present section presents its

design as a combustor, and no attention is given to its function as drilling

tool. The work performed on the latter task is discussed in section 5.4.

In the following paragraphs, the combustion chambers used previously in

our laboratory will be presented with a focus on the most successful of them.

Based on these designs, the new objectives of the combustion chamber are

52


outlined. A short review of the literature on flame stabilization follows,

which formed the basis of the subsequent detailed design.

3.2.1 Combustion chambers used previously in the

LTR

The combustion chambers used in the previous supercritical water oxida-

tion projects at the LTR had rocket propellant chambers as the starting

point [30, 29]. Self-ignition of the mixtures was used and various chambers

were compared in terms of combustion stability, wall compatibility and re-

action efficiency. The dual requirement of self-ignition and high reaction

stability led to chambers that were not optimized in terms of throughput

and power density [29].

All these chambers utilized coaxial or radial fuel injection patterns, which

resulted to distinct mixing behaviors. In the former, fuel and oxygen flow

in two concentric tubes with the fuel stream typically in the central tube

and oxygen in the annulus. Mixing takes place at the end of the inner

tube due to the shear mixing zone between the two streams. In the radial

fuel injection pattern, the fuel stream flows in the perpendicular direction

to the oxygen flow and the mixing of the reactants is enhanced from the

impingement of the fuel on the inner wall of the annulus.

The geometrical and operational parameters of the three most interesting

coaxial chambers are presented in Fig.3.4. The effect of the shear mix-

ing intensity and the average residence time in the chamber were studied

by changing the fuel injection velocity and the size of the recess volume.

Higher velocity differences between the two streams intensified the shear

mixing, while at the same time they reduced the average residence time

in the chamber. At the initial experiments, no recess (R in the BLC in

Fig.3.4) was used and consequently no ignition occurred. The addition of

15mm of recess length in chamber Nr.1 offered a longer residence time for

the reactants before being quenched by the parallel flow of water. This

53


Cooling waterOxygenFuel

2015

6 1684

D

RLk LaD

a

dk

20208 10

44

1

2 104

0.24

24

Operational characteristics [29, 30]

mO2 mf Cf vf vO2 tr

[ g/h] [ g/h] [%wt.] [m/s] [m/s] [ s]

1.14 2.13 30 0.535 3.523 0.041

1.20 2.13 30 11.54 0.412 0.107

1.10 2.15 25 138.5 3.430 0.015

Figure 3.4: Chambers with coaxial fuel injection nozzles. They are num-bered from the upper to the lowest.

resulted in ignition and very stable combustion behavior. Chamber Nr.2

demonstrated stable combustion as well, while the operational limit of this

chamber configuration was reached with chamber Nr.3. In this case, the

intensification of the mixing shear layer that resulted from the higher fuel

velocities could not compensate for the lower residence time in the chamber.

While the flames resulting from the coaxial fuel injection nozzles had the

typical characteristics of diffusion flames, the radial fuel injection chambers

resulted in a fundamentally different combustion development. The radial

injection of the fuel stream introduced a premixing step to the combus-

tion process. Furthermore, the fuel injection nozzle played the role of the

bluff body, thus contributing considerably to the pronounced combustion

stability of these chambers. Weber [30] experimented with several con-

54


2020

15

6 16148

8 10

4

64

24

42

Cooling waterOxygenFuel

Operational characteristics [29, 30]

mO2 mf Cf vf vO2 tr

[ g/h] [ g/h] [%wt.] [m/s] [m/s] [ s]

1.04 2.17 16.5 6.124 0.459 0.109

1.34 2.10 30 6.067 0.592 0.222

Figure 3.5: Chambers with radial fuel injection nozzles. Both fuel injec-tion nozzles had 12 radial holes and the fuel jets impinged on the innersurface of the oxygen annulus, thus resulting in very good mixing.

struction parameters of the radial fuel injection nozzles like the diameter

of the injection holes, their number and their distance from the tip of the

nozzle. The two most important designs of these chambers are presented in

Fig.3.5, with their respective geometrical and operational characteristics.

The vast majority of the experiments was conducted with two configura-

tions of the nozzle with 12 radial holes. Despite their remarkable flame

stability - flames continued burning even for fuel temperatures of 80 ◦C- many of these chambers demonstrated sonic phenomena and vibrations

during operation. Weber attributed these phenomena to the back flow of

fuel in the upstream direction of the oxygen annulus, which extended the

flame front in this direction. The combination of this up-flow expansion

with pressure fluctuations induced from the feed pumps was held respon-

sible for the observed screaming and chugging phenomena.

55


3.2.2 Design concept of the combustion chamber

The low reactants temperatures of the new system rule out the self-ignition

of the flames. An igniter must be inserted and withdrawn from the chamber

through a central hole in the new fuel injection nozzle. The changes in the

flow field caused by the igniter movement and the increased fuel capacity of

the chamber make combustion stability a challenging issue. Furthermore,

a flexible chamber design is needed so that an easy adjustment of its outlet

geometry will be possible. This way the characteristics of the flame jets,

subsequently used as drilling tools, will be optimized easier in the future

drilling experiments. Therefore, the dump combustors used in the past as

stabilization components [30] can not be utilized, and novel methods must

be used in order to increase the chamber stability.

The central idea of the new chamber design was to combine the positive

features of both previous designs. The new chamber must have a pre-

mixing step of the reactants, and the same stabilization mechanism as the

radial injection nozzles. At the same time, the impingement of the fuel

stream on the inner wall of the chamber should be minimized in order

to avoid combustion instabilities and reduce the cooling of the mixture

from this particular cooled wall. All these objectives can be achieved by

injecting the fuel at an angle to the axis of the oxygen flow. Through

this angle the mixing recirculation zones, typical of a bluff body flow, can

be controlled. They could be placed at the point where the igniter will

be inserted, away from the cooled walls of the chamber. Once ignition is

achieved the igniter would be removed from the flame zone thus introducing

an additional, weak central fuel jet to the system. The resulting flame will

have mixed characteristics of a diffusion and a pre-mixed flame, depending

on the geometrical parameters and the operating conditions.

The main characteristics of the new design can be summarized as follows:

• An igniter is inserted in the chamber through a central hole in the

fuel injection nozzle.

56


• The fuel is injected with an angle to the oxygen flow, so that a recir-

culation zone is produced in the region where the igniter is inserted.

This injection angle is additionally chosen in such a way that the

impingement of the fuel jet on the chamber walls can be minimized.

• The volume of the combustion chamber must be larger than in the

previous projects, in order to accommodate higher combustion ther-

mal loads.

• The chamber and fuel injection nozzles must facilitate their easy

adaptation.

d1d2d3

d4

φ1

φ2

l1 l2

d2 26

Figure 3.6: Parametric design of the combustion chamber. d1 : igniterhole diameter, φ2 : fuel injection angle, φ2 : oxygen flow broadening angle, d4 : fuel injection holes diameter, d2, d3 : oxygen annulus diameters, l1 :injection holes position, l2 : recess length.

Fig.3.6 shows the conceptual design of the combustion chamber, based

on the aforementioned considerations. Although every parameter has an

influence on the combustion reaction and must be defined, some of them

have structural limitations and cannot take all the possible values. To

define these parameters, an extensive literature study on the bluff body

flame stabilization and qualitative flow and mixing simulations has been

carried out.

57


3.2.3 Bluff body, flame stabilization analysis

The fundamental difference among the combustion chambers presented in

section 3.2.1, is that the coaxial chambers resulted in diffusion flames,

whereas the radial fuel injection chambers produced flames resembling pre-

mixed flames. Since the intended design will have characteristics from both

diffusion and pre-mixed flames, the bluff body stabilization mechanisms of

both cases were examined.

Bluff body stabilization of diffusion flames

Diffusion flames have typically a central and an annular jet, and the central

tube of the concentric arrangement plays the role of the bluff body. Fuel is

commonly injected through the inner tube and oxygen flows in the annulus.

Roquemore et al.[41] and Dano et al.[42] defined three types of flow in

the wake of the bluff body and used the central to annular jet ratiovfuel

vair

as a characteristic parameter. When the annular jet dominates the wake

flow, the central jet is constrained to recirculate toward the bluff body

surface. Two eddies are thus produced that rotate counter-clockwise and

clockwise respectively. The resulting two stagnation points of the flow are

located in the center-line of the combustion chamber. As the central jet

velocity gets higher than the annular one, it dominates the wake flow and no

stagnation point is located along the center-line. Finally, the intermediate

state, where neither jet is dominating and the eddies are tangent to each

other, produces only one stagnation point on the center-line. According

to Minx and Kremer [43], a stable chamber operation is achieved when

the stoichiometric line of the mixture lies inside the recirculation zone.

Likewise, the flame stability limit is reached once this line falls on the

border of this zone.

When confinement is introduced in the problem [44], like in our case, a

stabilization mode called wall attached flame arises for a fuel jet velocity in

the transition region. In this case the whole recirculation zone becomes non-

58


reacting and the flame is attached on the wall of the combustion chamber.

The stabilization studies generally concentrate on the process parameters

that could lead to an extinction of the flame like:

• Central fuel jet (vjet) and concentric air (vair) velocities and their

turbulence characteristics.

• The blockage factor and the shape of the bluff body. The Blockage

Factor (BF) is defined as the ratio of the cross-sectional area of the

blockage (bluff body area) to the free flow area.

• Confinement of the flame.

Bluff body stabilization of premixed flames

In premixed flames the bluff body does not contribute to the mixing of the

reactants but it works as an obstacle in the flow. A strong radial component

is added to the velocity field, and a pressure drop is produced behind the

bluff body. The pressure difference between the wake of the body and the

rest of the flow results in a typical back flow and a recirculation. Mironenko

[45] and Zukoski and Marble [46] divided the flow behind a bluff body in the

direct flow region of fresh mixture, the reverse flow zone and the turbulent

shear layer between them. According to them, combustion cannot take

place in the recirculation zone because it is always full with combustion

products. Furthermore, the temperature of the mixture in the free flow

region is too low for combustion to be possible. They concluded that the

reaction is possible only in the turbulent shear layer between the two regions

and they developed the following stabilization mechanism:

• The fresh combustible is heated by mixing with the combustion prod-

ucts in the shear layer between the free stream and the recirculation

zone.

• If sufficient gas can be ignited until the end of the recirculation zone,

a propagating flame is established and stable combustion is reached.

59


Fetting et al. [47] confirmed these mechanisms, by showing that the flames

burned within the shear layer after the tip of the bluff body and that higher

free flow velocities shifted the ignition point downstream in this layer. In

conjunction with these considerations, Loeblich (see [48]) showed that the

spreading of the flame is determined by the heat flux from the flame to the

fresh gas at the boundary of the recirculation zone.

Winterfeld [48] investigated the shape and the dimensions of the recircula-

tion zone for three different bluff body geometries, arriving at the following

conclusions concerning its geometrical characteristics:

• Wider and longer recirculation zones result when the bluff body in-

duces a higher radial velocity component.

• The free stream velocity has a rather small influence on the width of

the recirculation zone.

• The difference of the width of the recirculation zone with and without

combustion is minimal. By contrast, its length is considerably greater

when a flame is present.

• The length of the flame reaches its maximum for mixtures near their

stoichiometric conditions, irrespective of the geometry of the bluff

body.

Simple Stabilization models and criteria

Several stabilization models have been developed over time, considering the

recirculation zone and the processes taking place in it and on its boundaries.

The CSTR Model ([45],[49]) is based on the thermal ignition models de-

veloped from Frank Kamenetskii [50]. It treats the recirculation zone of

a bluff body as a CSTR reactor and the ignition conditions are computed

with the corresponding thermal ignition theory. This flame stability crite-

rion states that for a stable flame, the residence time of the mixture in the

shear layer of the recirculation zone has to be higher than its ignition time.

The ignition time is approximated with the thermal ignition theory [50].

60


A similar model comparing characteristic time scales of the flow and the

reaction is the Peclet number stabilization criterion. This criterion essen-

tially assumes that stable operation occurs, when the flow velocity and the

laminar flame propagation velocity are matched [51, 52].

3.2.4 Combustion system construction

The most challenging aspect of the construction was the realization of the

forced ignition of hydrothermal flames, because no experience existed in

this field. Moreover, increased flexibility was crucial because both the fuel

injection nozzle and the chamber outlet had to be easily adaptable.

Fig.3.7 presents the most simple design of the system, which was used

during the experiments presented in chapter 4. It consists of three compo-

nents, the fuel injection nozzle, an adapter and the combustion chamber

assembly. Fuel and oxygen enter the vessel from the respective ports on

its upper flange. The fuel injection nozzle, which comes in contact only

with the small flange of the vessel, defines the fuel flow conditions in the

chamber. Similarly, the adapter forms the connection between upper flange

and chamber assembly. Its inner diameter and the outer diameter of the

fuel injection nozzle define the annulus, through which oxygen enters the

combustion chamber. The latter is formed from two, welded concentric

tubes that form the annulus, in which the cooling water flows. Besides, the

shape of this passage defines the injection conditions of the cooling water

in the main vessel volume.

Several geometrical limitations are imposed on the combustion system de-

sign, mainly due to the vessel geometry, which are summarized as follows:

• The outer diameter of the chamber cannot exceed 26mm, which is the

diameter of the hole in the large upper flange of the vessel. By choos-

ing a minimum wall thickness of 1mm, the chamber inner diameter

cannot be higher than 20mm.

• The fuel injection nozzle is inserted in the system through a hole in

61


CW1inlet

Combustion chamber

Adapter

Fuel injection nozzle

Oxygen inlet

Fuel inlet

Figure 3.7: Design of the combustion chamber in the high pressure vessel.

the upper small flange of the vessel. Given the diameter of this hole

is 10mm, the maximum outer diameter of the fuel injection nozzle is

limited to 9.5mm.

• The central hole of the fuel injection nozzle was matched to the ge-

ometry of commercial igniters and thus it was chosen equal to 4mm.

A series of 2D simulations was performed to better understand the influence

of the chamber geometry on the temperature and concentration distribu-

tions in it. For these simulations only the cold flow (no combustion) was

considered and temperature dependent fluid properties were used. The

simulated operational cases are presented in Tab.3.1.

The simulation defined the geometrical parameters of the combustion sys-

tem that were relevant to its mixing performance. Thus, the distance of

62


Table 3.1: Cold (no combustion), 2-D simulation cases for the combustionchamber design.

Case Nr. 1 2 3 4 5 6

P [ kW] 20 20 50 50 80 80mf [ kg/h] 26.71 6.68 66.79 16.7 53.43 26.72Cf [wt.%] 10 40 10 40 20 40mO2 [ kg/h] 10 10 14 14 22.3 22.3

the injection holes from the tip of the fuel injection nozzle, their diameter

and their angle to its axis were defined from the simulations. Furthermore,

the inner diameter of the adapter was considered because it influences the

oxygen injection velocity.

The finalized design of the fuel injection nozzle can be seen in Fig.3.8. The

fuel stream is injected at an angle of 30◦ to the axis of the chamber. The

distance of the injection holes from the nozzle tip is 6.5mm, while the total

length of the nozzle is 115mm. The entire nozzle was built in two pieces to

make the construction of the difficult injection geometry easier. Laser rapid

prototyping was used for the nozzle tip made of EOS CobaltChrome MP1.

The tip was built layer after layer from a powder of the material, which

was rapidly sintered by a lased beam. In comparison, the main body of the

nozzle was conventionally machined and made of Inconel 625. Finally, the

two pieces were laser welded to produce the final geometry.

Detail E Section View

30

1156.5 E12 X 1mm

47.5

9.5

Figure 3.8: Drawing of the fuel nozzle.

Fig.3.9 shows the dimensions of the combustion chamber. The inner tube,

has diameters of 15.8mm and 21.3mm, and it is made of Inconel 625;

63


heated fuel and oxygen are injected, mixed and ignited in this volume. The

outer tube has diameters 22.5mm and 26mm and it is made of stainless

steel.

262615.8

15.8

100100

22.5

22.5

21.3

21.3

Figure 3.9: Drawing of the combustion chamber.

3.3 High pressure auxiliaries

Simple, spatially defined temperature measurements, supplemented with

chemical analysis of the combustion products have sufficed during most of

the past SCWO projects. However, positioning of measuring probes and

rock samples in the vessel volume, while under pressure, is necessary for

the experiments of the present thesis. Furthermore, the igniter has to be

inserted and withdrawn from the combustion chamber, while the whole

system is operating at high pressure and high temperature.

The positioning devices’ concept is similar to that developed from Prikop-

sky [53]. The positioning devices consist of two parts, arranged like a

double telescopic tube. A fixed part is threaded directly on the pressure

vessel, and a moving part is fixed to a servo motor that is responsible for

accurate positioning. The positioning tube of the system is mechanically

connected to this moving part. This tube is guided through the fixed part

of the device and is sealed on it with a glide ring. The sealing principle is

similar to that of the stems of high pressure needle valves.

64

3.3 High pressure auxiliaries

Trav

el d

ista

nce

= 10

0 m

m

Upper positionLower position

Sealing of cables

Movable part

Sealing

Connection to the reactor

(a) Upper positioning device.

Pressure vessellowe flange

Trav

el d

ista

nce

= 35

0 m

m

Connection for heat fluxsensors and rock probes

Connection - servo motor

Upper position Lower position

Cooling waterconnection

(b) Lower positioning device.

Figure 3.10: The devices used for the positioning of probes in the vessel,while under pressure. The accuracy of the positioning is typically 0.1mm.

The smaller positioning device is installed on the upper part of the pressure

vessel and is presented in Fig.3.10(a). Its positioning tube has an inner

diameter of 3.2mm, a stroke of 100mm and it is sealed on its outer diameter

with a composite, gliding ring (glass-teflon). Cooling fins are machined in

its periphery to keep the temperature of the sealing low, thus allowing a

maximum operating temperature of 420 ◦C.

65


On the lower side of the pressure vessel a second, larger device is installed,

with a travel distance of 350mm and a positioning tube with diameters of

8mm and 14mm respectively. The same sealing technique is used for this

positioning tube too, but in this case only teflon sealing rings are used and

no cooling fins are necessary. The operating force of the device is 4.8 kN

and the whole positioning system has been dimensioned for 5.5 kN.

66

Chapter 4

The ignition project

4.1 The scientific problem - motivation

The field implementation of hydrothermal flames for spallation drilling

requires a methodology to reliably ignite flames using the available un-

derground equipment. Self-ignition cannot be achieved in the conditions

of a typical bore hole, where the reactants temperature cannot exceed

100-250 ◦C. At the same time, underground heating of the reactants is

not an option, due to the very high electrical power necessary for it. Thus,

forced ignition remains the only way to operate a hydrothermal flame in

2.5 km and its investigation is indispensable for the implementation of the

technology.

An equally strong motivation for the performed experiments was the sci-

entific aspect of the forced ignition mechanisms in hydrothermal flames.

Broad research has been conducted on these flames in the 90s and 2000s,

when the self-ignition of various combustible mixtures was investigated in

batch reactors ([54],[55],[56]). Besides, diffusion flames were continuously

operated, in a similar way to conventional combustors ([30],[29]). Despite

the fact that many characteristics of these flames have been clarified [53],

research on their forced ignition was only limited and non systematic [28].

Finally, the heat transfer conditions in the trans-critical thermodynamic

67

Chapter 4 The ignition project

region of fluids offer a fundamental motivation on the analysis of forced ig-

nition. Research by Laurendeau [57] and Mullen et.al [58] presented critical

approaches on forced ignition in environments with strong convective heat

transfer characteristics. The expected complex heat transfer in our com-

bustion chamber offered the chance to extend this data to regions where

the convective heat transfer coefficient is strongly temperature dependent.

Generally, spark and hot-surface ignition are the mechanisms typically used

in the relevant applications [51]. The former has been used for highly

volatile combustible mixtures and premixed combustion applications, while

the latter has been primarily implemented for the ignition of diffusion

flames. SI internal combustion engines offer the most typical application

of spark ignition, while most of the industrial furnaces still make use of

hot-surface ceramic igniters. The dielectric constant and heat capacity of

water in its trans-critical region demonstrate high values in comparison to

conventional gaseous combustibles. As a consequence, the implementation

of spark ignition in the present project would require very high voltages

to be fed at a high pressure environment. Hence, hot surface ignition was

chosen as our ignition methodology.

4.1.1 Ignition measurement goals

The two parameters that define the ignition conditions of a combustible

mixture, are the temperature and heat flux of the heated surface at the

point of ignition. The current thesis focused on the physical phenomena

during the forced ignition of hydrothermal flames. Thus, case relevant pa-

rameters like the geometrical characteristics of the combustion chamber,

its cooling methodology and the oxygen to fuel mass ratio (λ) were kept

constant throughout the investigation. Since the pressure of the system

would only change the quantitative and not the qualitative aspects of the

measurements, its influence was not considered either. Consequently, the

experiments focused on the influence of the combustible mixture tempera-

ture, its composition and the combustion chamber load. The combustion

68

4.1 The scientific problem - motivation

chamber load is defined as the consumed fuel power divided by the chamber

volume and it is a function of the fuel flow rate and its ethanol concentra-

tion.

As of the time of writing, no ignition study has yet considered the direct

link between the ignition conditions and the heat transfer coefficient of

the flow. Most of the existing studies take this parameter indirectly into

account, by changing the velocity and the turbulence levels of the flow.

By contrast, the convective heat transfer of trans-critical fluids is more

complex and the study of its direct influence on the ignition parameters

is relevant. The ignition project was therefore divided in two consecutive

measurement series. In a first stage the average convective heat transfer

coefficient was measured in the combustion chamber with a ternary mixture

of water-ethanol-nitrogen as a model of the combustible mixture. The

second stage comprised the ignition measurements at similar points to the

former experiments.

4.1.2 The measurement concepts

During the heat transfer and the ignition experiments, a heated surface

(the igniter) is inserted in the flow of a ternary mixture.

The definition of the convective heat transfer coefficient in eq.4.1 is the

starting point of the respective measurements.

q = h · (Tsurf − Tbulk) (4.1)

In order to perform a measurement of h, the rest of the parameters in eq.4.1

must be measured. The average surface temperature of the igniter was

measured indirectly, by measuring its electrical resistance. The correlation

of its resistance with its temperature was then calibrated in a separate

experiment. Likewise, the heat flux was calculated from the voltage and

69


current values measured on the igniter, and its surface area (see eq.4.2).

q =V · IA

(4.2)

The bulk temperature was measured during both the heat transfer and the

ignition experiments by a K-type thermocouple with a diameter of 3mm.

This thermocouple was inserted in the chamber from below with the lower

positioning device (see Fig.3.10(b)).

Finally, the ignition experiments aimed also at the simultaneous measure-

ment of the heat flux and the surface temperature of the igniter at the

point of ignition. As a result, the same conceptual methods with the heat

transfer coefficient measurements were used.

4.2 Hot surface ignition state of the art

Hot-surface ignition has been the topic of numerous investigations. Case

studies on various combustible mixtures, stagnant or flowing over surfaces,

form a very useful background for the adaptation of the technology to

hydrothermal flames. Theoretical and experimental studies, relevant to

the aims of the thesis, are summarized in the following two sections. This

summary is qualitative because quantitative data of these studies have no

practical importance for the task at hand.

4.2.1 Thermal ignition theory

Semenov [59] and Frank-Kamenetskii [50] developed two simplified models

for the thermal explosion limits of a combustible mixture in a vessel. Both

models attempt to define a criterion for the onset of combustion, based on

the governing equations of the system and its boundary conditions.

The Semenov theory deals with a combustible that is confined in a vessel

70


and has a uniform temperature distribution. Heat is produced from the

slowly advancing combustion reaction and is dissipated to the surroundings

through convection. The model states that an explosion takes place once

the rate of heat production becomes higher than that of heat dissipation.

The resulting Semenov criterion defines a critical value of the temperature

difference between the combustible and its surroundings, above which an

explosion occurs.

Frank-Kamenetskii studied a similar problem of a combustible in a vessel

that had a non-uniform temperature distribution, while the wall of the

vessel had a fixed temperature. Only heat conduction was considered in

the mass of the combustible and the heat transfer to the surroundings

was assumed to occur at an infinite rate (infinite Biot number). The model

predicts that an explosion occurs when the boundary values problem of heat

conduction in the combustible results in no real temperature distribution.

Based on this assumption, a value of a parameter δc is defined as the

explosion limit and each case study with a higher δ value will eventually

lead to an explosion.

The simplifications that the models used in order to facilitate the analyti-

cal solution of the governing equations, made them very popular until the

late seventies. However, the same assumptions and simplifications led to

discrepancies between their predictions and the observed physical phenom-

ena. Gustafson et al. ([60],[61]) and Bazley et al. [62] additionally showed

that the critical phenomena predicted from the Frank-Kamenetskii model

did not correspond to the solutions they obtained from numerically solving

the boundary values problem.

This fact led many researchers to search for alternative ignition criteria, by

solving the problem numerically. Laurendereau [57], and Merzhanov et al.

[63] were some of the numerous researchers using the so-called Damkohler

or Van’t Hoff criterion for ignition. They calculated the temperature and

concentration distributions in the boundary layer of the flow and argued

that ignition occurs when the heated surface and the fluid have the same

temperature. Unfortunately, this criterion fails to predict realistic ignition

71


conditions when the heated surface operates with a constant heat flux or

when the wall temperature approaches the adiabatic flame temperature.

In the first case, a temperature difference is always necessary to transfer

heat form the surface to the fluid. In the other case the resulting flame

temperature would be lower than that of the surface.

Chen and Faeth [64, 65] studied the ignition of a combustible that was

heated from a vertical plate. They calculated the temperature distribution

in the laminar boundary layer, but instead of stopping their simulations on

the surface, they expanded them in the wake of the plate. As an ignition

criterion, they used the development of the concentration and the temper-

ature fields in the wake flow. In the ignition cases, the temperature of the

combustible increased in the wake until it reached a plateau, equal to the

adiabatic wall temperature. A similar approach was followed by Siccama

et al. [66], and Bolk et al. [67] for the ignition of ethene-air mixtures with a

hot wire. Their results support the approach of simulating the whole wake

of the flow to define the ignition conditions of a system.

4.2.2 Experimental studies on hot surface ignition

Most of the experimental studies on hot wire ignition are carried out either

in stagnant gases or in closed vessels similar to the classical combustion

bombs. Mullen et al. [58] investigated systematically the ignition of com-

bustibles flowing over heated wires. They used a variety of fuels and ana-

lyzed the influence of several parameters on the temperature of the hot wire

at the point of ignition. While Mullen et al. used an indirect way to mea-

sure the temperature of the heated wire, Buckel et al. [68], who worked

on the flammability limits of hydrogen - oxygen mixtures, measured its

temperature directly through its electrical resistance. Besides, Bolk and

Westertep [67] investigated the influence of convection on the flammability

limits of ethene-air-nitrogen mixtures. They also used a heated wire, but

its temperature was measured with an infrared thermometer.

In addition to the hot wire experiments, studies on the ignition with heated

72


plates have been also conducted. Toong [69] concentrated on the laminar

flow of a combustible over a heated plate. He argued that ignition takes

place in the boundary layer of the flow and the free stream participates in

the reaction only when the flame starts to propagate. On the other hand,

Ono et al. [70] studied the ignition of combustibles flowing over a vertical

heated surface, with their flow being a result only of natural convection.

Like Toong, they investigated only the cases with laminar flow, so that

they could compute the conditions in its boundary layer.

The majority of the experimental studies concentrated on a limited number

of parameters influencing the ignition process, and the general trends may

be summarized as follows:

• An increase in the flow velocity leads to an increase of both the surface

ignition temperature and the respective heat flux. The same trend

holds for the turbulence intensity in the flow of the combustible.

• Large heated surfaces led in most of the cases to lower ignition tem-

peratures.

• Mixtures with higher bulk temperatures always have lower surface

ignition temperatures.

• The lowest ignition temperatures are observed around the stoichio-

metric point of the mixtures.

• Increasing the pressure of the mixtures led in most of the cases to

lower surface ignition temperatures.

All these qualitative remarks come from studies performed under the as-

sumption of constant fluid properties during the experiments. By contrast,

the very strong properties variation of trans-critical fluids dominates all the

heat and mass transfer phenomena. Hence, a brief summary of the heat

transfer literature in trans-critical fluids is necessary for the interpretation

of the experimental data of the thesis.

73


4.3 Convective heat transfer in supercritical

fluids - literature review

Most of the studies on the convective heat transfer of supercritical fluids

consider simple tubular and annular geometries and pure fluids, mainly wa-

ter or carbon dioxide. Although the general, physical principles known from

conventional heat transfer problems apply to supercritical fluids, the very

strong variation of the fluid properties around their critical point introduces

two additional heat transfer modes. Depending on the geometrical setup

and the flow arrangement, a region of enhanced and one of deteriorated

convective heat transfer are observed. In the following paragraphs, a short

presentation of the fluid properties of water and its mixtures with ethanol,

oxygen and nitrogen will be presented. A phenomenological description

and an analysis of the convective heat transfer phenomena in supercritical

fluids will conclude this section.

4.3.1 Physical properties of water and its mixtures

with ethanol, oxygen and nitrogen in the

trans-critical region

Although water behaves as a typical liquid and gas, this is not the case

when its pressure and temperature approach their critical values (pcrit =

220,64 bar, Tcrit = 374 ◦C). Changes in its molecular structure cause sig-

nificant variations of its thermal and transport properties in a very narrow

temperature region. Its properties as functions of temperature for a pres-

sure of 250 bar are shown in Fig.4.1(a), where they are normalized with

their value at 0 ◦C.

A rapid change of all the properties from liquid-like to gas-like values is

observed in a temperature interval of approximately 40 ◦C. These changes

during the transition from sub-critical liquid to supercritical gas are similar

to the ones between liquid water and water steam. The fundamental dif-

74

4.3 Convective heat transfer in supercritical fluids - literature review

0 100 200 300 400 500 600 700 8000

0.2

0.4

0.6

0.8

1

1.2

T [◦C]

ε ε 0,

ρ ρ0,

η η0,κ κ0

ερηκ

(a) Water properties at 250 bar, normalized with their values at 0 ◦C [71].Here ε is the dielectric constant, η is the viscosity, ρ the density and κ thethermal conductivity.

210220

230240

250260

270

350360

370

390380

400410

420

200400600800

100012001400

p[bar]T[°C]

cp [kJ kg K ]1600-1 -1

(b) Specific isobaric heat capacity of water, adapted from [72].

Figure 4.1: Water thermodynamic and transport properties [71, 72].

75


ference between a normal evaporation and the aforementioned transition is

that the fluid is experiencing a phase transition with no coexistence of these

two phases. This transition requires a continuous change of the properties,

which must adapt abruptly from the values of the liquid to these of the

supercritical fluid. The behavior of the specific isobaric heat capacity is

the most striking. It demonstrates a localized maximum at a temperature

that is defined as the pseudo-critical temperature of the fluid. As it can be

seen from Fig.4.1(b), both the peak value of the cp and the pseudo-critical

temperature are functions of pressure.

As already mentioned, the heat transfer and ignition measurements of this

thesis are carried out with ternary mixtures of water-ethanol-nitrogen and

water-ethanol-oxygen. Although the properties of all the components in-

dividually are known in their trans-critical regions, this is not the case

for the mixtures themselves. Nevertheless, data for the binary mixtures

water-oxygen [73], water-nitrogen [74], water-ethanol [75, 76] and ethanol-

nitrogen [77] can be found in literature. This data may provide at least

qualitative indications on the expected critical and pseudo-critical points

of each ternary mixture.

Japas and Franck studied the properties of trans-critical mixtures of water

with oxygen [73] and nitrogen [74]. Both mixtures demonstrate a type

II critical curve, which starts from the critical point of water and with

increasing pressure first moves to lower temperatures up to the value 640K

for oxygen and 639K for nitrogen. A further increase in the pressure of

the system leads to an increase of its critical temperature as shown in

Fig.4.2. In other words, when working at 260 bar an addition of oxygen or

nitrogen results in a mixture with a lower critical temperature than that of

water. Moreover, it is clear from Fig.4.2 that both mixtures have almost

identical behaviors, and one gas may replace the other without a significant

impact on the properties of the resulting mixture. Rogak [78, 79] performed

extensive measurements of the specific isobaric heat capacity of the water-

oxygen mixture at a pressure region directly relevant to this thesis. He

showed that an addition of 8% vol. of oxygen lowers the pseudo-critical

76


p [M

Pa]

50

100

150

200

250

T [°C]227 277 327 377

(a) Water-Nitrogen.

p [M

Pa]

50

100

150

200

250

T [°C]227 277 327 377

(b) Water-Oxygen.

Figure 4.2: Binary mixtures water-oxygen and water-nitrogen [73, 74].Isopleths (mole % water) in the pressure-temperature diagram, where thedashed line is the projection of the critical curve on the respective plane.

point of the mixture by approximately 20 ◦C at a pressure of 250 bar.

The most recent study on the properties of trans-critical water-ethanol

mixtures was published from Bazaev et al. [76]. Their findings on the

critical point of mixtures with molar ethanol concentrations of 20%, 50%

and 80% can be seen in Fig.4.3. The critical line of the mixtures connects

the critical points of the two constituents, without any minima or maxima.

The addition of ethanol shifts the critical point of the mixture to lower

pressure and temperature values.

77


0 0.2 0.4 0.6 0.8 15

9

13

17

21

p C[M

Pa]

EtOH mole fraction

(a) Critical pressure.

227

EtOH mole fraction 0 0.2 0.4 0.6 0.8 1

267

307

347

T C [°

C]

(b) Critical temperature.

Figure 4.3: Critical pressure and temperature of the binary mixture ofwater and ethanol and various ethanol mole concentrations [76].

4.3.2 Convective heat transfer phenomena in

supercritical fluids

Typically, the convective heat transfer is measured under the assumption

that the dependence of the fluid properties on the temperature and pres-

sure can be neglected. A side effect of this assumption is that the weak

dependence of the heat transfer coefficient on the fluid and surface temper-

atures is also neglected. Apparently, the constant properties assumption

does not hold when the fluid is crossing its critical (or pseudo-critical) point

in the course of the experiment. The drastic changes in the fluid properties

introduce two additional heat transfer behaviors. Depending on the flow

and temperature fields, a heat transfer enhancement and a deterioration

mode appear in addition to normal convection.

Tanaka et al. [80] and Yamagata et al. [81] studied respectively the heat

transfer of carbon dioxide and water crossing their critical point, while flow-

ing in vertical, heated, circular tubes. Both studies used tubes that were

directly, electrically heated and their outer temperature was measured with

thermocouples in many points along their length. Their inner wall temper-

ature resulted from the heat conduction equation for cylindrical coordinates

by assuming uniform heat production in the tube.

78


Yamagata et al. observed an increase of the convective heat transfer coef-

ficient in certain areas along the tube, for relatively high mass fluxes and

low heat fluxes. These areas moved along the tube depending on the inlet

temperature of the fluid and the heat and mass flux. They also reported

that an increase of pressure reduced the maximum value of the heat trans-

fer coefficient, which also occurred at higher bulk temperatures. Likewise,

higher heat fluxes caused a decrease of this maximum value and its oc-

currence at lower fluid temperatures. A further increase of the heat flux

brought about an even stronger decline of the heat transfer coefficient and

led to the phenomenon known as heat transfer deterioration. During this

phenomenon, high wall temperatures were observed at fluid temperatures

similar to the ones where heat transfer enhancement took place. Further-

more, the fluids behaved in a different way depending on whether the flow

was upward or downward. As a matter of fact, the upward flow showed a

higher tendency to develop deteriorated heat transfer conditions.

Tanaka et al. observed similar phenomena in the flow of carbon dioxide, but

went a step further and described the phenomena taking place in the vicin-

ity of the heated wall. They concentrated on the regions where enhanced

heat transfer was observed and argued that it may be explained with the

typical turbulent convection concept. They showed that when the fluid

that occupies the buffer layer of the turbulent boundary layer is close to

its pseudo-critical point, heat transfer enhancement occurs. The increased

cp of this fluid forces heat to be transferred with a very small temperature

gradient between the wall and its bulk and results in a heat transfer co-

efficient increase. As the heat flux is further increased, the volume of the

boundary layer occupied by the high cp fluid is reduced, because the heat

is sufficient to produce a considerable temperature difference. Thus, the

temperature profile approaches the typical profile for convective heat trans-

fer and the heat transfer coefficient falls to normal values. This concept

explains the behavior of the peak values of the heat transfer coefficient, and

their correlation with the pseudo-critical point of the fluids.

Licht et al.[82] performed measurements of the convective heat transfer

79


Increasing Density Layer

AcceleratingDensity Layer

High specific heat

Localized high specific heat

Het

ated

Wal

lTemperature

Flow Direction

a) Normal heat transfer(No property variation)

b) Heat transfer enhancement(Large G, low q)

c) Impairment of enhancedheat transfer(Large G, increasing q)

d) Deteriorating heat transfer(Low G)

e) Recovering heat transfer(Low G)

x

T

Figure 4.4: Conceptual model of the heat transfer enhancement and de-terioration mechanisms in supercritical fluids (from [82]).

of water flowing in annuli with circular and square cross sections. They

extended their initial experiments with velocity field measurements, using

laser Doppler velocimetry in a subsequent work [83]. Their heat transfer

enhancement model was identical to that of Tanaka et al. (see Fig.4.4).

They additionally stated that during heat transfer enhancement no signif-

icant changes take place in the flow field of the fluid near the wall. Conse-

quently, the phenomena of heat transfer enhancement and its impairment

for higher heat fluxes can be considered as purely thermal and take place

in the turbulent boundary layer of the flow. On the other hand, they dis-

covered significant changes in the flow field near the wall of their square

channel during heat transfer deterioration. They divided the occurrences

in three steps and highlighted the importance of buoyancy in the disruption

of the flow field:

• In the first step, the fluid near the surface is heated rapidly and

changes to its supercritical state. The lower density of this fluid

80

4.4 The igniter modules

layer accelerates the flow in the near-wall region. As a result, the

radial velocity profile across the annulus becomes flat and the velocity

gradients are concentrated close to the wall. This change reduces the

momentum and heat exchange between the two layers of the flow.

Further rise of the temperature of the film leads to a gradual decrease

of its density and finally to a self sustaining phenomenon.

• In the second stage, the considerable velocity difference between the

near-wall film and the bulk leads to a gradual increase of their heat

and momentum exchange. Thus, the temperature and density gradi-

ents between the two regions are diminished and the temperature of

the near-wall layer is reduced.

• In the third part, the bulk crosses its critical temperature and the flow

in the annulus becomes a classical turbulent gas flow. This change

leads to an increase and a stabilization of the heat transfer coefficient.

Summarizing, the heat transfer deterioration is caused by the combined

effect of buoyancy and the phase change of the fluid. Moreover, it has

been interpreted from many researchers in the past as a pseudo-film boiling

phenomenon [84].


The ignition modules, their design and their behavior during the measure-

ments, are the focus of this section. Although many igniters can be found

in the respective market, our specifications posed difficulties on their direct

implementation in the plant. The igniters must be able to continuously

operate in a temperature up to 1000 ◦C, while for short intervals they must

withstand temperatures reaching 1500 ◦C. In parallel, supercritical mix-

tures are known to be chemically very aggressive and the igniters must be

made of materials able to withstand these fluids. A further challenge was

the choice of a material that would allow the measurement of its mean

surface temperature through a measurement of its resistance.

81


In most technical applications, the electrical connection of the igniters is a

minor issue. In our application the connecting wires are fed through high

pressure (260 bar) and high temperature (420 ◦C). Besides, the very high

temperatures of the hydrothermal flames make the removal of the igniters

from the combustion chamber after ignition necessary. As a result the

ignition module should provide the necessary mechanical stability for this

task.

In the course of the project ceramic and metal igniters have been imple-

mented in the plant. The former operated with AC 230 V, while the latter

used DC voltage with a maximum value of 65V. The following paragraphs

describe the design of the two ignition modules, the implemented measure-

ment principles for their resistance and their calibration.

4.4.1 Ceramic ignition module

The ceramic igniter is a cylindrical body with a 4mm diameter consisting

primarily of hot pressed silicon nitride (HPSN). Its electrical resistance is

integrated in its body by a sintering procedure and it consists of approx-

imately 80 % vol. silicon nitride while the rest is additives, MoSi2 and

TaN . The outer surface of the igniter is electrically insulating, its total

length is 90 mm, while its heated zone is 40mm long.

The temperature dependence of its resistance was calibrated up to 520 ◦Cin a high temperature oven. A typical calibration line is presented on

Fig.4.5, together with its confidence intervals, computed according to the

orthogonal regression method [85]. Higher temperatures were not possible

during calibration because the material used for the electrical contacts of

the igniter melts above 550 ◦C. However, reference calibration data from

the construction company suggested that the linear dependence is valid

up to 1000 ◦C [86]. The temperature during the calibration was measured

with two K-type thermocouples (±2.5◦C), positioned near the igniter and

its resistance was measured with a simple digital multi-meter (±0.8%).

82


300 320 340 360 380 400 420 440 460 480 50060

65

70

75

80

T[◦C]

Rign[O

hm]

MeasurementsRegressionConfidence interval

Figure 4.5: Ceramic igniter calibration curve, R = b0 + b1 ·Tw.

The resistance - temperature correlation that results from the calibration

experiments gives the values of the temperature in the core of the igniter.

This value is not the same with the surface temperature of the igniter, when

it operates in a strongly convective environment. As a result, the thermal

resistance of the electrically insulating layer of HPSN must be also taken

into account. In the experiments of this thesis, the thermal conductivity

of this layer was estimated from the values reported by Watari et al. [87]

at the same temperature as the core. Accordingly, the value of the surface

temperature was corrected by assuming simple conduction in cylindrical

coordinates.

Igniter oxidation and measurements correction

After the first commissioning experiments of the ceramic igniter, silicon ni-

tride oxidation was identified. The oxidation phenomenon, which is studied

in various publications ([88],[89],[90]), is modeled from the chemical reac-

tion below. The same publications argued that the respective reaction rate

83


follows the Arrhenius kinetics law.

Si3N4 + 6H2O � 3SiO2 + 4NH3

The silicon oxide layer that was a product of the oxidation introduced an

additional thermal resistance layer between the surface and the fluid and

affected the heat transfer coefficient measurements.

0 1 2 3 4 5 64800

4900

5000

5100

5200

5300

5400

5500

h[W

m−2K

−1]

h

0 1 2 3 4 5 60

5

10

15

20

25

30

35

Lay

erThickness[μm]

Time [hr]

Figure 4.6: Reference measurements for the oxidation of the ce-ramic igniter. Conditions: mf=40 kg/h, mN2=130 kg/h, Tbulk=326 ◦C,q=0.5MW/m2.

The growth of the layer was studied at THX2 = 385 ◦C, by measuring the

heat transfer coefficient of a fresh igniter and by repeating the measurement

every fifteen minutes for six hours. The changes of the heat transfer coeffi-

cient were attributed to the additional thermal resistance of the SiO2 layer,

the thickness which was calculated from one - dimensional heat transfer in

cylindrical coordinates and is presented in Fig.4.6. Finally, a function was

fitted to the measured data in order to calculate the growth of the layer

with time and correct the heat transfer measurements of section 4.5.

84


4.4.2 Coil igniter

A coiled wire was chosen as an alternative to the ceramic igniter, because

several metals are less prone to surface oxidation from supercritical water

than ceramics. The coil consists of NiCr 60/15, has a suitable temperature

dependence of its resistance and can withstand high temperatures [68]. The

outer diameter of the coil is 2.5mm, while the wire thickness is 0.4mm,

leaving an inner diameter of approximately 1.7mm. Its total length is

30mm and its electrical resistance is 7.5 Ω at 20 ◦C.

Figure 4.7: NiCr 60/15 coil igniter before its connection with its leadwires.

The resistance-temperature function of the igniter was calibrated in a high

temperature oven between 500 ◦C and 1000 ◦C, and the calibration function

can bee seen in Fig.4.8. The calibration temperatures were chosen in the

range expected for the ignition temperatures to simplify the fitting of a

curve through the data. Consequently, a simple second order polynomial

function was applied to describe the change of the material resistance with

temperature.

As can bee seen from Fig.4.8, the small change of the igniter resistance with

increasing temperature - 0.7mΩ/◦C - makes the temperature dependence

of the lead-wires resistance important. As a result of space limitations in

85


high pressure vessel no direct four point (with a kelvin bridge) resistance

measurement was possible. Instead, an experimental model of the lead

wires was used that consisted only of the lead wires and it was simultane-

ously calibrated with the igniter to correct the calibration. All else equal,

the resistance of the wires was not the same during the ignition experiments

and the calibration. To solve this problem, the wires model was inserted

in the high pressure vessel and its resistance was measured at exactly the

same operating points as in the ignition experiments. This methodology

additionally considered the thermoelectric voltages produced from the dif-

ferent materials of the lead wires, by using the same contact polarity during

calibration and operation.

500 600 700 800 900 10008

8.1

8.2

8.3

8.4

8.5

8.6

T [◦C]

Rign[O

hm]

MeasurementsRegressionConfidence interval

Figure 4.8: Coil igniter calibration curve, R = a1 + a2 ·T + a3 ·T 2.

The temperature in the calibration was measured with a K-type thermo-

couple (±2.5 ◦C) and the resistance measurements were performed with a

high accuracy digital multi-meter (±0.034%). The error of the temperature

calculation from the resistance measurement varied between 3% and 5%.

86


4.4.3 Electrical connection of the igniters

In order to address the high temperature in which the lead wires must

operate, their length is divided in a high (200-420 ◦C) and a low temper-

ature region (40-200 ◦C). The former consists of 400mm of alumel wires

(�0.8mm), while the latter is made of 350mm of kapton-insulated copper

wires (AWG 20). The alumel wires are insulated from each other through

an alumina capillary tube with two inner holes (�0.82mm). The lower

end of the copper wires is welded to the upper end of the alumel ones, the

lower ends of which are then welded on the igniter. In the case of the coil

igniter, the lower lead wire is led through the coil in a alumina capillary

(see also Fig.4.7). The ceramic igniter is constructed with prefabricated

connections, on which the lead wires are directly welded. The whole lead

cable constellation (Fig.4.9) has an electrical resistance of 0.543Ω at 20 ◦C.

copper wires :350 mmalumel wires :400 mm

Igniter :30 mm

Figure 4.9: Coil igniter constellation with its lead cables. The same cableswere used for both igniter types.

In order to insert and withdraw the igniter from the combustion chamber

we used the upper positioning device (Fig.3.10(a)). The lead wires are fed

through the moving tube of the device and the copper wires are then sealed

at its end with a high pressure gland. This gland fixes the igniter module

on the moving part of the positioning device. By pulling or pushing the

upper end of the copper cables we are able to move the whole module up

and down. This movement requires the mechanical force transfer from the

copper wires to the rest of the assembly, which puts pressure on the junction

of the two wire types. To avoid bending, the length of the copper wires is

chosen so that this junction is always inside the 3.2mm positioning tube.

Given the outer diameter of the alumina capillary is 2.8mm, no space is

87


available for the bending of the copper wires on this junction. As a result,

the whole constellation reacts as a rigid body when force is exerted on it.

The resulting voltage feeding system can conduct a maximum current value

of 9A (AC) and a maximum AC-voltage of 230V. Throughout the heat

transfer experiments, the AC voltage to the ceramic igniter was controlled

manually with a transformer. For the ignition experiments, the coil igniter

was fed with DC voltage from a DC EA-PS 8065-10 T power supply unit,

which had a maximum rated voltage of 65V and a maximum current of

10A.

4.4.4 Data acquisition electronics

As already mentioned, both measuring campaigns required the simultane-

ous measurement of the consumed electrical power and the surface temper-

ature of the igniters.

The high electrical resistance of the ceramic igniter and its strong linear

temperature dependence made these measurements possible through a sim-

ple measurement of its voltage drop and its current.

On the contrary, the weak temperature dependence of the coil igniter re-

sistance required the use of a Wheatstone bridge for the measurement of

its resistance. The igniter with its lead cables formed one of the four legs

of the bridge, the final configuration of which is presented in Fig.4.10. The

values of the other resistances of the bridge were selected so that the bridge

would consume the least power possible (see Tab.4.1).

Before each experiment the bridge was balanced at atmospheric tempera-

ture, while the igniter was inserted in a water bath to avoid heating it. The

excitation voltage of the bridge (V b) was set at 5V and the potentiometer

resistance (Rp in Fig.4.10) was adjusted until the bridge output voltage

(V 2) was zero.

The resistance between points A and B resulted from the solution of the

88


Figure 4.10: Resistance bridge for the coiled igniter resistance measure-ment (Ri is the igniter resistance, Rl is the lead wire resistances and thevalues of the other resistances can be seen on Tab.4.1).

the bridge circuit (eq.4.3 & eq.4.4) and the measurement of the bridge

excitation V b and output V 2 voltages. The igniter resistance is equal to

the value RAB reduced by the resistance of the lead cables, which was

Table 4.1: Resistance values for the bridge circuit, measured with anaccuracy of 0.034%.

Name Value (Ω) Temperature dependence ( ppm/◦C)R1 0.9735 100R3 9994.8 no temperature changeR4 999.64 no temperature changeRp 0-500 no temperature change

89


measured in a separate experiment (see 4.4.2).

RAB =R1 ·

(V2

Vb+A

)1− V2

Vb−A

(4.3)

A =R3

R3 +R4 +Rp(4.4)

Table 4.2: Accuracy of the electrical measurements on the bridge circuit.

Parameter Measurement accuracyV1 2mVV2 0.25mVVb 30mVI 20mA

The current flowing through the igniter leg of the bridge was indirectly

measured from the voltage drop V 1 across the known resistance R1. This

resistance has a very weak temperature dependence, it was air-cooled and

its temperature was recorded with a K-type thermocouple.

This measurement offers two possibilities for the computation of the con-

sumed electrical power from the igniter. The first is based on the measured

igniter resistance and the known current flowing through it (eq.4.5).

Pign =

(V 1

R1

)2

· (RAB −Rlead) (4.5)

The second calculation is a result of an energy balance in the entire cir-

cuit, since all the currents and the electrical resistances are known for each

operational point.

90

4.5 Heat transfer experiments in the combustion chamber

4.5 Heat transfer experiments in the

combustion chamber

To avoid accidental ignitions during the heat transfer measurements in the

ternary mixtures ethanol-water-oxygen, either oxygen or ethanol had to be

replaced by an inert fluid. At the same time, this replacement should have

a minimal influence on the physical properties of the system because the

heat transfer measurements were designed to model the subsequent igni-

tion experiments. As analyzed in section 4.3.1, the behavior of supercritical

mixtures of water with nitrogen or oxygen is almost identical. Moreover,

many studies on LOX jets in supercritical oxygen for the cryogenic propul-

sion engines have been carried out with nitrogen instead of oxygen [91]. On

the other hand, if we had replaced ethanol with another alcohol the result-

ing mixture would have been considerably different from the combustion

mixture (water,ethanol, oxygen) [92, 93, 94].

4.5.1 Measurement procedure

The heat transfer measurements were performed at a constant pressure of

260±2 bar and the fluids were heated at temperatures 350 - 420 ◦C. Four

fuel compositions were used (Cf = 0%, 10%, 20% and 30%wt.) and the

operational conditions of the measurements are presented on Tab.4.3.

The mass flow rate of the CW1 stream was kept equal to 150 kg/h through-

out the experiments, while the igniter was inserted 45mm in the con-

trol volume of the measurements (see Fig.4.11(b)). Two heat flux values,

0.5 & 0.7 (±0.05)MW/m2, were used for each operational point to study

the influence of this parameter on the heat transfer coefficient.

The measurement procedure comprises the following steps:

1. Once the predefined fluid temperatures were reached, power was fed

to the igniter.

91


Table 4.3: Operational conditions for the heat transfer experiments.

Cf mf mN2

[%wt.EtOH ] [ kg/h] [N l/min]

0

10

130203040

10

10

130203040

20

10 1302030 18035 220

30

10 1301520 18025 220

2. The power was increased stepwise to a value resulting in a heat flux

of 0.5MW/m2.

3. The resulting voltage, current and the bulk temperature were simul-

taneously recorded for sixty seconds.

4. The power was adjusted to 0.7MW/m2 and the second measurement

was carried out in the same way as the first.

5. Once the sixty seconds of the second measurement elapsed, the power

to the igniter was reduced in small steps to avoid its breakage due to

thermal stresses.

6. After switching off the power, the value of the bulk temperature was

recorded again for sixty seconds.

Since the heating of the igniter resulted in small but measurable changes in

the bulk temperature of the mixtures, an equivalent temperature was used

92


Igniter

Cooling waterinlet:350 (kg/h),20°C

Oxygen Inlet:λ:1.2 , 390°C

Fuel inlet:20,30 (kg/h)370°C-420°C

Thermocouple

Tbulk

Pressure vessel space: 260 bar

(a) Combustion chamber with opera-tional conditions and the coil igniter.

9.512

2

4570

15.821.322.5

26

(b) Combustion chamber with theceramic igniter. The dimensionsare given in mm.

Figure 4.11: The assembly used during the ignition experiments. Theigniter is only schematically illustrated for the case of the ceramic igniter.The coil igniter protruded 8mm less in the combustion chamber, but theposition of the thermocouple was kept the same for both of them.

for the interpretation of the data. This temperature was the mean value

of the bulk temperatures with and without heating that were measured

during steps 3 and 6 respectively. Furthermore, a radiation correction of

the surface temperatures was accounted for, by computing the radiation

heat exchange of two concentric cylinders. The igniter surface was taken

as the inner cylinder and the inner wall of the combustion chamber as the

outer cylinder. The igniter surface was considered always at its measured

temperature and the data on the emissivity of silicon nitride was taken

from Ravindra et al. [95]. The temperature of the outer cylinder was

always taken equal to that of the cooling water CW1, and the emissivity

93


50 100 150 200 250 3000

50

100

150

200

Time [s]

Vign[V

]

Figure 4.12: Voltage ramp example for a heat transfer measurement.

of Inconel 625 for 25 ◦C was used.

4.5.2 Heat transfer coefficient results

The heat transfer experiments were designed as a preparation for the igni-

tion experiments. Their primary objective was to measure the heat transfer

coefficient in the combustion chamber, in order to optimize the dimension-

ing of the igniters. As the pseudo-critical points of the various ternary

mixtures could only be estimated from the available literature data, their

measurement was a secondary aim of the experiments. In parallel with this

measurement we aimed at the identification of the heat transfer mode of

the mixtures at the operational points of the experiments. It was crucial

for the design of the subsequent ignition experiments to find out, whether

the mixtures were at their heat enhancement or heat deterioration mode.

Finally, the composition and the mass flow rate of the mixtures were pa-

rameters easily adjustable with the plant and the study of their influence

on the heat transfer conditions would support the future ignition strategy

of the facility.

94


The heat transfer experiments have shown that the mixtures were at their

heat enhancement region for the utilized heat fluxes. Their pseudo-critical

points were found between 293-330 ◦C, depending on their composition.

Our results were qualitatively consistent with the data in the respective

literature (section4.3), even though a quantitative comparison was not pos-

sible.

The results for a water-nitrogen and a water-ethanol-nitrogen mixture are

presented in Fig.4.13(a) and Fig.4.13(b) respectively. Both systems dis-

played an enhancement of the heat transfer coefficient for the low heat flux,

followed by its impairment when the heat flux was increased. An increase

of the heat flux by 40% almost halved the peak value of the heat transfer

coefficient for the binary mixture, whereas it reduced that of the ternary

mixture by approximately 25%. As a consequence of the used measure-

ment methodology, the error of a heat transfer coefficient measurement is

inversely proportional to the square of the temperature difference between

the surface and the fluid. As high values of heat transfer coefficients lower

this temperature difference, the errors of the measurements for the low heat

flux in Fig.4.13(a) are considerably higher.

To observe the influence of the ethanol concentration on the heat transfer

coefficient, the flow rate of the fuel and nitrogen streams were kept constant,

and the fuel concentration was changed (Cf= 0%, 10%, 20%). The results

of these measurements are presented on Fig.4.14.

An analysis of the data in Fig.4.13 shows that the heat transfer coefficient of

the respective ternary mixture peaks at 293 ◦C, while that of the binary one

has its maximum value at 330 ◦C. Both cases show a decrease of the pseudo-

critical temperature in comparison to that of water. Similarly, the data in

Fig.4.14 reveals that the use of a fuel stream with Cf = 10%wt. shifts

the pseudo-critical temperature from around 322 ◦C of the simple water-

nitrogen binary mixture to 317 ◦C. Likewise, when 20%wt. of the fuel

stream is ethanol this temperature becomes 307 ◦C. This fall of the pseudo-critical temperature of the mixtures due to ethanol addition is qualitatively

and quantitatively consistent with that of the critical temperature of simple

95


300 310 320 330 340 350 360 3700

1000

2000

3000

4000

5000

6000

7000

8000

Tbulk [◦C]

h[ W

m−2K

−1]

q=

[0.5MWm−2

]

q=[0.7MWm−2

]

(a) Water-nitrogen mixture: Cf = 0% wt., mf = 40 kg/h, mN2=130N l/min.

270 275 280 285 290 295 300 305 310 315 3200

500

1000

1500

2000

2500

3000

3500

Tbulk [◦C]

h[ W

m−2K

−1]

q=

[0.5MWm−2

]

q=[0.7MWm−2

]

(b) Water-ethanol-nitrogen mixture: Cf = 30% wt., mf = 25 kg/h, mN2=

130N l/min.

Figure 4.13: Influence of the heat flux on the convective heat transfercoefficient for a water-nitrogen and a water-ethanol-nitrogen mixture.

96


280 285 290 295 300 305 310 315 320 325 330 3350

500

1000

1500

2000

2500

3000

Tbulk [◦C]

h[ W

m−2K

−1]

Cf = 0%wt.Cf = 10%wt.Cf = 20%wt.

Figure 4.14: Convective heat transfer coefficient values for water-ethanol-nitrogen mixtures with different fuel compositions (Cf=0%, 10%, 20%).Conditions: mf= 20 kg/h, mN2= 130N l/min, q= 0.7MW/m2.

water-ethanol mixtures [75, 76].

The impact of the nitrogen concentration was studied by keeping the ni-

trogen flow rate and the ethanol concentration constant and changing the

fuel flow rate. This methodology changed the concentration of all the three

constituents of the mixture and would generally not isolate the influence of

individual components on the characteristics of the mixture. However, the

changes of the nitrogen concentration in the resulting mixtures were much

higher than these of the other two constituents. As a result the qualitative

study of its influence on the heat transfer coefficient and the pseudo-critical

temperature of the mixtures was possible. The results of these measure-

ments are presented in Fig.4.15(a) for a water-nitrogen and in Fig.4.15(b)

for a water-ethanol-nitrogen mixture.

All the mixtures in Fig.4.15 indicate that the addition of nitrogen in water

or in a binary water-ethanol mixture reduces its pseudo-critical tempera-

ture. In Fig.4.15(a), a mixture with a nitrogen concentration of 18% wt.

demonstrates a reduction of the pseudo-critical temperature by approxi-

97


280 290 300 310 320 330 340 350 360 3700

500

1000

1500

2000

2500

3000

3500

4000

Tbulk [◦C]

h[ W

m−2K

−1]

mf = 20

[kgh−1

]

mf = 30[kgh−1

]

mf = 40[kgh−1

]

(a) Water-nitrogen mixtures: mN2= 130N l/min, q= 0.7MW/m2.

280 290 300 310 320 330 340 350 360 3700

500

1000

1500

2000

2500

3000

3500

4000

Tbulk [◦C]

h[ W

m−2K

−1]

mf = 20

[kgh−1

]

mf = 30[kgh−1

]

mf = 40[kgh−1

]

(b) Water-ethanol-nitrogen mixtures: Cf=10%wt, mN2=130N l/min, q=

0.7MW/m2.

Figure 4.15: Influence of the mixture composition on the convective heattransfer coefficient for water-nitrogen and water-ethanol-nitrogen mixtures.

98


mately 53 ◦C compared to that of water. By the same token, mixtures

with nitrogen concentrations of 23% wt. and 31% wt. have pseudo-critical

temperatures 335 ◦C and 320 ◦C respectively. The similarity of our results

with these of Rogak [78] underlines the validity of the assumption that the

mixtures of water with oxygen and nitrogen have very similar properties in

their trans-critical region.

Following the analysis of section 4.3, we can argue that the changes of the

heat transfer coefficient follow these of the specific isobaric heat capacity

around the pseudo-critical point of the mixtures. In this context, the in-

fluence of ethanol and nitrogen addition on the cp of the mixtures can be

qualitatively observed from their influence on the heat transfer coefficient.

Hence, Fig.4.14 shows that in general ethanol addition lowers the peak

value of the cp around the pseudo-critical point. The same conclusion can

be drawn from the comparison of the heat transfer impairment between the

ternary and the binary mixtures in Fig.4.13. The impairment experienced

from the binary mixture is much stronger because the peak value of its cpis higher than that of the ternary mixture. The same trends can be seen

for the addition of nitrogen to water or to binary mixtures of water and

ethanol. As presented in Fig.4.15 higher nitrogen concentrations lead to

lower peaks of the cp of the mixtures. Again both effects, for ethanol and

nitrogen addition, are consistent with the respective literature.

Finally, a quantitative comparison of the heat transfer coefficients reported

here, with the values reported from Yamagata et al. [81] and Rogak [79],

shows that our values are 30-70% lower. The differences with the results of

Yamagata et al. should be most probably attributed to the use of water-

ethanol-nitrogen mixtures in our measurements, whereas they used only

water. On the contrary, the heat flux values used here are two to five

times higher from the ones implemented by Rogak, which could explain

the observed differences in the results. Moreover, we must stress that our

experiments reported the average values of the heat transfer coefficient over

the surface of the igniter. In addition to that, we did not perform our mea-

surements in a tube but in a highly turbulent flow at the wake of a bluff

99


body. These facts hinder the analysis of the local heat transfer coefficient

and make the quantitative comparison to existing literature extremely dif-

ficult.

4.5.3 Conclusions of the heat transfer measurements

in the combustion chamber

The presented experiments give a clear view of the heat transfer conditions

in the combustion chamber. Moreover they present valuable quantitative

and qualitative data for the ignition experiments that are presented in the

next section.

In this stage of the project we were able to develop our ignition strategy by

taking into account the experimental results and the respective literature

from section 4.2.1. The experiments have outlined the heat flux regions,

in which the mixtures demonstrate heat transfer enhancement, impairment

and potential deterioration. Depending on the still unknown chemical be-

havior of the combustion reactions, one of these heat transfer regions could

be chosen to achieve ignition. In case very high surface temperatures are

needed, the mixtures should be heated above their pseudo-critical tempera-

ture in order to ensure a low heat transfer coefficient. This choice could be

adapted accordingly if lower temperatures would suffice, in order to avoid

very high reactant temperatures.

As the bulk temperature of a mixture plays a significant role in the ignition

process, the capacity of the combustion chamber cooling system was also

of crucial importance. Apart from the flow rate of cooling water, this

intensity proved to be a function of the mixture composition, thus offering

an additional controlling parameter for the ignition experiments.

Finally, although the ceramic igniter failed as a long term ignition source,

the acquired data served as a basis for the dimensioning of the coil igniter.

The resistance and geometry of the latter have been adapted to the mea-

sured heat transfer coefficients, so that surface temperatures of at least

100

4.6 Ignition experiments

800 ◦C could be reached with the available equipment.



Tab.4.4 presents the operational points of the plant, chosen for the igni-

tion experiments. In the course of the experiments this list was adapted,

because we were not able to ignite mixtures with an ethanol concentration

below 12.5%. Furthermore, the mixtures with mf=30 kg/h and Cf=17.5%

and mf=20 kg/h and Cf=20% led to severe ignition phenomena and were

performed only for a limited number of cases.

Table 4.4: Operational conditions for the ignition experiments.

mf Cf mO2 THX2 THX1

[ kg/h] [%wt. EtOH] [N l/min] [ ◦C] [ ◦C]

20

7.50% 4510% 60

12.50% 7515% 9017.5% 105 420,410,400,395

39020% 119 390,385,380,370

30

7.50% 6710% 90

12.50% 11115% 13417.5% 155

The experimental procedure started with heating the reactants. Once the

desired reactant temperatures were reached, ethanol was introduced in the

fuel stream. As soon as the same concentration was established throughout

the fuel tubing, the oxygen flow rate was set to the value corresponding

101


0 1 2 3 4 5 6 7 8 90.32

0.34

0.36

0.38

0.4

0.42

0.44

Time [s]

BridgeVou

t[V

]

Figure 4.16: Ignition occurrence, identified by the step in the bridge out-put voltage.

to λ = 1.2. The thermocouple was then positioned in the combustion

chamber for two minutes and the bulk temperature of the mixture was

recorded. Following the temperature measurement, the thermocouple was

removed and the power to the igniter was switched on. The current of the

igniter was increased in 0.2A steps, while each value was applied for 20 s

and three steps were performed. After the third step, the power to the

igniter was switched off for one minute to minimize heating-up of the lead

wires. Ignition was identified from the abrupt increase of the bridge output

voltage (see Fig.4.16). Directly after ignition the power to the igniter was

switched off and the igniter was withdrawn from the combustion chamber.

The output of the bridge was recorded with a 10Hz frequency and the

reported ignition temperature is the average value of the last three before

ignition.

102


4.6.2 Ignition experiments results

Figures 4.17 and 4.19 show the influence of the bulk temperature and the

fuel concentration (Cf ) on the ignition power and temperature respectively.

In the course of the ignition experiments, the combustion chamber load

- defined as the consumed fuel power divided by the chamber volume -

could be changed either by adjusting the fuel concentration, its mass flow

rate, or eventually both. For the experiments dealing with the influence of

the chamber load on the ignition parameters, only the mass flow rate was

adjusted. Hence, figures 4.18 and 4.20 present the impact of the mass flow

on the ignition conditions of a fuel stream with Cf= 15%wt.

Specifically, figures 4.17(a), 4.18(a) and 4.17(b) present the ignition power

for mixtures with an increasing ethanol concentration and the same fuel

mass flow rate (mf= 20 kg/h). A closer look at these figures leads to the

expected conclusion that mixtures with higher ethanol concentration re-

quire a lower ignition power. An increase of this concentration from 12.5

%wt. to 17.5 %wt. was sufficient to almost halve the ignition power. Fig-

ures 4.17(b) and 4.18(a) present a further common behavior of the ignition

power, which stays almost constant for low bulk temperatures and starts

falling above a certain temperature. This temperature value is not con-

stant for all the mixtures, but it shows a change consistent with that of

their pseudo-critical temperature. A higher ethanol concentration lowers

this temperature exactly as it lowered the pseudo-critical temperatures of

the mixtures in section 4.5. It is thus plausible to claim that this temper-

ature approximates the pseudo-critical temperature of each mixture.

The behavior of the ignition power may be thus explained from the tran-

sition of the mixtures from sub-critical to supercritical. As long as the

mixture is below its pseudo-critical temperature the liquid fuel stream is

not miscible with the gaseous oxygen stream. Consequently, a two-phase

mixture is formed and a surface tension between the two phases is present.

In fact, the power inserted from the igniter to the mixture is the sum of the

power needed for its phase transition near the heated surface and for its

103


300 305 310 315 320 325 330 335 340 34560

80

100

120

140

160

180

200

220

240

Tbulk[◦C ]

Qign[W

]

(a) Ignition power for the fuel stream with Cf = 12.5%wt.

300 305 310 315 320 325 330 335 340 34560

80

100

120

140

160

180

200

220

240

Tbulk[◦C ]

Qign[W

]

(b) Ignition power for the fuel stream with Cf = 17.5%wt.

Figure 4.17: Comparison of the ignition power of fuel stream mixtureswith Cf= 12.5% wt. and 17.5% wt. and mf= 20 kg/h.

104


subsequent ignition. When the mixture temperature exceeds its pseudo-

critical value a single phase mixture is formed and all the electrical power

from the igniter is consumed on the ignition of this mixture. At even

higher temperatures the mixture behaves like a gaseous combustible and

the ignition power falls with increasing bulk temperature.

A comparison with the existing literature of forced ignition becomes more

interesting, when ignition temperatures are considered (Figures 4.19 and

4.20). For all the cases, the ignition temperature starts from high values in

the two-phase region of the mixtures, it then drops to a minimum around

their pseudo-critical point and it starts increasing again for higher bulk

temperatures. These results are in contrast to conventional forced ignition

experiments. The data of Mullen et al. and Laurendeau [57, 58] displayed

a steady decrease of the ignition temperature for increasing bulk tempera-

tures of the combustible. Additionally, they pointed out that higher con-

vective heat transfer coefficients generally result in higher ignition temper-

atures. According to the results of section 4.5, the heat transfer coefficient

has its maximum values around the pseudo-critical point of the mixtures,

whereas the lowest ignition temperatures are observed at this point. Figures

4.18 and 4.20 also reveal that an increase in the flow rate of the combustible

causes practically no change on the ignition temperatures, but only on the

ignition power. This fact indicates that the observed ignition phenomena

(and the resulting temperatures) are dominated by the properties variation

of the mixtures and not by the turbulence levels of the flow in the com-

bustion chamber. This is also the fundamental difference to the ignition

experiments found in the literature, where the change of the heat transfer

coefficient is attributed primarily to changes in the turbulence levels of the

flow.

The variation of the ignition temperature as a function of the bulk tem-

perature can be explained by virtue of the mechanisms behind the increase

of the heat transfer coefficient in supercritical fluids. The phenomena can

be divided in three bulk temperature regions, the first of which is the

sub-critical, two-phase region. In this region, the two phase nature of the

105


300 305 310 315 320 325 330 335 340 34560

80

100

120

140

160

180

200

220

240

Tbulk[◦C ]

Qign[W

]

(a) Ignition power for mixtures with mf = 20 kg/h.

300 305 310 315 320 325 330 335 340 34560

80

100

120

140

160

180

200

220

240

Tbulk[◦C ]

Qign[W

]

(b) Ignition power for mixtures with mf = 30 kg/h.

Figure 4.18: Comparison of the ignition power of fuel mixtures with Cf=15%wt. and mf= 20,30 kg/h. This data presents the influence of thecombustion chamber load on the ignition power.

106


mixture makes high heat fluxes necessary for ignition. Since the mixtures

are in the near-critical zone, higher heat fluxes lower the heat transfer

coefficient of the flow and lead to heat transfer impairment. Hence, the

mixtures ignite at high temperatures that are comparable with the ones in

the conventional forced ignition of two-phase mixtures.

The second temperature region is the vicinity of the pseudo-critical point.

In this region, heat transfer enhancement occurs (see section 4.5) and heat is

transferred from the surface to the fluid with a small temperature difference

between them. Hence, the temperature of a considerable fluid mass near

the heated surface has almost the same temperature with the surface. If

the surface temperature exceeds the self-ignition of the mixture and this

fluid amount is sufficient to sustain and propagate a flame, ignition might

occur even though the surface temperature is low in comparison to that

expected from conventional forced ignition.

Finally, in the supercritical one-phase region the mixtures behave more like

turbulent gaseous mixtures. The higher the bulk temperature of the mix-

ture the more it will resemble a perfect gas and its ignition temperatures

will approximate the ones of conventional gases. Furthermore, all the mix-

tures are already close to their self-ignition temperature. Thus, an increase

of their bulk temperature is generally expected to decrease the ignition

power as was already demonstrated in the respective figures.

In conclusion, the contrast with the conventional ignition studies stems

from the mechanisms underlying the increase of the heat transfer coeffi-

cient. Higher heat transfer coefficients in turbulent flow mean stronger

dissipation of heat in the fluid and thinner boundary layers. The heat

transfer enhancement in supercritical fluids has a completely different ori-

gin and must not be confused with classical convection phenomena.

The presented conceptual model also explains the variation of the mini-

mum ignition temperatures for various ethanol concentrations. These tem-

peratures fall slightly for increasing ethanol concentration (figures 4.17(a),

4.18(a) and 4.17(b)), a fact that could be attributed to the expected de-

107


300 305 310 315 320 325 330 335 340 345450

500

550

600

650

700

750

800

Tbulk[◦C ]

Tign[◦C]

(a) Ignition temperatures for the fuel stream mixture with Cf=12.5%wt.

300 305 310 315 320 325 330 335 340 345450

500

550

600

650

700

750

800

Tbulk[◦C ]

Tign[◦C]

(b) Ignition temperatures for the fuel stream mixture with Cf=17.5%wt.

Figure 4.19: Comparison of the ignition temperatures of fuel stream mix-tures with Cf= 12.5% wt. and 17.5% wt. and mf= 20 kg/h.

108


crease of the self-ignition temperatures for higher ethanol concentrations.

On the contrary, a comparison of the ignition temperatures away from the

pseudo-critical point, reveals a much weaker dependence on the ethanol

concentration than that of the respective ignition power. This general be-

havior is consistent with the data presented from Weber [30] and Wellig

[37] on the self-ignition of methanol-oxygen hydrothermal flames. Unfortu-

nately, there is no systematic data on the forced ignition of hydrothermal

flames, and the data presented from Steinle [28], cannot be compared to

that produced in the current work.

Finally, as the stability of hydrothermal flames has been the main focus of

many investigations in the past, it could be stated here that the resulting

flames were stable. With some exceptions, they continued to burn after

their ignition even for bulk temperatures as low as 270 ◦C. Only mixtures

with mf = 30 kg/h, sub-critical bulk temperatures and Cf = 12.5%wt.

resulted in flames that did not continue to burn after their ignition.

4.6.3 Conclusions of the ignition experiments

The influence of the bulk temperature, the fuel composition and the com-

bustion chamber load on forced ignition of trans-critical water-ethanol-

oxygen mixtures was investigated. The enhanced heat transfer phenomena

in trans-critical fluids led to unexpected results; the ignition temperature

had high values below the pseudo-critical point of the mixtures, it decreased

to a minimum around this point and it increased again for higher mixtures

temperatures. Temperature values between 500 ◦C and 850 ◦C and power

values between 60W and 300W were measured on the igniter.

The values of the ignition power for two phase mixtures have proven the

feasibility of hot surface forced ignition for a future field implementation

of hydrothermal spallation drilling. Electrical power in the same order of

magnitude could be provided from an underground generator. In case this

power is limited, the flexible design of the igniter allows the reduction of its

surface area, so that the same heat flux values can be reached with lower

109


300 305 310 315 320 325 330 335 340 345450

500

550

600

650

700

750

800

Tbulk[◦C ]

Tign[◦C]

(a) Ignition temperatures for mf = 20 kg/h.

300 305 310 315 320 325 330 335 340 345450

500

550

600

650

700

750

800

Tbulk[◦C ]

Tign[◦C]

(b) Ignition temperatures for mf = 30 kg/h.

Figure 4.20: Comparison of the ignition temperatures of fuel stream mix-tures with Cf = 15% wt. and mf = 20,30 kg/h. This data presents theinfluence of the combustion chamber load on the ignition temperatures.

110

4.7 Flame temperature profile measurements

electrical power.

Finally, the durability of the igniters has been demonstrated as well, and

ignition modules have been used in extreme conditions for more than 150

operating hours and 100 ignitions without replacement.


4.7.1 Scientific problem, motivation and aims

The primary objective of the flame axial temperature measurements was

the performance characterization of the new pilot plant and its combus-

tion chamber. Another very important aspect of these measurements was

their relevance to the impingement heat transfer experiments. The value

of the heat transfer coefficient of an impinging flame is calculated from

the measured heat flux and a reference fluid temperature, which must be

consistently defined and measured. In the present thesis, the flame temper-

ature at a point inside the combustion chamber is defined as this reference

fluid temperature.

Prikopsky [53] performed flame temperature measurements with the goal

to define the length of the flames in his system. His axial flame temper-

atures showed the characteristic development inside a coaxial combustion

chamber. The new combustion chamber produces a mixed flow and a very

intense recirculation zone, which is expected to result in different flame

temperature profiles. In particular, the temperature values will give an in-

dication about the position and the intensity of this bluff body recirculation

zone.

111



Flame temperatures were measured with the same thermocouple installed

for the measurement of the bulk temperature during the ignition and heat

transfer experiments. After ignition, the thermocouple was positioned

5mm below the exit of the chamber and it traveled 35mm, in 1mm steps,

thus measuring the temperature up to a point 30mm inside the cham-

ber. The starting and ending points of the measurements are indicated in

Fig.4.21.

(a) Starting position; x=0mm. (b) Ending position; x=35mm.

Figure 4.21: Starting and ending positions of the thermocouple for theaxial flame temperature profile measurements.

The temperature was recorded at each point for 60 s with a frequency of

1Hz. The reported values are average values without accounting for any

correction due to radiation. Once the measurements for one flame were

finished, the flame was put out by reducing the concentration of ethanol in

112


the fuel stream. The new operational conditions of the plant were set and

the whole procedure of ignition - temperature measurement - extinction

was repeated.

4.7.3 Experimental results

The measurements focused on the dependence of the flame temperature on

the reactants temperature before ignition, the fuel ethanol concentration

and the combustion chamber load. The combustion chamber load was

adjusted in the same way as during the ignition experiments, by changing

the fuel flow rate and keeping its concentration constant. On the other

hand, the influence of the fuel concentration was so strong that it could be

observed at a constant chamber load and a constant fuel flow rate.

Fig.4.22 shows that an increase in the bulk temperature of the mixture

resulted in an increase of the temperatures in the flame. This is consistent

with a simple energy balance in the combustion chamber, with the assump-

tion of complete combustion in it. A higher mixture temperature before

ignition results always to a higher adiabatic flame temperature.

Most of the measurements showed a fairly constant temperature in the

chamber up to approximately 10mm before its exit, where the cooling

water started to quench the flame. This effect was more pronounced for

lower chamber loads, where the flame had a lower energy content. Another

remarkable result was that flames with an average temperature of approx-

imately 700 ◦C continued to burn after their ignition, thus demonstrating

the successful design of the combustion chamber.

The effect of the fuel concentration on the resulting flame temperatures is

presented in Fig.4.23. In Fig.4.23(a), the fuel flow rate was kept constant

but the flow rate of oxygen was changed to keep a constant λ, as the fuel

concentration varied. Therefore, the flames presented on this figure had

different chamber loads and the respective mixtures had slightly different

bulk temperatures before ignition. Nevertheless, the rise of the fuel con-

113


0 5 10 15 20 25 30 350

200

400

600

800

1000

1200

x [mm]

Tflame[◦C]

Tbulk = 338 [◦C ]Tbulk = 333 [◦C ]Tbulk = 329 [◦C ]Tbulk = 308 [◦C ]

(a) Axial flame temperatures for the mixtures with: mf = 20 kg/h and P =22.5 kW.

0 5 10 15 20 25 30 350

200

400

600

800

1000

1200

x [mm]

Tflame[◦C]

Tbulk = 338 [◦C ]Tbulk = 333 [◦C ]Tbulk = 329 [◦C ]Tbulk = 308 [◦C ]

(b) Axial flame temperatures for the mixtures with: mf = 30 kg/h and P =33.7 kW.

Figure 4.22: Influence of the bulk temperature before ignition on the axialflame temperature for fuel streams with Cf = 15% wt.

114


centration raised the flame temperature in all the cases. In a similar way,

the results in Fig.4.23(b) illustrate the influence of the fuel concentration

on the flame temperature for a constant chamber load of 29 kW. Likewise,

a clear increase of the flame temperature is seen for higher ethanol concen-

trations in the fuel. This increase is caused by a reduction of the thermal

capacity of the combustion products, due to the lower water concentration

in them. Since the same amount of energy is released from the combustion

reaction, the final temperature of these products is increased.

As already mentioned, the effect of the chamber load on the flame temper-

atures was studied by changing the fuel flow rate while keeping its compo-

sition constant. In fact, an increase of this flow rate always resulted in an

increase of the bulk temperature of the mixtures before ignition. Therefore,

Fig.4.24(a) and Fig.4.24(b) present the influence of the chamber load for a

constant bulk and a constant THX2 respectively.

The flame temperatures of fuel streams with mf = 20 kg/h and 30 kg/h

are compared for THX2 = 420 ◦C in Fig.4.24(b). In this case, a small offset

of the flame temperatures to higher values was observed for the increased

chamber load. This slight increase can be attributed on the one hand on

the higher bulk temperatures of the mixtures due to the higher fuel flow

rate. On the other hand, an increase of the fuel mass flow rate increases

the energy content of the combustion products. Since the cooling intensity

of the combustion chamber was kept constant throughout the experiments,

the final temperature in the chamber rose.

Likewise, Fig.4.24(a) shows the flame temperatures of two fuel streams

with mf = 20 kg/h and 30 kg/h but with the same bulk temperature of

338 ◦C. This comparison reveals an additional effect of the combustion

chamber operation. The flame temperature of the 30 kg/h case declined for

the positions closer to the fuel injection nozzle. This nozzle has a central

hole, through which the igniter is inserted in the combustion chamber and

removed after ignition. In fact, the igniter partly blocks this hole but a

weak, central fuel jet is still produced through it. This jet is stronger for

higher flow rates of the fuel stream. If these higher flow rates are combined

115


0 5 10 15 20 25 30 350

200

400

600

800

1000

1200

x [mm]

Tflame[◦C]

Cf = 12.5%wt.Cf = 15%wt.Cf = 17.5%wt.

(a) Axial flame temperatures for the mixtures with: mf = 20 kg/h and THX2

= 390 ◦C.

0 5 10 15 20 25 30 350

200

400

600

800

1000

1200

x [mm]

Tflame[◦C]

Cf = 12.5%,mf = 30[kghr−1

]

Cf = 17.5%,mf = 20[kghr−1

]

(b) Axial flame temperatures for a constant combustion chamber load of P= 29 kW and THX2 = 400 ◦C.

Figure 4.23: Influence of the fuel stream composition on the axial flametemperature for either constant mf or constant combustion chamber load.Some profiles are incomplete because the maximum operation temperatureof the thermocouple was reached.

116


with low reactants temperatures, like in Fig.4.24(a), the reaction rate in

the flame front is decreased and the flame front moves downstream in the

combustion chamber. Thus, the regions near the fuel injection nozzle had

lower temperatures due to the quenching effect of this central fuel jet.

These effects could provide an explanation for the weak vibrations expe-

rienced while operating the plant with fuel stream mixtures of Cf = 15%

wt. and low reactants temperatures. Most probably, the low combustion

intensity led to a rapid upstream and downstream movement of the flame

front, which was experienced as weak vibrations from the operator point

of view.

4.7.4 Conclusions of the axial flame temperature

measurements

Although the axial flame temperature measurements were limited to the

cases, for which the temperature stayed below 1200 ◦C, the general trends

and capabilities of the plant and its combustion chamber have been identi-

fied. Temperatures well above 1000 ◦C have been reached with low ethanol

concentration and these values exceeded 1200 ◦C when fuel stream mixtures

with Cf = 20% were used.

Both the fuel flow rate and its composition proved to be very efficient con-

trolling parameters for the flame temperatures and the load of the chamber.

These two parameters will also be central for all the future flame imping-

ing experiments. Their control is easy and can be performed online, thus

offering a valuable operational tool.

The chamber loads of the presented experiments were limited to 40 kW,

mainly due to safety reasons. A maximum outlet temperature of the

pressure vessel of approximately 80 ◦C was observed with a cooling wa-

ter (CW1) flow rate of 350 kg/h. Given this flow rate can be increased to

650 kg/h with the present setup, it is concluded that the maximum power

of the plant can be safely reached and continuously operated.

117


0 5 10 15 20 25 30 350

200

400

600

800

1000

1200

x [mm]

Tflame[◦C]

mf = 30[kghr−1

]

mf = 20[kghr−1

]

(a) Axial flame temperatures for Tbulk = 338 ◦C and Cf = 12.5% wt.

0 5 10 15 20 25 30 350

200

400

600

800

1000

1200

x [mm]

Tflame[◦C]

mf = 30[kghr−1

]

mf = 20[kghr−1

]

(b) Axial flame temperatures for THX2 = 420 ◦C and Cf = 15% wt.

Figure 4.24: Influence of the combustion chamber load on the axial flametemperature for a constant Tbulk and a constant THX2.

118

Chapter 5

Design of the hydrothermal

spallation drilling tool

The design of the combustion system, presented in section 3.2, concen-

trated on the combustion reaction and its efficiency. Likewise, the design

of the drilling tool focuses on the flame-jet formation and the optimiza-

tion of hydrothermal flame impingement heat transfer. The behavior of

hydrothermal flames as free jets in liquid water is different from that of

conventional flames utilized from Rauenzahn [19] and Wilkinson [20]. In

order to approach the problem of water entrainment, and to tackle the

absence of data on the impingement heat transfer of hydrothermal flames,

the following experimental plan has been devised:

• In a first step several configurations of the combustion chamber outlet

are constructed.

• These nozzles are subsequently characterized according to the quality

of the achieved control of the entrainment phenomenon.

• The nozzles with the most promising performance during these ex-

periments are selected and used in drilling tests.

• The flame impingement heat transfer of the nozzles that demonstrate

an acceptable drilling performance is subsequently analyzed.

• The data is then combined to adapt and optimize the nozzle design.

119

Chapter 5 Design of the hydrothermal spallation drilling tool

Heat flux

Drilling

Entrainment

Figure 5.1: Concept for the design procedure of the drilling tool.

The procedure of Fig.5.1 must be repeated several times before the drilling

tool is optimized. The present chapter presents the work performed on all

the individual steps of this design loop. In order to carry out the necessary

heat transfer experiments, heat flux sensors are needed that are able to

operate near a hydrothermal flame. As no such sensors were found in the

market, their design and calibration was part of the aforementioned design

procedure and is presented in sections 5.1 - 5.3. The first impingement

experiments with hydrothermal flames, and the characterization of their

entrainment behavior, when operated as free jets in water, is the topic

of section 5.4. The initial optimization of the drilling tool geometrical

characteristics was an essential part of these experiments. In conclusion,

the first drilling experiments that were carried out with the best version of

this design are briefly presented in section 7.2.1.

120

5.1 Heat flux sensor development


5.1.1 Scientific problem - sensor requirements

The conditions near a hydrothermal flame impose extreme requirements on

heat flux sensors, especially on their chemical compatibility and their tem-

perature robustness. No heat flux sensors able to fulfill these requirements

are commercially available and custom made sensors must be designed,

constructed and calibrated.

The highest operating temperature of a sensor defines the closest point to

the flame where a heat flux measurement can be performed. At the same

time, Rauenzahn [19] reported that the highest temperature rise on the rock

surface during spallation is approximately 500 ◦C. As the sensors will most

probably have a higher heat conductivity than the spalled rock surface and

will be additionaly cooled, their maximum design temperature was set to

700 ◦C. Similarly, heat flux values between 500 kW/m2 and 2000 kW/m2

are expected in hydrothermal spallation drilling [19, 25]. Thus, the sensors

design should ensure that the maximum expected heat flux will not cause

temperatures, which will exceed their maximum operating temperature.

A topic that came up after the heat transfer measurements in the combus-

tion chamber, is the chemical compatibility of the materials. Supercritical

fluids in general, and specifically supercritical water, are known to be chem-

ically aggressive. The fast oxidation of many ceramic materials that are

commercially used as substrates of heat flux sensors should be considered

in the design of the sensor [88],[90].

5.1.2 Heat flux sensors state of the art

Heat flux is a physical quantity that can only be measured indirectly and

every heat flux measurement is intrusive. The measurement concept shown

in Fig.5.2 is the basis of most heat flux measurement methods.

121


qradqrad qconv

qcond

Sensor Basis

Figure 5.2: Energy conservation in a control volume around a heat fluxsensor.

The simplest heat flux measurement is based on the recording of the tem-

perature difference inside a material on two points very close to each other.

The heat flux results from eq. 5.1 under the assumption of one - dimen-

sional heat conduction between these points.

q =κ

δ· (T1 − T2) (5.1)

Godefroy et al. [96] constructed thin film sensors that follow this principle.

Their sensors covered a very small surface area and comprised hundreds

of thermocouples that were connected in series. Half of the thermocouple

junctions were covered with a thermal insulating material. A temperature

difference between the covered and uncovered junctions resulted when heat

was flowing through the sensor. The output voltage of the sensor was

proportional to this temperature difference and hence to the incident heat

flux. Similar sensors were presented from Hager and Diller [97, 98] and

Holbeg and Diller [99], each producing directly a voltage proportional to

the incident heat flux. Fralick et al. [100] and Stefanescu et al. [101]

implemented a similar idea on a Wheatstone bridge, by covering two of its

resistances with a thermal insulating material (see section 5.1.5).

An alternative working principle is that of the circular foil sensor [102]. This

sensor type consists of a thin circular foil made of a thermoelectric material

122


that is welded at the periphery of a cylinder usually made of copper. This

weld forms the first leg of a thermocouple, while its second leg is formed by

a wire of the same material as the cylinder that is welded in the center of the

foil. The resulting thermocouple produces a voltage that is proportional to

the temperature difference between the center and the periphery of the foil.

If the back side of the foil is additionally insulated, the incident heat flux

contributes only to the measured temperature difference. Consequently,

the produced voltage is a direct function of the incident heat flux. This

sensor has been initially designed to work at purely radiative heat transfer

conditions. When used in a convective environment its sensitivity depends

on the convective heat transfer coefficient [103, 104]. The sensor is thus

inadequate for the measurements of the present thesis due to their very

strong convective character.

Sujay et al. [105] presented an interesting design of a high temperature

sensor that generated a thermoelectric voltage in a direction perpendicular

to that of the heat flux. Gifford et al. [106] developed this idea further and

Pullins et al. [107] calibrated the sensor up to 1000 ◦C. In all these cases

the sensor plays also the role of the thermal resistance between the points

where the thermo-junctions are placed.

Transverse thermoelectric sensors also produce a thermoelectric voltage

perpendicular to the heat flux. They are built from materials with either

a natural or an artificial anisotropy of their thermal and electrical prop-

erties. Knauss et al. [108, 109, 110, 111] investigated the behavior of a

naturally anisotropic sensor, while several researchers [112, 113, 114] pre-

sented applications of sensors with an artificial anisotropy. Finally Mann

[115] presented a thorough investigation of an artificially anisotropic ele-

ment as a heat flux sensor. The detailed analysis of this sensor type is the

topic of section 5.1.4.

123


5.1.3 Choice of the heat flux sensors working

principles

The working principle of a sensor is decisive for its size, its construction

and its operation. In turn all these aspects can have a significant impact

on the measurements, due to their intrusive character. Childs [116] pre-

sented the thermal and flow disruption phenomena that are caused from

the presence of a sensor. Both phenomena change the heat flux compared

to that without the presence of a sensor, and this difference must be at the

very least estimated and if possible minimized.

The sensor specifications described in section 5.1.1 narrow the number of

possible sensors for the application in hydrothermal flames. Since the first

generation of sensors focused on their robustness, the sensors presented

from Gifford et al. [106] might be a very good choice. However, the rela-

tively large size of these sensors and the complexity of their construction

would hinder their integration in the spallation drilling system and finally

ruled them out.

As a result, the construction of a transverse Seebeck sensor made of chromel

and alumel was favored. The sensor sensitivity is proportional to its length

with its thickness having only a weak influence. As a result, extreme ro-

bustness can be achieved without making sacrifices on the sensor sensitivity.

Thermal and flow disruption could be also minimized because the sensor

functions as its own thermal resistance and its thermal conductivity can be

adapted to that of its substrate material.

Although the transverse sensor offers a good solution for the measurements

in question, an additional sensor based on the Wheatstone Bridge circuit

has been constructed [117]. The operational characteristics of this sensor

would offer a comparison benchmark for the measurements of the transverse

sensor.

124


5.1.4 Transverse type anisotropic sensor

The general operation principle of the transverse Seebeck sensors is illus-

trated in Fig.5.3. Unlike thermopile sensors they do not require a thermal

resistive layer and they can be as thin as one micrometer or as thick as

3mm. Hence, they can achieve either extremely short response times and

little temperature disruptions or extreme robustness respectively. More-

over, they provide a great design flexibility and the possibility to place

them on complex geometries.

α

δ

l

Heat

Current

Figure 5.3: Illustration of the transverse Seebeck effect. Heat that flowsin the perpendicular direction to the sensor surface produces a voltage inthe transverse direction.

There are natural anisotropic thermoelectric materials, like YBCO, but an

anisotropy can be also produced artificially by a structure of tilted alternat-

ing layers of thermoelectric materials. Since most of the natural anisotropic

thermoelectric materials cannot withstand temperatures above 350 ◦C, thealternative of an artificial anisotropic sensor was chosen.

Sensor modeling

The modeling of the artificial, transverse thermoelectric sensors starts from

a simple layered construction (see Fig.5.4) of two materials A and B. The

thermal (κ) and electrical properties (S, σ) of the simple layered mate-

125


rial in the x and y axes result from the individual material properties

(SA,B, κA,B, σA,B) and from the thickness ratio of the layers (p) [112].

Sx = SAσA+pSBσB

σA+pσBκx = κA+pκB

1+p σx = σA+pσB

1+p

Sy = SAκB+pSBκA

pκA+κBκy = κAκB(1+p)

pκA+κBσy = σAσB(1+p)

pσA+σB

The properties tensor (X) of the tilted material is computed by a rotation

of the properties tensor of the simple layered material by the tilt angle of

the layers. In this tensor, X is one of the material properties.

X =

⎛⎝Xx · cos2α+Xy · sin2α 0 1

2 · (Xx −Xy) · sin(2α)0 Xx 0

12 · (Xx −Xy) · sin(2α) 0 Xx · sin2α+Xy · cos2α

⎞⎠

S ,κ ,σ

y

x

αMaterial AMaterial B

yyy

S ,κ ,σ xxx

Figure 5.4: Modeling of the transverse Seebeck sensor. Both materials areanisotropic, but only the tilted material is producing a transverse voltage.

A heat flux measurement has the goal to connect the voltage of the sensor

in the x direction with the incident heat flux. However, the incident heat

flux produces a three - dimensional temperature field in the mass of the

sensor. From the definition of the Seebeck coefficient it follows that the

electric field in the sensor due this temperature field will be given from

126


eq.5.2.�E = S • �∇T (5.2)

Thus, the component of the voltage vector in the x direction results from

eq.5.3.

Ex =(Sxcos

2α+ Sysin2α

) ∂T

∂x+

(1

2(Sx − Sy)sin(2α)

)∂T

∂y(5.3)

The first term of the right hand side of eq.5.3 is a function of the tem-

perature gradient in the x direction. The second term in this equation is

connected to the temperature gradient in the same direction with the heat

flux. The output voltage of the sensor in the x direction results from the

integration of eq.5.3 with respect to x.

Vx =

∫ l

0

Exdx (5.4)

From the equations 5.3 and 5.4, we can identify two behaviors of the sensor

depending on its temperature field. If the sensor has a constant temper-

ature in the x direction, the first term of the right hand side in eq.5.3 is

canceled out. The incident heat flux produces only a vertical temperature

gradient and the output voltage is directly connected its total value. This

case corresponds to the one - dimensional heat transfer model, used typi-

cally in all the respective publications [112, 113]. By contrast, an arbitrary

temperature profile can be produced in the x direction. Hence, a percent-

age of the total incident heat flux will flow in this direction and produce

an additional voltage. If the integration of the two terms of eq.5.3 leads

to values with the same sign, the total heat flux is represented from the

measured voltage. In the opposite case the output voltage will represent

the weighted difference of the heat flux flowing in these two directions and

the respective Seebeck coefficients will be the weighting factors. Therefore,

a very careful calibration of the sensor is imperative, in order to account

127


for these phenomena.

Sensor element

CeramicBasisImpingement

Plate

Cooling system

Fixation screw

Figure 5.5: Exploded drawing of the total sensor assembly.

Transverse sensor design and construction

The sensor model has shown that its careful construction is essential for its

adequate operation. The final sensor assembly consists of six components

and its exploded drawing is presented in Fig.5.5. The sensing element is

vacuum soldered on its ceramic basis, which insulates it electrically from the

metal components of the assembly. This basis is in turn vacuum soldered

on the metal cooling system of the sensor and these components are fixed

with a screw in the impingement plate.

In a first step, a stack of thin flat wires (8mm x 0.8mm) is welded by

diffusion welding. The sensor geometry is cut out from the stack by an

eroding cutter, so that the two materials are tilted by an angle α as shown

in Fig.5.6(a). The tilt angle was calculated according to the considera-

tions of Babin et.al on its optimal value [112]. The resulting sensor has a

thickness of 2mm which provides the necessary robustness and facilitates

128


4.3

8

0.849°

1

2

0.5

(a) Schematic of the stacked materials and the sensing element after erosion cutting (di-mensions in mm).

Chromel Tav surface Alumel

Vtrans

Tav

(b) Schematic of the lead wires connections onthe sensor.

Figure 5.6: The geometry of the sensing element and the connections ofits lead wires.

the construction of the ceramic basis. Finally, the protruding, thin plates

provide a clear surface to solder the sensor on its ceramic basis.

The electrical connections of the sensor are realized with three wires, pre-

sented in Fig.5.6(b), that are laser-welded on the sides of the sensor to

measure its transverse voltage and its average temperature. Two chromel

wires are connected to the chromel plates on the sensor sides and measure

the transverse voltage. An alumel wire is welded on the alumel plate, next

to one of these chromel plates. This third wire forms a thermocouple with

the neighboring chromel wire that measures the average sensor temperature

129


Epoxy filling WireCeramic Basis

SensorCeramic capillary Cooling water

channels

Soldered surface

(a) Details of the sensor construction and assembly.The soldering surfaces are visible along with the ce-ramic capillaries and the epoxy protection of the leadwires. The inner thread of the metal basis forms theconnection to the lower positioning device of the highpressure plant.

20

5

(b) Illustration of the cooling system of the sensor afterDeuchert [118]. Four cooling channels with relativeangles of 90◦ are used.

Figure 5.7: The sensor assembly and its cooling system.

130


between the two tilted plates.

The cooling system configuration has a strong influence on the quality of the

measurements. A very intense cooling rate increases the three - dimensional

heat transfer effects, while a weak cooling leads to overheating. Deuchert

[118] studied the geometrical and flow characteristics of the cooling system

and argued that a compromise between these two effects should be found.

The structural result of his suggestions is shown in Fig.5.5. Cooling water

flows through the positioning device, is impinging on the point directly

below the sensor and it flows through the four cooling channels of the

system.

The manufacturing process of the sensor (Fig.5.7(a)) starts with the weld-

ing of the lead wires on the sensing element. These wires are consequently

led through the holes of the ceramic basis, which is in turn vacuum soldered

with the sensing element. The lead wires that protrude from the resulting

body are pulled though the holes of the metal basis and are insulated from

it with three ceramic capillaries. Once the ceramic and metal bases are

soldered, the lower ends of these three holes are filled with epoxy, which is

kept at low temperatures through its direct contact with the cooling water.

This construction method provides increased mechanical stability to the

wires and guarantees their insulation from the other sensor components.

Sensor materials

The working principle and the geometrical details of a sensor have to be

followed by a careful selection of its materials. A wrong choice of the

materials can lead to a severe disruption of the temperature field and the

measurement. The sensing element consists of chromel and alumel, because

of their low price, their high temperature compatibility and their chemical

resistance [119].

Waag [120] and Wang [121] studied the influence of the sensor materials

on its thermal and mechanical behavior. Their studies led to the choice

131


0 100 200 300 400 500 600 700 800 900 10000

10

20

30

40

50

60

T [◦C]

κ[ W

m−1K

−1]

κyyκxx and κzzκAL2O3

(a) Thermal conductivity of the sensing element and its ceramic basis.

0 100 200 300 400 500 600 700 800 900 10000

5

10

15

T [◦C]

α[10−6K

−1]

(b) Thermal expansion coefficient of KOVAR.

Figure 5.8: Properties of the materials used for the sensor constructionand the sensing element.

132


of KOVAR for the metal parts of the sensor and aluminum oxide for the

ceramic basis. The thermal conductivity of the sensor and its ceramic basis

is presented in Fig.5.8(a) as a function of temperature, whereas the thermal

expansion coefficient of aluminum oxide is approximately 8.4 10−6◦C−1.

Finally, KOVAR has an equivalent heat conductivity with stainless steel,

while its thermal expansion coefficient can be seen in Fig.5.8(b).

5.1.5 Thin film resistance sensor

As an alternative to the robust design of the transverse Seebeck sensor,

a thin film sensor was constructed. The main reason for the construction

of the second sensor was the possibility to repeat the measurements of the

transverse sensor and to investigate its influence on the actual heat transfer

coefficient. A further motive was to address the open questions about the

transient heat transfer conditions and the way they influence the drilling

process.

Sensor working principle

The sensor is a thin film Wheatstone resistance bridge, similar to that

presented from Fralick [100] and Stefanescu [101]. Two of the bridge re-

sistances, which are positioned diagonally opposed to each other in the

schematic of the bridge (Fig.5.9), are covered with a heat insulating layer.

At room temperature and with no heat flux present, the resistances have

the same temperature but not the same value, due to limitations of their

construction methodologies. Owing to this difference the bridge output is

equal to an offset voltage that is a function only of the excitation voltage of

the sensor. When heat flows through the sensor, the two uncovered resis-

tances have a higher temperature than the covered ones. The temperatures

of the resistances can be computed with the assumption of one - dimen-

sional heat transfer, because the individual sensor layers are very thin (<

3 �m). The difference of these temperatures results in different values of

133


the resistances and the bridge is brought further off-balance thus producing

an additional output voltage. The output voltage of the sensor results from

eq.5.5 by replacing the resistances with their temperature functions from

eq.5.6 and by making the assumption that T1 = T3 and T2 = T4.

R1

R2 R3

R4

Vin Vout

Covered resistance

Figure 5.9: Schematic of the electrical circuit of the thin film sensor.

Ri = Ri0[1 + α(Ti − T0)] (5.5)

Vout =

(R1R3 −R2R4

(R2 +R3)(R1 +R4)

)Vin (5.6)

Sensor design, construction and materials

Holmberg and Diller [99] presented some important construction details

of their thin film sensor configuration. Even though our sensor does not

use thermocouples, its geometrical arrangement borrowed many of their

solutions. The stress relieving mechanisms they used and the way they dis-

tributed their thermocouple junctions on the available surface were adapted

in our thin film design.

The sensor consists of five almost identical resistances, four of which form

134


the Wheatstone bridge of the sensor and one measures the temperature of

the uncovered resistances. Meander was chosen as their two - dimensional

shape and its pattern was spread equally in the whole measuring surface

(see Fig.5.10(a)). Great caution was given in the positioning of the sensor

lead contacts, which were placed at the periphery of the substrate plate to

avoid overheating them (see Fig.5.10(b)).

Platinum was chosen as the material of the resistances and aluminum oxide

was selected as the substrate material. Furthermore, the thermal insulation

is made of silicon dioxide, whereas gold was applied for the contacts of the

sensor. To conclude, a protective layer of aluminum oxide was deposited

over the whole sensor surface.

The studies of Stalder [122] on the geometry of the resistances led to the

construction of sensors with three resistance thickness values 100, 200, 300

nm and 30 �m width. Moreover, the meander pattern resulted in a resis-

tance length 65.78mm, which was equal for all the resistances. In order to

maximize the sensor sensitivity, we exhausted the potential of the available

construction methods and we built a 3 �m silicon dioxide insulating layer,

while the aluminum oxide protection layer has a thickness of 500 nm.

Sensor cooling system and assembly

After some initial attempts to calibrate the transverse sensor, it became

clear that the cooling of the thin film sensor had to be more intense. A metal

cooling system similar to the one of the transverse sensor was designed and

is presented on Fig.5.11. It consists of two metal plates, one of which

supports the sensor plate while the other is responsible for the positioning

of the whole assembly in the calibration and the high pressure facility. The

four cooling channels of the sensor assembly are located in the upper metal

plate, 2mm below its surface.

The lead cables of the sensor are led through nine holes that are drilled in

its upper plate. They are fixed in a machined passage at the back of this

135


R1R1

R2 R3R3

R4

Vin Vout

Covered resistance

Surface temperature measurement

4.53mm

R1R2 R3

R4

(a) Two - dimensional geometry of the thin film sensor.

Contacts for the temperature resistance measurement

1

12

289

3

3

4

4

5

5

67

(b) Picture of the thin film sensor.

Figure 5.10: 2-D geometry of the thin film sensor with its contacts.

136

5.2 Calibration plant and methodology for heat flux sensors

GG

Lead contact holes

F F

Upper sensor surface

Cable

M1 thread connectionCooling channel

Figure 5.11: Cooling and fixation system of thin film sensor.

plate with high temperature silicone, which additionally seals them from

the cooling water. Small threads (M1) are attached on the top of each

cable and simultaneously realize the electrical connection of the sensor to

the cables and its mechanical fixation to the upper metal plate.

5.2 Calibration plant and methodology for

heat flux sensors

5.2.1 Scientific problem - calibration requirements

The construction of custom made heat flux sensors, makes their calibration

in an environment similar to their intended operation necessary. The sen-

sors were constructed for the measurement of the convective heat transfer

coefficients of impinging hydrothermal flames. All the same, the values

137


of this coefficient that can be achieved with supercritical fluids cannot be

reached by any other fluid type. Therefore, air jet impingement could be

used and the sensors could be calibrated at surface temperatures, compara-

ble to the temperatures of their intended operation. In addition to that, the

calibration facility must provide various testing conditions and support the

optimization of the sensors before their implementation in a high pressure

environment.

The following list summarizes the specifications of the calibration plant:

• A high temperature - high velocity impinging air jet should be used

as a heat flux source.

• The cooling of the sensors must be flexible and make both water and

air cooling available.

• The mass flow of the air must be directly controlled, to provide flexible

control of the convective heat transfer coefficient.

• Heat fluxes as high as 0.5MW should be reached by the plant.

• The calibration error must be comparable with the values reported

in the relevant literature (see section 5.2.2).

5.2.2 Heat flux sensor calibration state of the art

Heat flux calibrations are generally performed with either the transfer or

the absolute method [123, 124]. The absolute calibration connects the

sensor signal directly with a primary standard. Conversely, the transfer

method uses a previously calibrated sensor as a reference and transfers its

calibration function to the sensor to be calibrated. Generally, all modes

of heat transfer can be used (conduction, radiation and convection) for

calibration purposes, but when a mode is chosen the other two have to be

minimized to avoid their interference in the measurements.

The absolute radiative calibration of heat flux sensors is generally per-

formed by inserting the sensor in a black body cavity, while the tempera-

138


ture of the latter is measured. The incident heat flux is calculated from the

cavity temperature and the radiative heat transfer equation. The sensor

sensitivity results from the sensor signal and the known incident heat flux.

The optical technology division and the physics laboratory of the NIST in

the USA have published exhaustive data on both types of radiative heat

flux calibration [124, 125, 126].

While in radiation calibration the radiation wave lengths can add pertur-

bations in the resulting sensitivity, in convective calibration the mode of

convection and the exact flow conditions may have this result. Holmberg et

al. [127, 128] presented a transfer calibration methodology based on shear

flow convection over a heated plate. In their facility, a cold flow of air was

used to cool down the plate, in which a transfer sensor and the sensor to

be calibrated were installed. Borell and Diller [129] developed a similar

absolute calibration facility that used the stagnation flow of cold air on a

heated plate. The temperature of the plate, at the region where the sensor

was installed, was kept constant by adjusting its heating. The heat flux

from the plate, through the sensor and to the flow was set by the heating

system and it was used directly for the calculation of the sensor sensitivity.

Finally, Gifford et al.[130] presented a stagnation flow calibration facility,

which used a transfer calibration methodology. In their case, air was heated

and it was directed through a T-junction before it impinged on two plates,

which were positioned opposite to each other. Each of the two sensors (ref-

erence and to be calibrated) was mounted on one plate. The calibration

was performed by assuming that the flow through the T-junction led to the

same convective heat transfer coefficient on both sensors. The surface tem-

perature of the sensors was measured and the definition of the convective

heat transfer coefficient was used to transfer the calibration function.

Conduction offers the easiest way to calibrate heat flux sensors. ASTM

has published the only official heat flux sensor calibration standard based

on this heat transfer mode [131]. An absolute calibration is performed by

mounting the sensor between two plates, one of which is heated with a

defined flux while the other is guarded to minimize the heat losses. The

139


temperature and the voltage of the sensor are measured and its sensitivity

is defined through the used heating power.

5.2.3 The calibration concepts

Transfer calibration in a stagnation convection environment has been cho-

sen as a concept in the current thesis. We adapted the methodology of

Gifford et al.[130] because air entrainment in high temperature air jets has

a large impact on their behavior [132]. The new concept makes use of the

reproduction of heat flux values in consecutive experiments. We developed

two calibration methods, which are based respectively on the matching of

the heat flux and the heat transfer coefficient between the sensors.

The heat flux matching between the two sensors requires the matching of

their surface temperatures for the same jet. This is possible only if their

construction results in similar thermal resistances between the jet and their

cooling water. In detail, the following experimental steps are taken during

this calibration methodology:

• The sensor to be calibrated is exposed to a specified jet. Its output

voltage and its surface temperature are recorded.

• The transfer sensor is exposed to the same jet and its cooling is ad-

justed to match its surface temperature with that of the other sensor

in the preceding experiment. Once this is achieved, the heat flux is

measured.

• Since both sensors have the same surface temperature, similar size

and the jets are exactly the same (depending on the ability of the

plant to reproduce two identical jets), the heat flux in both cases is

the same. The recorded heat flux from the transfer sensor and the

voltage from the sensor to be calibrated are then connected to form

the sensitivity function of the latter (see eq.5.7).

140


Ssens =Vsens

qref(5.7)

It is obvious that the calibration procedure cannot be limited to sensors

constructed in a similar way to the transfer sensor of the calibration facility.

If a surface temperature match is not possible, the temperature difference

is minimized and the concept presented from Gifford et al. [130] is used.

The steps followed in this case are:

• The sensor to be calibrated is exposed to a specified jet. Its output

voltage Vsens and its surface temperature Tssens are recorded.

• The transfer sensor is exposed to the same jet and its surface temper-

ature is adjusted to a value as close as possible to that of the other

sensor in the preceding experiment.

• The heat flux qref and the surface temperature Tsref are recorded

and the heat transfer coefficient is calculated from eq.5.8.

qref = h · (Tf − Tsref ) (5.8)

The value of the jet temperature in the outlet of the heater is the

reference fluid temperature during this calculation.

• Since both sensors operated with the same heat transfer coefficient,

the heat flux for the sensor to be calibrated is calculated from eq.5.8.

qsens = h · (Tf − Tssens) (5.9)

The sensitivity Ssens of the sensor results from the measured voltage

of the first step and the calculated heat flux from the third step.

Equation 5.10 expresses the sensitivity as a function only of measured

parameters.

Ssens =Vsens

qref· Tf − Tsref

Tf − Tssens(5.10)

The last method is based on the assumption that the convective heat trans-

141


fer coefficient is only a function of the flow properties. This assumption

holds as long as the difference of the surface temperature between the sen-

sors does not lead to a considerable change of the flow. This could be caused

only by an impact of the surface temperature on the material properties

of the fluid. Air properties, and mainly its density, viscosity and thermal

conductivity, stay fairly constant inside a range of 50 ◦C, irrespective of themean temperature value. Thus, a surface temperature difference between

the two sensors in this order of magnitude, would allow the implementation

of this calibration principle.

5.2.4 The convection calibration setup

The plant consists of two sub-systems and its PI diagram is presented in

Fig.5.12(a). One sub-system is responsible for the production and control

of the air jet and the other realizes the positioning and the cooling of the

sensors.

The jet control sub-system comprises a thermal mass flow controller and

a tubular air heater. The mass flow control has a full scale accuracy of

±1.5% and a reproducibility of ±0.5%, while the installed electrical power

of the heater is 3 kW. A PID controller sets its outlet temperature that

is measured with a K-type thermocouple and controlled with an accuracy

of ±3 ◦C. The heater can reach a maximum temperature of 760 ◦C at

mass flow rates of 10Nm3/h and its outlet geometry can be adjusted. We

installed a solid stream nozzle made of Hastelloy B with a 5.33mm diameter

to carry out all the calibration experiments.

The sensors sub-system consists of the construction for their positioning,

the transfer sensor, an infrared thermometer and all the necessary tub-

ing for their cooling. An HFM-8 E/H thermopile sensor, manufactured

by Vatell Corporation [97], is used as a transfer sensor. The 2 �m thick

thermopile is made of Nichrome (80% Ni, 20% Cr) and Constantan (55%

Cu, 45% Ni) and is produced on an aluminum nitride substrate, which is

in turn embedded in a nickel housing. The sensor can operate at temper-

142


FIC

TI

TI

FI

TI

TIC

Air Heater

QI

Impi

ngem

ent p

late

Air Supply 6 bar

Cold water10°C, 7 bar

Hot water45°C, 7 bar

TI

TI

Cooling water systemAir supply system

(a) Piping and instrumentation diagram of the setup.

(b) Top view of the calibration plant.

Figure 5.12: The stagnation convection calibration plant.

143


0 2 4 6 8 10 12 14 16 18 200

50

100

150

200

250

300

350

400

mair[Nm3h−1]

q[kW

m−2]

SOD = 5.3 cmSOD = 10.3 cmSOD = 15.3 cm

(a) Plant performance data measured with the reference sensor.

5 7.5 10 12.5 15180

210

240

270

300

330

360

q[kW

m−2]

5 7.5 10 12.5 155

5.1

5.2

5.3

5.4

5.5

5.6

Uncertainty

[%]

mair[Nm3h−1]

Uncertainty

(b) Uncertainty of the reference heat flux measurements, for an SOD of 5.3 cm.For the error calculation see [133].

Figure 5.13: Performance of the stagnation calibration plant.Tjet =500 ◦C, Tcool = 10 ◦C at Vcool=89L/h.

144

5.3 Calibration of the sensors

atures up to 700 ◦C and the outer diameter of its actual sensing area is

approximately 4.8mm. According to the manufacturer data, the heat flux

accuracy is ±5% and its response time is 10 �s [97]. Likewise, the infrared

thermometer has a spectral response 8...14 �m and it records the surface

temperature of the sensors at the jet stagnation point. The accuracy of the

device is either 1% or 1 ◦C (whatever is higher for the measured temper-

ature) and the reproducibility of its measurements is 0.5% or 0.5 ◦C. The

comparability of the surface temperature measurements during the cali-

bration experiments was achieved by coating all the sensors with the same

black paint with a known emissivity. Finally, Fig.5.12(b) presents a picture

of the positioning and cooling sub-systems.

Fig.5.13(a) shows the heat flux values measured with the reference sensor at

different stand-off distances (SOD) and air volume flow rates. Depending

on the chosen SOD and jet parameters, heat flux values between 20 kW/m2

and 600 kW/m2 are achieved, thus providing a broad calibration spectrum.


In the context of the calibration of the sensors we investigated the influence

of the flow conditions and their surface temperature on their sensitivity.

Parameters like the jet temperature (Tjet), the air volumetric flow rate

(Vair) and the stand-off distance (SOD) were varied during the experiments

to obtain a wide range of heat transfer coefficients and sensor temperatures.

The method of heat transfer coefficient matching was applied for both

sensors, due to difficulties in their surface temperature matching. The

plant operated at each jet temperature for at least 15 minutes before the

measurements started, to avoid errors due to the thermal elongation of the

air heater. For each calibration point measurements were taken at a 10Hz

frequency, while the duration of each measurement session was 90 s. Finally,

a high precision nano-voltmeter (Agilent 34420A) was used throughout the

experiments to perform all the resistance and voltage measurements of the

145


sensors.

5.3.1 Transverse sensor calibration

The experimental plan of the transverse sensor calibration can be seen in

Tab.5.1. The measurements were performed at three stand-off distances,

typical for the future experiments in the high pressure plant.

Table 5.1: Calibration plan for the transverse sensor. For each SOD, allthe jet temperatures were examined, and for each temperature all the airmass flow rates.

SOD [ cm] Tair [ ◦C] mair [Nm3/h]

1.2, 2.2, 3.2

400 4500 5600 6700 7

10

The calibration results are presented in Fig.5.14, along with a simple one -

dimensional model of the sensor. The relatively thick layers of aluminum

oxide and KOVAR behind the sensing element, resulted in surface tempera-

tures reaching 420 ◦C, while the maximum heat flux value was 535 kWm−2.

The temperature measured with the sensor’s internal thermocouple, was

always 20-30 ◦C below its surface temperature. The resulting sensitivity

values are in the range 3-6.5 �Vcm−2W−1, and lie in the same order of

magnitude predicted by both one and three - dimensional models of Waag

[120] and Pavese [134] respectively.

The experimental data shows a decrease of the sensor sensitivity with in-

creasing temperature. In comparison, the one - dimensional model predicts

the opposite trend even though the computed and measured values are of

the same order of magnitude. A further observation in Fig.5.14(b), is that

146


60 80 100 120 140 160 180 200 22015

20

25

30

35

40

45

50

55

Vsens [μV]

q[W

cm−2]

SOD = 1.2cmSOD = 2.2cmSOD = 3.2cm

(a) Output voltage as a function of the incident heat flux for three SODvalues.

100 150 200 250 300 350 4000

1

2

3

4

5

6

7

Tsens[◦C]

S[μVcm

2W

−1]

SOD = 1.2cmSOD = 2.2cmSOD = 3.2cmModel

(b) Sensitivity as a function of the sensor surface temperature. The temper-ature values correspond to the average sensor temperatures measured fromits internal thermocouple.

Figure 5.14: Calibration results of the transverse heat flux sensor.

147


the sensor sensitivity shows an offset for an SOD value of 3.2 cm compared

to that for for SODs 1.2 cm and 2.2 cm. This is also obvious in Fig.5.14(a),

where the measured output voltages are approximately 20-60 �V lower for

this SOD value.

The discrepancy between the model and the experimental data can be at-

tributed to the shortcomings of the one - dimensional model, which is not

able to account for the three - dimensional phenomena and the structural

details of the sensor. The voltage in the x direction that is produced from

the temperature gradient in the same direction is not included in the model.

Moreover, the high operating temperatures prohibited the use of a thermal

conduction grease in the surfaces between the individual sensor compo-

nents. As a consequence, the resulting thermal blockage is reducing the

temperature gradient in the sensor and consequently its sensitivity.

In order to understand the different behavior of the sensor for SOD=3.2 cm,

we must concentrate on its three - dimensional model and on the behavior

of the air jet.

Ex =(Sxcos

2α+ Sysin2α

) ∂T∂x

+

(1

2(Sx − Sy)sin(2α)

)∂T

∂y⇐⇒

Vx =

∫ l

0

(Sxx

∂T

∂x+ Sxy

∂T

∂y

)dx (5.3)

As already explained in the model of the sensor, its voltage in the x direction

is computed from eq.5.3. It is obvious that the temperature distribution

in the x direction has an impact on the sensor output voltage. The values

of the Seebeck coefficients Sxx and Sxy in this equation are presented in

Fig.5.15(b). Furthermore, Fig.5.15(a) shows the radial distribution of the

convective heat transfer coefficient measured with the reference sensor for

SOD 2.2 cm and 3.2 cm. It is apparent that the distribution at SOD 3.2 cm

causes an almost linear temperature profile in the x direction of the sen-

sor. On the contrary, this distribution at SOD 2.2 cm results in almost no

temperature gradient in this direction. Considering that these differences

are weighted with the Seebeck coefficients Sxx and Sxy and then integrated

148


0 0.5 1 1.5 2 2.5 3 3.5 4400

450

500

550

600

650

700

750

800

Radial distance [mm]

h[W

m−2K

−1]

SOD = 2.2 mmSOD = 3.2 mm

(a) Radial distribution of the convective heat transfer coefficient for twoSOD values and mair=5Nm3/h, Tjet=300 ◦C. The value zero of theabscissa corresponds to the stagnation point.

0 50 100 150 200 250 300 350 400 450−5

0

5

10

15

20

25

T [◦C]

S[μVK

−1]

SxxSxy

(b) Seebeck coefficients of the sensor as a function of temperature. Sxx

refers to the voltage in the x direction due to temperature differences inthis direction and Sxy to temperature differences in the y direction (seeeq.5.3).

Figure 5.15: Factors influencing the sensitivity offset of the sensor ob-served in Fig.5.14 for an SOD of 3.2mm.

149


over x, we conclude that this effect is responsible for the behavior of the

output voltage.

To conclude the calibration results we present the difference of the surface

temperature between the transfer and the transverse sensor in Fig.5.16.

This difference was always below the set limit of 50 ◦C, and the assumptions

of the utilized calibration methodology hold.

100 150 200 250 300 350 4000

5

10

15

20

25

30

35

Tsens[◦C]

ΔTsu

rf[◦C]

Figure 5.16: Surface temperature difference between the transverse andthe transfer sensor, for all the calibration points.

5.3.2 Thin film resistance sensor calibration

Although the initial plan was to use the same calibration conditions for

both sensors, the thermal stresses in the substrate plate of the thin film

sensor broke it twice. Therefore, the calibration plan of Tab.5.2 was used,

so that the sensor would not be subjected to great strain.

The experiments of the current thesis were performed with sensor number

5 from the second production batch of the thin film sensors. Its resistances

150


Table 5.2: Calibration plan for the thin film sensor. For each SOD, all thejet temperatures are examined, and for each temperature all the air massflow rates.

SOD [ cm] Tair [ ◦C] mair [Nm3/h]

3.2 400, 500456

4.2 400, 500, 600

4567

4.7 500, 600567

had a thickness of 200 nm and their values are presented in Tab.5.3. The

offset voltage of the sensor was 0.264348V for an excitation voltage of

10V. Besides, the calibration of the resistance that measured its surface

temperature (RT) resulted in the linear temperature dependence of eq.5.11,

where the resistance is in Ω and the temperature in ◦C.

Table 5.3: Resistance values of the thin film sensor at 20 ◦C.

Name R1 R2 R3 R4 RTValue ( kΩ) 2.307 2.438 2.328 2.449 2.451

RT = 10.531 ·T + 2201.6 (5.11)

The calibration results are presented in Fig.5.17, together with the equiv-

alent thermal resistance model of the sensor. Sensitivity values between

140 - 260 �Vcm−2W−1 were measured in a surface temperature range be-

tween 90-150 ◦C. The heat fluxes varied between 250-500 kWm−2, and

151


5 5.5 6 6.5 7 7.5 8 8.5 9 9.520

25

30

35

40

45

50

Vout[mV ]

q[W

cm−2]

SOD=3.2SOD=4.2SOD=4.7Model

(a) Output voltage as a function of the incident heat flux.

70 80 90 100 110 120 130 140 150 160 170140

160

180

200

220

240

260

Tsurf [◦C]

S[μVcm

2W

−1]

SOD=3.2SOD=4.2SOD=4.7Model

(b) Sensitivity as a function of the sensor surface temperature.

Figure 5.17: Calibration and modeling results for the thin film resistancesensor.

152


the respective output voltage was between 5-9mV. The data presented

in Fig.5.17(b) and the results of the sensor model show that its sensitivity

drops with increasing surface temperature. Moreover, no dependence of the

sensitivity of the sensor from the SOD of the jet was observed. Although

the calibration temperatures were lower than the initial calibration goal,

the results give strong indications that the sensor is working as expected

and potential adaptations could allow its implementation in the high pres-

sure plant.

RAl2O3

RAl2O3 RSiO2

Rsubstr

Rsubstr

Rkovar

Rkovar

Heat flux - h1

Alumina substrate

KOVAR plate

Cooling water flow - h21/h2

1/h21/h1

1/h1

Pt - Ts2Al2O3SiO2

Figure 5.18: Equivalent thermal resistance model of the thin film sensor.

q =Tf − Tcw

1h1

+ 1h2

+RAl2O3 +Rsubstr +RKOVAR

(5.12)

Ts1 = Tf − q · ( 1

h1+RAl2O3@Ts2) (5.13)

Ts2 = Tf − q · ( 1

h1+RAl2O3@Ts2 +RSiO2@Tmean) (5.14)

A graphical representation of the equivalent thermal resistance model can

be seen in Fig.5.18. The model results that are depicted in Fig.5.17(b) were

computed with the experimentally measured heat transfer coefficient. It

assumes one - dimensional heat transfer and computes the temperatures in

the various points of the sensor. The values of the parameters of the cooling

water flow, h2 and Tcw, were adjusted until Ts2 was as close as possible to its

experimental value. The temperatures of the uncovered (Ts2) and covered

153


(Ts1) resistances were calculated from eq.5.14 and eq.5.13 the sensor output

voltage was computed from eq.5.6. In all the computations, temperature

dependent material properties were considered (see section 5.1.4) and the

temperatures calculated from the model were used in their computation.

In conclusion Fig.5.19 presents the difference of the surface temperatures of

the two sensors during the calibration. Again, this temperature difference

was between 0-20 ◦C and the constant heat transfer coefficient calibration

is valid.

90 100 110 120 130 140 1500

5

10

15

20

25

Tsens[◦C]

ΔTsu

rf[◦C]

Figure 5.19: Surface temperature difference between the thin film and thetransfer sensor.

154

5.4 Drilling tool design - Flame impingement temperatures

5.4 Drilling tool design - Flame impingement

temperatures

5.4.1 Scientific and technical problem

The operation of hydrothermal flame-jets as free jets in a liquid water

bath is characterized by the strong water entrainment. This phenomenon

quenches the flame-jets and affects considerably the efficiency of the tech-

nology. A hydrothermal spallation drilling tool should be able to control

but not necessarily minimize the entrainment rates. Spallation drilling re-

quires moderate surface temperatures that do not exceed the brittle-plastic

transition temperature of the rock samples. The main advantage of trans-

critical fluids is their very high convective heat transfer coefficients near

their pseudo-critical point. Hence, the combination of moderate impinge-

ment temperatures with high heat transfer coefficients seems to be the

optimal way to operate a hydrothermal spallation drilling tool.

Water entrainment

Oblique impingement

Flameimpingement

Figure 5.20: Illustration of the transfer phenomena during the injectionof a hydrothermal flame-jet in a water bath.

In order to achieve these goals, expertise from three different scientific fields

is required. First of all, the influence of water entrainment on the flame

impingement and its limits must be studied. Secondly, underwater welding

is a technology that is facing similar challenges as in our case. A water

curtain is used to protect the impingement region of gas jets and to provide

155


a suitable environment for the welding procedure. Yet, the impingement

of this water curtain will still cause an inward flow towards the stagnation

point of the flame. Studies on the oblique impingement of submerged water

jets can therefore help to reduce this inward flow and are necessary for the

drilling tool design. Fig.5.20 summarizes the three scientific components of

the drilling tool design process.

The general technical goals of the drilling tool design can be thus summed

up as follows:

• A way to control water entrainment in the flame must be found.

• The drilling tool has to be able to cause spallation for stand-off dis-

tances of at least 20-30mm.

• The rock area affected from the flame impingement must be con-

trolled, so that only a small part of it is rapidly heated-up.

5.4.2 Literature review and design considerations

The length of flame-jets and their efficiency as heat transfer devices depend

strongly on the environment in which they are injected. Usually flames are

injected in air that has a lower temperature and a higher density. This

density and momentum differences produce a turbulent shear layer between

the jet and the surrounding air. The resulting turbulent mixing causes

a flow of air into the jet, the momentum and temperature of which are

gradually reduced. Ricou and Spalding [40] quantified the entrainment

mass flow rate in a jet for the simple case of isothermal jets injected in a

fluid with an arbitrary density. They directly measured the entrainment

flow rate through a porous wall and developed a simple correlation for it

(eq.5.15). In eq.5.15, m0 is the initial jet mass flow rate, ρ0 and ρ∞ are

the densities of the jet and the surrounding air respectively and d0 is the

jet exit diameter.

m

m0= 0.32 · x

d0

(ρ∞ρ0

) 12

(5.15)

156


Entrainment studies for similar configurations, like the one of the current

thesis, were performed by Weimer et al. [135] and Kerney et al. [136] for

steam jets injected in a cold water bath. They both developed correlations

for the jet penetration length, starting from simple condensation models.

Kerney et al. [136] considered the steam jet as a cavity with clearly defined

boundaries, and analyzed the process as a mass and heat transfer prob-

lem. They assumed that condensation takes place at the interface, while

the released heat is transferred in the water bath only by convection. Af-

ter implementing the entrainment theory of Ricou and Spalding [40] they

came up with a semi-empirical correlation of the jet penetration length.

Despite the shortcomings of their theory, they identified the most impor-

tant parameters of the process. Weimer et al. [135] argued that no such

clearly defined boundary can exist and modeled the jet as a combination of

a steam expansion jet and a two phase jet. They defined a driving potential

for the condensation process and developed the semi-empirical correlation,

presented in eq.5.16. The correlation connects the penetration length of

the jet (L) with its exit diameter (d0), the density ratio of the fluids (ρ∞ρs

),

the mass flow velocity ratio (G0

Gs) and the condensation driving potential

B.

L

d0= 15.45 ·

(G0

Gs

) 12

(1 + 2.25 ·B)(ρ∞ρs

) 12 ·B

(5.16)

B =hf − h∞hs − hf

Although these two studies form the basis for the understanding of direct

steam condensation jets, the high pressure of the hydrothermal spallation

drilling process complicates the problem. In particular, the transition from

a supercritical fluid to a sub-critical liquid is fundamentally different from

the condensation procedure. Besides, hydrothermal flames are gas mix-

tures, so this transition takes place not only for water but also for carbon

dioxide.

157


Rothenfluh et al. [137] investigated the length of a supercritical water

jet, injected in a liquid water bath at elevated pressures. They studied

the influence of the jet temperature, its mass flow and the size of its exit

diameter on the resulting jet length. They concluded that the influence of

these parameters on the jet length is weak and that its average value was

equal to the injector diameter. These results were attributed on the one

hand to the high density differences and on the other hand to the absence

of any condensation potential and are summarized in eq.5.17.

L

d0= 7.323 · 2 +B

4 ·B ·(

ρ∞ρ0

) 12

(5.17)

B =hc − h∞h0 − hc

It becomes clear form eq.5.16-5.17, that the design goals of the hydrother-

mal spallation drilling tool can be only achieved if this condensation driving

potential, the density ratio or both are controlled.

Underwater welding and cutting of metal sheets are technical problems,

for which the entrainment of water in the working area must be minimized

or avoided. Several technologies have been developed to respond to this

need, ranging from direct welding in water (wet welding), to the total

extraction of water from the working surface by means of a suction cavity

[138]. A variation of dry underwater welding is the so-called dry cavity

formation technology. This method uses a combination of a high velocity

water curtain and a central gas jet to extract the water from the working

surface.

Tamura et al. [139] presented a model for the cavity formation, the details

of which are shown in Fig.5.21. It initially considers only the water curtain

operation without the inner gas jet. The impingement of the water curtain

produces a pressure difference between the inner (A) and the outer space

(B), which is calculated with basic non-compressible fluid dynamics. A

stable cavity - for the operation with both jets - can be formed only when

158


Dd

AB

JA (QA) JB

V0,J0

QBdθ

GasWater Water

Figure 5.21: Illustration of the dry cavity formation model with the useof a water curtain nozzle. Adapted from [139].

the outer pressure is higher than the inner one for the operation without the

inner gas jet. The additional presence of this jet raises the inner pressure

and exhausts its fluids through the water curtain to the outer space.

Matsumoto et al. [140] implemented the same cavity formation methodol-

ogy for underwater metal sheet cutting. They investigated the influence of

the stand-off distance, the water curtain angle and the water and gas flow

rates on the quality of the cavity. They concluded that 61◦ is the optimal

oblique impingement angle of the water curtain (θ in Fig.5.21) and that

higher inner jet velocities lead to stable cavities for higher stand-off dis-

tances. Zhang et al.[141] had a similar setup but for a laser beam welding

application. Their work resulted in the same optimal water curtain angle

and extended that of Matsumoto et al. by dividing the stable cavities in

two subcategories depending on the amount of entrained water.

As mentioned, the oblique impingement of the water curtain could result

in water flowing towards the impingement region of the flame. The water

curtain can be modeled as a submerged two - dimensional slot jet, imping-

ing obliquely on a flat surface. Beltaos [142] developed an analytical tool

159


for the prediction of the pressure and shear stress fields produced on an

impingement plate by the oblique impingement of plane turbulent jets. His

results have shown a shift of the stagnation point by a distance s from the

intersection of the jet axis with the impingement plate. Finally, he argued

that the ratio of momentum flux in the positive (index B) and the negative

(index A) directions is a function only of the impingement angle (eq.5.18).

MB

MA=

1+ cosθ

1− cosθ(5.18)

Chin and Agarwal [143] studied the mass transfer coefficient of similar jets

and also reached similar conclusions on the eccentricity of the stagnation

point. Moreover, Akansu et al. [144] studied the heat transfer and flow

field of an oblique impinging air jet. Their results show that the stand-off

distance plays a dominant role in defining the stagnation point position,

by modifying the impingement angle. This publication however fails to

attribute this effect to the pronounced jet entrainment before its impinge-

ment.

5.4.3 Design of the combustion chamber nozzle

During the ignition experiments, the combustion chamber consisted only

of two concentric tubes and the cooling water was injected in the parallel

direction to the flame-jet. The outlet diameter of the chamber was almost

16mm leading to very low flame-jet velocities. The major goal of the new

design was to control water entrainment by reducing the density of the fluid

in the vicinity of the flame-jet.

Fig.5.22 presents the conceptual design of the new nozzle and its geomet-

rical characteristics. The cooling water is diverted from the flame-jet, by

injecting it to an angle in respect to the axis of the chamber. Further-

more, the exit diameters of the central and the annular jets control the jet

velocities and their ratio.

160


d1 φd2 d3

Figure 5.22: Conceptual design of the drilling tool. d1 : flame-jet outletdiameter, d2&d3 : cooling water injection annulus diameters, φ : watercurtain angle.

On the other hand, the water injection angle affects the volume confined

between the water curtain and the impinging plate, which determines the

heat needed to lower the density of the water around the flame. The larger

this volume is, the more energy is necessary for the density reduction of the

surrounding fluid and the less efficient the process becomes. Conversely, a

small curtain angle will decrease this volume but it will increase the inward

flow of water due to oblique impingement. This effect may quench the

flame exactly on the point of its impingement. Obviously a design must be

found that balances these two counteracting effects.

Based on these background considerations and on the data presented in

section 5.4.2, six nozzles with three jet exit diameters and four water curtain

angles were built. In all cases the annulus diameters were kept constant

(d2 = 23mm and d3 = 24mm) due to limitations imposed from the pressure

vessel geometry. Tab.5.4 presents the combinations of these parameters.

Table 5.4: Nozzle design parameters (see also Fig.5.22).

Nozzle Nr. 1 2 3 4 5 6Flame-jet diameter: d1(mm) 10 5 7.5 7.5 7.5 7.5Water curtain angle: φ(◦) 30 30 30 60 90 120

161


5.4.4 Experimental setup

The measurements focused on the two - dimensional temperature profiles

that resulted from the impingement of hydrothermal flames on a flat, metal

plate. Their goals can be summarized in the following list:

• The maximum impingement temperature should be measured at var-

ious stand-off distances, to approximate the flame length at each op-

erational point.

• The symmetry of the impingement temperature profiles must be mon-

itored.

• The measurements must examine the size of the surface area on the

plate, were temperatures relevant for spallation drilling experiments

are produced.

A matrix of ten K-type thermocouples with a diameter of 1mm, was welded

on the surface of a stainless steel plate and their arrangement is presented

in Fig.5.23. In order to focus the measurements to an area relevant for

spallation drilling the equivalent diameter of the thermocouple matrix was

chosen equal to 15mm.

Even though the application of the impingement plate seems quite simple,

several major problems were encountered during the installation of the first

plate. Four of the ten thermocouples were burned during laser welding

and the experiments with nozzles 1 to 3 were carried out without them.

The sealing of the thermocouples was also challenging and a whole set

was destroyed in the first attempt. Slight adaptations in the assembly

procedure and the sealing materials solved the problem and eventually

made the measurements with ten thermocouples possible.

162


2.5

1.05

Thermocouple positions

Thermocouple holes

Holes geometry details

90

120°

Figure 5.23: Thermocouples array for the measurement of the two - di-mensional impingement temperature profile of hydrothermal flames.


The effect of six parameters on the impingement temperatures was analyzed

during the experimental investigation. These parameters can be divided

in two groups, referring to the geometry of the problem and the plant

operation. The former group included the jet diameter, the water curtain

angle and the stand-off distance (SOD) between the plate and the nozzle.

The operational parameters were the combustion chamber load, the cooling

water flow rate and the fuel stream composition. The combustion chamber

load is defined as the consumed fuel power divided by the chamber volume

and it is a function of the fuel flow rate and its ethanol concentration. In a

similar way as in the ignition experiments, the chamber load was changed

by varying the fuel flow rate and keeping its composition constant. Hence,

the changes of the chamber load affected only the flame-jet exit velocity

and not its outlet temperature, which is primarily influenced from the fuel

composition (see section 4.7).

The impact of the central jet velocity and the water curtain angle on the

impingement temperatures was examined with nozzles 1-3 and 3-6 respec-

tively. For each nozzle and SOD, two values of the fuel composition and

163


Table 5.5: Measurement plan for the flame impingement experiments.

Cf SOD mf mo2 mcw1

[% wt. EtOH] [mm] [ kg/h] [NL/min] [ kg/h]

12.5 40,30,25,20,1520 74

350475

30 111350475

20 40,30,25,20,15

20 119350475

30 177350475

40 236350475

four values of its flow rate were used, resulting in five values of the cham-

ber load. Five SOD values and two values of the cooling water flow rate

were applied for each flame operational point and the detailed measurement

points are presented in Tab.5.5.

At the beginning of each experiment the plate was positioned at an SOD

= 250mm, to avoid any interaction with the ignition of the flame. After

ignition, the defined flame operation point and the corresponding mini-

mum cooling water flow rate were set. The plate was slowly brought at the

highest SOD and the system was left for 120 s to reach steady state. The

temperatures on the plate were recorded for 120 s and once the measure-

ments for both cooling water flow rates were performed, the next SOD was

chosen.

Tab.5.6 shows the calculated velocity and Reynolds numbers for each flame

operating point and each jet diameter. For these calculations it was as-

sumed that the jet consisted of the products of a complete combustion.

The resulting jets were modeled as mixtures of perfect gases, while the

properties of each gas were calculated after [145] at the flame temperatures

presented in section 4.7.

164


Table 5.6: Jet parameters at the outlet of the combustion chamber nozzle.The material properties were calculated after [145] at the flame tempera-tures presented in section 4.7.

d1Fuel stream

Jet velocity Jet ReynoldsCf mf

[ mm] [% wt. EtOH] [ kg/h] [m/s] [-]

10

12.520 1.86 1871130 2.80 28067

2020 3.62 1735730 5.44 2603540 7.25 34713

7.5

12.520 3.31 2494830 4.97 37422

2020 6.44 2314230 9.67 3471340 12.89 46284

5

12.520 7.46 3742230 11.19 56133

2020 14.50 3471330 21.75 5207040 29.00 69426

5.4.6 Experimental results and impingement

temperature profiles

The impingement temperatures showed the expected behavior as a function

of the SOD.For higher SOD values, lower temperatures were observed and

the jets were almost fully mixed with the surrounding water. As the SOD

decreased, higher temperatures and larger differences between the thermo-

couples were measured. With few exceptions, an increase of the cooling

water flow rate resulted in a decline of the impingement temperatures. On

the contrary, higher water curtain angles increased the temperatures on the

plate. Finally the impingement temperatures were a non monotone func-

tion of the jet velocity. The temperatures increased for rising jet velocities

165


10 15 20 25 30 35 40 4550

100

150

200

250

300

350

SOD [mm]

T[◦C]

(a) Temperature profiles for nozzle 3, with Cf=20%, P=45 kW andmcw=350 kg/h.

10 15 20 25 30 35 40 4550

100

150

200

250

300

350

SOD [mm]

T[◦C]

(b) Temperature profiles for nozzle 3, with Cf=20%, P=29 kW andmcw=475 kg/h.

Figure 5.24: Influence of the SOD on the temperature profile of two flameswith a different combustion chamber load. The positions of the thermo-couples can be seen in the upper right part of the diagrams.

166


up to a different value for each nozzle, where they leveled off. A further

rise of the jet velocity revealed a plateau of the impingement temperatures,

which was followed by a further increase for still higher combustion cham-

ber loads. As the performed experiments were limited to the comparison

of only two values of the fuel concentration, its influence was observed

for only one operational point. As a consequence, the weak dependence

of the impingement temperatures on the fuel concentration could not be

generalized.

The impact of the SOD on the impingement temperatures for two chamber

loads is presented in Fig.5.24. Both sub-figures show that water entrain-

ment was strong enough, at SOD values between 25-40mm, to fully mix

the flame-jets with the surrounding water. A uniform temperature distri-

bution resulted on the plate, which had an average value proportional to

the chamber load. By contrast, higher temperatures were observed around

certain thermocouples, when the jet had a high momentum and energy

content and the SOD values were between 15-20mm. Fig.5.24(a) presents

an experiment where the high-energy central jet raised the temperatures

on the plate up to 320 ◦C. Conversely, the thermal content of the flame-jet

in Fig.5.24(b) was not enough to produce any changes of the temperatures

on the plate, even at an SOD of 15mm. The entrainment was so intense

that the flame-jet was fully quenched directly after its injection.

The second investigated geometrical parameter was the jet exit diameter,

which influences the flame-jet exit velocity. Fig.5.25 presents the tempera-

ture profiles for nozzles 1-3, which had a constant water curtain angle and

three values of the jet exit diameter. The colors in this figure correspond

to the three nozzles, while the flame-jet characteristics for each of the exit

velocities can be found in Tab.5.6.

The values presented in Fig.5.25(a) confirm the observation that the jet was

fully mixed with the surrounding water at high stand-off distances. Again,

the impingement temperatures showed a linear increase with increasing

chamber load, while no differences were observed between the individual

thermocouples at a certain operational point. When the SOD was equal to

167


0 5 10 15 20 25 3050

100

150

200

250

300

350

Vjet[ms−1]

T[◦C]

(a) Temperature profiles for SOD = 40mm. The two first points foreach nozzle correspond to the mixtures Cf = 12.5%wt.

0 5 10 15 20 25 3050

100

150

200

250

300

350

Vjet[ms−1]

T[◦C]

(b) Temperature profiles for SOD=15mm. The two first points for eachnozzle correspond to the mixtures Cf = 12.5%wt. The second and thethird velocity values for each nozzle have equivalent chamber loads.

Figure 5.25: Influence of the jet exit velocity and the combustion cham-ber load on the temperature profiles of nozzles 1-black, 2-red, 3-blue.Cf=12.5% or 20%, P=18-60 kW, mcw=350 kg/h.

168


15mm (see also Fig.5.25(b)), nozzles 2 and 3 caused an initial increase of the

plate temperatures with increasing jet exit velocity. This trend continued

up to flame-jet velocities of approximately 10m/s for nozzle 3 and 15m/s

for nozzle 2. For higher velocities the temperatures leveled off and a slight

increase was observed for nozzle 2 at very high velocities.

Two counter-acting phenomena, caused from the increase of the chamber

load, are responsible for this effect. In fact, an increase of the chamber

load raises the energy content of the jet, but it is simultaneously increas-

ing its velocity. The higher jet velocity intensifies its mixing shear layer

with the surrounding water and it increases the cooling intensity of water

entrainment. Nonetheless, these competing effects are not strengthened at

the same extent. For very low chamber loads the jet energy content is low

and the water entrainment quenches the jet to a very low temperature. For

low jet velocities, its higher energy content compensates for the rise of the

entrainment intensity and the temperatures on the plate increase. The fact

that the temperatures level off for some velocities is an indication that the

two competing mechanisms are balanced. At the same time, the further

increase of the impingement temperatures caused from nozzle 2 at very

high jet velocities, leads us to believe that a saturation of the entrainment

phenomenon takes place in this velocity region.

The water curtain angle was the most interesting geometrical parameter,

due to its influence on the entrainment intensity. Fig.5.26 presents a com-

parison of the temperature profiles of nozzles 3 and 4 at an SOD of 15mm.

The larger water curtain angle of nozzle 4 (Fig.5.26(b)) diverted the water

from the flame-jet and reduced the cooling intensity of the entrainment

effect. Higher impingement temperatures were measured and the increase

of the jet velocity led to their saturation around 350 ◦C. The same effect

was observed on the profiles of both nozzles 5 and 6, which had higher

water curtain angles. An important observation concerns the temperature

at which the saturation of the impingement profiles was observed. Fig.5.27

presents the specific isobaric heat capacity of a gas mixture that results

from the complete combustion of a fuel mixture with Cf = 12.5% wt. Its

169


5 6 7 8 9 10 11 12 13 14 1550

100

150

200

250

300

350

Vjet[ms−1]

T[◦C]

(a) Temperature profiles for nozzle 3 - φ = 30◦.

5 6 7 8 9 10 11 12 13 14 1550

100

150

200

250

300

350

Vjet[ms−1]

T[◦C]

(b) Temperature profiles for nozzle 4 - φ = 60◦.

Figure 5.26: Influence of the water curtain angle on the temperatureprofiles for SOD=15mm, Cf=20%, P=30, 45, 60 kW, mcw=350 kg/h.

170


200 250 300 350 400 450 5000

2

4

6

8

10

12

14

16

18

T[◦C]

c p[kJkg−1K

−1]

Figure 5.27: Specific isobaric heat capacity of the combustion productsof a fuel stream with Cf=12.5%. The calculation of the mixture propertieswas performed with REFPROP [146], with its built in mixing rules andmaterial properties.

pseudo-critical temperature is estimated to be in the region where the sat-

uration of the temperature profiles was observed. The results in Fig.5.26

suggest that water entrainment was sufficient, already by a jet velocity

of 10m/s, to cool the jet from its exit temperature to its pseudo-critical

temperature. Once the jet enters this temperature region, higher cooling

capacities are necessary for its further cooling, due to its high cp. The rise

of water entrainment through the higher jet velocity is not sufficient to

change the jet temperature and the impingement profile levels off around

the pseudo-critical temperature of the jet. As the data does not give any in-

dication about the system behavior for higher chamber loads, two possible

scenarios can be identified.

1. The entrainment effect has approximated its maximum cooling in-

tensity, due to the increased water curtain angle. As a result, higher

chamber loads will cause a disproportionate increase of the flame-jet

thermal content in comparison with the augmentation caused to en-

171


trainment. For a range of flame-jet velocities, this effect will not be

translated to higher surface temperatures, due to the high cp values

of the flame-jet. Once this energy input overcomes the pseudo-critical

region of the flame-jet, the impingement temperatures will rise again.

2. A balance has developed between the aforementioned counter-acting

effects. In this case, a further increase of the chamber load will not

cause any changes to the temperature profile. The maximum impinge-

ment temperature will be equal to the pseudo-critical temperature of

the flame-jet, and a new solution must be found to control water

entrainment.

4 5 6 7 8 9 10 1150

100

150

200

250

300

350

400

450

Vjet[ms−1]

T[◦C]

Figure 5.28: Impingement experiments with nozzle 5 but with d1 = 10mmand φ = 90◦. The black marks are the experiments with Cf = 15%wt.,and the red with Cf = 17.5%wt. SOD = 15mm, P = 39-78 kW.

After finalizing this experimental series, the jet diameter of nozzle 5 was

increased to 10mm and further tests were conducted to explore these two

scenarios. Chamber loads up to 78 kW were applied that led to maximum

impingement temperatures of 420 ◦C. The results that are presented in

Fig.5.28 support the first scenario, since high chamber loads resulted in a

further increase of the impingement temperatures.

172


10 15 20 25 30 35 40 4550

100

150

200

250

300

350

SOD [mm]

T[◦C]

(a) Impingement temperatures for mcw=350 kg/h and all the investi-gated SOD values.

10 15 20 25 30 35 40 4550

100

150

200

250

300

350

SOD [mm]

T[◦C]

(b) Impingement temperatures for mcw=475 kg/h and all the investi-gated SOD values.

Figure 5.29: Temperature profiles for nozzle 3, with Cf=20%, P=60 kWand for two cooling water flow rates.

173


Finally, an increase of the cooling water mass flow led in most cases to

an increased entrainment and a strong quenching of the flame-jets. In

contrast to this general trend, some experiments demonstrated a marginal

increase of the impingement temperatures for higher cooling water flow

rates. Fig.5.29 presents the temperature profiles of nozzle 3 at 60 kW for

two cooling water flow rates. The increase of the cooling water flow rate

resulted in lower impingement temperatures for all the SOD values except

from that of 15mm, where the higher cooling water flow raised the imping-

ing temperatures. This marginal increase must be attributed to similar

phenomena with these reported during underwater welding. The stronger

water jet is producing an overpressure in the outer volume of the water

curtain. The combination of this overpressure with the strong inner jet led

to higher impinging temperatures.

5.4.7 Conclusions of the impingement experiments

The flame impingement experiments resulted in 400 temperature profiles,

which gave a clear picture of the physical behavior of the system. The

results have shown that an increase of the water curtain angle reduced the

entrainment of water in the flame-jet to a certain extent. As the experi-

ments with a water curtain angle (φ in 5.22) of 120◦ caused major pressure

fluctuations in the pressure vessel a value of φ = 90◦ was chosen for the

future drilling tools.

Another interesting parameter was the degree of local concentration of

high temperatures on the impinging plate. The optimal jet operation led

to profiles concentrated in a very small region. The imprint of the mea-

sured temperatures provides a strategy for the operation of the combustion

chamber, if approximately 10% of the total rock surface should be affected

from the flame during drilling.

Finally the connections between the jet velocity, water entrainment, the

energy content of the jet and the geometrical parameters of the nozzle

design were clarified. The future experiments should concentrate on the

174


operation of nozzles with jet exit diameters around 10mm and combustion

chamber loads approximately 60-90 kW.

175

Chapter 6

Thesis conclusions

The previous chapters gave a detailed account of the first attempt to use

hydrothermal flames as heat transfer and drilling tools. Many of the chal-

lenges of the technology, presented in Fig.6.1, have been addressed and

overcome. The forced ignition of hydrothermal flames has been technically

solved and scientifically studied. The implemented combustion chamber

accommodated high thermal loads and its optimized design will serve as a

benchmark for the scale-up of the technology. Finally, two heat flux sensors

have been manufactured and calibrated in a custom-made calibration facil-

ity. Initial tests with one of these sensors in the high pressure vessel proved

its ability to operate at the conditions it was intended to. In the present

chapter the technical and scientific results of the thesis are summarized.

6.1 Summary of the technical results

Pilot plant construction

The central objective of the thesis was to design, built and commission the

high pressure pilot plant and the first hydrothermal spallation drilling pres-

sure vessel. After more than 500 operating hours, no accidents or major

technical problems occurred. This is attributed to the accuracy of the de-

sign decisions, described in chapter 2, and specifically to the safety analysis

177

Chapter 6 Thesis conclusions

Entrainment• High density differences••

•

•

High temperature differences

→Rapid temperature decayEntrainment control

Heat transfer coefficient• Crucial for spallation performance

• Heat flux sensors developement• Dependency on operation conditions

Rock mechanics - Drilling procedure• Hole size modeling

• Fracture model and drilling velocity predictions

• Interaction flame - rock

Combustion in an aqueous environment

Forced ignition of hydrothermal flamesTemperature and power control

Figure 6.1: Hydrothermal spallation drilling challenges.

of the plant. Furthermore, only minor adaptations of the plant equipment

were necessary, largely due to wear of the system or expectations, which

did not materialize. The materials of the oxygen tubing have been further

adapted to the oxidizing environment they are operate in. After trying

three alternatives for the pressure control valve, the optimal solution has

been found in the form of a three stem valve, which reduced the pressure

in steps.

In conclusion, the maximum fuel power in the experiments conducted so

far was 78 kW, which is close to the goal of 120 kW. At the same time, the

plant operation gave no indications so far that its nominal power capacity

could not be safely reached in the future.

178

6.1 Summary of the technical results

Ignition module construction

Another essential objective of the thesis was the invention of a forced ig-

nition module for hydrothermal flames. The realization of this module

allowed for a plant operation at lower reactants temperatures and it con-

siderably increased its operational safety.

Two types of igniters have been employed, a ceramic body made of silicon

nitride and coiled metal wire. The ceramic igniters have shown very strong

oxidation from supercritical water and their use has not been pursued fur-

ther. Nonetheless, their implementation forced us to face the challenge of

feeding an electrical resistance with 230V and a maximum electrical power

of 394W, while it was operating in a vessel at 260 bar and 420 ◦C. On the

other hand, most of the coiled wire igniters operated for at least 30 h, while

the latest so far performed approximately 100 ignitions and operated 150 h.

Heat flux sensors construction

Two innovative heat flux sensors have been developed for the measurements

in hydrothermal flames and cutting edge technologies have been applied in

their construction.

The fruitful cooperation with the department of Joining and Sensors Tech-

nology of the Paul Scherrer Institute led to the construction of a transverse

Seebeck sensor, which was calibrated up to temperatures of 420 ◦C. High

temperature (1200 ◦C) soldering technologies were used for the connection

of the metal and ceramic parts of this sensor. Initial tests in the high pres-

sure vessel and at combustion chamber loads up to 78 kW and SOD of

15 mm have proven the ability of the sensor to withstand the very harsh

operational conditions in the vicinity of a hydrothermal flame.

Ten thin film sensors were the outcome of the close cooperation with EM-

Bremmen. Their shortcomings during their high temperature calibration,

made a new design of their plate necessary. All the same, the used thin

film technologies proved viable for the intended temperatures.

179


6.2 Summary of the scientific results

Heat transfer measurements in the combustion chamber

The measurements of the convective heat transfer coefficient in the combus-

tion chamber of the plant provided valuable input for the igniters’ dimen-

sioning. These measurements were carried out with the ceramic igniters

and with ternary mixtures of water, ethanol and nitrogen.

All the mixtures demonstrated enhanced or normal convective heat trans-

fer characteristics and heat transfer coefficients up to 7000W/m2K were

measured. The pseudo-critical points of the mixtures were found between

293-330 ◦C, depending on their composition. In particular, the use of a

fuel stream with Cf = 20%wt. shifted the pseudo-critical point of the

respective water-nitrogen binary mixture from around 322 ◦C to 307 ◦C.Likewise, the pseudo-critical temperature of a water-nitrogen mixture with

18% wt. nitrogen was approximately 53 ◦C lower than that of water. The

addition of ethanol or nitrogen in the mixtures lowered the peak value of

the cp and the heat transfer coefficient.

Even though a quantitative comparison of the data with the literature was

not possible, the results were qualitatively consistent.

Forced ignition measurements

The successful realization of forced ignition in hydrothermal flames pro-

duced scientific data that sheds a new light on the ignition characteristics

of supercritical fluids.

Ignition temperatures between 450-800 ◦C and ignition powers between

60-300W have been measured. The trans-critical mixtures demonstrated a

minimum ignition temperature around their pseudo-critical point. Further-

more, an increase of the turbulence level and the velocity in the combustion

chamber had no effect on the ignition temperatures and a weak one on the

ignition power. We also observed that the ignition temperature approxi-

mated the self-ignition temperature of the combustible, when the latter was

180

6.2 Summary of the scientific results

around its pseudo-critical point. These phenomena demonstrate consider-

able differences from the results of conventional ignition studies. In order

to explain these discrepancies, we developed a qualitative ignition concept

and we argued that the phenomena should be attributed to the strongly

varying properties of the mixtures around their pseudo-critical point.

In conclusion, our results indicate that the forced ignition of hydrothermal

flames will be feasible in a field experiment, with the power that state-of-

the-art down-hole generators are able to produce.

Flame impingement measurements

The impingement experiments were an essential part of the drilling tool

design. The two - dimensional temperature profiles that resulted from the

impingement of free hydrothermal flame-jets on a flat surface were mea-

sured. Combinations of the nozzle operational and geometrical parameters

were sought, which would lead to a maximum impingement surface tem-

perature of 400 ◦C at an SOD of 20mm. In the end, this target had to be

modified, reaching 420 ◦C at an SOD of 15mm, which proved to be enough

for the commencement of spallation on rock probes.

The tests have shown an initial increase of the impingement temperatures

for increasing jet power and velocity. A region of constant temperatures

followed and a further increase of the jet power resulted to a raise of the

temperatures on the plate. This behavior was explained with classical en-

trainment rules and by considering the flame-jet properties variation around

its pseudo-critical point.

The experiments with four water curtain angles concluded that an increase

of the water curtain angle reduces the water entrainment in the jet. In

general, higher cooling water flow rates caused a decline of the impinge-

ment temperatures, with the exception of some cases where an increase

was observed. Moreover, the operation with a water curtain angle of 120◦

resulted to severe pressure fluctuations in the vessel and an angle of 90◦

was chosen as the optimal for our case.

181


The conclusion to be drawn, was that the flame-jet should operate with a

low Cf and high mf . As a result, its pseudo-critical point will be closer

to that of water and high heat transfer coefficients will be available for a

variety of operational points. The resulting flexibility will make the opti-

mization of the drilling tool more effective.

182

Chapter 7

Thesis outlook

In the course of the thesis, several ideas have been developed that could not

be further pursued, mostly due to time constraints. A summary of these

ideas that remained on the conceptual level but are worth considering in

the further steps of the project will be presented in the first section of

this chapter. The second section concentrates on the future experiments

on hydrothermal spallation drilling and presents several proposals on how

these experiments could be carried out.

7.1 Sensor project outlook

Along with the sensors that we constructed and calibrated, two innovative

heat flux measuring concepts have been developed. Both concepts could

be also useful to other projects that require similar heat flux measurements

and are presented in the following sections.

7.1.1 Heat flux measured with optical fibers

The measurement of temperature, pressure or deformation with optical

fibers is already an established method, used mainly in applications where

electrical signals are difficult or impossible to operate. In most measuring

183

Chapter 7 Thesis outlook

devices a discontinuity of the fiber refractive index is introduced at the

measurement point. The comparison of the light that is sent though the

fiber with its reflection at the measurement point provides information on

the measured quantity. An in depth analysis of interferometric fiber optic

sensors can be found in the work of Greywall [147].

Optical fiber interferometry has been already used to measure heat flux,

through the measurement of a temperature difference inside a fiber. McPher-

son et al. [148] formed a Fabry-Perot cavity between a Bragg grating and

the end of an optical fiber to measure transient heat flux. Furthermore,

Shen et al. [149] tried to measure steady state heat flux by placing two

Bragg gratings in the same fiber at a distance of 85mm from each other.

Their experiments showed that considerably smaller gratings and lower

distances between them are needed for this method to be viable.

T vacuum

n(T1)*ds1(T1)

ds2

ds1

amb

T1

n(T2)*ds2(T2) T2

Figure 7.1: Principle of operation of the modified Oxsensis sensor, for themeasurement of heat flux. A comparison of the reflections at the boundariesof the cavity gives the average temperature in each layer (ds1 and ds2).

Oxsensis is an English company that is producing a pressure and temper-

ature fiber optic sensor. The sensor is working with a sapphire optical

fiber, on the end of which a Fabry-Perot cavity is spliced. For purposes

of the project, Oxsensis has been asked to support a feasibility analysis of

the use of their sensor in heat flux measurements. A pressure sensor has

been modified to measure the average fiber temperature before and after

184


its cavity, and the principle of its operation is illustrated in Fig.7.1.

The new sensor concept was tested in the calibration plant and a linear

increase of the temperature difference with increasing incident heat flux

was observed. However, the value of this difference was comparable with

the error of the individual temperature measurements and no reliable heat

flux measurement could be carried out. In view of the proven ability of the

sensor to operate in very harsh environments [150], the measuring principle

could still work for the intended conditions, if some requirements would be

fulfilled:

• The effect of accumulation of dust or soot on the sensor should be

technically minimized.

• The error of the individual temperature measurements must be lower

than their difference, by at least one order of magnitude.

• Direct cooling of the cavity and structural changes in the sensor hous-

ing could increase the aforementioned temperature difference and the

accuracy of the heat flux measurement.

• Finally, the cross correlation of temperature and pressure must be

eliminated or calibrated, so that the sensor could be used in a high

pressure environment.

7.1.2 Heat flux sensors made of ceramic

thermoelectric oxides

Ceramic thermoelectric oxides are artificial materials that have an ex-

tremely high Seebeck coefficient and are able to operate at very high tem-

peratures. The group of Professor Weidenkaff in the EMPA - Dubendorf

- has a long experience with these materials, especially with a focus on

renewable energy production [151, 152]. A brief project was initiated in

cooperation with this research group, in order to design and construct a

novel sensor that would take advantage of these materials. Fig.7.2 presents

185


the concept of a sensor based on the synthetic thermoelectric material

La1.98Sr0.02CuO4, which has an average Seebeck coefficient of 200 �V/◦C.

Q

d1

2

KOVAR

La Sr CuO 1.98 0.02 4

Figure 7.2: Principle of operation of the thermoelectric oxide sensor. Thetemperature difference between the two lead wires is producing the sensingsignal. The small size of the sensor allows the direct connection of thissignal to the incident heat flux.

The material, which is produced in the form of a powder, is prepared as

a slurry and cast in a groove on the surface of a ceramic substrate. Two

wires made of KOVAR, are fixed in two lateral positions in the mass of the

thermoelectric material at a given distance from each other. The whole

assembly should be sintered at approximately 1000 ◦C and the resulting

body would have the two lead wires embedded in it. When heat flows

through the body the temperature difference between the two connection

points of the wires will produce a thermoelectric voltage. If this distance is

kept low enough to minimize three - dimensional heat transfer effects, this

voltage will be proportional to the incident heat flux. The resulting sensor

should be able to operate at temperatures reaching 1000 ◦C, depending on

its ceramic substrate.

186


7.1.3 Optimization of the existing sensors

Initial tests in the high pressure vessel have proven that the transverse

sensor is be able to perform the necessary heat transfer measurements.

Nevertheless, the cooling system of the sensor can be improved by reducing

the thermal resistance between its cooling water and its upper surface.

This change can be achieved by resizing the individual sensor parts and

by applying a thermal conductive paste on their contact surfaces. Apart

from the structural adjustments, the addition of a second thermocouple

to measure the average temperature at another point would be a further

improvement. The two thermocouples will provide data on the temperature

distribution in the x direction and a juxtaposition of the sensor operational

conditions and its calibration will more reliable.

Figure 7.3: Optimized design of the thin film sensor. The substrate platehas now a diameter of 20mm, and the resistance temperature measurementis realized as a two point measurement.

The thin film sensor fulfilled its design specifications only partially. The

underestimation of the thermal stresses on its substrate plate led to its

breakage during calibration. This problem can be solved by reducing the

size of the plate and bringing the contact screws closer to the measurement

point. However, this would increase the flow disruption caused by the

187


sensor, which must in turn be taken into account during calibration. Fig.7.3

shows a possible design of the new ceramic plate of the sensor. This plate

could be vacuum soldered on a thin metallic plate made of KOVAR, which

could in turn be flush mounted on an impingement plate.

7.2 Drilling tool design outlook

In addition to the operational optimization of the existing design of the

drilling tool, several ideas have been developed for a future design. The

experiments presented in this thesis, clarified the role of the jet exit diam-

eter, thermal power and velocity, and also the role of the injection angle

for the cooling water.

In the future designs, the injection strategy of the cooling water (CW1)

should be made more flexible. This cooling water is essential for the main

vessel chamber, but its entire flow rate does not necessarily have to reach

the lowest point of the nozzle. The simplest way to divert a proportion of

it to the main vessel space would be to drill holes in the outer chamber

tube. As a result, the cooling intensity of the main vessel space would not

change, but the temperatures around the nozzle exit would be higher.

However, this option does not provide a way to control the amount of water

that is injected at each point of the vessel. This control could be achieved

by making use of the two inlet ports of CW1 in the vessel head (presented

in Fig.A.3). The combustion chamber cooling system could separate these

ports inside the vessel, and a valve could control the flow rate of water

in them. The water reaching one port could be led directly to the lowest

point of the nozzle, while the other port would feed the peripheral injection

points. This design would also facilitate the future spallation drilling and

heat transfer measurements in concave cavities. During these experiments,

it will be necessary to divert a portion of the water from the cavity, in order

to optimize the heat transfer conditions in it.

188


A more sophisticated design of the combustion chamber is presented in

Fig.7.4. Here, the combustion products are separated in two jets, a central

and an annular one. The central jet is the same as in the former case, while

the annular one could have the same cone angle as the water injection or

even a larger one. The annular jet functions as a first entrainment frontier

and reduces the direct and very intense flow of cold water in the central jet.

A combination of this nozzle with the controlled diversion of water from

the flame injection point could provide all the necessary tools to adapt the

impingement heat transfer conditions for spallation of various rock types.

d d4 5 φ1 φ2

Figure 7.4: Optimized design the combustion chamber outlet. Two flameinjection ports are integrated to reduce the direct entrainment of water inthe central jet.

7.2.1 Initial drilling experiments

After assessing the final impingement results we were ready to proceed with

the first drilling experiments, which were conducted with Iragna granite

samples with a diameter of 83mm. While the operational conditions re-

mained the same (see Tab.7.1) for three distinct experiments, the exposure

time of the rock sample was varied.

Each time, the flame was ignited with a much lower chamber load, which

was then slowly increased to the value of Tab.7.1. The SOD between the

rock sample and the chamber nozzle was 100mm at the beginning and it

189


Table 7.1: Nozzle and operational parameters during the first drillingexperiments.

Rock mf Cf mcw SOD d1 φ Time[ kg/h] [%wt.] [ kg/h] [mm] [mm] [grad] [ min]

Iragna 1 60 17.5 350 15 10 90 1Iragna 2 60 17.5 350 15 10 90 10Iragna 3 60 17.5 350 15 10 90 30

Figure 7.5: Rock samples after the first three drilling experiments. Theexposure time is from left to right 1, 10, 30min and the small cavities arevisible on the two samples on the right side.

was rapidly reduced to 15mm just after ignition. The rock samples after

the three experiments are presented on Fig.7.5.

All the experiments demonstrated successful spallation of the rock samples

and the produced spalls were in the form of sand-like particles. As a matter

of fact, the enlargement of the cavity between experiments 1 and 2 has

shown that the process was continuous. However, its negligible enlargement

between experiments 2 and 3 indicated that spallation has most probably

stopped after a certain time (most probably between 1 -10min) due to a

heating-up of the rock probe.

The initial drilling experiments demonstrated the feasibility of spallation

drilling with hydrothermal flames. It will be the goal of a next research

190


project to optimize the design of the combustion chamber in order to

achieve higher penetration depths and rates. Even these preliminary ex-

periments have clearly proven that the temperature and heat flux values

provided by free hydrothermal flame-jets can induce spallation in granite

samples.

7.2.2 Future impingement heat transfer experiments

The impingement heat transfer and the drilling experiments on flat sur-

faces are only the starting point of the investigation. In a field experiment

a hole will be already present when the spallation drill bit will start its

operation. If hydrothermal spallation drilling is to be further pursued and

implemented on an industrial scale, the conditions in the concave surface

must be measured and adapted. Thus, once the experiments on the flat

surfaces will be finalized and the spallation conditions will be defined for

a certain rock type, the design loop (Fig.5.1) of the drilling tool should

focus on concave surfaces. As Rauenzahn [19] showed in his work, the flow

conditions in the concave version of the problem are considerably different.

Accordingly, the positions of the thermocouples for the measurement of the

new impingement profiles must be adequate and the heat flux sensors must

be adapted to the concave version of the problem.

Apart from the experiments in the concave geometry, a concept for the

interpretation of the heat flux measurements is needed. It is unlikely that

the sensors will have the same heat conductivity with the rock samples,

which will also be different for each rock type. As a result, the surface

temperature of the sensors will not be the same as that of the rock samples,

even though the same flame-jets will be applied. Moreover, we know that

the convective heat transfer coefficient of supercritical fluids is a strong

function of their temperature. This difference of the surface temperatures

will lead to different fluid temperatures and heat transfer coefficients. Thus,

a measurement with a sensor will be useless unless a method can be found

to interpret it for the conditions prevailing in the rock experiments. As a

191


detailed analysis of the methods to interpret this data exceeds the scope of

the thesis, only some initial thoughts are briefly presented:

• The first methodology takes advantage of the Biot number to esti-

mate the surface temperature on the rock. It relies on the sensor

measurements and the ratio of the two heat conductivities.

• The second method makes use of a simulation of the temperature

distribution in the rock samples.

– In a first step the heat transfer coefficient is measured with the

heat flux sensor as a function of the surface temperature.

– Then, it is assumed that the rock and the sensor develop the

same surface temperature during the experiments. The respec-

tive heat transfer coefficient and the reference fluid temperature

are taken as boundary conditions and the temperature distribu-

tion in the rock is computed numerically.

– The resulting surface temperature is used to adapt the boundary

condition of the heat transfer coefficient and the simulation of

the rock sample is repeated.

– This procedure is repeated until the difference between two con-

secutive values of the surface temperature is negligible. The

results of the final temperature distribution and the correspond-

ing boundary conditions will give the actual heat flux in the rock

experiments.

7.2.3 Future drilling measurements

The most important objective of the hydrothermal spallation drilling project

is the examination of the feasibility of this technology. The first question

of a feasibility analysis, i.e. whether the idea works, has been satisfactorily

answered. The results of this thesis have proven that drilling with hy-

drothermal flames is possible. Once the capability is explored and proven,

the next questions refer to the economical and field implementation of the

192


new technology. In order to prove that the technology is technically and

economically viable, a series of further questions has to be explored:

• The experiments in the laboratory must prove, that drilling holes

with diameters comparable to the ones of conventional bore-holes is

feasible.

• At least an estimation of the achieved drilling velocities in the labo-

ratory is needed.

• Once this data becomes available, a detailed and in depth analysis

of the economics of the procedure must follow. Through this analy-

sis, the most significant obstacles for the technology will emerge and

further research will focus on overcoming them.

• In case the technology will seem economically viable further field

experiments should follow.

The future drilling experiments must focus on answering the aforemen-

tioned questions and several adaptations of the existing plant will be nec-

essary.

The current pressure vessel is designed with a focus on simple one-off

drilling experiments, during which the flame and the rock are static. A

flame is ignited, the rock is positioned at a defined SOD from the flame

nozzle and a drilling experiment is performed. The thorough experimental

investigation of the spallation drilling process requires an extension of this

procedure. The system should be adapted to accommodate a continuous

relative movement of the flame and the rock while drilling. This movement

can be achieved with the lower positioning device, provided two prerequi-

sites are fulfilled:

• The distance between the rock surface and the nozzle must be con-

tinuously measured during the experiments.

• Additional safety measures must be taken to avoid hitting the nozzle

with the rock.

193


In a first stage, a safety mechanism should be integrated in the rock holding

plate. This plate must take the form of a sandwich of two plates connected

with springs to each other. A contact switch can be installed between the

plates, and once the rock is pressed from the combustion chamber nozzle

the switch will reverse the movement of the positioning device. With this

mechanism, simple experiments can be performed, where the first estima-

tions of the hole diameter and the drilling velocity could be attempted.

On a second development stage, a continuous, online distance measurement

must be implemented in the system. The resulting measurement device will

provide the SOD values online during the experiments and reliable data on

the achieved drilling velocities inside concave surfaces could be produced.

Several distance measurement methodologies could be integrated in the

system, with the use of ultrasound measurements being the most favor-

able. A typical ultrasound sensor could be fitted in the tube of the lower

positioning device. The sensor could sent ultrasound waves through the

rock probes, which would be reflected from the nozzle tube. A calibration

of the system with and without a flame in operation should be possible,

resulting to a reliable operation during continuous drilling.

The implementation of such a continuous distance measurement would open

a new chapter in the drilling experiments, by allowing for the actual online

measurement of drilling velocity. The experiments with the flat surfaces

can be then used only for the initial characterization of different rock types.

Once the values of the heat flux and the surface temperature necessary for

spallation will be defined, the experiments would proceed with pre-drilled

rock samples, in which the drilling velocities could be online measured. This

way the situation in the actual bore-hole will be modeled experimentally,

and the optimization of the drilling tool will be possible.

In parallel to the development of a distance sensor, the design and con-

struction of a new pressure vessel, able to accommodate longer rock probes,

could ease these experiments. The length of the current vessel limits the

possibilities for continuous drilling experiments, since the drilled holes can-

not exceed the depth of 15 cm. The adaptation of the current vessel should

194


be a relatively easy task, because both its upper and lower flanges can be

kept and only a longer vessel body has to be constructed.

While the findings of this thesis are restricted to the aforementioned initial

drilling experiments, the plant has proved to be not only operational but

also flexible and capable of further adaptations. It therefore forms the

basis for further research, which can address many of the questions that

still remain open.

195

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Appendix A

The design of the pressure

vessel

Starting from a very simple sketch of the vessel and gradually defining

many details, a series of seven design steps were made in total. During this

procedure the general rough mechanical stability calculations have been

performed, and in parallel the safety analysis refined some aspects of the

vessel design. The following paragraphs present the first and the last design

drawings along with the main features of the vessel.

A.1 First concept

The first drawing of the vessel along with three details are presented in

Fig.A.1 and Fig.A.2 respectively.

In this initial stage, most of the solutions were based on the existing vessels

in the LTR. Thus, only one port is foreseen for each cooling water stream, as

can be seen in sections D-D and C-C in Fig.A.1. Section C-C in the same

figure and detail G in Fig.A.2 present the two small sapphire windows,

which have the same viewing angle and depth in the inner volume. The

oxygen inlet port is placed on the same plane and its angle to the fuel inlet

direction can be seen in detail G of Fig.A.2.

215

Appendix A The design of the pressure vessel

400

140

100

Figure A.1: First general drawing of the pressure vessel. The detailsdesignated here are presented in Fig.A.2.

A construction feature, important for the experimental modeling of the

drilling procedure, is the flow pattern of the fluids exiting the vessel. The

simulation of the upward fluid flow in an actual bore hole is modeled by

locating the outlet ports of the main vessel volume on its upper part. These

outlet ports can be seen on section view A-A in Fig.A.1 and detail F in

Fig.A.2. Furthermore, three thermocouple ports are arranged on the side

wall of the vessel on the same section plane with the outlet ports, in order

to monitor the temperature of the cooling mantle water (CW2).

In this initial stage, the combustion chamber design, the centering of the

fuel injection nozzle and the injection system of the cooling water are only

216

A.1 First concept

Combustion products outlet

CW1 inlet

CW2 outlet

Detail F

Detail E

Detail G

Fuel inlet

Oxygen inlet

Figure A.2: Details referring to the first general drawing of the pressurevessel (Fig.A.1).

conceptually illustrated. Similarly, the cooling mantle tube is fixed on the

vessel top side, making its maintenance and replacement difficult, and no

pressure balancing mechanism between the two vessel volumes is foreseen.

In all the preceding supercritical water oxidation projects, the implemented

pressure vessels had no direct overpressure release mechanism. A conse-

quence of this practice is the lack of a bursting disc directly connected to

the vessel volume. In the same context, the pressure was measured and

controlled just before the pressure controller of the previous plants and not

directly on the pressure vessels. In all these cases this was acceptable, due

to the short tubes between the vessel and the pressure controller. As this is

not the case in the new plant, this absence of a direct pressure transducer

on the vessel was corrected in the subsequent design steps.

217

Appendix A The design of the pressure vessel

A.2 The final design of the pressure vessel

The final pressure vessel design is presented in the drawings of Fig.A.3-A.6.

On the section in Fig.A.3, the injection of the combustion chamber cooling

water can be seen. As Prikopsky [53] argued, the uneven distribution of this

stream - injected through only one port in his case - could lead to flames

issuing in a direction with a slight angle to the axis of their combustion

chamber. Weber [30] and La Roche [29] tried to solve similar problems

with special inserts in the flow, adding a radial component in its velocity

field. Two water inlet ports diametrically opposed to each other are used

in the new vessel to address this issue. The safety analysis of the vessel

Pressure measurement,and bursting discWindow90

140

1 X 12

CW1 inlets

Fuel

inle

t

Thermocouple ports Pressure balancing holes

Figure A.3: Section A-A of the final design of the pressure vessel.

resulted in the addition of two pressure measurements (a mechanical and

an electronic) and a bursting disc at the lower side of the vessel wall. A

further addition of this analysis were the twelve holes (1mm diameter) in

the mantle tube in order to balance the pressure between the two vessel

volumes.

Fig.A.4 presents the plane of the vessel, on which the oxygen and the CW2

ports are situated. Although the water ports are presented in this drawing

to be diametrically opposed to each other, they have an offset of 15mm

in the finalized vessel. This offset was introduced to increase the cooling

218

A.2 The final design of the pressure vessel

capacity of the mantle, by adding a radial velocity component to the flow.

Furthermore, the mantle tube is now fixed on the lower side of the vessel

and its installation and replacement are much easier.

Oxy

gen

inle

t Cooling mantle fixation

Window

CW2 inletCW2 outlet

CW2 outletCW2 inlet

400

100

Figure A.4: Section C-C of the final design of the pressure vessel.

With the long term use of the vessel in mind we optimized the visual access

to the flame by adding in total four windows to it. Two windows are placed

on the vessel head and two on its side walls. The side windows are kept

closed with blinds during normal operation, and their future use will require

a replacement of the metal mantle tube by one made of glass. They are

positioned at 100mm (see Fig.A.4) and 90mm from the vessel head (see

Fig.A.3). In the finalized design of the upper windows we have placed them

on two planes with two viewing angles, to provide extended optical access

of the flame. Their viewing depths are 80mm and 70mm and on of them

is presented in Fig.A.4. Unfortunately, mechanical stability regulations

forced us to finally place them on the same vessel plane. In Summary. if

all four windows are used the observation of the flame at three plains and

four different points is possible.

Irrespective of vessel arrangement, the drilling process taking place in it

will produce rock particles and their management is of capital importance.

Apart from the scientific significance of the particle size distribution, the

protection of the pressure controller should be solved. In order to gather

as many of the produced particles as possible, two collection stages have

219

Appendix A The design of the pressure vesselW

indo

w

80

Figure A.5: Section B-B of the final design of the pressure vessel.

been implemented. The first filtering level is implemented at the outlet

tubes of the vessel in the form of coarse filters with a filtering threshold

of 1mm, as shown in Fig.A.6. The second filtering stage is realized from

the filter installed before the pressure controller, which was presented in

section 2.1.1.

Coarse Filter

D

combustion chamber outlet

60

Figure A.6: Section H-H of the final design of the vessel.

220

List of publications

Journal publications

1. P. Stathopoulos, F. Hofmann, T. Rothenfluh and P. Rudolf von Rohr,

Calibration of a Gardon Sensor in a High-Temperature High Heat

Flux Stagnation Facility. Experimental Heat Transfer 25, 222-237,

2012.

2. P. Stathopoulos, K. Ninck and and P. Rudolf von Rohr, Heat transfer

of supercritical mixtures of water, ethanol and nitrogen in a bluff body

annular flow. The Journal of Supercritical Fluids 70, 112 - 118, 2012.

3. P. Stathopoulos, K. Ninck and and P. Rudolf von Rohr, Hot-wire

ignition of ethanol - oxygen hydrothermal flames. Combustion and

Flame - under review.

Conference contributions

1. P. Stathopoulos, T. Rothenfluh, M. Schuler and P. Rudolf von Rohr,

Assisted Ignition of Hydrothermal Flames in a Pilot Hydrothermal

Spallation Drilling Plant. Stanford Geothermal Workshop, Stanford

University - Palo Alto CA, 2012.

Curiculum vitae

Name: Panagiotis Stathopoulos

Date of birth: March 6th, 1982

Nationality: Greek

Marital status: Married

01.2010 - 04.2013 Doctoral studies at the institute of Process

Engineering, D-MAVT, ETH Zurich

01.2009 - 12.2009 Teaching assistant at the institute of Process

Engineering, D-MAVT, ETH Zurich

02.2008 - 12.2008 Energy systems engineer at HELECTOR SA

10.2007 - 02.2008 Research assistant at the National technical

university of Athens (NTUA) for the project

RECOFUEL

02.2007 - 02.2008 Military service

10.2005 - 02.2007 Studies in the program “Energy production

and management” of the NTUA (academic ti-

tle Msc.)

07.2006 - 10.2006 Internship at the National Power Corporation

at the power station of Lavrion

03.2005 - 10.2005 Research assistant at the Laboratory of Inter-

nal Combustion Engines of the NTUA

10.1999 - 02.2005 Mechanical engineering studies at the NTUA

(academic degree Dipl.Eng.)

10.2004 - 11.2004 Internship at the Hellenic Register Of Ship-

ping

07.2002 - 08.2002 Internship at the National Power Corporation

at the power station of Lavrion

Zurich, April 2013

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