design of a test set-up for experimental investigation of

Stijn Daelman

supercritical conditions.forced convective heat transfer to fluids operating atDesign of a test set-up for experimental investigation of

Academic year 2013-2014Faculty of Engineering and ArchitectureChairman: Prof. dr. ir. Jan VierendeelsDepartment of Flow, Heat and Combustion Mechanics

Master of Science in Electromechanical EngineeringMaster's dissertation submitted in order to obtain the academic degree of

Counsellor: Marija LazovaSupervisor: Prof. dr. ir. Michel De Paepe

The author gives permission to make this master dissertation available for consultation and to

copy parts of this master dissertation for personal use.

In the case of any other use, the limitations of the copyright have to be respected, in particular

with regard to the obligation to state expressly the source when quoting results form this master

dissertation.

Ghent, June 2014

The author

S. Daelman

”Wisdom is not a product of schooling but of the lifelong attempt to acquire it.”

-

Albert Einstein

The reason a lot of people do not recognize opportunity is because it usually goes around wearing

overalls looking like hard work.

-

Thomas A. Edison

Preface

This master dissertation is a final conclusion of my career as an engineering student at Ghent

University.

Even through the relative short timespan of my studies, technology and science have changed a

lot. The energy market and the technologies related to it not the least. Putting the knowledge

gained during the past years to a practical use within this context was something I was looking

forward to for my dissertation.

The design and building of a new test set-up was engineering in all its possible aspects. The

combination of the theoretical, mathematical, technical, economic and especially social aspects

of engineering were an enrichment of my education.

I would like to thank a lot of people and friends that have been important during my studies

and especially during this final year.

First of all, I would like to thank all the other people of the Department of Flow, Heat and

Combustion Mechanics. Especially my supervisor Prof. dr. ir. M. De Paepe and my counsellor

Marija Lazova for the trust they had to let me work on a project this big and important.

Special thanks also go out to Hugo Bellinck, Patrick De Pue, Adriaan Lebbe and Yves Maenhout

for the technical support and to Annie Harri and Griet Blonde for the administration. I know it

has been a lot of work!

Furthermore I would like to thank the other students in the lab. Marijn Billiet, Thomas Deruyter,

Piet Scheerlinck, Ben D’Haeger, your company and support are well appreciated, especially when

things didn’t go as expected.

Finally, I would like to thank my parents for giving me the opportunity to complete my studies

and for all their support!

Stijn Daelman

Ghent, June 2014

i

Design of a test set-up for experimental investigation of forced convective

heat transfer to fluids operating at supercritical conditions.

Stijn Daelman

Master’s dissertation submitted in order to obtain the academic degree of Master of Science in

Electromechanical Engineering

Supervisor: Prof. dr. ir. Michel De Paepe

Counsellor: Marija Lazova

Department of Flow, Heat and Combustion Mechanics

Chairman: Prof. dr. ir. Jan Vierendeels

Faculty of Engineering and Architecture

Ghent University

Academic year 2013-2014

Summary

The goal of this master’s dissertation is to develop a test set-up that is able to perform measure-

ments on the heat transfer process to refrigerants in supercritical conditions.

In chapter 1, an introduction will be given on the current energy challenges and how Organic

Rankine Cycles (ORCs) can provide a solution to these challenges.

In chapter 2, a general overview is given of the different types of ORCs and why the use of

a heat transfer process to fluids at supercritical conditions can be important to improve their

performance. Furthermore, the general selection procedure of working fluids is clarified. A review

of past work and existing correlations for forced convective heat transfer and existing set-ups

concludes this chapter.

Chapter 3 gives a description of the actual developed set-up and its general working principles.

A more detailed version of how the set-up was designed and why certain choices have been made

can be found in chapter 4, which discusses the design procedures.

The importance of the data-acquisition and control hardware is discussed in chapter 5. The

general ideas for the control software are discussed in chapter 6, whereas chapter 7 gives more

information on the operational procedures.

In chapter 8, the processing of the acquired data and the modified Wilson plot technique for

determination of the heat transfer coefficient are discussed.

Finally an error analysis is performed on the calibration of the temperature sensors and the

calculations that are used in the processing of the data.

Keywords

Design, Test set-up, Supercritical, Forced convective heat transfer, Organic Rankine Cycle

iii

Extended abstract

v

EINDWERKEN 2014 Onderzoeksgroep Technische Thermodynamica en Warmteoverdracht Vakgroep Mechanica van Stroming, Warmte en Verbranding - UGent

Paper number: SD

EXPERIMENTAL FACILITY TO STUDY FORCED CONVECTIVE HEAT TRANSFER TO WORKING FLUIDS FOR ORGANIC RANKINE CYCLES AT SUPERCRITICAL

CONDITIONS

Daelman Stijn, Lazova Marija, De Paepe Michel Department of Flow, Heat and Combustion Mechanics

Ghent University Sint-Pietersnieuwstraat 41 – B9000 Gent – Belgium

E-mail: [email protected]

ABSTRACT Organic Rankine Cycles (ORC) have proven their use to

convert ‘waste’ heat (e.g. from process industry) and other low-temperature heat sources (e.g. solar thermal) to usable electrical energy. The efficiency of these cycles however, has still room for improvement. Transcritical ORCs, in which the heat transfer from the heat source to the cycle’s working fluid happens at supercritical conditions, can provide a higher energy conversion efficiency.

In order to design these transcritical cycles, the heat transfer process to the working fluids at supercritical conditions has to be studied in order to provide useful correlations for design calculations of such a heat exchanger. In order to get to these correlations, a test set-up that provides data on this heat transfer process has been designed.

This paper reports on the design study in terms of operational parameters, technical limitations, technical implementations and analytical methods used in such a set-up. The results obtained during this study will be used in the ORCNext project to design a supercritical heat exchanger to be applied in a transcritical ORC.

INTRODUCTION

A shift towards the use of low-temperature energy sources, such as flue gasses, have led to a wider use of Organic Rankine Cycle. A common problem in ORCs however, is the exergy destruction in the heat transfer process in the evaporator within the ORC. The temperature mismatch between working fluid and heat source, such as visualised in Figure 1 is the mean reason for this exergy destruction.

In order to provide a solution, a transcritical cycle can be used. In this cycle architecture, the heat transfer during the evaporation phase takes place at supercritical pressures, providing a better thermal match between heat source and working fluid. This is indicated in Figure 1.

The lack of knowledge of the heat transfer to the applied working fluids within these transcritical ORCs leads to an over- dimensioning of heat exchangers, thus resulting in a lower

NOMENCLATURE Thermodynamic symbols Cp [kJ/kgK] Heat capacity h [W/m²K] Convection coefficient k [W/mK] Conduction coefficient Nu [-] Nusselt number P [bar] Pressure Pr [-] Prandtl number q [W/m3] Heat flux R [m2K/W] Thermal resistance Re [-] Reynolds number T [K] Temperature v [m³/kg] Specific volume Abbreviations GWP Glowal Warming Potential ODP Ozone Depletion Potential ORC Organic Rankine Cycle Subscripts b Bulk c Critical min Minimum w Wall 0 Ambient or reference

economic feasibility of the transcritical ORCs, which can be a threat to this technology within the current energy markets. Consequently, further application of ORCs necessitates the development of a heat transfer correlation for the applied working fluids under the supercritical conditions. Thanks to these correlations, heat exchangers for heat input in transcritical ORCs can be dimensioned in an appropriate way, thus lowering the costs of such installations. This can result in more efficient energy cycles, using low-temperature energy sources, whilst being economic sustainable as well.

Figure 1 Temperature profiles.

PROPERTIES OF SUPERCRITICAL HEAT TRANSFER Supercritical heat transfer to fluids has already been studied

by other researchers. However, the fluids in question were mainly water, helium and CO2. The working fluids that would be used within transcritical ORCs have not been tested yet. This doesn’t mean that the results and findings of these former studies are of no use for these working fluids. It can be assumed that the general properties that have been found for supercritical heat transfer to water, helium and CO2 are also applicable to the working fluids of interest used in transcritical ORC applications.

During previous studies, thermophysical properties of fluids

at supercritical conditions have been mapped. These are represented on PV-diagrams and Ts-diagrams, such as represented for R125 on figures Figure 2 and Figure 3.

M. Bazargan [1] made a very complete review in terms of

reviewed documents on the past experimental research on heat transfer to supercritical fluids.

Figure 2: PV-diagram of R125.

Figure 3: Ts-diagram of R125.

In the reviewed research findings, the effects of buoyancy

forces and acceleration effects come forward. These two aspects tend to be able to have a significant impact on the heat transfer process in certain conditions. However, before discussing how these effects influence the heat transfer process, a brief discussion of their existence is at hand.

When neglecting the pressure losses over the heat

exchanger, the heat transfer process can be modelled as a variation of temperature along an isobar in the PV-diagram.

The variation of the fluid’s properties along this isobar are studied as well, and the results have proven to be quite interesting.

At near-critical conditions (80 bar, 40°C for CO2), the variations of the heat capacity Cp and the density ρ show large variations for rather small variations in temperature. This can be seen on Figure 4, which is generated by Liao and Zhao [1] in their study of supercritical CO2.

Figure 4: Cp and ρ variations in function of temperature near

the critical point.

To discuss the effect of buoyancy and acceleration, a horizontal tube with a fluid at supercritical conditions is being used as an example. In this tube, a differentiation between the upper flow layers and the lower flow layers will be noticeable. This is a result of the buoyancy effect and will also result into a temperature difference between top and bottom side of the tube.

The hotter, less dense fluid will raise to the upper section of the tube, whereas the colder, denser fluid remains at the lower section. The influence of this phenomenon on the heat transfer process will be larger for larger tube diameters and for lower mass flow rates of the working fluid in the tube.

To be able to measure the buoyancy effect in the newly designed set-up, temperature measurements at each measuring point are executed at three locations along the tube’s circumference.

Next to the buoyancy effect, the rapid change of density will result into an acceleration of the flow because for a constant mass flow rate, the same cross-section will be available, but a larger volumetric flow will be established.

EXISTING CORRELATIONS FOR SUPERCRITICAL HEAT TRANSFER

By using a variety of existing heat transfer correlations for supercritical heat transfer, derived from the previously studied fluids, an acceptable estimation of the heat transfer coefficient (HTC) can be made, which is used to determine the working range of the set-up and the component’s specifications.

The correlation of Petukhov, Kranoschekov and Protopopov

for water and CO2 at supercritical conditions [1] has proven to show a good predicting performance. About 85% of their data and data of experiments by other researchers matches their correlation within a deviation of 15%.

(1)

( )( )

0,35 0,11

0,,

0, 0,52/3

210

Re Pr8

12,7 Pr 1 1,07

) 1,64

8

1,82log (Re

p b bb b

p b w w

bb

bb

b

CNu

C

f

Nuf

f

Nu λ µλ µ

−

= − +

=

−

=

Swenson, Carver and Kakarala [1] derived a correlation for

water in horizontal tubes:

(2) w

Nu hdk

=

(3)

0,923 0,6133

0,231

4,59 w b w

w w b w

w

b

DG H HNu eT T k

µµ

ρρ

− −= ⋅ −

⋅

A correlation from Shitsman [1], for helium, CO2 and water

in horizontal tubes in the Dittus-Boelter form is: (4) 0,8 0,80,023Re Prb b minNu = With Prmin the smaller one of Prb and Prw .

A fourth and final correlation that is used in the pre-design phase is the one from Jackson and Fewster [1]. This correlation is in the Dittus-Boelter form for constant properties convention:

(5)

0,30,50,820,0183Re Pr w

b bb

Nu ρρ

=

In this equation, any effect of buoyancy is being neglected.

EXTERNAL BOUNDARY CONDITIONS Because of the future use in a low temperature heat source

ORCs, several boundary conditions on the parameter testing range exist. These limitations apply to the working fluid, geometric properties of the heat exchanger, pressure levels, flow rates and the power to be transferred.

The selection of the working fluids to be studied is done at

hand of a range of parameters. First of all, there are thermo-physical parameters, such as the critical temperature and critical pressure of the working fluid. These properties give a direct indication whether the fluid is a valid candidate to be used in supercritical ORC’s. A third thermophysical property is the fluid type: wet, dry or isentropic. This last property is of importance when including a rotating expander in the cycle. In the study of heat transfer, all three types will be used.

The safety properties of the working fluids have a very

significant influence on the choice of refrigerants. For now, only refrigerants of the A1 and A2L ASHRAE 34 safety group will be studied. These have a low toxicity and no or low flammability.

Finally, the environmental aspects of the refrigerants have

been taken into account. This includes a study of the ozon depletion potential (ODP) and the global warming potential (GWP), which both have to be as low as possible.

When considering a general ORC-cycles that use industrial waste heat, the most practical implementation of the heat exchangers is in the form of a shell and tube heat exchanger, using thermal oil to transfer the heat from the heat caption elements to the actual electricity generating cycle. These heat exchangers are commonly used in a horizontal configuration, thus the focus of the study will be on a horizontal heat transfer test section.

Keeping the practical aspect in mind, the tube diameters that

are tested are also within the range of practical applications. The complete circuit is constructed out of 1 1/8th inch tubing, including the current test section. For future studies of smaller diameters, only the tubing of the test section has to be replaced.

The goal of the flow rates to be tested are derived from the

mass fluxes G, since this is more relevant for testing in a single-tube set-up. The mass fluxes range from 100 kg/m²s to 1000 kg/m²s. However, for certain conditions, the minimum and maximum will have to be altered a bit, in order to be compliant with the limitations of the components in the set-up.

A final result of the practical application, is to use a tube-in-

tube test section, which uses thermal oil in the annulus and the working refrigerant in the inner tube. This simulates the practical conditions in the ORC’s perfectly, but is harder to do measurements on. Another solution might have been a test section which is uniformly heated thanks to a Joule-heating method, in which the electrical resistance of the tube is used as a heat generating source. Other set-ups using this method, however, make notice of high wall temperatures. These can have a large effect on the heat transfer to the refrigerant in the boundary layer. Therefore the final decision has been made to use the primary described tube-in-tube approach.

NUMERICAL METHOD A numerical solution approach was followed in order to

calculate the heat transfer in the test section. The heat transfer at the side of the supercritical fluid is calculated with the different correlations that are listed above. By using this range of correlations, a reasonably window with the order of magnitude for the heat transfer has been found.

These calculations are also used to find an optimal length for the set-up. For this length, a trade-off has to be made between pressure drop and temperature difference over the test section, whilst keeping available space in mind.

The heat transfer over the complete test section is calculated using following formula’s:

(6) LMTDQ UAF T= ∆

(7) 1 1 1

2

o

i LMTD

i i oo

dlnd

UA h A L h A Qπλ

∆ = + + =

In equation (6), F is equal to one, due to the counterflow setup of the test section. In equation (7), no fouling factor is taken into account since the set-up will be used in a controlled lab environment and the focus of the study is on the heat transfer coefficient ih at the inner side of the tube.

The heat transfer over the test section is calculated by discretizing the test section into multiple smaller sections, over which the temperature change, and thus the change of properties is limited. The following calculations are used implemented in Engineering Equation Solver (EES):

Gnielinski correlation for convection coefficient oh in the

annulus:

(8)

( )

( )

2

2/3

2/3

( ) 3,28

( 1000)2 1

1 12,

1,58

7 P2

r 1

hydraulic

ln Ref Re Pr D

uL

Nf

f −⋅ −

− + + −

=

=

Correlations from Petukhov et al., Swenson et al., Shitsman

et al. and Jackson and Fewster, such as listed in formula’s (1) to (5) are used to calculate the inner convection coefficient ih .

The results of these calculations and measurements on the setup are used to provide first impression of their applicability with the working fluids at hand.

PROCESSING OF RESULTS The modified Wilson plot method of Briggs and Young is

used to find a new correlation for the inner convection coefficient in the test section.

For a turbulent flow, the Nusselt numbers at the inner and outer side of the tube are:

(9)

0,141/3

0,141/3

Pr

Pr

ai ii i i i

i w

aoo

o w

oo o o

Nu

N

h D C Rek

ku h D C Re

µµ

µµ

=

=

=

=

By rearranging these expressions into expressions of ih and

oh and substituting these expressions into equation (7), following expression is found:

(10)

0,141/3

0,141/3

1 iwall

ai i i i i

w

o

ao o o o o

w

D RUA muk AC Re Pr

muD

muk A C Re Prmu

= +

+

In general, a can be taken equal to 0,8 . Rearranging the

equation to the form Y mX b= + enables the use of linear regression on the data to determine m and b .

(11)

0,141/3

0,141/3

0,141/3

1

1 1

aow o o o

o w o

aoo o o

o w o

i oaii i i

i w i

kR A Re PrUA D

k A Re PrD

C Ck A Re PrD

µµ

µµ

µµ

− =

⋅ +

The coefficients achieved through this regression can then

be implemented again in equation (9) to find the corresponding Nusselt numbers.

EXECUTION OF DESIGN AND SET-UP BUILDING

During the transition phase from a theoretical design, based upon calculations, simulations and operational strategies to the actual building of the set-up, several challenges are met.

This is also one of the aspects that was of importance during this project, since these challenges will also be met on large-scale set-ups and are easier to anticipate when working on a lab scale.

It is very much recommended to work closely together

with the suppliers of the components that are needed in the set-up, since supercritical fluids are not yet commonly used in current installations and might pass the boundaries that are set by the manufacturer of the component.

The set-up at Ghent University is divided into three

circuits: the main circuit, which contains the working fluid that is being studied, and two supporting circuits. The supporting circuits provide the necessary heating and cooling capacity to the main circuit.

Figure 5 Schematic layout.

REFRIGERANT CIRCUIT

The refrigerant circuit on Figure 6 consists around the 4m long tube-in-tube test section, in which the working fluid is heated and its temperature is continuously monitored.

Figure 6 Refrigerant circuit.

The inlet pressure of the test section are controlled by the

circulation pump and the pressure control relief valve. The inlet temperature of the test section is controlled by

two tube-in-tube preheaters, which are necessary because the working fluid has to be condensed before entering the circulation pump, with on/off control and an electrical preheater which has a finer PID temperature control.

The test section is a 4m long tube-in-tube heat exchanger

in which the working fluid flows through the inner tube and thermal oil flows through the annulus as heating fluid.

Measurements of pressure and temperature of the working fluid are performed directly at in- and outlet of the test section. Over the total length of the test section, temperature measurements are performed on the heating fluid and on the wall of the inner tube.

By monitoring the wall temperature at a 333mm interval, the temperature gradient over the test section is visualized. In order to measure the influence of the buoyancy effect, three measurements are performed at interval point: Top, side and bottom. From the temperature difference between the three point, the range of working conditions in which the buoyancy effect plays a significant role on the heat transfer process can be determined.

Figure 7 Thermocouple configuration in the tube-in-tube test

section.

An accumulator at the suction side of the circulation pump provides a NPSH, large enough to provide a stable flow. The pump is a diaphragm pump with 5 internal diaphragms, which results into minimal flow and pressure pulsations.

Table 1 Properties of working fluids. Property HFC-134a HFC-125 HFO-1234yf Chemical formula CF3CH2F C2HF5 C3F4H2

cT 101,03°C 66,02°C 94,7°C

cp 4060,3kPa 3618kPa 3382,2kPa ODP 0 0 0 GWP 1430 3500 4

HEATING CIRCUIT The heating circuit is filled with low-viscosity thermal oil.

This enables the use of a tube-in-tube heat exchanger as test section. The heat transfer from the heat source to the interface wall between heating fluid and working fluid has to be sufficiently large to minimize its influence on the total thermal resistance. This can only be achieved by using high velocities in the annulus of the heat exchanger. By using a low-viscosity fluid as heat source, these velocities can be achieved whilst keeping the pressure drop in the annulus within practical limits.

The heating capacity in the test section is controlled by means of a bypass valve which controls the flow through the annulus. COOLING CIRCUIT

The cooling circuit provides a cold source for cooling and condensation of the working fluid in the low-pressure section of the refrigerant circuit. This is required to meet the specified conditions at inlet of the pump.

The cooling circuit consists of a 37kW chiller with on/off thermostatic control that is connected to an insulated buffer vessel of 900l which stores a 30% Ethyleneglycol – 70% water mixture. This solution provides a stable cooling temperature and results into economical operation of the chiller.

As for the heating circuit, the cooling capacity is also controlled by a regulated bypass.

OPERATIONAL PROCEDURES AND CONTROL MECHANISMS

A set-up at elevated temperatures and pressures requires a fail-safe control system. This has to be mechanically included in the design and the software also has to be adapted to guarantee safe and accurate operation.

To provide an extra safety layer, a headless control unit

with its own real-time operating system is used, in order to avoid errors or safety issues due to communication errors between the control unit and the computer that is used for measurements. This control unit determines the operational conditions based on measurements of specific parameters. These parameters are processed and adaptations in control are made to achieve or maintain the desired operational conditions.

CONCLUSION Heat transfer to fluids at supercritical conditions has been

researched over the past decades. The possible application of the derived correlations for use with heat transfer to working fluids of ORCs needs verification and adjustments. A new test facility at Ghent University has been developed and is briefly described in this paper.

The design and construction of a new test set-up always

proves to be a challenge. For an increase in size and complexity, the challenges also prove to become harder to solve. A close collaboration between the scientists, engineers, suppliers and the constructor, proves to result into a good combination of insights, which are necessary for completion of the project.

The next step is to further develop and implement the operational strategies of the set-up.

REFERENCES [1] E. K. a. V. P. B. Petukhov, „An investigation of heat

transfer to fluids flowing in pipes under supercritical conditions,” in Developments in Heat Transfer Part, vol. 3, 1961, pp. 569-578.

[2] M. Bazargan, „Forced convection heat transfer to turbulent flow of supercritical water in a round horizontal tube,” 2001.

[3] S. L. a. T. Zhao, „Measurements of Heat Transfer Coefficients From Supercritical Carbon Dioxide Flowing in Horizontal Mini/Micro Channels,” Journal of Heat Transfer, vol. 124, nr. 3, p. 413, 2002.

[4] J. C. a. C. d. K. H. Swenson, „Heat transfer to supercritical water in smooth-bore tubes,” Journal of Heat Transfer, vol. 87, p. 477, 1965.

[5] M. Shitsman, „Heat transfer to supercritical helium, carbon dioxide and water: Analysis of thermodynamic and transport properties and experimental data,” Cryogenics, 1974.

[6] J. J. a. J. Fewster, „Forced convection data for supercritical pressure fluids,” Simon Engineering Laboratory, University of Manchester, 1975.

Contents

Preface i

Extended abstract v

1 Introduction 1

1.1 Low grade heat as an energy source . . . . . . . . . . . . . . . . . . . . . . . . . 2

1.2 Organic Rankine Cycles for conversion of low grade heat . . . . . . . . . . . . . . 3

2 Literature 5

2.1 Supercritical conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5

2.2 Organic Rankine Cycles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7

2.2.1 Classification of ORCs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7

2.3 Selection of working fluids . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11

2.3.1 Thermophysical properties at supercritical conditions . . . . . . . . . . . 11

2.3.2 Safety properties . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17

2.3.3 Environmental properties . . . . . . . . . . . . . . . . . . . . . . . . . . . 18

2.3.4 Working fluids candidates . . . . . . . . . . . . . . . . . . . . . . . . . . . 19

2.4 Heat transfer to fluids at a supercritical state . . . . . . . . . . . . . . . . . . . . 20

2.4.1 Overview of experimental research on heat transfer to supercritical fluids 20

2.4.2 Effects of buoyancy and acceleration . . . . . . . . . . . . . . . . . . . . . 22

2.4.3 Heat transfer correlations . . . . . . . . . . . . . . . . . . . . . . . . . . . 26

2.5 Existing set-ups for testing heat transfer to SC fluids . . . . . . . . . . . . . . . . 28

2.5.1 Freon Thermal Hydraulic Experimental Loop . . . . . . . . . . . . . . . . 28

2.5.2 Supercritical test set-up Tsinghua University of Beijing . . . . . . . . . . 31

2.5.3 Supercritical test set-up Seoul National University . . . . . . . . . . . . . 33

3 Test set-up description 37

3.1 Test set-up cycle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37

3.2 Description of the test facility . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38

3.3 Refrigerant circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 41

3.3.1 Test section . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 41

xiii

Contents xiv

3.3.2 Pressure regulating expansion valve . . . . . . . . . . . . . . . . . . . . . 42

3.3.3 Mechanical pressure relief valve . . . . . . . . . . . . . . . . . . . . . . . . 42

3.3.4 Low pressure section shut-off valve . . . . . . . . . . . . . . . . . . . . . . 42

3.3.5 Cooler/Condenser . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43

3.3.6 Accumulator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43

3.3.7 Circulation pump . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43

3.3.8 Tube-in-tube preheaters . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43

3.3.9 Electrical preheater . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43

3.4 Heating circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 44

3.5 Cooling circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 45

4 Design Procedure 47

4.1 Determination of test conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . 48

4.2 General required calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 49

4.2.1 Pressure drop characteristics . . . . . . . . . . . . . . . . . . . . . . . . . 49

4.2.2 Mixing valve characteristics . . . . . . . . . . . . . . . . . . . . . . . . . . 50

4.3 Refrigerant circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 52

4.3.1 Material selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 52

4.3.2 Test section design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 53

4.3.3 Circulation pump . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 57

4.3.4 Mass flow meter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 60

4.3.5 Preheaters . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 61

4.3.6 Pressure control systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . 62

4.3.7 Cooler/Condenser . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 65

4.3.8 Accumulator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 65

4.4 Heating circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 70

4.4.1 Heating fluid selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 70

4.4.2 Main heating unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 70

4.4.3 Flow rate control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 71

4.5 Cooling circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 72

4.5.1 Chiller and buffer vessel . . . . . . . . . . . . . . . . . . . . . . . . . . . . 72

4.5.2 Cooling circulation pump . . . . . . . . . . . . . . . . . . . . . . . . . . . 72

4.5.3 Cooling flow rate control . . . . . . . . . . . . . . . . . . . . . . . . . . . . 72

4.6 Test set-up assembly . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 73

5 Data-acquisition hardware 75

5.1 DAQ controller . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 75

5.2 Temperature sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 76

5.2.1 Thermocouples . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 76

5.2.2 Pt100 temperature sensors . . . . . . . . . . . . . . . . . . . . . . . . . . 78

Contents xv

5.3 Pressure sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 78

5.3.1 Accuracy . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 78

5.4 Coriolis mass flow meters . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 79

5.4.1 Accuracy . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 79

5.5 Sensor and driver tags . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80

5.5.1 Temperature sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80

5.5.2 Pressure sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80

5.5.3 Mass flow meter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80

6 Software 81

6.1 PC-software . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81

6.1.1 User interface . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81

6.1.2 CoolProp . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 82

6.2 Control software . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83

6.2.1 Pressure build-up . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83

6.2.2 Pressure drop . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83

6.2.3 Uncontrolled temperature raise . . . . . . . . . . . . . . . . . . . . . . . . 83

6.2.4 Pump inlet conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83

7 Operational procedures 85

7.1 Pressure test . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 85

7.2 System charging . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 85

7.3 System start-up . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 86

7.4 System shut-down . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 86

7.5 Heat balance test . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 86

7.6 Absolute limitations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 87

7.6.1 High pressure section . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 87

7.6.2 Low pressure section . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 88

8 Data reduction 89

8.1 Data acquisition and calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . 89

8.1.1 Annulus . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 89

8.1.2 Wall . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 89

8.1.3 Inner tube . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 90

8.2 Modified Wilson plot . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 91

8.2.1 Experimental expressions for the Nusselt number . . . . . . . . . . . . . . 91

9 Error analysis 93

9.1 Accuracy of temperature sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . 93

9.2 Accuracy of the heat transfer coefficient . . . . . . . . . . . . . . . . . . . . . . . 94

9.2.1 Local heat flux . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94

Contents xvi

9.3 Accuracy of the Wilson plot . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95

10 Conclusion 97

A Hardware 99

A.1 Plate heat Exchanger . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 99

A.2 Mixing valve diagram . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 109

A.3 Mass flow meter error analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . 111

B Process flow diagram 119

C Test section piece- and assembly drawings 123

D Complete set-up 3D renders 133

E Software 139

E.1 EES-simulation code . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 139

List of symbols

Abbreviations

Symbol Description

ASHRAE American Society of Heating, Refrigerating

and Air-Conditioning Engineers

BSL Best Straight Line

DAQ Data-Acquisition

EES Engineering Equation Solver

FEM Finite Element Method

FS Full Scale

FTHEL Freon Thermal Hydraulic Experimental Loop

GWP Global-Warming Potential

HMI Human-Machine Interface

HTC Heat Transfer Coefficient

NPSHa Net Positive Suction Head available

NPSHr Net Positive Suction Head required

NPSH Net Positive Suction Head

ODP Ozone Depletion Potential

ORC Organic Rankine Cycle

SC Supercritical

SCF Supercritical Fluid

SCWR Supercritical Water Reactors

UI User Interface

xvii

List of symbols xviii

Thermodynamic Symbols

Symbol Units Description

G[kg/m2s

]Mass flux

m [kg/s] Mass flow rate

Q[m3/s

]Volumetric flow rate

Cp [kJ/kg ·K] Specific heat capacity

f [−] Friction factor

G[kg/m2 · s

]Mass flux

g[m/s2

]Gravitational acceleration

Gr [−] Grashof number

h[W/m2 ·K

]Convection coefficient

i [kJ/kg] Specific enthalpy

k [W/m ·K] Conduction coefficient

L [m] Length

Nu [−] Nusselt number

p [Pa] Pressure

Pr [−] Prandtl number

Q [W ] Transferred heat

q[W/m2

]Heat flux

Re [−] Reynolds number

s [J/kg ·K] Entropy

T [C] Temperature

List of symbols xix

Greek Symbols

Symbol Units Description

β [1/K] Volumetric thermal expansion coefficient

∆ [−] Difference

η [−] Efficiency

λ [W/m ·K] Thermal conductivity

µ [Pa · s] Dynamic viscosity

ν[m2/s

]Kinematic viscosity

Subscripts

Symbol Description

b Bulk

C Cold

crit Critical

H Hot

h Hydraulic

i Inner

in Inlet

m Mean

o Outer

out Outlet

pc Pseudo-critical

th Threshold

TS Test Section

w Wall

List of symbols xx

Chapter 1

Introduction

This master’s dissertation is a part of the ORCNext project [1]. The ORCNext project focuses

on a next generation of Organic Rankine Cycle (ORC) technologies.

The ORCNext consortium focusses on two main technical reasons why the application of ORCs

is not as widespread as it potentially could be.

First of all, cycle efficiencies for installations using a low-temperature heat source, such as the

ORC, are quite low. This results into an ineffective recovery of the heat and consequently a low

electrical energy production.

Secondly, efficient systems for small power ranges should be investigated. This could open a

potential for a lot of small scale systems at energy sources that are too small to use for today’s

installations.

This master’s dissertation makes part of the study on the cycle efficiencies. The use of a

transcritical cycle, a new cycle architecture, is a promising step forwards. During this master’s

dissertation, a test set-up for experimental studies about the heat transfer to fluids at a super-

critical state is designed and built. The obtained results contribute to a deeper insight in the

heat transfer process to fluids operating at supercritical conditions, which can be used to design

more efficient heat exchangers.

1

Chapter 1. Introduction 2

1.1 Low grade heat as an energy source

According to the data of BP’s statistical review of world energy use [2], displayed in figure 1.1,

world primary energy use grew by 1,8 % in 2012. This figure indicates that oil, with 33,1 % of

the global energy consumption, remains the world’s leading fuel. In relative numbers however,

oil lost market share for 13 years in a row, whilst hydroelectric and renewable energy sources

reach their record shares at 6,7 % and 1,9 % of global primary energy use.

Figure 1.1: World energy use in million tonnes oil equivalent [2].

Because of the increasing energy cost and the awareness of the impact on the climate, recent

evolutions in industry aim for a higher energy efficiency of individual systems. The increase in

worldwide energy use can be easily explained since the amount of systems is still rising at a high

rate, especially in Asia, Africa, the Middle East and South-America.

In industrial applications, one of the most striking phenomena within the context of energy use,

is the amount of waste heat. This is one of the effects of the second law of thermodynamics. This

law states that there cannot exist a system that has a heat input and a work output without

an additional heat output. However, when a heat-source temperature drops below a certain

temperature (e.g. 370 C for steam generation), thermal efficiency becomes uneconomically low.

Statistical investigations indicate that this low-grade waste heat accounts for about 50 % of the

total heat generated in industries [3]. Therefore, research has been done on the application of

this low grade (low-temperature) heat.

A possible application of this waste heat is the use of an ORC, which is able to use this

low-temperature heat for the production of electricity.


1.2 Organic Rankine Cycles for conversion of low grade heat

The Organic Rankine Cycle is an adapted execution of the Clausius-Rankince Cycle. Instead of

using water as working fluid in the cycle, an organic working fluid is used as the cycle’s working

medium. This makes it possible to use heat sources at low temperature as energy input of the

cycle, whereas this heat might otherwise have been wasted into the atmosphere.

Figure 1.2 shows the layout for a simple ORC. As within the steam cycle, the working fluid

(being an organic working fluid instead of water) is compressed with a pump, (super)heated in

the evaporator, expanded in a turbine or expander and condensed in the condenser. In this

process, the turbine or expander drives an electric generator. The remaining heat from the

condenser can be used for heating applications or can be cooled down using a cooling circuit,

depending on its temperature.

Condenser

Generator

Turbine/Expander

Pump

Evaporator

Figure 1.2: Layout of a simple ORC.

Over the past decade, the research on ORC installations has been intensified because it shows to

be the most promising technology to convert low-temperature heat sources into power. Common

heat sources that are used today include solar energy, geothermal energy, biomass and waste

thermal energy from processes.

Chapter 2

Literature

During the design of a new experimental test set-up, many papers, surveys and reviews provide

parts of the necessary information. This clarifies the need for a broad view on different subjects

in literature. However, a more in depth approach is also necessary considering the main focus of

this study.

In this chapter, both aspects will be handled. First of all, an introduction to supercritical

conditions and a more general overview of types of ORCs will be given in order to clarify the

context of this research. The selection procedure for transcritical ORC working fluids will also

be discussed.

The research on heat transfer to fluids at supercritical conditions will be elaborated in more

depth. Finally some existing experimental set-ups will be reviewed in order to gain first insights

in the layout and principles behind these experimental facilities.

2.1 Supercritical conditions

A Supercritical Fluid (SCF) is easily defined as a fluid at a temperature and pressure above its

critical point. At these conditions, there is no real distinction present between the liquid phase

and the gas phase, so there is no real phase transition in this supercritical region.

On the phase diagram of CO2 on figure 2.1 a distinction between the different conditions based

on pressure and temperature is visualized. This chemical component is used in the solid, liquid,

gas and supercritical ’phases’ and its properties at all conditions are already quite well known.

5

Chapter 2. Literature 6

Temperature (K)

Pre

ssure

(bar

)

200 250 300 350 400

1

10

100

1000

10000

solid

liquid

gas

supercriticalfluid

triple point

critical point

Figure 2.1: Pressure-temperature phase diagram for CO2.

On this figure, the different areas are shown where CO2 exists as solid, gas, liquid and as SCF.

The curves on the figure represent temperatures and pressures at which two phases can exist in

an equilibrium. At the triple point, even an equilibrium between three phases is possible.

The properties of a fluid at supercritical conditions generally lie between those of a gas and a

liquid. By changing either the pressure or the temperature, the properties of the fluid can be

adapted to be more liquid- or more gas-like.

In power generation, Supercritical Water Reactors (SCWR) are being introduced in nuclear

systems in order to be able to work with a higher temperature of the heat source, thus resulting

in a higher theoretical efficiency (based on the Carnot efficiency of expression 2.1).

η =W

QH= 1− TC

TH(2.1)

Supercritical CO2 is also being investigated to be used in new heat-pump system applications.

This application also proves to provide a solid base in technical equipment for the use of SCFs in

ORCs.


2.2 Organic Rankine Cycles

Most low-grade heat sources, such as waste heat, will show a gliding temperature profile. Two

main reasons can be found for this phenomenon: The temperature level of the heat source is

not high enough or the amount of heat in the source is limited (e.g. a limited mass flow of flue

gasses) [4].

The fact that the temperature of the heat source drops as soon as heat is removed from the

source, proves that the heat source is a sensible heat source. A latent heat source would show a

region where heat is removed while the source remains at a constant temperature.

2.2.1 Classification of ORCs

Different types of ORCs are available. Cycle classification can be mainly based on where the

thermodynamic mechanisms take place in respect to the saturation curve on the Ts-diagram.

A major distinction between the different types of ORCs can be made when studying the heat

addition (evaporation) phase. The temperature profiles in the evaporator, such as represented

on figure 2.2 show how three types can be defined based on this thermodynamic process.

Figure 2.2: Temperature variation in the vapour generator for ORC, binary mixtures and transcritical

cycles[5].

The cycles, linked to these different temperature profiles, are elaborated in the following sections.

A first indication, why the transcritical ORC is particularly interesting, can already be seen when

comparing the mean temperature differences between the waste heat source and the working

fluid on figure 2.2.


2.2.1.1 Subcritical

A subcritical Organic Rankine Cycle can best be compared with the classical Rankine cycle. The

main change in working conditions is the use of an organic fluid instead of water as working fluid

for the cycle.

Ordinary working fluids

One of the limitations of this subcritical cycle using the standard working fluids (e.g. R134a)

is the constant temperature of evaporation. This makes the subcritical cycle less suitable for

sensible heat sources, such as waste heat, due to large irreversibilities. The use of zeotropic

mixtures, such as discussed below, and supercritical fluids (sections 2.2.1.2 and 2.2.1.3) can

reduce this problem.

T

S

12

3 45

6

6’7

Figure 2.3: Schematic of subcritical cycle for isentropic working fluids.

Zeotropic mixtures

Instead of using the simple, standard working fluids, zeotropic mixture working fluids can also be

used. These mixtures have so-called ’temperature slip’ in evaporation and condensation processes.

This results into a potential for reducing exergy destruction due to temperature differences and

thus higher efficiencies [6].

T

S

12

3 4 5

66’

7

Figure 2.4: Schematic of cycle for zeotropic mixtures.


2.2.1.2 Supercritical

The subcritical cycle can be adapted to a complete supercritical one. In this supercritical

thermodynamic power cycle, as proposed by Feher [7], the heat reception and rejection take place

at a pressure that’s higher than the working fluid’s critical pressure.

T

S

Figure 2.5: Schematic of supercritical cycle.

2.2.1.3 Transcritical

The heat transfer processes in transcritical cycles happen at different conditions. Heat is added

to the working fluid at supercritical pressure, whilst heat rejection in the condenser takes place

at subcritical pressure. In literature however, the differentiation between supercritical and

transcritical cycles is not always very clear. When referring to a supercritical cycle, most of

the times, the transcritical cycle is meant, but with the focus on the supercritical heat adding

mechanism.

T

S

12

3

4

4’5

Figure 2.6: Schematic of transcritical cycle for isentropic working fluids.


The main reason for the use of Supercritical (SC) ORCs is the better thermal match with the

heat source. Because the working fluid is at supercritical pressure, there is no two-phase region

as is the case with subcritical heat transfer. This better thermal match in the evaporator section

results in less irreversibility[8].

Chen [9] studied the performance of a system in case of a supercritical Rankine cycle using

CO2 as a working fluid and a subcritical ORC with R123 as working fluid. In this study, the

transcritical CO2 cycle shows a higher potential than the subcritical ORC, due to a better

matching of the temperature glide between heat source and working fluid.


2.3 Selection of working fluids

Many different organic fluids are available to use. Fluids that will be used in an ORC cycle

will have to be selected carefully, based on different selection criteria. These criteria can be

divided into three main groups: Thermophysical-, safety- and environmental properties. The

specifications and demands for each of these properties are discussed below.

2.3.1 Thermophysical properties at supercritical conditions

The thermophysical properties of a fluid can be studied on both PV-diagrams and Ts-diagrams

such as presented for R125 in figures 2.7 and 2.8. On the PV-diagram of figure 2.7, for pressures

lower than the critical pressure Pcrit, discontinuities can be found where the isotherms intersect

with the saturation line. The same phenomenon can be found on the Ts-diagram on figure 2.8

for isotherms intersecting the saturation curve below the critical temperature Tcrit. Between the

two discontinuities on an isotherm or isobar, a phase change occurs at constant pressure and

temperature respectively, where the slope is equal to zero.

The isotherm for the critical temperature shows a slope equal to zero at the critical point, where

the pressure is equal to the critical pressure. The same applies for the critical pressure isobar.

Above these critical pressure and temperature, isotherms and isobars don’t show any discontinuity.

This suggests a continuous transition from a liquid-like to a gas-like fluid.

When moving over a supercritical isobar, the complete fluid conditions can be determined with

only one temperature measurement. This is a very useful property of fluids at supercritical

conditions, since a combination of pressure- and temperature measurement defines the complete

condition of the fluid.


10-4 10-3 10-2 10-110-1

100

101

102

103

v [m 3/kg]

100°C 75°C

50°C

25°C

R125

Liquid

Liquid-VapourMixture

Vapour

Critical point

Pcrit

SaturatedLiquid

SaturatedVapour

Figure 2.7: PV-diagram of R125.

0,60 0,80 1,00 1,20 1,40 1,60 1,80-100

-75

-50

-25

0

25

50

75

100

s [kJ/kg-K]

3000 kPa 3500 kPa

0,2 0,4 0,6 0,8

103 R125

0,60 0,80 1,00 1,20 1,40 1,60 1,80 2,00-100

-75

-50

-25

0

25

50

75

100

125

150

s [kJ/kg-K]

1 bar

25 bar

50 bar 75 bar

R125

Tcrit

Liquid

Liquid-VapourMixture

Vapour

Critical point

SaturatedLiquid

SaturatedVapour

Figure 2.8: Ts-diagram of R125.


2.3.1.1 Critical temperature

First of all, the critical temperature Tcrit of the selected working fluid needs to be lower than the

temperature of the heat source.

Tcrit also needs to be higher than the temperature of the cold source that is used in the condenser

in order to achieve a transcritical cycle.

For a supercritical pressure, a pseudo-critical temperature Tpc can be found. This is the

temperature where the specific heat capacity rises to a peak. On figure 2.12, this pseudo-

critical temperature for CO2 at 80 bar and 100 bar can be observed to be about 37 C and 47 C

respectively.

2.3.1.2 Critical pressure

The critical pressure Pcrit of the working fluid is one of the main factors of influence for the

design of a cycle. This pressure determines the mechanical properties of several components such

as the thickness of tubes in the cycle, the type of pump to use, etc.

Should the critical pressure be too high, the cost of mechanical components can get too high for

the ORC to be economical feasible.

2.3.1.3 Saturation vapour curve: classification of fluid types

To determine whether a refrigerant is well suited to be used in ORC installations, the saturation

vapour curve should be studied.

The determination of the refrigerant type depends on the slope of the saturation curve in the

Ts-diagram.

Three types of refrigerants can be distinguished:

• Wet:dT

ds< 0

• Dry:dT

ds> 0

• Isentropic:ds

dT=∞

Since the value of dT/ds leads to infinity for isentropic fluids, the use of the inverse of the slope

is advised in order to make classification of working fluids easier to compare. The inverse of the

slope (ds/dT) can be used to express how ’wet’ or ’dry’ a fluid is, avoiding big values in the

near-isentropic region.


Wet fluids

A wet fluid will result in a two-phase state at the end of the expansion if overheating is limited.

R125 is as an example for this type of fluids.

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0,50 0,75 1,00 1,25 1,50 1,75 2,00 2,25-100

-75

-50

-25

0

25

50

75

100

125

150

s [kJ/kg-K]

2400 kPa

1000 kPa

320 kPa

70 kPa 0,2 0,4 0,6 0,8

R125

Figure 2.9: Ts-diagram of R125 as a wet fluid.

Dry fluids

A dry fluid will result in a dry vapour state at the end of the expansion, even when almost no

overheating is applied. Isopentane can be used as an example for this type of fluids.

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-2,0 -1,5 -1,0 -0,5 0,0 0,5 1,0-50

0

50

100

150

200

250

s [kJ/kg-K]

2300 kPa

1000 kPa

350 kPa

90 kPa

0,2 0,4 0,6 0,8

Isopentane

Figure 2.10: Ts-diagram of Isopentane as a dry fluid.


Isentropic fluids

An isentropic fluid has a quasi-vertical saturation vapour curve at the expansion side of the

diagram. The saturation vapour curve coincides with a theoretical isentropic expansion line.

R134a can be used as an example for this type of fluids.

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-0,25 0,00 0,25 0,50 0,75 1,00 1,25 1,50 1,75-100

-50

0

50

100

150

200

250

s [kJ/kg-K]

2700 kPa

1100 kPa

340 kPa

70 kPa 0,2 0,4 0,6 0,8

R134a

Figure 2.11: Ts-diagram of R134a as an isentropic fluid.


2.3.1.4 Property variations at supercritical conditions

Neglecting pressure losses in a supercritical heat exchanger, the heat transfer process can be

modelled as a temperature variation along an isobar. The variations of other fluid properties

along this isobar then can be studied. For fluids at supercritical conditions, most studies have

been performed on water, CO2 and helium.

As an example for the behaviour of the specific heat Cp, figure 2.12 shows the results of Liao and

Zhao [10]. During their study they noted that one of the most important characteristics of fluids

at supercritical pressures is the rapid variation of their physical properties with a small change

of temperature and pressure. This phenomenon is especially noticeable near the pseudocritical

point (the temperature at which the specific heat reaches its peak value for a given pressure).

Figure 2.12: Variations of the specific heat Cp and the density ρ of CO2 at p=80 bar and 100 bar [10].


2.3.2 Safety properties

The Designation and Safety Classifications of Refrigerants of American Society of Heating,

Refrigerating and Air-Conditioning Engineers (ASHRAE) [11] assigns each refrigerant to a

certain safety group, based on the toxicity and flammability.

Table 2.1: ASHRAE 34 Classification.

Low toxicity High toxicity

High flammability A3 B3

Low flammability A2 B2

No flame propagation A1 B1

The ASHRAE 34 classification is indicated using two symbols. The first symbol is a letter, which

indicates the toxicity of the refrigerant. The second symbol is a number, which indicates the

flammability.

A class A refrigerant has not been identified as toxic at a volume concentration level of 400 ppm

or lower. For a class B refrigerant, there is evidence of the refrigerants toxicity.

A class 1 refrigerant does not show flame propagation when tested at atmospheric conditions

(101,3 kPa and 21 C).

A class 2 refrigerant has a lower flammability limit that is higher than 0,10 kg/m3 at atmospheric

conditions and its heat of combustion is less than 19 MJ/kg.

A class 3 refrigerant either has a lower flammability limit that is higher than 0,10 kg/m3 at

atmospheric conditions or its heat of combustion is less than 19 MJ/kg.

For the design of the test set-up, Pcrit is also considered as a property affecting safety. Refrigerants

whose Pcrit is too high will be set aside because of these high pressures.


2.3.3 Environmental properties

Next to the possible toxicity of refrigerants, two other main factors that have an impact on the

environment are evaluated:

First of all, the Ozone Depletion Potential (ODP) is an indicator of the degradation to the ozone

layer, caused by the gas. As a reference for this indicator, the ODP of T-11 is fixed at 1 ·

Secondly the Global-Warming Potential (GWP) is evaluated. This factor compares how much

heat is trapped by a certain mass of the gas to the amount of heat that is trapped by the same

amount of CO2 over a specified time interval.

GWPgas =

∫ T0 agas [gas] dt∫ T

0 aCO2[CO2] dt

(2.2)

In this calculation, agas is the instantaneous radiative efficiency of the specified gas, the value

between square brackets is the concentration of the gas and T is the time horizon.


2.3.4 Working fluids candidates

The different properties discussed are represented in following table. In order to test the behaviour

of different types of working fluids, a wet, a dry and an isentropic fluid will be selected.

Table 2.2: Overview of refrigerants and their main properties.

Name Type Tcrit [C] pcrit [bar]ASHRAE 34

ODPGWP

safety group (100 yr)

PFC-116 Wet 19,88 30,5 A1 0 11900

HFC-23 Wet 26,14 48,3 A1 0 14800

R-747 - CO2 Wet 31,1 73,8 A1 0 1

HC-170 - ethane Wet 32,18 48 A3 0 20

HFC-41 Wet 44,13 59 - 0 92

HFC-125 Wet 66,02 36,2 A1 0 3500

HFC-410A - 70,2 47,9 A1 0 2088

PFC-218 Isentropic 71,89 26,8 A1 0 8830

HFC-143a Wet 72,73 37,64 A2 0 4470

HFC-32 Wet 78,11 57,83 A2 0 550

HFC-407C - 86,79 45,97 A1 0 1800

HC-290 - propane Wet 96,65 42,47 A3 0 20

HFC-134a Isentropic 101,03 40,60 A1 0 1430

HFC-227ea Dry 101,74 29,29 A1 0 3220

PFC-3-1-10 Dry 113,18 23,2 - 0 8600

HFC-152a Wet 113,5 44,95 A2 0 124

PFC-C318 Dry 115,2 27,78 A1 0 10250

HCFC-124 Isentropic 122,28 36,2 A1 0,03 620

HC-600a - iso-butane Dry 135,05 36,47 A3 0 20

HCHF-142b Isentropic 137,11 40,6 A2 0,04 2400

HFC-236ea Dry 139,22 34,12 - 0 1370

PFC-4-1-12 Dry 147,41 20,5 - 0 9160

HC-600 - n-butane Dry 152 37,95 A3 0 20

HFC-245fa Isentropic 154,05 36,4 B1 0 900

HFC-245ca Dry 174,42 39,25 A1 0 693

HCFC-21 Wet 178,33 51,8 B1 0,01 210

HCFC-123 Isentropic 183,7 36,68 B1 0,02 77

HFO - 1234yf Dry 94,7 33,82 A2L 0 4

HC-601 - n-pentane Dry 196,5 33,64 A3 0 20

HCFC-141b Isentropic 204,2 42,49 A2 0,12 725


2.4 Heat transfer to fluids at a supercritical state

Up to now, many data on heat transfer under supercritical pressures have been obtained. The

main focus of previous studies however was on water, CO2 and helium. Few experimental data

for hydrocarbons which are used in ORCs is available. In order to make a thorough analysis

and models for the heat transfer to these fluids, an extensive set of data on the thermophysical

properties and the heat transfer mechanisms to these fluids are needed.

This does not diminish the value of previously performed research however. The peculiarities

observed and the results gained during these studies are most likely also applicable to the working

fluids of interest for the transcritical ORC.

2.4.1 Overview of experimental research on heat transfer to supercritical

fluids

In 2001 M. Bazargan made a very complete review in terms of reviewed documents about

the phenomenon of heat transfer to supercritical fluids[12]. This review is used as a historical

guideline for following discussion.

Back in 1957, Bringer and Smith were the pioneers on experimental research for heat transfer to

supercritical fluids [13]. They found out that existing empirical and semi-theoretical correlations

did not give accurate results. The rapid variations of thermal conductivity, viscosity and density

were identified as the main reasons for the deviation between experimental results and expec-

tations from the correlation. As a final conclusion, existing constant property correlations in

the form of Nu = a ·RebPrc (Dittus-Boelter) were found to be inadequate to predict the heat

transfer coefficient near the critical region.

The first empirical correlation for heat transfer to supercritical fluids was the one from Miropolski

and Shitsman in 1957 [14]. This empirical relationship was the first to fit experimental data

in the supercritical region and required only minor changes to the Dittus-Boelter correlation.

Prmin, the lesser of the Prandtl number at the wall and in the bulk fluid was used instead of

using only the bulk Prandtl number. This correlation however, is limited to fluids with a Prandtl

number around unity.

In order to incorporate case studies where the temperature, thus the fluid properties, varied

noticeable, existing correlations needed to be adapted.

Krasnoshchekov, Protopopov and Petukhov [15]proposed following correlation for forced convec-

tive heat transfer to water and CO2 in supercritical conditions (see section 2.4.3.1).

Relatively speaking, this is still one of the most accurate correlations [12][16].


Most of the experimental studies up to this point considered vertical flow. In 1964, Vikrev and

Lokshin conducted one of the earliest major studies of heat transfer to supercritical water in

horizontal flow [17]. This also was the first attempt to quantitatively formulate deterioration of

heat transfer coefficients in supercritical conditions.

In 1966, Shitsman was one of the first to investigate the effect of density changes (buoyancy)

during horizontal flow [18]. Therefore, he measured temperatures at top and bottom along the

tube section over which heat transfer took place. Temperature differences up to 250 C were

observed, denoting the large possible impact of buoyancy forces.

In order to quantify the significance, the product of the Grashof and Prandtl number was

taken. However, no similar experiments had been conducted before. This meant that Shitsman’s

correlation for buoyancy could not be verified against any existing data but his own. Nevertheless,

his study had a major impact on further work.

It took until 1976 to study the effects of buoyancy and local acceleration due to density variation

more closely. Adebiyi and Hall [19] measured the wall temperature at top and bottom surface of

a tube for a wide range of heat fluxes and mass flow rates. Due to the relatively large diameter

of 22,1 mm of the horizontal tube, free convection could occur.

They noticed that the buoyancy effect does not only cause a heat transfer deterioration at the top

surface of the tube, but also an improvement at the bottom surface. Furthermore, buoyancy-free

and buoyancy-dependent cases were distinguished.

From the late 1970’s and onwards, Jackson and Hall [20] [21] [22] compared several correlations

with experimental data. They also proposed a correlation to take the effect of buoyancy on heat

transfer into account.

The existence of buoyancy in horizontal flows results into a non-symmetry of the flow around the

central axis. This adds to the complexity of both experimental, analytical and numerical research.


2.4.2 Effects of buoyancy and acceleration

Since the primary focus of this research is on a horizontal layout, the effects of buoyancy and

acceleration will be of importance. In this section, these effects are discussed more thoroughly.

In the heat transfer process to fluids above their critical pressure, certain particularities can be

observed. Small variations in fluid pressure and temperature can result in enormous changes in

thermophysical properties, as discussed before in section 2.3.1.4. This variation of properties

results in special features of the heat transfer process.

The rather big density change for an increase in temperature of the working fluid results into

two main phenomena affecting the heat transfer.

First of all, buoyancy or Archimedes forces interact with the fluid, resulting in a temperature

and density gradient in a cross-section of the flow.

Secondly, the overall heating of the working fluid results into a lowering mean density of the

working fluid in the axial direction. In order to maintain a constant mass flow rate, this results

into a thermally driven acceleration of the flow.

The effect of density of the working fluid on the internal flow an thus the heat transfer phe-

nomenon suggest that orientation of the tube should certainly be taken into account. In terms of

the new test set-up at Ghent University, the main primary focus is on horizontal tubes. This

same focus is applied to the remainder of the discussion of the heat transfer phenomenon.

Depending on the tube diameters, the main factor of influence can differ. For large diameter

tubes, the buoyancy effect becomes the main factor of importance, as for small diameter tubes,

the flow acceleration effect becomes very important.


2.4.2.1 Buoyancy effect

Unlike for two-phase flow, there are no different phases, thus different heat transfer mechanisms,

are present in the working fluid.

However, in a horizontal tube there is a certain differentiation between the upper flow layers and

the lower flow layers due to the buoyancy effect. This results into a difference between the fluid

temperatures (and wall temperatures) between the top and bottom sides of the pipe. The hotter,

less dense, fluid raises to the upper section of the pipe, whereas the denser, colder fluid remains

at the lower section.

As stated before, the influence of this phenomenon is biggest for pipes with a relative large

diameter. Even in small diameter pipes, the non-uniformity of the temperature can reach a

substantial value.

The influence of buoyancy also further increases for lower mass flow rates [23].

The result of the temperature difference at the top and bottom due to this buoyancy effect is

noteworthy. In the results of Bazargan et al.[23], using water at supercritical conditions, the

local heat transfer coefficients at the bottom surface are up to 2,5 times larger than those at the

top surface.

It should also be noted that the influence of the buoyancy effect diminishes once the critical

temperature (Tcrit,water = 374 C) is achieved.

(a) Wall and bulk temperature variations

with bulk enthalpy.

(b) Differences of local heat transfer coef-

ficients at top and bottom surfaces.

Figure 2.13: Test results for P = 24,4 MPa, G = 340 kg/m2s, q = 300 kW/m2.

Measurement of buoyancy effect

As suggested above, the temperature difference between the top and bottom surface of the tube

can be used as a measure of buoyancy effect. In order to get accurate results, this temperature


measurement should be first performed with an isothermal flow. In an isothermal flow there is

no density gradient, thus providing a case which can be used to as a reference for the relative

measuring errors.

To have a more qualitative measure of buoyancy, the Grashof number can be used.

Gr =gβ (Tw − Tb) d3

ν2(2.3)

The Reynolds numbers for low flow rates are small. In these conditions, the radial temperature

profile and density gradient become greater. This causes the Grashof number to become greater

to. The dimensionless number Gr/Re2, which is used for measurement of the buoyancy thus

increases too.

Bazargan [12] compared several test cases in order to determine whether the effect of buoyancy

is negligible or not. In case of normal, subcritical pressure flows, buoyancy can not be neglected

for Gr/Re2 > 1. For supercritical flows however, this criterion cannot be used, as can be seen on

figure 2.14.

Table 2.3: Summary of the test cases by Bazargan [12].

q” ≈ 300 kW/m2 G ∆Twall,max Twall,min for Peak HTC bottom Peak HTC

in all cases [kg/m2s] [C] Tw,top = Tw,bottom [kW/m2K] bottom/top

Case I 340 70 670 12 2,75

Case II 432 35 570 15 2,15

Case III 646 12 460 24 1,5

Case IV 965 4 405 37 1,2

Figure 2.14: Variations of Gr/Re2 with bulk enthalpy for various G (q”=300 kW/m2) [12].


On figure 2.14, it can be observed that the buoyancy is still noticeable for values of Gr/Re2

lower than 0,1. The peak values of Gr/Re2 however vary for the different cases. This requires a

more fine-tuned parameter to distinguish buoyancy-affected and buoyancy-free conditions.

Petukhov et al. studied the effect of buoyancy in horizontal flow [24]. A threshold value for the

Grashof number, Grth was derived. Below this threshold, buoyancy effects can be neglected.

Later, Petukhov and Polyakov refined this threshold according to following definition [25]:

Grth = 3 · 10−5Re2,75b Pr0,5[1 + 2, 4Re

−1/8b

(Pr

2/3 − 1)]

(2.4)

In this equation, Pr is defined as:

Pr =iw − ibTw − Tb

· µbλb

(2.5)

With the assumption that q” = λ∂T∂x near the wall, the expression for the Grashof number

(equation 2.3) is modified to:

Grq =gβq”d4

ν2bλb(2.6)

where

β =1

ρfilm· ρb − ρwTw − Tb

(2.7)

Finally Petukhov defined the criterion that buoyancy can be neglected when Grq < Grth.

2.4.2.2 Flow acceleration effect

The fast variation of density results into a flow acceleration of the working fluid. At a continuous

mass flow rate, a lowering of the density results in a raise of volumetric flow rate over the heated

sections. Compared to the buoyancy effect, the acceleration term becomes more important in

tubes with a small diameter. For large diameter tubes, the buoyancy term is dominant.

In between the two extremes, for moderate tube sizes, the acceleration effect can generally be

neglected with respect to the buoyancy effect.

Ankudinov and Kurganov showed that the acceleration effect can cause heat transfer deterioration

[26].


2.4.3 Heat transfer correlations

Over time, several heat transfer correlations have been developed. Most of these are for water,

helium or CO2 or a selection of these fluids. Nevertheless, these correlations can be put to good

use in the design procedure. A comparison of results, given by different correlations, indicates

the order of magnitude for several aspects of the test facility.

In what follows, four different correlations are given. These correlations are implemented in EES

and used in chapter 4 to estimate the heat transfer in the test section for several conditions.

2.4.3.1 Petukhov, Krasnoshchekov and Protopopov

The correlation of Petukhov, Krasnoshchekov and Protopopov [15] for water and CO2 at super-

critical conditions has proven to show a good predicting performance. About 85 % of their data

and data of experiments by other researchers matches their correlation within a deviation of

±15 %.

In order to avoid an extensive listing of different correlations and their application range, only

this correlation will be discussed in more detail, since it was used to give a first estimation for

sizing the test set-up.

Nub = Nu0,b

(Cp

Cp,b

)0,35(λbλw

)−0,33( µbµw

)0,11

(2.8)

Nu0,b =

fb8RebPr

12, 7

(fb8

)0,5 (Pr

2/3 − 1)

+ 1, 07

(2.9)

f = (1, 82log10 (Reb)− 1, 64)−2 (2.10)

This correlation is assumed to be valid within following ranges:

2 · 104 <Reb < 8, 6 · 105

0, 85 <Prb < 65

0, 90 <µbµw

< 3, 60

1, 00 <kbkw

< 6, 00

0, 07 <Cp

Cp,b< 4, 50


2.4.3.2 Swenson, Carver and Kakarala

The correlation of Swenson et al. [27] is applicable for water in horizontal tubes:

Nu =hD

kw= 0, 00459

(DG

µw

)0,923(Hw −Hb

Tw − Tb· µwkw

)0,613

·(ρwρb

)0,231

(2.11)

2.4.3.3 Shitsman

The correlation of Shitsman [28] is applicable for helium, CO2 and water in horizontal tubes and

is of the Dittus-Boelter form:

Nub = 0, 023Re0,8b Pr0,8min (2.12)

In the above expression, Prmin is the smaller one of Prb and Prw.

2.4.3.4 Jackson and Fewster

The correlation of Jackson and Fewster [29] is based upon the Dittus-Boelter form for constant

properties convection:

Nub = 0, 0183Re0,82b Pr0,5(ρwρb

)0,3

(2.13)

The effect of buoyancy is neglected in the equation.


2.5 Existing set-ups for testing heat transfer to SC fluids

2.5.1 Freon Thermal Hydraulic Experimental Loop

Kang et al. [30] performed an experimental study on the heat transfer characteristics during the

pressure transients under supercritical pressures. These experiments have been performed on the

Freon Thermal Hydraulic Experimental Loop (FTHEL).

Goal of the research

Supercritical pressure fluids have already been utilized in the field of fossil-fired power plants.

Previous studies have been focused on the steady-state heat transfer regime to investigate peculiar

heat transfer characteristics and develop the Heat Transfer Coefficient (HTC). A thorough study

for pressure transient conditions have not been performed.

2.5.1.1 Working fluid

Table 2.4: Overview of the working fluids for the FTHEL.

Working fluid Tc pc

R134a 101,08 C 40,59 bar

2.5.1.2 Test set-up

The test set-up used for the experiments is the Freon Thermal Hydraulic Experimental Loop

(FTHEL), as represented in figure 2.15.

The set-up is a closed hydraulic loop in which two pumps, connected in series, assure the flow

through the system. The pumps are non-seal canned motor pumps, that integrate the motor

and pump in a single unit to avoid leakage. The outlet pressure of the pump is stabilized by

accumulators using N2 as a pressure regulator.

Two preheaters are meant to heat the working fluid up to the desired entrance temperature of

the test section.

The test section is a 2000 mm long tube with an inner diameter of 9,4 mm. The test section is

heated uniformly in axial direction by Joule heating. This heating method consists of an electrical

current that is applied to the test section. The electrical resistance of the tube causes the tube

to heat under influence of this current. The power for heating the test section is supplied by a

DC-source at 60 V and a current of up to 12 000 A.

In order to evaluate the heat transfer to the supercritical fluid, the wall temperature is used. 39

Chromel-alumel thermocouples are equally spaced silver-soldered to the outer wall to monitor its

temperature. Ceramic wool and tape are used as insulation material for the test section in order

to avoid heat loss during the experiments.


Figure 2.15: Schematic diagram of the FTHEL facility.

2.5.1.3 Test conditions

The test condition range for the steady state tests are represented in table 2.5.

Table 2.5: Test condition range for the tests on the FTHEL.

Parameter Lower limit Higher limit

Tin 50 C 100 C

p 41 bar 45 bar

Q 10 kW/m2 160 kW/m2

G 600 kg/m2s 2000 kg/m2s

Steady state

The steady state experiments are performed to provide a reliable heat transfer database. These

experimental results will be used as a reference case and are thus performed under a wide range

of experimental parameters.


Pressure transient

The experiments considering the pressure transient are performed under two different conditions:

pressure increase and pressure decrease. During these experiments, the other parameters were

kept at a constant value:

• Constant mass flux

• Constant test section inlet temperature

• Constant applied heat flux

The pressure itself was varied from 3,8 MPa to 4,5 MPa or vice versa at a transient rate from

1,1 kPa/s up to 13,6 kPa/s.


2.5.2 Supercritical test set-up Tsinghua University of Beijing

At the Tsinghua University of Beijing, a test set-up has been built to determine the flow and heat

characteristics for supercritical fluids. The main focus of this set-up lies within the application of

supercritical fluids to use as a ’third fluid’ cooling system for the combustion chamber of liquid

rocket engines, whereas the two other fluids are the fuel and liquid oxygen [31].

2.5.2.1 Working fluids

Table 2.6: Overview of the working fluids for Tsinghua University test set-up.

Working fluid Tc pc

R22 96,14 C 49,9 bar

Ethanol 241 C 63 bar

2.5.2.2 Test set-up

The test set-up is a closed-loop cycle. The working fluid is added to the cycle from a container,

through an accumulator and a filter before being pressurized by a supercritical fluid pump. This

pump boosts the pressure of the fluid up to the desired pressure level, which is closely controlled

by the manostat.

In the preheater, the working fluid is heated to the desired inlet temperature for the test section.

The test section is a vertical stainless steel tube of 152 mm and an inner tube diameter of

1,004 mm in which the working fluid is heated by the use of Joule heating. In order to heat the

working fluid by Joule heating, an electrical current is applied to the stainless steel tube. The

electrical resistance of the tube causes the tube to heat in a axial-uniform manner. The wall

temperature of the tube is measured using 15 equidistant K-type thermocouples. To assure a

uniform distributed temperature in the bulk fluid at the measuring points, mixers are installed

in front of the Pt-100 temperature sensors.

The heated working fluid is cooled down by the sub-cooler after leaving the test section. The

subcooled fluid then flows through a decompression valve, which controls the system pressure

together with the manostat. The decompressed fluid flows back through the accumulator to the

supercritical pump.


Figure 2.16: Schematic of Tsinghua University test set-up.


Table 2.7: Test condition range of the Tsinghua University test set-up.


T 25 C 200 C

p 55 bar 100 bar

Q 110 kW/m2 1800 kW/m2

G 600 kg/m2s 2000 kg/m2s


2.5.3 Supercritical test set-up Seoul National University

In Korea a test set-up for a gas cooling process has been developed in a cooperative project

between the Digital Appliance Research Laboratory of LG Electronics Inc., The School of

Mechanical and Aerospace Engineering of Seoul National University and the Department of

Mechanical Engineering from Korea University [32].

2.5.3.1 Working fluid

Table 2.8: Overview of the working fluids for the Korean test set-up.

Working fluid Tc pc

CO2 31,1 C 73,8 bar

2.5.3.2 Test set-up

The test set-up consists out of four loops:

The closed refrigerant loop, a hot water loop for preheating of the working fluid in the preheat

heat exchanger, a primary cooling water loop for the test section and a secondary cooling loop

using ethylene glycol/water for subcooling of the refrigerant.

In the refrigerant loop, CO2 is added to the circuit by a syringe pump which controls the inlet

pressure of the test section by adjusting the amount of CO2 in the system. The tubes of this

refrigerant loop are made out of copper and have an inner diameter of 7,73 mm.

The experiments are conducted at various mass fluxes and inlet pressures of CO2. In order to

control the mass flux, the magnetic gear circulation pump is driven by a variable speed electric

motor.

A Coriolis type mass flow meter measures the flow rate in the refrigerant loop in order to control

the pump to the desired flow rate.

After exiting the mass flow meter, the working fluid enters the preheaters. The first preheater

is heated electrically, which is controlled manually with a variable voltage source. The second

preheater is a hot water preheater. These preheaters are controlled in such a way that the fluid

temperature rises to the desired inlet temperature for the test section.

In the test section, the working fluid is cooled down in a tube-in-tube heat exchanger. The

working fluid flow through the inner tube and cold water flows through the annulus with an

inner diameter of 16 mm. The test section itself is composed out of eight subsections, which are

connected in series. One of such components is shown in figure 2.17. Each of these units has a

heat transferring length of 470 mm. This adds up to a total heat transferring test section length

of 3760 mm.


At the inlet and outlet of each test section unit, temperatures of CO2 and cooling water are

measured by T-type thermocouples, directly inserted in the flow.

For each unit, the effect of buoyancy on the temperature distribution is measured by three

thermocouples. These are attached at top, side and bottom of the heat exchanging surface of the

inner tube.

Figure 2.17: Schematic of one unit of the test section.


Table 2.9: Test condition range of the Tsinghua University test set-up.


Tin,CO250 C 80 C

p 75 bar 88 bar

GCO2225 kg/m2s 450 kg/m2s

Tin,water 7 C 12 C

Gwater 60 g/s 120 g/s


Figure 2.18: Schematic of the Korean set-up for in-tube cooling of CO2.

Chapter 3

Test set-up description

This chapter gives a general introduction to the new test set-up and describes its working princi-

ples. More detailed information of the design principles and choices can be found in chapter 4.

The new test set-up at Ghent University shares several aspects of the existing set-ups such as

discussed in section 2.5. Of the previously discussed set-ups, the new facility at Ghent University

best resembles the test facility of Seoul National University. However, there are some important

differences.

First of all, the Korean set-up is built to study heat transfer in a cooling regime of the working

fluid whereas the new test facility at Ghent University will be used to study the heating process

of the working fluid.

A second major difference between the Korean set-up and the new set-up of Ghent University is

the incorporation of an expansion process in the design. This was not planned to be included

on beforehand, but became necessary in order to find a suitable pump for circulation of the

refrigerant. The fluid thus does not remain at supercritical conditions throughout the whole

circuit.

3.1 Test set-up cycle

The test set-up cycle can be compared with an actual transcritical power cycle. In the test set-up

cycle however, the expansion process is achieved by means of an expansion valve instead of a

mechanical power generator such as an expander or a turbine.

37

Chapter 3. Test set-up description 38

3.2 Description of the test facility

A first and important part of the set-up was the determination of the working conditions. These

have a major impact on the design of the set-up. Since the research on heat transfer to refrigerants

at supercritical conditions is in an early stage, the choice is made to be able to test a selected

range of conditions. To be able to adapt the installation for testing different layouts of the test

section (e.g. orientation, tube diameter), the design is made modular. This however adds a

certain level of complexity to the set-up.

Finally, the aim is to obtain accurate results from the test set-up. This however conflicts with the

aim for a wide operating range, so certain trade-offs or special measures will have to be made.

The general design choices that have been made are listed below:

• Horizontal set-up

• Pressure: up to 1,2 pcrit,max = 50 bar (R134a)

• Refrigerant mass flux G

– Minimum: 100 kg/m2s

– Maximum: 1000 kg/m2s

• Volumetric flow rate refrigerant (derived from G and ρ(T ))

– Minimum: 0,1536 m3/h (R134a Tin = 50 C)

– Maximum: 1,945 m3/h (R134a Tin = 90 C)

• Test section

– Length: 4000 mm

– Diameter: 1 1/8inch = 28,57 mmOD

– Wall thickness: 1,9 mm

– Inner diameter: 24,77 mm

– Maximum system pressure: 50 bar

– Maximum test section heating power: 10 kW

• Tube-in-tube preheater

– Power: 20 kW

• Electrical preheater

– Power: 10 kW


Most of these values however (such as power, hot fluid flow rate,...) are based on simulations in

EES and have to be handled carefully. In order to compensate for the deviation of the correlation,

a design factor of 20 % is taken into account. If the existing correlations prove to give accurate

results, this design factor results into an over-dimensioning of the test set-up. The spare capacity

can be used to test a broader range of conditions.

The set-up consists out of three main fluid circuits such as represented on figure 3.1:

• A refrigerant circuit, containing the test section.

• A Heating circuit, which is used to heat the refrigerant in the test section.

• A cooling circuit, which removes heat from the refrigerant circuit.

Preheater

Hot water system

Cold water system

CoolerTest section

Cooling circuit Refrigerant circuit Heating circuit

Figure 3.1: Simplified schematic of the fluid loops in the set-up.

Following sections of this document will briefly explain the components and their function for

each of the three fluid circuits, following the process flow diagram shown on figure 3.2 and

appendix B.


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Acc

um

ula

tor

Oil

hea

ter

Figure 3.2: Schematic of the set-up.


3.3 Refrigerant circuit

The refrigerant circuit is the main circuit of the test facility. It contains the working fluid to be

tested. These working fluids are still mainly used as working fluids in refrigeration applications,

thus giving this circuit its name. Due to the recent developments of ORCs, the use of the term

refrigerant for these chemical compounds should be reviewed. Throughout this dissertation

however, the widely accepted term refrigerant will still be used to indicate the working fluid that

is being studied.

3.3.1 Test section

The test section forms the core of the set-up. It is executed as a counterflow tube-in-tube heat

exchanger which is fitted with K-type thermocouples to monitor the temperature over the test

section. This layout is represented on figure 3.3.

The use of a tube-in-tube heat exchanger is quite different from most existing set-ups. However,

it is a better approach to simulate the actual heat transfer such as it would take place in the

actual heat exchangers of an ORC.

In a set-up with an electrically heated test section, the wall temperature would be quite a lot

higher than in practical situations, thus affecting the convective heat transfer coefficient at the

refrigerant side.

Working fluid inletWorking fluid outlet

Heating fluid inlet

Heating fluid outlet

Thermocouples

Figure 3.3: Schematic of the test section with thermocouple measuring points.

Since the fluid is at a supercritical state, the enthalpy of the fluid can be calculated from

measuring the temperature and pressure of the fluid. This is not the case for heat transfer to a

two-phase flow, which also needs a measurement of the vapour quality in order to determine the

enthalpy state.

The inner tube contains the supercritical refrigerant. Hot thermal oil (Therminol ADX10) flows

through the annulus of the heat exchanger to deliver heat to the refrigerant.


By using thermal oil, the actual use of a heat exchanger tube from a shell and tube heat exchanger

of an industrial applied ORC can be simulated. This type of process, using thermal oil, typically

has less problems with the temperature mismatch between hot source and working fluid such as

would be the case with flue gasses. However, the temperature dependency can also be thoroughly

studied by finely varying the inlet temperature of the test section.

Bulk temperature measurements of both refrigerant and heating fluid take place at inlet and

outlet of the test section.

In the test section itself, it is not possible to take bulk temperature measurements of the

supercritical refrigerants.

Three outer wall temperatures of the inner tube and three bulk temperature measurements of

the thermal oil are used to calculate the local heat transfer coefficient (equation 4.3.2.1).

3.3.2 Pressure regulating expansion valve

The pressure in the test section is controlled by a mechanically controlled expansion valve,

which uses the pressure at the inlet of the test section, which is the controlled variable, as input

parameter for its position.

In collaboration with the pump the expansion valve separates the high pressure section of the

test set-up from the low pressure section. The need for a low-pressure section came forward

during the selection process of a circulation pump for this set-up. The inlet pressure of the pump,

best fitting for this application, is limited to 35 bar, which is well below some of the operating

pressures in the test section.

In case of failure of the expansion valve, or if the reaction time of this valve proves to be

inadequate, an electrically operated valve can be used to bypass the expansion valve.

3.3.3 Mechanical pressure relief valve

The pressure relief valve is a mechanical last line of defence in case of a pressure build-up in the

system. In such case, the software control is programmed to lower the system pressure as fast as

possible within the safe operating area. Should this control software fail, the relief valve will

open and refrigerant will flow out of the system into a buffer vessel that is used to capture the

excess refrigerant.

3.3.4 Low pressure section shut-off valve

This valve ensures that the pressure at inlet of the circulation pump and expansion accumulator

does not raise above 35 bar and is operated by means of an electronically controlled pilot valve.


3.3.5 Cooler/Condenser

The cooler/condenser is a heat exchanger that cools down the refrigerant to a temperature just

below entrance conditions of the test section or condenses the expanded working fluid to a liquid

state, depending on the operational conditions in the test section.

It ensures that the condition of the working fluid at the pump inlet is compliant with the

prescribed conditions.

3.3.6 Accumulator

The accumulator in the low pressure section of the refrigerant circuit is required to compensate

the working volume of the circuit when heating up the working fluid from ambient conditions to

the actual operating conditions. During this process, the density of the working fluid decreases

significantly, which would result into a high pressure rise if no expansion volume is available.

3.3.7 Circulation pump

The circulation pump is responsible for the circulation of working fluid through the system. It is

also used to pressurize the working fluid to the pressure that is desired in the test section.

3.3.8 Tube-in-tube preheaters

The tube-in-tube preheaters use the rundown of the thermal oil from the test section to preheat

the working fluid going to the test section. The test section and the two tube-in-tube preheaters

can be considered as one tubular heat exchanger, which is split up into three parts to save space

and break up boundary layers within the tubes.

Each of the two preheaters can be bypassed by means of a manual three-way valve. Whether

a preheater should be bypassed or not depends on the working conditions in the test section.

These preheaters are only controlled by the three-way valve which results into an on/off-control

of each of the preheaters

3.3.9 Electrical preheater

The preheater is mainly used during the start-up of the testing facility and for more fine control

of the preheating than what is possible with the tube-in-tube preheaters.

At start-up, the refrigerant is at ambient temperature level. By circulating it through the

electrical preheater, the time to get to a thermal regime can be lowered significantly.

Once the required conditions of the working fluid are achieved, the duty of the electrical preheater

will be lowered. From then on, the preheater assures the desired entrance temperature of

the test section is achieved and it takes up a heating duty between 0 W and 10kW, which is

PID-controlled.


3.4 Heating circuit

In order to achieve the required temperatures in the test section, the temperature of the heating

fluid should be higher than 100 C. Therefore it is not possible to use a hot water circuit at

atmospheric pressure. Two possible solutions were considered:

A first possibility would be the use of a pressurised hot water circuit. This however complicates

the design of the heating circuit.

A second possibility is the use of thermal oil as a hot fluid. This oil can be easily heated using

electrical elements.

The use of a thermal oil heater is certainly preferable, since this can work at unpressurised

conditions. This simplifies the installation of the test set-up in the laboratory.

Electrical heater

Circulation pump

Figure 3.4: Schematic layout of the heating circuit.

By using an integrated solution of a thermal oil heater and oil circulation pump, the thermal oil

unit can easily be used on other set-ups in the laboratory.


3.5 Cooling circuit

The cooling circuit provides the cooler with a cold mixture of ethylene glycol and water. The

cooler cools down the refrigerant to a state just below entrance conditions for the test section.

The layout of this subsystem is represented in figure 3.5. Three key components of this subsystem

are the chiller, the buffer vessel and the pump. The cooler/condenser itself is considered as a

part of the refrigerant circuit.

Cooler/Condenser

Chiller

Buffer vessel

Figure 3.5: Schematic layout of the cooling circuit.

The cooling circuit is already available in the laboratory and is shared with other set-ups. This

predefines the specifications of the cooling circuit. The chiller has a capacity of 37 kW and the

buffer vessel has a content of 900 l.

Chapter 4

Design Procedure

The previous chapter gave a brief overview of the complete set-up and its working principles. In

this chapter, the design procedures and certain design choices will be elaborated. This makes it

easier to understand why certain solutions were applied.

The structure of this chapter is based on the chronological order of design, since certain design

choices have influence on the later choices. A functional differentiation as in chapter 3 will also

be followed.

First of all, the boundary conditions due to the selection of working fluids are listed, then the

actual sizing of the equipment is done.

The sizing of the equipment are based on the operational conditions in the test section itself.

By varying the operational conditions, an estimation for the maximum heat transfer in the test

section can be calculated. This heat transfer rate is used to calculate the necessary dimensions of

the remaining components of the installation (e.g. buffer vessel volume, heating power, cooling

power,...)

47

Chapter 4. Design Procedure 48

4.1 Determination of test conditions

The initial phase of the design starts with the determination of the test conditions, as already

stated in section 3.2.

Table 4.1: Overview of the refrigerant properties.

Property HFC-134a HFC-125 HFO-1234yf

Chemical formula CF3CH2F C2HF5 C3F4H2

Tc 101,03 C 66,02 C 94,7 C

pc 4060,3 kPa 3618 kPa 3382,2 kPa

ODP 0 0 0

GWP (100y) 1430 3500 4

These working fluid properties determine the general test conditions and boundary conditions

for the design. The maximum working pressure in the test section is set at about 125 % of the

maximum critical pressure of the working fluids that are going to be used. This results into a

design pressure of 50 bar.

The maximum working temperature will be achieved in the test section itself. Determination

of the mass flow rate is done more arbitrary, based on the velocity of the refrigerant flow in

the piping and the cross-section of the test section. A mass flux ranging from 100 kg/m2s to

1000 kg/m2s is pursued.

Table 4.2: Overview of the design working conditions.

G 100 -1000 kg/m2s

T 20 - 120 C

p 0 - 50 bar

The value range for T and p are fixed for the selection of the different components. The mass flux

working range however is subjected to limitations of the components in the system. Although

this range is wanted to be kept as wide as possible, it might have to be narrowed down.


4.2 General required calculations

In order to do the calculations for the test section, preheaters and cooling water circuit a few

general calculation methods are discussed first. These contain the calculation of pressure drop

and flow rate control using mixing valves. The actual calculation methods and design procedures

follow later in this section.

4.2.1 Pressure drop characteristics

In order to select proper valves and pumps, the flow rate through the system and the correspond-

ing pressure drop are calculated. Therefore, the pressure drop characteristics for the system are

calculated.

These pressure drop characteristics give the pressure drop corresponding to a certain flow rate.

For a straight tube, following equation has to be used:

∆p = 4fL

Dh

ρv2m2

(4.1)

In this expression, vm is the mean flow velocity in the tube, Dh is the hydraulic diameter and f

is the friction coefficient.

The mean velocity can easily be calculated from the flow rate and the cross section of the tube:

vm =Q

A(4.2)

The formula for the friction factor depends on the flow type.

For a turbulent flow (Re > 2300), f is calculated using the Fanning formula:

f = (1, 58ln(Re)− 3, 28)−2 (4.3)

For laminar flow (Re < 2300):

f =16

Re(4.4)

The pressure drop is further increased by the presence of bends, valves and T-pieces. In order to

use equation 4.1, the equivalent length of these line components can be used.

Table 4.3: Equivalent length for common line components.

Component Le/D

90 deg turn 32

Ball valve: full bore 3

Ball valve: reduced bore 25

T-piece: line flow 20

T-piece: branch flow 67


4.2.2 Mixing valve characteristics

As stated in the description, the flow rate of thermal oil to the test section and the flow rate of

cooling mixture to the cooler are controlled by a mixing valve and bypass. This allows that part

of the total flow goes through the bypass without any heat transfer occurring.

The amount of bypassed flow is controlled by a three-way mixing valve. This type of valve only

allows flow going from 1 to 3 and from 2 to 3, as represented on figure 4.1

3

1

2

Figure 4.1: Installation of mixing valves for bypass application.

In order to select the proper mixing valve, the pressure drop characteristics are studied using the

formulas that are formulated in section 4.2.1.

The pressure drop characteristic of the valves is expressed as a Kvs-value. This value expresses

the flow rate in m3/h that correlates with a pressure drop of 1 bar over the valve.

In order to dimension the mixing valve, the valve authority is used. This property is expressed

as:

a =∆p1

∆p1 + ∆p2(4.5)

Where ∆p1 is the pressure drop over a full open valve and ∆p2 is the pressure drop over the

other components of the circuit when the valve is full open.

In an ideal situation, the pressure drops over the valve and over the remainder of the circuit are

equal. This situation would result into a = 0, 5.

Using figure A.1, the appropriate valve is selected. In most cases, the pressure drop of the valve

does not match the pressure drop of the remainder part of the circuit. This would result in a

valve authority that is either larger or smaller than 0,5. In terms of controllability, the smaller

valve should be selected, resulting into a higher valve authority.


The characteristic of the valve is divided into two parts: from 1 to 3 and from 2 to 3.

The characteristic from 1 to 3 is logarithmic and the one from 2 to 3 is linear.

1+2 3

1 3

2 3

Figure 4.2: Mixing valve characteristics.

Specific dimensioning of the valves used in the set-up is described in respective sections. For the

thermal oil circuit, the selection procedure can be found in section 4.4.3. For the cooling water

circuit, the selection procedure can be found in section 4.5.3


4.3 Refrigerant circuit

For the refrigerant circuit, which contains the working fluid that is being studied, the choice of

the tube diameter has a large influence on the further design. Because of the practical application

of the heat transfer process in SC heat exchangers for ORCs, a diameter of 1 1/8 inch is chosen

to be the maximum diameter for which the parameter range from table 4.2 is completely applicable.

In order to determine the influence of the diameter on the heat transfer process, the actual

diameter in the test section will be varied during the study of the heat transfer correlation.

The variation will be limited to practical applicable diameter range for the heat exchangers.

Should the interest for the study of larger diameters raise, the mass flux will be limited due to

the limited volumetric flow rate of the circulation pump.

4.3.1 Material selection

In terms of material selection, a trade-off between several aspects has to be made.

For experimental accuracy, the thermal conductivity of the tube material should be as high as

possible. This assures that the heat transfer from the hot fluid to the working fluid is not limited

by the tube material, but by the convection coefficients.

Secondly the safety aspect of the test set-up is of major importance. The tubing will have to

withstand the pressure of the working fluid at the elevated temperatures.

4.3.1.1 Material selection procedure

Copper

Copper has an excellent thermal conductivity, which is preferable in the experimental set-up,

since it will increase the relative proportion of the thermal resistance of convection in the tube

to the total thermal resistance. Despite of its superior thermal conductivity in relation to the

other materials, the safety aspect of copper can not be guaranteed. This resulted in the choice

not to use standard copper tubing.

Stainless steel

The second material of choice is stainless steel. In contrast to copper, the mechanical properties

of the material are superior for use in high-pressure conditions. The thermal conductivity of the

material however, is a factor 22 lower.

Copper alloys

In order to combine the mechanical properties of the different materials, alloys should be taken

into account. Since copper was most beneficial for conductivity and machinability, copper alloys

are preferred. These alloys have better mechanical properties for use at high pressures and

elevated temperatures.


4.3.2 Test section design

The main goal of the test set-up is to determine the convective heat transfer coefficient for the

working fluids of transcritical ORCs at their supercritical conditions.

This section includes the calculation method for the heat transfer in the test section. To be able

to use this calculation method, several factors have to be set:

• Test section length

• Inlet temperature working fluid

• Temperature resolution of the test section (∆T )

• Inlet temperature of the heating fluid

• Working fluid mass flow rate

• Heating fluid mass flow rate

The length of the test section is fixed at 4 m. This length is set arbitrary in order to have a ∆T

over the test section that is as large as possible, whilst keeping the dimensions of the complete

set-up within a reasonable range. The large ∆T over the test section is preferred in order to get a

low relative measuring error. The test section is a tube-in-tube heat exchanger with counterflow

flow regimes in central pipe and annulus, as represented on figure 4.3.

Figure 4.3: Schematic representation of the flow in the test section.

The temperature resolution on the test section determines the temperature difference of the in-

and out-flowing working fluid.


As discussed more extensively in the previous chapter, the test section is constructed as a

counterflow tube-in-tube heat exchanger.

To determine the heat transfer coefficient U, following equations will be used.

Q = U ·A · F ·∆TLMTD (4.6)

1

UA=

1

hiAi+

ln

(dudi

)2πλL

+1

Auhu=

∆TLMTD

Q(4.7)

In equation 4.6, F = 1 due to the counterflow set-up.

In equation 4.7, there is no mention of fouling coefficients since the test set-up will be a closely

controlled environment. Furthermore, all coefficients except for hi, the convective heat transfer

coefficient at the inner tube, are known or can be measured. This equation will be used to

calculate hi.

4.3.2.1 Calculation of heat transfer in the test section

To be able to size the complete test set-up, a first estimation of the heat transfer in the test section

is needed. This heat transfer differs for different inlet conditions (inlet temperatures, flow rates,...).

In these simulations, the flow rates and the inlet temperatures of the working fluid and the heating

fluid will be specified in advance. The outlet temperatures and net heat transfer will be calculated.

Calculations are implemented in Engineering Equation Solver (EES). The complete EES-script

can be found in appendix E.1, where the calculations are done for R125, with an inlet temperature

of 70 C at a pressure of 1, 05 · pcrit and an oil flow of 10 m3/h with an inlet temperature of

130 C.

Initial guesses for the different parameters in the equation system are not represented in this

appendix, but have to be included in the actual script. Without good initial guesses, EES tends

to have difficulties with converging the numerical solution. Especially near the critical point, a

good initial guess of the values and heat transfer rates is of importance due to the rapid property

changes of the working fluid within this region.


Internal convection coefficient

Based on the existing correlations, as discussed in section 2.4.3, the internal heat transfer

coefficient hi can be found.

For the inner tube, following thermodynamic dimensions are defined:

Hydraulic diameter:

dhyd = di (4.8)

Thermal diameter:

dt = di (4.9)

The results of the different correlations for the internal convection coefficient have to be consid-

ered, since there is no indication yet that one of them results into the correct convection coefficient.

By using the results of a range of different correlations and adding a safety factor to its calculated

value, a numerical range in which the actual value of the internal convection coefficient will lie,

can be found.

The heat transferred between two points is:

Q =Twall − Twf

Rtot(4.10)

The total thermal resistance Rtot is the sum of the conductive resistance of the tube Rcond and

the convective resistance Rconv for convective heat transfer from tube to the working fluid.

Rcond =ln(Do/D)

2λD (4.11)

Rconv =1

hiπDL(4.12)

The equations above can be rewritten to following form, that expresses the internal convection

coefficient as a function of the temperature difference between working fluid and outer wall

temperature and the local heat transfer.

1

hi=Twall − Twf

q− ln (Do/D)

2λD (4.13)


External convection coefficient

The flow in the outer tube is turbulent, thus the Gnielinksi correlation [33] can be used.

In order to simplify the expressions, following parameters are introduced.

Hydraulic diameter:

Dhyd =4A

Pwet=

4πD2

i − d2o4

π(do +Di)(4.14)

Thermal diameter:

Dt =4A

PHT=

4πD2

i − d2o4

πdo(4.15)

The Gnielinski correlation yields:

f = (1, 58 · ln(Re)− 3, 28)−2 (4.16)

Nu =

f

2(Re− 1000)Pr

1 + 12, 7

√f

2

(Pr2/3 − 1

)(

1 +

(Dhyd

L

)2/3)

(4.17)

In equation 4.17, the correction factor can be neglected because the length L is much larger than

the hydraulic diameter Dhyd.


4.3.3 Circulation pump

In contrast to an actual ORC-loop, this circuit does not require different pressure levels. This

means that the pump would only have to compensate for pressure losses. These are calculated in

section 4.2.1.

In order to disturb the measurements in the test section as less as possible, some additional

requirements are set for the pump:

• Sealing of the refrigerants

• Flow rate control → volumetric pump

• Minimal pulsations

4.3.3.1 Gear pump

The specifications required from the pump resulted into the choice of a gear pump for the

circulation pump, since it fulfils all requirements quite well.

Sealing of the refrigerants can also be achieved by using a magnetic coupling.

However, the use of refrigerants at supercritical conditions resulted into some additional challenges:

Static pressure

Even though the ∆P over the circulation pump is quite small, the static pressure on the housing

of the pump will be within a range up to 50 bar.

Viscosity

At supercritical conditions, the viscosity of the working fluids will lower to a level of Pa · s. In

case of a magnetic coupled gear pump, this viscosity is too low to guarantee proper lubrication

of the pump’s inner components.

Those two major disadvantages make it a lot harder to find a suitable solution for the test set-up.

Therefore, other options will have to be discussed as well.

4.3.3.2 Diaphragm pump

Due to the problems of finding a suitable gear pump for the use as a circulation pump in the

set-up, other options are also reviewed.

The low viscosity specification in combination with adequate sealing resulted into the choice of a

diaphragm pump. The pump of choice is a Hydra-Cell G15 diaphragm pump with 5 integrated

diaphragms.


The use of 5 diaphragms limits the peak-to-peak pulsations in the flow (figure 4.4). This is

beneficial for overall vibrations in the system, and especially for the lifetime of the brazed

components, which are subjected to fatigue.

t

p

∆p

Figure 4.4: Pressure pulsations with 5 shifted diaphragms.

This pump can cope with the different requirements for the test set-up. However, the use of this

pump also results into some additional restrictions:

• Pump oil temperature limited to 80 C

• Inlet pressure limited to 35 bar

Pump oil temperature

The limitation on the oil temperature is easily solved by implementing an oil cooler on the pump.

Inlet pressure

The limitation of the pump inlet pressure to 35 bar adds additional complexity to the set-up:

• An expansion valve is installed, which results into two pressure sections of the circuit: A

high pressure section from the pressure side of the pump up to the expansion valve and a

low pressure section from the expansion valve up to the circulation pump.

• For R134a and R125, the temperature-pressure combination results into fluid in the two-

phase region. This means that the cooler will have to be replaced by a condenser with a

higher capacity in order to prevent gas bubbles being trapped in the pump.


4.3.3.3 Pump conditions

Net Positive Suction Head

The conditions at the inlet of the pump are of major importance to avoid cavitation. The

conditions can be described in terms of the Net Positive Suction Head (NPSH).

The Net Positive Suction Head required (NPSHr)is 4,5 m. In order to avoid cavitation, the Net

Positive Suction Head available (NPSHa) should be at least 5,5 m and is expressed as:

NPSHa =

(piρg

+V 2i

2g

)− pvρg

(4.18)

with pi the pressure and Vi the velocity at the pump inlet and pv the vapour pressure.

Table 4.4: NPSHa for different conditions

R134a T [C] pv [bar] p [bar] ρ

[kg

m3

]G[kg/m2s

]v [m/s] NPSHa[m]

Charging 20 5,721 7 1243 1000 0,45 10,50

Testing 93 34,52 35 807,7 1000 0,70 6,08

R125 T [C] pv [bar] p [bar] ρ

[kg

m3

]G[kg/m2s

]v [m/s] NPSHa[m]

Charging 20 12,06 14 1221 1000 0,46 16,21

Testing 60 31,71 35 916,1 1000 0,61 36,63

R1234yf T [C] pv [bar] p [bar] ρ

[kg

m3

]G[kg/m2s

]v [m/s] NPSHa[m]

Charging 20 5,917 7 1110 1000 0,51 9,96

Testing 90 30,3 33,8 799,1 1000 0,70 59,98

Furthermore, the pre-pressure at inlet of the pump is only allowed during operation of the pump.

When the pump is not in use, the pressure has to be removed. If a pressure should remain, the

diaphragms can be unbalanced or even deform.

This is realised by isolating the pump from the remainder of the system and only draining the

isolated section of the pump and flexible hoses.


Velocity and turbulence

The maximum allowed liquid inlet velocity of the pump is 0,9 m/s, with as least turbulence as

possible.

The calculation of the inlet velocity is derived from the maximum mass flux (G = 1000 kg/m2s)

in the test section. By making use of the density ρ at the pump inlet conditions, such as defined

in table 4.4 and the different flow sections of the test section and pump inlet line.

vinlet =G

ρ

ATS

Ainlet(4.19)

=G

ρ

πD2TS

4πD2

inlet

4

(4.20)

In this calculation, the maximum diameter of the test section (1 1/8”) and a suction line with a

diameter of (1 1/2”) are used to obtain the numerical results in table 4.4.

Vibrations

In order to separate vibrations that are generated by the pump from the rest of the system,

flexible hoses are used at inlet and outlet of the pump.

4.3.4 Mass flow meter

The mass flow meter is a Coriolis type meter, of which the working principle is explained more

extensively in section 5.4. The measuring device is a tube shaped as an Ω, but is enclosed within

a casing.

For fluids at supercritical conditions, there is no real indication of how it should be mounted.

For liquids, the system is mounted with the Ω inverted, so with the enclosure underneath the

piping. For gasses, the device is turned upside down, so the enclosure is above the piping.

The reason for these orientations is to avoid trapping of non-typical fluids (gasses and liquids

respectively) in the Ω-shaped tube.

For supercritical fluids, the behaviour is more difficult to anticipate. This is being evaluated when

doing test runs on the installation. If the mass flow meter indicates a flow when the circulation

pump is shut down, the orientation of the mass flow meter clearly does not fit the preferred

conditions for supercritical fluids.


4.3.5 Preheaters

Due to the limitation of pump inlet pressure, the working fluid is being condensed before entering

the circulation pump. This results into a large heat loss from the working fluid.

In order to achieve the desired inlet temperature of the test section, the working fluid is preheated

in two different stages: once in a tube-in-tube heat exchanger and once in an electrical preheater.

4.3.5.1 Tube-in-tube heat exchangers

Two tube-in-tube preheaters heat up the working fluid once it is pressurised again. These

preheaters use the rundown of thermal oil coming from the test section as a heat source.

In order to provide a controllability of the heat input, each of these preheaters can be bypassed

at the thermal oil side. This results into an effective preheater length of either 0 m, 4 m or 8 m.

Depending on the actual conditions of the refrigerant, each of those tube-in-tube preheaters have

a heating capacity of up to about 10 kW.

In order to guarantee the desired oil inlet temperature in the test section during measurements,

the flow rate has to be lowered, or one of the tube-in-tube preheaters has to be bypassed. At

this point, a preheater that can be controlled more precisely is required.

4.3.5.2 Electrical in-line heater

The electrical preheater is used in order to control the inlet temperature of the working fluid, en-

tering the test section, more accurately than what is possible by using the tube-in-tube preheaters.

At 10 kW, the maximum heating capacity of this electrical preheater is the same as one of the

tube-in-tube preheaters. This enables to control the effective total heating capacity to any level

between 0 W and 30 kW. Nonetheless, during measurements, part of this heating capacity is

located in the test section itself and one of the tube-in-tube preheaters is bypassed.


4.3.6 Pressure control systems

4.3.6.1 Motorized expansion valve

The pressure in the test section is controlled by a motorized expansion valve. The inlet pressure

of the test section is the main operational parameter for control of this valve.

A motor operated valve has been chosen since this type of valve can keep the valve open, inde-

pendent of the flow conditions.

Selection of the valve is done using DIRcalc software that is provided by Danfoss. An extraction

of this software for the selection of the expansion valve can be found on figure 4.5.

Figure 4.5: DIRcalc valve selection software.

The value for Max. load info should always be kept above 10 % throughout the complete test

range in order to guarantee proper functioning of the valve. In the figure above, this results into

the choice of a ICMTS20-A for the given conditions.


For the other operating points, following data can be found:

Table 4.5: Expansion valve selection data through DIRcalc.

Refrigerant R134

Mass flow 230 kg/h

Product Max. load info [%]

ICMTS20-A 50,8

ICMTS20-B 8,75

ICMTS20-B66 13,2

Refrigerant R134

Mass flow 1800 kg/h

Product Max. load info [%]

ICMTS20-B 68,5

ICMTS20-B66 103

ICMTS20-C 41,4

In case of a sudden pressure rise in the high pressure section, the valve is steered completely

open. If the pressure doesn’t drop quick enough, this might indicate improper functioning of

the valve. An additional electrically controlled valve provides a bypass of the expansion valve,

connecting the high pressure circuit with the low-pressure circuit and the pressure relief valves.

4.3.6.2 Low pressure section shut-off valve

The shut-off valve provides a mechanical separation between the high- and the low pressure

sections of the refrigerant loop. The main function of the valve is to protect the low pressure

section from overpressure (at the accumulator and inlet of the pump). Should the bypass of the

expansion valve open, this valve is also closed to prevent pressure build-up in the low pressure

section.

In contradiction to the expansion valve and its bypass, this shut-off valve is not mechanically

operated. A coil operated pilot is used, which further steers the valve open or close. This means

that a minimum load is required in order to open the valve and keep it open.

This is a deliberate design choice to protect the installation in case of a small leak of the expansion

valve, resulting in a pressure rise of the low pressure section. A leakage flow however, will not be

large enough to open this shut-off valve.

The use of an electrically operated coil ensures controllability through the control software and

hardware.

Selection of the best fitting valve for this application is also done using the DIRcalc software.


Table 4.6: Shut-off valve selection data through DIRcalc.

Refrigerant R134

Mass flow 230 kg/h

Product Min. rec. load [%] ∆pmax [bar]

ICS25-5 52,9 0,142

ICS25-10 100 0,121

Refrigerant R134

Mass flow 1800 kg/h

Product Min. rec. load [%] ∆pmax [bar]

ICS25-5 6,77 1,38

ICS25-10 13,9 0,325

In case of an ICS25-10, the mass flow rate of 230 kg/h is still feasible to get the valve opened.

However, in the future, smaller diameters in the test section will be studied, with the same mass

flux range, this will result into lower mass flows. Therefore the ICS25-5 is a better solution,

keeping future testing ranges in mind.

At the maximum flow of 1800 kg/h however, the pressure drop over an ICS25-5 becomes quite

large with 1,38 bar. The influence of this pressure drop on the system is less important because

the valve is placed in the low-pressure section.

4.3.6.3 Pressure relief valves

The pressure relief valve provides a pure mechanical safety mechanism, should all other mecha-

nisms (software or combination of software and hardware) fail.

The relief valve is placed in the section between the expansion valve and the shut-off valve and is

set to a pressure of 40 bar.

Figure 4.6: Pressure relief valve.

On figure 4.6 the relief valve is connected to the working fluid circuit by the connection on the

left and connected to a gas container by the connection on the bottom. This container captures

the working fluid, should any by evacuated through the valve.


4.3.7 Cooler/Condenser

The heat that’s added in the test section has to be removed from the refrigerant circuit in order

to get to the desired inlet conditions.

However, for some combinations of working fluid and inlet conditions of the pump, a full con-

densation of the working fluid will be necessary (e.g. for R134a). The plate heat exchanger has

been calculated for the maximum mass flux and test section outlet temperature. Cooling of the

working fluid due to expansion is not taken into account, in order to provide enough cooling

capacity.

The final calculations are executed by Alfa Laval (Appendix A.1). From these calculations, and

the pressure drop characteristics of the cooling water tubing, the characteristics of the pump can

be found.

The choice for the condenser is a CBXP27 plate heat exchanger from Alfa Laval with 45 plates.

At its working temperature, the heat exchanger is able to work with pressures up to 90 bar.

4.3.8 Accumulator

The main function of the accumulators is to account for the volumetric expansion of the refrigerant

from start-up of the experimental facility to test conditions.

The accumulator is a pressure vessel containing two fluids: The working fluid and nitrogen (N2),

separated by a diaphragm. Figure 4.7 shows a cross-section of the accumulator. The nitrogen

is situated in the upper part of the accumulator and the working fluid in the lower part. The

nitrogen has a certain pre-charge pressure, which determines the possible volumetric expansion

and end-pressure.


P2

2V

C

P1

1V

B

P0

0V

A V

Figure 4.7: Stages of the accumulator [34].

Table 4.7: Legend for symbols used on figure 4.7.

V0 Nitrogen capacity of the accumulator

P0 Initial preload of the accumulator

V1 N2 gas volume at the minimum pressure

P1 N2 gas pressure at the minimum pressure

V2 N2 gas volume at the maximum pressure

P2 N2 gas pressure at the maximum pressure

∆V Stored volume of working fluid between P1 and P2

The volumetric expansion capacity of the accumulator should be rather high. Near critical

conditions, the density of the working fluids varies rapidly with minor temperature changes. This

local change in density has a major influence on the complete refrigerant circuit.

In order to determine the necessary volumetric expansion capacity, the refrigerant circuit is

divided into two parts: a high pressure part, which is at testing pressure; and a low-pressure

part, which is at the inlet pressure of the circulation pump.

The volume of these circuits mainly consists of the tube volume, which can easily be calculated.

Through symmetric location of pump and expansion valves, these volumes can be taken equal.

V =l

2· π ·D

2

4(4.21)

= 12, 5 · d2iπ

4(4.22)

= 6 l (4.23)


To account for additional volume of the refrigerant section, caused by line components such as

the mass flow meters, heat exchangers, couplings,... this value is rounded up to 10 l for each of

the two pressure sections.

The most extreme variations in volume will be between charging and the upper testing temperature.

These conditions determine the necessary volumetric expansion.

Charging

For the high-pressure section, the present mass is calculated at testing pressure and ambient

temperature:

mc,hp = V · ρc,hp (4.24)

For the low-pressure section, the present mass is calculated at pump inlet pressure and ambient

temperature. The initial pressure is determined by the pressure ratio of the accumulator (fixed

pmax = 35 bar limited by the pump) and by the refrigerant properties, since the refrigerant has

to be in the liquid phase.

mc,lp = V · ρc,lp (4.25)

From equations 4.24 and 4.25, the necessary mass of working fluid is calculated.

mwf = mc,hp +mc,lp (4.26)

Testing

For the high-pressure section, the present mass is calculated at testing pressure and testing

temperature:

mt,hp = V · ρt,hp (4.27)

The remaining mass of working fluid is located in the low-pressure section:

mt,lp = mwf −mt,hp (4.28)

From this mass and the average temperature in the low-pressure section, the necessary volume

of this section is calculated:

Vtest,crit =mtest,lp

ρtest,lp(4.29)

=mwf −mtest,hp

ρtest,lp(4.30)

∆Vexp = Vtest,lp − Vcharge,lp (4.31)

The results of these calculations are displayed in table 4.8 for each of the refrigerants.


Table 4.8: Volumetric expansion of working fluid.

R134a Charging T [C] p [bar] ρ

[kg

m3

]V [l] m [kg]

High pressure 20 40,6 1243 10 12,43

Low pressure 20 7 1226 10 12,26

R134a Testing critical T [C] p [bar] ρ

[kg

m3

]V [l] m [kg]

High pressure 115 40,6 221,1 10 2,21

Low pressure 93 35 807,7 27,8 22,48

R125 Charging T [C] p [bar] ρ

[kg

m3

]V [l] m [kg]

High pressure 20 36,2 1247 10 12,47

Low pressure 20 14 1221 10 12,21

R125 Testing critical T [C] p [bar] ρ

[kg

m3

]V [l] m [kg]

High pressure 75 36,2 271 10 2,71

Low pressure 60 35 916,1 24 21,97

R1234yf Charging T [C] p [bar] ρ

[kg

m3

]V [l] m [kg]

High pressure 20 33,8 1119 10 11,19

Low pressure 20 7 1110 10 11,10

R1234yf Testing critical T [C] p [bar] ρ

[kg

m3

]V [l] m [kg]

High pressure 105 33,8 221,7 10 2,21

Low pressure 90 33,8 799,1 25,13 20,08

The results of table 4.8 clearly indicate the effect of heating of the working fluid on its volume

at constant pressure. This shows the importance of the volumetric expansion capacity of the

accumulators.

For all working fluids, the charging pressure of the low pressure circuit is kept as low as possible.

This is limited by either the compression ratio of the accumulator, or by the saturation conditions

of the working fluids. The latter is the case for R125, which has to be pressurized up to 14 bar

at 20 C to get to the liquid conditions.

For R134a and R1234yf, the minimum pressure is 7 bar. This is derived from the maximum inlet

pressure of 35 bar from the pump and the maximum pressure ratio of the accumulator.

Based on the above data and formula 4.31, an expansion capacity of 16 l should be available.


Accumulator size in litre

ALC

stored volume

Basic sizing chart for accumulator used in energy storage.

Figure 4.8: Selection diagram for accumulator [34].

On figure 4.8, the selection procedure for the accumulator is visualized.

On the left side, the selection line starts of at the pressure ratio of 5 and goes up to the isothermal

curve, where it continues horizontally from its intersection point with this isothermal curve.

On the right side, the selection line starts of at the stored volume ∆V of 16 l.

The intersection point of the two curves indicates that an accumulator of 24,5 l does fit the

required conditions. However, in order to be on the safe side and to have a wider working range,

an accumulator of 32 l is being used in the set-up.


4.4 Heating circuit

The heating circuit provides hot thermal oil that is used in multiple parts of the set-up. First

of all, it is used in the test section which is constructed as a tube-in-tube heat exchanger with

thermal oil in the annulus. The remaining heat in the thermal oil, once it has left the test section,

is being used in the tube-in-tube preheaters.

The use of these preheaters is controlled by two three-way valves. So the preheating length can

be varied from 0 m to 8 m in steps of 4 m.

4.4.1 Heating fluid selection

Table 4.9: Properties of different heating fluid candidates at 90 C.

Density Thermal conductivity Heat capacity Dynamic viscosity

[kg/m3] [W/m ·K] [J/kg ·K] ·10−3[Pa · s]Water 965 0,6613 4204 0,3145

Therminol 66 961,1 0,114 1803 4,55

Therminol ADX10 808 0,115 2180 1,26

Therminol ALD 821 0,108 2170 4,14

Based on this data, a choice of thermal oil has to be made. Following properties are required:

• High density in order to lower the volumetric flow rate

• High thermal conductivity

• High heat capacity

• Low dynamic viscosity

Based on these requirements, Therminol ADX10 is the best fluid of choice.

4.4.2 Main heating unit

The heating circuit is an integrated solution from Vulcanic. This unit includes an oil circulation

pump, a heat exchanger, buffer vessel, thermostats and expansion vessel.


P T1MN

B1

ECH

P

UtilisationUsing

UtilisationUsing

TH1TC2

PA

NB

AC

BCV

R5

P = Motor pump assemblyE C H = Heat exchangerPA = Automatic air bleedingT H1 = S afety thermostatP T 1 = Temperature sensorT C 2 = Limiter thermocoupleAC = TankNB = Low level switchR 5 = Level tapB 1 = Draining bushB C V = S peed breaking boxMN = Manometer

Figure 4.9: Schematic layout of the thermal oil heating unit.

4.4.3 Flow rate control

The mass flow rate of the thermal oil is monitored using a Coriolis mass flow meter. The working

principle of this instrument is described more extensively in section 5.4

The flow rate to the test section and the preheaters is regulated by means of a bypass. This

bypass allows to guide a part of the flow rate directly to the thermal oil buffer vessel without

running through the heat exchangers.

The general procedure for selecting this valve is described in section 4.2.2.

For the selection of this specific valve, figure A.1 is being used. The flow rate and maximum

pressure drop characteristics for the section that is being bypassed are:

Q 10 m3/h

∆pmax 0,81 bar

Based on the diagram from figure A.1, a mixing valve with Kvs 10 is necessary for this application.

Therefore a Danfoss VRB3 with Kvs 10 is chosen.

Since the length of the preheaters is variable, an additional valve in the circuit can be used in

order to induce an additional pressure drop in the thermal oil tubing, so the same pressure drop

can be achieved as in the case of full use of the preheaters. This ensures that one mixing valve is

appropriate for a variety of conditions.

Whether this regulating valve is really necessary, will be evaluated on the set-up itself.


4.5 Cooling circuit

The cooling circuit provides the cooler with a cold mixture of ethylene glycol and water in a

30 %/70 % volumetric ratio. The circuit has been previously represented in figure 3.5.

4.5.1 Chiller and buffer vessel

The cold source is a vessel of 900 l, which is also already available in the lab. The vessel is cooled

down by a 37 kW chiller, situated outside the building. The cooled water coming from the chiller,

has a temperature of 5 C.

There are several reasons for the use of a buffer vessel. First of all it increases the safety of the

system. Should the chiller break down, the remaining cooling capacity of the buffer vessel can be

used to shut down the test facility in a quick and easy manner. Secondly the vessel increases the

stability of the cooling liquid temperature. A third and final argument to the use of a buffer

vessel is the possibility to use the chiller in an easy ’on/off’ control system. This yields a higher

efficiency of the chiller unit itself.

4.5.2 Cooling circulation pump

The pump is a centrifugal pump, of which the flow rate is linked to the pressure it delivers. The

flow rate of cooling liquid to the cooling heat exchanger can be regulated by means of a three-way

mixing valve.

4.5.3 Cooling flow rate control

In order to control the flow rate independently of the pump- and heat exchanger characteristics,

an extra valve is added to the circuit in order to control the amount of bypass of the heat exchanger.

For the selection of this specific valve, figure A.1 is being used. The flow rate and maximum

pressure drop characteristics for the section that is being bypassed are:

Q 8 m3/h

∆pmax 1,75 bar

Based on diagram from figure A.1, a mixing valve with Kvs 6,3 is necessary for this application.

Therefore a Danfoss VRB3 with Kvs 6,3 is chosen.

The flow rate of cooling liquid is controlled by a PID in LabVIEW in order to slightly subcool

the supercritical working fluid compared to the test section inlet temperature.


4.6 Test set-up assembly

In appendix D, some 3D renders of the complete set-up without insulation are included. This

gives an overview of the practical implementation of the installation.

Chapter 5

Data-acquisition hardware

For this new test set-up, the Data-Acquisition (DAQ) and control hardware is very important. It

is the foundation for the accuracy of the final results and also ensures proper and safe operations

of the installation.

5.1 DAQ controller

Since control of the complete set-up is most important, a real-time controller is used to monitor

and control the operational parameters of the set-up.

In collaboration with National Instruments, a CompactRIO system has been chosen. The

real-time controller ensures that the set-up is always ’in control’ thanks to its on-board CPU and

real-time operating system (RTOS). The most common way for a DAQ system to fail is through

the communication between the actual DAQ-unit and the data analysis computer. By using an

on-board control unit, safety issues due to communication failure are eliminated.

The CompactRIO is equipped with analog input and output cards that are used to measure and

control the sensors (e.g. pressure transducers) and drivers (e.g. valve drivers) on the system.

This results into a reconfigurable I/O (RIO) architecture.

A user interface with communication to the compactRIO and data analysis is provided on the

actual measurement pc.

User

ComputerNetwork compactRIO

Figure 5.1: Configuration of the control and DAQ system.

75

Chapter 5. Data-acquisition hardware 76

5.2 Temperature sensors

Since heat transfer is studied in the set-up, temperature sensors are of major importance.

Two different types of temperature sensors are used: K-type thermocouples and Pt100 resistive

temperature sensors.

Depending on the location, desired accuracy and time scale, a choice between these two sensors

is made.

Following sections give a short introduction of the properties of the sensors and their implemen-

tation in the set-up.

5.2.1 Thermocouples

The thermocouples that are implemented in this set-up are type K thermocouples, which are

resistant to the relative high temperatures in the test section. The metals used in this type are

chromel and alumel.

Most of the thermocouples included in the set-up are implemented in the test section itself.

Each 333 mm, three thermocouples are attached to the inner tube of the test section, in order

to measure the temperatures at top, side and bottom. Even though the thermal conductivity

of the inner tube will even out the temperature profile, this implementation should result in a

measurable temperature difference which is used to visualize the buoyancy effect.

The implementation of these thermocouples is visually represented on figures 5.2 and 5.3.

Figure 5.2: Implementation of thermocouple on the inner tube in the test section.

To fix the thermocouples on the inner tube, a groove is milled into the surface of the tube, which

is filled with thermal paste in order to provide a good contact between thermocouple and tube.

For each axial location, three thermocouples are fixed in place and pressed inside the groove

by an ear clamp. Finally, a strip of heat shrink is used in order to ensure that no thermal

paste is washed away by the thermal oil. This heat shrink also provides a small amount of

thermal insulation between tube and thermal oil. This results into a shorter effective heat transfer

length, but also ensures that the actual wall temperature is measured and not the oil temperature.

On figure 5.3, three thermocouples in the annulus are also represented. These are placed each

1000 mm and measure the temperature variation of the heating fluid.


Figure 5.3: Configuration of thermocouples on the inner tube in the test section.

Due to their small size (1 mm), the thermocouples have a low thermal inertia and are ideal to

measure rapid temperature variations. However, the accuracy of thermocouples is limited, so the

temperature variations have to be large enough.

5.2.1.1 Accuracy

Each thermocouple is calibrated over its working range. This calibration is achieved by using a

calibrated reference probe (Fluke 1523), which gives the temperature of a controlled calibration

oven. Using the least-square error method, a regression curve for the temperature, measured by

the thermocouple, is generated.

Over the temperature range of 20 C to 100 C, the behaviour of the thermocouples is best

represented by a second order curve Y = Ax2 +Bx+ C, where Y is the actual temperature and

x is the temperature as measured by the thermocouple.

Calibrations are executed in order to achieve a standard deviation, as represented by equation

5.1 [35], lower than 0,1 C.

σTC =

√∑(Yi −Ax2i −Bxi − C

)2N − 3

(5.1)

Where N represents the number of measured values per temperature, per thermocouple.


5.2.2 Pt100 temperature sensors

A Pt100 is a resistive temperature device (RTD), of which the resistance is dependent on the

sensor’s temperature.

Table 5.1: Temperature-resistance correlation of the Pt100 sensors.

Temperature C Resistive value Ω

0 100

20 107,79

40 115,54

60 123,24

80 130,89

100 138,50

120 146,21

The thermal inertia of this type of sensor is significantly larger than the inertia of the smaller

thermocouples. The main benefit of using Pt100 sensors is their increased accuracy and physical

robustness.

5.2.2.1 Accuracy

The same calibration principle as with thermocouples is used for the Pt100 sensors. However, in

this case a linear regression curve is sufficient.

5.3 Pressure sensors

At inlet and outlet of the test section, an absolute pressure sensor is installed. These sensors are

used to control the pressure in the test section and give an indication of the pressure drop over

the test section.

For high accuracy pressure drop measurements, it is advisable to use a differential pressure sensor

that can cope with the high static pressure. This is not included yet in the set-up, so at first

only the absolute pressures at inlet and outlet are measured.

A third pressure sensor, with a lower accuracy, is installed in the low-pressure section, to monitor

the overall pressure in this part of the circuit.

5.3.1 Accuracy

The pressure sensors at the test section has an accuracy of ±0,1 % Full Scale (FS) Best Straight

Line (BSL). The pressure sensor in the low-pressure section has an accuracy of ±0,2 % FS BSL.


5.4 Coriolis mass flow meters

The Coriolis mass flow meter directly measures the mass flow through the device instead of the

more standard volumetric flow meters.

In a Coriolis mass flow meter, two measuring tubes act as an oscillating system. The tubes are

clamped at both ends and driven by an actuator. Vibration sensors measure the mechanical

oscillations of the two tubes. The frequency and phase shift of the oscillations of the two tubes

are dependent on the actual mass flow through the measuring device.

Figure 5.4: Coriolis mass flow meter without casing.

5.4.1 Accuracy

Accuracy analysis of the Coriolis mass flow meters has been executed by Rheonik. The summary

of this analysis is displayed in appendix A.3.


5.5 Sensor and driver tags

Each sensor and driver in the system has a tag, which is added to the wiring of the instrument.

These tags are also used in the software.

5.5.1 Temperature sensors

Thermocouples have tag numbers in the range from 100 to 499. These are subdivided in the 100

range for the thermocouples located at the top of the test section, the 200 range for the bottom

positions and the 300 range for the side positions. The 400 range of thermocouples are reserved

for bulk oil temperature measurements and temperature measurements in other parts of the

system.

Table 5.2: List of thermocouples that are not in the test section.

Sensor Description Tag

Thermocouple TC bulk oil pos. 4 402









Thermocouple TC condensor/cooler outlet 411

Thermocouple TC pump outlet 412

Thermocouple TC el. preheater in 413

Thermocouple TC el. preheater out 414

Thermocouple TC 415

5.5.2 Pressure sensors

All wires from the pressure sensors are given a tag in the 500 range.

5.5.3 Mass flow meter

All wires from the mass flow meters are given a tag in the 600 range.

Chapter 6

Software

The software in LabVIEW is divided into two main parts: The PC-software, containing the

Human-Machine Interface (HMI) and the control software on the compactRIO.

In this section, merely the principles of the software are described. The software that has been

developed during the time period of this master dissertation is mainly for testing, debugging and

reviewing of the hardware and their controls. No process control software has been developed

yet.

6.1 PC-software

The pc software provides the user interface for interaction with the set-up (HMI) and is used to

process the acquired data.

6.1.1 User interface

The User Interface (UI) is divided into three main sections, between which can be switched

through tabs: Settings, Operational, Measurements.

Settings

The Settings tab is used to start up communication with the compactRIO unit, set the desired

operational parameters,...

Operational

The Operational tab provides a schematic representation of the different flows in the installation

and the main data, such as acquired by the sensors. This overview is represented in figure 6.1. It

is clear that only the main data is represented, in order to provide a quick overview of how the

installation is behaving.

81

Chapter 6. Software 82

Figure 6.1: Human-machine interface of the installation.

Measurements

The Measurements tab provides a graphical representation of the measurements. These repre-

sentations are both direct numeric values as visualisation of the measurements, indicators for

steady-state behaviour,...

6.1.2 CoolProp

CoolProp [36] is an open-source, cross-platform, free property database based in C++ that

includes pure fluids, pseudo-pure fluids, and humid air properties.

During the design of the set-up, this software has been used in order to easily obtain the properties

of the discussed refrigerants at different conditions. In the operational control software and DAQ

software, CoolProp is also being used in order to define set points and trip points based on the

fluid that is being studied.


6.2 Control software

As discussed before in chapter 5, control of the complete set-up and process conditions is of

major importance for safety reasons.

These control mechanisms have to be implemented in the software that drives the set-up.

Following sections indicate a possible safety problem, describes how to detect it and what

measures should be taken by the software in order to ensure control of the set-up.

6.2.1 Pressure build-up

If the pressure in the high-pressure section raises too quickly, this is probably due to the relatively

slow reaction time of the pressure control valve (opening time). Therefore, the flow rate of the

pump is lowered.

6.2.2 Pressure drop

A sudden pressure drop in the system indicates a leakage of refrigerant out of the set-up. The

control software continuously monitors the pressures for sudden drops and results into an alarm

and shut-down of the system when this should occur.

6.2.3 Uncontrolled temperature raise

In case of an uncontrolled or undesired temperature raise, above the maximum set point of

130 C, proper functioning of several components cannot be ensured. A temperature above this

limit can also affect the integrity of several components that are used in the set-up (e.g. thermal

paste, sealants,...).

6.2.4 Pump inlet conditions

In order to ensure proper functioning of the pump, the conditions of the refrigerant at the inlet

of the pump have to be monitored continuously, such as discussed in section 4.3.3.3

Chapter 7

Operational procedures

Because of the relative complexity of the set-up, there is a broad range of procedures that have

to be followed when working on the set-up. Especially when executing special work procedures

such as charging the set-up with a new refrigerant or when emptying the set-up to a vacuum.

7.1 Pressure test

Before making the set-up operational, several pressure tests will have to be conducted. These

include:

• Vacuum test

• Low pressure (35 bar) pressure test with nitrogen of the complete setup

• High pressure (50 bar) pressure test of the high pressure section

7.2 System charging

To charge the system with refrigerant, a certain procedure will have to be followed. First of all

the system needs to be drained. This is done using a vacuum pump. The combined filters/dryers

will remove the moisture from the system.

To achieve the supercritical pressure in the system, the use of a boosting compressor/pump is nec-

essary. This unit assures that the critical pressure is reached, so no two-phase flow will be present.

The amount of mass present in the system can be calculated from the working fluid’s density at

ambient temperature and working pressure. Combined with the volume of the test circuit (with

empty accumulator), this results into the amount of working fluid mass in the system.

85

Chapter 7. Operational procedures 86

ρcharging = ρ(Tamb, Ptest) (7.1)

mwf = Vsystem · ρcharging (7.2)

The amount of refrigerant in the set-up varies slightly with the ambient temperature, at which

the fluid is before testing.

7.3 System start-up

To get to the desired test section inlet temperature, the system is started. During this phase, the

working fluid is circulated through its loop, whilst being heated by the electrical preheater and

by the thermal oil in the test section. Initially, the cooler is completely bypassed and the cooler

capacity will be increased as soon as the test section inlet temperature is reached.

7.4 System shut-down

At a shut-down of the system, no more heat is added to the system. The system is cooled down

more quickly by keeping the circulation pump in use and by using the cooler/condenser. Once

the temperature is lowered adequately, the circulation pump can be shut down as well

Once this is done, following steps have to be executed:

• Isolate the pump and flexible hoses from the remaining of the circuit

• Drain the isolated section of remaining refrigerants

7.5 Heat balance test

Before starting the main experiments, a heat balance test should be carried out. This test uses

single-phase flow conditions and results in an estimate of the heat loss in the test section. During

this test, all instrumentation is checked for proper working.


7.6 Absolute limitations

The installation is designed for the working fluids and conditions that have been discussed in the

previous chapters. However, this does not limit the use of the set-up to these specific operational

aspects. The possibility that other working fluids are going to be developed and studied is very

real. Depending on the properties of those new working fluids, they might be suitable for testing

on the set-up of Ghent University with little or no changes to the set-up.

The absolute limitations and material compatibilities with the components used in the set-up

should always be checked before tests with new working fluids are conducted. In order to ease

this, tables 7.1 and 7.2 list the maximum pressure and temperature for the components that are

used.

7.6.1 High pressure section

Table 7.1: Absolute pressure and temperature limits of components of the high pressure section.

High pressure section

Component pmax[bar] Tmax[C]

Flexible pressure line 61,6 230

Tubing 120 200

Mass flow meter 120 350

Pressure sensor 60 125

Test section 50 130

Expansion valve 140 120

Overall limitations 50 120

The overall limitation on pressure is the pressure in the test section itself. This is the result of

machining of grooves in the surface of the inner tube of the test section. Because of the possible

inaccuracies in machining, a large safety factor is applied. An extensive Finite Element Method

(FEM) calculation and practical tests on the workpiece itself can raise this limitation.

The overall limitation on temperature is dependent on the expansion valve.


7.6.2 Low pressure section

Table 7.2: Absolute pressure and temperature limits of components of the low pressure section.

High pressure section

Component pmax[bar] Tmax[C]

Pressure relief valve 40 100

Tubing 120 200

Shut-off valve 54 120

Condenser 80 150

Pressure sensor 60 125

Accumulator 35 150

Flexible suction line 35 100

Overall limitations 35 100

With 35 bar, the maximum pressure of the accumulator is rather low. Nevertheless, this is not a

mechanical limit in terms of strength of the accumulator itself since it can withstand pressures

up to 330 bar. The limitation is more of an operational point of view, since the pressure ratio of

the accumulator

(pmax

pmin

)is limited to avoid over-compression of the internal bladder. This has

been discussed more extensively in section 4.3.8.

The overall limitation on temperature is dependent on the pressure relief valve and the flexible

suction line of the pump.

Chapter 8

Data reduction

8.1 Data acquisition and calculations

The supercritical conditions of the working fluid pose extra challenges in order to gain correct

measurements of fluid properties. Especially the measurements of the working fluid itself are

difficult to perform in practice.

8.1.1 Annulus

The bulk temperature in the annulus is the average of the temperatures measured by the

thermocouples in the bulk flow.

Th,b =T011 + T012 + T013

3(8.1)

8.1.2 Wall

The outer wall temperature Two is directly measured by the thermocouples installed at the top,

side and bottom of the wall.

The inner wall temperature is calculated from the outer wall temperature using a heat conduction

model of the cylindrical tube. This model however, assumes a uniform heat transfer over the

surface of the tube section.

For each section, the heat transfer from the heating fluid to the working fluid can be calculated

by following expressions.

Qhf = Qwf + Qloss (8.2)

The heat loss Qloss is by the heat transfer through the outer tube and the insulation. In this

model the thermal resistance of the insulation is the dominant factor, so the model can be

simplified to heat conduction through the insulation.

89

Chapter 8. Data reduction 90

Qloss =2πλinsulationLsection

ln

(Do,insulation

Di,insulation

) (Tb − Ta) (8.3)

The heat transferred to the working fluid in the corresponding section is expressed as:

Qwf = Qhf − Qloss (8.4)

= mwf ·∆Hwf (8.5)

The heat flux Qwf flows through the wall of the inner tube by conduction. This results in

following equation:

Qwf =2πλtubeLsection

ln

(dodi

) (Tw,o − Tw,i) (8.6)

From this equation, the inner wall temperature can be found.

8.1.3 Inner tube

The local fluid bulk temperature can be calculated from the thermophysical properties which are

already available. Therefore, the property package CoolProp is implemented into the LabVIEW

control environment.

The pressure drop for each (virtual) subsection of the test section can be modelled. Combining

the subsection’s inlet and outlet pressure results in an average pressure Pavg for the section.

Twf,b = Twf (Pavg, Hwf ) (8.7)

The enthalpy Hwf of the working fluid is derived from the heat transfer equation 8.5 applied to

the subsection.


8.2 Modified Wilson plot

The modified Wilson plot method of Briggs and Young will be used ([37],[38],[39]). For turbulent

flow, the Nusselt number is represented in following forms:

Nui =hiDi

ki= CiRe

ai Pr

1/3i

(µ

µw

)0,14

i

(8.8)

Nuo =hoDT

ko= CoRe

aoPr

1/3o

(µ

µw

)0,14

o

(8.9)

These definitions are separate for annulus on the outer side of the tube (subscript o) and the

inner side of the tube (subscript i).

The heat transfer coefficients can be expressed as an thermal resistance equation:

1

UA=

1

hiAi+Rw +

1

hoAo(8.10)

With UA derived from the LMTD equation:

UA =Qw

∆TLMTD=mwfcp,wf∆Twf

∆TLMTD(8.11)

The thermal resistance of the wall Rw in equation 8.10 is expressed as:

Rw =ln(Do/Di)

2πLk(8.12)

By rearranging expressions 8.8 and 8.9 to expressions of hi and ho and substituting these equations

into equation 8.10, following expression is achieved:

1

UA=

Di

kiAiCiReai Pr1/3i

(µ

µw

)0,14

i

+Rw +Do

koAoCoReaoPr1/3o

(µ

µw

)0,14

o

(8.13)

The power factor a is generally equal to 0,8. A first rearrangement of equation 8.13 yields:

(1

UA−Rw

)[koDo

AoReaoPr

1/3o

(µ

µw

)0,14

o

]=

koDo

AoReaoPr

1/3o

(µ

µw

)0,14

o

kiDiAiReai Pr

1/3i

(µ

µw

)0,14

i

· 1

Ci+

1

Co(8.14)

This equation in the form Y = mX + b can be used in combination with linear regression on the

data in order to find m and b, thus Ci and Co

8.2.1 Experimental expressions for the Nusselt number

Based on the coefficients that are found by using above method, the final expression for the

Nusselt number can be derived.

Chapter 9

Error analysis

In order to determine the heat transfer coefficient, following measurements are performed:

• Mass flow rate

• Temperatures

• Pressures

These properties can be measured directly. The accuracy of the results depends only on the

measurement devices. During the calculation procedure for derived properties (e.g. pressure

drop, specific heat, heat transfer coefficient), these direct measurement errors propagate. The

accuracy of the results should be determined using an error analysis. This analysis reflects the

influence of measurement errors on the final results.

As a general approach, the error on a calculated value q, depending on actual measured parameters

(x,y,...) , is calculated as follows:

∆q(x, y, ...) =

√(∂q

∂x∆x

)2

+

(∂q

∂y∆y

)2

+ ... (9.1)

In this equation, the ∆ indicates the absolute measurement error of a parameter.

9.1 Accuracy of temperature sensors

All thermocouples and Pt100 sensors are calibrated before they are implemented in the system.

Each temperature sensor is calibrated within the range from 20 C to 100 C. This calibration is

achieved through providing a stable temperature environment by means of a dry block temper-

ature calibrator (DBC150), of which the temperature is checked by means of an independent

calibrated reference thermometer (Fluke 1523).

93

Chapter 9. Error analysis 94

The dry block calibrator is stabilised in steps of 10 C within the calibration range. Once a

stable environment has been achieved, 100 temperature measurements are executed for each

thermocouple. For each measurement, the temperature of the reference thermometer is also logged.

Based on the gained data, a calibration curve for the thermocouples and Pt100 sensors can be

calculated. In this calibration curve, the measured temperature of a thermocouple is taken as

known data (x), and the temperature of the reference thermometer is taken as the final result (y).

In order to achieve a standard deviation of the calibration curve from the reference temperatures

smaller than 0,1 C, a second order calibration curve is constructed using the following form:

y = Ax2 +Bx+ C (9.2)

The factors A, B and C that result from this calibration are implemented in the control and

DAQ-software, so the corrected temperatures are used.

The error from the linear regression on the measured temperature is expressed as:

∆TC =

√1

n− 3

∑Yi −AXi −BXi − CN

i=1 (9.3)

In which Yi is the actual temperature, as measured by the reference thermometer.

In the error analysis, the systematic error of the reference temperature sensor has also to be

taken into account. For the Fluke 1523, this error is 0,064 C [40].

δtotal =√

∆2TC + ∆2

systematic

=√

∆2TC + ∆2

Fluke

9.2 Accuracy of the heat transfer coefficient

The calculation of the accuracy of the heat transfer coefficient can be divided into several

intermediate results:

9.2.1 Local heat flux

The enthalpy change at position z can be written as:

dh(z)

dz= Cp

dT

dz(9.4)

The local heat flux:

q(z) =mCp

dT

dzπD

(9.5)


Using equation 9.1, this yields:

∆q(z) =

√(∂q

∂m∆(m)

)2

+

(∂q

∂T∆T

)2

(9.6)

The internal convection coefficient hi from equation 4.13 can be rewritten as:

hi =q

Twall − Twf −Rq(9.7)

Application of formula 9.1 yields:

∆hi =

√√√√√√√√(

−q(Twall − Twf −Rq)2

∆Twall

)2

+

(q

(Twall − Twf −Rq)2∆Twf

)2

+

(Twall − Twf

(Twall − Twf −Rq)2∆q

)2(9.8)

9.3 Accuracy of the Wilson plot

The accuracy of the expression for the Nusselt number, derived using the Wilson plot method,

is also expressed by means of equation 9.1. This method is applied to equations 8.8 and 8.9.

Generally, this results into following expression:

∆Nu =

√√√√√√√√√√

[Reai Pr

1/3

(µ

µw

)0,14

∆C

]2+

[CaRea−1Pr1/3

(µ

µw

)0,14

∆Re

]2

+

[CRealn(Re)Pr1/3

(µ

µw

)0,14

∆a

]2+

[CRea

1

3Pr−2/3

(µ

µw

)0,14

∆Pr

]2 (9.9)

The error on Re and Pr can be determined using equation 9.1. The error on the coefficients C

and a has to be determined using a error calculation method for linear regressions of the form

Y = mX + b

∆m =

√1

N − 2

∑(yi − b−mXi)

Ni=1

√N

N∑x2 − (

∑x)2

(9.10)

∆b =

√1

N − 2

∑(yi − b−mXi)

Ni=1

√ ∑x2

N∑x2 − (

∑x)2

(9.11)

Chapter 10

Conclusion

The main goal of this dissertation was the design of a new test set-up for the study of heat

transfer to working fluids of ORCs at supercritical conditions. Starting from a blank page, this

has proven to be a challenging job. Throughout the course of this dissertation, a lot of aspects of

scientific research and engineering have been executed.

During the first part of the dissertation, the idea has been made more specific. Keeping the future

use of the supercritical heat transfer to organic fluids in ORCs in mind, the initial boundary

conditions such as represented on page 38 were defined.

Based on the initial boundary conditions and existing research, as discussed in chapter 2, the

research idea has been specified furthermore, including the study of the buoyancy and acceleration

effects. The existing set-ups to study different properties of fluids at supercritical conditions

were also used as a source to get a general idea of what the set-up could look like.

The specifications and conditions for what is going to be studied on the actual set-up at Ghent

University have then been used to perform a range of simulations. These simulations include the

heat transfer process at supercritical conditions, for which the correlations discussed in chapter 2

have been used. A simplified set-up has also been simulated in order to get a better view on how

the working fluid will behave throughout the circuit.

These simulations resulted into a good estimation of the order of magnitude for different compo-

nents (e.g. heating power, flow rates, pumping power,...).

Starting from the theoretical design that has been developed throughout the study phase and

simulation phase of this dissertation, companies and suppliers were contacted in order to convert

the theoretical design to a design that could be executed and for which all necessary components

were commercially available.

97

Chapter 10. Conclusion 98

This phase of translating the original design to a design that can be executed, has proven to be

one of the most challenging parts of the project. The selected working fluids are not commonly

used at supercritical conditions, which results into a lack of experience for the companies involved.

The limitations of certain components implemented additional boundary conditions into the

design and the calculations. This required the first design to be adapted several times until all

specifications were met again.

During the construction phase, further adaptations to the design have been made due to practical

considerations. These aspects are often easily forgotten during the initial design and require a

certain amount of practical experience as an engineer.

The next steps that should be undertaken are the leakage- and pressure tests such as described in

chapter 7. Furthermore, the operational strategies and control systems have to be implemented

in the control software of the set-up.

Appendix A

Hardware

A.1 Plate heat Exchanger

99

.

CBXP27Brazed Plate Heat Exchanger

.

General informationAlfa Laval introduced its first brazed plate heat exchanger(BHE) in 1977 and has since continuously developed andoptimized its performance and reliability.

Brazing the stainless steel plates together eliminates the needfor gaskets and thick frame plates. The brazing material sealsand holds the plates together at the contact points ensuringoptimal heat transfer efficiency and pressure resistance. Theplate design guarantees the longest possible life.

The design options of the brazed heat exchanger areextensive. Different plate patterns are available for variousduties and performance specifications. You can choose astandard configuration BHE, or a unit designed according toyour own specific needs. The choice is entirely yours.

Typical applications- HVAC heating/cooling- Refrigerant applications- Industrial cooling/heating- Oil cooling

CO2 refrigerant applications- Suction gas heating- Oil cooling- Evaporating- Economizing- Sub cooling- Condensing

Working principlesThe heating surface consists of thin corrugated metal platesstacked on top of each other. Channels are formed betweenthe plates and corner ports are arranged so that the two mediaflow through alternate channels, usually in countercurrent flowfor the most efficient heat transfer process.

Standard designThe plate pack is covered by cover plates. Connections arelocated in the front or rear cover plate. To improve the heattransfer design, the channel plates are corrugated.

Particulars required for quotationTo enable Alfa Laval’s representative to make a specificquotation, specify the following particulars in your enquiry:- required flow rates or heat load- temperature program- physical properties of liquids in question- desired working pressure- maximum permitted pressure drop

Examples of connections

Externalthreaded

Internalthreaeded

Soldering Welding

* More connections are available on request.

CBXP27 - PED approval pressure/temperature graph*

CBXP27 - KHK 150°C (302°F) approval pressure/temperature graph*

Standard dimensions and weight*

A measure mm = 13 + (2.4 * n) (+/-3 mm)A measure inch = 0.51 + (0.09 * n) (+/-0.12 inch)Weight** kg = 2 + (0.13 * n)Weight** lb = 4.41 + (0.29 * n)

(n = number of plates)* Excluding connections

Standard data

Min. working temperature see graphMax. working temperature see graphMin. working pressure vacuumMax. working pressure see graphVolume per channel, litres (ga) 0.05 (0.013)Max. particle size mm (inch) 1.2 (0.05)Max. flowrate* m3/h (gpm) 14 (61.6)Min. nbr of plates 6Max. nbr of plates 150* Water at 5 m/s (16.4 ft/s) (connection velocity)

Standard materials

Cover plates Stainless steelConnections Stainless steelPlates Stainless steelBrazing filler Copper

Standard dimensionsmm (inch)

S3S3 S2S2

S4S4 S1S1

111 (4.37)

50 (1.97) A

250

(9.8

4)

310

(12.

2)

.

For exact values please contact your local Alfa Laval representative

.

PCT00128EN 1202 Alfa Laval reserves the right to change specifications without prior notification.

How to contact Alfa LavalUp-to-date AlfaLaval contact details forall countries are always available on ourwebsite on www.alfalaval.com

Brazed Plate Heat Exchanger Technical Specification Model : CBXP27-26L-F (32871 3758 0) Item : R134a condensor Date : 12/03/2014 ___________________________________________________________________________________ Hot Side Cold side Primary side Secondary side Fluid R134a 30.0% Eth.glycol Mass flow rate kg/h 1736 12380 Fluid Condensed/Vapourized kg/h 1736 0.000 Inlet temperature °C 95.0 15.0 Dew p. °C 94.2 / Outlet temperature(vapor/liquid) °C 94.2/266.6 10.0 Operating pressure(In/Out) bara 35.0/35.0 Pressure drop kPa 2.03 73.3 Velocity connection(In/Out) m/s 3.14/0.818 5.58/5.86 Heat Exchanged kW 63.59 Heat transfer area m² 0.60 OHTC service W/(m²*K) 1221 Fouling resistance*10000 m²*K/W 0.20 Margin % 8.37 Mean Temperature Difference K 86.8 Relative directions of fluids Countercurrent Number of passes 1 1 Materialplate/ brazing Alloy 316 / Cu ConnectionS1 (Cold-Out) Threaded (Internal)/ 1" ISO 228/1-G (V24) Alloy 304 ConnectionS2 (Cold-In) Threaded (Internal)/ 1" ISO 228/1-G (V24) Alloy 304 ConnectionS3 (Hot-Out) Threaded (Internal)/ 3/4" NPT ConnectionS4 (Hot-In) Threaded (Internal)/ 3/4" NPT Pressure vessel code PED Design pressure at 90.0 Celsius Bar 90.0 90.0 Design pressure at 225.0 Celsius Bar 75.0 75.0 Design temperature °C -196.0/225.0 Overall length x width x height mm 120 x 111 x 310 Net weight, empty / operating kg 5.68 /6.96 ___________________________________________________________________________________ Performance is conditioned on the accuracyof customers data and customers abilityto supply equipment and products in conformity therewith.

Appendix A. Hardware 109

A.2 Mixing valve diagram

Flow Rate

l/sec m3/h

FLOW Pressure drop kPa (100 kPa = 1bar = ~ 10 m H2O)

Δp m

ax

Figure A.1: Selection diagram for mixing valves.


A.3 Mass flow meter error analysis

Rheonik Flowmeter Summary(Printdate: 2-jan-2014 - Version: 2.1)

Tag No/Project: Universiteit Gent

Application Data

Fluid/Gas Name: Viscosity:R125 0,0002 Pa sMax. Rate: Density:30,0000 kg/min 1.251,00 kg/m³Max. Pressure: Max. Temp.:50,00 bar 110,00 Celsius

Application Data in Rheonik Standard Units

Max. Rate: Viscosity:30,0000 kg/min 0,1668 cPMax. Pressure: Density:50,00 bar 1.251,00 kg/m³

Max. Temp.: 110,00 Celsius

Your Selection

Meter Name: RHM12 [M#12] P-Sealless [PF0] - wt 1.00 [P1] -- ID 630Material: 1.4571 (316Ti) [M1] (std.)Connection: 1 inch ANSI 600 RF/SF [A3] (std.)

Your Given Rate

Flow Rate Mass Flow Accuracy ∆p Velocity Reynoldskg/min % Rate bar m/s Number

30,0000 0,16 0,12 2,54 187241

More Rates


2,0000 0,50 0,00 0,17 124835,1111 0,19 0,00 0,43 319008,2222 0,18 0,01 0,70 51318

11,3333 0,18 0,02 0,96 7073614,4444 0,17 0,03 1,23 9015317,5556 0,17 0,04 1,49 10957120,6667 0,17 0,06 1,75 12898823,7778 0,17 0,08 2,02 14840626,8889 0,17 0,10 2,28 16782430,0000 0,16 0,12 2,54 187241

Repeatability is 4 times better than the accuracy / linearity.For gas depending on pressure.

www.rheonik.com - the mass flowmeter experts



Application Data

Fluid/Gas Name: Viscosity:R125 0,0000 Pa sMax. Rate: Density:30,0000 kg/min 352,00 kg/m³Max. Pressure: Max. Temp.:50,00 bar 110,00 Celsius


Max. Rate: Viscosity:30,0000 kg/min 0,0267 cPMax. Pressure: Density:50,00 bar 352,00 kg/m³


Your Selection


Your Given Rate


30,0000 0,16 0,35 9,04 1061033

More Rates


2,0000 0,50 0,00 0,60 707365,1111 0,19 0,01 1,54 1807698,2222 0,18 0,03 2,48 290802

11,3333 0,18 0,05 3,42 40083514,4444 0,17 0,09 4,35 51086817,5556 0,17 0,12 5,29 62090120,6667 0,17 0,17 6,23 73093423,7778 0,17 0,22 7,17 84096726,8889 0,17 0,28 8,11 95100030,0000 0,16 0,35 9,04 1061033





Application Data

Fluid/Gas Name: Viscosity:R134a 0,0002 Pa sMax. Rate: Density:30,0000 kg/min 1.245,00 kg/m³Max. Pressure: Max. Temp.:50,00 bar 110,00 Celsius


Max. Rate: Viscosity:30,0000 kg/min 0,2210 cPMax. Pressure: Density:50,00 bar 1.245,00 kg/m³


Your Selection


Your Given Rate


30,0000 0,16 0,12 2,56 144686

More Rates


2,0000 0,50 0,00 0,17 96465,1111 0,19 0,00 0,44 246508,2222 0,18 0,01 0,70 39655

11,3333 0,18 0,02 0,97 5465914,4444 0,17 0,03 1,23 6966417,5556 0,17 0,05 1,50 8466820,6667 0,17 0,06 1,76 9967323,7778 0,17 0,08 2,03 11467726,8889 0,17 0,10 2,29 12968230,0000 0,16 0,12 2,56 144686





Application Data

Fluid/Gas Name: Viscosity:R134a 0,0000 Pa sMax. Rate: Density:30,0000 kg/min 352,00 kg/m³Max. Pressure: Max. Temp.:50,00 bar 110,00 Celsius




Your Selection


Your Given Rate


30,0000 0,16 0,34 9,04 1591549

More Rates


2,0000 0,50 0,00 0,60 1061035,1111 0,19 0,01 1,54 2711538,2222 0,18 0,03 2,48 436202

11,3333 0,18 0,05 3,42 60125214,4444 0,17 0,08 4,35 76630217,5556 0,17 0,12 5,29 93135120,6667 0,17 0,16 6,23 109640123,7778 0,17 0,21 7,17 126145026,8889 0,17 0,27 8,11 142650030,0000 0,16 0,34 9,04 1591549





Application Data

Fluid/Gas Name: Viscosity:Therminol 20 gr 0,0092 Pa sMax. Rate: Density:140,0000 kg/min 857,00 kg/m³Max. Pressure: Max. Temp.:5,00 bar 130,00 Celsius




Your Selection


Your Given Rate


140,0000 0,16 0,40 5,35 8980

More Rates


6,0000 0,50 0,00 0,23 38520,8889 0,19 0,01 0,80 134035,7778 0,18 0,02 1,37 229550,6667 0,17 0,07 1,94 325065,5556 0,17 0,10 2,51 420580,4444 0,17 0,15 3,07 516095,3333 0,17 0,20 3,64 6115

110,2222 0,17 0,26 4,21 7070125,1111 0,16 0,33 4,78 8025140,0000 0,16 0,40 5,35 8980





Application Data

Fluid/Gas Name: Viscosity:Therminol 130 gr 0,0007 Pa sMax. Rate: Density:140,0000 kg/min 765,00 kg/m³Max. Pressure: Max. Temp.:5,00 bar 130,00 Celsius




Your Selection


Your Given Rate


140,0000 0,16 0,28 5,99 119601

More Rates


6,0000 0,50 0,00 0,26 512620,8889 0,19 0,01 0,89 1784535,7778 0,18 0,02 1,53 3056550,6667 0,17 0,04 2,17 4328465,5556 0,17 0,07 2,81 5600480,4444 0,17 0,10 3,44 6872395,3333 0,17 0,14 4,08 81443

110,2222 0,17 0,18 4,72 94162125,1111 0,16 0,22 5,36 106882140,0000 0,16 0,28 5,99 119601



Appendix B

Process flow diagram

119

Vacuum pump

Circulation pump

Filter/Dryer

Electricalpreheater

Cooler/Condenser

T TT

T

T

T

F

PP

Refrigerant

T

F

Chiller

Safety reliefvalve

Pressure controlexpansion valve

P

Tube-in-tube preheater 1

Tube-in-tube preheater 2

Test Section

Low pressure section shut-off

valve

T T

Buffer vessel

Accumulator

Oil heater

Appendix C

Test section piece- and assembly

drawings

123

65

53

57

65 A

A

G2"

50

32,5

25

8,8

15 B

B

SECTION A-A

8,7

SECTION B-B

G 1/8 BSPT

G 1/8 BSPT

G 1/8 BSPT

Firm:Scale: 1:1

Title: Measuring block type 1

UGent

Drawing/Design: Stijn Daelman

65

53 57

A

A

G2"

15

22 22 B

B

CC

SECTION A-A

18,8

18,

8

18,8

M6

SECTION B-B

SECTION C-C

G 1/8"BSPTG 1/8"BSPT

G 1/8"BSPT

G 1/8"BSPT

Firm:Scale: 1:1

Title: Measuring block type 2

UGent


65

A

A

40,6

60,3 G2"

28

,6

53

15 15

4,2

5

41

34

,9

SECTION A-A

Firm:Scale: 1:1

Title: End cap for tube-in-tube heat exchangers

UGent


A CF

1 3

7 8 9 10 11 12 13 14 15 16 17 3

1

6

4 4 5 4 4 5 4 4 5 4 4

18

BB

DETAIL A SCALE 1 : 2

SECTION B-B SCALE 1 : 2

2

DD

DETAIL C SCALE 1 : 2

SECTION D-D SCALE 1 : 2

2

50

GG

DETAIL F SCALE 1 : 2

SECTION G-G SCALE 1 : 2

ITEM NO. PART NUMBER QTY.

1 Outer tube top 22 Sealing ring 2

3 Tee 2

4 Small measuring block 8

5 Large measuring block 3

6 Outer tube 1 17 Outer tube 2 18 Outer tube 3 19 Outer tube 4 110 Outer tube 5 111 Outer tube 6 112 Outer tube 7 113 Outer tube 8 114 Outer tube 9 115 Outer tube 10 116 Outer tube 11 117 Outer tube 12 118 Inner tube 1

Department of Flow, Heat and Combustion MechanicsApplied Thermodynamics and Heat Transferwww.ugent.be/ea/floheacomUGent

Title: Assembly drawing - Test Section

Scale: Firma:

Drawing ID:Proofread:Drawing/Design:Stijn Daelman

Date:

Appendix D

Complete set-up 3D renders

133

Appendix E

Software

E.1 EES-simulation code

1 ”EES Simulat ion to s ize a tube−in−tube heat exchanger ”

2

3 ”WORKING CONDITIONS”

4

5 ”Hot Source ”

6 ”Use o f Therminol ADX10 as hot f l u i d . This i s not inc luded in EES data s e t s

yet , so manual implemented c o r r e l a t i o n s are being used . ”

7

8 T hf in =130 [C] ” I n l e t temperature o f the heat ing f l u i d ”

9 P hf in= 1e5 [ Pa ] ”Water at 2 bar to be ab le to go to temperatures

h igher than 100C”

10 Rho hf = Rho mean b hf [ 1 ] ” Density o f the heat ing f l u i d at i n l e t

temperature ”

11 v h f i n = m dot hf /Rho hf/S an ” Entrance v e l o c i t y o f the heat ing f l u i d ”

12 Q dot hf = m dot hf /Rho hf∗ convert ( ’ 1/ s ’ ; ’ 1/h ’ )

13 Q dot hf = 10 [mˆ3/h ]

14

15 T hf b mean = ( T hf in+T hf out ) /2 ”Mean temperature o f the heat ing

f l u i d ”

16 P hf mean = ( P h f i n+P hf out ) /2 ”Mean pre s su r e o f the heat ing f l u i d

”

17 Cp b hf = ( Cp mean b hf [1 ]+ Cp mean b hf [N−1]) /2 ”Average heat capac i ty o f

the heat ing f l u i d ”

18

19 ”Working f l u i d ”

20 WF$ = ’ r125 ’ ”Working f l u i d ”

21

22 T wf in = 70 [C] ” I n l e t temperature working f l u i d ”

23 T wf out = T wf in+DeltaT wf ” Outlet temperature o f the working f l u i d ,

139

Appendix E. Software 140

de f ined by i n l e t temperature and temperature r e s o l u t i o n ”

24

25 P wf in = P c r i t (WF$) ∗1 ,05 ” s u p e r c r i t i c a l working pr e s su r e ”

26 G = m dot wf/ S in ”Mass f l u x in the inner tube ”

27 G= 1000 [ kg/mˆ2−s ]

28 Rho wf in = Density (WF$;T=T wf in ;P=P wf in ) ” Density o f the working

f l u i d at i n l e t ”

29 v in w f = G/ Rho wf in ” I n l e t v e l o c i t y working f l u i d ”

30 Q dot wf = m dot wf/Rho mean b wf [N−1]∗ convert ( ’ 1/ s ’ ; ’ 1/h ’ )

31

32

33 T wf b mean = ( T wf in+T wf out ) /2 ”Mean temperature o f the working

f l u i d ”

34 P wf mean = ( P wf in+P wf out ) /2 ”Mean pre s su r e o f the working f l u i d

”

35 Cp b wf = Cp(WF$;T=T wf b mean ; P=P wf mean ) ”Average heat capac i ty o f

the working f l u i d ”

36

37

38 ” Ca l cu l a t i on o f s t a t e po in t s at in− and o u t l e t ”

39 h wf out = Enthalpy (WF$;T= T wf out ; P=P wf [ 1 ] ) ” c a l c u l a t e enthalpy o f

working f l u i d at the o u t l e t us ing the correspond ing temperature and

pr e s su r e ”

40 h wf in=Enthalpy (WF$;T=T wf in ;P=P wf [N] ) ” c a l c u l a t e enthalpy o f

working f l u i d at the i n l e t us ing the correspond ing temperature and

pr e s su r e ”

41

42

43 ”T IN T HEAT EXCHANGER”

44

45 ” Dimensions ”

46 d o = (1+1/8) ∗25 ,4∗10ˆ(−3) [m] ” Outside diameter o f the tube ”

47 t = 1 ,9∗10ˆ(−3) [m] ” Thickness o f tube ”

48 d i = d o−2∗t ” I n s i d e diameter o f inner tube ”

49

50 D out = (2+1/8) ∗10ˆ(−3) ∗25 ,4 [m] ”Outer tube out s id e diameter ”

51 t ou t = 1 ,47∗10ˆ(−3) [m]

52 D in= D out−2∗ t ou t ”Outer tube i n s i d e diameter ”

53

54 L = 4 [m] ” Total length o f the t e s t s e c t i o n ”

55

56 S an = ( pi∗D inˆ2/4−pi∗d o ˆ2/4) ”Annulus f low cross s e c t i o n ”

57 S in = ( pi∗ d i ˆ2/4) ”Tube f low f low c r s o s s s e c t i o n ”

58


59 ”Thermodynamical dimensions ”

60

61 D h = 4∗pi ∗( D inˆ2−d o ˆ2) /( pi ∗( d o+D in ) ∗4) ” Hydraul ic diameter ”

62 D t = 4∗pi ∗( D inˆ2−d o ˆ2) /( pi∗d o ∗4) ”Thermal diameter ”

63

64 lambda tube = 260 [W/m−K] ”Thermal conduc t i v i ty o f K65”

65

66 ” D i s c r e t i s a t i o n o f the heat exchanger ”

67

68 N = 20 ”10 D i s c r e t e s e c t i o n s o f 20 cm”

69

70 T hf in = T hf [ 1 ] ” Point 1 i s the i n l e t o f the heat ing f l u i d ”

71 T hf out = T hf [N] ” Point N i s the o u t l e t o f the heat ing f l u i d ”

72 T wf in = T wf [N] ” Point N i s the i n l e t o f the working f l u i d ”

73 T wf out = T wf [ 1 ] ” Point 1 i s the o u t l e t o f the working f l u i d ”

74

75 P hf in = P hf [ 1 ] ” I n l e t p r e s su r e o f the heat ing f l u i d ”

76 P hf out = P hf [N] ” Outlet p r e s su r e o f the heat ing f l u i d ”

77

78 P wf in = P wf [N] ” I n l e t p r e s su r e o f the working f l u i d ”

79 P wf out = P wf [ 1 ] ” Outlet p r e s su r e o f the working f l u i d ”

80

81 L[1]=0 ” F i r s t po int at beg inning o f heat exchanger ”

82 L s e c t i o n= L/(N−1) ”Length o f one s e c t i o n ”

83 A i = d i ∗pi∗ L s e c t i o n ”Heat exchanging area in the WF s i d e ”

84 A o = d o∗pi∗ L s e c t i o n ”Heat exchanging area in the HF s i d e ”

85

86 Q tot [ 1 ] = 0 ”No heat exchanged at s t a r t i n g po int ”

87 Q netto = Q tot [N] ”Net heat exchanged”

88

89 ” Ca l cu l a t i on o f heat t r a n s f e r per d i v i s i o n ”

90

91 d u p l i c a t e i =1;N−1

92

93 ”Heat balance for each s e c t i o n ”

94

95 T hf mean b [ i ] = ( T hf [ i ]+ T hf [ i +1]) /2 ”Mean temperature heat ing

f l u i d per d i v i s i o n ”

96 T wf mean b [ i ] = ( T wf [ i ]+T wf [ i +1]) /2 ”Mean temperature working

f l u i d per d i v i s i o n ”

97

98 Q[ i ] = m dot hf ∗Cp mean b hf [ i ] ∗ ( T hf [ i ]−T hf [ i +1])

99 Q[ i ] = m dot wf∗Cp mean b wf [ i ] ∗ ( T wf [ i ]−T wf [ i +1])

100


101 Q[ i ] = HTC mean hf [ i ]∗A o ∗( T hf mean b [ i ]−T hf mean w [ i ] ) ”Heat

t r a n s f e r heat ing f l u i d to tube ”

102 Q[ i ] = HTC mean wf [ i ]∗ A i ∗( T wf mean w [ i ]−T wf mean b [ i ] ) ”Heat

t r a n s f e r tube to working f l u i d ”

103 Q[ i ] = 2∗pi∗ lambda tube∗ L s e c t i o n / ln ( d o / d i ) ∗( T hf mean w [ i ]−T wf mean w [ i

] ) ”Heat t r a n s f e r through tube ”

104

105 Q tot [ i +1] = Q tot [ i ]+Q[ i ]

106

107 L [ i +1] = L [ i ]+ L s e c t i o n

108 h wf [ i ] = Enthalpy (WF$;T=T wf [ i ] ; P=P wf [ i ] ) ” c a l c u l a t e enthalpy

o f working f l u i d at the cor re spond ing temperature and pre s su r e ”

109 s wf [ i ] = Entropy (WF$; h=h wf [ i ] ; P=P wf [ i ] ) ” c a l c u l a t e entropy

o f working f l u i d at the cor re spond ing temperature and pre s su r e ”

110

111

112 ”GNIELINSKI” ”The G n i e l i n s k i c o r r e l a t i o n i s used to c a l c u l a t e

the HTC of the heat ing f l u i d ”

113

114 ” Ca lcu la te HTC of heat ing f l u i d area per d i v i s i o n ”

115 HTC mean hf [ i ] = Nu mean hf [ i ]∗ lambda mean b hf [ i ] / D t ”Heat t r a n s f e r

c o e f f i c i e n t heat ing f l u i d ”

116

117 ” Pressure r e c a l c u l a t e d v ia p r e s su r e drop”

118 P hf [ i +1] = P hf [ i ] − Dp mean hf [ i ] ” Pres sure in the beg inning

o f next s e c t i o n ”

119

120 ” Ca lcu la te p r o p e r t i e s o f heat ing f l u i d per d i v i s i o n ”

121 Dyn visc mean b hf [ i ] = (exp (645 ,13/( T hf mean b [ i ]+117 ,8) −2 ,662) ) ∗Rho mean b hf [ i ]∗10ˆ(−6) ”Dynamic v i s c o s i t y o f the heat ing

f l u i d ”

122 Rho mean b hf [ i ] =870 ,297 −0 ,684497∗T hf mean b [ i ]+5 ,18441∗10ˆ(−5)∗T hf mean b [ i ]ˆ2−1 ,0695∗10ˆ(−6)∗T hf mean b [ i ] ˆ3 ” Density o f the

heat ing f l u i d ”

123 lambda mean b hf [ i ] = −0 ,000123∗ T hf mean b [ i ] − 9 ,161∗10ˆ(−8)∗T hf mean b

[ i ]ˆ2+0 ,1265 ”Thermal conduc t i v i ty o f the heat ing f l u i d ”

124 Cp mean b hf [ i ] = (0 ,00392 ∗ T hf mean b [ i ] −1 ,5∗10ˆ(−6)∗ T hf mean b [ i ] ˆ2

+ 1 ,8363) ∗10ˆ3 ” S p e c i f i c heat o f the heat ing f l u i d ”

125

126 ” Ca lcu la te Nu, Re and Pr−number o f heat ing f l u i d per d i v i s i o n ”

127 Nu mean hf [ i ] = ( ( f h f [ i ] / 8 ) ∗( Re mean b hf [ i ]−1000)∗Pr mean b hf [ i ] )

/(1+12 ,7∗( f h f [ i ] / 8 ) ˆ(1/2) ∗ ( ( Pr mean b hf [ i ] ) ˆ(2/3)−1) ) ” G n i e l i n k s i

c o r r e l a t i o n ”

128 f h f [ i ] = (1 ,82∗ log10 ( Re mean b hf [ i ] ) −1 ,64)ˆ(−2) ”


f r i c t i o n f a c t o r for working f l u i d ”

129 Re mean b hf [ i ] = Rho mean b hf [ i ]∗ v mean b hf [ i ]∗D h/ Dyn visc mean b hf [ i ]

”Re−number heat ing f l u i d ”

130 Pr mean b hf [ i ] = Cp mean b hf [ i ]∗ Dyn visc mean b hf [ i ] / lambda mean b hf [ i ]

”Pr−number heat ing f l u i d ”

131

132 ” Ca lcu la te v e l o c i t y o f heat ing f l u i d i n s i d e the tube per d i v i s i o n ”

133 v mean b hf [ i ] = m dot hf /( Rho mean b hf [ i ]∗ S an ) ” v e l o c i t y

working f l u i d in the annulus ”

134

135

136 ”PETUKHOV−KRANOSCHEKOV” ”The Petukhov−Kranoschekov c o r r e l a t i o n i s

used to c a l c u l a t e the s u p e r c r i t i c a l HTC of the working f l u i d ”

137

138 ” Ca lcu la te HTC of working f l u i d area per d i v i s i o n ”

139 HTC mean wf [ i ] = Nu mean wf [ i ]∗ lambda mean b wf [ i ] / d i ”Heat t r a n s f e r

c o e f f i c i e n t working f l u i d ”

140

141 ” Pressure r e c a l c u l a t e d v ia p r e s su r e drop”

142 P wf [ i +1] = P wf [ i ] + Dp mean wf [ i ] ” Pressure in the beg inning

o f next s e c t i o n ”

143

144 ” Ca lcu la te p r o p e r t i e s o f working f l u i d per d i v i s i o n ”

145 Dyn visc mean b wf [ i ] = v i s c o s i t y (WF$; T=T wf mean b [ i ] ; P=P wf [ i ] ) ”

Dynamic v i s c o s i t y o f the working f l u i d ”

146 Rho mean b wf [ i ] = Density (WF$; T=T wf mean b [ i ] ; P=P wf [ i ] ) ” Density o f

the working f l u i d ”

147 lambda mean b wf [ i ] = Conduct iv i ty (WF$; T=T wf mean b [ i ] ; P=P wf [ i ] ) ”

Thermal conduc t i v i ty o f the working f l u i d ”

148 Cp mean b wf [ i ] = Cp(WF$; T=T wf mean b [ i ] ; P=P wf [ i ] ) ” S p e c i f i c heat o f

working f l u i d ”

149

150 ” Ca lcu la te Nu, Re and Pr−number o f working f l u i d per d i v i s i o n ”

151 Nu mean wf [ i ] = Nu mean wf 0 [ i ] ∗ ( Cp mean wf [ i ] / Cp mean b wf [ i ] ) ˆ (0 ,35) ∗(

lambda mean b wf [ i ] / lambda mean w wf [ i ] ) ˆ(−0 ,33) ∗( Dyn visc mean b wf [ i ] /

Dyn visc mean w wf [ i ] ) ˆ (0 ,11) ”Adapted Nu−number for s u p e r c r i t i c a l

working f l u i d s ”

152 Nu mean wf 0 [ i ] = ( ( f w f [ i ] / 8 ) ∗Re mean b wf [ i ]∗ Pr mean wf [ i ] ) /(1 ,07+12 ,7∗(

f w f [ i ] / 8 ) ˆ(1/2) ∗ ( ( Pr mean wf [ i ] ) ˆ(2/3)−1) ) ”Nu−number for

working f l u i d ”

153 f w f [ i ] = (1 ,82∗ log10 ( Re mean b wf [ i ] ) −1 ,64)ˆ(−2)

” F r i c t i o n f a c t o r for working f l u i d ”

154 Re mean b wf [ i ] = Rho mean b wf [ i ]∗ v mean b wf [ i ]∗ d i / Dyn visc mean b wf [ i ]

”Re−number working f l u i d ”


155 Pr mean wf [ i ] = Cp mean wf [ i ]∗ Dyn visc mean b wf [ i ] / lambda mean b wf [ i ]

”Pr−number working f l u i d ”

156

157 ” Ca lcu la te EXTRA p r o p e r t i e s o f working f l u i d per d i v i s i o n ”

158 h mean w [ i ] = Enthalpy (WF$;T=T wf mean w [ i ] ; P=P wf [ i ] ) ”Mean wal l enthalpy

o f the working f l u i d to determine the mean Cp per d i v i s i o n ”

159 h mean b [ i ] = Enthalpy (WF$;T=T wf mean b [ i ] ; P=P wf [ i ] ) ”Mean bulk enthalpy

o f the working f l u i d to determine the mean Cp per d i v i s i o n ”

160 Cp mean wf [ i ] = (Cp(WF$; T=T wf mean b [ i ] ; P=P wf [ i ] )+Cp(WF$; T=T wf mean w

[ i ] ; P=P wf [ i ] ) ) /2 ”Mean Cp working f l u i d ”

161 lambda mean w wf [ i ] = Conduct iv i ty (WF$; T=T wf mean w [ i ] ; P=P wf [ i ] ) ”

Thermal conduc t i v i ty o f the working f l u i d at the wa l l ”

162 Dyn visc mean w wf [ i ] = v i s c o s i t y (WF$; T=T wf mean w [ i ] ; P=P wf [ i ] ) ”

Dynamic v i s c o s i t y o f the working f l u i d at the wa l l ”

163

164 ” Ca lcu la te v e l o c i t y o f working f l u i d i n s i d e the tube per d i v i s i o n ”

165 v mean b wf [ i ] = m dot wf /( Rho mean b wf [ i ]∗ S in ) ” v e l o c i t y working

f l u i d in 1 tube ”

166

167

168 ” Pressure drop”

169 Dp mean hf [ i ] = 4∗ f h f [ i ] ∗ ( L s e c t i o n /D h) ∗( Rho mean b hf [ i ] ∗ ( v mean b hf [ i

] ˆ 2 ) /2) ” p r e s su r e drop heat ing f l u i d per d i v i s i o n ”

170 Dp mean wf [ i ] = 4∗ f w f [ i ] ∗ ( L s e c t i o n / d i ) ∗( Rho mean b wf [ i ] ∗ ( v mean b wf [ i

] ˆ 2 ) /2) ” p r e s su r e drop working f l u i d per d i v i s i o n ”

171

172 end

173

174

175 Delta P wf = P wf in−P wf out

176 Dyn visc mean wf plot = Dyn visc mean w wf [ 1 ]

Bibliography

[1] “ORCNext.”

http://www.orcnext.be/.

[2] “BP Statistical Review of World Energy June 2013.”

bp.com/statisticalreview.

[3] T. Hung, T. Shai, and S. Wang, “A review of organic Rankine cycles (ORCs) for the

recovery of low-grade waste heat,” Energy, vol. 22, no. 7, pp. 661–667, 1997.

[4] Y. Chen and Y. Chen, “Novel Cycles Using Carbon Dioxide as Working Fluid New Ways to

Utilize Energy from Low-Grade Heat Sources,” 2006.

[5] E. Cayer, N. Galanis, M. Desilets, H. Nesreddine, and P. Roy, “Analysis of a carbon dioxide

transcritical power cycle using a low temperature source,” Applied Energy, vol. 86,

pp. 1055–1063, July 2009.

[6] K. Yang, H. Zhang, Z. Wang, J. Zhang, F. Yang, E. Wang, and B. Yao, “Study of zeotropic

mixtures of ORC (organic Rankine cycle) under engine various operating conditions,”

Energy, vol. 58, pp. 494–510, Sept. 2013.

[7] E. Feher, “The supercritical thermodynamic power cycle,” Energy conversion, 1968.

[8] B. Saleh, G. Koglbauer, M. Wendland, and J. Fischer, “Working fluids for low-temperature

organic Rankine cycles,” Energy, vol. 32, pp. 1210–1221, July 2007.

[9] Y. Chen, P. Lundqvist, A. Johansson, and P. Platell, “A comparative study of the carbon

dioxide transcritical power cycle compared with an organic Rankine cycle with R123 as

working fluid in waste heat recovery,” Applied Thermal Engineering, vol. 25, no. 6,

pp. 2142–2147, 2006.

[10] S. M. Liao and T. S. Zhao, “Measurements of Heat Transfer Coefficients From Supercritical

Carbon Dioxide Flowing in Horizontal Mini/Micro Channels,” Journal of Heat Transfer,

vol. 124, no. 3, p. 413, 2002.

[11] A. Ashrae, W. J. Brock, J. M. Calm, and R. G. Richard, Designation and Safety

Classifications of Refrigerants, 2000.

145

http://www.orcnext.be/

bp.com/statisticalreview

Bibliography 146

[12] M. Bazargan, “Forced convection heat transfer to turbulent flow of supercritical water in a

round horizontal tube,” no. August, 2001.

[13] R. P. Bringer and J. M. Smith, “Heat transfer in the critical region,” AIChE Journal, vol. 3,

no. 1, pp. 49–55, 1957.

[14] Z. Miropolsky and M. Shitsman, “Heat transfer to water and steam at variable specific heat

(in near critical region),” J Tech Phys, vol. 27, no. 10, pp. 2359–2372, 1957.

[15] B. Petukhov, E. Krasnoshchekov, and V. Protopopov, “An investigation of heat transfer to

fluids flowing in pipes under supercritical conditions,” ASME International Developments

in Heat Transfer Part, vol. 3, pp. 569–578, 1961.

[16] I. L. Pioro, H. F. Khartabil, and R. B. Duffey, “Heat transfer to supercritical fluids flowing

in channels - empirical correlations (survey),” Nuclear Engineering and Design, vol. 230,

pp. 69–91, May 2004.

[17] Y. V. Vikrev and V. Lokshin, “An experimental study of temperature conditions in

horizontal steam generating tubes at supercritical pressures,” Teploenergetika, vol. 11, no. 3,

pp. 12–16, 1964.

[18] M. Shitsman, “The effect of natural convection on temperature conditions in horizontal

tubes at supercritical pressures,” Teploenergetika, vol. 13, no. 7, pp. 51–56, 1966.

[19] G. Adebiyi and W. Hall, “Experimental investigation of heat transfer to supercritical

pressure carbon dioxide in a horizontal pipe,” International Journal of Heat and Mass

Transfer, vol. 19, no. 7, pp. 715–720, 1976.

[20] W. Hall, “Heat transfer near the critical point,” Advances in Heat Transfer, 1971.

[21] J. Jackson and W. Hall, “Forced convection heat transfer to fluids at supercritical pressure,”

Turbulent forced convection in channels and bundles, vol. 2, pp. 563–611, 1979.

[22] J. Jackson and W. Hall, “Influences of buoyancy on heat transfer to fluids flowing in

vertical tubes under turbulent conditions,” Turbulent forced convection in channels and

bundles, vol. 2, pp. 613–640, 1979.

[23] M. Bazargan, D. Fraser, and V. Chatoorgan, “Effect of Buoyancy on Heat Transfer in

Supercritical Water Flow in a Horizontal Round Tube,” Journal of Heat Transfer, vol. 127,

no. 8, p. 897, 2005.

[24] B. Petukhov, A. Polyakov, V. Kuleshov, and Y. Shekhter, “Turbulent flow and heat transfer

in horizontal tubes with substantial influence of thermo-gravitational forces,” in Proc. 5th

Int. Heat Transfer Conf., Tokyo, vol. 3, pp. 164–168, 1974.

Bibliography 147

[25] B. Petukhov, A. Polyakov, and B. E. Launder, “Heat transfer in turbulent mixed

convection,” 1988.

[26] V. Ankudinov and V. Kurganov, “Intensification of deteriorated heat transfer in heated

tubes at supercritical pressures.,” HIGH TEMP., vol. 19, no. 6, pp. 870–874, 1982.

[27] H. Swenson, J. Carver, and C. d. Kakarala, “Heat transfer to supercritical water in

smooth-bore tubes,” Journal of Heat Transfer, vol. 87, p. 477, 1965.

[28] M. Shitsman, “Heat transfer to supercritical helium, carbon dioxide, and water: Analysis of

thermodynamic and transport properties and experimental data,” Cryogenics, 1974.

[29] J. Jackson and J. Fewster, “Forced convection data for supercritical pressure fluids,” Simon

Engineering Laboratory, University of Manchester, 1975.

[30] K.-H. Kang and S.-H. Chang, “Experimental study on the heat transfer characteristics

during the pressure transients under supercritical pressures,” International Journal of Heat

and Mass Transfer, vol. 52, pp. 4946–4955, Oct. 2009.

[31] P.-X. Jiang, C.-R. Zhao, and B. Liu, “Flow and heat transfer characteristics of r22 and

ethanol at supercritical pressures,” The Journal of Supercritical Fluids, vol. 70, pp. 75–89,

Oct. 2012.

[32] S. H. Yoon, J. H. Kim, Y. W. Hwang, M. S. Kim, K. Min, and Y. Kim, “Heat transfer and

pressure drop characteristics during the in-tube cooling process of carbon dioxide in the

supercritical region,” International Journal of Refrigeration, vol. 26, pp. 857–864, Dec. 2003.

[33] V. Gnielinski, “New equations for heat and mass transfer in turbulent pipe and channel

flow,” Int. Chem. Eng, vol. 16, no. 2, pp. 359–368, 1976.

[34] Parker Hannifin, “Bladder Accumulators EHV from 250 to 690 bar.”

[35] J. Taylor, “An introduction to error analysis: the study of uncertainties in physical

measurements,” 1997.

[36] V. L. J. W. Ian Bell, Sylvain Quoilin, “Coolprop.”

[37] A. I. E. Sherbini and A. M. Jacobi, “Modified Wilson-Plot Technique for Heat Exchanger

Performance : Strategies for Minimizing Uncertainty in Data Reduction PERFORMANCE :

STRATEGIES FOR MINIMIZING UNCERTAINTY IN,” 2004.

[38] J. W. Rose, “Heat-transfer coefficients, Wilson plots and accuracy of thermal

measurements,” Experimental Thermal and Fluid Science, vol. 28, pp. 77–86, Jan. 2004.

[39] J. Fernandez-Seara, F. J. Uhıa, J. Sieres, and A. Campo, “A general review of the Wilson

plot method and its modifications to determine convection coefficients in heat exchange

devices,” Applied Thermal Engineering, vol. 27, pp. 2745–2757, Dec. 2007.

[40] Hart Scientiffic, “1523,1524 Reference Thermometer Technical Guide.”

[41] B. Petukhov, “Heat transfer and friction in turbulent pipe flow with variable physical

properties,” vol. 6 of Advances in Heat Transfer, pp. 503 – 564, Elsevier, 1970.

[42] A. Polyakov, “Heat transfer under supercritical pressures,” vol. 21 of Advances in Heat

Transfer, pp. 1 – 53, Elsevier, 1991.

[43] M. van Belleghem, “Ontwerp van een proefstand voor tweefasige stromingen in de

koeltechniek,” Master’s thesis, 2006.

[44] S. Garimella, “High Condensing Temperature Heat Transfer Performance of Low Critical

Temperature Refrigerants,” Air-Conditioning and Refrigeration Technology . . . ,

no. September, 2006.

[45] H. Yamaguchi, X. Zhang, K. Fujima, M. Enomoto, and N. Sawada, “Solar energy powered

Rankine cycle using supercritical CO2,” Applied Thermal Engineering, vol. 26,

pp. 2345–2354, Dec. 2006.

[46] X.-R. Zhang, H. Yamaguchi, and D. Uneno, “Experimental study on the performance of

solar Rankine system using supercritical CO2,” Renewable Energy, vol. 32, pp. 2617–2628,

Dec. 2007.

[47] a. Borsukiewiczgozdur and W. Nowak, “Comparative analysis of natural and synthetic

refrigerants in application to low temperature Clausius-Rankine cycle,” Energy, vol. 32,

pp. 344–352, Apr. 2007.

[48] a. Schuster, S. Karellas, E. Kakaras, and H. Spliethoff, “Energetic and economic

investigation of Organic Rankine Cycle applications,” Applied Thermal Engineering, vol. 29,

pp. 1809–1817, June 2009.

[49] H. Caniere, “Flow pattern mapping of horizontal evaporating refrigerant flow based on

capacitive void fraction measurements,” 2009.

[50] X. Wang and L. Zhao, “Analysis of zeotropic mixtures used in low-temperature solar

Rankine cycles for power generation,” Solar Energy, vol. 83, pp. 605–613, May 2009.

[51] H. Chen, D. Y. Goswami, and E. K. Stefanakos, “A review of thermodynamic cycles and

working fluids for the conversion of low-grade heat,” Renewable and Sustainable Energy

Reviews, vol. 14, pp. 3059–3067, Dec. 2010.

[52] S. Lecompte, H. Huisseune, and M. De Paepe, “Thermo-economic analysis of waste heat

recovery technologies for low grade waste heat,” 2013.

[53] S. Yu, H. Li, X. Lei, Y. Feng, Y. Zhang, H. He, and T. Wang, “Experimental investigation

on heat transfer characteristics of supercritical pressure water in a horizontal tube,”

Experimental Thermal and Fluid Science, vol. 50, pp. 213–221, 2013.

List of Figures

1.1 World energy use in million tonnes oil equivalent [2]. . . . . . . . . . . . . . . . . 2

1.2 Layout of a simple ORC. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3

2.1 Pressure-temperature phase diagram for CO2. . . . . . . . . . . . . . . . . . . . . 6

2.2 Temperature variation in the vapour generator for ORC, binary mixtures and

transcritical cycles[5]. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7

2.3 Schematic of subcritical cycle for isentropic working fluids. . . . . . . . . . . . . . 8

2.4 Schematic of cycle for zeotropic mixtures. . . . . . . . . . . . . . . . . . . . . . . 8

2.5 Schematic of supercritical cycle. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9

2.6 Schematic of transcritical cycle for isentropic working fluids. . . . . . . . . . . . . 9

2.7 PV-diagram of R125. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12

2.8 Ts-diagram of R125. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12

2.9 Ts-diagram of R125 as a wet fluid. . . . . . . . . . . . . . . . . . . . . . . . . . . 14

2.10 Ts-diagram of Isopentane as a dry fluid. . . . . . . . . . . . . . . . . . . . . . . . 14

2.11 Ts-diagram of R134a as an isentropic fluid. . . . . . . . . . . . . . . . . . . . . . 15

2.12 Variations of the specific heat Cp and the density ρ of CO2 at p=80 bar and

100 bar [10]. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16

2.13 Test results for P = 24,4 MPa, G = 340 kg/m2s, q = 300 kW/m2. . . . . . . . . . 23

2.14 Variations of Gr/Re2 with bulk enthalpy for various G (q”=300 kW/m2) [12]. . . 24

2.15 Schematic diagram of the FTHEL facility. . . . . . . . . . . . . . . . . . . . . . . 29

2.16 Schematic of Tsinghua University test set-up. . . . . . . . . . . . . . . . . . . . . 32

2.17 Schematic of one unit of the test section. . . . . . . . . . . . . . . . . . . . . . . . 34

2.18 Schematic of the Korean set-up for in-tube cooling of CO2. . . . . . . . . . . . . 35

3.1 Simplified schematic of the fluid loops in the set-up. . . . . . . . . . . . . . . . . 39

3.2 Schematic of the set-up. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 40

3.3 Schematic of the test section with thermocouple measuring points. . . . . . . . . 41

3.4 Schematic layout of the heating circuit. . . . . . . . . . . . . . . . . . . . . . . . 44

3.5 Schematic layout of the cooling circuit. . . . . . . . . . . . . . . . . . . . . . . . . 45

4.1 Installation of mixing valves for bypass application. . . . . . . . . . . . . . . . . . 50

149

List of Figures 150

4.2 Mixing valve characteristics. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 51

4.3 Schematic representation of the flow in the test section. . . . . . . . . . . . . . . 53

4.4 Pressure pulsations with 5 shifted diaphragms. . . . . . . . . . . . . . . . . . . . 58

4.5 DIRcalc valve selection software. . . . . . . . . . . . . . . . . . . . . . . . . . . . 62

4.6 Pressure relief valve. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 64

4.7 Stages of the accumulator [34]. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 66

4.8 Selection diagram for accumulator [34]. . . . . . . . . . . . . . . . . . . . . . . . . 69

4.9 Schematic layout of the thermal oil heating unit. . . . . . . . . . . . . . . . . . . 71

5.1 Configuration of the control and DAQ system. . . . . . . . . . . . . . . . . . . . 75

5.2 Implementation of thermocouple on the inner tube in the test section. . . . . . . 76

5.3 Configuration of thermocouples on the inner tube in the test section. . . . . . . . 77

5.4 Coriolis mass flow meter without casing. . . . . . . . . . . . . . . . . . . . . . . . 79

6.1 Human-machine interface of the installation. . . . . . . . . . . . . . . . . . . . . 82

A.1 Selection diagram for mixing valves. . . . . . . . . . . . . . . . . . . . . . . . . . 109

List of Tables

2.1 ASHRAE 34 Classification. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17

2.2 Overview of refrigerants and their main properties. . . . . . . . . . . . . . . . . . 19

2.3 Summary of the test cases by Bazargan [12]. . . . . . . . . . . . . . . . . . . . . . 24

2.4 Overview of the working fluids for the FTHEL. . . . . . . . . . . . . . . . . . . . 28

2.5 Test condition range for the tests on the FTHEL. . . . . . . . . . . . . . . . . . . 29

2.6 Overview of the working fluids for Tsinghua University test set-up. . . . . . . . . 31

2.7 Test condition range of the Tsinghua University test set-up. . . . . . . . . . . . . 32

2.8 Overview of the working fluids for the Korean test set-up. . . . . . . . . . . . . . 33

2.9 Test condition range of the Tsinghua University test set-up. . . . . . . . . . . . . 34

4.1 Overview of the refrigerant properties. . . . . . . . . . . . . . . . . . . . . . . . . 48

4.2 Overview of the design working conditions. . . . . . . . . . . . . . . . . . . . . . 48

4.3 Equivalent length for common line components. . . . . . . . . . . . . . . . . . . . 49

4.4 NPSHa for different conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . 59

4.5 Expansion valve selection data through DIRcalc. . . . . . . . . . . . . . . . . . . 63

4.6 Shut-off valve selection data through DIRcalc. . . . . . . . . . . . . . . . . . . . . 64

4.7 Legend for symbols used on figure 4.7. . . . . . . . . . . . . . . . . . . . . . . . . 66

4.8 Volumetric expansion of working fluid. . . . . . . . . . . . . . . . . . . . . . . . . 68

4.9 Properties of different heating fluid candidates at 90 C. . . . . . . . . . . . . . . 70

5.1 Temperature-resistance correlation of the Pt100 sensors. . . . . . . . . . . . . . . 78

5.2 List of thermocouples that are not in the test section. . . . . . . . . . . . . . . . 80

7.1 Absolute pressure and temperature limits of components of the high pressure section. 87

7.2 Absolute pressure and temperature limits of components of the low pressure section. 88

151

design of a test set-up for experimental investigation of

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