Master Thesis

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

Stijn Daelman

Supervisor: Prof. dr. ir. Michel De Paepe Counsellor: Marija Lazova

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

Department of Flow, Heat and Combustion Mechanics Chairman: Prof. dr. ir. Jan Vierendeels Faculty of Engineering and Architecture Academic year 2013-2014

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

Stijn Daelman

Supervisor: Prof. dr. ir. Michel De Paepe Counsellor: Marija Lazova

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

Department of Flow, Heat and Combustion Mechanics Chairman: Prof. dr. ir. Jan Vierendeels Faculty of Engineering and Architecture Academic year 2013-2014

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 Blond´efor 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

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 measurements 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 NOMENCLATURE Organic Rankine Cycles (ORC) have proven their use to Thermodynamic symbols convert ‘waste’ heat (e.g. from process industry) and other low- Cp [kJ/kgK] Heat capacity temperature heat sources (e.g. solar thermal) to usable electrical h [W/m²K] Convection coefficient energy. The efficiency of these cycles however, has still room k [W/mK] Conduction coefficient for improvement. Transcritical ORCs, in which the heat Nu [-] Nusselt number transfer from the heat source to the cycle’s working fluid P [bar] Pressure happens at supercritical conditions, can provide a higher energy Pr [-] Prandtl number conversion efficiency. q [W/m3] Heat flux In order to design these transcritical cycles, the heat transfer R [m2K/W] Thermal resistance process to the working fluids at supercritical conditions has to Re [-] Reynolds number be studied in order to provide useful correlations for design T [K] Temperature calculations of such a heat exchanger. In order to get to these v [m³/kg] Specific volume correlations, a test set-up that provides data on this heat transfer process has been designed. Abbreviations This paper reports on the design study in terms of GWP Glowal Warming Potential operational parameters, technical limitations, technical ODP Ozone Depletion Potential implementations and analytical methods used in such a set-up. ORC Organic Rankine Cycle The results obtained during this study will be used in the ORCNext project to design a supercritical heat exchanger to be Subscripts applied in a transcritical ORC. b Bulk c Critical INTRODUCTION min Minimum A shift towards the use of low-temperature energy sources, w Wall such as flue gasses, have led to a wider use of Organic Rankine 0 Ambient or reference 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 economic feasibility of the transcritical ORCs, which can be a reason for this exergy destruction. threat to this technology within the current energy markets. In order to provide a solution, a transcritical cycle can be Consequently, further application of ORCs necessitates the used. In this cycle architecture, the heat transfer during the development of a heat transfer correlation for the applied evaporation phase takes place at supercritical pressures, working fluids under the supercritical conditions. Thanks to providing a better thermal match between heat source and these correlations, heat exchangers for heat input in transcritical working fluid. This is indicated in Figure 1. ORCs can be dimensioned in an appropriate way, thus lowering The lack of knowledge of the heat transfer to the applied the costs of such installations. This can result in more efficient working fluids within these transcritical ORCs leads to an over- energy cycles, using low-temperature energy sources, whilst dimensioning of heat exchangers, thus resulting in a lower being economic sustainable as well.

Figure 1 Temperature profiles. Figure 3: Ts-diagram of R125. PROPERTIES OF SUPERCRITICAL HEAT TRANSFER Supercritical heat transfer to fluids has already been studied In the reviewed research findings, the effects of buoyancy by other researchers. However, the fluids in question were forces and acceleration effects come forward. These two mainly water, helium and CO2. The working fluids that would aspects tend to be able to have a significant impact on the heat be used within transcritical ORCs have not been tested yet. This transfer process in certain conditions. However, before doesn’t mean that the results and findings of these former discussing how these effects influence the heat transfer process, studies are of no use for these working fluids. It can be assumed a brief discussion of their existence is at hand. that the general properties that have been found for supercritical heat transfer to water, helium and CO2 are also applicable to the When neglecting the pressure losses over the heat working fluids of interest used in transcritical ORC exchanger, the heat transfer process can be modelled as a applications. variation of temperature along an isobar in the PV-diagram. The variation of the fluid’s properties along this isobar are During previous studies, thermophysical properties of fluids studied as well, and the results have proven to be quite at supercritical conditions have been mapped. These are interesting. represented on PV-diagrams and Ts-diagrams, such as At near-critical conditions (80 bar, 40°C for CO2), the represented for R125 on figures Figure 2 and Figure 3. variations of the heat capacity Cp and the density ρ show large variations for rather small variations in temperature. This can M. Bazargan [1] made a very complete review in terms of be seen on Figure 4, which is generated by Liao and Zhao [1] reviewed documents on the past experimental research on heat in their study of supercritical CO2. transfer to supercritical fluids.

Figure 4: Cp and ρ variations in function of temperature near Figure 2: PV-diagram of R125. the critical point.

To discuss the effect of buoyancy and acceleration, a horizontal tube with a fluid at supercritical conditions is being 0,923 0,613 used as an example. In this tube, a differentiation between the −3 DG  Hw− H bwµ  upper flow layers and the lower flow layers will be noticeable. (3) Nu= 4,59 e  ⋅  This is a result of the buoyancy effect and will also result into a µw TTk wb− w  temperature difference between top and bottom side of the tube. 0,231 ρ The hotter, less dense fluid will raise to the upper section of ⋅ w the tube, whereas the colder, denser fluid remains at the lower  ρb 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. A correlation from Shitsman [1], for helium, CO2 and water To be able to measure the buoyancy effect in the newly in horizontal tubes in the Dittus-Boelter form is: designed set-up, temperature measurements at each measuring 0,8 0,8 point are executed at three locations along the tube’s (4) Nub = 0,023Reb Pr min circumference. With Pr the smaller one of Pr and Pr . Next to the buoyancy effect, the rapid change of density will min b w result into an acceleration of the flow because for a constant mass flow rate, the same cross-section will be available, but a A fourth and final correlation that is used in the pre-design larger volumetric flow will be established. phase is the one from Jackson and Fewster [1]. This correlation is in the Dittus-Boelter form for constant properties convention: EXISTING CORRELATIONS FOR SUPERCRITICAL 0,3 HEAT TRANSFER 0,5 0,82 ρw By using a variety of existing heat transfer correlations for (5) Nubb= 0,0183Re Pr  supercritical heat transfer, derived from the previously studied ρb fluids, an acceptable estimation of the heat transfer coefficient (HTC) can be made, which is used to determine the working In this equation, any effect of buoyancy is being neglected. range of the set-up and the component’s specifications. EXTERNAL BOUNDARY CONDITIONS The correlation of Petukhov, Kranoschekov and Protopopov Because of the future use in a low temperature heat source for water and CO2 at supercritical conditions [1] has proven to ORCs, several boundary conditions on the parameter testing show a good predicting performance. About 85% of their data range exist. These limitations apply to the working fluid, and data of experiments by other researchers matches their geometric properties of the heat exchanger, pressure levels, correlation within a deviation of 15%. flow rates and the power to be transferred.

0,35 0,11 C λµ   The selection of the working fluids to be studied is done at p bb hand of a range of parameters. First of all, there are thermo- Nubb= Nu0,    λµ Cpb,  w  w  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 fb Reb Pr supercritical ORC’s. A third thermophysical property is the 8 fluid type: wet, dry or isentropic. This last property is of (1) Nu0,b = 0,5 2/3 fb importance when including a rotating expander in the cycle. In 12,7( Pr−+ 1) 1,07 the study of heat transfer, all three types will be used. 8 −2 f = (1, 82 log (Re )− 1, 64 ) The safety properties of the working fluids have a very 10 b significant influence on the choice of refrigerants. For now, only refrigerants of the A1 and A2L ASHRAE 34 safety group Swenson, Carver and Kakarala [1] derived a correlation for will be studied. These have a low toxicity and no or low water in horizontal tubes: flammability.

hd Finally, the environmental aspects of the refrigerants have (2) Nu = been taken into account. This includes a study of the ozon kw 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 In equation (6), F is equal to one, due to the counterflow waste heat, the most practical implementation of the heat setup of the test section. In equation (7), no fouling factor is exchangers is in the form of a shell and tube heat exchanger, taken into account since the set-up will be used in a controlled using thermal oil to transfer the heat from the heat caption lab environment and the focus of the study is on the heat elements to the actual electricity generating cycle. These heat transfer coefficient h at the inner side of the tube. exchangers are commonly used in a horizontal configuration, i The heat transfer over the test section is calculated by thus the focus of the study will be on a horizontal heat transfer discretizing the test section into multiple smaller sections, over test section. which the temperature change, and thus the change of Keeping the practical aspect in mind, the tube diameters that properties is limited. The following calculations are used implemented in Engineering Equation Solver (EES): are tested are also within the range of practical applications. th The complete circuit is constructed out of 1 1/8 inch tubing, including the current test section. For future studies of smaller Gnielinski correlation for convection coefficient ho in the diameters, only the tubing of the test section has to be replaced. annulus: −2 The goal of the flow rates to be tested are derived from the f = (1, 58 ⋅−ln( Re ) 3, 28) mass fluxes G, since this is more relevant for testing in a single- f − 2/3 tube set-up. The mass fluxes range from 100 kg/m²s to (8) (Re 1000) Pr D 1000 kg/m²s. However, for certain conditions, the minimum Nu = 2 1+ hydraulic  and maximum will have to be altered a bit, in order to be f 2/3 L 1+− 12,7 (Pr 1)  compliant with the limitations of the components in the set-up. 2

A final result of the practical application, is to use a tube-in- Correlations from Petukhov et al., Swenson et al., Shitsman tube test section, which uses thermal oil in the annulus and the et al. and Jackson and Fewster, such as listed in formula’s (1) to working refrigerant in the inner tube. This simulates the practical conditions in the ORC’s perfectly, but is harder to do (5) are used to calculate the inner convection coefficient hi . measurements on. Another solution might have been a test The results of these calculations and measurements on the section which is uniformly heated thanks to a Joule-heating setup are used to provide first impression of their applicability method, in which the electrical resistance of the tube is used as with the working fluids at hand. a heat generating source. Other set-ups using this method, however, make notice of high wall temperatures. These can PROCESSING OF RESULTS have a large effect on the heat transfer to the refrigerant in the The modified Wilson plot method of Briggs and Young is boundary layer. Therefore the final decision has been made to used to find a new correlation for the inner convection use the primary described tube-in-tube approach. coefficient in the test section. For a turbulent flow, the Nusselt numbers at the inner and outer side of the tube are: NUMERICAL METHOD 0,14 hD µ A numerical solution approach was followed in order to Nu = ii= C Rea Pr1/3 calculate the heat transfer in the test section. The heat transfer i ii i kiwµ at the side of the supercritical fluid is calculated with the (9) 0,14 different correlations that are listed above. By using this range hDo o a 1/3 µ of correlations, a reasonably window with the order of Nuo= = Coo Re Pro  magnitude for the heat transfer has been found. kow µ  These calculations are also used to find an optimal length for the set-up. For this length, a trade-off has to be made By rearranging these expressions into expressions of h and between pressure drop and temperature difference over the test i section, whilst keeping available space in mind. ho and substituting these expressions into equation (7), The heat transfer over the complete test section is calculated following expression is found: using following formula’s:

1 Di (6) Q = UAF ∆ T = 0,14 + Rwall LMTD UA mu k AC Rea Pr1/3 do iii i i  ln  muw 11di 1∆LMTD (10) (7) =+ += Do UA h A2πλ L h A Q + 0,14 ii o o  a 1/3 mu kooo A C Re o Pr o  muw

In general, a can be taken equal to 0,8 . Rearranging the REFRIGERANT CIRCUIT The refrigerant circuit on Figure 6 consists around the 4m equation to the form Y= mX + b enables the use of linear long tube-in-tube test section, in which the working fluid is regression on the data to determine m and b . heated and its temperature is continuously monitored.

0,14 1 ko a 1/3 µ −=Rw A ooo Re Pr  UA D µ owo 0,14 (11) ko a 1/3 µ Aooo Re Pr  D µ 11 owo ⋅+ 0,14 CC ki a 1/3 µ io Aiii Re Pr  Diwµ i

The coefficients achieved through this regression can then Figure 6 Refrigerant circuit. be implemented again in equation (9) to find the corresponding Nusselt numbers. 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 EXECUTION OF DESIGN AND SET-UP BUILDING two tube-in-tube preheaters, which are necessary because the During the transition phase from a theoretical design, working fluid has to be condensed before entering the based upon calculations, simulations and operational strategies circulation pump, with on/off control and an electrical preheater to the actual building of the set-up, several challenges are met. which has a finer PID temperature control. This is also one of the aspects that was of importance during this project, since these challenges will also be met on The test section is a 4m long tube-in-tube heat exchanger large-scale set-ups and are easier to anticipate when working on in which the working fluid flows through the inner tube and a lab scale. thermal oil flows through the annulus as heating fluid. Measurements of pressure and temperature of the working It is very much recommended to work closely together fluid are performed directly at in- and outlet of the test section. with the suppliers of the components that are needed in the set- Over the total length of the test section, temperature up, since supercritical fluids are not yet commonly used in measurements are performed on the heating fluid and on the current installations and might pass the boundaries that are set wall of the inner tube. by the manufacturer of the component. By monitoring the wall temperature at a 333mm interval, the temperature gradient over the test section is visualized. In The set-up at Ghent University is divided into three order to measure the influence of the buoyancy effect, three circuits: the main circuit, which contains the working fluid that measurements are performed at interval point: Top, side and is being studied, and two supporting circuits. The supporting bottom. From the temperature difference between the three circuits provide the necessary heating and cooling capacity to point, the range of working conditions in which the buoyancy the main circuit. effect plays a significant role on the heat transfer process can be determined.

Figure 7 Thermocouple configuration in the tube-in-tube test Figure 5 Schematic layout. section.

An accumulator at the suction side of the circulation pump CONCLUSION provides a NPSH, large enough to provide a stable flow. The Heat transfer to fluids at supercritical conditions has been pump is a diaphragm pump with 5 internal diaphragms, which researched over the past decades. The possible application of results into minimal flow and pressure pulsations. 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 Table 1 Properties of working fluids. described in this paper. Property HFC-134a HFC-125 HFO-1234yf Chemical formula CF3CH2F C2HF5 C3F4H2 The design and construction of a new test set-up always proves to be a challenge. For an increase in size and T 101,03°C 66,02°C 94,7°C c complexity, the challenges also prove to become harder to pc 4060,3kPa 3618kPa 3382,2kPa solve. A close collaboration between the scientists, engineers, ODP 0 0 0 suppliers and the constructor, proves to result into a good combination of insights, which are necessary for completion of GWP 1430 3500 4 the project.

The next step is to further develop and implement the

operational strategies of the set-up. 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 REFERENCES wall between heating fluid and working fluid has to be sufficiently large to minimize its influence on the total thermal [1] E. K. a. V. P. B. Petukhov, „An investigation of heat resistance. This can only be achieved by using high velocities transfer to fluids flowing in pipes under supercritical in the annulus of the heat exchanger. By using a low-viscosity conditions,” in Developments in Heat Transfer Part, vol. 3, fluid as heat source, these velocities can be achieved whilst 1961, pp. 569-578. keeping the pressure drop in the annulus within practical limits. [2] M. Bazargan, „Forced convection heat transfer to turbulent The heating capacity in the test section is controlled by flow of supercritical water in a round horizontal tube,” means of a bypass valve which controls the flow through the 2001. annulus. [3] S. L. a. T. Zhao, „Measurements of Heat Transfer Coefficients From Supercritical Carbon Dioxide Flowing in COOLING CIRCUIT Horizontal Mini/Micro Channels,” Journal of Heat The cooling circuit provides a cold source for cooling and Transfer, vol. 124, nr. 3, p. 413, 2002. condensation of the working fluid in the low-pressure section of [4] J. C. a. C. d. K. H. Swenson, „Heat transfer to supercritical the refrigerant circuit. This is required to meet the specified water in smooth-bore tubes,” Journal of Heat Transfer, vol. conditions at inlet of the pump. 87, p. 477, 1965. The cooling circuit consists of a 37kW chiller with on/off thermostatic control that is connected to an insulated buffer [5] M. Shitsman, „Heat transfer to supercritical helium, carbon vessel of 900l which stores a 30% Ethyleneglycol – 70% water dioxide and water: Analysis of thermodynamic and mixture. This solution provides a stable cooling temperature transport properties and experimental data,” Cryogenics, and results into economical operation of the chiller. 1974. As for the heating circuit, the cooling capacity is also [6] J. J. a. J. Fewster, „Forced convection data for supercritical controlled by a regulated bypass. pressure fluids,” Simon Engineering Laboratory, University of Manchester, 1975. 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.

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 supercritical 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. Chapter 1. Introduction 3

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.

Turbine/ Expander

Generator

Evaporator Condenser

Pump

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 1. Introduction 4 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

10000

solid

1000 supercritical fluid

liquid 100 critical point Pressure (bar) 10 gas triple point

1 200 250 300 350 400 Temperature (K)

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).

W T η = = 1 − C (2.1) QH TH

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. Chapter 2. Literature 7

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. Chapter 2. Literature 8

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 5 3 4

6 2 1 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 4 5 3

6 6’ 2 7 1

Figure 2.4: Schematic of cycle for zeotropic mixtures. Chapter 2. Literature 9

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.

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.

4 2 1 4’ 5

Figure 2.6: Schematic of transcritical cycle for isentropic working ﬂuids. Chapter 2. Literature 10

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. Chapter 2. Literature 11

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. Chapter 2. Literature 12

R125 103

102 Critical point 100°C Pcrit 75°C

50°C

Liquid 25°C Vapour 101

Liquid-Vapour Mixture 100 Saturated Vapour Saturated Liquid 10-1 10-4 10-3 10-2 10-1 v [m 3/kg]

Figure 2.7: PV-diagram of R125.

3500 kPa R125 3000 kPa 10150

1275 75 bar 50 bar

1050 Critical point 275 Tcrit

50 25 bar Liquid Vapour -25

-50 Liquid-Vapour 0,2 0,4 0,6 0,8 -725 Mixture

-1-050 1 bar 0,60 0,80 1,00 1,20 1,40 1,60 1,80 -75 Saturated Saturated Liquid s [kJ/kg-K] Vapour -100 0,60 0,80 1,00 1,20 1,40 1,60 1,80 2,00

s [kJ/kR125g-K] 103 Figure 2.8: Ts-diagram of R125. Chapter 2. Literature 13

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: dT • Wet: < 0 ds dT • Dry: > 0 ds ds • Isentropic: = ∞ 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. Chapter 2. Literature 14

File:H:\Thesis\EES\Simulations\TS.EESWet fluids 18/11/2013 20:50:28 Page 2 EES Ver. 9.496: #3479: For use only by Michel De Paepe, Ghent University, Belgium 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.

R125 150

125

100

50 2400 kPa 25 1000 kPa 0

-25 320 kPa

-50 70 kPa 0,2 0,4 0,6 0,8 -75

-100 0,50 0,75 1,00 1,25 1,50 1,75 2,00 2,25 s [kJ/kg-K]

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

File:(Untitled) 18/11/2013 20:55:10 Page 2 Dry fluids EES Ver. 9.496: #3479: For use only by Michel De Paepe, Ghent University, Belgium 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.

Isopentane 250

200

2300 kPa 150

1000 kPa 100

350 kPa 50

90 kPa 0 0,2 0,4 0,6 0,8

-50 -2,0 -1,5 -1,0 -0,5 0,0 0,5 1,0 s [kJ/kg-K]

Figure 2.10: Ts-diagram of Isopentane as a dry ﬂuid. Chapter 2. Literature 15

Isentropic ﬂuids

File:H:\Thesis\EES\Simulations\TS.EESAn isentropic ﬂuid has a quasi-vertical saturation vapour curve at the expansion side18/11/2013 of the 20:50:29 Page 3 EES Ver. 9.496: #3479: For use only by Michel De Paepe, Ghent University, Belgium diagram. The saturation vapour curve coincides with a theoretical isentropic expansion line. R134a can be used as an example for this type of ﬂuids.

R134a 250

200

150

100 2700 kPa

50 1100 kPa

0 340 kPa

70 kPa -50 0,2 0,4 0,6 0,8

-100 -0,25 0,00 0,25 0,50 0,75 1,00 1,25 1,50 1,75 s [kJ/kg-K]

Figure 2.11: Ts-diagram of R134a as an isentropic ﬂuid. Chapter 2. Literature 16

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]. Chapter 2. Literature 17

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. Chapter 2. Literature 18

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.

R T 0 agas [gas] dt GW Pgas = (2.2) R T a dt 0 CO2 [CO2]

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. Chapter 2. Literature 19

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.

◦ ASHRAE 34 GWP Name Type Tcrit [ C] pcrit [bar] ODP 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 Chapter 2. Literature 20

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 · RebP rc (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.

P rmin, 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 convective 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]. Chapter 2. Literature 21

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. Chapter 2. Literature 22

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 phenomenon 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. Chapter 2. Literature 23

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 (b) Diﬀerences of local heat transfer coef- with bulk enthalpy. 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 Chapter 2. Literature 24 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.

3 gβ (Tw − T ) d Gr = b (2.3) ν2

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].

2 q” ≈ 300 kW/m G ∆Twall,max Twall,min for Peak HTC bottom Peak HTC 2 ◦ 2 in all cases [kg/m s] [ C] Tw,top = Tw,bottom [kW/m K] 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]. Chapter 2. Literature 25

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]:

−5 2,75 0,5 h −1/8 2/3 i Grth = 3 · 10 Reb P r 1 + 2, 4Reb P r − 1 (2.4)

In this equation, P r is deﬁned as:

iw − i µ P r = b · b (2.5) Tw − Tb λb

With the assumption that q” = λ∂T ∂x near the wall, the expression for the Grashof number (equation 2.3) is modiﬁed to:

gβq”d4 Grq = 2 (2.6) νb λb where 1 ρ − ρw β = · b (2.7) ρfilm Tw − Tb

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]. Chapter 2. Literature 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 supercritical 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.

0,35 −0,33 0,11 Cp λb µb Nub = Nu0,b (2.8) Cp,b λw µw   fb  RebP r  Nu  8  0,b =  0,5  (2.9) f 2/3 12, 7 b P r − 1 + 1, 07 8 −2 f = (1, 82log10 (Reb) − 1, 64) (2.10)

This correlation is assumed to be valid within following ranges:

4 5 2 · 10

0, 85

2.4.3.2 Swenson, Carver and Kakarala

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

0,923 0,613 0,231 hD DG Hw − H µw ρw Nu = = 0, 00459 b · · (2.11) kw µw Tw − Tb kw ρb

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: 0,8 0,8 Nub = 0, 023Reb P rmin (2.12)

In the above expression, P rmin is the smaller one of P rb and P rw.

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: 0,3 0,82 0,5 ρw Nub = 0, 0183Reb P r (2.13) ρb The eﬀect of buoyancy is neglected in the equation. Chapter 2. Literature 28

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. Chapter 2. Literature 29

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. Chapter 2. Literature 30

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. Chapter 2. Literature 31

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. Chapter 2. Literature 32

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

2.5.2.3 Test conditions

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

Parameter Lower limit Higher limit T 25 ◦C 200 ◦C p 55 bar 100 bar Q˙ 110 kW/m2 1800 kW/m2 G˙ 600 kg/m2s 2000 kg/m2s Chapter 2. Literature 33

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. Chapter 2. Literature 34

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.

2.5.3.3 Test conditions

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

Parameter Lower limit Higher limit T ◦ ◦ in,CO2 50 C 80 C p 75 bar 88 bar G˙ / 2 / 2 CO2 225 kg m s 450 kg m s ◦ ◦ Tin,water 7 C 12 C

G˙ water 60 g/s 120 g/s Chapter 2. Literature 35

Figure 2.18: Schematic of the Korean set-up for in-tube cooling of CO2. Chapter 2. Literature 36 Chapter 3

Test set-up description

This chapter gives a general introduction to the new test set-up and describes its working principles. 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 )) 3 ◦ – Minimum: 0,1536 m /h (R134a Tin = 50 C) 3 ◦ – Maximum: 1,945 m /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 Chapter 3. Test set-up description 39

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.

Cooling circuit Refrigerant circuit Heating circuit

Cold water Test section Hot water Cooler system system

Preheater

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. Chapter 3. Test set-up description 40 r e t a e h

l i O r l e a F T T t c i a r e t c h e e l r E p T F P T p m u p

m u u c a V p m u p

1 n 2

o r i r e t e t a t l a a u e n e c h o h r i i e T e t r C r c p p

e S e

b t b s u u t e t - - T n n i i - - r e e o b t b u a u l T T u m u c c A P T / r r e T e y t r l i D F t n a P r e g i r f r e e / s R r n T e l e o d o n C o C

l f

e f o v e r o l r t T - a f t u n v e s u

o i e s l c h n

v e e s e l o

r r e i v

a r l n p s y

v u a o t n i s v w e a t s f o c p e a L x e r S s e P l e s s r e e l v l

i r h e C f f u B

Figure 3.2: Schematic of the set-up. Chapter 3. Test set-up description 41

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.

Heating fluid inlet Thermocouples

Working fluid outlet Working fluid inlet

Heating fluid outlet

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. Chapter 3. Test set-up description 42

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. Chapter 3. Test set-up description 43

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. Chapter 3. Test set-up description 44

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. Chapter 3. Test set-up description 45

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

Buffer vessel

Chiller

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 3. Test set-up description 46 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. Chapter 4. Design Procedure 49

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 corresponding 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:

L ρv2 ∆p = 4f m (4.1) Dh 2

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:

Q˙ v = (4.2) m A 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): 16 f = (4.4) Re 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 Chapter 4. Design Procedure 50

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 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: ∆p1 a = (4.5) ∆p1 + ∆p2

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. Chapter 4. Design Procedure 51

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 Chapter 4. Design Procedure 52

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. Chapter 4. Design Procedure 53

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. Chapter 4. Design Procedure 54

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 · ∆TLMT D (4.6) d ln u 1 1 d 1 ∆T = + i + = LMT D (4.7) UA hiAi 2πλL Auhu Q˙

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 ◦ 3 of 70 C at a pressure of 1, 05 · pcrit and an oil ﬂow of 10 m /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. Chapter 4. Design Procedure 55

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 considered, 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:

Twall − Twf Q = (4.10) Rtot

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.

ln(Do/D) R = D (4.11) cond 2λ 1 Rconv = (4.12) hiπDL 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 Twall − Twf ln (Do/D) = − D (4.13) hi q 2λ Chapter 4. Design Procedure 56

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: D2 − d2 4π i o 4A 4 Dhyd = = (4.14) Pwet π(do + Di) Thermal diameter: D2 − d2 4π i o 4A 4 Dt = = (4.15) PHT πdo The Gnielinski correlation yields:

f = (1, 58 · ln(Re) − 3, 28)−2 (4.16) f (Re − 1000)P r 2/3! Dhyd Nu = 2 1 + (4.17) rf L 1 + 12, 7 P r2/3 − 1 2 In equation 4.17, the correction factor can be neglected because the length L is much larger than the hydraulic diameter Dhyd. Chapter 4. Design Procedure 57

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. Chapter 4. Design Procedure 58

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.

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. Chapter 4. Design Procedure 59

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:

2 pi Vi pv NPSHa = + − (4.18) ρg 2g ρg with pi the pressure and Vi the velocity at the pump inlet and pv the vapour pressure.

Table 4.4: NPSHa for diﬀerent conditions ◦ kg 2 R134a T [ C] pv [bar] p [bar] ρ G kg/m s v [m/s] NPSHa[m] m3 Charging 20 5,721 7 1243 1000 0,45 10,50 Testing 93 34,52 35 807,7 1000 0,70 6,08 ◦ kg 2 R125 T [ C] pv [bar] p [bar] ρ G kg/m s v [m/s] NPSHa[m] m3 Charging 20 12,06 14 1221 1000 0,46 16,21 Testing 60 31,71 35 916,1 1000 0,61 36,63 ◦ kg 2 R1234yf T [ C] pv [bar] p [bar] ρ G kg/m s v [m/s] NPSHa[m] m3 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. Chapter 4. Design Procedure 60

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.

G ATS vinlet = (4.19) ρ Ainlet 2 πDTS G 4 = 2 (4.20) ρ πDinlet 4 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. Chapter 4. Design Procedure 61

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, entering 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. Chapter 4. Design Procedure 62

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, independent 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. Chapter 4. Design Procedure 63

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. Chapter 4. Design Procedure 64

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 mechanisms (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. Chapter 4. Design Procedure 65

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 condensation 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. Chapter 4. Design Procedure 66

P0 P1 P2

V0 V1 V2

A B C 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.

l π · D2 V = · (4.21) 2 4 d2π = 12, 5 · i (4.22) 4 = 6 l (4.23) Chapter 4. Design Procedure 67

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:

mtest,lp Vtest,crit = (4.29) ρtest,lp mwf − mtest,hp = (4.30) ρtest,lp

∆Vexp = Vtest,lp − Vcharge,lp (4.31)

The results of these calculations are displayed in table 4.8 for each of the refrigerants. Chapter 4. Design Procedure 68

Table 4.8: Volumetric expansion of working ﬂuid.

kg R134a Charging T [◦C] p [bar] ρ V [l] m [kg] m3 High pressure 20 40,6 1243 10 12,43 Low pressure 20 7 1226 10 12,26 kg R134a Testing critical T [◦C] p [bar] ρ V [l] m [kg] m3 High pressure 115 40,6 221,1 10 2,21 Low pressure 93 35 807,7 27,8 22,48 kg R125 Charging T [◦C] p [bar] ρ V [l] m [kg] m3 High pressure 20 36,2 1247 10 12,47 Low pressure 20 14 1221 10 12,21 kg R125 Testing critical T [◦C] p [bar] ρ V [l] m [kg] m3 High pressure 75 36,2 271 10 2,71 Low pressure 60 35 916,1 24 21,97 kg R1234yf Charging T [◦C] p [bar] ρ V [l] m [kg] m3 High pressure 20 33,8 1119 10 11,19 Low pressure 20 7 1110 10 11,10 kg R1234yf Testing critical T [◦C] p [bar] ρ V [l] m [kg] m3 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. Chapter 4. Design Procedure 69

Accumulator size in litre

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. Chapter 4. Design Procedure 70

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[P a · 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. Chapter 4. Design Procedure 71

TH1 PA MN P = Motor pump assembly TC2 PT1 E C H = Heat exchanger PA = Automatic air bleeding NB TH1 = S afety thermostat Utilisation PT1 = Temperature sensor ECH Using TC 2 = Limiter thermocouple AC = Tank P NB = Low level switch R 5 = Level tap B1 = Draining bush AC R5 BC V = S peed breaking box MN = Manometer BCV Utilisation B1 Using

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. Chapter 4. Design Procedure 72

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. Chapter 4. Design Procedure 73

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 4. Design Procedure 74 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 Network compactRIO Computer

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 implementation 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. Chapter 5. Data-acquisition hardware 77

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.

s P 2 2 Yi − Axi − Bxi − C σTC = (5.1) N − 3

Where N represents the number of measured values per temperature, per thermocouple. Chapter 5. Data-acquisition hardware 78

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. Chapter 5. Data-acquisition hardware 79

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. Chapter 5. Data-acquisition hardware 80

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 bulk oil pos. 4 403 Thermocouple TC bulk oil pos. 4 404 Thermocouple TC bulk oil pos. 7 405 Thermocouple TC bulk oil pos. 7 406 Thermocouple TC bulk oil pos. 7 407 Thermocouple TC bulk oil pos. 10 408 Thermocouple TC bulk oil pos. 10 409 Thermocouple TC bulk oil pos. 10 410 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. Chapter 6. Software 83

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 6. Software 84 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 necessary. 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. Chapter 7. Operational procedures 87

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. Chapter 7. Operational procedures 88

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 pmax the accumulator is limited to avoid over-compression of the internal bladder. This has pmin 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.

T011 + T012 + T013 T = (8.1) h,b 3 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.

Q˙ hf = Q˙ wf + Q˙ loss (8.2)

The heat loss Q˙ loss 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

2πλ Lsection Q˙ = insulation (T − T ) (8.3) loss D b a ln o,insulation Di,insulation The heat transferred to the working ﬂuid in the corresponding section is expressed as:

Q˙ wf = Q˙ hf − Q˙ loss (8.4)

=m ˙ wf · ∆Hwf (8.5)

The heat ﬂux Q˙ wf ﬂows through the wall of the inner tube by conduction. This results in following equation:

2πλ Lsection Q˙ = tube (T − T ) (8.6) wf d w,o w,i ln o di 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. Chapter 8. Data reduction 91

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: 0,14 hiDi a 1/3 µ Nui = = CiRei P ri (8.8) ki µw i 0,14 hoDT a 1/3 µ Nuo = = CoReoP ro (8.9) ko µw o 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 1 1 = + Rw + (8.10) UA hiAi hoAo With UA derived from the LMTD equation:

Qw m˙ wf cp,wf ∆Twf UA = = (8.11) ∆TLMT D ∆TLMT D

The thermal resistance of the wall Rw in equation 8.10 is expressed as:

ln(Do/Di) Rw = (8.12) 2πLk

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 Di Do = + R + (8.13) UA 0,14 w 0,14 a 1/3 µ a 1/3 µ kiAiCiRei P ri koAoCoReoP ro µw i µw o The power factor a is generally equal to 0,8. A ﬁrst rearrangement of equation 8.13 yields: 0,14 ko a 1/3 µ " 0,14# AoReoP ro 1 ko µ Do µw 1 1 − R A ReaP r1/3 = o · + (8.14) UA w D o o o µ 0,14 C C o w o ki a 1/3 µ i o AiRei P ri Di µw i 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 8. Data reduction 92 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:

s ∂q 2 ∂q 2 ∆q(x, y, ...) = ∆x + ∆y + ... (9.1) ∂x ∂y 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 temperature 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: r 1 X N ∆TC = Yi − AXi − BXi − C (9.3) n − 3 i=1

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].

q 2 2 δtotal = ∆TC + ∆systematic q 2 2 = ∆TC + ∆F luke

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) dT = C (9.4) dz p dz The local heat ﬂux: dT mC˙ p q(z) = dz (9.5) πD Chapter 9. Error analysis 95

Using equation 9.1, this yields:

s ∂q 2 ∂q 2 ∆q(z) = ∆(˙m) + ∆T (9.6) ∂m˙ ∂T

The internal convection coeﬃcient hi from equation 4.13 can be rewritten as:

q hi = (9.7) Twall − Twf − Rq

Application of formula 9.1 yields:

v u 2 2 u −q q u 2 ∆Twall + 2 ∆Twf u (Twall − Twf − Rq) (Twall − Twf − Rq) ∆h = u (9.8) i u 2 t Twall − Twf + 2 ∆q (Twall − Twf − Rq)

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: v u " #2 " #2 u µ 0,14 µ 0,14 u ReaP r1/3 C CaRea−1P r1/3 Re u i ∆ + ∆ u µw µw ∆Nu = u (9.9) u " 0,14 #2 " 0,14 #2 u µ 1 µ t + CRealn(Re)P r1/3 ∆a + CRea P r−2/3 ∆P r µw 3 µw

The error on Re and P r 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

r s 1 X N N ∆m = (yi − b − mXi) (9.10) N − 2 i=1 N P x2 − (P x)2 r s P 2 1 X N x ∆b = (yi − b − mXi) (9.11) N − 2 i=1 N P x2 − (P x)2 Chapter 9. Error analysis 96 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 components (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 .

CBXP27 Brazed Plate Heat Exchanger

General information Alfa Laval introduced its first brazed plate heat exchanger (BHE) in 1977 and has since continuously developed and optimized its performance and reliability.

Brazing the stainless steel plates together eliminates the need for gaskets and thick frame plates. The brazing material seals and holds the plates together at the contact points ensuring optimal heat transfer efficiency and pressure resistance. The plate design guarantees the longest possible life.

The design options of the brazed heat exchanger are extensive. Different plate patterns are available for various duties and performance specifications. You can choose a standard configuration BHE, or a unit designed according to your 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 Examples of connections - Economizing - Sub cooling -Condensing

Working principles The heating surface consists of thin corrugated metal plates stacked on top of each other. Channels are formed between the plates and corner ports are arranged so that the two media External Internal Soldering Welding threaded threaeded flow through alternate channels, usually in countercurrent flow for the most efficient heat transfer process. * More connections are available on request.

Standard design The plate pack is covered by cover plates. Connections are located in the front or rear cover plate. To improve the heat transfer design, the channel plates are corrugated.

Particulars required for quotation To enable Alfa Laval’s representative to make a specific quotation, 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 CBXP27 - PED approval pressure/temperature graph* Standard data

Min. working temperature see graph Max. working temperature see graph Min. working pressure vacuum Max. working pressure see graph Volume 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 6 Max. nbr of plates 150 * Water at 5 m/s (16.4 ft/s) (connection velocity)

Standard materials

CBXP27 - KHK 150°C (302°F) approval pressure/temperature graph* Cover plates Stainless steel Connections Stainless steel Plates Stainless steel Brazing filler Copper

Standard dimensions mm (inch)

111 (4.37) 50 (1.97) A

S4 S1 Standard dimensions and weight*

Ameasuremm = 13 + (2.4 * n) (+/-3 mm) Ameasureinch = 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) 250 (9.84) 310 (12.2) * Excluding connections

S3 S2

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 Laval Up-to-date AlfaLaval contact details for all countries are always available on our website 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 x a m p Δ

FLOW Pressure drop kPa (100 kPa = 1bar = ~ 10 m H2O)

Figure A.1: Selection diagram for mixing valves.

Appendix A. Hardware 111

A.3 Mass ﬂow meter error analysis Rheonik Flowmeter Summary Tag No/Project: Universiteit Gent (Printdate: 2-jan-2014 - Version: 2.1)

Application Data

Fluid/Gas Name: R125 Viscosity: 0,0002 Pa s Max. Rate: 30,0000 kg/min Density: 1.251,00 kg/m³ Max. Pressure: 50,00 bar Max. Temp.: 110,00 Celsius

Application Data in Rheonik Standard Units

Max. Rate: 30,0000 kg/min Viscosity: 0,1668 cP Max. Pressure: 50,00 bar Density: 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 630 Material: 1.4571 (316Ti) [M1] (std.) Connection: 1 inch ANSI 600 RF/SF [A3] (std.)

Your Given Rate

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 30,0000 0,16 0,12 2,54 187241 More Rates

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 2,0000 0,50 0,00 0,17 12483 5,1111 0,19 0,00 0,43 31900 8,2222 0,18 0,01 0,70 51318 11,3333 0,18 0,02 0,96 70736 14,4444 0,17 0,03 1,23 90153 17,5556 0,17 0,04 1,49 109571 20,6667 0,17 0,06 1,75 128988 23,7778 0,17 0,08 2,02 148406 26,8889 0,17 0,10 2,28 167824 30,0000 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 Rheonik Flowmeter Summary Tag No/Project: Universiteit Gent (Printdate: 2-jan-2014 - Version: 2.1)

Application Data

Fluid/Gas Name: R125 Viscosity: 0,0000 Pa s Max. Rate: 30,0000 kg/min Density: 352,00 kg/m³ Max. Pressure: 50,00 bar Max. Temp.: 110,00 Celsius

Application Data in Rheonik Standard Units

Max. Rate: 30,0000 kg/min Viscosity: 0,0267 cP Max. Pressure: 50,00 bar Density: 352,00 kg/m³ Max. Temp.: 110,00 Celsius

Your Selection

Meter Name: RHM12 [M#12] P-Sealless [PF0] - wt 1.00 [P1] -- ID 630 Material: 1.4571 (316Ti) [M1] (std.) Connection: 1 inch ANSI 600 RF/SF [A3] (std.)

Your Given Rate

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 30,0000 0,16 0,35 9,04 1061033 More Rates

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 2,0000 0,50 0,00 0,60 70736 5,1111 0,19 0,01 1,54 180769 8,2222 0,18 0,03 2,48 290802 11,3333 0,18 0,05 3,42 400835 14,4444 0,17 0,09 4,35 510868 17,5556 0,17 0,12 5,29 620901 20,6667 0,17 0,17 6,23 730934 23,7778 0,17 0,22 7,17 840967 26,8889 0,17 0,28 8,11 951000 30,0000 0,16 0,35 9,04 1061033

Repeatability is 4 times better than the accuracy / linearity. For gas depending on pressure.

www.rheonik.com - the mass flowmeter experts Rheonik Flowmeter Summary Tag No/Project: Universiteit Gent (Printdate: 2-jan-2014 - Version: 2.1)

Application Data

Fluid/Gas Name: R134a Viscosity: 0,0002 Pa s Max. Rate: 30,0000 kg/min Density: 1.245,00 kg/m³ Max. Pressure: 50,00 bar Max. Temp.: 110,00 Celsius

Application Data in Rheonik Standard Units

Max. Rate: 30,0000 kg/min Viscosity: 0,2210 cP Max. Pressure: 50,00 bar Density: 1.245,00 kg/m³ Max. Temp.: 110,00 Celsius

Your Selection

Meter Name: RHM12 [M#12] P-Sealless [PF0] - wt 1.00 [P1] -- ID 630 Material: 1.4571 (316Ti) [M1] (std.) Connection: 1 inch ANSI 600 RF/SF [A3] (std.)

Your Given Rate

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 30,0000 0,16 0,12 2,56 144686 More Rates

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 2,0000 0,50 0,00 0,17 9646 5,1111 0,19 0,00 0,44 24650 8,2222 0,18 0,01 0,70 39655 11,3333 0,18 0,02 0,97 54659 14,4444 0,17 0,03 1,23 69664 17,5556 0,17 0,05 1,50 84668 20,6667 0,17 0,06 1,76 99673 23,7778 0,17 0,08 2,03 114677 26,8889 0,17 0,10 2,29 129682 30,0000 0,16 0,12 2,56 144686

Repeatability is 4 times better than the accuracy / linearity. For gas depending on pressure.

www.rheonik.com - the mass flowmeter experts Rheonik Flowmeter Summary Tag No/Project: Universiteit Gent (Printdate: 2-jan-2014 - Version: 2.1)

Application Data

Fluid/Gas Name: R134a Viscosity: 0,0000 Pa s Max. Rate: 30,0000 kg/min Density: 352,00 kg/m³ Max. Pressure: 50,00 bar Max. Temp.: 110,00 Celsius

Application Data in Rheonik Standard Units

Max. Rate: 30,0000 kg/min Viscosity: 0,0248 cP Max. Pressure: 50,00 bar Density: 352,00 kg/m³ Max. Temp.: 110,00 Celsius

Your Selection

Meter Name: RHM12 [M#12] P-Sealless [PF0] - wt 1.00 [P1] -- ID 630 Material: 1.4571 (316Ti) [M1] (std.) Connection: 1 inch ANSI 600 RF/SF [A3] (std.)

Your Given Rate

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 30,0000 0,16 0,34 9,04 1591549 More Rates

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 2,0000 0,50 0,00 0,60 106103 5,1111 0,19 0,01 1,54 271153 8,2222 0,18 0,03 2,48 436202 11,3333 0,18 0,05 3,42 601252 14,4444 0,17 0,08 4,35 766302 17,5556 0,17 0,12 5,29 931351 20,6667 0,17 0,16 6,23 1096401 23,7778 0,17 0,21 7,17 1261450 26,8889 0,17 0,27 8,11 1426500 30,0000 0,16 0,34 9,04 1591549

Repeatability is 4 times better than the accuracy / linearity. For gas depending on pressure.

www.rheonik.com - the mass flowmeter experts Rheonik Flowmeter Summary Tag No/Project: Universiteit Gent (Printdate: 2-jan-2014 - Version: 2.1)

Application Data

Fluid/Gas Name: Therminol 20 gr Viscosity: 0,0092 Pa s Max. Rate: 140,0000 kg/min Density: 857,00 kg/m³ Max. Pressure: 5,00 bar Max. Temp.: 130,00 Celsius

Application Data in Rheonik Standard Units

Max. Rate: 140,0000 kg/min Viscosity: 9,1900 cP Max. Pressure: 5,00 bar Density: 857,00 kg/m³ Max. Temp.: 130,00 Celsius

Your Selection

Meter Name: RHM20 [M#20] P-Sealless [PF0] - wt 1.00 [P1] -- ID 810 Material: 1.4571 (316Ti) [M1] (std.) Connection: 2 inch ANSI 150 RF/SF [A1] (std.)

Your Given Rate

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 140,0000 0,16 0,40 5,35 8980 More Rates

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 6,0000 0,50 0,00 0,23 385 20,8889 0,19 0,01 0,80 1340 35,7778 0,18 0,02 1,37 2295 50,6667 0,17 0,07 1,94 3250 65,5556 0,17 0,10 2,51 4205 80,4444 0,17 0,15 3,07 5160 95,3333 0,17 0,20 3,64 6115 110,2222 0,17 0,26 4,21 7070 125,1111 0,16 0,33 4,78 8025 140,0000 0,16 0,40 5,35 8980

Repeatability is 4 times better than the accuracy / linearity. For gas depending on pressure.

www.rheonik.com - the mass flowmeter experts Rheonik Flowmeter Summary Tag No/Project: Universiteit Gent (Printdate: 2-jan-2014 - Version: 2.1)

Application Data

Fluid/Gas Name: Therminol 130 gr Viscosity: 0,0007 Pa s Max. Rate: 140,0000 kg/min Density: 765,00 kg/m³ Max. Pressure: 5,00 bar Max. Temp.: 130,00 Celsius

Application Data in Rheonik Standard Units

Max. Rate: 140,0000 kg/min Viscosity: 0,6890 cP Max. Pressure: 5,00 bar Density: 765,00 kg/m³ Max. Temp.: 130,00 Celsius

Your Selection

Meter Name: RHM20 [M#20] P-Sealless [PF0] - wt 1.00 [P1] -- ID 810 Material: 1.4571 (316Ti) [M1] (std.) Connection: 2 inch ANSI 150 RF/SF [A1] (std.)

Your Given Rate

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 140,0000 0,16 0,28 5,99 119601 More Rates

Flow Rate Mass Flow Accuracy ∆p Velocity Reynolds kg/min % Rate bar m/s Number 6,0000 0,50 0,00 0,26 5126 20,8889 0,19 0,01 0,89 17845 35,7778 0,18 0,02 1,53 30565 50,6667 0,17 0,04 2,17 43284 65,5556 0,17 0,07 2,81 56004 80,4444 0,17 0,10 3,44 68723 95,3333 0,17 0,14 4,08 81443 110,2222 0,17 0,18 4,72 94162 125,1111 0,16 0,22 5,36 106882 140,0000 0,16 0,28 5,99 119601

Repeatability is 4 times better than the accuracy / linearity. For gas depending on pressure.

www.rheonik.com - the mass flowmeter experts Appendix A. Hardware 118 Appendix B

Process ﬂow diagram

119

T P T T T P T

Test Section

T Electrical preheater

Pressure control T expansion valve Tube-in-tube preheater 2

F Safety relief valve

Low pressure section shut-off valve Tube-in-tube preheater 1

Cooler/ Condenser Accumulator

T P F Oil heater

Buffer vessel Filter/ Dryer Circulation pump

Vacuum pump Refrigerant

Chiller

Appendix C

Test section piece- and assembly drawings

123

50 65 25 65

57 53 A A

8,8 32,5 G2" Scale: 1:1 Scale: 15 SECTION A-A Title: Measuring block type 1 1 type block Measuring Title: Firm: Daelman Stijn Drawing/Design: UGent B B G 1/8BSPT SECTION B-B G 1/8BSPT G 1/8BSPT 8,7

15 C 22 SECTION C-C SECTION A-A B B G 1/8"BSPT G 1/8"BSPT 22 G 1/8"BSPT G 1/8"BSPT C G2" Scale: 1:1 Scale:

65 Title: Measuring block type 2 2 type block Measuring Title: Firm: Daelman Stijn Drawing/Design:

UGent 53 57

A 18,8 18,8 18,8 SECTION B-B M6

18,8

15 15 A 34,9 28,6

5 A 4,2

65 41

SECTION A-A

40,6

Title: End cap for tube-in-tube heat exchangers Drawing/Design: Stijn Daelman

Scale: 1:1 Firm: 60,3 G2" UGent

18 A 6 F 7 8 9 10 11 12 13 14 15 16 17 C 3

1 3 4 4 5 4 4 5 4 4 5 4 4 1

G G B B D D

ITEM NO. PART NUMBER QTY. DETAIL F DETAIL A DETAIL C 1 Outer tube top 2 SCALE 1 : 2 SCALE 1 : 2 SCALE 1 : 2 2 2 Sealing ring 2 2 3 Tee 2 Small measuring 4 block 8 Large measuring 5 block 3 6 Outer tube 1 1 7 Outer tube 2 1 8 Outer tube 3 1 SECTION G-G SECTION B-B SECTION D-D 9 Outer tube 4 1 SCALE 1 : 2 SCALE 1 : 2 SCALE 1 : 2 10 Outer tube 5 1 11 Outer tube 6 1 Date: Drawing/Design: Proofread: Drawing ID: Stijn Daelman 12 Outer tube 7 1 13 Outer tube 8 1 14 Outer tube 9 1 Title: Assembly drawing - Test Section 15 Outer tube 10 1 16 Outer tube 11 1 Scale: Firma: Department of Flow, Heat and Combustion Mechanics 17 Outer tube 12 1 Applied Thermodynamics and Heat Transfer 18 Inner tube 1 UGent www.ugent.be/ea/floheacom

Appendix D

Complete set-up 3D renders

133

Appendix E

Software

E.1 EES-simulation code

1 ”EES Simulation to size a tube−in−tube heat exchanger” 2 3 ”WORKING CONDITIONS” 4 5 ”Hot Source” 6 ”Use of Therminol ADX10 as hot fluid. This is not included in EES datasets yet, so manual implemented correlations are being used.” 7 8 T h f in =130 [C] ”Inlet temperature of the heating fluid” 9 P h f in=1e5 [Pa] ”Water at 2 bar to be able to go to temperatures higher than 100◦C” 10 Rho hf = Rho mean b hf[1] ”Density of the heating fluid at inlet temperature” 11 v h f i n = m dot hf / Rho hf / S an ”Entrance velocity of the heating fluid” 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 h f i n+T hf out)/2 ”Mean temperature of the heating f l u i d ” 16 P hf mean = ( P h f i n+P hf out)/2 ”Mean pressure of the heating fluid ” 17 Cp b hf = ( Cp mean b hf [1]+ Cp mean b hf [N−1])/2 ”Average heat capacity of the heating fluid” 18 19 ”Working fluid” 20 WF$= ’r125’ ”Working fluid” 21 22 T wf in = 70 [C] ”Inlet temperature working fluid” 23 T wf out = T wf in+DeltaT wf ”Outlet temperature of the working fluid ,

139 Appendix E. Software 140

defined by inlet temperature and temperature resolution” 24 25 P wf in = P c r i t (WF$) ∗1,05 ”supercritical working pressure” 26 G = m dot wf / S in ”Mass flux in the inner tube” 27 G= 1000 [kg/mˆ2−s ] 28 Rho wf in = Density(WF$;T=T wf in ;P=P wf in) ”Density of the working fluid at inlet” 29 v i n w f = G/ Rho wf in ”Inlet velocity working fluid” 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 of the working f l u i d ” 34 P wf mean = ( P wf in+P wf out)/2 ”Mean pressure of the working fluid ” 35 Cp b wf = Cp(WF$;T=T wf b mean ; P=P wf mean) ”Average heat capacity of the working fluid” 36 37 38 ”Calculation of state points at in− and o u t l e t ” 39 h wf out = Enthalpy(WF$;T= T wf out ; P=P wf[1]) ”calculate enthalpy of working fluid at the outlet using the corresponding temperature and p r e s s u r e ” 40 h w f in=Enthalpy (WF$;T=T wf in ;P=P wf[N]) ”calculate enthalpy of working fluid at the inlet using the corresponding temperature and p r e s s u 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 of the tube” 47 t = 1 ,9∗10ˆ( −3) [m] ”Thickness of tube” 48 d i = d o −2∗t ”Inside diameter of inner tube” 49 50 D out = (2+1/8) ∗10ˆ( −3) ∗25,4 [m] ”Outer tube outside diameter” 51 t o u t = 1 ,47∗10ˆ( −3) [m] 52 D in= D out−2∗ t out ”Outer tube inside diameter” 53 54 L=4[m] ”Total length of the test section” 55 56 S an = ( pi ∗ D inˆ2/4−pi ∗ d oˆ2/4) ”Annulus flow cross s e c t i o n ” 57 S i n = ( pi ∗ d iˆ2/4) ”Tube flow flow crsoss section” 58 Appendix E. Software 141

59 ”Thermodynamical dimensions” 60 61 D h = 4∗ pi ∗(D inˆ2−d o ˆ2) /( pi ∗( d o+D in ) ∗4) ”Hydraulic 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 conductivity of K65” 65 66 ”Discretisation of the heat exchanger” 67 68 N= 20 ”10 Discrete sections of 20 cm” 69 70 T h f i n = T hf[1] ”Point 1 is the inlet of the heating fluid” 71 T hf out = T hf[N] ”Point N is the outlet of the heating fluid” 72 T wf in = T wf[N] ”Point N is the inlet of the working fluid” 73 T wf out = T wf[1] ”Point 1 is the outlet of the working fluid” 74 75 P h f i n = P hf[1] ”Inlet pressure of the heating fluid” 76 P hf out = P hf[N] ”Outlet pressure of the heating fluid” 77 78 P wf in = P wf[N] ”Inlet pressure of the working fluid” 79 P wf out = P wf[1] ”Outlet pressure of the working fluid” 80 81 L[1]=0 ”First point at beginning of heat exchanger” 82 L section= L/(N−1) ”Length of one section” 83 A i = d i ∗ pi ∗ L section ”Heat exchanging area in the WF side” 84 A o = d o ∗ pi ∗ L section ”Heat exchanging area in the HF side” 85 86 Q tot[1] = 0 ”No heat exchanged at starting point” 87 Q netto = Q tot[N] ”Net heat exchanged” 88 89 ”Calculation of heat transfer per division” 90 91 duplicate i=1;N−1 92 93 ”Heat balance for each section” 94 95 T hf mean b [ i ] = ( T hf [ i ]+ T hf[i+1])/2 ”Mean temperature heating fluid per division” 96 T wf mean b [ i ] = ( T wf [ i ]+T wf[i+1])/2 ”Mean temperature working fluid per division” 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 Appendix E. Software 142

101 Q[ i ] = HTC mean hf [ i ] ∗ A o ∗(T hf mean b [ i ]−T hf mean w [ i ] ) ”Heat transfer heating fluid to tube” 102 Q[ i ] = HTC mean wf [ i ] ∗ A i ∗(T wf mean w [ i ]−T wf mean b [ i ] ) ”Heat transfer tube to working fluid” 103 Q[ i ] = 2∗ pi ∗ lambda tube ∗ L section/ln(d o / d i ) ∗(T hf mean w [ i ]−T wf mean w [ i ]) ”Heat transfer 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$;T=T wf [ i ] ; P=P wf[i]) ”calculate enthalpy of working fluid at the corresponding temperature and pressure” 109 s wf[ i ] = Entropy(WF$;h=h wf [ i ] ; P=P wf[i]) ”calculate entropy of working fluid at the corresponding temperature and pressure” 110 111 112 ”GNIELINSKI” ”The Gnielinski correlation is used to calculate the HTC of the heating fluid” 113 114 ”Calculate HTC of heating fluid area per division” 115 HTC mean hf [ i ] = Nu mean hf [ i ] ∗ lambda mean b hf [ i ] / D t ”Heat transfer coefficient heating fluid” 116 117 ”Pressure recalculated via pressure drop” 118 P hf [ i +1] = P hf [ i ] − Dp mean hf[i] ”Pressure in the beginning of next section” 119 120 ”Calculate properties of heating fluid per division” 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 viscosity of the heating 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 of the heating fluid” 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 conductivity of the heating fluid” 124 Cp mean b hf[i] = (0,00392 ∗ T hf mean b [ i ] −1 ,5∗10ˆ( −6)∗ T hf mean b [ i ] ˆ 2 + 1 ,8363) ∗10ˆ3 ”Specific heat of the heating fluid” 125 126 ”Calculate Nu, Re and Pr−number of heating fluid per division” 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 hf[i]/8)ˆ(1/2) ∗ ( ( Pr mean b hf[i])ˆ(2/3) −1)) ”Gnielinksi correlation” 128 f hf[i] = (1,82∗ log10 ( Re mean b hf [ i ] ) −1 ,64) ˆ(−2) ” Appendix E. Software 143

friction factor for working fluid” 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 heating fluid” 130 Pr mean b hf [ i ] = Cp mean b hf [ i ] ∗ Dyn visc mean b hf[ i]/lambda mean b hf [ i ] ”Pr−number heating fluid” 131 132 ”Calculate velocity of heating fluid inside the tube per division” 133 v mean b hf [ i ] = m dot hf /( Rho mean b hf [ i ] ∗ S an) ”velocity working fluid in the annulus” 134 135 136 ”PETUKHOV−KRANOSCHEKOV” ”The Petukhov−Kranoschekov correlation is used to calculate the supercritical HTC of the working fluid” 137 138 ”Calculate HTC of working fluid area per division” 139 HTC mean wf [ i ] = Nu mean wf [ i ] ∗ lambda mean b wf [ i ] / d i ”Heat transfer coefficient working fluid” 140 141 ”Pressure recalculated via pressure drop” 142 P wf [ i +1] = P wf [ i ] + Dp mean wf[i] ”Pressure in the beginning of next section” 143 144 ”Calculate properties of working fluid per division” 145 Dyn visc mean b wf[i] = viscosity(WF$; T=T wf mean b [ i ] ; P=P wf [ i ] ) ” Dynamic viscosity of the working fluid” 146 Rho mean b wf[i] = Density(WF$; T=T wf mean b [ i ] ; P=P wf[i]) ”Density of the working fluid” 147 lambda mean b wf[ i ] = Conductivity(WF$; T=T wf mean b [ i ] ; P=P wf [ i ] ) ” Thermal conductivity of the working fluid” 148 Cp mean b wf [ i ] = Cp(WF$; T=T wf mean b [ i ] ; P=P wf[i]) ”Specific heat of working fluid ” 149 150 ”Calculate Nu, Re and Pr−number of working fluid per division” 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 supercritical working fluids” 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 wf[i]/8)ˆ(1/2) ∗ ( ( Pr mean wf[i])ˆ(2/3) −1) ) ”Nu−number for working fluid” 153 f wf[i] = (1,82∗ log10 ( Re mean b wf [ i ] ) −1 ,64) ˆ(−2) ”Friction factor for working fluid” 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 fluid” Appendix E. Software 144

155 Pr mean wf [ i ] = Cp mean wf [ i ] ∗ Dyn visc mean b wf[ i]/lambda mean b wf [ i ] ”Pr−number working fluid” 156 157 ”Calculate EXTRA properties of working fluid per division” 158 h mean w[ i ] = Enthalpy(WF$;T=T wf mean w [ i ] ; P=P wf[i]) ”Mean wall enthalpy of the working fluid to determine the mean Cp per division” 159 h mean b[ i ] = Enthalpy(WF$;T=T wf mean b [ i ] ; P=P wf[i]) ”Mean bulk enthalpy of the working fluid to determine the mean Cp per division” 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 fluid” 161 lambda mean w wf[ i ] = Conductivity(WF$; T=T wf mean w [ i ] ; P=P wf [ i ] ) ” Thermal conductivity of the working fluid at the wall” 162 Dyn visc mean w wf[i] = viscosity(WF$; T=T wf mean w [ i ] ; P=P wf [ i ] ) ” Dynamic viscosity of the working fluid at the wall” 163 164 ”Calculate velocity of working fluid inside the tube per division” 165 v mean b wf [ i ] = m dot wf /( Rho mean b wf [ i ] ∗ S in) ”velocity working fluid 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 ]ˆ2)/2) ”pressure drop heating fluid per division” 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 ]ˆ2)/2) ”pressure drop working fluid per division” 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/.

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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