development of a novel high temperature heat pump system

Development of a Novel High Temperature Heat Pump System

Entwicklung eines neuartigen

Hochtemperatur-Wärmepumpensystems

Der Technischen Fakultät der Friedrich-Alexander-Universität

Erlangen-Nürnberg zur

Erlangung des Doktorgrades Dr.-Ing. vorgelegt von

Florian Reißner aus Nürnberg

Als Dissertation genehmigt von der Technischen Fakultät der Friedrich-Alexander-Universität Erlangen-Nürnberg Tag der mündlichen Prüfung: 30.06.2015 Vorsitzende des Promotionsorgans: Prof. Dr.-Ing. habil. Marion Merklein

Gutachter: Prof. Dr.-Ing. Jürgen Karl Ao.Univ.-Prof. Dipl.-Ing. Dr.techn. René Rieberer

Vorwort Die vorliegende Arbeit wurde im April 2014 schriftlich fertiggestellt und im Juni 2015 von der Technischen Fakultät der Friedrich-Alexander-Universität Erlangen-Nürnberg als Dissertation angenommen. Sie entstand innerhalb einer Industriepromotion in Kooperation des Lehrstuhls für Energieverfahrenstechnik und der zentralen Forschungsabteilung der Siemens AG (Corporate Technology). Der Leiter des Lehrstuhls für Energieverfahrenstechnik, Herr Prof. Jürgen Karl, ist zugleich mein Doktorvater. Ihm gilt mein besonderer Dank für die Betreuung, welche stets von einem ausgewogenen Maß an Freiräumen und konstruktiven Ratschlägen geprägt war. Auch als externer Doktorand, habe ich mich am Lehrstuhl jederzeit wohl gefühlt und war in soziale Aktivitäten eingebunden. Weiter gilt mein besonderer Dank Herrn Prof. René Rieberer für die schnelle Erstellung des Zweitgutachtens und die fruchtbaren Diskussionen auf verschiedenen Konferenzen. Auf Seiten der Siemens AG möchte ich besonderen Dank aussprechen an Herrn Dr. Bernd Gromoll. Er hat mich in den ersten beiden Jahren meiner Promotionszeit betreut und war mit seinem jahrzehntelangen Erfahrungsschatz auf dem Fachgebiet immer ein hervorragender Ratgeber. Nach seiner Pensionierung hat Herr Dr. Vladimir Danov die Betreuung übernommen. Ihm gilt ebenfalls mein besonderer Dank, vor allem für das Korrekturlesen. Weiter bedanke ich mich besonders bei Herrn Dr. Jochen Schäfer, welcher als Forschungsgruppenleiter fortwährend die Arbeit begleitet hat und das Thema über die Dissertation hinaus vorangetrieben hat. Meiner Frau Soyeon danke ich außerordentlich für die große Geduld und die harmonische und beständige Atmosphäre zu Hause, woraus ich Kraft für meine tägliche Arbeit gewinnen konnte. Ihre spontane und unbeschwerte Art half mir vielmals die Balance zwischen Freizeit und Arbeit zu wahren. Größten Dank schulde ich meinen Eltern Michaela und Hans. Sie haben mir von Beginn an meines Lebens alle erdenkliche Unterstützung zukommen lassen. Sie haben es stets geschafft, dass ich mich immerzu ohne Sorgen auf meinen beruflichen Werdegang konzentrieren konnte. Dadurch erlangte ich große Zuversicht und Ausdauer, was mich schließlich bis zum Doktortitel geführt hat. Nürnberg im September 2015 Florian Reißner

Table of contents 1

Table of contents

Table of contents 1

Abstract 3

Deutsche Zusammenfassung 4

1. Introduction and aim of the thesis 5

2. Fundamentals and trends of heat pump technology 8

2.1 Historic development ................................................................................. 9

2.2 Classification of different heat pump types .............................................. 11

2.3 Vapor compression heat pumps ............................................................... 12

2.3.1 Basics of the technology .................................................................. 13

2.3.2 Cycle control.................................................................................... 18

2.3.3 Impact on environment..................................................................... 20

2.4 Working fluids ......................................................................................... 21

2.4.1 Summary of important properties ..................................................... 25

2.4.2 Future trends of working fluids ........................................................ 26

2.5 High temperature heat pumps .................................................................. 29

3. Evaluation of working fluids for high temperature heat pumps 36

3.1 Evaluation criteria.................................................................................... 36

3.2 Considered working fluids and cycle independent properties ................... 37

3.3 System simulation.................................................................................... 42

3.3.1 Heat pump cycle types and corresponding working fluids ................ 42

3.3.2 Calculation software, procedure and parameters ............................... 47

3.3.3 Results and discussion...................................................................... 51

3.4 Selection of the working fluids for experimental investigation ................. 59

4. Operation strategies for overhanging working fluids 62

4.1 Nature of the overhanging behavior ......................................................... 62

4.1.1 Quantification by the slope of the saturated vapor line...................... 62

4.1.2 Linear correlations and their coverage on molecule classes .............. 63

4.2 Consequences for the operation in heat pumps ......................................... 67

4.2.1 Definition of the minimum required superheat ................................. 68

4.2.2 Internal heat exchanger as heat source for the additional superheat ... 70

2

4.2.3 Limiting temperature lift as stationary boundary condition ............... 71

4.3 Summary of linear correlations and their significance for cycle design..... 74

4.4 Control strategies and optimization possibilities ....................................... 75

5. Dynamic simulation of the start-up 82

5.1 Importance of the dynamic simulation ..................................................... 82

5.2 Determining variables and limiting cases ................................................. 84

5.3 Simulation methodology .......................................................................... 86

5.4 Results and discussion ............................................................................. 89

6. Experimental evaluation of selected working fluids 92

6.1 Setup of the functional model .................................................................. 93

6.1.1 Principal design for overhanging working fluids .............................. 93

6.1.2 Component selection ........................................................................ 95

6.1.3 Data acquisition system .................................................................... 97

6.2 Experimental procedure ........................................................................... 98

6.2.1 Operation characteristics of the functional model ............................. 98

6.2.2 Characterization of determining factors .......................................... 101

6.2.3 Investigation of the theoretically predicted limiting temperature lift 103

6.2.4 Determination of the experimental coefficient of performance ....... 104

6.2.5 Optimization of the coefficient of performance by reduction of

superheat ........................................................................................ 105

6.3 Results and discussion ........................................................................... 108

6.3.1 Relevance of the determining factors .............................................. 108

6.3.2 Verification of the limiting temperature lift .................................... 110

6.3.3 Potential of the optimization possibilities ....................................... 111

6.3.4 Experimental coefficient of performance ........................................ 114

7. Summary and outlook 121

Nomenclature 124

List of tables 127

List of figures 129

Appendix 135

References 138

Abstract 3

Abstract

The range of possible application of heat pumps in industrial processes is limited due to maximum supply temperatures of about 100 °C. Increasing the supply temperatures will increase the range of application and the efficiency of industrial processes. The main reason for the limited temperatures is the absence of adequate working fluids and corresponding heat pump systems. Legislative authorities restrict the usage of established working fluids. Only working fluids without ozone depletion potential (ODP) and with a low global warming potential (GWP) are future-proof.

The aim of this thesis is to develop a novel high temperature heat pump system using future-proof (no ODP, low GWP) and safe (nonflammable, non-toxic) working fluids.

An evaluation investigates a large variety of working fluids. Only two working fluids fulfill the evaluation criteria and are selected for the development of the novel system.

The selected working fluids have a complex molecular structure and show the phenomenon of a distinct overhanging behavior. This causes condensation during compression, which could damage the compressor of the heat pump. As a consequence, standard heat pump cycles are not suitable anymore.

Operation strategies for alternative cycle designs are developed. New terms like the “limiting temperature lift” define, characterize and classify the operation of overhanging working fluids.

A new functional model is developed and built up based on the operation strategies. The concept of the functional model is proven by stable operation of the selected working fluids at condensation temperatures up to 140 °C. The experimental results indicate a promising performance of this novel heat pump system with future-proof working fluids.

4

Deutsche Zusammenfassung

Die Anwendbarkeit von Wärmepumpen in industriellen Prozessen ist durch maximale Versorgungstemperaturen von etwa 100 °C begrenzt. Durch Erhöhung der maximalen Versorgungstemperaturen von Wärmepumpen kann die Bandbreite der Einsatzmöglichkeiten und somit die Effizienz von industriellen Prozessen erhöht werden. Der Hauptgrund für die begrenzten Versorgungstemperaturen ist der Mangel an geeigneten Arbeitsfluiden und zugehörigen Wärmepumpensystemen. Der Gesetzgeber schränkt den Einsatz von etablierten Arbeitsfluiden ein. Zukunftssicher einsetzbar sind nur Arbeitsfluide ohne Ozonabbaupotential (ODP) und mit geringem Erderwärmungspotential (GWP).

Das Ziel dieser Arbeit ist die Entwicklung einer neuartigen Hochtemperatur-Wärmepumpe unter dem Einsatz zukünftig einsetzbarer (kein ODP, geringer GWP) und sicherer (nicht brennbarer, nicht toxischer) Arbeitsfluide.

Eine Vielzahl von Arbeitsfluiden wird in dieser Arbeit evaluiert. Nur zwei Arbeitsfluide erfüllen die Evaluationskriterien und werden für die Entwicklung des neuartigen Systems ausgewählt.

Die ausgewählten Arbeitsfluide weisen eine komplexe Molekularstruktur auf und zeigen das Phänomen einer stark überhängenden Taulinie. Dies bedingt eine mögliche Kondensation während der Kompression und kann zu Schäden am Verdichter der Wärmepumpe führen. Als Konsequenz sind Standard Wärmepumpensysteme nicht geeignet.

Es werden Betriebsstrategien für alternative Kreislaufgestaltungen entwickelt. Neue Begriffe wie der „Grenztemperaturhub“ werden definiert, um den Betrieb von Arbeitsfluiden mit überhängender Taulinie zu charakterisieren und klassifizieren.

Ein neues Funktionsmuster wird basierend auf den Betriebsstrategien entwickelt und aufgebaut. Die Machbarkeit des Funktionsmusterkonzepts wird durch einen stabilen Betrieb der ausgewählten Arbeitsfluide bei Kondensationstemperaturen bis 140 °C nachgewiesen. Die experimentellen Ergebnisse zeigen vielversprechende Leistungsdaten dieses neuartigen Wärmepumpensystems mit zukünftig einsetzbaren und sicheren Fluiden.


1. Introduction and aim of the thesis

Introduction

The International Energy Agency states [1]:

“Heat pump technologies are widely used for upgrading low-temperature free heat from renewable sources, such as air, water, ground and waste heat, to useful temperatures. They are used for residential and commercial space and water heating, cooling, refrigeration and in industrial processes. […] 1.8 billion tones of CO2 per year could be saved by heat pumps, corresponding to nearly 8% of total global CO2 emissions.”

The stated possible CO2 emission reduction is even higher than the reduction

goals of the Kyoto protocol [2]. About half of the worldwide primary energy consumption serves for heating

purposes and fossil fuels are the main source for this thermal energy [3, 4]. These fossil fuels have to be replaced in the future for a CO2 emission free heat supply. The challenge of replacing fossil fuels intensifies by the expansion of renewable energy. Mostly electricity generating technologies are installed as renewable energy source like photovoltaic and wind power. However, they do not comprise combined heat and power generation. Nevertheless, the heat production of conventional power generation is still required and has to be replaced somehow. There are several sustainable heat sources like the combustion of biomass or geothermal sources. However, the capacity of such sources is limited.

The heat pump technology shows no capacity limitation. Heat sources for the heat pump like ambient air, ground, sea or lake water are infinitely available and renewable. Only the driving force of the heat pump indirectly causes CO2 emissions. Most heat pumps use electric energy as driving force. In the European Union, the average efficiency in electricity generation is 44% [5]. This means, that a heat pump with > 2.3 times generated thermal energy in relation to consumed electric energy reduces CO2 emissions (at the same fuel mix). The performance of state-of-the-art heat pump technology is far above this limit and heat pumps are therefore widely disseminated. In the European Union over 750’000 heat pumps are installed every year [6].

This large number refers mainly to heat pump units for residential heating. The dissemination of industrial heat pumps is much lower [7]. Other conditions are valid

6

for industrial processes. Usually higher temperatures are needed, which lowers the efficiency of a heat pump. Consequently, the economical benefit decreases or even diminishes, depending on fossil fuel and electricity prices. In Scandinavian countries, where a lot of cheap electricity from hydropower exists [8], the dissemination of industrial heat pumps is much higher than in other countries.

The temperature limit of commercially available heat pump technology is around 100 °C [7]. This temperature limit should be elevated, so that more industrial processes can be supplied with sustainable thermal energy from heat pumps. For example, the generation of steam would be possible with 120 °C, which is a universal thermal energy carrier in industrial processes. The main challenge is the choice of the working fluid and the development of corresponding new high temperature heat pump systems.

Strict limitations by law forbid the use of established working fluids in the future. The aim of the legislative authorities is to minimize the usage of working fluids, which destroy the ozone layer and contribute to global warming [9]. Especially, for the high temperature range, there are currently no future-proof working fluids commercially available, which are also safe in operation (non-flammable, non-toxic).

The chemical manufacturers try to synthesize molecules, which fit to the demands of environment protection. For example, functional groups inserted into the molecules lower the contribution to global warming. Thus, the complexity of the working fluids’ molecular structure increases. However, the thermodynamic behavior alternates as well. The alternated properties have to be accounted for operation of these working fluids in heat pumps. Aim and structure of the thesis

The aim of this thesis is to develop a novel high temperature heat pump system operating with future-proof and safe working fluids. The targeted temperature range of the heat supply is > 120 °C. In doing so, working fluids are for the first time investigated in this thesis. The goal is to establish these working fluids in a new high temperature heat pump system and fully describe their thermodynamic behavior related to the heat pump operation.

Chapter 2 describes the fundamental knowledge of the heat pump technology starting with a historical review and a classification of different heat pump types. The heat pump type vapor compression is explained in detail followed by a subsection about many aspects of working fluids. In the end, high temperature and industrial heat pumps are outlined.


Chapter 3 evaluates working fluids for the high temperature range theoretically. The environmental and safety properties and the thermodynamic properties function as evaluation criteria. A system simulation of different theoretical heat pump cycles determines the thermodynamic properties. In the end of the chapter, working fluids are selected for the experimental evaluation in chapter 6.

Chapter 4 develops and establishes novel strategies for the operation of the selected working fluids from chapter 3. The working fluids show alternated thermodynamic properties. Boundary conditions for the operation are ascertained and new terms are defined to characterize them for steady state operation.

Chapter 5 analyzes the non-steady behavior of the new high temperature heat pump system by dynamic simulations. Especially, the start-up procedure of the heat pump is important due to the alternated thermodynamic properties.

Chapter 6 evaluates the performance of the selected working fluids from chapter 3 experimentally. At first, the development and the setup of the novel high temperature heat pump system is presented. The aim of this chapter is to proof the concept of the new system experimentally and to determine the performance of the selected working fluids. Determining factors are characterized and developed operation strategies of chapter 4 are applied for the experiments.

8

2. Fundamentals and trends of heat pump technology

The basic principle of heat pumps is to “pump” heat from a source with a low temperature level to a sink with a high temperature level (see Figure 2.1). According to the second Law of Thermodynamics formulated by Rudolf Clausius in the year 1865 [10]:

“Heat can never pass from a colder to a warmer body without some other change, connected therewith, occurring at the same time.”

Consequently, a heat pump needs an additional driving force to transfer heat from a low to a high temperature reservoir. This leads to an additional energy flow.

Figure 2.1 Basic principle of a heat pump

As stated by Cube et al. [11] the described principle is terminological divided

in heat pump and chiller, depending on which side of the heat transfer is utilized. In case of a chiller, the removal of heat from the low temperature reservoir is the intended usage. In case of a heat pump, it is the heat supply to the high temperature reservoir. In fact, there is no difference in the working principle whether it is a chiller or a heat pump (see Figure 2.2).


Figure 2.2 Heat pump and chiller working principle; Dashed boxes indicate the utilized side

2.1 Historic development

Natural resources like biomass, natural gas and coal store chemical energy and deliver heat through combustion processes. That is why historically there was no need to develop alternative heat supply systems. Nevertheless, there is no adequate natural resource (e.g. transport of arctic ice) for a cold supply below the ambient temperature. Hence, the historic development of technologies for this principle (see Figure 2.1) originated from the need for cold supply (synonym: refrigeration) [11].

The Scottish physician William Cullen is said to be the first person who applied artificial refrigeration. In 1756, he evacuated a vessel of diethyl ether with a pump. The diethyl ether evaporated due to the low pressure and absorbed heat from the surroundings. In doing so, Cullen achieved to produce a small amount of ice from liquid water [12, 13].

The first commercial machines for refrigeration were built around 1850 by the ideas of William Thomson (later known as Lord Kelvin) [14]. The decisive aspect was to use a continuous process (cycle), which allowed a commercialization to happen. Started with air as working fluid in the 1850s, the development of refrigeration technologies quickly grew with the use of other working fluids [12]. Carl von Linde was the first person to use ammonia in 1876 [15], and from the 1920s on, ammonia was established for the use in closed chiller cycles. From this time on,

10 2.1 Historic development

the demand for cold production was met. It revolutionized the quality of human life [12]. For example, the production, processing and trade of food changed considerably. Finally, today it is common to use machines like household refrigerators and car air-conditioning in everyday life.

The first heat pumps, which were intentionally used to supply heat, were applied from the 1920s on. Krauss in 1921 [16] was probably the first person, who considered heat pumps based on the knowledge of William Thomson’s work [14]. With the fast development of chiller technology, heat pumps were also commercialized from that time on [12, 15]. A famous and early example is a heat pump installed in the Town Hall of Zurich in 1938. It used river water as heat source (7°C) and delivered a heat load of 100 kW of 60 °C hot water for space heating [12, 17].

Heat pumps often could not compete with oil or natural gas furnaces for heat generation. Especially in the time of low oil prices from 1950 to 1972, the development of heat pump technology stagnated. In 1973 and 1979 the first and second oil crisis boosted the dissemination of heat pumps (see Figure 2.3) and around 4 million heat pump units were installed in that period worldwide. But after the oil price collapse in 1981 the heat pump market broke down again. It took 15 years until the heat pump market grew again (see Figure 2.3).

Figure 2.3 Sold number of heat pumps units in Germany from 1978 to 2012 [18]

0

10'000

20'000

30'000

40'000

50'000

60'000

70'000

80'000

Sold

num

ber o

f hea

t pum

p un

its

Year


The societal understanding of environmental problems led to the conviction that saving primary energy is a necessity. This resulted in committing climate protection goals in the Kyoto protocol from 1997. From the mid 1990s, the heat pump market grew constantly [17]. In the European Union around 755´000 heat pump units were installed in 2012 and over 5.4 million units accumulated since 2005 [6].

Today, the heat pump technology is said to be one of the most promising possibilities to reach climate protection goals. It will play a major role for the transition from fossil fuel to renewable energy in the heating sector [1].

2.2 Classification of different heat pump types

There are several different heat pump types based on the principle of Figure 2.1. According to Cube et al. [11] their driving force distinguishes them. Heat pumps can be mechanically or thermally driven (see Figure 2.4)

Figure 2.4 Heat pump types with continuous process distinguished by driving force

All mentioned thermally driven heat pump types make use of a pair of

working fluids. In spite of intensive efforts, there are only two pairs of working fluids established. These are ammonia/water and water/lithium bromide [11]. Ziegler et al. [19] and Schaber and Römich [20] mentioned crucial drawbacks like toxicity, flammability and corrosiveness. Furthermore they stated that there are currently no other efficient and safe to operate pairs of working fluids available. There is some research for new pairs of working fluids, in which ionic liquids are involved. However, that research still focuses mainly on ascertaining thermophysical properties of these fluids [21, 22, 23] and measured efficiencies are low [24]. Therefore, a

12 2.3 Vapor compression heat pumps

thermally driven heat pump is not selected for the realization of the experimental tasks and is not part of this thesis.

The commercial significance of mechanically driven heat pumps is much larger than that of thermally driven heat pumps. More than 90% of all worldwide installed heat pumps are mechanically driven [25].

The first continuously operated heat pumps used air as working fluid (see section 2.1). These are Brayton heat pumps. The essential characteristic is that the working fluid always keeps its gaseous state during operation. Compared to vapor compression systems, the efficiency of such systems is low. The heat transfer shows no phase change and thus is inefficient. Moreover, the mass flow must be much higher to achieve the same heat load, which consumes more energy in the compressor. Recent research of Brayton heat pumps considers noble gases as working fluid, at which the practical feasibility to reach adequate efficiency has still to be shown [26].

Stirling heat pumps also utilize a working fluid in its gaseous state. The Stirling heat pump is, like the Brayton heat pump less efficient than the vapor compression heat pump [27]. They are used only for niche applications, e.g. in the field of cryogenics. There are no suitable working fluids with very low boiling points, which can be used in vapor compression heat pumps. With helium as working fluid in a Stirling heat pump (Cryocooler), it is possible to reach temperatures below -200 °C [28].

It would be technically possible to build a Brayton or Stirling heat pump for high temperatures, but due to the expected low efficiency, they are not selected for this thesis.

The by far most used and advanced heat pump type is the vapor compression heat pump. It utilizes a working fluid in its gaseous and liquid state during operation. The heat from the source and to the sink transfers with a phase change of the working fluid (evaporation and condensation) [29]. Working fluids are designed specifically for the use in vapor compression heat pumps. Thus, the possibility to reach a better efficiency is higher. Therefore, the vapor compression type is selected to realize this thesis and is explained in detail in section 2.3.

2.3 Vapor compression heat pumps

As stated in section 2.2 the most used and advanced heat pump type is the vapor compression heat pump. This technology is described extensively in literature.


This chapter shows a description of the main important aspects of this technology based on Cube et al. [11]. 2.3.1 Basics of the technology

The vapor compression heat pump bases on the reversed Rankine cycle. It is a

counter clockwise thermodynamic cycle (see Figure 2.5). The working fluid absorbs heat from the heat source at low temperature by evaporation (4 1). The compressor is mechanically driven and compresses the working fluid (1 2). At high pressure, the working fluid condenses and releases heat at high temperature to the sink (2 3). After condensation, the working fluid expands to evaporation pressure and closes the cycle (3 4).

Figure 2.5 Schematic diagram of a vapor compression heat pump

An electric drive usually provides the mechanical energy flow of the

compressor. The efficiency of the vapor compression heat pump is defined as the

Compressor

2

Expansionvalve

3

4 1

Qcond

Qevap

Evaporator

Condenser

Heat sink(high temperature

reservoir)

Heat source(low temperature

reservoir)

Electric energyreservoir

PelEmechanical MElectric

drive


ratio of the heat load (Qcond) and the electric power (Pel) and is called coefficient of performance (COP).

COP = Qcond

Pel (2.1)

The Carnot cycle describes the theoretical maximum of the COP. It is an ideal

cycle for comparison without any losses. It consists of four reversible changes of state:

Isothermal evaporation (4 1) Isentropic compression (1 2) Isothermal condensation (2 3) Isentropic expansion (3 4).

The Carnot COP depends only on the temperature levels of evaporation and condensation and is defined as:

COPCarnot = Tcond

Tcond - Tevap (2.2)

The denominator of equation (2.2) is the difference of the condensation temperature and the evaporation temperature. This term is the temperature lift. It describes how much the heat pump lifts the temperature of the heat source. TLift = Tcond - Tevap (2.3)

The Carnot cycle is ideal and, of course, not realizable in vapor compression heat pumps. A rule of thumb is that around 50% of the Carnot COP is achieved in commercial heat pumps [30]. The reasons are several losses (deviations from the Carnot cycle) within the changes of state:

Evaporation (4 1) Liquid droplets inside the compressor can cause mechanical damage by droplet erosion. Consequently, the working fluid has to evaporate at least to saturated vapor. To be sure, that no liquid droplets enter the compressor, the working fluid superheats after evaporation for about 2 to 10 K.

Compression (1 2) The compression occurs in the vapor phase. An assumed isentropic compression along an isentrope leads to a compression end temperature higher


than the condensation temperature of Carnot. Furthermore, there are volumetric losses, heat transfer and friction leading to a non-isentropic compression. This is regarded by the isentropic compressor efficiency ( s). The compressor outlet temperature of a real compression can be lower than the isentropic compression (at e.g. s = 0.8).

Condensation (2 3) The working fluid cools (desuperheats) down to the condensation temperature before the actual condensation starts. In addition, the working fluid subcools after the condensation for about 5 K to avoid undesirable evaporation by pressure losses in the pipe following the condenser.

Expansion (3 4) An actual isentropic expansion would imply an expansion machine, which recovers the lost work. However, the recoverable work in vapor compression heat pumps is very low, compared to the compression work. Hence, the effort to include components like turbines is usually not legitimate. Consequently, simple expansion valves function as throttle. The expansion then follows an isenthalpic change of state.

All these deviations from the Carnot cycle together lead to a theoretical cycle depending on a particular working fluid and cycle parameters (Tcond, Tevap, superheat, subcool, s). This theoretical cycle is depicted in a logarithmic pressure enthalpy diagram (log p,h) to easily retrace all mentioned changes of state (see Figure 2.6). It is the standard phase diagram to visualize vapor compression heat pump cycles.


Figure 2.6 Log p,h diagram of a theoretical vapor compression heat pump cycle; Working fluid

R134a; Tcond = 60 °C; Tevap = 20 °C; s = 0.8; Superheat = 5 K; Subcool = 5 K; Isenthalpic expansion

The saturated liquid line and the saturated vapor line meet at the critical point

and enclose the two-phase region (phase boundary). In that region, the working fluid is in its liquid and vapor phase. The liquid phase situates at lower enthalpies (left of the saturated liquid line) and the vapor phase at higher enthalpies (right of the saturated vapor line). The evaporation and condensation take place inside the phase boundary. The superheat usually comes from the heat source inside the evaporator. Likewise the desuperheat and the subcool are transferred to the heat sink. Together with this corresponding single phase heat transfers the specific heat loads, respectively specific power, can be directly read out of the diagram as enthalpy differences (distances parallel to the enthalpy axis). Qevap = h1 - h4 (2.4)

Qcond = h2 - h3 (2.5) Pcomp = h2 - h1 (2.6) The theoretical COP is accordingly:

COPtheoretical = h2 - h3

h2 - h1 (2.7)

1

2

4

3

2030

4050

6070

8090

100

10

1.75

1.8

1.85

1.71

1

10

100

150 200 250 300 350 400 450 500

p [b

ar]

h [kJ kg-1]

Isotherm [°C]Isentrop [kJ kg-1 K-1]Heat pump cycle

Liquidphase

VaporphaseLiquid+Vapor

phase

Critical point

Condensation

EvaporationExpa

nsio

n

Superheat

Subcool Desuperheat


It is the most important value to describe the thermodynamic suitability of a working fluid for vapor compression heat pumps. It is denominated as theoretical, because it compares the fluid-specific COP, depending only on theoretical cycle parameters and not on equipment based aspects like losses by heat transfer or pressure drop. Thus, the theoretical COP intends not to forecast real COP values. It compares working fluids to each other independent of used equipment.

In order to calculate the enthalpies of equation (2.7), each point of state has to be defined. All points of state in single phase determine by a function of two state variables [31]. The enthalpies are as follows:

h1 = f pevap; Tevap + Superheat (2.8)

h2 = h2s-h1

s + h1 (2.9)

h2s = f pcond; s1 (2.10)

h3 = f pcond; Tcond - Subcool (2.11)

h4 = h3 (2.12) An additional value to describe the thermodynamic suitability of a working fluid is the volumetric heating capacity (VHC): VHC = (h2 - h3) 1 (2.13) The VHC describes the generated heat load per processed volume. It gives an idea of the size of the compressor. The higher the VHC, the smaller compressor is needed. Larger compressors need more power and have higher heat losses for a given heat load. Consequently, the VHC influences the achievable experimental COP. Common values of the VHC of ordinary heat pump applications are between 3000 and 6000 kJ m-3. The lower practical limit of the VHC depends strongly on the compressor type and is said to be between 500 [32] and 1000 kJ m-3 [33].

Together with the theoretical COP, an adequate estimation of the efficiency of a particular working fluid is possible [34]. A real vapor compression heat pump cycle has additional losses like pressure losses, heat losses and at the compressor volumetric, mechanical and electrical losses. Especially the pressure losses are significant for the power uptake of the compressor.


In the theoretical cycle, the compressor has to raise the pressure according to the pressure ratio given by the desired evaporation and condensation temperature:

pratio = pcondpevap

(2.14)

The suction pressure is lower than the evaporation pressure due to pressure losses in the evaporator and the suction line. The same applies analog for the pressure side, at which the compressor has to overcome the pressure losses in the pressure line and the condenser. 2.3.2 Cycle control

The vapor compression heat pump cycle has four main components, as shown in Figure 2.5 (Evaporator, Compressor, Condenser, Expansion valve). The state of the working fluid in each component and the connecting pipes has to be controlled somehow. This section describes the basic state of technology for the cycle control.

Working fluid flow

The heat load in the condenser and evaporator and the compressor power

result mainly from the flow of the working fluid, generated by the compressor. There are several possibilities to control the flow and the following listing shows some examples:

Change of compressor speed (variable adjustment by e.g. frequency converter) [35]

Prevention of full closing of suction valve by mechanical fixture (sucked working fluid is moved back to the suction line) [36]

Reduction of the suction chamber volume by adjusting spacers [36] Hot gas bypass of pressure line to suction line [35]

Temperatures and pressures

The condensation temperature and pressure arise from the temperature of the heat sink.


The evaporation temperature should always be as near as possible to the temperature of the heat source to minimize the pressure ratio. The prerequisite is that the vapor in the suction line completely evaporates and slightly superheats (see section 2.3.1). It is state of technology to control the superheat at the exit of the evaporator by adjusting the flow through the expansion valve. The controlling system needs information about pressure and temperature of the working fluid. There are two common and widespread technologies to control the superheat (thermostatic and electronic). Thermostatic expansion valve The thermostatic expansion valve is a passive component, which adjusts only by pressures and indirectly by temperatures. It consists of two main parts, which are the valve itself and the sensor. A capillary pipe connects them (see Figure 2.7).

Figure 2.7 Schematic of the function of a thermostatic expansion valve; p1: Sensor pressure, p2:

Evaporator pressure, p3: Spring pressure

The position of the membrane defines the degree of opening of the valve.

There are three pressures acting on the membrane: Sensor pressure (p1), evaporator pressure (p2), spring pressure (p3).

The easiest configuration is, when the heat pump cycle working fluid is also inside the sensor. Consequently, the same pressure acts inside the sensor, inside the capillary pipe and in the valve at the upper side of the membrane. This pressure is equal to the saturation pressure of the sensors’ temperature. During stationary operation, the pressures are in equilibrium. If the heat source temperature rises, the superheat rises as well, thus producing a higher pressure at the upper side of the membrane. Hence, the valve opens more and more working fluid flows through the valve. The evaporator has to evaporate a higher amount of working fluid and the


superheat decreases accordingly. The same happens vice versa, when the heat source temperature is decreasing with a followed decrease of the superheat. The spring can be adjusted to add a static superheat to the dynamically controlled superheat of the sensor [36, 37].

Usually more sophisticated thermostatic expansion valves are in use nowadays. Desorbing gas (different from the working fluid) from a solid sorbent in the sensor often generates the occurring pressure. In doing so, additional functions as a maximum operating pressure can be included as well [36, 38]. Electronic expansion valve An electronic expansion valve needs the same information like the thermostatic to control the superheat at the exit of the evaporator. Temperature and pressure are measured and a controller calculates the superheat. Algorithms control the opening of the valve by positioning a step motor. For each working fluid, there are different algorithms in use. Extensive testing is required to develop these algorithms and controller manufacturers [39] keep them secret. 2.3.3 Impact on environment

The aim of using heat pump technology for heating purposes is to reduce CO2 emissions and costs. The extent of reduction significantly depends on the compressor and drive efficiency. Furthermore, if a high share of renewable electricity is present, the CO2 factor (kg CO2 emission per kWh thermal energy) can be very low. In Scandinavian countries, for example, hydropower generates a large part of the electricity [8] and thus the CO2 factor is low.

In many other countries, fossil fuel is the main primary energy carrier for the electricity generation. In the European Union, the average efficiency in electricity generation is 0.44 [5]. The reciprocal value corresponds to the minimum COP, which has to be realized to reduce CO2 emissions (COP > 2.3) compared to heat generation by fossil fuel.

Figure 2.8 shows the CO2 factor of different COPs (average over whole year) compared to a gas burner in Germany. The projected progress until 2050 bases on the electricity mix forecast data of the German Aerospace Center (DLR) [40].


Figure 2.8 CO2 factor of gas burner and heat pump at different COPs operated with electricity mix

in Germany with forecast data of the DLR until 2050 [40]

Zotter and Rieberer state similar CO2 factors for heat pump operation with the Austrian electricity mix. A heat pump with a COP of 3.7 emits 0.09 kg kWhth

-1 and a gas burner 0.24 kg kWhth

-1 [41]. The CO2 emissions by electricity consumption of a heat pump are indirect

emissions. There are also direct emissions of the working fluid, which contribute to global warming. The mass of direct emissions is much smaller than the indirect emissions. However, standard working fluids show a global warming potential (GWP) that is more than 1000 times higher than the potential of CO2 (see section 2.4). The working fluid can emit through leakages of heat pump components and general handling (e.g. maintenance, transport and waste disposal without fluid recovery).

2.4 Working fluids

The working fluid is the most important part of a vapor compression heat pump. Every chemical compound can function as working fluid, if it condenses and evaporates at technical feasible pressures. In other words, the vapor pressure line must fit to the desired temperatures of a particular application. However, just a few working fluids established for commercial usage.

As stated, ammonia functions as working fluid in large-scale since the 1920s (see section 2.1). Moreover, others like Methyl chloride, Sulfur dioxide or ethers were

0.25

0.20

0.15

0.10

0.05

02010 2020 2030 2040 2050

Gas burnerHeat pump COP = 2.5Heat pump COP = 3.0Heat pump COP = 3.5Heat pump COP = 4.0

CO

2fa

ctor

[kg

kWh t

h-1]

22 2.4 Working fluids

used. These are natural chemical compounds or were known from other applications. This first generation (1830-1930) of working fluids had problems with toxicity, flammability and operation stability [42]. In 1930, the first working fluid (R12), specifically designed for heat pumps, was introduced. From that time on synthetic working fluids developed for the use in vapor compression heat pumps [43].

The second generation (1930-1990) of working fluids is the Chlorofluorocarbons (CFCs) and the Hydrochlorofluorocarbons (HCFCs). They showed much better safety and durability [42]. However, in the 1970s it was discovered that a huge amount of CFCs and HCFCs accumulated in the atmosphere destroying the ozone layer. This led to the Montreal protocol from 1987 signed by 197 countries, which phases out the use of these chemical compounds [44]. Today, all chemical compounds with a significant ozone depletion potential (ODP) are forbidden. The ODP is a relative compound-specific value, defined on R11 (ODP = 1) [45].

The third generation of working fluids (1990-2010s) developed because of the Montreal protocol. Hydrofluorocarbons (HFCs) have no chlorine content and thus the ODP is zero. These working fluids are standard technology today. However, they show a global warming potential (GWP) and are listed in the Kyoto protocol from 1997 [2] as green house gases. The GWP is a relative compound-specific value, defined on carbon dioxide (GWP = 1).


Figure 2.9 Development of working fluid usage in vapor compression heat pumps by generations;

Adapted from [42]

The GWP of working fluids for vapor compression heat pumps is a very

important value because there are recent restrictions by law for their usage. Especially the European Union with its strict green house gas emission reduction goals is a pioneer in reducing the application of HFCs. For example the use of working fluids in new car air-conditioning systems with a GWP higher than 150 is forbidden since 2013 [46]. There are further general restrictions forbidding the use of working fluids with a GWP higher than 150 in chillers (new products for industrial use) from 2022 on. Moreover, the whole production capacity together with the amount of import of HFCs reduces by -40% until 2020 and -80% until 2030 [9]. These restrictions are valid since the ratification of the new F-Gas regulation in the EU Directive 517/2014.

During the work for this thesis, the transition from the third generation to the fourth generation of working fluids started. The challenge of the fourth generation is to protect the ozone layer (ODP = 0) as well as the climate (low GWP) while also meeting the demands of a high thermodynamic efficiency. Mostly, Hydrofluoroolefins (HFOs) serve to achieve this. Until today, they still did not significantly penetrate the application of working fluids in the heat pump market [47, 48].

Table 2.1 lists common working fluids of each generation. The thermodynamic suitability has no significant difference among the four mentioned working fluids. Thus, the thermodynamic properties are not the main reason for the decision to use or

1st Generation1830-1930s“Whatever worked”NH3, SO2, Ethers, …

2nd Generation1930-1990s“Safety and durability”e.g. CFCs, HCFCs

3rd Generation1990-2010s“Ozone protection”e.g. HFCs

4th Generation2010-?“Climate protection”e.g. HFOsMontreal protocol 1987

ODP compounds banned

Kyoto protocol 1997Reduction of GWP compounds


not to use a particular working fluid. The environment- and safety properties are very important as well. The lack of and need for new alternative working fluids can be explicitly seen at the example that R1234yf shall replace R134a. Flammability is condoned to use a low GWP working fluid. There are no other adequate low GWP replacements for R134a commercially available.

R744 was also developed for heat pumps and chillers. However, after decades of research and development, there is still no significant market penetration (e.g. car air-conditioning). The indirect emissions exceed the direct emissions. Consequently, the advantage of the low GWP of R744 is futile. Table 2.1 Thermodynamic, environmental and safety properties of historically momentous

working fluids each representing one generation; COPtheoretical, VHC calculated at Tevap = 20 °C; Tcond = 60 °C; Subcool = 5 K; Superheat = 5 K s = 0.8; Colors are red if toxic, flammable, ODP > 0, GWP > 2500; Orange if 2500 > GWP > 150; Rest is green [45, 49]

Generation 1 2 3 4 Example fluid Ammonia R12 R134a R1234yf Fluid type Natural CFC HFC HFO COPtheoretical 5.97 5.85 5.74 5.51 VHC [kJ m-3] 8071 4144 4424 4028 Toxic Yes No No No Flammable Yes No No Yes ODP 0 0.9 0 0 GWP < 1 8500 1300 4

For the sake of completeness it has to be said, that the mentioned working

fluids and given time periods shall just give an idea about the situation on working fluids. These are the most common working fluids and trends of usage. Of course, there are several other working fluids like hydrocarbons or mixtures like R410A or R407C. Some fluids are used for generations like ammonia, which is still used today in a variety of applications.


2.4.1 Summary of important properties

The following list summarizes the most important working fluid properties. They are the key attributes for the evaluation of the working fluids for the high temperature range in chapter 3.

Coefficient of performance The most important thermodynamic property is the theoretical COP. It indicates the maximum achievable COP of a working fluid (see section 2.3.1). Volumetric heating capacity The VHC is important for the compressor design and influences the achievable experimental COP (see section 2.3.1). Critical temperature For the high temperature range, the critical temperature should be clearly higher than the targeted condensation temperature [11, 50]. The closer the condensation temperature is to the critical temperature, the smaller is the condensation enthalpy and this negatively influences the COP.

It is also possible to operate transcritical or supercritical heat pump cycles (e.g. with R744) [51]. Here the temperature of the generated heat is higher than the critical temperature. Hence, the heat transfer to the heat sink is sensible and a large temperature glide at the heat sink is required to realize reasonable COPs. Section 3.3.1 considers the difference of these cycles. Pressure ratio As mentioned in section 2.3.1, the pressure ratio should be as low as possible to minimize the compressor power. Equation (2.14) and the vapor pressure line give the pressure ratio. Pressure level The pressure level is important for safety aspects and the material effort of the equipment. It is an advantage of a working fluid, when the pressure level is low.


Ozone depletion potential As shown in section 2.4, the ODP of working fluids must be zero or very low. Otherwise, authorities do not grant permission for usage. Global warming potential For future-proof usage of working fluids, the GWP should be as low as possible. This thesis focuses on research for applications in long-term future. Upcoming restrictions by law will phase out the usage of working fluids with a high GWP (see section 2.4). Consequently, a low GWP is strongly advisable. Flammability The flammability of a working fluid is a serious safety issue and should be avoided if possible. This thesis distinguishes working fluids in flammable and non-flammable. Subgroups like mildly flammable working fluids are not considered.

The European safety norm TS 95006 [52] regulates the usage of heat pump systems filled with flammable working fluids with a charge of smaller than 150 g. For larger charges, national safety norms like the VDMA 24020 [53] apply. The allowed charge limit depends on several aspects like type of location, type of person presence and room volume and goes up to 25 kg. Large industrial heat pumps need working fluid charges of several hundreds of kilograms. There are no norms available for such cases. If the working fluid is flammable, enormous and expensive safety equipment is mandatory. Toxicity Toxicity is a serious safety issue as well. Non-toxic working fluids are of course preferable to avoid hazard to persons. This thesis applies the European Union directives 67/548/EG and 1999/45/EG to distinguish toxic and non-toxic working fluids. If a working fluid has at least one health hazard classification (e.g. skin or eye irritation, acute toxicity, carcinogenicity, germ cell mutagenicity or aspiration hazard it is denominated as toxic. 2.4.2 Future trends of working fluids

The aim of working fluid development is to find “perfect” chemical compounds. Ideally, they should show every important property (section 2.4.1) in a positive way. As shown in section 2.4, the working fluid usage developed already


over more than 100 years. After natural working fluids, methane derivatives were developed. There are 15 statistical possible molecules with chlorine and fluorine as substitutes for hydrogen. Figure 2.10 shows four (of 15) well-known examples of methane derivatives. They all show at least one negative environment or safety property and this counts for all 15 statistical possibilities as well.

Figure 2.10 Examples of methane derivatives; Colors are red if ODP > 0, GWP > 2500, flammable;

Orange if 150 < GWP < 2500

Subsequently, ethane (e.g. R134a) and propane (e.g. R1234yf) derivatives

developed, which are state of the technology today. The number of possible molecules increases greatly with the number of carbon atoms. However, “perfect” working fluids were not discovered among them.

The complexity of the molecules increases during the ongoing search for new working fluids. Recent filings for patents of companies like DuPont and Arkema include working fluids based on butane derivatives [54, 55]. This emphasizes the trend towards complex molecules.

The molar mass naturally increases with molecule complexity. In addition, the thermodynamic behavior changes as well. Table 2.2 shows sketches of T,s diagrams. The shape of the phase boundary varies. The working fluids distinguish by the slope of the saturated vapor line. On the one hand, the saturated vapor line can show a negative slope and on the other hand a positive slope at least in a part of its progression.

Different terminology exists in literature for this behavior. Working fluids with positive slopes are called “wet” and “bell-shaped” [56]. Negative slopes are called “dry”, “re-entrant” [57], “skewed” [58] and “overhanging” [56]. The terms

C?

?

?

?

CF

Cl

H

F

CF

F

F

F

CF

F

H

H

CF

F

Cl

Cl

R32

R12

R14

R22

High ODPVery high GWP

FlammableHigh GWP

High ODPHigh GWPHigh GWP

Very high GWPFlammable


“dry” and “wet” are cycle type dependent and thus are not clearly defined. If other cycles like the Organic Rankine cycle are considered, the terminology is vice versa [59]. The terms “bell-shaped” and “overhanging” express the shapes of the phase boundary in a pictographic way. Therefore, these terms are used throughout this thesis. Table 2.2 Exemplary working fluids of each generation with trend towards complex molecular

structure and overhanging behavior

Generation 2 3 4 -

Example fluid

R22 R134a R1234yf e.g. LG6,

R1336mzz Year of commer-cialization

1936 [60] 1990 [61] 2010 [47] -

Molecular structure

-

Carbon atoms 1 2 3 4

Molar mass [g mol-1]

86 102 114 >150

Phase boundary shape

Bell-shaped Bell-shaped Overhanging Strongly

overhanging

Sketches of T,s diagrams with trend of slope

The overhanging behavior may cause problems during compression. If the compression takes place near the saturated vapor line, condensation could occur (see Figure 4.4). Liquid droplets form and the compressor could take damage by droplet erosion. Morrison G. [58] proved the reason for the overhanging behavior. With the alkane series, he showed that the degree of the overhanging behavior depends on the molecule complexity. He derived equation (2.15).

CFF

Cl

HCFF

FC

FH

H

CH

HC

F

C

F

F

F

0dsdT 0

dsdT 0

dsdT 0

dsdT


dSdT sat.vap.

=1T

cv+pT V

dVdT sat.vap.

(2.15)

On the left side is the inverse slope of the saturated vapor line in a T,s diagram. The two terms on the right side, pressure and volume term, are almost the same for different alkanes and the product is negative [58]. The main influence on the slope comes from the isochoric heat capacity. The more chemical bonds a molecule has, the larger is the capacity to store energy due to the increased amount of modes of motion. The slope is negative when the pressure and volume term on the right side of the plus sign shows a higher absolute value than the isochoric heat capacity term. This counts for methane, ethane and propane. Butane is the first alkane, which shows a positive slope at a certain temperature range. Here the isochoric heat capacity term is higher than the pressure and volume term. The operation of overhanging working fluids in heat pumps is a challenge and is subject of chapter 4, in which operation strategies for these working fluids are developed.

2.5 High temperature heat pumps

The term high temperature heat pump (HTHP) refers to heat pumps, which suppose to generate heat at high temperatures. The terminology in literature about the temperature level is not consistent. The border to distinguish heat pumps from HTHPs ranges from 60 °C [62] to 100 °C [15]. There is also the term very high temperature heat pump (140 °C) [63]. In this thesis, 70 °C function as border to classify a HTHP. Pennartz stated this border as chairperson of the “High Temperature Industrial Heat Pumps” session during the “2011 International Congress of Refrigeration” [64].

Heat pumps in the building sector for space heating and hot water do usually not need temperatures above 60 °C [65]. Consequently, HTHPs are used solely for industrial purposes or district heating networks.

Industrial applications demand heat at a wide range of temperatures [66]. This heat demand should be as much as possible met by technologies with less usage of fossil fuel. Heat pumps can recover industrial waste heat, which cannot be used directly due to low temperatures. A part of the waste heat can be lifted again to higher temperatures by heat pumps for industrial heat supply [41].

30 2.5 High temperature heat pumps

The problem of the commercial state of heat pump technology is that the temperature limit of heat supply is at about 100 °C [65, 67, 7]. If the temperature range of heat pumps extends up to 140 °C, the application potential could be more than doubled from 270 PJ to 600 PJ per year of industrial heat demand only in Germany [68]. There are several industrial branches, in which HTHPs with a higher temperature than commercially available are applicable. These branches are for example: Pulp and paper, cement, food and beverage, metal and rubber, plastics industry and city utilities [63, 66, 69, 70]. This application potential is the motivation to develop HTHPs for higher temperatures than commercially possible today. Several research and development projects work on the extension of the applicable temperature range of HTHPs (see Table 2.4). Dissemination and applications

Until the 1990s HTHPs were already in use at temperatures up to 130 °C. They were operating mainly with the working fluid R114 [15, 50, 71], which was later forbidden [72] by the Montreal protocol due to its ODP (see section 2.4).

Table 2.3 shows a variety of exemplary HTHPs, which are currently in operation. Supply temperatures range from 70 to 100 °C. There are small HTHPs with a heat load of few hundred kilowatts to few megawatts for heating industrial processes. Large HTHPs with a heat load higher than 10 MW usually operate in district heating networks. Worth mentioning is a 100 MW heat pump used in the Moscow district heating network [73]. It is probably the world’s largest heat pump with a single compressor.


Table 2.3 Variety of exemplary HTHPs in operation

Industry branch

Country Process Tsupply [°C]

Qsupply [kW]

Reference

Metal Germany Degreasing and phosphating baths

70 260 [7]

Wood Canada Softwood drying 100 480 [74] Food Netherlands Cleaning and scalding 80 440 [75] Food Switzerland Meat processing 90 1'000 [51] Food Sweden Gelantine production 70 1'400 [76] Paper Norway Paper processing 80 2'000 [76] City utilities Norway District heating 90 14'000 [77] City utilities Sweden District heating 80 27'000 [78] City utilities Russia District heating 100 100'000 [73]

Usually, very large HTHPs are unique designs. They are designed for

particular large-scale applications. Nevertheless, there are several manufactures with series production of HTHP models (see Figure 2.11), which are adapted to their application when integrated into an industrial process.

Figure 2.11 HTHPs in series production of several manufacturers with model name if available;

Sorted by maximum supply temperature; Supply heat load is plotted partly logarithmic; References are from top to bottom: [79, 80, 81, 82, 83, 84, 85, 86, 65]

70

80

90

100

110

1 10 100 1'000 10'000

T sup

ply,

max

[°C]

Qsupply [kW]10'000 15'000 20'000 25'000

Qsupply [kW]

Simaka, Simacovery, GermanyOchsner, Austria

Advansor, DenmarkMayekawa, Eco Cute, Japan

Star Refrigeration, Neatpump, UKFriotherm, Unitop, Switzerland

Thermea, thermeco2, Germany

York, Titan OM, USACombitherm, Germany

//

100

90

80

T sup

ply,

max

[ºC

]


Situation in the field of research

There are several research projects, which try to extend the applicable temperature range of heat pumps. The state of these projects is that functional models, respectively prototypes are built to demonstrate the technical feasibility of such HTHPs. Table 2.4 lists recent research projects.

Table 2.4 Research and development projects for HTHPs

Organization, Company

Country Working fluid

Tcond [°C]

Qcond [kW]

Reference

University Tianjin China R245fa/R600 100 5 [62] Johnson Control, Electricité de France France R245fa 100 450 [87]

Dürr, Combitherm, University Stuttgart Germany R245fa 110 60 [7]

University Lyon, Electricité de France France Water 120 350 [67]

Kobe Steel Japan R245fa 120 370 [88]

Electricité de France France ECO3 (blend with R245fa) 140 200 [63, 89]

University Erlangen- Nuremberg, Siemens Germany LG6, MF2 140 10 [33, 90]

The aim of these projects is to develop heat pump systems, which can operate

in industrial environment to replace fossil fuel fired heat supply. These projects are motivated by a large application potential. For example, Bobelin et al. [63] show the application potential in different industry branches in France (see Figure 2.12).


Figure 2.12 Distribution of heat demand per industrial branch in France [63]

The majority of the projects from Table 2.4 tries to develop heat pump systems for the working fluid R245fa. The main challenge is to develop compressor technology for this working fluid and to investigate the achievable performance. Different compressor types are investigated and developed: Scroll- [63], Screw- [87, 88] and Piston-type compressors [7]. Figure 2.13 shows the investigated experimental performance for two projects from Table 2.4. The measured experimental COP results vary from 40% to 60% of the Carnot COP and are therefore in the range of commercial heat pumps.

7000 GWh a-124002200200018001600140012001000

800600400200

0

Hea

t dem

and

[GW

ha-1

]

Pulp &paper

Food &beverage

Sugar Iron &steel

Rubber &plastic

Dairy

60-69 °C70-79 °C80-89 °C90-99 °C100-119 °C120-139 °C


Figure 2.13 Photograph and experimental results from functional models working with R245fa of

research projects for high temperature heat pumps; Left side (A): Peureux et al. [89]; Right side (B): Wolf et al. [7]

From a thermodynamic point of view, R245fa is a suitable working fluid (see

section 3.3.3). However, as stated by Hanni and Rieberer [72], R245fa shows a GWP of 950 that has to be considered. The use of R245fa will be strongly limited in future (see section 2.3.3). These projects started before the restrictions by law became valid in May 2014.

There are only two projects from Table 2.4, which work with low GWP working fluids. This emphasizes the lack of suitable low GWP working fluids for the high temperature range. One of them is water. However, water vapor shows a very low density and thus the achievable COP is low. For example, Chamoun et al. achieved a maximum Carnot COP value of 38% [67] and Hackensellner 34% [15].

CO

P5

4

3

2

1

0

Compressor drive frequ. [Hz] 25 50 60 70

Tevap [°C] 504035 45 55

T con

d[°

C]

138

130

8060

120

110

100

90

A B

Tevap = 65 °C; Tcond = 110 °C


The second project of Table 2.4 that works with low GWP working fluids is part of this thesis. Reissner et al. [33, 90] presented results at condensation temperatures up to 140 °C with low GWP working fluids with Carnot COP values up to 55%.

All working fluids from Table 2.4 are included in the evaluation in chapter 3 and section 3.2 introduces them.

36 3.1 Evaluation criteria

3. Evaluation of working fluids for high temperature heat pumps

The aim of this chapter is to show the suitability of various working fluids for the

use in high temperature heat pumps. Section 3.1 presents the evaluation criteria. Section 3.2 lists and introduces all considered working fluids. The system simulation in section 3.3 evaluates the working fluids thermodynamically. Different heat pump cycles are distinguished and simulated for the corresponding working fluids in section 3.4. In the end of this chapter, working fluids are selected based on the evaluation criteria for the experimental part of this thesis (see chapter 6).

3.1 Evaluation criteria

Several properties of the working fluids function as evaluation criteria. Sections 2.3 and 2.4 introduce and explain these properties. They are distinguished in cycle independent and cycle dependent properties. Cycle independent properties:

Critical temperature (Tcrit) Ozone depletion potential (ODP) Global warming potential (GWP) Flammability Toxicity Availabilty (important for experimental evaluation)

Cycle dependent properties:

Theoretical COP (COPtheoretical) Volumetric heating capacity (VHC) Pressure ratio (pratio) Pressure level (plevel)


3.2 Considered working fluids and cycle independent properties

The literature discusses several working fluids for high temperature heat pumps. It is claimed, that all worldwide commercially available and suitable working fluids are included in the evaluation. Additionally, three new working fluids are identified during this thesis, which are for the first time introduced for HTHPs.

The following list gives a general description of the working fluids and Table 3.1 summarizes the cycle independent properties. Hydrochlorofluorocarbons (HCFCs) R114 (CClF2-CClF2) was used for supply temperatures up to 130 °C until the 1990s (see section 2.5). The Montreal protocol phased out the usage and it is forbidden today (see section 2.4). It is included in the evaluation for comparison reasons. Hydrofluorocarbons (HFCs) R245fa (CF3-CH2-CHF2) is currently the most common working fluid for the development of HTHPs. There are several research projects working with this working fluid (see section 2.5). Advantages are non-toxicity and non-flammability. A disadvantage is its high GWP of 950. It is commercially available from a few manufacturers like Honeywell and Linde [91]. R236fa (CF3-CH2-CF3) was developed as replacement for R114. It is non-flammable and non-toxic but the GWP is extremely high with a value of 9400 [49]. The availability is strongly limited and it is used only by some appliers like the American Navy [30]. R236ea (CF3-CHF-CHF2) was developed as replacement for R114 and is used in a few research applications [92]. A disadvantage is its high GWP of 1410 [93].

38 3.2 Considered working fluids and cycle independent properties

R365mfc (CF3-CH2-CF2-CH3) was developed from Solvay for high temperature applications and is used as foaming agent. The GWP is high with a value of 890 and it is flammable [45]. There are two approaches to avoid the flammability by mixing R365mfc with other working fluids. On the one hand it is mixed with R227ea [50] and on the other hand with a Perfluoropolyether, which is known as SES36 [45]. It is possible to overcome the flammability, however the GWP rises to 1110 [94] respectively 3126 [95]. Fluorocarbons (FCs) R31-10 (CF3-CF2-CF2-CF3) and R41-12 (CF3-CF2-CF2-CF2-CF3) They are perfluorinated alkanes and are used in other applications. They are included for comparison for working fluids with high molar masses. Disadvantages are the very high GWP (7000, 7500) and the toxicity of R41-12 [96, 97]. Hydrofluoroolefins (HFOs) R1234yf (CF3-CF=CH2) Honeywell and DuPont developed it as low GWP replacement for R134a. The automotive industry counts on this working fluid for air conditioning systems. A problem is its flammability (see section 2.4). R1234ze(E) (CHF=CH-CF3) Honeywell commercialized this low GWP working fluid around 2010 and first heat pumps are operating with it. For example, Friotherm built a heat pump with 16 MW heating capacity for an energy company in Stockholm. A disadvantage is its flammability [98].


R1336mzz(Z) (CF3-CH=CH-CF3) is a very recent development of DuPont especially for the high temperature range. It was introduced since 2011 under the code name DR-2. Since October 2013 the chemical formula of this working fluid is revealed as R1336mzz(Z) and some data on thermodynamic properties are available [32, 99]. The GWP is low with a value of 9. It is included in the evaluation as far as possible with the public available data. It is still not commercialized and physically not available without confidential agreements. Hence, it is not included in the selection process for the published experimental part of this thesis. Hydrochlorofluoroolefins (HCFOs) R1233zd(E) (CHCl=CH-CF3) is a very recent development of Honeywell especially for the high temperature range. Like HCFCs it contains a chlorine atom, nevertheless the ODP is extremely low (0.0003), so that it is not forbidden by the Montreal protocol. Moreover it is non-flammable, non-toxic and the GWP is low with a value of 6 [100]. Data on thermodynamic properties are freely available since July 2013. It is therefore included in the evaluation but this working fluid is still not commercialized and physically not available. Hence, it is not included in the selection process for the experimental part of this thesis. Hydrofluoroethers (HFEs) RE347mcc (CF3-CF2-CF2-O-CH3) is developed and distributed by 3M under the trade name Novec 7000. It is used as heat transfer fluid in the electronics industry [101].

40 3.2 Considered working fluids and cycle independent properties

Natural working fluids R718 (Water, H2O) has of course perfect safety and environmental properties but the VHC is very low (see section 3.3.3). There have been many research projects on water as working fluid. Gromoll [102] and Hackensellner [15] showed that the achievable real COP is too low for an efficient operation. Nevertheless, recent research projects (see Table 2.4) still use water as working fluid. It is included in the evaluation for comparison reasons. R717 (Ammonia, NH3) See section 2.4 R744 (Carbon dioxide, CO2) is a natural working fluid. The big advantages are the cheap availability and the low GWP. (Cyclic) Hydrocarbons (HCs) R600a (Isobutane, CH(CH3)3), R601 (Pentane, CH3-CH2-CH2-CH2-CH3) and Cyclopentane (-CH2-CH2-CH2-CH2-CH2-) Hydrocarbons are sometimes used for small heat pumps with small working fluid charges. These three hydrocarbons show high critical temperatures and are included for comparison [96]. Unclassified LG6 is non-flammable, non-toxic and the GWP is 1. Reissner et al. [33] presented an experimental research project with this working fluid in high temperature heat pumps. MF2 is non-flammable and non-toxic and the GWP is 1. Data on thermodynamic properties and sampling material for experimental testing are available for this thesis but the code name MF2 must be used throughout this thesis.


MF3 is non-flammable and non-toxic and the GWP is < 10. Data on thermodynamic properties are available but the code name MF3 must be used throughout this thesis. ECO3 is a code name for a working fluid blend (containing R245fa) presented from Bobelin et al. with a high GWP of 950 [63]. The exact composition of this working fluid is not given in literature and data on thermodynamic properties are not available. Hence, ECO3 is not evaluated in this thesis. Table 3.1 Heat pump cycle independent evaluation criteria of considered working fluids; Colors

are red if toxic, flammable, ODP > 0.0005, GWP > 2500; Orange if 2500 > GWP > 150, Tcrit < 130 °C; Rest is green; Sources according to list in 3.2

Working fluid

Tcrit [°C]

ODP GWP Flammable Toxic Available

R114 146 0.85 9200 no no no R245fa 154 0 950 no no yes R236fa 125 0 9400 no no yes R236ea 139 0 1410 no - yes R365mfc 187 0 890 yes no yes R31-10 113 0 7000 no no yes R41-12 147 0 7500 no yes yes R1234yf 95 0 4 yes no yes R1234ze 109 0 6 yes no yes R1336mzz 171 0 9 no no no R1233zd 166 0.0003 6 no no no RE347mcc 165 0 420 no no yes LG6 169 0 1 no no yes R718 374 0 <1 no no yes R717 132 0 <1 yes yes yes R744 31 0 1 no no yes R600a 135 0 20 yes no yes R601 197 0 11 yes yes yes Cyclopentane 239 0 11 yes no yes MF2 148 0 1 no no yes MF3 >190 0 <10 no no no

42 3.3 System simulation

The cycle independent properties of the working fluids alone are not significant enough for the selection. Hence, they are discussed together with the results of the system simulation in section 3.4, in which the working fluids are selected.

3.3 System simulation

The system simulation evaluates the working fluids thermodynamically. The thermodynamic suitability is important for the selection of working fluids for the experimental part of this thesis. The aim is to determine the cycle dependent properties (COPtheoretical, VHC, pratio, and plevel) for the temperature range (Tevap = 40 - 100 °C; Tcond = 90 - 150 °C).

3.3.1 Heat pump cycle types and corresponding working fluids

Different cycle types have to be simulated depending on the properties of the working fluid and the intended application. Three decisive aspects define the corresponding cycle type. The following subsections explain the three aspects. Critical temperature

The considered working fluids (see Table 3.1) have a wide range of different critical temperatures (31 - 374 °C). Different heat pump cycle types have to be simulated depending on the critical temperature of the working fluid and the condensation temperature. Working fluids with a high critical temperature simulate in a subcritical cycle type. If the condensation temperature is higher than the critical temperature, the cycle type is transcritical. If the evaporation temperature is also higher than the critical temperature, the cycle type is supercritical (see Figure 3.1).


Figure 3.1 Sketches of p,h diagrams of subcritical, transcritical and supercritical heat pump cycles

( ) at the standard simulation point with subcool = 5 K (see section 3.3.1); Axes are not scaled

A standard cycle simulation point is defined to compare all working fluids and

the different cycle types at Tevap = 80 °C and Tcond = 130 °C. At this standard simulation point, the working fluids classify as followed:

Subcritical R114, R245fa, R236ea, R365mfc, R41-12, R1336mzz, R1233zd, RE347mcc, LG6, R718, R717, R600a, R601, Cyclopentane, MF2, MF3

Transcritical R236fa, R31-10, R1234yf, R1234ze

Supercritical R744

At other simulation points, some working fluids change between the cycle

types subcritical and transcritical. R744 is always supercritical, because the lowest considered evaporation temperature is 40 °C. An exception is R1234yf, which is supercritical at Tevap > 95 °C.

The transcritical and supercritical cycle types show heat transfers with gliding temperatures. The working fluids in this cycle types show no phase change and assume to take up heat to a temperature as high as the evaporation temperature plus the superheat (85 °C, see Figure 3.1). The same counts for the heat generation, where the temperature decreases to a temperature equaling the condensation temperature minus the subcool (125 °C, see Figure 3.1). Consequently, the working fluids show

Enthalpy

Pres

sure

Transcritical SupercriticalSubcriticale.g. R1234zee.g. R245fa e.g. R744

125 °C85 °C 125 °C85 °C

125 °C85 °C


the same temperatures at the exit of the heat exchangers after heat uptake and heat generation independent of the cycle type.

Subcool cases

The heat sink temperature glide is different for various applications. Two

contrary examples are on the one hand a thermal phase change material storage, which is loaded at 125 °C (aligned to standard simulation point) and a hot water preparation from 20 °C fresh water to 125 °C. Therefore, two different subcool cases are simulated).

Subcool case 1 has a low degree of subcool (5 K at standard simulation point, see Figure 3.1), which ensures that all the condensation enthalpy is transferred to the heat sink. This case is more significant because the generated heat of the condenser fully transfers to the heat sink independent of the heat sinks’ temperature glide.

Subcool case 2 is simulated with a high degree of subcool (35 K at standard simulation point, see Figure 3.2). In doing so, only reasonable subcool temperatures are considered. For example, at the standard simulation point, the minimum temperature of the evaporation heat source would be 85 °C. Heating fresh water from 20 °C would at first be done up to 80 °C with the heat source directly and then to higher temperatures with the heat pump.

Figure 3.2 Sketches of p,h diagrams of subcritical, transcritical and supercritical heat pump cycles

( ) at the standard simulation point with subcool = 35 K (see section 3.3.1); Axes are not scaled

Enthalpy

Pres

sure

Transcritical SupercriticalSubcriticale.g. R1234zee.g. R245fa e.g. R744

125 °C85 °C 125 °C85 °C

125 °C85 °C95 °C 95 °C95 °C


Internal heat exchanger An internal heat exchanger (IHX) subcools the working fluid further after

condensation and transfers the heat to the suction gas to achieve a larger superheat (see Figure 4.7). Such IHX are usually used to improve the COP.

This system simulation considers the utilization of an IHX. The theoretical COP improves for all considered working fluids except for R718 and R717. Therefore, the system simulation considers an IHX for all working fluids except the two mentioned ones. Cycle types

Altogether ten different cycle types are simulated, which are distinguished by the critical temperature, the subcool case and the usage of an IHX (see Figure 3.3).


Figure 3.3 Heat pump cycle types for system simulation with corresponding working fluids; Some

fluids are named several times due to multiple operation points; IHX is internal heat exchanger

The supercritical cycle types (number 9 and 10) have no differentiation by the usage of an IHX, because only R744 and R1234yf are simulated. Here the IHX is always beneficial for the theoretical COP.


3.3.2 Calculation software, procedure and parameters Calculation software The Software Refprop in version 9.1 serves to calculate the thermodynamic properties of the working fluids. It uses the Reference Database 23 of the National Institute of Standard and Technology [93]. The working fluids LG6, MF2 and MF3 from the list in section 3.2 are used under code names throughout this thesis. The thermodynamic data is available under non-disclosure agreements with the manufacturers. The confidential thermodynamic data is delivered by the manufacturers within fluid files, which can be directly imported into Refprop.

Following thermodynamic properties are calculated with this software: Enthalpy, pressure, temperature, density, entropy and heat capacity. Single phase calculations are done by input of two state variables, whereas calculations at saturated states are done by input of one state variable. Calculation procedure and parameters

For the calculation of the cycle dependent evaluation criteria (COPtheoretical, VHC, pratio and plevel) the equations (2.4) - (2.14) are used. The numbering of the state points always starts at the compressor inlet (suction side) for all cycle types. For the cycle types with an IHX the state point numbers will change compared to them in equations (2.4) - (2.14). This is apparent by comparing Figure 2.6 and Figure 3.4.


Figure 3.4 Log p,h diagram of a theoretical vapor compression heat pump cycle with IHX at

subcool case 1 (cycle type 1); Working fluid R245fa; Standard simulation point Tcond = 130 °C; Tevap = 80 °C; s = 0.8; Superheat = 5 K; Subcool = 5 K; Isenthalpic expansion;

TIHX = 5 K

Cycle types with an IHX need additional equations for calculation. The

additional enthalpy through subcool from state point 3 to state point 4 transfers for superheating the state point 6 to state point 1:

h1 = f pevap; T3 - 5 K (3.1)

h4 = h3 + h6 - h1 (3.2) A minimum temperature difference between the two sides within the IHX of 5 K is assumed (see equation (3.1)).

The subcool case 1 always calculates with a subcool (in the condenser) of 5 K. The temperature of the subcooled (in the condenser) working fluid for the subcool case 2 is assumed to be at least 10 K higher than the evaporation heat source. Consequently, the subcool at the standard simulation point is 35 K. The difference is recognizable by comparing Figure 3.4 with Figure 3.5. A large amount of the subcool transfers to the heat sink and cannot transfer in the IHX. Thus, state point 1 superheats less in Figure 3.5.

1.8 1.85 1.91.95

6070

8090

100110

120130

140

150

3

30

250 300 350 400 450 500 550 600

p [b

ar]

h [kJ kg-1]





1

234

5 6

100

Evaporation

Expa

nsio

n

Superheatin evaporator

Subcoolin IHX Desuperheat

Condensation

Superheatin IHX

Subcoolin condenser


Figure 3.5 Log p,h diagram of a theoretical vapor compression heat pump cycle with IHX at

subcool case 2 (cycle type 3); Working fluid R245fa; Standard simulation point Tcond = 130 °C; Tevap = 80 °C; s = 0.8; Superheat = 5 K; Subcool = 5 K; Isenthalpic expansion;

TIHX = 5 K

At cycle type 3 an additional distinction has to be made. The full subcool of

35 K cannot transfer to the heat sink for some working fluids. A certain part of the subcool has to transfer inside the IHX. These working fluids are simulated with the maximum possible amount of subcool transferred to the heat sink. At the standard simulation point these working fluids are: R365mfc, R41-12, RE347mcc, LG6, R601, MF2 and MF3.

1.8 1.85 1.91.95

6070

8090

100110

120130

140

150

3

30

250 300 350 400 450 500 550 600

p [b

ar]

h [kJ kg-1]





1

234

56

100

Evaporation

Expa

nsio

n

Superheatin evaporator

Subcoolin IHX Desuperheat

Condensation

Superheatin IHX

Subcoolin condenser


Determination of pressure levels of the different cycles

The evaporation and condensation pressures of the subcritical cycle types 1 - 4 and the evaporation pressures of the transcritical cycle types 5 - 8 are given by the working fluid’s vapor pressure line. Other pressures (in transcritical and supercritical state points) have to be determined. The aim is to find the optimum pressures, where the theoretical COP is at its maximum. An optimization modeling algorithm [103] determines the pressures by finding the maximum possible theoretical COP. The algorithm calculation includes following constraints: T2 Tcond (3.3) 0.1 bar < plow < 200 bar (3.4) pcrit < phigh < 200 bar (3.5)

phigh > plow (3.6)

The lower border of the lower pressure (0.1 bar in equation (3.4)) changes to

the critical pressure for the simulation of the supercritical cycles. With the optimized pressures, the cycle simulation conducts like the subcritical cycle types after equations (2.4) - (2.14) and (3.1) - (3.2). For all calculations, an isentropic compressor efficiency is assumed to be 0.8, which is common for theoretical calculations [41]. The superheat through the evaporator is 5 K and all cycle types are calculated for the considered temperature range (Tevap = 40 - 90 °C; Tcond = 90 - 150 °C) with an increment of 10 K.


3.3.3 Results and discussion Theoretical coefficient of performance

As stated in 2.3.1, the theoretical COP is fluid-specific and its value arises from certain assumed cycle parameters (e.g. superheat = 5 K; isentropric efficiency = 0.8). It is usual to compare working fluids in this way independent of heat pump equipment design. Thus, this evaluation does not intend to forecast COP values for the experimental evaluation of this thesis. Instead, this evaluation shall give a detailed view on the thermodynamic suitability of working fluids for the high temperature range in general.

The theoretical COP shows a characteristic progression with increasing condensation temperature at a constant temperature lift. Figure 3.6 shows the progression for the exemplary working fluid R245fa over a large temperature range. At this point it is reminded, that the pressure level of the transcritical and supercritical state points are determined by a COP maximizing algorithm (see section 3.3.2 subsection 3).

Figure 3.6 Progression of the theoretical COP over a large condensation temperature range;

Exemplary working fluid R245fa; Superheat = 5 K; Subcool = 5 K; Tlift = 50 K; Tincrement = 10 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; Pressures in the transcritical and supercritical range are fixed by COP maximizing algorithm (see section 3.3.2 subsection 3),

The COP is at first increasing to a maximum in the subcritical range. Then it

is decreasing still in subcritical range and after reaching the critical temperature it is

1.52.02.53.03.54.04.55.05.56.06.5

0 50 100 150 200 250

CO

P the

oret

ical

Tcond [°C]

Subcritical Transcritical Supercritical

SubcriticalTranscriticalSupercritical

Evaluatedtemp. range


continuing to decrease in the transcritical and then supercritical range. Figure 3.7 shows the theoretical COP of all evaluated working fluids. A part of the described theoretical COP progression from Figure 3.6 (evaluated temperature range) is shown for each working fluid depending on their critical temperature. For the exemplary working fluid R245fa, it is the subcritical range near the critical point with the COP maximum.

Figure 3.7 Theoretical COP of all considered working fluids over the temperature range Tcond = 90 -

150 °C; Superheat = 5 K; Subcool = 5 K; Tlift = 50 K; Tincrement = 10 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; No IHX for R718, R717

Two contrary examples are R744 and R718. R718 has the by far highest

critical temperature (374 °C) of all considered working fluids. The considered temperature range has a large difference (> 200 K) to the critical temperature of R718. Consequently, the shown progression of the theoretical COP in Figure 3.7 is always increasing in the subcritical range. R744 has the by far lowest critical temperature below the considered temperature range and thus the theoretical COP progression is only decreasing in the supercritical range.

Working fluids with higher critical temperatures show higher theoretical COPs. R744 with the lowest critical temperature shows the lowest theoretical COP.

1.5

2.0

2.5

3.0

3.5

4.0

4.5

5.0

5.5

6.0

6.5

90 100 110 120 130 140 150

CO

P the

oret

ical

Tcond [°C]

R718R601MF3

R114

RE347mccR1336mzz

R41-12

SubcriticalTranscriticalSupercritical


At higher theoretical COPs, the transition from subcritical to transcritical cycles is shifted to higher temperatures starting with R1234yf (100 °C) and ending with R41-12 (140 °C). All working fluids with still higher theoretical COPs (above R41-12 in image plane) are subcritical. An exception is R718 at low condensation temperatures (90 - 110 °C) because of its very high critical temperature.

As stated in 2.4.1, the critical temperature should be clearly higher than the condensation temperature for good performance. The results of the system simulation in Figure 3.7 verify this for the high temperature range.

More detailed is the question, where the maximum of the theoretical COP is, compared to the critical temperature. Figure 3.7 shows the maximum of the progression of the theoretical COP for the working fluids: R600a, R236ea, MF2, R114, R41-12, R245fa, R1233zd, RE347mcc, and LG6. Figure 3.8 shows the theoretical COP progression over the distance of the condensation temperature to the critical temperature. Black points mark the maximum.

Figure 3.8 Progression of the theoretical COP over the distance of the condensation temperature to

the critical temperature; Tcond = 90 - 150 °C; Superheat = 5 K; Subcool = 5 K; Tlift = 50 K; Tincrement = 10 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K

3.5

4.0

4.5

5.0

5.5

6.0

-20 -10 0 10 20 30 40 50 60 70 80

CO

P the

oret

ical

Tcrit - Tcond [K]

LG6

R245fa

R114R236ea

R600a

RE347mcc

MF2

R1233zd

R41-12

SubcriticalTranscritical


All theoretical COP maxima are between temperature distances from 33 to 42 K (marked by blue vertical lines), at an average of 36 K. This verifies the statement in section 2.4.1 about the critical temperature influence on the COP. Coming from this result, it is possible to use the evaluation criteria “critical temperature” more detailed. It shall not only be high [11], it shall show a distinct distance to the targeted condensation temperature. Moreover, it is possible to recommend at which condensation temperatures a working fluid should be used to reach maximum efficiency.

Table 3.2 and Figure 3.9 list the results of the standard simulation point for both subcool cases.


Table 3.2 Heat pump cycle dependent evaluation criteria of considered working fluids at standard simulation point: Tcond = 130 °C; Tevap = 80 °C; Superheat = 5 K; s = 0.8; isenthalpic expansion; TIHX = 5 K; No IHX for R718, R717; Subcool case 1 (5 K) and 2 (35 K); R1336mzz is estimated (~) from [32]

Working fluid

COP VHC

[kJ m-3] pratio

plevel

[bar] COP

VHC [kJ m-3]

pratio plevel

[bar] Subcool case 1 (5 K) Subcool case 2 (35 K)

R114 5.18 5667 2.7 9-25 6.78 7079 2.7 9-25 R245fa 5.41 5639 3.0 8-23 6.86 6905 3.0 8-23 R236fa 4.49 7025 2.9 12-36 7.06 9265 2.6 12-32 R236ea 5.07 6345 2.9 10-29 6.81 8092 2.9 10-29 R365mfc 5.85 2961 3.2 4-11 6.58 3287 3.2 4-11 R31-10 4.05 5536 2.9 12-34 6.97 7710 2.6 12-31 R41-12 5.27 3237 3.1 5-15 5.88 3542 3.1 5-15 R1234yf 3.62 10038 2.5 25-64 6.73 13607 2.0 25-51 R1234ze 3.92 9323 2.7 20-54 7.54 12602 2.1 20-41 R1336mzz ~5.6 ~3400 ~3.1 ~4-14 ~6.6 ~3900 ~3.1 ~4-14 R1233zd 5.56 4779 2.9 7-19 6.84 5728 2.9 7-19 RE347mcc 5.61 3154 3.1 4-13 6.35 3502 3.1 4-13 LG6 5.63 2111 3.3 3-9 5.96 2216 3.3 3-9 R718 5.81 744 5.7 0-3 6.10 781 5.7 0-3 R717 4.55 24992 2.6 41-109 5.84 32114 2.6 41-109 R744 2.06 14518 1.5 135-200 2.75 21560 1.6 128-200 R600a 4.87 7118 2.5 13-34 6.60 9189 2.5 13-34 R601 5.89 2925 3.0 4-11 6.63 3251 3.0 4-11 Cyclo- pentane

6.05 2227 3.2 3-8 6.92 2534 3.2 3-8

MF2 5.11 3551 3.0 6-17 5.83 3965 3.0 6-17 MF3 5.85 1153 3.8 1-5 5.98 1177 3.8 1-5

The theoretical COP increases by raising the subcool. Figure 3.9 shows both subcool cases. The increase from subcool case 1 (5 K) to subcool case 2 (35 K) varies significantly among the working fluids. For example, the theoretical COP of R245fa increases by +27%, whereas R1234yf increases by +86%.


Figure 3.9 Theoretical COP of all considered working fluids at the standard simulation point: Tcond

= 130 °C; Tevap = 80 °C; Superheat = 5 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; No IHX for R718, R717; Subcool case 1: 5 K; Subcool case 2: 35 K; Tmin.2-sat.vap. = 5 K; R1336mzz is estimated from [32]

The different regions in the log p,h diagram (see Figure 3.10) where the

subcool takes place, explain this diverse behavior. The three working fluids R245fa, R1234yf and R744 serve for explanation. At the standard simulation point R245fa is subcritical and the subcool takes place from the saturated liquid line into the liquid phase. The isotherms in this region proceed approximately parallel, meaning that a constant amount of enthalpy transfers per Kelvin. The same counts for R744, which is supercritical. The distance between the isotherms here is approximately constant. In contrast to that, the subcool of R1234yf transfers near the critical point. Here the isotherms proceed with large deviations in the distances to each other. Consequently, the enthalpy difference of the subcool is larger at the same temperature difference. This is illustrated by the arrows (see Figure 3.10), which show the same temperature difference but are of different length, which is correlated to the enthalpy difference.

0.00.51.01.52.02.53.03.54.04.55.05.56.06.57.07.58.0

CO

P the

oret

ical

Subcool case 1 (5 K) Subcool case 2 (35 K)


Figure 3.10 Sketch of a log p,h diagram with location of the region of subcool compared to the phase

boundary for the exemplary working fluids R245fa, R1234yf and R744; Temperature range of subcool is same for all three working fluids

Figure 3.11 Increase of the theoretical COP depending on degree of subcool; Tcond = 130 °C; Tevap =

80 °C; Superheat = 5 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; For exemplary working fluids R245fa, R1234yf, R744

Hence, the theoretical COP increases in different extent depending on the cycle type. Transcritical cycles improve significantly and nonlinear to the degree of

Enthalpy

Pres

sure

R245fa

R1234yf

R744— Isotherm

R1234yf

CO2

R245fa

012345678

5 10 15 20 25 30 35

CO

P the

oret

ical

Subcool [K]

+33%

+27%

+86%

Case 1 Case 2

R744


subcool (e.g. R236fa, R31-10, R1234yf and R1234ze). Subcritical cycles and supercritical cycles improve for about one third linear to the degree of subcool. Volumetric heating capacity

Figure 3.12 shows the VHC of the considered working fluids. R717 and R744 are not displayed. Due to their very high pressure level (> 100 bar) at the considered temperature range, the density is very high and so is the VHC. Their VHC is about 25’000 and 15’000 respectively. Similar behavior is shown by the working fluids in transcritical cycles, which show predominantly higher VHCs than the working fluids in subcritical cycles.

Figure 3.12 VHC of all considered working fluids over the temperature range Tcond = 90-150 °C;

Superheat = 5 K; Subcool = 5 K; Tlift = 50 K; Tincrement = 10 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; No IHX for R718

The practical limit for efficient compressor design is below 500-1000 kJ m-3

(see 2.3.1). R718 and MF3 are predominantly below that limit and thus are not suitable. LG6 and Cyclopentane are closely above the limit and are at higher

0

1000

2000

3000

4000

5000

6000

7000

8000

9000

10000

11000

12000

90 100 110 120 130 140 150

VH

C [k

J m-3

]

Tcond [°C]

R365mfc

R245faR114

R236eaR31-10

RE347mccR41-12

SubcriticalTranscritical


temperatures in the range of usual VHC values. Consequently, they are suitable at higher temperatures.

The VHC values at the subcool case 2 do not significantly change compared to case 1 and thus are not shown over the whole temperature range. Table 3.2 shows them for the standard simulation point.

3.4 Selection of the working fluids for experimental investigation

The results of the system simulation (section 3.3) and the safety and environmental properties (section 3.2) are the basis for the selection. The aim is to find thermodynamic suitable, future-proof (no ODP, low GWP) and safe working fluids for experimental investigation. Following working fluids are excluded:

R744 (carbone dioxide) shows a too low theoretical COP and a very high pressure level. R717 (ammonia) is flammable and toxic. Moreover, the COP is low and it has a very high pressure level [72]. R236fa, R31-10, R1234yf, R1234ze and R600a show theoretical COPs below 5. They can be operated efficiently only with a high degree of subcool and so are depending on special requirements to the heat sink. This significantly narrows possible applications. The COP limit of ~5 arises from section 2.3.1 (COPexperimental = 50% of COPtheoretical) and section 2.3.3 (COP > 2.3 for CO2 factor reduction) R41-12 has a very high GWP and will be forbidden in near future. Moreover, it is flammable. R718 (water) and MF3 show a too low VHC.

60 3.4 Selection of the working fluids for experimental investigation

R365mfc, R601 and Cyclopentane show good thermodynamic suitability. However, flammability is a serious safety issue and these working fluids are excluded in the first place for the experimental work.

Table 3.3 shows the remaining working fluids. They are all non-flammable, non-toxic and have no significant ODP or very high GWP. Table 3.3 Evaluation criteria at standard simulation point: Tcond = 130 °C; Tevap = 80 °C; Superheat

= 5 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; Subcool case 1 (5 K); Tmin.2-sat.vap. = 5 K; R1336mzz is estimated (~) from [32]; Colors are red if not available, orange if GWP > 150, Rest is green

Working fluid

GWP Available COP VHC

[kJ m-3] pratio

plevel

[bar] R245fa 950 yes 5.41 5639 3.0 8-23 R236ea 1410 yes 5.07 6345 2.9 10-29 R1336mzz 9 no ~5.60 ~3400 ~3.1 ~4-14 R1233zd 6 no 5.56 4779 2.9 7-19 RE347mcc 420 yes 5.61 3154 3.1 4-13 LG6 1 yes 5.63 2111 3.3 3-9 MF2 <10 yes 5.11 3551 3.0 6-17

The COP varies around 5.4 with maximum deviations of 7% among the

working fluids (Table 3.3) with LG6 having the highest COP. In contrast to that, LG6 has the lowest VHC with -50% compared to the average of the working fluids. The working fluids with a high GWP (R245fa, R236ea) show the best VHC values.

The pressure ratio of all working fluids is around three and is therefore not crucial for the selection. At the pressure level, LG6 has a light advantage with a level lower than 10 bar.

All working fluids from Table 3.3 can be used in mid-term future (< 2020). The phase-out plans of the European Union, however will force to use low GWP working fluids in long-term future. Therefore, R245fa, R236ea and RE347mcc are not selected, because this thesis focuses on research for long-term future (> 2020).

R1336mzz and R1233zd are the best choice by thermodynamic suitability but are not commercially available before 2015. They should be considered for future research work. LG6 and MF2 are selected for further experimental evaluation (see chapter 6). They are the only future-proof and safe working fluids with good theoretical thermodynamic suitability, which are commercially available in 2012.


This evaluation and selection validates the trend towards complex molecules in working fluid usage (see section 2.4.2). Only the working fluids R1336mzz, R1233zd, LG6 and MF2 are thermodynamic suitable, safe and future-proof. They show complex molecules with high molar masses. Chapter 4 investigates the consequences for heat pump operation.

62 4.1 Nature of the overhanging behavior

4. Operation strategies for overhanging working fluids

The evaluation and selection of the working fluids (chapter 3) confirm the trend towards complex molecules (section 2.4.2). Both selected working fluids (section 3.4) show a distinct overhanging behavior. It is probable that more and more complex working fluids operate in heat pumps in the future.

As stated in section 2.4.2, the prevention of condensation during the compression is mandatory. Therefore, it is important to investigate and understand the overhanging behavior for a reliable operation with these working fluids.

This chapter analyses in detail the overhanging behavior (section 4.1) and defines new terms for its characterization (section 4.2). The summary and formulation of boundary conditions (section 4.3) shows the capabilities of these working fluids and provides the basis for operation strategies. The overhanging behavior requires altered heat pump cycle control and leaves room for optimization possibilities. Section 4.4 investigates these two aspects.

4.1 Nature of the overhanging behavior 4.1.1 Quantification by the slope of the saturated vapor line

Different working fluids vary significantly in their extent of the overhanging behavior. The slope of the saturated vapor line is a good indicator to quantify this extent. Working fluids, which are near the transition from bell-shaped to overhanging, show extremely high slopes. This is a problem for comparison of working fluids. Consequently, the inverse slope (IS) of the saturated vapor line (see equation (4.1)) is more useful [56].

Lai and Fischer [56] define a mean slope for each working fluid by taking the minimum and the maximum entropy of the positive slope range (see Figure 4.1). However, this thesis focuses on the high temperature range. If a specified temperature is considered, there is a deviation from the definition to the actual slope of the tangent on the saturated vapor line. Therefore, the IS for a specified temperature (e.g. standard simulation point: Tevap = 80 °C) serves in this thesis to describe the overhanging behavior (see equation (4.1) and Figure 4.1).


IS = ssat.vap.

T=

ssat. vap., T + 1 K - ssat. vap.,T - 1 K

2 K (4.1)

Figure 4.1 Different definitions of the slope of the saturated vapor line; Lai and Fischer is reference

number [56]

4.1.2 Linear correlations and their coverage on molecule classes

The heat capacity of the alkane series increases with the molar mass (see section 2.4.2). Larger molecules are capable of storing more energy per mol due to the increased amount of modes of motion. Figure 4.2 shows this correlation for the working fluids evaluated in chapter 3. These working fluids are from a variety of different molecule classes. The trend line shows a linear correlation for the classes: Hydrofluoroolefin (HFO), Hydrofluorocarbon (HFC), Hydrofluoroether (HFE), Perfluorocarbon (FC), and other (LG6, MF2, MF3), marked by black dots. The trend line equation is as follows:

cv [J mol-1 K-1] = 0.79 J K-1g-1 M + 30.9 J mol-1 K-1 (4.2)

Tem

pera

ture

Entropy

A: Definition according toLai and Fischer

B: Own definition usedin this thesissmax

smin

Ttarget


Figure 4.2 Correlation of molar mass and isochoric heat capacity for working fluids of evaluation;

cv is calculated at 80 °C saturated vapor; Trend line is created for HFOs, HFCs, HFEs, FCs, LG6, MF2 and MF3 (black dots); — Trend line; --- Parallel shifted trend lines for working fluids with chlorine content

These working fluid classes consist of a carbon chain with fluorine and

hydrogen atoms. Inserted functional groups like ether (HFE), olefin (HFO), perfluorinated carbons (FC) or other (LG6, MF2, MF3) do not interfere with the linear correlation. Despite the deviating molecular structure, the equation (4.2) is valid by approximation. For example, compared to HFCs, HFOs loose two atoms (less molar mass) and two molecule bonds (less heat capacity). Both losses balance each other.

R114 and R1233zd deviate from equation (4.2) by 52 respectively 30 g mol-1 (see Figure 4.2). They also consist of a carbon chain with fluorine and hydrogen, but have chlorine atoms in their molecules as well. Chlorine has a high molar mass, thus the molar mass of the molecules is increasing much. However, the number of molecule bonds remains the same and the heat capacity is not increasing. This is the reason for the deviation of R114 and R1233zd to higher molar masses respectively lower heat capacities.

With consideration of the chlorine content, it is possible to apply equation (4.2) to R114 and R1233zd. If chlorine atoms (35 g mol-1) theoretically substitute the fluorine or hydrogen atoms (mean value 10 g mol-1), the deviation compared to the trend line reduces. Without consideration, the deviation is 33% respectively 21%.

0

50

100

150

200

250

300

350

0 50 100 150 200 250 300 350 400

c v[J

mol

-1K

-1]

M [g mol-1]

MF3LG6

R41-12MF2

R31-10RE347mcc

R114 M = 52 g mol-1)R1233zd M = 30 g mol-1)

R718

R717

HFCs:R236faR236eaR365mfcR245fa

HFOs:R1234zeR1234yfHCs:

R601CyclopentaneR600a

nCl = 1

nCl = 2

nCl = 0


With consideration, the deviation is 1% respectively 3% (see Table 4.1). With this in mind, it is possible to modify equation (4.2) to:

cv [J mol-1 K-1] = 0.79 J K-1g-1 (M - nCl 25 g mol-1) + 30.9 J mol-1 K-1 (4.3) In this form, equation (4.3) represents the parallel shifted trend lines in Figure 4.2. Table 4.1 Consideration of the chlorine content for compatibility of the working fluids R114 and

R1233zd with equation (4.2)

Working fluid R114 R1233zd

Molecular structure

M [g mol-1] 171 130

cv

[J mol-1 K-1] 125.0 110.3

cv after (4.2) with M [J mol-1 K-1] 166.0 133.6

Deviation from trend line 33% 21% MF/H for Cl

[g mol-1] 121 105

cv after (4.2) with MF/H for Cl

[J mol-1 K-1] 126.5 113.9

Deviation from trend line 1% 3%

The HCs deviate from the trend line. They have no fluorine content and thus shift towards lower molar masses.

According to equation (2.15), the IS correlates to the isochoric heat capacity (cv ~ IS) and from equation (4.2) the isochoric heat capacity correlates to the molar mass (M ~ cv). Figure 4.3 shows these two correlations of the three terms M, cv and IS.

CC l

F

F

C

F

C l

F

CC l

HC

H

C

F

F

F


Figure 4.3 Correlation of inverse slope, molar mass and isochoric heat capacity for working fluids

of evaluation, which follow equation (4.2) and Tcrit > 130 °C; Molar mass is calculated with equation (4.2) and chlorine content is considered after Table 4.1; — Trend line

The IS increases linear with the isochoric heat capacity and the molar mass by

approximation. Consequently, equation (2.15) is also valid for the considered high temperature working fluids. The trend line equation is as follows:

IS [J mol-1 K-2] = 2.83 10-3 cv K-1- 98.1 J mol-1 K-2 (4.4)

Equations (4.2) and (4.4) show that the overhanging behavior in terms of IS is directly depending on the molar mass. It is possible to determine the overhanging behavior only by knowing the molar mass without any further thermodynamic properties. This is valid for the mentioned molecule classes in the range of the saturated vapor line with a positive slope.

y = 353.4x + 98.113

90

140

190

240

290

340

390

100

150

200

250

300

350

0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7

M [g

mol

-1]

c v[J

mol

-1K

-1]

IS [J mol-1 K-2]

MF3

LG6R41-12

MF2

RE347mcc

R365mfc

R1233zdR114

R245faR236ea

400

350

300

250

200

150

100


4.2 Consequences for the operation in heat pumps

After the description of the overhanging behavior in section 4.1, the question is which consequences arise for the operation in heat pumps. It is apparent, that working fluids with a distinct overhanging behavior are not operable in a standard heat pump cycle. The compression proceeds along the isentropes with the deviation of the isentropic efficiency. Figure 4.4 shows a bell-shaped and an overhanging working fluid in p,h diagrams. The isentropes of the bell-shaped working fluid emerge from the 2-phase region, thus the compression cannot enter the 2-phase region. In contrast to that, the isentropes of the overhanging working fluid cross the 2-phase region and the compression ends in partly liquid state (see state point 2 in Figure 4.4).

Figure 4.4 Bell-shaped (R134a) and overhanging (LG6) working fluid in a standard heat pump

cycle shown in p,h diagrams

Compression end (2)in 2-phase region

Bell-shaped(e.g. R134a) Overhanging (e.g. LG6)Compression end (2)

in vapor phase

1

2

4

3

7

20

3040

5060

7080

90100

10

1.75

1.8

1.85

3250 300 350 400 450 500

p [b

ar]

h [kJ kg-1]


50

asdf

1.71Isotherm [°C]Isentrop [kJ kg-1 K-1]Heat pump cycleasdf



1.71

6070

8090

100110

120130

140150

1.5 1.6

1.65

1

50

300 350 400 450

p [b

ar]

h [kJ kg-1]


1

23

4

1.55

68 4.2 Consequences for the operation in heat pumps

Consequently, overhanging working fluids need alternative heat pump cycles for a stable operation. The state point 1 of the overhanging working fluid in Figure 4.4 needs large superheat so that state point 2 is in vapor phase as well. 4.2.1 Definition of the minimum required superheat

Working fluids with a high IS need a larger superheat than working fluids with a low IS. The term “minimum required superheat” (minSH) is here defined to characterize the extent of superheat (see Figure 4.5 and equations (4.5) - (4.9)). It is the minimum superheat required at state point 1 to reach a 5 K distance to the saturated vapor line at the compressor outlet. The 5 K distance implies to prevent entering the 2-phase region by minor fluctuations.

Figure 4.5 Scheme of the definition of the minimum required superheat (minSH); 2

compression process with isentropic efficiency 0.8; The diagram is only for explanation purpose and consists of no quantitatively information

minSH = T1 - Tevap (4.5)

T1 = f pevap; h1 (4.6)

h1 = h2 -h2 - h1s

s (4.7)

h2 = f Tcond + 5 K; pcond (4.8)

h1s = f pevap; s2 (4.9)

Entropy

minSH [K]

T2 - Tsat.vap. at p 5 K

s = 0.8

Tem

pera

ture

Entropy

2

1

2


Figure 4.6 shows the minSH of the working fluids from Figure 4.3 as a function of the IS. There is a linear correlation of the working fluids with a high minSH by approximation.

Figure 4.6 Correlation of the minimum required superheat (minSH) and inverse slope for working

fluids of evaluation, which follow equation (4.2) and Tcrit > 130 °C; minSH is calculated for standard simulation point: Tcond = 130 °C; Tevap = 80 °C; s = 0.8; --- Limit of superheat to be provided by evaporator; — Trend line for working fluids with high minSH

The equation of the trend line is as follows: minSH [K] = 37.6 IS K2 mol J-1 + 12.8 K (4.10)

A superheat by the evaporator heat source of up to 10 K is possible in standard heat pump cycles (dashed red line in Figure 4.6). The working fluids R1233zd, R245fa, R114 ad R236ea require less than 10 K of minSH. Consequently, they need no alternative heat pump cycle. In contrast to these working fluids, all others need a minSH of higher than 15 K. The inlet temperature of the evaporator heat source must be significantly higher than the evaporation temperature to provide such a high superheat. Therefore, working fluids with a high minSH need an alternative cycle with an additional heat source.

Ultimately, the term minSH classifies working fluids in standard cycle operable and alternative cycle operable. The IS defines the border for cycle alteration,

0

5

10

15

20

25

30

35

40

0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7

min

SH [K

]

IS [J mol-1 K-2]

MF3LG6

R41-12MF2RE347mcc

R365mfc

R1233zdR114

R245fa

R236ea


which situates between 0.1 and 0.2 J mol-1 K2 and is here assumed to be 0.15 J mol-1

K-2 for the following sections.

4.2.2 Internal heat exchanger as heat source for the additional superheat

The easiest way to reach the minSH is to use an external heat source like an electric heater or steam. However, an external heat source consumes additional energy, which reduces the COP of the whole heat pump system. It is better to use a heat source, which comes from within the cycle without additional energy consumption.

An internal heat exchanger (IHX) subcools (3 4) the condensed working fluid and superheats (6 1) the evaporated working fluid (see Figure 4.7). Usually the purpose of an IHX is to increase the temperature of the suction line to prevent water of the surrounding air from condensing. Another purpose is to improve the COP by increasing the evaporator pressure with a stable superheat [104].

In contrast to these purposes, in which the IHX is an optional apparatus, the IHX for alternative cycle design (IS > 0.15 J mol-1 K-2) is indispensible.


Figure 4.7 Heat pump cycle with internal heat exchanger (IHX) and overhanging working fluid

(e.g. LG6) in p,h diagram; Transferred heat by IHX from subcool (3 4) to superheat (6 1); Calculated at standard simulation point: Tcond = 130 °C; Tevap = 80 °C;

s = 0.8; Subcool = 5 K; Superheat = 5 K; Isenthalpic expansion; TIHX = 5 K; T,Q diagram of IHX heat transfer

4.2.3 Limiting temperature lift as stationary boundary condition

The amount of heat from the subcool in the IHX must be enough to superheat the evaporated working fluid up to the minSH. The maximum transferable heat amount ( hIHX, transferable) depends on the working fluid and the operation point. The IS varies slightly with the temperature (in the positive regime). Thus, the transferable heat amount varies slightly as well. Another reason is the increasing heat capacity (+11%) of the vapor with increasing evaporation temperature from 30 to 90 °C. The influence of the temperature lift is much larger. The higher the temperature lift, the higher is the transferable heat amount (see Figure 4.8).

Overhanging (e.g. LG6) Compression end (2)in vapor phase

70

80

90

100

110

120

130

140

0% 50% 100%

T [°

C]

Q [%]

6

1

4

3TIHX

Tcond

Tevap


Figure 4.8 Transferable heat amount through IHX at different operation points: Tevap = 30 - 90 °C;

Superheat = 5 K; Subcool = 5 K; Isenthalpic expansion; TIHX = 5 K; Working fluid LG6

The reason for the increase is the rising temperature difference of the condenser outlet and evaporator outlet (10 K at TLift = 20 K; 50 K at TLift = 60 K). The minSH corresponds to a minimum required heat amount ( hrequired). The transferable heat amount has to be higher than the required heat amount for operation with the IHX only. If the required heat amount is higher than the transferable heat amount, an additional external heat amount ( hexternal) is required.

hexternal = hrequired - hIHX, transferable (4.11)

5

10

15

20

25

30

35

40

45

20 25 30 35 40 45 50 55 60

h IH

X, t

rans

fera

ble[k

J kg-1

]

TLift [K]

90 °C70 °C50° CT = 30° Cevap


Figure 4.9 Heat amount terms (transferable, required and external) versus the temperature lift;

Calculated at standard simulation point: Tevap = 80 °C; s = 0.8; Subcool = 5 K; Superheat = 5 K; Isenthalpic expansion; TIHX = 5 K; Working fluid LG6

The transferable and the required heat amount show a point of intersection (see Figure 4.9). The temperature lift at this intersection is here defined as new term “limiting temperature lift” (LTL). The LTL is a working fluid-specific term and specifies whether a heat pump cycle with an overhanging working fluid is operable with an IHX alone or not.

Figure 4.10 shows the correlation of the LTL to the IS. The LTL increases linear with IS by approximation. The trend line equation is as follows: LTL [K] = 40.8 IS K3 mol J-1 + 13.4 K (4.12)

Consequently, the overhanging behavior in terms of IS is also an informative

value for the feasibility of targeted operation points.

0

510

15

20

25

3035

40

45

20 25 30 35 40 45 50 55 60

h [k

J kg-1

]

Tlift [K]

h

h

h

Limiting temperature lift (LTL)

external

required

IHX, transferable

74 4.3 Summary of linear correlations and their significance for cycle design

Figure 4.10 Correlation of the limiting temperature lift (LTL) and inverse slope for considered

working fluids of evaluation, which follow equation (4.2) and Tcrit > 130 °C; LTL is calculated for standard simulation point: Tevap = 80 °C; s = 0.8; — Trend line

4.3 Summary of linear correlations and their significance for cycle design

The IS quantifies the overhanging behavior. The molar mass correlates linear to

the IS: M ~ cv ~ IS (equations (4.2) and (4.4)).

The new term minSH correlates linear to the IS (equation (4.10)). At the border of IS = 0.15 J mol-1 K-2 working fluids distinguish in standard cycle operable and alternative cycle operable. The alternative cycle needs an additional heat source and the IHX is recommended as heat source.

The new term LTL as well correlates linear to the IS (equation (4.12)). With the LTL, it is possible to state whether an targeted operation point is feasible with the IHX alone or whether an additional external heat source is necessary.

15

20

25

30

35

40

45

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7

LTL

[K]

IS [J mol-1 K-2]

MF3

LG6

R41-12MF2

RE347mccR365mfc

R1233zd

R114

R245faR236ea


Figure 4.11 Graphical summary of linear correlations and their significance on cycle design;

Numbers in brackets refer to equations

Ultimately, only with the knowledge of the molar mass of a working fluid, it is

possible to predict the necessary cycle design. These correlations base by approximation on the considered high temperature working fluids of different molecule classes in this thesis. Others might deviate more, but it is very probable that these correlations also count for other high temperature working fluids of same molecule classes.

For example, R1336mzz is a HFO with a molar mass of 164 g mol-1 and the minSH at the standard simulation point is approximately 18 K ± 2 K (inaccuracy in graphical readout) [99]. The predicted minSH by equations (4.2), (4.4) and (4.10) is 19.1 K and so fits exactly.

4.4 Control strategies and optimization possibilities In standard heat pump cycles, a thermostatic or electronic expansion valve

controls the flow through the expansion by the superheat at the exit of the evaporator (see section 2.3.2 and Figure 4.12 A). In alternative cycles with an IHX, the control variable changes from the superheat at the evaporator outlet to the superheat at the compressor outlet. The working fluid at the compressor outlet shall superheat by at least 5 K distance to the saturated vapor line (see definition of minSH in section 4.2.1). Depending on the working fluids’ LTL and the operation point, different strategies for control are applicable.

cv

minSH

LTL

Border0.15 J mol-1 K-2

M IS(4.2) (4.3)

Standard cycle

Alternative cyclewith IHX

TLift > LTL

TLift < LTL

Operation point isfeasible with IHX

Additional externalheat source

76 4.4 Control strategies and optimization possibilities

Figure 4.12 Differences in control of A: Standard heat pump cycle and B: Alternative heat pump

cycle with IHX; SH: Superheat

Temperature lift lower than the limiting temperature lift (TLift < LTL)

If the temperature lift of a targeted operation point is below the LTL, an

external heat source is necessary. The maximum outlet temperature of the IHX vapor side (state point 1) depends only on the condenser outlet (state point 3). Consequently, a superheat increase inside the evaporator is not helpful. This would decrease only the transferred heat amount in the IHX, but the outlet temperature (state point 1) would be the same. Thus, the evaporator controls simply by its superheat, like in the standard cycle.

An additional control mechanism is necessary for the compressor outlet superheat. Three control mechanisms are suggested here. Calculations with LG6 at the standard simulation point and at different temperature lifts (equations (2.4) - (2.12), (3.1) - (3.2) and (4.5) - (4.9)) show the external energy consumption. A condenser heat load of 10 kW sets the reference point for comparison of the different control mechanisms.

The easiest way is to integrate a trace heater between the IHX and the compressor (see Figure 4.13). The trace heater adjusts its heat load to control the superheat at the compressor outlet.

A: Standard heat pump cycle B: Alternative cycle with IHX

Com-pressor

2

Expansionvalve

3

4

1

Evaporator

Condenser

5 6

IHX

SH? ?


Figure 4.13 Control mechanism #1: Standard evaporator control and trace heater for compressor

outlet control; SH: Superheat

Another possibility is a hot gas bypass, which shifts a part of the pressure vapor

to the suction vapor. The mixing results in a higher superheat, which originates from an increased compressor power due to a higher mass flow (see Figure 4.14).

Figure 4.14 Control mechanism #2: Standard evaporator control and hot gas bypass compressor

outlet control; SH: Superheat

The compression outlet temperature increases with the pressure ratio [105]. It is

possible to reduce the evaporation pressure for an increase of the pressure ratio. Consequently, the compression outlet temperature rises. The reduction of the evaporation pressure easily adjusts by the opening of the expansion valve. The temperature of the heat source for the evaporator stays the same, thus the superheat in the evaporator rises. However, the compressor inlet temperature is the same due to the IHX. Only the compressor inlet pressure reduces.


Figure 4.15 Control mechanism #3: Evaporator pressure adjustment to control compressor outlet;

SH: Superheat

The three control mechanisms show different additional external energy

consumptions (see Table 4.2). However, the energy form and its efficiency need a more detailed consideration. Table 4.2 Additional external energy consumption of the three control mechanisms; Working fluid

LG6

Tevap

[°C]

Tcond

[°C]

Additional external energy consumption #1 #2 #3 [kW] [kW] [kW]

80 66 100 0.80 0.69 1 0.52 3 80 76 110 0.23 0.20 2 0.14 4

80 120 0 5 0 5 0 5

1: mcompressor +105% 2: mcompressor +21% 3: pevap reduced 2.65 1.77 bar

4: pevap reduced 2.65 2.37 bar

5: Tlift > LTL; External heat source is not necessary

Control mechanism #3 shows the lowest energy consumption. It is an additional

compressor power and the mechanical and electrical efficiency have to be accounted for a certain heat pump system. The same counts for the control mechanism #2.

The heat transfer of the trace heater also shows a certain efficiency. Moreover, it is important if electric energy or for example, steam (thermal energy) serves for heating.


With all these considerations, it is possible to determine an adequate external heat source and control mechanism for a certain heat pump application.

Temperature lift higher than the limiting temperature lift (TLift > LTL)

A high temperature lift results in unnecessarily high superheat of the compressor outlet. The transferable heat amount through the IHX exceeds largely the required heat amount (see Figure 4.9). This situation yields freedom to the control of such operation points, which leaves room for optimization possibilities. The aim is to increase the achievable experimental COP.

Figure 4.16 shows an operation point with a temperature lift of 60 K. The compressor outlet temperature (state point 2) is 16.2 K superheated. According to the definition of the minSH (see section 4.2.1), this superheat can be reduced to 5 K.

Additionally to the introduced log p,h diagram (Figure 2.6), this diagram shows the isochors (green lines) to investigate the progression of the density (inverse specific volume).

Figure 4.16 Heat pump cycle with working fluid LG6 in p,h diagram; Calculated at: Tcond = 130 °C;

Tevap = 70 °C; s = 0.8; Subcool = 5 K; Superheat = 5 K; Isenthalpic expansion; TIHX = 5 K

34

5 66050

70

1.55

80

80

90

90

100110

120130

140150

1.5 1.6

1.65

10

20

40

70100

1

10

250 300 350 400 450

p [b

ar]

h [kJ kg-1]

isotherm [°C]isentrop [kJ kg-1 K-1]

heat pump cycleisochor [kg m-3]

Isotherm [°C]Isentrop [kJ kg-1 K-1]

Heat pump cycleIsochor [kg m-3]

50

h2,3

h2,1

1

2’

1’ 1

2


The reduction of the superheat leads to altered state points (1 1’ and 2 2’). The important parameters for the COP are the condenser and compressor enthalpy differences ( h2,3 and h2,1) and the compressor inlet density ( 1), which influences the VHC. The changes of these parameters due to the altered state points are as follows:

h2,3 - 15% h2,1 - 5% 1 + 4%

The change of the compressor enthalpy difference ( h2,1) and the density are

beneficial. However, the reduced condenser enthalpy difference ( h2,3) is a disadvantage. Hence, this is an antidromic optimization issue and the real effect of this optimization possibility needs experimental evaluation (see section 6.2.5 and 6.3.3). Two methods for reducing the superheat are suggested here and are used for the experimental evaluation: Method 1: Increase of the subcool (see Figure 4.17)

If the subcool before the IHX increases (3 3’), there is less transferable heat amount for the IHX left. Consequently, state point 1 superheats less and as a result state point 2 as well. This method has the advantage, that if the additional subcool fits to the heat sink, the disadvantage of the reduced enthalpy difference diminishes ( h2,3 + 3% instead of – 15%). Method 2: Partly shifted evaporation in IHX (see Figure 4.18)

If the evaporation takes place partly in the IHX (6 6’), the heat amount from subcool in the IHX for the superheat is smaller. Consequently, state point 1 superheats less and as a result state point 2 as well.

Both methods are experimentally evaluated (section 6.2.5) and discussed (section 6.3.3) in the experimental part of this thesis.


Figure 4.17 Optimization method 1: Increase of subcool; Heat pump cycle with working fluid LG6

in p,h diagram; Calculated at: Tcond = 130 °C; Tevap = 70 °C; s = 0.8; Subcool = 16.9 K; Superheat = 5 K; Isenthalpic expansion; TIHX = 5 K

Figure 4.18 Optimization method 2: Partly shifted evaporation in IHX; Heat pump cycle with

working fluid LG6 in p,h diagram; Calculated at: Tcond = 130 °C; Tevap = 70 °C; s = 0.8; Subcool = 5 K; Superheat = 0 K (evaporator outlet is in 2-phase state); Isenthalpic expansion; TIHX = 5 K

34

5 66050

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1.55

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80

90

90

100110

120130

140150

1.5 1.6

1.65

10

20

40

70100

1

10

250 300 350 400 450

p [b

ar]

h [kJ kg-1]





50

2’

1’

’ 32

1

34

5 66050

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1.55

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90

100110

120130

140150

1.5 1.6

1.65

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p [b

ar]

h [kJ kg-1]





50

2’

1’’ 6

2

1

82 5.1 Importance of the dynamic simulation

5. Dynamic simulation of the start-up

Chapter 4 analyses the consequences for the operation of overhanging working fluids. Operation characteristics and boundary conditions are identified and are valid for stationary operation points.

Another question is the topic of this chapter: What is the influence of the overhanging behavior on the heat pump start-up procedure? A simple start is not possible, due to immediate working fluid condensation in the compressor.

This chapter analyses the start-up by dynamic simulations. The aim is to identify dynamic boundary conditions of heat pump operation with overhanging working fluids.

Section 5.1 describes the importance and the surrounding conditions of this chapters’ question. Section 5.2 refers to the variables, which influence the start-up progression. In doing so, two limiting cases for the start-up are illustrated. Section 5.3 explains the methodology and the execution of the dynamic simulation. In the end, section 5.4 shows the results and discussion.

5.1 Importance of the dynamic simulation

At idle state, all heat pump components have the temperature of the ambient air, for example 20 °C. During the compressor start, the internal heat exchanger (IHX) delivers no heat, because it is still cold. Consequently, the minimum required superheat (minSH) is not present at start-up. At the beginning, the compressor, the condenser, connecting pipes and then the IHX must heat up (see Figure 5.1 orange marks). These components have a thermal inertia, which leads to a delay of the transferred heat to the heat sink (condenser) and to the suction gas (IHX).


Figure 5.1 Heat pump cycle of overhanging working fluids for the dynamic simulation; Orange:

Components with thermal inertia, which delay the heat transfer to the suction gas

The delay for the heat sink is usual for all working fluids. However, the delay

for the IHX is a problem for fluids with high IS, because they are not superheated to the minSH right after the compressor start.

External heat sources like a trace heater (see Figure 4.13) are required for the start-up to avoid liquid compression. This consumes additional energy, which has to be considered. Especially for heat pumps, which turn on and off frequently (e.g. in a smart grid), the external energy consumption could lower the economical benefit.

Another interesting question is the delay time of the condenser heat transfer to the heat sink. At short operation times, the external energy consumption of the start-up might foil the transferred heat through the condenser. Here the aim is to identify a minimum operation time. It describes the minimum time, which needs to operate after the start-up so that the external energy consumption becomes insignificant. For the results, less than 5% of external energy compared to the condenser heat amount are considered to be insignificant.

84 5.2 Determining variables and limiting cases

5.2 Determining variables and limiting cases

The transferred energies during the start-up and the start-up time (idle state to targeted operation point) are the basic parameters for the evaluation of the start-up procedure. Figure 5.2 shows the heat pump cycle with the transferred energies in orange text.

Figure 5.2 Transferred energies during start-up procedure in orange text and determining variables

in green marks

The evaporator heat source for supplying the evaporation heat is not important

because it is usually free of cost (e.g. ambient air, waste heat). The external energy supply to reach the minSH is of great interest. The amount of this energy depends on the extent of the overhanging behavior and has to be accounted for the start-up. The electric energy for the compressor is required for the start-up independently of the overhanging behavior.

Compressor

2

Expansionvalve

3

4

1

5 6

IHX

Condenser

Eexternal

Tinlet, heat sink

Evaporator

Tinlet, heat sink

Eel

Qcond

Qevap


The condensation heat during the start-up can supply heat to the heat sink when the temperatures are suitable. However, this depends strongly on the characteristic of the heat sink. The heat sink could be a thermal energy storage, which discharged completely and has the temperature of the ambient air. Consequently, the heat sink inlet temperature of the condenser rises slowly. In contrast to that, the heat sink could be a phase change material storage. It requires heat at a temperature slightly below the targeted operation point right after start-up of the heat pump. Here, the heat sink inlet temperature of the condenser rises quickly. The possible different heat sink characteristics are accounted in the dynamic simulation by the use of a variety of heat sink temperature ramps.

Finally, the heat sink inlet temperature and the external energy to reach the minSH are the main determining variables. Figure 5.2 shows the two variables in green marks.

Two limiting cases arise with these two variables. With a very low temperature ramp of the heat sink inlet, the condensation pressure keeps low. The IHX has time to heat up and delivers the full heat for minSH when the limiting temperature lift (LTL) is reached. Thus, less external energy is required and the time to a targeted operation point is longer. If the targeted operation point with its condensation pressure is reached immediately, more external energy is required. In this case, the heat for minSH is provided immediately by the external energy source without support of the IHX at the very start. Figure 5.3 shows a scheme of four steps of each limiting case with the start-up from idle state to a targeted operation point.

86 5.3 Simulation methodology

Figure 5.3 Time dependent start-up progression for two limiting cases

The aim of the dynamic simulation is to ascertain the dependency of the two

variables for the start-up procedure.

5.3 Simulation methodology Software and thermodynamic working fluid data

The software ChemCAD with the add-on Dynamics of the manufacturer Chemstations serves for the dynamic calculations of this chapter. All dynamic simulations conduct with LG6 as exemplary working fluid. LG6 is not included in the ChemCAD chemical database. Therefore, the thermodynamic data of LG6 are edited for the supported neutral file import of ChemCAD.

Limiting case 1:• Slow• Less Qexternal

Enthalpy

Pres

sure

Idle state

Intendedoperation point

Tim

eLimiting case 2:• Fast• High Qexternal

Qexternal

QIHX


Flowsheet coding The cycle of a HTHP is programmed into ChemCAD (see Figure 5.4).

ChemCAD uses operation units and streams to build up a flowsheet. The streams are numbered in squares and the units are parallel numbered in circles. The blue marked streams (1 to 6) are the major state points with same numbers according to Figure 5.1. The following units represent the heat pump components:

Evaporator Unit 7 Compressor Unit 8 Condenser Unit 1 Expansion Unit 12 IHX Unit 10/14

The rest of the flowsheet is required for controlling the units and streams. Table 5.1 summarizes the programmed functions of the units.

Figure 5.4 Flowsheet of ChemCAD for dynamic simulations

The dynamic simulation shall preferably reproduce a real HTHP. Therefore,

parameters of the functional model (see section 6.1) are adopted. The mass and material of heat exchangers serve to determine the thermal inertia. Moreover, the displacement volume of the compressor is used.

88 5.3 Simulation methodology Table 5.1 Unit functions of ChemCAD flowsheet for dynamic simulation

Unit Function 1 Condenses and subcools (5 K) stream 16 to stream 3

2 Feed-forward: Sets pressure out at unit 8 equal to stream 14

3 No flowsheet function (checking for operator)

4 Feed-forward: Sets temperature of stream 14 (at saturated vapor state) equal to stream 8 plus 10 K

5 Feed-backward: Controls total mass rate of stream 20 to adjust total actual volume rate of stream 1 equal to 12.142 m3 h-1 (displacement volume of real compressor of section 6.1.2 with 80% volumetric efficiency)

6 Feed-backward: Controls heat duty of unit 10 to apply minSH at stream 21 by fitting function: T21 – 0.5452*T14 = 0.4628*T6; T must be specified in degree Rankine

7 Evaporates and superheats (5 K) stream 26 to stream 6

8 Specifies isentropic efficiency = 0.8

9 No flowsheet function (checking for operator)

10 Superheats stream 6 to stream 21

11 Simulates thermal inertia of condenser of section 6.1.2 (Steel AISI316; m = 3.8 kg; cp = 0.50 kJ kg-1 K-1) with water stream of same thermal inertia

12 Expands isenthalpic stream 23 to stream 5

13 Feed-backward: Controls heat duty of unit 14 to apply possible subcool depending on operation point at stream 4 by fitting function: T4 – 0.2239*T3 = 0.7871*T6; T must be specified in degree Rankine

14 Subcools stream 30 to stream 4

15 Feed-forward: Sets temperature of stream 17 equal to stream 18

16 Feed-forward: Sets pressure out at unit 12 equal to stream 20

17 Adjusts temperature ramp of stream 20

18 Adjusts temperature ramp of stream 8

19 Simulates thermal inertia of IHX of section 6.1.2 (Steel AISI316; m = 4.6 kg; cp = 0.50 kJ kg-1 K-1) with water stream of same thermal inertia

20 Feed-forward: Sets temperature of stream 29 equal to stream 31


Simulation parameters

Each dynamic simulation starts from idle state with a temperature of all units and streams of 20 °C. The targeted operation point is: Tevap = 80 °C; Tcond = 130 °C.

The heat source for the evaporator is the same for all simulations. It has a constant temperature ramp of 6 K min-1 and heats up the evaporator from the start at 20 to 80 °C in 10 min.

Eleven dynamic simulations are conducted with different temperature ramps of the heat sink inlet to account for the different possible heat sink characteristics (see section 5.2). The temperature ramps are 100, 67, 50, 33, 25, 20, 17, 14, 13, 11 and 10 K min-1. These ramps correspond to time periods of 1 to 10 min to the targeted condensation temperature.

Constant temperature ramps do not occur in reality. However, these constant ramps serve as simplified assumption due to implementation characteristics in ChemCAD.

Instead of the heat sink, the thermal inertia of the condenser might (especially right after the start) condense and subcool the working fluid. As a result, the working fluid can show a lower temperature than the heat sink. In this case, the heat transfer would be vice versa from the heat sink to the working fluid. Simulation points with this phenomenon are considered. The transferred heat from the heat sink to the working fluid counts as external energy as well.

5.4 Results and discussion

Figure 5.5 shows the results of the dynamic simulations. The longer the time respectively the lower the temperature ramp, the lower is the external energy demand. It declines significantly with increasing start-up times at the simulations with short start-up times (at 1, 1.5 and 2 min). Here, the condensation temperature reaches quickly the target and unit 10 needs much external energy to superheat the working fluid to the minSH. The longer the time, the more heat is provided by the IHX, when the thermal inertia is overcome. The extent of the decline decreases (at 3 and 4 min) and diminishes (at 5, 6 and 7 min).

90 5.4 Results and discussion

Figure 5.5 Results of the dynamic simulation: External energy during start-up over time to targeted

condensation temperature with ratio to transferred heat in condenser for start-up time; Electric energy consumption of the compressor drive; Minimum operation time including start-up time for 5% external energy compared to condenser heat amount

After 8, 9 and 10 min, the external energy is increasing again. There is a reason

for this phenomenon. The low temperature ramp of the heat sink causes a low temperature lift for a long period of the start-up compared to low start-up times (see Figure 5.6; 1.4 and 6 min). At this period, the temperature lift is lower than the LTL and thus the IHX cannot provide the required heat for minSH, in spite of the fact, that no further heat for the thermal inertia is needed. Consequently, the external energy rises and the temperature lift during the start-up must be considered.

Tramp of heat sink inlet [K min-1]100 50 33 25 20 17 14 13 11 10

100120140160180200220

Qex

tern

al[k

J]

300350400450500550600

E el[

kJ]

10.0

12.0

14.0

16.0

18.0

0 1 2 3 4 5 6 7 8 9 10 11Min

. ope

ratio

n tim

e [m

in]

Time to Tcond = 130 °C [min]

468

101214

Qex

tern

alQ

cond

-1[%

]


Figure 5.6 Temperature ramps of condenser and evaporator for the dynamic simulations with ramps

to Tcond = 130 °C in 4 and 10 min; Temperature ramps are constant as simplified assumption; Colored areas mark the period with TLift < LTL; LTL of LG6 is 34 K (see Figure 4.9)

Finally, there is an optimum for the start-up time at 7 min. However, this is only

the optimum, where the external energy is lowest. The external energy is required due to the overhanging behavior. Additionally, the compressor requires electric energy. It decreases with increasing start-up times (see Figure 5.5). Consequently, the effect of a higher energy consumption at fast start-ups is strengthened.

If short start-up times are preferred due to conditions of a particular application (or heat sink characteristic), shorter times are possible with a higher amount of external energy. Figure 5.5 also shows the external energy in relation to the condenser heat of the start-up. Up to 13% (at 1 min) compared to the condenser heat are consumed for the external energy of the start-up. If the heat pump turns off right after the start-up, these 13% significantly reduce the overall performance.

A longer operation time after the start-up can compensate the higher consumption. If a relation of e.g. 5% shall be reached to minimize the negative influence, a minimum operation time results (see Figure 5.5). The heat pump can start very fast in 1 min to the targeted condensation temperature but needs to run for about 16 min to reduce a negative effect on the overall performance. The optimum start-up time with the lowest external energy at 7 min, needs to run only 11 min for the same effect on the performance.

2030405060708090

100110120130140

0 1 2 3 4 5 6 7 8 9 10

T ram

p[°

C]

Time [min]

TLift < LTL

1.4 min

6.0 min

34 K

34 K


6. Experimental evaluation of selected working fluids

The working fluids LG6 and MF2 are the selected ones from the theoretical evaluation in chapter 3. They both show a low GWP, thus the direct emissions (see section 2.3.3) are very low. It would be beneficial for the future dissemination of high temperature heat pumps (HTHP), if such working fluids also show low indirect emissions due to a good performance. Therefore, it is necessary to investigate the experimental performance.

There are no experimental investigations of these two fluids known from literature. The developed operation strategies (see chapter 4) are applied here due to the overhanging behavior of these fluids. The application of these new strategies in interaction with the working fluids needs experimental verification. Moreover, the low VHC of LG6 is a challenge for a good performance.

A novel functional model of a HTHP system is developed for the experimental investigation of this thesis. The planning, design, construction and setting-up operation constitute the main effort of this thesis. Goals and structure of the experimental evaluation

The first main goal is to prove the feasibility of overhanging working fluids in the newly developed functional model. The second main goal is to determine the experimental performance (COPexperimental) of the working fluids LG6 and MF2 at condensation temperatures higher than 120 °C.

Section 6.1 presents the principle design of the functional model including all components and the data acquisition system. Section 6.2 describes the general operation characteristics and the experimental procedure of the different investigations:

Influence of the compressor speed and condenser heat load Experimental verification of the LTL Performance optimization at high temperature lifts Measuring of the experimental COP

Section 6.3 shows the results and discussion of the different experimental

investigations.


Requirements on the functional model The main challenge of the functional model is the requirement to operate at

high condensation temperatures (>120 °C). This is beyond the limits of standard equipment of heat pump technology. Especially the compression end temperature can rise up to 170 °C at these condensation temperatures. The compressor and all other components need to handle these temperatures.

The size of the functional model bases on the requirement that a heat load at the condenser shall reach a maximum of 10 kW. The design of all components complies with this heat load maximum.

The evaporation and condensation temperatures (and corresponding pressures) must be adjustable to measure a variety of different operation points. In addition, the heat load must be adjustable as well.

The investigation of the optimization possibilities demands to actively shift the transferred heat loads between the heat exchangers (see section 4.4). Therefore, a new system is developed and integrated in the functional model. Details about this system are in section 6.2.5.

6.1 Setup of the functional model 6.1.1 Principal design for overhanging working fluids

The main heat pump cycle consists of a cycle with an integrated internal heat exchanger (IHX) (see Figure 4.7 and Figure 6.1). The IHX is necessary because of the overhanging behavior of the working fluids LG6 and MF2. Additionally, an electric trace heater is integrated between the IHX vapor outlet and the compressor inlet. This trace heater works according to the control mechanism #1 (see Figure 4.13).

94 6.1 Setup of the functional model

Figure 6.1 Piping and instrumentation diagram of the functional model; Main heat pump cycle is

highlighted blue

Ret

urn

line


Figure 6.2 Photograph of the functional model during construction (not insulated) and 3D sketch

6.1.2 Component selection Compressor The compressor is the heart and driving force of the heat pump. The majority of heat pump compressors is lubricated. These lubrication systems are tailored to particular working fluids. Each established working fluid has its own lubricants. Manufacturers guarantee a stable operation of their compressors only for certain combinations of lubricants and working fluids. Therefore, the introduction of a new working fluid to an existing compressor series requires extensive testing and development. Especially the miscibility of the lubricant in the working fluid is crucial. Entrained lubricant in the compressor outlet needs to return to the compressor. Although lubricant separators with separation efficiencies of larger than 99% exist, the lubricant would accumulate in the cycle over a long period without returning. That is why miscibility is favorable. These requirements diminish when using a lubricant-free compressor. This would ease the introduction of new working fluids significantly. However, there are

Compressor

Condenser

Lubricantseparator

IHX

Expansion valveEvaporator

96 6.1 Setup of the functional model

no compressors commercially available for this power range, which meet the demands of the desired high temperatures in the functional model. Consequently, a lubricated compressor must be used. It is the model F3 of the manufacturer GEA Bock with a Siemens drive (see Figure 6.3). The drive directly couples to the compressor shaft and has a maximum power of 5.5 kW. The F3 is a single stage piston compressor with two cylinders and a displacement volume of 0.23 liter.

Figure 6.3 Compressor GEA Bock F3 with Siemens drive

The allowed compressor speed (revolution per minute, rpm) is 900 to 1800 min-1. A frequency inverter (Siemens model G120P) adjusts the compressor speed within the mentioned limits. The frequency inverter also measures the actual drive power and passes the value to the data acquisition system. A lubricant sump heater is present in the compressor for the start-up. Lubricant separator

The lubricant separator is the model 903 of the manufacturer Temprite. It is a coalescent filter, which separates the lubricant droplets from the vapor of the compressor outlet. The separation process works only with the working fluid in the vapor phase. However, overhanging working fluids condense at a cold start. Therefore, an additional heating system is wrapped around the lubricant separator including the connecting pipes.

The separated lubricant returns through a pipe directly connected to the lubricant reservoir of the compressor.


Heat exchangers

The Condenser and the IHX are plate type heat exchangers of the manufacturer SWEP. The condenser is a B15x24 and the IHX is a B26Ax24 model. They are designed with the manufacturers’ software SSP G7 [106].

The evaporator is a coaxial double pipe heat exchanger of the manufacturer Wieland (model WKE24). This type is selected to ensure full evaporation in each channel.

The maximum heat load of all heat exchangers is according to the requirement of the functional model to achieve a maximum heat load of 10 kW in the condenser. All heat exchangers are overdesigned with 30% of additional heat transfer area. Expansion valve The use of standard thermostatic or electronic expansion valves is not possible because they are available only for established working fluids. An automatic expansion valve is integrated as expansion unit. It is the model AEL6.0 of the manufacturer Honeywell. The advantage is that the evaporator pressure can be adjusted independent of the evaporator superheat. This is necessary for the investigation of optimization possibilities (see section 6.2.5) 6.1.3 Data acquisition system

All measuring points of the functional model (see Figure 6.1) and the control points centrally process by a data acquisition system. A cDAQ system together with the software LabView of the manufacturer National Instruments serves.

A code is programmed (see Appendix B), which enables LabView to connect to Refprop and calculate all needed data in real time. In doing so, it is possible to display a log p,h diagram of the functional model operating cycle in real time (see Appendix A) with an updating time of 1 second. All measured and calculated values export every 10 seconds in a log file for further analysis. Several sensors serve for the measuring points:

Pressure (measuring points 1, 2, 3, 4, 5, 6, 7): Siemens, pressure transmitter Sitrans P200 Accuracy of measurement: 0.25% of measuring range limit (±0.04 bar)

98 6.2 Experimental procedure

Temperature (measuring points 1, 2, 3, 4, 5, 6, 7, 8a, 9a, 10, 11): Titec, thermocouple type J Accuracy of measurement: Tolerance class one: ±1.5 K

Temperature (measuring points 1a, 2a, 8, 9): Jumo, thermocouple type K Accuracy of measurement: Tolerance class one: ±1.5 K

Flow, volumetric (measuring points 5, 8a, 10): PKP, turbine flow meter DR08-15 Accuracy of measurement: 1% of measuring range limit (±0.2 l min-1)

Flow, volumetric (measuring point 8): Siemens, magnetic flow meter Sitrans F M MAG 1100 Accuracy of measurement: 0.2% of measured value (±0.02 l min-1)

The experimental COP is the most important measured value. Therefore, the condenser heat load is measured in three different ways in parallel to minimize measurement errors. Two times at the water side (8-9 and 8a-9a) and one time at the working fluid side (3-4 with flow at 5) of the condenser. Together with the electric power of the drive, this provides to two outer COPs (water side) and one inner COP (fluid side). In the case of an ideal heat transfer from the working fluid to the water, the inner and outer COPs should be the same. A deviation of 8% due to heat losses and measuring uncertainty is tolerated. Only measured operation points below that limit include in this thesis.

6.2 Experimental procedure 6.2.1 Operation characteristics of the functional model Adjustment of operation points

Except for the start-up, the aim of the experimental procedure is to adjust different operation points with precise and stable state points. Several controlled variables (see Table 6.1) yield the operation points:


Table 6.1 Controlled variables of the functional model with corresponding manipulated variables; Numbers according to state points in Figure 6.1

Controlled variable Manipulated variable Temperature (10) Superheat (7)

Automatic expansion valve opening Tevap, pevap, Superheat (7)

Compressor speed Qcond, Qevap, QIHX, Pel,comp Temperature (8) Tcond, pcond

Fluid charge Subcool

Ptrace heater Superheat of suction gas (1)

With the controlled variables from Table 6.1, all in the following described

experimental investigations are operable. Each operation point operates for at least 10 minutes after stabilizing of all values. Figure 6.4 shows a typical progression for an abridgement of monitored values from the start-up to a stable operation point. In between, several other operation points are operated shortly for checking of previous measured operation points.

Figure 6.4 Abridgement of monitored values at functional model operation from start-up to stable

operation point

The highlighted stable operation point is the one at Tevap = 40 °C and Tcond =

90 °C of the working fluid MF2. The start-up usually takes 30 minutes. After adjusting the controlled variables for a certain operation point, the functional model stabilizes within few minutes. At the highlighted operation point, the COP fluctuates

0

1

2

3

4

5

6

7

0

20

40

60

80

100

120

140

0 0.5 1 1.5 2 2.5 3

See legendT [°

C]

Time [h]

Left axis:Tcond [°C]Tevap [°C]

Right axis:Flowfluid [l min-1]COPinnerQevap [kW]Qcond,fluid [kW]Qcond,water [kW]Pel,compressor [kW]

Start-up and adjusting of operation pointsStablepoint

25 min


with a standard deviation of 2%. Figure 6.14 and Figure 6.15 include the standard deviation of all measured operation points indicated as margin.

Operable temperature range of LG6 and MF2 The fixed displace volume of the compressor together with the allowed range of the compressor speed leads to an operation limit of the working fluids. The density of the suction vapor fixes the mass flow and the enthalpies are fixed as well by a certain operation point. Consequently, the heat load of the heat exchangers is fixed. The density (and mass flow) rises with the evaporation temperature and pressure. Thus, there is an upper limit of the evaporation temperature for each fluid, where the maximum heat load of the condenser (10 kW) is exceeded (see Figure 6.5). The limits on the left and right hand side of the operable zone arise from the targeted temperature lifts, which shall be between 30 to 60 K. Evaporation temperatures below 40 °C are not intended for HTHPs. The upper condensation temperature is limited by the compressor lubricants’ thermal stability limit. The operation point Tevap = 70 °C; Tcond = 100 °C cannot be operated with the working fluid MF2 due to the maximum condenser heat load (10 kW). Here the mass stream increases compared to the operation point Tevap = 70 °C; Tcond = 110 °C due to the lower pressure ratio in a way that the maximum condenser heat load is exceeded.

Figure 6.5 Operable temperature range of the working fluids LG6 and MF2 in the functional model

30

40

50

60

70

80

90

100

110

60 70 80 90 100 110 120 130 140 150

T eva

p[°

C]

Tcond [°C]

MF2

LG6

Tevap < 40 °C

Qcond > 10 kW


Start-up procedure The overhanging behavior of the working fluids LG6 and MF2 demands a cautious start-up, which avoids liquid compression. Like in chapter 5, it is possible to use the trace heater and a certain temperature ramp for the heat sink. The necessity of energy saving during the start-up is not present in the laboratory facility of the functional model. Therefore, the trace heater, the lubricant separator heater and the lubricant sump heater turn on for at least 30 minutes before the start of the compressor. Additionally, the evaporator and the condenser heat up with the water cycles before the start of the compressor as well. When the compressor starts, the trace heater adjusts to ensure that the state point 1 always fulfills the requirement of the minimum required superheat (minSH). If the superheat at state point 2 decreases below 5 K distance to the saturated vapor line, the temperature ramp of the condenser heat sink is lowered. 6.2.2 Characterization of determining factors

The experimental investigations compare different operation points to each other. Therefore, it is necessary to neglect determining factors, which would influence parameters, for example the COP, only at some operation points. Compressor speed Electric drives and compressors show different efficiencies depending on the speed. Consequently, this influence has to be investigated and later be considered in the discussion. An experimental series with LG6 is conducted at operation points of Tevap = 70 °C; Tcond = 120 °C; Superheat of 5 - 10 K; Subcool of 5 - 10 K with different compressor speeds from 900 - 1400 min-1. The increment of the rpm is 100 min-1 leading to six operation points.

102 6.2 Experimental procedure Table 6.2 Operation points of experimental series for determining factor: Compressor speed

Operation point Tevap [°C]

Tcond [°C]

Superheat [K]

Subcool [K]

rpm [min-1]

1 70 120 5 - 10 5 - 10 900 2 70 120 5 - 10 5 - 10 1000 3 70 120 5 - 10 5 - 10 1100 4 70 120 5 - 10 5 - 10 1200 5 70 120 5 - 10 5 - 10 1300 6 70 120 5 - 10 5 - 10 1400

Condenser heat load

The targeted maximum condenser heat load is 10 kW. Depending on the operation point (mainly the evaporation temperature), different heat loads (2 - 10 kW) are required. At operation points with high heat loads, the area of the condenser might limit the heat transfer and influence the COP negatively compared to low heat loads. Therefore, the condenser area is overrated by 30% and thus a negative influence is unlikely.

Nevertheless, an experimental series is conducted with LG6 to ensure that there is no influence of the condenser heat load respectively the transfer area. A constant condenser heat load of 6.75 kW adjusts by the compressor speed at three operation points: Tcond = 120, 130 and 140 °C; TLift = 50 K; Superheat of 5 K; Subcool of 7 K. Table 6.3 Operation points of experimental series for determining factor: Condenser heat load


Tcond [°C]

Superheat [K]

Subcool [K]

rpm [min-1]

1 70 120 5 7 900 2 80 130 5 7 1200 3 90 140 5 7 1500

Heat losses Naturally, heat losses increase at operation points with higher temperatures, due to the increased temperature difference to the ambient. Therefore, the heat exchangers and all pipes are insulated thermally.


However, it is not possible to insulate the compressor. The operation temperature limit of the compressor requires to cool the compressor head. Especially, at operation points with high condensation temperatures, the compressor head exceeds the temperature limit. Insulation would damage the compressor. Consequently, the changing influence of the heat losses at the compressor for different condensation temperatures cannot be prevented. 6.2.3 Investigation of the theoretically predicted limiting temperature lift

The limiting temperature lift (LTL) is theoretically predicted in section 4.2.3. An experimental series is conducted to confirm the LTL with LG6. Operation points with Tevap = 40 - 90 °C; Tcond = 70 - 140 °C; Superheat of 5 - 10 K; Subcool of 5 - 10 K; rpm = 1100 min-1 are measured. The increment of the evaporation and condensation temperature is 10 K, this leads to 23 operation points.

In order to prevent damage to the compressor each operation point is at first operated with the trace heater. The trace heater turns off by hand when a stable temperature at the compressor outlet is reached. If the temperature at the compressor outlet is then decreasing below a temperature difference of 5 K to the saturated vapor line, the trace heater turns on again. If the temperature difference is stable (for at least 10 minutes) at a value higher than or equal to 5 K the trace heater is kept off.

104 6.2 Experimental procedure Table 6.4 Operation points of experimental series for investigation of theoretically predicted

temperature lift and determination of the experimental COP


Tcond [°C]

Superheat [K]

Subcool [K]

rpm [min-1]

1 40 70 5 - 10 5 - 10 1100 2 40 80 5 - 10 5 - 10 1100 3 40 90 5 - 10 5 - 10 1100 4 40 100 5 - 10 5 - 10 1100 5 50 80 5 - 10 5 - 10 1100 6 50 90 5 - 10 5 - 10 1100 7 50 100 5 - 10 5 - 10 1100 8 50 110 5 - 10 5 - 10 1100 9 60 90 5 - 10 5 - 10 1100

10 60 100 5 - 10 5 - 10 1100 11 60 110 5 - 10 5 - 10 1100 12 60 120 5 - 10 5 - 10 1100 13 70 100 5 - 10 5 - 10 1100 14 70 110 5 - 10 5 - 10 1100 15 70 120 5 - 10 5 - 10 1100 16 70 130 5 - 10 5 - 10 1100 17 80 110 5 - 10 5 - 10 1100 18 80 120 5 - 10 5 - 10 1100 19 80 130 5 - 10 5 - 10 1100 20 80 140 5 - 10 5 - 10 1100 21 90 120 5 - 10 5 - 10 1100 22 90 130 5 - 10 5 - 10 1100 23 90 140 5 - 10 5 - 10 1100

6.2.4 Determination of the experimental coefficient of performance

The experimental COP is determined with LG6 and MF2. Both working fluids are operated in experimental series at all possible operation points (see Figure 6.5). With an increment of 10 K of the evaporation and condensation temperature (intersection points of gridlines) MF2 has 15 and LG6 has 23 operation points. All these operation points are operated with rpm = 1100 min-1; Superheat of 5 - 10 K; Subcool of 5 - 10 K. Table 6.4 lists the operation points for this series for LG6. For


MF2, these are the same operation points except the operation points 13 and 17-23 from Table 6.4 (compare Figure 6.5)

Mean values of the measured COP values (10 second time step, see section 6.1.3) for the measuring period are taken for analysis and the results in section 6.3.4. 6.2.5 Optimization of the coefficient of performance by reduction of superheat

Section 4.4 describes optimization possibilities at operation points with temperature lifts higher than the LTL as antidromic optimization issue. This issue requires experimental investigation.

Furthermore, section 4.4 describes two methods, which are able to reduce the superheat (see Figure 4.17 and Figure 4.18). Experimental series are conducted for each method. Both series operate at a high temperature lift of 60 K at Tevap = 70 °C; Tcond = 130 °C; rpm = 1100 min-1.

Method 1: Increase of subcool The aim is to increase the subcool at state point 3 (respectively 4 in Figure 6.1), so that less heat is transferable in the IHX (see Figure 4.17). Usually, an additional heat exchanger after the condenser serves to obtain a high subcool. However, with such a system, the subcool is fixed and not actively capable of being influenced. For this thesis, a newly developed system serves to actively adjust the subcool. The working fluid charge of the functional model cycle can be varied during the operation with this system. A hydraulic cylinder is connected to the cycle before and after the automatic expansion valve at the high pressure and low pressure side (see Figure 6.6).


Figure 6.6 Scheme of system for operation-variable working fluid charge

The shaft height adjusts the volume of the fluid in the container. If the volume decreases, the fluid charge of the functional model cycle increases and vice versa. Thus, it is possible to adjust the fluid charge steplessly variable. Surplus working fluid usually accumulates in a receiver or the pipe above the expansion valve (see Figure 6.7 A, only pipe is shown). However, there is no receiver integrated in the functional model. Consequently, surplus liquid working fluid accumulates at the bottom of the condenser (see Figure 6.7 B). The higher the height of the liquid level in the condenser, the higher is the area for subcool and thus the subcool of the working fluid exiting the condenser. Hence, it is possible to adjust the subcool with this system by the working fluid charge. The experimental series of method 1 is conducted at three different operation points with subcool of 5, 10 and 15 K. Table 6.5 Operation points of experimental series for optimization possibility method 1: Increase

of subcool


Tcond [°C]

Superheat [K]

Subcool [K]

rpm [min-1]

1 70 130 5 5 1100 2 70 130 5 10 1100 3 70 130 5 15 1100

Flui

d co

ntai

ner

(hyd

raul

ic c

ylin

der)


Figure 6.7 Scheme of a one-plate heat exchanger as condenser with A: Only condensing working

fluid and B: Condensing and subcooling working fluid

Method 2: Partly shifted evaporation in IHX The idea of method 2 is to shift a part of the evaporation into the IHX. The heat of the subcool partly serves for evaporation and not only for the superheat. Thus, the working fluid superheats less. In order to shift a part of the evaporation into the IHX, the water inlet temperature of the evaporator (T-10) is decreased intentionally. An experimental series is conducted with T-10 = 84, 82, and 81 °C, which leads to a superheat at the evaporator outlet of 10, 5 and 0 K (at 0 K with a vapor fraction < 1). Table 6.6 Operation points of experimental series for optimization possibility method 2: Partly

shifted evaporation in IHX


Tcond [°C]

Superheat [K]

Subcool [K]

rpm [min-1]

1 70 130 10 5 1100 2 70 130 5 5 1100 3 70 130 0 5 1100


6.3 Results and discussion

The whole concept of the functional model HTHP proves by stable operation with both working fluids (LG6 and MF2). It is possible to operate the overhanging working fluids due to several cycle design alterations based on the operation and control strategies from chapter 4. The start-up and a large variety of different operation points are feasible. Condensation temperatures up to 140 °C are achieved, which is far above state-of-the-art.

This section starts with the relevance of the determining factors (section 6.3.1). It is important to characterize these influences before different operation points compared to each other. The verification of the limiting temperature lift (LTL) follows in section 6.3.2. Here, the theoretically predicted LTL is investigated experimentally to prove the operation strategies. Section 6.3.3 discusses the optimization possibilities. The most important result of the experimental evaluation is the measured COP and its discussion is in the end of this chapter (see section 6.3.4). 6.3.1 Relevance of the determining factors Compressor speed (rpm)

Figure 6.8 shows the results of the experimental series from section 6.2.2 subsection 1. The condenser heat load increases with the compressor speed. Both values increase by approximately 50% for the whole range. This is as expected, because the enthalpies of all state points are fixed for the whole series. Only the mass flow of the working fluid changes with the compressor speed and thus the heat load.


Figure 6.8 Influence of the compressor speed on operation points; measured at Tevap = 70 °C; Tcond =

120 °C; Superheat = 5 K; Subcool = 7 K; Working fluid LG6

The experimental COP decreases with the compressor speed from 4.04 (rpm =

900 min-1) to 3.43 (rpm = 1400 min-1). The decrease of the COP at higher compressor speed explains by changing

drive efficiency of the electric drive and the compressor. Moreover, the cross-section area of the pipes and the heat exchanger channels is constant. Thus, the pressure loss is higher at higher compressor speed because the velocity of the working fluid is increased due to the higher mass flow respectively volume flow. This leads to a higher pressure ratio for the compressor and thus to a lower volumetric efficiency and a higher compressor power.

This influence has to be neglected by comparing different operation points only at the same compressor speed. Consequently, all COP measurements (see 6.3.4) operate at a constant compressor speed of 1100 min-1. Condenser heat load

Figure 6.9 shows the results of the experimental series from section 6.2.2 subsection 2. It is possible to adjust the condenser heat load at 6.75 kW (fluctuating from 6.55 to 6.90 kW) at all three operation points. Therefore, the compressor speed is set to 917, 1187 and 1493 min-1 for the condensation temperatures 140, 130 and 120 °C.

4.0

4.5

5.0

5.5

6.0

6.5

2.5

3.0

3.5

4.0

4.5

5.0

800 900 1000 1100 1200 1300 1400 1500

Qco

nd[k

W]

COP e

xper

imen

tal

rpm [min-1]


Figure 6.9 Influence of the condenser heat load on operation points; measured at Tcond = 120, 130,

140 °C; TLift = 50 K, Superheat = 5 K, Subcool = 7 K; Working fluid LG6

The condenser heat load is constant. Nevertheless, the COP is decreasing with

the same compressor speed range like in Figure 6.8. This verifies that the condenser heat transfer area is not limiting the COP. There could be an influence of the different evaporation temperatures (mainly due to varying VHC respectively ), but this is not the case (see Figure 6.14). The COP at these operation points in Figure 6.14 varies only between 3.72 and 3.80. Consequently, the main influence on the COP for this experimental series is caused by the compressor speed. 6.3.2 Verification of the limiting temperature lift

Figure 6.10 shows the results of the experimental series from section 6.2.3 and the theoretically predicted LTL from section 4.2.3. The grey points indicate operation points at which the trace heater is required. At these operation points, the distance of the compression end point to the saturated vapor line fell below 5 K after turning off the trace heater and the trace heater was turned on again. Black points mark operation points at which the distance is higher than 5 K without the trace heater (only with the IHX).

8008108208308408508608708808909009109209309409509609709809901000101010201030104010501060107010801090110011101120113011401150116011701180119012001210122012301240125012601270128012901300131013201330134013501360137013801390140014101420143014401450146014701480149015001510152015301540155015601570158015901600

0

2

4

6

8

10

2.5

3.0

3.5

4.0

4.5

5.0

800 1000 1200 1400 1600

Tcond [°C]

Qco

nd[k

W]

CO

P exp

erim

enta

l

rpm [min-1]

140 130 120


Figure 6.10 Verification of the LTL; Theoretically predicted LTL from section 4.2.3; Points

measured at Tevap = 40 – 90 °C; Tcond = 70 – 140 °C; Tincrement = 10 K; Superheat = 5 - 10 K; Subcool = 5 - 10 K; rpm = 1100 min-1; Working fluid LG6

The theoretical LTL situates between the grey (lower TLift) and the black

operation points (higher TLift). This verifies that the calculated transferable heat amount over the IHX (see section 4.2.2) is also valid for real (experimental) cases. Consequently, the operation strategies and suggested cycle designs (depending on molar mass, see Figure 4.11) are applicable for real HTHPs. 6.3.3 Potential of the optimization possibilities

Figure 6.11 and Figure 6.13 show the results of both experimental series of section 6.2.5. The absolute values of the results are in tables in Appendix C. Method 1: Increase of subcool (see Figure 6.11)

The system for operation variable working fluid charge (see section 6.2.5) works properly. It is possible to adjust the liquid level inside the condenser and thus the subcool. The influence of the subcool on the internal heat exchanger (IHX) heat load is clear (see Figure 6.11 left hand side). It decreases by -22% from Subcool = 5 K to 15 K. Consequently, less heat transfers and the compressor inlet temperature (T1, suction gas in Figure 6.1) decreases by -7% (from 118 to 109 °C). Figure 6.11 shows the theoretical changes of the mentioned parameters on the right hand side as well. All changes show the same trend for the theoretical and experimental values.

20

30

40

50

60

70

30 40 50 60 70 80 90 100 110

T Lift

[K]

Tevap [°C]

Limiting temperature lift (LTL)

Noheater

Heaterrequired


The density and the condenser heat load are both positively changing with the result, that the COP improves by +10%.

Figure 6.11 Experimental results (left) and theoretical changes (right) of optimization possibilities

method 1: Increase of subcool

The higher condenser heat load and COP are of course positive. However, the

increased subcool causes a lower temperature at which the condenser heat load transfers. The condensation takes place at the same temperature at 130 °C but the desuperheat and the subcool are at lower temperatures (see Figure 6.12). Consequently, the improvement of method 1 is only valid for matching heat sinks. If the whole condenser heat load shall transfer to a heat sink at for example 125 °C, method 1 is not advantageous.

-30%

-25%

-20%

-15%

-10%

-5%

0%

5%

10%

-30%

-25%

-20%

-15%

-10%

-5%

0%

5%

10%

5 10 15

Dev

iatio

n

Subcool [K]

Q cond suction gas

T suction gasQ IHXCOP

-30%

-25%

-20%

-15%

-10%

-5%

0%

5%

10%

5 10 15Subcool [K]

Q cond suction gas

T suction gasQ IHX

Experimental Theoretical


Figure 6.12 Q, T diagram of the condenser heat load of operation points for optimization possibilities

method 1: Increase of subcool; Dashed lines are the heat sink (water) temperature progresses at the secondary side of the condenser

Method 2: Partly shifted evaporation in IHX (see Figure 6.13)

It is possible to decrease the superheat by decreasing the heating water inlet temperature (T10 in Figure 6.1), when the evaporator pressure is kept constant. The inlet of the IHX gas side then shows a lower temperature respectively a vapor fraction smaller than 1. Consequently, the heat load of the IHX increases (see Figure 6.13). The experimental results show this behavior and the IHX heat load increases by +20%. This is lower than the theoretical change (+29%). The reason for the difference is the theoretical assumption, that the temperature of the IHX liquid outlet (T5 in Figure 6.1) decreases to the evaporation temperature with 5 K difference (75 °C). This is not achieved in this experiment due to the IHX heat transfer characteristics. The IHX is not designed for evaporation and very low temperature differences.

The IHX liquid outlet temperature fell down to 80.7 °C instead of 75 °C (see Appendix C). For the same reason, the compressor inlet temperature (T1, suction gas) decreases less (116 °C) than theoretically predicted (108 °C) and less than with method 1 (109 °C) as well.

95100105110115120125130135140

0 1 2 3 4 5

T [°

C]

Qcond [kW]

Subcool = 5 KSubcool = 10 KSubcool = 15 KHeat sink at subcool = 5 KHeat sink at subccol = 10 KHeat sink at subcool = 15 K


Figure 6.13 Experimental results (left) and theoretical changes (right) of optimization possibilities

method 2: Partly shifted evaporation in IHX

Nevertheless, there is an improvement of the COP, which is lower (+7%) than

with method 1. However, the improvement is independent of the heat sinks’ temperature characteristic, because the subcool at the condenser is constant for this method 2. 6.3.4 Experimental coefficient of performance

Figure 6.14 and Figure 6.15 show the results of the experimental series of section 6.2.4. Besides the condensation temperature axis, all operation points are marked by the evaporation temperature and temperature lift. The margin is the standard deviation (see section 6.2.1).

-15%-10%

-5%0%5%

10%15%20%25%30%

-15%-10%-5%0%5%10%15%20%25%30%

0 5 10

Dev

iatio

n

Superheat [K]

Q IHX suction gas

T suction gasCOP

-15%-10%-5%0%5%10%15%20%25%30%

0 5 10Superheat [K]

Q IHX suction gas

T suction gas

Experimental Theoretical


Figure 6.14 Experimental COP of working fluid LG6 at all measured operation points; Margin is

standard deviation (see section 6.2.1)

Figure 6.15 Experimental COP of working fluid MF2 at all measured operation points; Margin is

standard deviation (see section 6.2.1)

The COP decreases with increasing temperature lift at same evaporation

temperatures. This is valid for all measured operation points of both working fluids.

80 °C70 °C60 °C

50 °C

40 °C

90 °C

2

3

4

5

6

7

60 70 80 90 100 110 120 130 140 150

CO

P exp

erim

enta

l

Tcond [°C]

30 K

40 K

60 K

50 K

TevapTLift

70 °C60 °C50 °C40 °C

2

3

4

5

6

7

60 70 80 90 100 110 120 130 140

CO

P exp

erim

enta

l

Tcond [°C]

30 K

40 K

60 K

50 K

Tevap

TLift


Section 3.3.3 describes the theoretical COP progression at a constant temperature lift. Figure 6.16 shows this theoretical COP progression together with the experimental results for both working fluids at a temperature lift of 50 K. The experiments show the same trend of this progression. At first, the COP is increasing to a maximum with a certain distance to the critical point, followed by a decline towards the critical point. The maximum of the theoretical and experimental COP is at the same operation points for both working fluids and situates at the theoretically predicted distance to the critical point (see Figure 3.8).

Figure 6.16 Theoretical and experimental COP at a temperature lift of 50 K of the working fluids

LG6 and MF2 with Carnot maximum

However, the distance of the experimental COP to the theoretical COP varies

significantly among the different operation points. This distance explains descriptively by comparing the experimental COP to the Carnot COP. As mentioned in section 2.3.1 commercial heat pumps reach COP values around 50% of the Carnot COP. Figure 6.17 shows the percentage of the Carnot COP of all measured experimental COP of the working fluids LG6 and MF2.

MF2 LG6

2.0

3.0

4.0

5.0

6.0

7.0

8.0

9.0

80 90 100 110 120 130 140 150

CO

P

Tcond [°C]

experimentalexperimental

MF2theoretical

theoreticalLG6

Carnot

40-90 50-100 60-110 70-120 80-130 90-140Tevap-Tcond [°C]


Figure 6.17 Experimental COP as percentage of Carnot COP of all measured operation points of the

working fluids LG6 and MF2

The majority of the operation points situates in the common range around

50%. There are 28 operation points out of 38 in the range 45% to 55%. The functional model shows a good performance in accordance to commercial products at these operation points. The losses compared to the Carnot COP are common and base on pressure drop, heat losses, isentropic, volumetric, electric and mechanic efficiency.

The ten operation points below 45% of the Carnot COP show a lower performance with the lowest at 34%. Nine of these ten operation points belong to the working fluid LG6 and show a low VHC. As mentioned in section 2.3.1, the VHC has an influence on the achievable experimental COP because it influences the compressor efficiency. Practical limits of the VHC mentioned in literature are values of 500 and 1000 kJ m-3 (see section 2.3.1). The experiments with the functional model of this thesis indicate the same range. Eight operation points show a VHC below 1000 kJ m-3 and a Carnot COP below 45%. Finally, the functional model indicates a minimum VHC of 1500 kJ m-3 is required to surely avoid a negative influence.

The negative influence mainly determines by the compressor characteristics. The question is how efficient a compressor can process a certain working fluid stream to the targeted pressure. Figure 6.18 shows the required electric power per processed mass stream in dependency of the VHC for the working fluid LG6.

20%25%30%35%40%45%50%55%60%

0 500 1000 1500 2000 2500 3000

CO

P Car

not

VHC [kJ m-3]

MF2LG6

Low distance tothe critical pointLow VHC

CO

P exp

erim

enta

lCO

P Car

not-1


Figure 6.18 Electric power per processed mass stream in dependency of the VHC for all measured

operation points of the working fluid LG6 sorted by the temperature lift

The performance of the functional models’ compressor shows a clear dependency on the VHC. The required electric power for processing the same mass stream increases significantly with decreasing VHC. At a temperature lift of 50 K the required power increases by a factor of 1.8 of the measured VHC range. This explains the low Carnot COP values of the operation points at low VHC in Figure 6.17.

The deviations between the different temperature lifts in Figure 6.18 result from the volumetric efficiency. At lower temperature lifts, the pressure ratio is lower and thus the volumetric efficiency is higher.

The working fluid MF2 shows only one operation point below 45% (see Figure 6.17). This is the operation point at the highest applied condensation temperature (130 °C, see Figure 6.15). The distance to the critical point is the lowest (18 K) among all operation points. The low percentage of the Carnot COP at this operation point verifies the mentioned required distance to the critical point (see section 2.4.1 and Figure 3.8).

The two determining parameters (VHC and distance to the critical point) are

discussed separately in more detail in the next subsections. Influence of the VHC

The increase of the experimental COP is larger than the increase of the

theoretical COP (see Figure 6.16). This counts especially for the working fluid LG6.

10

15

20

25

30

35

500 1000 1500 2000 2500

P elm

-1[k

J kg-1

]

VHC [kJ/m³]

Tlift = 60 K

50 K40 K

30 K


The experimental COP increases by +39% at a temperature lift of 50 K, but the theoretical COP only by +5%. The increase for MF2 is +5% respectively +1%. The reason for this different behavior is that the described theoretical progression (see section 3.3.3) is without consideration of the VHC. However, LG6 shows a low VHC at low evaporation temperatures (see Figure 6.19) due to the low density, which leads to low efficiencies.

Figure 6.19 VHC and distance to the critical point of measured operation points at a temperature lift

of 50 K for the working fluids LG6 and MF2

At an evaporation temperature of 40 °C, LG6 shows a VHC below the

practical limit (see section 2.3.1) of 1000 kJ m-3. Here the Carnot COP is only 38% at a temperature lift of 50 K. At higher evaporation temperatures, the Carnot COP rises to 47%, which is common for heat pumps (see section 2.3.1). Here the VHC is higher than 1500 kJ m-3.

In contrast to that, MF2 shows a VHC higher than 1500 kJ m-3 already at evaporation temperatures of 40 °C and the Carnot COP is higher than 50%. That is why, there is no large gap in the increase of the theoretical and the experimental COP (+1% respectively +5%).

Consequently, the VHC has a strong influence on the experimental COP. Figure 6.16 shows the theoretical and experimental COP at a temperature lift of 50 K. Despite the lower theoretical COP of MF2 compared to LG6, the experimental COP is higher than the COP of LG6. This is valid only for VHC values lower than 1500 kJ m-3. Thus, the limit of the VHC for achieving experimental COP values in

0

10

20

30

40

50

60

70

80

90

0

500

1000

1500

2000

2500

3000

3500

4000

80 90 100 110 120 130 140 150

T crit

-Tco

nd[K

]

VH

C [k

J m-3

]

Tcond [°C]

MF2

LG6


the range of 50% of the Carnot COP is around 1500 kJ m-3. At lower VHC values, the functional model is operable, but a lower efficiency has to be accepted.

Two operation points of LG6 are compared to emphasize this further. Usually, a higher temperature lift leads to a lower COP. This is not valid for the operation points Tevap - Tcond = 40 - 70 °C (Tlift = 30 K) and Tevap - Tcond = 80 - 130 °C (Tlift = 50 K) in Figure 6.14. Despite of the higher temperature lift, the COP is higher for the operation point 80 - 130 °C due to the very low VHC at Tevap = 40 °C Influence of the distance to the critical point

Section 3.3.3 shows, that the distance of the condensation temperature of an operation point to the critical temperature is influencing the COP. The COP decreases with decreasing distance in direction of higher condensation temperature. The results of the experiments verify this behavior (see Figure 6.16). Figure 3.8 shows the theoretical COP as a function over the distance. The maximum of the theoretical COP of MF2 is at a distance of 41 K and LG6 at 36 K. These distances are in the same range for the experimental results (see Figure 6.16). MF2 shows a maximum at a condensation temperature of 100 °C (distance = 45 K) and LG6 at 130 °C (distance = 39 K). The resolution of the experiments (Tincrement = 10 K) is, of course, not enough for an exact match, but the range of the distance for the maximum COP is confirmed. Conclusion on experimental coefficient of performance

Finally, the experimental series for the determination of the experimental COP show promising performances of the working fluids MF2 and LG6. This functional model is the first one worldwide, which is build to function with these working fluids and the performance is in the same range like commercial products. The two working fluids complement one another. The aim of the choice is to use a working fluid with a VHC higher than 1500 kJ m-3. At values higher than this border, the negative influence on the COP is avoided and the compressor shows a higher efficiency. MF2 is more efficient at lower temperatures (Tcond < 120) and LG6 is recommended for usage at higher temperatures (Tcond > 120 °C). Thus, for each desired temperature range a matching working fluid is established by this thesis.

The remarkable achievement of these experiments are the very high temperatures (140 °C) with a good efficiency (COP = 3.7) at a high temperature lift (50 K). Moreover, only future-proof working fluids (non-toxic, non-flammable, ODP = 0, GWP = 1) are used. This is achieved worldwide for the first time in this thesis.


7. Summary and outlook

Summary This thesis aims to develop a novel high temperature heat pump (HTHP) system

using environmentally friendly and safe working fluids. In the first step of this thesis, a large variety of different working fluids is evaluated. It is claimed, that all worldwide commercially available and suitable working fluids for the high temperature range are included in the evaluation. The result of the evaluation is that only two working fluids (LG6, MF2) fulfill all evaluation criteria. They are selected for the experimental investigation and both working fluids are for the first time presented for HTHP.

There is a trend towards complex molecules in working fluid usage. For example, the global warming potential lowers by the insertion of particular functional groups into the molecules, leading to a high molar mass. This trend is proven by the results of the evaluation. Both selected working fluids show a distinct overhanging behavior due to the high molar mass. This involves that the saturated vapor line shows a low positive slope. As a consequence, this kind of working fluids are not operable in a standard heat pump cycle.

Therefore, new operation strategies for overhanging working fluids are developed. New terms like the “limiting temperature lift” are defined to characterize the working fluids and their boundary conditions for operation. Several principles are ascertained to classify the working fluids by the new terms. In doing so, it is possible to conclude to the adequate cycle design of each particular working fluid only by knowing the molar mass.

An additional question is the start-up procedure of a heat pump with overhanging working fluids. Especially for heat pumps with frequent start-ups, it is important to know the boundary conditions of the start-up. An additional energy consumption or a delay of the heat generation for the heat sink might negatively influence the economic benefit. Several dynamic simulations of the start-up procedure are conducted. The results show that there is an optimum start-up time with the lowest additional energy consumption. This optimum depends mainly on the heat sink characteristic of a particular application.

A novel functional model is developed, which operates with the selected working fluids. It is build based on the developed operation strategies. The whole concept of the functional model is proven by stable operation. It is possible to operate the

122

overhanging working fluids with condensation temperatures up to 140 °C, which is far above state-of-the-art.

The main result is the experimental COP. In spite of the fact, that the functional model is the first one worldwide working with these working fluids, the efficiency is in the same range like commercial heat pumps (about 50% of the Carnot COP). The two selected working fluids complement one another. MF2 shows higher COP at lower condensation temperatures (< 120 °C) and LG6 at higher temperatures (> 120 °C). At the highest condensation temperature (140 °C) and at a temperature lift of 50 K, the experimental COP is 3.7. This is achieved with an environmentally friendly (ODP = 0, GWP = 1) and safe (non-flammable, non-toxic) working fluid.

Finally, a novel high temperature heat pump system is developed for the operation of overhanging working fluids. Two new working fluids are established within this system. The environmental properties guarantee the usage in the future and the promising performance supports the prospective dissemination of HTHPs.

Outlook

This thesis lays the foundation for the operation of overhanging working fluids in

HTHPs. Further required work has to deal with larger prototypes of the functional model. The functional model operated for a total of 200 hours. Here, long-term tests are required to ensure a stable operation over years with acceptable maintenance. Long-term material incompatibility or thermal stress could influence the stability in a negative way.

The compressor type plays a major role for the performance. The targeted application range of HTHPs (Qsupply > 500 kW) is much larger than the maximum heat load of the functional model (Qsupply < 10 kW). It is most likely that other compressor types are implemented in large-scale applications. It has to be shown that the same or even a better performance is feasible in heat pumps working with, for example, turbo compressors.

Another big challenge is the industrial process integration of HTHPs. The dissemination will always depend on the locally connected occurrence of heat source and sink. If high supply temperatures (e.g. 140 °C) at the sink are required, heat sources with higher temperatures than available in ambient air or lake or sea water are necessary. Here, a detailed survey on locally available waste heat and industrial heat demand is essential.

The development of the novel HTHP system started within this thesis. The laboratory scale investigation and proof of concept is completed. This project


continues after and beyond this thesis at the Siemens company. The main issues for the continuation are the scale-up design and the industrial process integration.

124

Nomenclature Roman c Heat capacity J mol-1 K-1 COP Coefficient of performance - E Energy kJ h Enthalpy (specific) kJ kg-1

IS Inverse slope of saturated vapor line J mol-1 K-2 LTL Limiting temperature lift K M Molar mass g mol-1

minSH Minimum required superheat K P Power kW p Pressure bar Q Quantity of heat kJ Q Heat load kW rpm Revolution per minute (compressor speed) min-1

s Entropy (specific) kJ kg-1 K-1 SC Subcool K SH Superheat K T Temperature (difference) / (absolute) K / °C VHC Volumetric heating capacity kJ m-3 Greek

Efficiency - Density kg m-3

Subscript comp Compressor cond Condensation crit At critical point el Electric

Nomenclature 125

evap Evaporation F/H for Cl Chlorine substituted by fluorine and hydrogen IHX at internal heat exchanger min.2 at compression end point (state point 2) with minimum required

superheat

s Isentropic sat.liq. At saturated liquid sat.vap. At saturated vapor v Isochoric Abbreviations COP Coefficient of performance GWP Global warming potential HTHP High temperature heat pump IHX Internal heat exchanger IS Inverse slope of the saturated vapor line LTL Limiting temperature lift minSH Minimum required superheat ODP Ozone depletion potential Working fluid classes CFC Chlorofluorocarbon FC Fluorocarbon HC Hydrocarbon HCFC Hydrochlorofluorocarbon HCFO Hydrochlorofluoroolefin HFC Hydrofluorocarbon HFE Hydrofluoroether HFO Hydrofluoroolefin

126

Working fluids R12 Dichlorodifluoromethane R14 Tetrafluoromethane R22 Chlorodifluoromethane R31-10 Decafluoropropane R32 Difluoromethane R42-12 Dodecafluorobutane R114 1,2-Dichloro-1,1,2,2-tetrafluroethane R134a 1,1,1,2-Tetrafluoroethane R236ea 1,1,1,2,3,3-hexafluoropropane R236fa 1,1,1,3,3,3-hexafluoropropane R245fa 1,1,1,3,3-pentafluoropropane R365mfc 1,1,1,3,3-Pentafluorobutane R407C Mixture of R32/R125/R134a in % 23/25/52 R410A Mixture of R32/R125 in % 50/50 R600a 2-Methylpropane (Isobutane) R601 Pentane R717 Ammonia R718 Water R744 Carbon dioxide R1233zd 1-Chloro-3,3,3-trifluoro-1-propene R1234yf 2,3,3,3-Tetrafluoro-1-propene R1234ze Trans-1,3,3,3-tetrafluoro-1-propene R1336mzz Cis-1,1,1,4,4,4-hexafluoro-2-butene RE347mcc 1-Methoxyheptafluoropropane

List of tables 127

List of tables Table 2.1 Thermodynamic, environmental and safety properties of historically

momentous working fluids each representing one generation; COPtheoretical, VHC calculated at Tevap = 20 °C; Tcond = 60 °C; Subcool = 5 K; Superheat = 5 K s = 0.8; Colors are red if toxic, flammable, ODP > 0, GWP > 2500; Orange if 2500 > GWP > 150; Rest is green [45, 49] ............................................................................................ 24

Table 2.2 Exemplary working fluids of each generation with trend towards complex molecular structure and overhanging behavior ................... 28

Table 2.3 Variety of exemplary HTHPs in operation........................................ 31

Table 2.4 Research and development projects for HTHPs ................................ 32

Table 3.1 Heat pump cycle independent evaluation criteria of considered working fluids; Colors are red if toxic, flammable, ODP > 0.0005, GWP > 2500; Orange if 2500 > GWP > 150, Tcrit < 130 °C; Rest is green; Sources according to list in 3.2 .............................................. 41

Table 3.2 Heat pump cycle dependent evaluation criteria of considered working fluids at standard simulation point: Tcond = 130 °C; Tevap = 80 °C; Superheat = 5 K; s = 0.8; isenthalpic expansion; TIHX = 5 K; No IHX for R718, R717; Subcool case 1 (5 K) and 2 (35 K); R1336mzz is estimated (~) from [32] .................................................................... 55

Table 3.3 Evaluation criteria at standard simulation point: Tcond = 130 °C; Tevap = 80 °C; Superheat = 5 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; Subcool case 1 (5 K); Tmin.2-sat.vap. = 5 K; R1336mzz is estimated (~) from [32]; Colors are red if not available, orange if GWP > 150, Rest is green ............................................................................................ 60

Table 4.1 Consideration of the chlorine content for compatibility of the working fluids R114 and R1233zd with equation (4.2) ................................... 65

Table 4.2 Additional external energy consumption of the three control mechanisms; Working fluid LG6 ..................................................... 78

Table 5.1 Unit functions of ChemCAD flowsheet for dynamic simulation ....... 88

Table 6.1 Controlled variables of the functional model with corresponding manipulated variables; Numbers according to state points in Figure 6.1 ........................................................................................................ 99

Table 6.2 Operation points of experimental series for determining factor: Compressor speed .......................................................................... 102

128

Table 6.3 Operation points of experimental series for determining factor: Condenser heat load ....................................................................... 102

Table 6.4 Operation points of experimental series for investigation of theoretically predicted temperature lift and determination of the experimental COP .......................................................................... 104

Table 6.5 Operation points of experimental series for optimization possibility method 1: Increase of subcool ........................................................ 106

Table 6.6 Operation points of experimental series for optimization possibility method 2: Partly shifted evaporation in IHX .................................. 107

List of figures 129

List of figures Figure 2.1 Basic principle of a heat pump ........................................................... 8

Figure 2.2 Heat pump and chiller working principle; Dashed boxes indicate the utilized side ........................................................................................ 9

Figure 2.3 Sold number of heat pumps units in Germany from 1978 to 2012 [18] . ........................................................................................................ 10

Figure 2.4 Heat pump types with continuous process distinguished by driving force ................................................................................................ 11

Figure 2.5 Schematic diagram of a vapor compression heat pump ..................... 13

Figure 2.6 Log p,h diagram of a theoretical vapor compression heat pump cycle; Working fluid R134a; Tcond = 60 °C; Tevap = 20 °C; s = 0.8; Superheat = 5 K; Subcool = 5 K; Isenthalpic expansion .................................... 16

Figure 2.7 Schematic of the function of a thermostatic expansion valve; p1: Sensor pressure, p2: Evaporator pressure, p3: Spring pressure ........... 19

Figure 2.8 CO2 factor of gas burner and heat pump at different COPs operated with electricity mix in Germany with forecast data of the DLR until 2050 [40] ......................................................................................... 21

Figure 2.9 Development of working fluid usage in vapor compression heat pumps by generations; Adapted from [42] ................................................... 23

Figure 2.10 Examples of methane derivatives; Colors are red if ODP > 0, GWP > 2500, flammable; Orange if 150 < GWP < 2500............................... 27

Figure 2.11 HTHPs in series production of several manufacturers with model name if available; Sorted by maximum supply temperature; Supply heat load is plotted partly logarithmic; References are from top to bottom: [79, 80, 81, 82, 83, 84, 85, 86, 65] ........................................................... 31

Figure 2.12 Distribution of heat demand per industrial branch in France [63] ...... 33

Figure 2.13 Photograph and experimental results from functional models working with R245fa of research projects for high temperature heat pumps; Left side (A): Peureux et al. [89]; Right side (B): Wolf et al. [7] ...... 34

Figure 3.1 Sketches of p,h diagrams of subcritical, transcritical and supercritical heat pump cycles ( ) at the standard simulation point with subcool = 5 K (see section 3.3.1); Axes are not scaled ...................................... 43

Figure 3.2 Sketches of p,h diagrams of subcritical, transcritical and supercritical heat pump cycles ( ) at the standard simulation point with subcool = 35 K (see section 3.3.1); Axes are not scaled .................................... 44

130

Figure 3.3 Heat pump cycle types for system simulation with corresponding working fluids; Some fluids are named several times due to multiple operation points; IHX is internal heat exchanger .............................. 46

Figure 3.4 Log p,h diagram of a theoretical vapor compression heat pump cycle with IHX at subcool case 1 (cycle type 1); Working fluid R245fa; Standard simulation point Tcond = 130 °C; Tevap = 80 °C; s = 0.8; Superheat = 5 K; Subcool = 5 K; Isenthalpic expansion; TIHX = 5 K ........................................................................................................ 48

Figure 3.5 Log p,h diagram of a theoretical vapor compression heat pump cycle with IHX at subcool case 2 (cycle type 3); Working fluid R245fa; Standard simulation point Tcond = 130 °C; Tevap = 80 °C; s = 0.8; Superheat = 5 K; Subcool = 5 K; Isenthalpic expansion; TIHX = 5 K ........................................................................................................ 49

Figure 3.6 Progression of the theoretical COP over a large condensation temperature range; Exemplary working fluid R245fa; Superheat = 5 K; Subcool = 5 K; Tlift = 50 K; Tincrement = 10 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; Pressures in the transcritical and supercritical range are fixed by COP maximizing algorithm (see section 3.3.2 subsection 3), ................................................................................... 51

Figure 3.7 Theoretical COP of all considered working fluids over the temperature range Tcond = 90 - 150 °C; Superheat = 5 K; Subcool = 5 K; Tlift = 50 K; Tincrement = 10 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; No IHX for R718, R717......................................................................... 52

Figure 3.8 Progression of the theoretical COP over the distance of the condensation temperature to the critical temperature; Tcond = 90 - 150 °C; Superheat = 5 K; Subcool = 5 K; Tlift = 50 K; Tincrement = 10 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K ........................................... 53

Figure 3.9 Theoretical COP of all considered working fluids at the standard simulation point: Tcond = 130 °C; Tevap = 80 °C; Superheat = 5 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; No IHX for R718, R717; Subcool case 1: 5 K; Subcool case 2: 35 K; Tmin.2-sat.vap. = 5 K; R1336mzz is estimated from [32] ..................................................... 56

Figure 3.10 Sketch of a log p,h diagram with location of the region of subcool compared to the phase boundary for the exemplary working fluids R245fa, R1234yf and R744; Temperature range of subcool is same for all three working fluids .................................................................... 57

List of figures 131

Figure 3.11 Increase of the theoretical COP depending on degree of subcool; Tcond = 130 °C; Tevap = 80 °C; Superheat = 5 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; For exemplary working fluids R245fa, R1234yf, R744 ................................................................................. 57

Figure 3.12 VHC of all considered working fluids over the temperature range Tcond = 90-150 °C; Superheat = 5 K; Subcool = 5 K; Tlift = 50 K; Tincrement = 10 K; s = 0.8; Isenthalpic expansion; TIHX = 5 K; No IHX for R718 ........................................................................................................ 58

Figure 4.1 Different definitions of the slope of the saturated vapor line; Lai and Fischer is reference number [56] ...................................................... 63

Figure 4.2 Correlation of molar mass and isochoric heat capacity for working fluids of evaluation; cv is calculated at 80 °C saturated vapor; Trend line is created for HFOs, HFCs, HFEs, FCs, LG6, MF2 and MF3 (black dots); — Trend line; --- Parallel shifted trend lines for working fluids with chlorine content .............................................................. 64

Figure 4.3 Correlation of inverse slope, molar mass and isochoric heat capacity for working fluids of evaluation, which follow equation (4.2) and Tcrit > 130 °C; Molar mass is calculated with equation (4.2) and chlorine content is considered after Table 4.1; — Trend line .......................... 66

Figure 4.4 Bell-shaped (R134a) and overhanging (LG6) working fluid in a standard heat pump cycle shown in p,h diagrams ............................. 67

Figure 4.5 Scheme of the definition of the minimum required superheat (minSH); 2 compression process with isentropic efficiency 0.8; The diagram

is only for explanation purpose and consists of no quantitatively information ...................................................................................... 68

Figure 4.6 Correlation of the minimum required superheat (minSH) and inverse slope for working fluids of evaluation, which follow equation (4.2) and Tcrit > 130 °C; minSH is calculated for standard simulation point: Tcond = 130 °C; Tevap = 80 °C; s = 0.8; --- Limit of superheat to be provided by evaporator; — Trend line for working fluids with high minSH ............................................................................................. 69

Figure 4.7 Heat pump cycle with internal heat exchanger (IHX) and overhanging working fluid (e.g. LG6) in p,h diagram; Transferred heat by IHX from subcool (3 4) to superheat (6 1); Calculated at standard simulation point: Tcond = 130 °C; Tevap = 80 °C; s = 0.8; Subcool = 5 K; Superheat = 5 K; Isenthalpic expansion; TIHX = 5 K; T,Q diagram of IHX heat transfer ......................................................................... 71

132

Figure 4.8 Transferable heat amount through IHX at different operation points: Tevap = 30 - 90 °C; Superheat = 5 K; Subcool = 5 K; Isenthalpic expansion; TIHX = 5 K; Working fluid LG6 .................................... 72

Figure 4.9 Heat amount terms (transferable, required and external) versus the temperature lift; Calculated at standard simulation point: Tevap = 80 °C;

s = 0.8; Subcool = 5 K; Superheat = 5 K; Isenthalpic expansion; TIHX = 5 K; Working fluid LG6 ..................................................... 73

Figure 4.10 Correlation of the limiting temperature lift (LTL) and inverse slope for considered working fluids of evaluation, which follow equation (4.2) and Tcrit > 130 °C; LTL is calculated for standard simulation point: Tevap = 80 °C; s = 0.8; — Trend line................................................ 74

Figure 4.11 Graphical summary of linear correlations and their significance on cycle design; Numbers in brackets refer to equations ........................ 75

Figure 4.12 Differences in control of A: Standard heat pump cycle and B: Alternative heat pump cycle with IHX; SH: Superheat ..................... 76

Figure 4.13 Control mechanism #1: Standard evaporator control and trace heater for compressor outlet control; SH: Superheat ................................... 77

Figure 4.14 Control mechanism #2: Standard evaporator control and hot gas bypass compressor outlet control; SH: Superheat ............................. 77

Figure 4.15 Control mechanism #3: Evaporator pressure adjustment to control compressor outlet; SH: Superheat ..................................................... 78

Figure 4.16 Heat pump cycle with working fluid LG6 in p,h diagram; Calculated at: Tcond = 130 °C; Tevap = 70 °C; s = 0.8; Subcool = 5 K; Superheat = 5 K; Isenthalpic expansion; TIHX = 5 K .......................................... 79

Figure 4.17 Optimization method 1: Increase of subcool; Heat pump cycle with working fluid LG6 in p,h diagram; Calculated at: Tcond = 130 °C; Tevap = 70 °C; s = 0.8; Subcool = 16.9 K; Superheat = 5 K; Isenthalpic expansion; TIHX = 5 K .................................................................... 81

Figure 4.18 Optimization method 2: Partly shifted evaporation in IHX; Heat pump cycle with working fluid LG6 in p,h diagram; Calculated at: Tcond = 130 °C; Tevap = 70 °C; s = 0.8; Subcool = 5 K; Superheat = 0 K (evaporator outlet is in 2-phase state); Isenthalpic expansion; TIHX = 5 K ................................................................................................... 81

Figure 5.1 Heat pump cycle of overhanging working fluids for the dynamic simulation; Orange: Components with thermal inertia, which delay the heat transfer to the suction gas.......................................................... 83

List of figures 133

Figure 5.2 Transferred energies during start-up procedure in orange text and determining variables in green marks ............................................... 84

Figure 5.3 Time dependent start-up progression for two limiting cases ............. 86

Figure 5.4 Flowsheet of ChemCAD for dynamic simulations ............................ 87

Figure 5.5 Results of the dynamic simulation: External energy during start-up over time to targeted condensation temperature with ratio to transferred heat in condenser for start-up time; Electric energy consumption of the compressor drive; Minimum operation time including start-up time for 5% external energy compared to condenser heat amount ..................................................................................... 90

Figure 5.6 Temperature ramps of condenser and evaporator for the dynamic simulations with ramps to Tcond = 130 °C in 4 and 10 min; Temperature ramps are constant as simplified assumption; Colored areas mark the period with TLift < LTL; LTL of LG6 is 34 K (see Figure 4.9) ....................................................................................... 91

Figure 6.1 Piping and instrumentation diagram of the functional model; Main heat pump cycle is highlighted blue .................................................. 94

Figure 6.2 Photograph of the functional model during construction (not insulated) and 3D sketch .................................................................................. 95

Figure 6.3 Compressor GEA Bock F3 with Siemens drive ................................ 96

Figure 6.4 Abridgement of monitored values at functional model operation from start-up to stable operation point ...................................................... 99

Figure 6.5 Operable temperature range of the working fluids LG6 and MF2 in the functional model ............................................................................ 100

Figure 6.6 Scheme of system for operation-variable working fluid charge ....... 106

Figure 6.7 Scheme of a one-plate heat exchanger as condenser with A: Only condensing working fluid and B: Condensing and subcooling working fluid ............................................................................................... 107

Figure 6.8 Influence of the compressor speed on operation points; measured at Tevap = 70 °C; Tcond = 120 °C; Superheat = 5 K; Subcool = 7 K; Working fluid LG6 ........................................................................ 109

Figure 6.9 Influence of the condenser heat load on operation points; measured at Tcond = 120, 130, 140 °C; TLift = 50 K, Superheat = 5 K, Subcool = 7 K; Working fluid LG6 .................................................................... 110

Figure 6.10 Verification of the LTL; Theoretically predicted LTL from section 4.2.3; Points measured at Tevap = 40 – 90 °C; Tcond = 70 – 140 °C;

134

Tincrement = 10 K; Superheat = 5 - 10 K; Subcool = 5 - 10 K; rpm = 1100 min-1; Working fluid LG6 ...................................................... 111

Figure 6.11 Experimental results (left) and theoretical changes (right) of optimization possibilities method 1: Increase of subcool ................ 112

Figure 6.12 Q, T diagram of the condenser heat load of operation points for optimization possibilities method 1: Increase of subcool; Dashed lines are the heat sink (water) temperature progresses at the secondary side of the condenser ............................................................................. 113

Figure 6.13 Experimental results (left) and theoretical changes (right) of optimization possibilities method 2: Partly shifted evaporation in IHX ...................................................................................................... 114

Figure 6.14 Experimental COP of working fluid LG6 at all measured operation points; Margin is standard deviation (see section 6.2.1) .................. 115

Figure 6.15 Experimental COP of working fluid MF2 at all measured operation points; Margin is standard deviation (see section 6.2.1) .................. 115

Figure 6.16 Theoretical and experimental COP at a temperature lift of 50 K of the working fluids LG6 and MF2 with Carnot maximum ..................... 116

Figure 6.17 Experimental COP as percentage of Carnot COP of all measured operation points of the working fluids LG6 and MF2 ..................... 117

Figure 6.18 Electric power per processed mass stream in dependency of the VHC for all measured operation points of the working fluid LG6 sorted by the temperature lift ......................................................................... 118

Figure 6.19 VHC and distance to the critical point of measured operation points at a temperature lift of 50 K for the working fluids LG6 and MF2 ...... 119

Appendix 135

Appendix A

136

B

Appendix 137

C Detailed results with absolute values of optimization possibilities for method 1 of section 6.3.3 Subcool [K] QIHX [kW] T-1 [°C] -1 [kg m-3] Qcond [kW] COP

5.35 3.04 117.9 18.32 4.48 3.10 10.33 2.73 113.7 18.69 4.70 3.25 14.98 2.37 109.4 18.71 4.80 3.40 Detailed results with absolute values of optimization possibilities for method 2 of section 6.3.3 Superheat [K] QIHX [kW] T-1 [°C] -1 [kg m-3] T-5 [°C] COP 10.50 2.89 118.8 18.09 89.49 3.16 5.80 3.10 117.9 18.23 86.06 3.21 0 3.48 116.2 18.57 80.66 3.40 Numbers in headlines refer to number of state point: e.g. T-1 is temperature at state point 1

138

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