Heat pumps : fundamentals and applications 978-3-319-62199-9, 3319621998, 978-3-319-62198-2

Table of contents :
Front Matter ....Pages i-viii
The Fundamentals (Walter Grassi)....Pages 1-14
Types of Compression Heat Pumps and Their Main Components (Walter Grassi)....Pages 15-71
Absorption Heat Pumps (Walter Grassi)....Pages 73-88
Operating Conditions (Walter Grassi)....Pages 89-111
The Refrigerants (Walter Grassi)....Pages 113-144
The External Sources: Water and Ground (Walter Grassi)....Pages 145-155
The Hybrid and Multipurpose Systems (Walter Grassi)....Pages 157-166
Additional Thermodynamic Remarks (Walter Grassi)....Pages 167-175

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Green Energy and Technology

Walter Grassi

Heat Pumps Fundamentals and Applications

Green Energy and Technology

More information about this series at http://www.springer.com/series/8059

Walter Grassi

Heat Pumps Fundamentals and Applications

123

Walter Grassi Department of Energy, Systems, Territory and Construction Engineering University of Pisa Pisa Italy

ISSN 1865-3529 Green Energy and Technology ISBN 978-3-319-62198-2 DOI 10.1007/978-3-319-62199-9

ISSN 1865-3537

(electronic)

ISBN 978-3-319-62199-9

(eBook)

Library of Congress Control Number: 2017946625 © Springer International Publishing AG 2018 This work is subject to copyright. All rights are reserved by the Publisher, whether the whole or part of the material is concerned, specifically the rights of translation, reprinting, reuse of illustrations, recitation, broadcasting, reproduction on microfilms or in any other physical way, and transmission or information storage and retrieval, electronic adaptation, computer software, or by similar or dissimilar methodology now known or hereafter developed. The use of general descriptive names, registered names, trademarks, service marks, etc. in this publication does not imply, even in the absence of a specific statement, that such names are exempt from the relevant protective laws and regulations and therefore free for general use. The publisher, the authors and the editors are safe to assume that the advice and information in this book are believed to be true and accurate at the date of publication. Neither the publisher nor the authors or the editors give a warranty, express or implied, with respect to the material contained herein or for any errors or omissions that may have been made. The publisher remains neutral with regard to jurisdictional claims in published maps and institutional affiliations. Printed on acid-free paper This Springer imprint is published by Springer Nature The registered company is Springer International Publishing AG The registered company address is: Gewerbestrasse 11, 6330 Cham, Switzerland

Preface

Heat pumps are a quite effective means to look forward to the enhancement of energy efficiency and savings. It is a very broad subject, and therefore, it is almost impossible to include all the related features in a unique volume. Far from being exhaustive, this volume is aimed at providing a detailed overview of the main topics that any professional needs to know, before either employing such machines in his designs or evaluating their energy performances. After a general description of the world market, the thermodynamic basic principles of heat pumps are recalled, emphasizing the effects of the internal and external irreversibilities on the heat pumps’ performances. The main components are analyzed, also concerning their reciprocal interactions and those with the thermal environment they are in contact with. In fact, heat pumps are complex systems which, in turn, interact with other complex systems constituted, on the one hand, by the indoor environment (internal source) and, on the other, by the outdoor environment (external source). Some details about the most used refrigerants are then provided, together with their thermophysical data. This is done with regard to the fluids used both in the compression and in the absorption heat pumps. Hybrid systems and 2-pipe and 4-pipe multipurpose systems are discussed, which constitute a very interesting technology for running thermal plant in an optimal way. The text tries to give an organic set of information and methods. Some numerical examples are provided for each treated subject, together with links and videos to help its understanding. Besides, products existing on market are often mentioned to give the interested reader a feel for the present status of technological application. According to the long experience gained by the author, this book can be useful to engineers involved in the field of building thermal installations and to students approaching this matter in energy engineering courses. Pisa, Italy

Walter Grassi

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Contents

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1 1 3 14

2 Types of Compression Heat Pumps and Their Main Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1 Main Components of Compression Heat Pumps . . 2.2 Compressor. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.3 Expansion Valve. . . . . . . . . . . . . . . . . . . . . . . . . . 2.4 The Liquid Receiver . . . . . . . . . . . . . . . . . . . . . . . 2.5 Evaporator and Condenser . . . . . . . . . . . . . . . . . . 2.6 Economizer and Vapor Injection . . . . . . . . . . . . . 2.7 The Four Way Reversing Valve . . . . . . . . . . . . . . 2.8 Engine Driven Heat Pumps (GHP) . . . . . . . . . . . . 2.9 Carbon Dioxide Heat Pumps . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

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15 15 16 38 44 46 55 59 60 65 70

3 Absorption Heat Pumps . . . . . . . . . . . . . . . . . . . . . . . 3.1 The Operating Principle . . . . . . . . . . . . . . . . . . . . 3.2 Some Features of Water–Ammonia Mixtures . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

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73 73 77 88

4 Operating Conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.1 Full Load and Partial Load Operation, the Balance Point . . . . 4.2 Comparison Among the Different Types of Heat Pump . . . . . 4.3 Further Features of Heat Pumps Operation . . . . . . . . . . . . . . .

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89 89 104 107

1 The Fundamentals . . . . . . . 1.1 General Features . . . . . 1.2 Working Principles . . . References . . . . . . . . . . . . . .

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vii

viii

Contents

5 The Refrigerants . . . . . . . . . . . . . . . . . . . . . . . . . 5.1 Properties of Some Refrigerants . . . . . . . . . . 5.2 Lubricating Oils . . . . . . . . . . . . . . . . . . . . . . 5.3 Table and Graphs of Some Refrigerants . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

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113 117 123 125 143

6 The External Sources: Water and Ground . . . . 6.1 Ground Water . . . . . . . . . . . . . . . . . . . . . . . . 6.2 Surface Water . . . . . . . . . . . . . . . . . . . . . . . . 6.3 Ground . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.4 The Ground Thermal Response . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

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145 145 147 147 151 154

7 The Hybrid and Multipurpose Systems . . . 7.1 The Hybrid System . . . . . . . . . . . . . . . . 7.2 The Multipurpose System . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . .

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157 157 160 166

8 Additional Thermodynamic Remarks . . . . . . . . . . . . 8.1 Thermodynamic Cycle . . . . . . . . . . . . . . . . . . . . . 8.2 First and Second Principles of Thermodynamics . 8.3 Phase Change of Pure Substances . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

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167 167 168 172 175

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Chapter 1

The Fundamentals

Abstract This chapter, at first, provides a synthetic picture of the present spread of heat pumps over the world market, also quoting some of the main producers. The most common types are shortly described too. In addition the thermodynamic fundamentals of their working principle are dealt with and the effect of irreversibilities on the heat pumps performances is stressed. The contribution of the different components to these irreversibilities is shortly illustrated together with that introduced by the unavoidable temperature difference between the evolving fluid and the external sources.

1.1

General Features

Heat pumps are an effective means of energy production in several fields of modern technology. To have an idea of the present situation we can refer to [1]. It reports a European market increase of 3.5% in 2014 with respect to 2013. Even if some countries recorded a decrease of the sold units, it was largely compensated by the top 10 markets led by France, Spain and Finland. In particular France (leading country) followed by Italy and Sweden reached more than hundred thousand units sold per year, while Finland, Germany, Norway and Spain exceeded fifty thousand of annual sold units. A fast increase of using heat pumps for sanitary water production is taking place both as stand alone units (heat pump and water storage tank in the same casing) or as heat pumps with separate tanks. Air has been and is the most diffused heat source so far, while larger heat pumps are increasingly employed for industrial and commercial uses, and for district heating. Air is still used, but also geothermal and hydrothermal sources are often employed. In some cases heat is provided by waste waters. To have a short insight on the more general state of the art in the world we can refer to [2]. It reports an increase of the world heat pump market of 7.2% by volume in 2013 with about two million units, jointly due to a recovery in Europe and to a strong increase of heat pump heaters in the USA. In terms of value in 2013 there was a © Springer International Publishing AG 2018 W. Grassi, Heat Pumps, Green Energy and Technology, DOI 10.1007/978-3-319-62199-9_1

1

2

1 The Fundamentals

decrease of 6.5% with respect to 2012, mainly due to an increasing competition among providers and a sale decrease of large power units. In 2013 heat pump heaters had a large diffusion with a market growth of 26.5%. Anyway this growth mainly occurred outside Europe, where air—water heat pumps and split systems dominated the market with small capacity machines, at the expenses of monobloc systems (sale decreased by 2%). Just the opposite occurred in China where these latter’s sales grew by around 14% and the worldwide increase is about 5%. Geothermal heat pumps performed poorly in 2013. High initial investment costs and lack of appropriate political support act as the major drawbacks. Despite of this they recorded a 5% increase in China and in the USA, while decreased by 1% in Europe. On the other hand, exhaust air heat pumps with heat recovery for energy saving in buildings are an emerging technology, in particular in Scandinavia, and is expected to expand at least in Northern Europe. Furthermore it is the case to stress how CO2 heat pumps had a significant rise in commercial and residential applications due to their environmental friendly features. At present, the major segments of market [3] are located in seven main regions: North America, South America, Eastern Europe, Western Europe, Asia Pacific, Japan and Middle East and Africa. Asia Pacific market is the fastest growing, while Europe holds the largest share at present. Reference [3] also pinpoints some of the major global market players in: Viessmann Group, Danfoss Group Global, Carrier Corporation, the Glen Dimplex Group, StiebelEltron, Bosch Thermotechnik GmbH, Panasonic Corporation, Mitsubishi Electric, NIBE energy systems, Geothermal International Ltd (GI), DeLonghi-Climaveneta, Airwell Group, and Enertech Group. Heat Pumps are classified according to several characteristic features. A first one consists in the type of their thermodynamic cycle and therefore in nature of fluids they use. On the basis of this we talk about vapor compression and absorption heat pumps. The former follow a traditional inverse thermodynamic cycle and use a compressor driven either by an electric motor or an engine. The used refrigerants have to be (at least should be) environment friendly, inert, chemically stable, neither flammable nor toxic, with low freezing temperatures and compatible with lubricating oils. Absorption heat pumps do not have a mechanical compressor. They use a mixture of two fluids with a different vapor pressure. The more volatile one evaporates and, then, recombines with the less volatile. The most common mixtures are water and lithium-bromide and water and ammonia. We can further differentiate heat pumps according to the type of source they heat exchangers interact with. The final fluid to be heated or cooled is the indoor air in most of the residential uses. The fluid flowing in heating, or cooling, devices can be the refrigerant itself (generally in case of short circuits) and we speak of direct expansion systems. Otherwise water is used to this aim, exchanging heat with the refrigerant in the heat pump heat exchangers. The most common outdoor heat source is air, but also surface water (rivers, lakes and sea), ground water and even the ground itself.

1.1 General Features

3

Lastly heat pumps can be used for sanitary water production only or to both heating and sanitary water production. Often they are used in combination with a backup device (boiler for winter heating). In this case we speak of hybrid systems.

1.2

Working Principles

First of all let us define a physical quantity useful to easily identify the thermodynamic performances of any thermodynamic cycle: the equivalent heat exchange average temperature, Tm,eq (K). Consider any transformation taking a system from point A to point B, as in Fig. 1.1a. The above temperature is equal to the ratio between the exchanged heat (subtended area by curve AB) and the entropy difference between B and A: RB Tds Tm;eq ¼ A sB  sA Doing so, it is possible to reduce a transformation to an isotherm where the actual heat exchange takes place. If now we refer to Fig. 1.1b and consider path AB we can see how segment A1 contributes with a lower Tm,eq than 12 and segment 2B contributes with a higher value of the same temperature. Therefore a thermodynamic cycle can be divided into equivalent Carnot sub-cycles at least for a preliminary estimation of its performances. This is one more reason to refer to this theoretical cycle in order to supply basic elements to enhance the understanding of some fundamental concepts concerning heat pumps.1 In the simplest configuration for residential uses a heat pump consists of an outdoor unit, containing compressor and a heat exchanger working as an evaporator in winter and a condenser in summer, and an indoor unit with another heat exchanger complementary to the previous one (condenser in winter and evaporator in summer). If subscripts C and F respectively indicate the hot and cold sources, the following relations hold (Figs. 1.2 and 1.3)2: First Principle of Thermodynamics QC þ QF ¼ L ðQC \0; L\0Þ Second Principle of Thermodynamics QC QF þ þ Sg ¼ 0 TC TF

1

Bear in mind we have referred to reversible transformations, while real transformations are not such. In particular lamination is absolutely irreversible, so that we can only say (if adiabatic) its final enthalpy is equal to its initial one and not isenthalpic. 2 Q > 0 if it is supplied to the system following the cycle and 0 (supplied to cycle), L < 0 (supplied to cycle), QC < 0 and, in addition |QC| > |QF|. It means that heat exchanged with the warmer source (indoor environment in winter and outdoor environment in summer) is larger than the one with the colder source. If a heat pump is used both in winter and summer, the net total energy exchange with the external source may approach zero, depending on the durations of the winter and summer periods. It is a favourable phenomenon for the external environment once the external source can accumulate and return this energy close to the user, as it occurs for geothermal heat pumps. Nevertheless the mechanical energy is provided to the heat pump all along the period of operation and the heat exchanged with the indoor environment has to be considered as the produced “useful effect”. A coefficient of performance, COP, is defined to characterize heat pumps performances as the ratio of heat exchanged with indoor environment and the

1.2 Working Principles

5

Τ

2 3

ΤC ΤF

ΤC

4

1

QC 3

s

2

condenser

L

Lamination valve 4

evaporatore

ΤF

compressor 1

QF

Fig. 1.2 Basic scheme of a compression heat pump and reference cycles in the planes p,h, top left and T, S. Source temperatures, different from the refrigerant’s ones, are also evidenced

SUMMER

indoor

condenser

indoor evaporator

evaporator

outdoor

WINTER

condenser outdoor

Fig. 1.3 Winter and summer operation with a four way reversion valve

6

1 The Fundamentals

mechanical work supplied to the machine. This is positive by definition and, when heat is exchanged with two sources, is given by3: Winter COPW ¼

QC 1 1   ¼ QF ¼ T L 1  jQC j 1  F 1  TC Sg jQC j

TC

Summer COPS ¼

QF 1 1  ¼ ¼  jLj jQC j  1 TC 1 þ TF Sg  1 QF

TE

QF

It clearly comes out that irreversibilities lower, even very much, the value of COP.4 The term Sg depends on the conditions of operation and TCSg, TFSg represent the energy losses caused by irreversibilities. The quantities: Winter cW ¼

TC Sg jQC j

Summer cS ¼

TF S g QF

are the ratios between the lost energy and the one exchange with the related source. The smaller is c the better is the expected performance. Figure 1.4 shows the effect of the above term, both as a reduction of COP and as a decrease of its sensitivity to outdoor temperature.  0 2  2     T C  TC TF  TF0 1 1 1 1 Sg ¼ jQC j 0   þ UF  QF ¼ UC TC TC TF TF0 TC0 TC TF0 TF with TC0 [ TC ;

  TF0 \TF ; jQC j ¼ UC TC0  TC ¼ UC DTC ;

  QF ¼ UF TF  TF0 .

Consider an internally reversible cycle, where irreversibilities are only located on the boundary as temperature differences between the evolving fluid, with equivalent temperatures TF’ (TC), and the heat sources. If UC and UF are the global heat transfer coefficient of the two heat exchangers we have: cW ¼

DTC DTF F UF DT TC Sg TC TF TF ¼ þ  DTC DTF C jQC j 1 þ DT 1  U C TC TF TC

For the sake of simplicity we refer to a sort of equivalent inverse “Carnot cycle” where irreversibilities are accounted for in Sg. 4 The name COP is today reserved only to the winter coefficient of performance, while for summer it is called the energy efficiency ratio, EER. 3

1.2 Working Principles

7 Winter (indoor temperature, TC =20°C)

(a)

COP

γ=0 γ=0.01

60

γ=0.05

50

γ=0.1

40

γ=0.2

30 20 10 0

-10

-8

-6

-4

-2

0

4

2

6

8

10

12

14

TF (ºC) Summer (indoor temperature, TF=24°C)

(b) 100 90

γ=0 γ=0.01 γ=0.05 γ=0.1

80

COP

70 60 50 40 30 20 10 0 27

29

31

33

35

37

39

41

43

45

47

49

TC (ºC)

Fig. 1.4 Trend of COP in winter (a) and summer (b) versus outdoor temperature for several values of c

cS ¼

DTC DTF C UC DT TF Sg TC TC ¼ TFDTF þ  DTC F QF 1  TF UF DT 1 þ TF TC

For a preliminary evaluation of the above quantities we can refer to the following values for an air/air heat pump: Seasons

TF (°C)

TC (°C)

Winter Summer

7 27

20 35

For the sake of simplicity, we can suppose UC = UF = U and DTC = DTF = DT (remember T is expressed in °C and T in K) and thus (see the trends in Fig. 1.5):

1 The Fundamentals

γ

8 0,11 0,1 0,09 0,08 0,07 0,06 0,05 0,04 0,03 0,02 0,01 0

winter summer

0

1

2

3

4

5

6

7

8

9

10

11

12

13

14

15

ΔT

Fig. 1.5 Trends of c versus DT in summer and winter

cW ¼ cS ¼

DT TC

1þ

DT TC

DT TF

1  DT TF

þ þ

DT TF DT TC DT TC DT TF

 

DT TF

1  DT TF DT TC

1þ

DT TC

What has been said so far emphasizes some basic thermodynamic aspects affecting the performances of a heat pump operating between two thermal sources. Of course referring to an equivalent inverse Carnot cycle is a convenient simplification to understand the fundamental working principles of this type of machine. Anyway, before going forward, it is the case to stress the main differences between the above reference scheme and the real behavior. They can be listed as follows (Fig. 1.6). • Fluids actually used can keep their temperatures reasonably constant as they undergo a phase change. Due to the close link between temperature and pressure in these transformations, a constant pressure is needed to have a constant temperature. This could be only achieved by neglecting friction losses in heat exchangers. In addition, due to the nature of some refrigerants, temperature changes occur also at constant pressure (Glide) as we will see in the following. Furthermore superheated vapor discharging from the compressor is cooled down, before reaching the saturated condition. Energy subtracted in this phase in a dedicated de-superheater is sometimes employed for different uses from ambient heating as the production of hot sanitary water. In the end liquid exiting from evaporator generally has some superheating to avoid liquid inlet into the compressor as well as liquid from evaporator is slightly subcooled not to have vapor in the expansion valve. • The expansion valve causes an unavoidable irreversibility, because it is not convenient to recover energy from the related pressure difference. • Compressor is characterized by friction losses, is not adiabatic and so on. All this leads to define the so called isentropic efficiency.

1.2 Working Principles

9

Outdoor source

Indoor source Winter condition

air air

water

water

ground Fig. 1.6 Heat pump interaction with thermal sources

• Thermal sources are in principle suppose with a uniform and constant temperature, but in reality this is not true. For instance in air cooled condenser, air temperature changes from time to time. This happens for any type of source in different ways. As shown in Fig. 1.9, outdoor sources can be air, surface water (rivers, lakes, sea) underground water, urban or industrial process wastewater and ground. The logarithmic average temperature of the fluid flowing in the related heat exchangers can be assumed as the reference source temperature. With regard to the indoor heat source, it can be both ambient air and water. The former case refers to a direct-expansion system, i.e., a system where refrigerant is in direct thermal contact with the either cooled or heated. This happens in the so called split systems. In the latter the heat exchange occurs between refrigerant and water of any hydronic system. The following combinations exist (Table 1.1): Figure 1.7 shows a split system.

Table 1.1 Heat pump nomenclature depending on heat sources

Outdoor source

Indoor source

Heat pump type name

Air Air Water Water Ground

Air Water Air Water Water

Air—Air Air—Water Water—Air Water—Water Geothermal

10

1 The Fundamentals

Indoor unit Outdoor unit

Fig. 1.7 A typical split system (Technibel)

As already said, a large temperature difference between the two sources has a negative effect on heat pumps performances. We might divide the total temperature jump into two smaller ones. Let us suppose to have to supply a power QC to the indoor environment, kept at a constant temperature TC, with a cold source at TF. We could use two stages, one working between the cold source and an intermediate temperature TC1′. It exchanges Q with the evaporator of the second stage, at TF2′. The exiting vapor goes to the second compression and, the to a condenser at temperature TC2′. If we still refer to Carnot cycles, as in Fig. 1.8 (lamination valves have been replaced by a reversible expander to eliminate irreversibilities) we get no advantage and the intermediate heat exchanger could only introduce irreversibilities. θ ’C2

θ θ ’C2

θC

θ ’C1

θ ’C2

Stage 2 θ’F2

θ ’F2

θ ’C1

θ ’C1

θ ’F2

Stage 1 θF

θ’F1

θ ’F1 S

θ’F1

Intermediate heat exchanger

Note – an expander is used in the reversible scheme instead of a lamination valve. Fig. 1.8 Two stage Carnot cycle

1.2 Working Principles

11

Cycle 2

θ

θ f Cycle 1

f’

g

b b’

c h

e a d

S

S

Fig. 1.9 Double stage cycle

Just as an example we describe the relations holding for a two stage cycle. Figure 1.9 refers to the temperature enthalpy plane to point out the various temperatures. Example 1.1 With reference to Fig. 1.9 we designate with subscript 1 the quantities referring to the lower temperature stage and with subscript 2 those referring to the higher temperature stage. If mk (k = 1 or 2) is the mass flow rate, h the enthalpy, QC the heat delivered to the indoor environment and Lk (k = 1 or 2) the compression work for each simple cycle we get: Stage 2   QC ¼ m2 hf  hg   L2 ¼ m2 he  hf and   hf  hg  COPðstage 2Þ ¼  he  hf Stage 1   Qc;1 ¼ m1 hf  he And the compression work   L 1 ¼ m 1 ð ha  hb Þ L 2 ¼ m 2 he  hf

L ¼ L1 þ L2

12

1 The Fundamentals

COP of the 2 stages combination in heating mode is:   m 2 hf  hg QC   ¼ COPð2stagesÞ ¼ L m 1 ð ha  hb Þ þ m 2 he  hf   hf  hg   ¼ m1 m 2 ð ha  hb Þ þ he  hf with

m1 he  hh ¼ m2 hb  hc

Therefore:   hf  hg QC   ¼ he hh COPð2 stagesÞ ¼ L h h ðha  hb Þ þ he  hf b

c

As already said, two fluids are used in these cases as, for instance R134a at high temperatures (e.g., 80 °C) in stage 2 and R410A at low temperature (e.g., −20 °C). Such a procedure can be used either for applications in cold places or for retrofit of already existing high temperature radiators.

We dedicate one more example to describe the effect of outdoor temperature time change on heat pump performances, with a constant indoor temperature. This is done referring to a very simplified model and, thus, gives only just an idea of the heat pump’s behavior. Let us suppose to keep the indoor temperature fixed at a constant set point value with a perfect thermostat. This mean neither time delay nor temperature band around the set point exist. On the other hand, outdoor temperature varies sinusoidally and instantaneously affects the heat exchange. Example 1.2 Let us refer to the obtained formulas in ideal conditions to emphasize some features that will dealt with elsewhere in more details. We graphically describe how COP is affected by the outdoor temperature. We impose this latter vary as a sinusoid around an average value Tmf and with an amplitude DTM. Tf ¼ Tm;f

 t  t  t  t  DTM 0 0 þ DTM sen2p sen2p ¼ Tm;f 1 þ s Tm;f s

DTM sinusoid amplitude s sinusoid period Of course the smaller is DTM/Tm,f the closer is the outdoor temperature to a constant value. Figure 1.10a shows COP trends versus time (hours of a day)

1.2 Working Principles

COP

(a)

13 outdoor souce air

outdoor source water

45 40 35 30 25 20 15 10 5 0 0

1 2

3 4

5 6

7 8

9 10 11 12 13 14 15 16 17 18 19 20 21 22 23

time (hour) outdoor source: air

COP/COP (Tmf)

(b) 1,8 1,6 1,4 1,2 1 0,8 0,6 0,4 0,2 0 0,98

0,985

0,99

0,995

outdoor source: water

1

1,005

1,01

1,015

1,02

Tf/Tmf

Fig. 1.10 a COP versus time in case of outdoor sources air (rhombs) and water (squares), b Ratio of instantaneous COP and COP(Tm,f) calculated at the average temperature, Tm,f versus T/Tm,f

in the cases of Tm,f = 8 °C e DTM = 5 °C (rhombs) and Tm,f = 10 °C e DTM = 1 °C (square). In Fig. 1.10b the ratio of the instantaneous COP to the one, COP(Tm,f), calculated at the average temperature, Tm,f, is reported versus the ratio between the temperature and its average daily value, Tm,f. This stresses the influence of the outdoor heat source. As above said, in the first case we refer to outdoor air, while in the second case the outdoor source might be water with a higher average temperature value and a lower temperature fluctuation. All this, herein evidenced for a single day, has a much greater importance if referred to seasonal performances. To account for this a seasonal COP is used (SCOP), defined as the ratio between the useful energy supplied during the related season and the energy that has to be provided to the heat pump to obtain this useful energy.

14

1 The Fundamentals

References 1. European Heat Pumps Market and Statistics 2015 by EHPA. 2. Growth in the world heat pump market August 2014 https://www.bsria.com/news/article/ growth-in-the-world-heat-pump-market/. 3. Heat Pumps Market: Global Industry Analysis and Opportunity Assessment 2015–2025. http:// www.futuremarketinsights.com/reports/heat-pumps-market.

Chapter 2

Types of Compression Heat Pumps and Their Main Components

Abstract The main components of compression heat pumps are treated herein. Their salient features, working principles and roles are dealt with, stressing their contribution to heat pumps operation and efficiency. Furthermore, products existing on the market are often referred to in order to allow the interested reader to have, at least, a rough idea of the available equipment, nowadays. Engine driven heat pumps, also named Gas Heat Pumps (GHP) are described in addition to the most commonly used electric heat pumps (EHP). Except for the driving motor, GHPs differ from EHPs both from the thermodynamic point of view, as they interact with three heat sources (they are a three-thermal-system), and for the possibility of using heat recovered by engine cooling. Last but not least, part of this chapter is devoted to describe CO2 heat pumps, due their peculiarity. In fact carbon dioxide has a very low critical temperature and, thus, they operate in hyper-critical condensions in most cases. Due to this a gas cooler is employed instead of a classical condenser.

2.1

Main Components of Compression Heat Pumps

From the scheme we have referred to so far, it clearly comes out that the main components of compression heat pumps are: • the compressor, that keeps the right pressure drop between evaporator and condenser to maintain the proper phase change temperatures to interact with the external (to the HP) sources; • the expansion valve, irreversibly taking the refrigerant from the condenser pressure to the one of the evaporator; • the condenser, where the superheated vapor coming from the compressor is de-superheated, first, then condensed to liquid with some degree of sub cooling to prevent vapor from entering the expansion valve; • the evaporator, where the mixture coming from the expansion device vaporizes. The exiting vapor can be either saturated (wet evaporator) or superheated (dry evaporator). In the former case a proper device (separator) is needed to prevent liquid from entering the compressor. In the latter case the vapor leaving the evaporator has a superheat of few degrees Celsius for the same purpose. © Springer International Publishing AG 2018 W. Grassi, Heat Pumps, Green Energy and Technology, DOI 10.1007/978-3-319-62199-9_2

15

16

2.2

2 Types of Compression Heat Pumps and Their Main Components

Compressor

Heat pumps mainly adopt volumetric compressors. They may be both reciprocating and rotary compressors. We will refer to the former ones to describe the main features of this type of device.

2.2.1

Reciprocating Compressor and Basic Concepts

Those we are referring in the following consist of cylinders where the vapor coming from the evaporator is sucked due to piston motion, and then released to the condenser. Figure 2.1 show the ideal cycle followed by a reciprocating compressor in the pressure volume plane. As well known, piston movement does not cover the whole cylinder volume. When it reaches the highest possible position, commonly referred to as top dead center, both suction and discharge valves are closed and discharge pressure, pM, is reached. At this point, a volume is left as no further compression is allowed due to the valve plate, named clearance volume, VN (V4 in Fig. 2.1). The smaller this volume, the more efficient the compression is. If VC (V3 in Fig. 2.1) is the cylinder volume at the end of compression, the volume of the fluid discharged, after the discharge valve opening, is VM = VC − VN. Then, the above valve closes and the re-expansion process of the clearance vapor occurs down to pressure pA (point 1 in the figure). The new volume VNA at pressure pA corresponds to the volume occupied by the clearance vapor VN at the discharge pressure.

1) Suction valve opens;

pM

4

1 -2) Gas introduced up to volume corresponding to bottom dead center (BDC);

3

2) Suction valve closes; 2 – 3) Compression; pA

2

1

3) Discharge valve opens; 3 -4 ) Gas is released down to top dead center (TDC); 4) Discharge valve closes;

VN VNA

VAS

4 – 1) Gas contained in clearance volume expands.

VA

Fig. 2.1 Reversible cycle of a reciprocating compressor

2.2 Compressor

17

The suction valve opens at pressure pA (BTD bottom dead center) and vapor enters the compressor. The volume VA = V2 − VN is the theoretical volume that could be sucked and VAS = V2 − VAN = VA − (VAN − VN) is the actual volume sucked by the compressor. The ratio VAS/VA is named the compressor volumetric efficiency. Example 2.1 Let us consider the isentropic compression of an ideal gas, in a cylinder with a clearance volume VN. Let us identify with VA the available volume at suction and with pA, TA, pB, and TB respectively the pressures and temperatures (K) at the suction and discharge points. The number of moles, nN, contained in the clearance volume is nN ¼

pM VN RTM

At suction, in the absence of clearance volume the numbers of moles that could be sucked would be: nA ¼

pA V A RTA

Actually (in the presence of clearance volume) we can suck a number of moles, nAS, equal to the difference between these latter diminished by the number of moles contained in the clearance volume. nAS ¼ nA  nN ¼

pA VA pA VNA pA VAS  ¼ RTA RTA RTA

The volumetric efficiency, ηv, of the compressor is: gv ¼

 VAS VN
¼ 1þ 1  b1=k VA VA

where b = pM/pA is the barometric compression ratio (simply called compression ratio) and k is the ratio between the gas specific heats or, more in general the polytrophic exponent. The clearance volume commonly varies between 2 and 5% of VA. Thus, if we assume a value of 5% and k = 1.4, values can be calculated by the following formula:   gv ¼ 1 þ 0:05 1  b0:714

18

2 Types of Compression Heat Pumps and Their Main Components

b

ηv

2 3 4 5 6 7 8

0.97 0.94 0.92 0.89 0.87 0.85 0.83

The volumetric efficiency reduction due to increasing of the pressure ratio can be easily explained as follows. The larger the compression ratio, the larger is the gas volume, VNA, after the expansion of the clearance volume. Consequently VAS decreases, as the mass sucked and then compressed during a cycle is given by q2VAS = ηvq2VA and the mass flow rate by m = ncyηv q2VA, if ncy is the number of cycles per second. The following figures show the trend of volumetric efficiency versus compression ratio for monatomic, diatomic and polyatomic ideal gases (Fig. 2.2) with VN/ VA = 0.05, and (Fig. 2.3) for a diatomic gas with different values of VAS/VA. It is, therefore, clear how both gas structure (even if ideal) and clearance volume affect volumetric efficiency. In reality, several phenomena have to be accounted for, causing irreversibilities. Among them we recall friction losses in the mechanical compressor’s components, heat losses along the compression (that we have previously assumed adiabatic), fluid friction losses and fluid leaks towards the suction valve through some seal flaws between the rotating (or alternating) and fixed parts of the casing. A very significant role is plaid by pressure losses in suction and discharge valves. They lower the suction pressure taking this to p2’ instead of the one required monoatomic gas

biatomic gas

polyatomic gas

volumetric efficiency

1 0,98 0,96 0,94 0,92 0,9 0,88 0,86 0,84 0,82

2

2,5

3

3,5

4

4,5

5

5,5

6

6,5

7

compression ratio

Fig. 2.2 Volumetric efficiency of an ideal gas (monatomic, diatomic and polyatomic) versus compression ratio

2.2 Compressor

19 VN/VA=0,01

VN/VA=0,05

VN/VA=0,1

volumetric efficiency

1

0,9 0,8

0,7 0,6

2

2,5

3

3,5

4

4,5

5

5,5

6

6,5

7

compression ratio

Fig. 2.3 Volumetric efficiency versus compression ratio for various VN/VA

by the heat exchange with the cold source p2 = pA, and cause the discharge pressure to be increased (to a value p3’) instead of the one required by the heat exchange with the hot source, p3 = pM. There exist some other reasons (see also the screw and scroll compressors) causing a difference between the pressures actually imposed by compressor and those imposed by external thermal sources. The actual pressure ratio occurring in the compressor, p3’/p2’ = b’, is called the internal compressor ratio and may be different from the previously defined b = p3/p2, also named the external compressor ratio. The above mentioned phenomena contribute to modify the volumetric efficiency and the work achievable. All this leads to introduce the isentropic efficiency, qc, defined as the ratio of the ideal enthalpy difference between discharge and suction, Δh, and the actual one, Δh’ (Fig. 2.4). qc ¼

Dh h3  h2 ¼ Dh0 ðh30  h3M Þ þ ðh3M  h2A Þ þ ðh2A  h2 Þ

In the case where external compression ratio is equal to the internal one, the work supplied to compressor in the presence of friction (la) is given by: Z3M l¼

"



pM vdp  la ¼ cp ðT2  T3M Þ ¼ cp T2 1  pA

# p1 p

2

Points 3 M and 2A respectively are at the same pressures as 3 and 2, p is the exponent of the real adiabatic transformation, and the work is negative as supplied to the system and T the absolute temperature (K). Recalling that, on the polytrophic 2’-3’1 1

Remember we are referring to an equivalent reversible transformation.

20

2 Types of Compression Heat Pumps and Their Main Components

p 3’ 3

pM

3M

Ideal transf. Real transf. 2

pA

2A 2’ Δh’ Δh h

p pM

3’ 3

p3’

Τ

pM

3’

3M

3M 3

pA

2

2

2A

pA 2A

p2’

2’

2’ S

v

Fig. 2.4 Ideal and real compression in the planes p,h, p,v and T,S

Z3M Tds ¼ la 2

We can display friction losses on the plane T, s as the area underneath the curve 2−3 M. Area 2−3−3 M represents the energy related to the compressed gas heating up: Z3M Areað2  3  3MÞ ¼

Tds  la ¼ cp ðT3M  T2 Þ  la 2

This phenomenon is called “thermal recovery”.

2.2 Compressor

21

Table 2.1 Values of k = cp/cv for some refrigerants

Fluid

k

T (°C)

p (bar)

NH3 CO2 R134a R437

1.31 1.29 1.11 1.15

0 27 30 25

1 1 1 1

To better describe the compressor technical features, a hydraulic efficiency is defined as: l þ la ¼ qy ¼ l

R 3M

 2 vdp  ¼
p1 pM p cp T 2 1  p A p

¼ p1 ¼ k k1

 p p1 pA v2 k k1 pA v2

1


p1 p pM pA


p1 pM p 1  pA

pk1 kp1

Such a parameter does not depend on compression ratio. Once the hydraulic or polytropic efficiency2 is known, the exponent of the polytropic curve can be obtained and viceversa. Therefore the isentropic efficiency can be written as: k1

qc ¼

h2  h3 T3  T2 b k 1 ¼ ¼ 1 k1 ; b ¼ ppMA h2  h3M T3M  T2 bqy k  1

It decreases with the compression ratio and depends on the type of fluid through k, ratio between the specific heats at constant pressure and volume. After recalling that k = 1.4 for standard air, some values of this parameter are given in Table 2.1 for four gas used in heat pumps. Figure 2.5 shows the theoretical trend (calculated here in) of the isentropic efficiency versus the compression ratio at constant values of the hydraulic efficiency, with k = 1.11 (continuous line), and k = 1.29 (dotted line). As already said, the pressure existing in the compressor, both at suction and discharge, is not the same as the one present in the circuit just before the suction and after the discharge. Therefore we introduced two compression ratios the internal, b’, and the external, b, ones. Three different cases can occur, for reciprocating, screw and scroll compressors, as follows.

The name “hydraulic efficiency” refers to the fact that the thermal recovery is negligible in the hydraulic machines, so that this efficiency is equal to 1. This parameter is also called “the polytropic efficiency” because a reference reversible polytropic is usually considered, with an average exponent equal to that of the actual transformation.

2

22

2 Types of Compression Heat Pumps and Their Main Components 0,95 0,9

isentropic efficiency

0,85 0,9

0,8

0,8 0,7

0,75

0,6 0,9

0,7

0,8 0,7

0,65

0,6

0,6 0,55 0,5 1,5

2

2,5

3

3,5

4

4,5

5

5,5

6

6,5

7

compression ratio

Fig. 2.5 Isentropic efficiency versus compression ratio for k = 1.11 (continuous line) and k = 1.29 (dotted line) and for several hydraulic efficiencies, as indicated in the legend

• b = b’—the discharge opening opens exactly when the refrigerant pressure (at the compressor exit) matches the one of the discharge line and the gas is immediately sent to it. This event very seldom occurs. • b’ < b—the compressed gas has not yet reached the pressure of discharge line, sub-compression. This causes a sudden refrigerant flow towards the compressor, with an abrupt pressure increase, over compression. After that gas is expelled to the circuit. • b’ < b—the compressed refrigerant pressure is larger than the one of the discharge line. This originates a sudden flow leaving the compressor. In both the last two cases some uncontrolled expansions occur and introduce additional losses. The most critical is the over compression, as gas expands within a larger volume while discharging. The trend of the isentropic efficiency versus the pressure ratio is qualitatively depicted in Fig. 2.6, and a peak of its value is easily identifiable. Figure 2.7 shows these trends for different types of compressors. They basically have the same shape even if some peculiar behavior occurs, depending on the type of compressor, in particular for screw machines.3 The above graphs show how the isentropic 3

A screw compressor can be optimized so that its isentropic efficiency is maximized in correspondence to a given value of compression ratio. Thus the maximum of the curve coincides with this ratio. During operation compression ratio can change (e.g., at partial load) and the curve can shift either leftward or rightward. To restore the optimum, compression ratio should be increased or decreased correspondingly.

2.2 Compressor

23

0,85

isentropic efficiency

0,8 0,75 0,7 0,65 0,6 0,55 0,5 0,45 0,4

2

2,5

3

3,5

4

4,5

5

5,5

6

compression ratio

Fig. 2.7 Typical trends of isentropic efficiency (top graph) and of volumetric efficiency (lower graph) versus compression ratio for reciprocating, screw and scroll compressors

isentropic efficiency

Fig. 2.6 Typical trend of a compressor isentropic efficiency

100%

reciprocating 50%

screw scroll

1

10

20

volumetric efficiency

compression ratio

100%

scroll

72%

screw 50%

reciprocating

1

7

10

14

20

compression ratio

efficiency could also increase with a reduction of the compression ratio. Furthermore the typical trends of the volumetric efficiency are displayed in the same figure. The thermodynamic cycle irreversibilities play a different role on heat pumps performances in winter and in summer. In fact, in winter, the useful effect (i.e., the useful heating output) h2 − h3, increases, due to the increase of the compression work, h2′ − h2. Thus the COP changes as below:

24

2 Types of Compression Heat Pumps and Their Main Components

ideal

case h2  h3 COPid ¼ h2  h1 real

case h2 0  h3 ð h2  h 3 Þ þ ð h2 0  h2 Þ COPre ¼ ¼ qc h20  h1 ð h2  h1 Þ 0 ðh2  h1 Þ  ðh2  h1 Þ ¼ qc COPid þ qc ð h2  h1 Þ ¼ qc COPid þ 1  qc On the other hand, in summer: ideal case 4 EERid ¼ hh12 h h1

real case 4 EERre ¼ hh10h h1 ¼ qc EERid 2

Figure 2.8 shows the trend of the relative change of the performance parameters (ratio of the difference between the real value minus the ideal one and the ideal value) versus the isentropic efficiency, to give an idea of its influence on summer and winter performances. Figure 2.9 qualitatively shows the difference between a cycle with an ideal compressor and a cycle with a real one. Another typical curve of volumetric compressor is the one linking the flow rate to the pressure head supplied. It is often provided as compression ratio versus volumetric flow rate, as in Fig. 2.10. For volumetric machines, the pressure supplied by compressors is practically independent from the flow rate, and only related to the circuit hydraulic characteristic curve. Thus the curve compression ratio-flow rate is usually represented by a COP

relative change

0 0,5 -0,1

0,55

0,6

0,65

0,7

EER

0,75

0,8

0,85

0,9

-0,2 -0,3 -0,4 -0,5 -0,6

isentropic efficiency

Fig. 2.8 Relative change of summer and winter performance coefficient versus isentropic efficiency

2.2 Compressor

25

Fig. 2.9 Cycles with ideal (continuous line) and real (dotted line) compressor

p

2

3

4

2’

1

h

Fig. 2.10 Compression ratio versus flow rate of a volumetric compressor. V is the volumetric flow rate, in this case

β n (round per second)

n3

D

n2

β2 n1

BD - Change of β at constant flow rate (V2). A

β1

V1

B

V2

C

V3

ABC - Flow rate (V) change at constant β.

V(m3/s)

straight line with a negative slope (not exactly vertical due to the volumetric efficiency decrease with increasing the compression ratio). Figure 2.10 also displays the effect of the number of rounds per second of the compressor shaft. An increase of this number moves the curve to the right, toward larger flow rates (from A to C in the figure) at the same compressor ratio. In this case the circuit hydraulic losses have to be decreased. If pressure has to be augmented at constant flow rate (from B to D in the figure), the pressure losses must increase. The corresponding decrease or augmentation of the pressure losses is obtained by opening or closing the metering device (expansion valve).

26

2 Types of Compression Heat Pumps and Their Main Components

The characteristic number of rounds per minute of compressors may be also very different for the different types. For example, for reciprocating compressors, they roughly go from a hundred for large and slow compressors with compression ratio 2–3, to a thousand for the smallest ones with a compression ratio around 10. In rotary compressors there might be several thousands rounds per minute.

2.2.2

Screw Compressors

Thanks to the technological progress in heating and cooling applications, rotary compressors are often employed instead of reciprocating compressors. Among other things this is due to their smaller size, larger silentness, smoothly running, low vibration and better control and modulation capability. They can be roughly divided in compressors with a single rotating axis (vane and scroll) and with two rotating axes. Among them we include vane, lobe, screw and scroll compressors. Vane and scroll compressors have a single shaft (single rotation axis), while the others can have both one and two axes. In general, screw compressors have two axes, i.e., two screws. Screw compressors are used for power values above 50 kW, where they have a better efficiency than the reciprocating ones, even if low power screw compressors (down to 2.25 kW) are available for small applications. Their working principle is shown in Fig. 2.11: two meshing helical screws of different diameters constitute the suction

Fluid sucked by the two counter-rotating screws axially moves and is compressed within the progressively reducing space between the screw threads.

discharge volume occupied by fluid Fig. 2.11 Screw compressor working scheme

2.2 Compressor

27

compressor rotors. Gas enters at the suction side and moves through the threads as the screws rotate. They force the gas to the discharge port at the end of the screws, progressively reducing the gas volume. Generally they have smaller compression ratios (b = 3:4) than the reciprocating compressors, they can be used with several stages in series. The most common configuration consists of a male rotor with four lobes and a female one with six indentations. Other possible configurations are 3(lobs)/5(indentations) and 5/7. Rotor diameters commonly ranges from 12 to 32 mm. Rotors are located in horizontal cylindrically shaped casings provided with suction and discharge ports. Lubricating oil is injected on the threads to prevent refrigerant leakages, thanks to the presence of an oil film. It is then recovered in an oil separator located close to the discharge port. Suction phase begins when the two moving rotors leave the suction port open. Fluid enters the compression region and moves along the screw axes. The suction port is closed by the engaging rotors and compression starts, with the discharge port closed. The ratio, vi, between the initial (suction) and final (discharge) volumes is the so called “intrinsic volumetric ratio”. Some typical values are 2.2; 2.6; 3.2; 4.4. A given compression ratio corresponds to each vi, depending on refrigerant properties. For a given fluid an optimal (top isentropic efficiency) compression ratio can be realized, by using an appropriate intrinsic volumetric ratio. What has been said above about over and sub compression holds also for this type of compressors.

2.2.3

Vane and Scroll Compressors

Vane and scroll compressors are mainly employed at the lowest power values. Figure 2.12 shows the scheme of a vane compressor. The rotor is eccentrically placed with respect to the casing. On this, suction and discharge ports are located, discharge

suction

Fig. 2.12 Vane compressor working scheme

28

2 Types of Compression Heat Pumps and Their Main Components

(a)

(b) fixed scroll

Orbiting movable scroll movable scroll (d)

(c)

P

high pressure pocket

P

discharge duct

low pressure pocket

Moving spiral, by orbiting on fixed scroll, figures (a) and (b), forms progressively smaller chambers (pockets), as in figures (c) and (d). Fig. 2.13 Working scheme of scroll compressor

without any valve. Sliding vanes are located on the rotor and pushed against the cylindrical casing by centrifugal forces produced by rotation. They originate chambers with a progressively decreasing volume from suction to discharge. A good continuity of the refrigerant flow is guaranteed, in this case too. Scroll compressors are basically constituted by two scrolls (spirals), a fixed and a movable one, sketched in Fig. 2.13. The latter is driven by a shaft that makes it orbit (not rotate) about the shaft axis. So, a chamber is formed, compression starts once the suction port is sealed off, progressively reducing the gas volume between the two scrolls. Seal between fixed and movable scroll is guaranteed by a lubricating oil film. As above said the chamber is in contact with suction, and fluid flows in. After a 90°-rotation, the scroll movement closes the suction port, refrigerant stays confined within the two scrolls and gradually compressed until it is released to the discharge duct.4 As all compressors without suction and discharge valves,5 they have larger isentropic and volumetric efficiencies than the reciprocating ones. They are commonly inserted in a hermetic shell together with the driving electric motor. A typical configuration is shown in Fig. 2.14. 4

Videos existing on You-tube may help clarify scroll compressor operation. Actually, a dynamic discharge valve can be adopted in particular in high pressure ratio applications typical of refrigeration. It is located at the scroll discharge port to prevent entry of high pressure gas into the scroll set during the unloaded state.

5

2.2 Compressor

29

space occupied by compressor

Space occupied by driving electric motor and auxiliaries.

Fig. 2.14 Typical external shape of a scroll compressor

By axially separating the two spirals (lifting the movable scroll) capacity reduces to zero. In the discharge phase the movable scroll moves 1 mm apart from the fixed one, annulling the gas flow rate (see: Copeland Scroll Digital™ Compressors). Generally a scroll compression has its own optimal compression ratio. When the actual compression ratio is lower than this one, over compression losses occur (e.g., half load condition [1]). At high compression ratios sub compression occurs, that can be prevented introducing a dynamic discharge valve, similar to those of reciprocating compressors. Compression is smooth and silent as very few moving part are involved. It makes this compressor very reliable. Bearing have to be carefully lubricated, while no oil injection in the compression process is needed. The compressor capacity is commonly controlled by an inverter. The application ranges of compressors can be briefly summarized in Table 2.2. We remark that for the use of ammonia open compressors are used (generally reciprocating and screw), due to its chemical aggressivity. Thus the driving motor is outside the compressor casing. Furthermore we recall the following definitions. • Open compressor—the driving motor is separated from the compressor, independently air-cooled and connected by a mechanical coupling. • Hermetic compressor—motor and compressor are inserted in the same casing and the motor is cooled by the same fluid circulating in the compressor. • Semi-hermetic compressor—a compressor directly coupled to the driving motor, in the same casing, but with a direct access, separate by the motor.

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2 Types of Compression Heat Pumps and Their Main Components

Table 2.2 Main types of compressors Type

Model

Capacity (kW)

Refrigerant

Application

Reciprocating

– Hermetic – Semi hermetic – Open

0.1/30 30/250 250/50

Industrial and commercial refrigerators, low temperature industrial refrigeration

Vane

Hermetic

0.75/3

Scroll

Hermetic

3.5/90

Screws

– Semi hermetic – open – Semi hermetic – Open

80/8000

R134a R404A R407A R407C R717 R744 R407C R410A R744 R407C R410A R407C R134a R717 R134a R410A

Single screw

2.2.4

100/500

Small refrigerators, portable air-conditioning, split systems Low and medium size air-conditioning Medium and large power air conditioning. Industrial refrigeration Medium and large power chillers for commercial and industrial climatization

Control of Compressors’ Operation

Compressors must be enabled to work off their nominal load. The on-off control is the simplest way to reach this goal, but it is the most energy consuming. In this way, the reference signal is a set point temperature, TSP, fixed by a thermostat. When this value is exceeded by ΔTSP (depending on thermostat accuracy) the compressor turns off. It starts again once the temperature achieves the value TSP − ΔTSP. In order to keep comfort conditions within the internal environment, ΔTSP should be as little as possible, but it could cause too many on- offs, thus stressing too much both compressor and driving motor during start-up phases. Besides, this would produce a COP decrease. Therefore some techniques have been implemented to work at partial load. In multicylinder6 reciprocating compressor one or more cylinders are made ineffective. This is done by bypassing fluid from suction to discharge of the cylinder we want to deactivate, by signals sent to a solenoid valve. So we obtain a step reduction of the active cylinders number as shown in Table 2.3. Therefore the number of on-off is reduced. The load to be supplied to the compressor does not proportionally decrease with the percent of reduction (e.g., a 33% reduction may correspond to 40% of the nominal load), as the ineffective cylinders are anyway operated by the crankshaft, consuming power.

The multiple cylinder compressor has also the advantage to keep the fluid flow smoother.

6

2.2 Compressor

31

Table 2.3 Active cylinder reduction in a reciprocating compressor Total number of cylinders

Active cylinders

Capacity (%)

Reduction (%)

4

4 2 6 4 2

100 50 100 67 33

no 50 no 1/3 2/3

6

volume shaping suction amount of fluid to discharge after compression

amount of fluid back to suction

slide oil

discharge

spring

Fig. 2.15 Scheme of a sliding valve for screw compressors. A piston activated by pressurized oil moves the slide towards suction or away from it, modifying the screw length engaged in compression

A method to obtain a continuous modulation (at least in a given range) consists in changing the rotation speed of the driving motor (being it an electric motor or an internal combustion engine). Electric heat pumps often use an inverter to control the electric motor. Such a device changes the feeding frequency from lower values than the mains one (50 or 60 Hz) to much higher frequencies. The main advantages are: a better achievable comfort, smoother start up, but, above all, an increase of the instantaneous and seasonal COP. It is even possible to gain a COP increase, respect to the nominal value, at a reduced flow rate. This is due to the use of oversized (in this case) heat transfer surfaces in comparison with the design nominal conditions. Thanks to this the temperature differences among heat exchangers and thermal sources shrink. Screw compressors. A typical control employed in screw compressors, using a slide valve, is outlined in Fig. 2.15.7 A slot, parallel to the screws axis, can be A duct can also be inserted to allow the fluid flow toward the economizer.

7

32

2 Types of Compression Heat Pumps and Their Main Components

gradually opened (or closed) by a sliding device. This is activated by the pressure exerted by an oil piston, depending on the actual operation requirements. The intake pressure acts on the left of the valve and the discharge pressure on the right. The right side contour of the slide is properly shaped to keep the intrinsic volumetric ratio practically constant within a given range (e.g., 70% of the full load). So a pretty much constant isentropic efficiency is obtained in the above range. When the slide is totally shifted to the intake side, the suction volume has its minimum value and compression takes place all along the screws length. By moving rightward (to discharge), the slide increases the suction volume, reducing the screw length got involved in compression. Thus the flow recirculating back to suction increases, while the one discharging decreases. As the consequence of this the actual sucked volume to be compressed lowers and the compression ratio increases. Viceversa, if the slide moves to the opposite side. The slide shift can be either stepwise or continuous, depending on the application requirements. The use of a continuous shift control is more convenient in the presence of fully variable loads. The stepwise configuration generally has four levels (% of the full load): • 10% minimum level determined by the oil injected in the compressor and commonly used only for start-up; • 50% • 75% • 100%, full load. In some cases both stepwise and continuous controls are feasible on the same compressor. Several types of the mentioned control method exist. For example: an additional flow rate bypass at partial loads can be adopted as well as a variable volumetric ratio, so that the compressor could always operate with the top isentropic ratio in correspondence to the required loads. Referring to Fig. 2.16, let us suppose to have a compressor characterized by curve 2 (isentropic efficiency vs. compression ratio), with an optimal compression ratio CR2 and constant intrinsic volumetric ratio. CR2 is the most frequently occurring value during daily operation. Anyway load variations can lead to different compression ratios, for instance CR1 or CR3. In this case a decrease of isentropic efficiency would occur on curve 2, i.e., if we work at constant volumetric ratio. Thus, in the case of rather frequent load changes, a compressor with a variable intrinsic volumetric ratio is suitable, where the trend of isentropic efficiency versus compression ratio passes trough CR1 and CR3. Anyway it is always recommended to contact the manufacturers. Scroll compressors. Capacity modulation, except for the on-off and variable speed methods, can be also obtained by axially distancing the movable scroll from the fixed one. Meanwhile the compressor keeps on rotating, with no significant power losses, at list down to a certain degree of modulation. The procedure is the following (at least the one adopted in Digital Scrolls by Copeland): in nominal conditions the movable scroll (lower scroll) is kept in place, keeping the nominal axial position. The scroll movement is activated by oil pressurized, or depressurized

2.2 Compressor

33 optimum compression ratios

Isentropic efficiency

80

CR1 CR3 CR2

curve with variable intrinsic volumetric ratio

70

60

50 2,0

1

curves with optimized intrinsic volumetric ratio 2,5

3,0

3,5

4,0

4,5

5,0

5,5

2

3

6,0

Compression ratio

Fig. 2.16 Trend of isentropic ratio of a screw compressor with optimized intrinsic volumetric ratio (curves 1, 2, 3) and with variable intrinsic volume ratio. CR1,2,3 are the points corresponding to top isentropic efficiencies

Fig. 2.17 Load modulation cycle of a scroll compressor

40% modulation full load

zero load

trough an electric control valve This valve, opens and closes, driven by a digital signal, separating the two scrolls axially by one millimeter or restoring the nominal axial position. When the two scrolls are in the nominal position the compressor works at full capacity. When they are separated it works at zero capacity. Modulation is achieved by varying the time of full and zero capacities. The cycles generally last from 10 to 30 s. Figure 2.17 shows a digital scroll modulation in a 20 s cycle where the compressor operates at full capacity for 8 s and 12 s at zero capacity.

2.2.5

Inverter Control

Inverter allows for varying the compressor rotation frequency in order to change flow rate at constant compression ratio and, therefore, at constant volumetric and isentropic efficiencies. The inverter used for a.c. electric motors transforms grid a.c.

34

2 Types of Compression Heat Pumps and Their Main Components

Inverter input

output

change of output voltage

change of frequency

PWM (Pulse Width Modulation) low voltage level t

t high voltage level t

t

Fig. 2.18 Inverter operation

into d.c. voltage. As an output it generates electric pulses, with different amplitude and frequency, simulating an a.c. voltage. The value of this latter is modulated by changing the signal amplitude, PMW (Pulse Width Modification), at a fixed frequency. The change of frequency of simulated voltage is obtained by varying pulses frequency, and, thus, the number of rounds per second of compressors, see Fig. 2.18. As the input a.c. voltage is converted in a d.c. voltage, at first, also a three-phase load can be fed by a single-phase voltage input. The output signal has a harmonic residual, causing electromagnetic noise, which may propagate in the surrounding environment. If V is the voltage applied to the motor, U the magnetic flux, f the frequency and C the torque applied to the rotor, the following relations hold: V / Ux P ¼ Cx V 2 ðUxÞ2 / x x2 x ¼ 2pf C/

Usually the magnetic flux is kept constant to avoid magnetic saturation of the iron nucleus with an increase of parasitic currents and consequent overheating (in hermetic compressors it would cause refrigerant overheating). To this purpose, a voltage proportional to frequency has to be applied and power grows up linearly

2.2 Compressor Fig. 2.19 Voltage versus frequency trends and torque behavior

35

V C C

constant torque

f V V=const

C

variable torque C C torque F frequency V voltage

f

with increasing frequency. The top achievable voltage is the one provided by the electric grid. Nevertheless frequency can be further increased, but doing so the relation between voltage and frequency is no longer linear and the torque decreases. In other cases, voltage is kept constant making the magnetic flux diminish, to compensate iron losses that increase with the frequency squared. Consequently power decreases with increasing frequency. The above two cases are sketched in Fig. 2.19. A better control can be achieved by using the so called “vector inverter, which can control both active (in phase with voltage) and reactive (90° out of phase) current components. For a better control, device, named “encoder”, may further be adopted. It tracks the turning of motor shafts to generate digital position and motion information. Dynamic power losses depend on the square of feeding voltage and on commutation frequency. As an example we report, in Table 2.4, some data related to an

Table 2.4 Some inverter data Typical useful mechanical power (kW)

5.5

7.5

11

15

18

…

45

Estimated power losses at nominal load (W) Efficiency

269

310

447

602

737

…

1636

0.96

0.96

0.96

0.96

0.96

…

0.96

36

2 Types of Compression Heat Pumps and Their Main Components

electric motor inverter. For a more complete overview of existing products reader can refer to [2] by Danfoss. The inverter efficiency is commonly above 92%.

2.2.6

The Compressor Operation Range

A dedicated region on the plane evaporation versus condensation temperature is provided for any compressor, for each given refrigerant. This is the compressor operating range. Out of this range manufacturers do not guarantee its performances. In Figs. 2.20, 2.21, 2.22, 2.23, 2.24 some of the above ranges are shown, obtained by elaborating the data supplied by Copeland (General-Product-Catalogue2014-IT_0.pdf). For more detailed information refer to [3]. Suction superheat is also specified in addition to the type of fluid used. The first three figures refer to scroll compressors with 10 °C superheat. The fourth (2.23) refers to reciprocating compressors, with four or six cylinders, provided with a

70

General-Product-Catalogue-2014-IT_0.pdf 65

R407C superheat 10K

60

Condensation temperature (°C)

55

50

45

40

35

30

25

-30

-25

-20

-15

-10

-5

20

0

5

10

15

20

25

30

Evaporation temperature (°C)

Fig. 2.20 Operation ranges of some scroll compressors by Copeland, each marked by a different line, using R407. Power of driving motor from 1.1 to 22 kW, capacities from 3.7 to 81.7 kW

2.2 Compressor

37 70

General-Product-Catalogue-2014-IT_0.pdf 65

R410A superheat 10K

60

Condensation temperature (°C)

55

50

45

40

35

30

25

-30

-25

-20

-15

-10

-5

20

0

5

10

15

20

25

30

Evaporation temperature (°C)

Fig. 2.21 Operation ranges of Copeland scroll compressors using R410 A as refrigerant. Motor powers from 1.4 to 44 kW, capacities from 5 to 160 kW

continuous inverter modulation. These figures give an idea of which are the main key factors influencing compressor performances. It is the case to stress some more that the cooling of electric motors is performed by the same refrigerant in hermetic compressors. As already said, Fig. 2.23 concerns reciprocating compressors with an external air cooled driving motor. With regard to this figure we remark: • The fan is the motor cooling fan. • The SGRT (Suction Gas Return Temperature) is the vapor temperature at compressor suction. The degree of superheat is given by the difference between this temperature and the evaporation one. • SH (Superheat) is the suction superheat. In the end we show the changes of operation range due to injection of vapor, taken from condenser exit, into the compression, with refrigerant R410A. Such a procedure widens this range, increasing heating potentiality and lowering discharge vapor temperature (improving isentropic efficiency) (referring to figs. 2.24 and 2.25).

38

2 Types of Compression Heat Pumps and Their Main Components General-Product-Catalogue-2014-IT_0.pdf

80 75

R134a superheat 10K

70 65 60

condensation temperature (°C)

55 50 45 40 35 30 25 20 15 10 5 -25

-20

-15

-10

-5

0

0

5

10

15

20

25

evaporation temperature (°C)

Fig. 2.22 Operation ranges of Copeland scroll compressors using R134a. Motor powers 1.5 to 22 kW, capacity 3.3 to 53.2 kW

2.3

Expansion Valve

As above said it is a metering device that feeds refrigerant to the evaporator, lowering its pressure from the condenser value to that of the evaporator, in order to keep suitable transformations temperatures for heat sources. In simplest applications it is obtained by a fixed bore capillary tube where the total-system charge flows in any operating condition. It has to be long enough to supply the total pressure drop at full flow rate and is generally helically coiled. Of course, such a device is not able to face load variations. Therefore several systems allowing for varying the discharge area of the valve according with the actual required load have been employed. The rationale is quite simple. As a consequence of a reduction of the heating power requested by the internal environment and at a constant flow rate, the condensation phase shifts the outlet point towards larger subcooling in the liquid. At the same time the superheat at evaporator exit increases. Both subcooling and

2.3 Expansion Valve

39

General-Product-Catalogue-2014-IT_0.pdf

60

R404A

55 50

d

a

40

c 35

b

30 25 20

condensation temperature (°C)

45

e

15 10 5

-60

-55

-50

-45

-40

-35

-30

-25

-20

-15

-10

-5

0

0

5

10

evaporation temperature (°C)

Fig. 2.23 Operation ranges of Copeland™ Stream Digital with Core Sense™ Diagnostics reciprocating compressors (4–6), with refrigerant R404 A. They use continuous modulation by an inverter from 50 to 100% (4 cylinders) and from 30 to 100% (6 cylinders), with the following characteristics(letters a, b, c, d, e refer to each graph): a 25 °C SGRT at 100% load or 0 °C SGRT + cooling fan, driving motor modulation at 33% (6 cylinders) and 50% (4 cylinders); b 25 °C SGRT with cooling fan and modulation at 33% (6 cylinders) or 50% (4 cylinders); c 0 °C SGRT with cooling fan and modulation at 33% (6 cylinders) or 50% (4 cylinders); d 25 °C SGRT at 100%; e SH > 20 °C at 100%

superheat increase as larger is the unbalance between the requested and the available power. It is, therefore, necessary to lower the flow rate in such a case. The valve discharge area has to be reduced. In many cases the actuating control signal comes from a sensor measuring vapor temperature at evaporator exit. This to keep vapor superheat at compressor suction fixed at a set-point. In this case we speak of thermostatic valve and this method is applied in dry evaporators (those where the exiting vapor is superheated). If Δpv is the valve pressure drop with a mass flow rate m, it can be set forth as Km2, where K is the corresponding flow coefficient. At a reduced flow rate, m’, the valve partially closes, keeping the pressure drop constant. The new flow coefficient

40

2 Types of Compression Heat Pumps and Their Main Components 125

General-Product-Catalogue-2014-IT_0.pdf

115

R744 (CO2) superheat 20K

105 95

Critical point:

85

condensation temperature (°C)

304K (31°C) 7,38MPa (73 bar)

75 65 55 45 35 25 15 5

-55

-50

-45

-40

-35

-30

-25

-20

-15

-10

-5

-5 0

5

10

15

20

25

-15 -25

evaporation temperature (°C)

Fig. 2.24 Refers to reciprocating compressors used with carbon dioxide. The graph (continuous line) on the top right regards compression in the hypercritical region, while the lower graph (dotted line) concerns a subcritical compression

has to become K’ = Δpv/m’2. Both K and K’ are two values of the flow characteristic of the installed valve. Example 2.2 A heat pump, working with R134a8 supplies a nominal heating load of 10 kW at 44 °C.9 Let us suppose liquid inlet into the expansion valve to be saturated. With data provided in the following Table 2.5 the mass flow rate is: m¼

8

QC 10 ¼ 0:063 kg/s ¼ hv  hl 158:7

The data are taken from NIST Chemistry Web Book. We suppose the de-superheating is used for sanitary hot water production. This may occur in offices, where sanitary water requirements are usually low.

9

2.3 Expansion Valve

41 R410A superheat 10K

Da General-Product-Catalogue-2014-IT_0.pdf 70 65 60

50 45 40 35

condensation temperature (°C)

55

no vapor injection

dry vapor injection

humid vapor injection.

30 25

-30

-25

-20

-15

-10

-5

20

0

5

10

15

20

25

30

evaporation temperature (°C)

Fig. 2.25 Change of operation range due to vapor injection

If the required power has a 10% decrease, the mass flow rate has to be reduced of the same % to keep the same conditions. If the lower temperature is −2 °C, lamination goes from 1130 to 272 kPa, with a pressure drop equal to 858 kPa. The relation between the new, K’, and the old, K, flow coefficients is: K 0
m 2 ¼ ¼ 1:23 m0 K First of all, we remark that the refrigerant flow rate cannot be reduced at one’s choice. A first constraint is imposed by the need of preventing liquid to flow into the compressor. Therefore vapor has to leave the evaporator (dry evaporator) with a certain degree of superheat (not less than 3–5 °C of static superheating). Thus the expansion valve receives a temperature signal from the evaporator exit, so that the flow rate be reduced if the detected value is lower than the fixed set-point, or increased if it is larger. The following definitions of superheating are given: • Static superheating: corresponding to this value the valve starts opening and the flow rate increase begins. • Opening superheating: is the value, larger than the static superheating, necessary to produce a given valve potentiality. • Operating superheating: is the sum of the two previous ones.

42

2 Types of Compression Heat Pumps and Their Main Components

Table 2.5 R134a data °C

kPa

vl (m3/kg)

vv (m3/kg)

hl (kJ/kg)

hv − hl (kJ/kg)

hv (kJ/kg)

−2 0 2 4 6 8 12 16 20 24 26 28 30 32 34 36 38 40 42 44 48

272.2 292.8 314.6 337.7 362.0 387.6 443.0 504.3 571.7 645.8 685.4 726.9 770.2 815.4 862.6 911.9 963.2 1016.6 1072.2 1130.1 1252.9

0.0007684 0.0007723 0.0007763 0.0007804 0.0007845 0.0007887 0.0007975 0.0008066 0.0008161 0.0008261 0.0008313 0.0008367 0.0008421 0.0008478 0.0008536 0.0008595 0.0008657 0.0008720 0.0008786 0.0008854 0.0008997

0.0744 0.0693 0.0647 0.0604 0.0564 0.0528 0.0463 0.0408 0.0360 0.0319 0.0300 0.0283 0.0266 0.0251 0.0237 0.0224 0.0211 0.0200 0.0189 0.0178 0.0160

49.17 51.86 54.55 57.25 59.97 62.69 68.19 73.73 79.32 84.98 87.83 90.70 93.58 96.48 99.40 102.33 105.29 108.27 111.26 114.28 120.39

200.12 198.60 197.07 195.53 193.95 192.36 189.11 185.74 182.28 178.70 176.87 175.00 173.09 171.16 169.18 167.17 165.12 163.01 160.88 158.69 154.16

249.29 250.46 251.62 252.78 253.92 255.05 257.29 259.47 261.60 263.68 264.70 265.69 266.67 267.64 268.58 269.50 270.41 271.28 272.14 272.97 274.55

A liquid separator can be inserted immediately after the evaporator, just to be sure and to have a low superheating. Actually these separators are used in the case of wet evaporators, where no superheating is required to increase efficiency. A further important parameter is the liquid subcooling at the expansion valve inlet. This is necessary to avoid vapor bubbles formation in the liquid line leading to the valve that would reduce its performances. A typical minimum subcooling value is 4 °C. The expansion valves are generally classified as: • Constant pressure expansion valve—also improperly called automatic expansion valve, it keeps the pressure inside the evaporator constant, no matter what the load inside the evaporator is. It does not allow the control of flow of refrigerant and, thus, this type of valve is not used when this control is needed. • Thermal (thermostatic) expansion valve—it controls the amount of refrigerant flow thereby controlling superheating at evaporator outlet. Thermal expansion valves are often generically referred to as “metering devices”. They are employed with variable thermal load. They operate according to the superheating at evaporator exit and to its pressure. This pressure has to be kept below a fixed threshold called MOP (Maximum

2.3 Expansion Valve

43

Operating Pressure) to avoid any abnormal operation of the compressor. In normal conditions (pressure below MOP) the expansion valve works according to the superheat, but once MOP is reached the valve orifice reduces preventing any further pressure increase. Figure 2.26 schematically shows the location (in the cycle) of the temperature sensor controlling the expansion valve. The related signal is collected by a bulb connected to the valve through a capillary tube. The fluid in the bulb contracts and expands according to the refrigerant superheat and cause changes in volume of a chamber of the valve, provided with an elastic membrane. It is attached to the valve stem moving up and down to increase or decrease the flow rate, as depicted in Fig. 2.27.

subcooled liquid p 3

2

3

2

4 4 superheated vapor

1

1

h

Fig. 2.26 Significant physical quantities and signals controlling operation of an expansion valve

fluid from measuring device (1) membrane

regulation spring

p4

Fig. 2.27 Scheme of a thermostatic valve

44

2 Types of Compression Heat Pumps and Their Main Components

On the top of the figure there is the membrane connected to the stem. The valve shutter is located on the other end of the stem and controls the valve orifice opening (in some types the orifice is interchangeable). One more chamber is placed underneath, where the evaporator pressure acts and a regulation spring is contained. The pressures below act on the membrane: • psup—corresponding to the overheating temperature. Its values grow up with the superheat. • pev—evaporation pressure (p4 in the figure). • pspr—spring pressure, settle at an appropriate value according to the desired static superheat. The resulting pressure acting on the membrane is psup − (pev + pspr). For a given refrigerant, to size and/or choice a thermostatic valve we need to know: • • • • •

Temperature and pressure, Tev and pev, of evaporator. Evaporator capacity. Condensation temperature and pressure, Tcond and pcond. Liquid temperature, Tl, at the valve inlet. Sum of pressure losses in the liquid line, distributor and evaporator Δpll.

The use of electronic expansion valves is becoming ever more common nowadays. In this type of valves the stem is controlled either by an electric motor (flow continuous modulation) or by a pulse controller, modulating the pulses duration (pulse flow modulation). They allow for a better flexibility then the traditional thermostatic ones with regard to: • MOP and, therefore, to the evaporator temperature. • Superheating, so reducing its value. • Possible injection into the evaporator of the optimum vapor flow at partial loads. This way, it is possible to keep instantaneously superheat at its minimum optimal value, thanks to the precision provided by the electronic control. In the case of continuous modulation, controller supplies a low voltage signal to the motor, capable of making rotor move either clockwise or anticlockwise. Pulse modulation provides proper windings with voltage pulses, axially moving a magnet connected with the valve stem. The valve can only work fully open or fully closed. Flow is regulated through pulses duration (see Fig. 2.28).

2.4

The Liquid Receiver

Generally a tank is placed at condenser outlet and upstream the expansion valve to store high pressure liquid leaving the condenser. It is sized to contain the whole refrigerant charge during the off-duty periods. Its purpose is to collect fluid when load fluctuations occur, thus allowing for flow rate modulation by the lamination valve.

2.4 The Liquid Receiver

45

Pulse control

Continuous control

flow rate

flow rate

time

time

Fig. 2.28 Control methods of electronic expansion valves

Fig. 2.29 Liquid receiver

to lamination

from condenser

A scheme of the receiver is illustrated in Fig. 2.29. It is a cylindrical steel tank with a pipe introducing the refrigerant coming from the condenser and an internal dip tube, ensuring that only 100% of liquid leaves the receiver. Of course this type of device is not used when a capillary tube is used instead an expansion valve, as no flow modulation is possible.

46

2.5

2 Types of Compression Heat Pumps and Their Main Components

Evaporator and Condenser

They can exchange heat with several types of indoor and outdoor sources. There is a widespread use of rather small systems (split systems), where both indoor and outdoor heat exchangers are cooled by air blown by properly sized axial fans. In this case we say we are using an air/air heat pump, in the sense than the refrigerant exchanges heat directly with air on both the heat exchangers. The internal source can also be water of a hydraulic system or sanitary water, while the external source can be water and even ground. With regard to the type of thermal sources heat pumps can be synthetically classified as: • Air/Air heat pumps, if both the sources are air. • Air/Water heat pumps, if the outer source is air and the inner one is water, as in water heating systems. • Water/Water heat pumps, if both the sources are water. In any case the first word refers to the outer source and the second one to the inner source. When the heat source/sink is air (air/air or air/water heat pumps) the air cooled heat exchanger mainly consists of a finned tube bundle with rectangular box headers on both ends of the tubes. Cooling air is provided by one or more fan. If refrigerant exchanges heat with water, plate and frame heat exchangers are used. They have corrugated metal plates to transfer heat between the fluids (Fig. 2.30). They may be welded, semiwelded and brazed (most commonly adopted in heat pumps). They are high heat transfer efficiency and compact10 heat exchangers.11 The plates are generally spaced by rubber sealing gaskets (Gasketed Plate Heat Exchangers GPHE) and are pressed to form troughs at right angles to the main direction of flow. Each fluid flows in gaps, each formed by two consecutive plates, 1.3–1.5 mm wide. Plates are compressed together in a rigid frame and form a set of parallel channels with alternating hot and cold fluids. They can easily be disassembled for cleaning and maintenance purposes as well as for inserting further elements. The plates can be also brazed (e.g., copper brazed) instead of welded, thus named Brazed Plate Heat Exchangers (BPHE). Plates are shaped to promote high levels of turbulence in order to increase heat transfer efficiency and self-cleaning.

10

A heat exchanger compactness is usually based on the value of two parameters: the hydraulic diameter DH ( the lowest of those employed for the two fluids) and the ratio between the heat transfer area and the volume where fluids flow S/V. The following definitions are given:

• Conventional heat exchangers for DH > 5 mm or S/V < 400 m2/m3. • Compact heat exchangers for 1 < DH(mm) < 5 or S/V > 400. 11 For a more detailed description of these heat exchanger the interested reader can also refer to [4].

2.5 Evaporator and Condenser

47

condensation up

refrigerant

water

The heat exchanger is formed by corrudated plates welded together. Channel are so created where fluids flow.

down

evaporation up

refrigerant

water

down Fig. 2.30 Scheme of a plate heat exchanger

vapor inlet

vapor to condensate

liquid inlet

cold to hot liquid

liquid exit

condensate discharge

Fig. 2.31 Scheme of a plate condenser. Vapor enters the larger duct then flowing through the gaps formed by the plates. It is collected by the two lower ducts. Cooling liquid flows in thermal contact with vapor

Some changes are adopted in two-phase applications to account for the difference between liquid and vapor specific volumes. Figure 2.31 shows a sketch of an Alfa-Laval condenser. Vapor enters a wider channel (to take into accounts its larger specific volume) and drops down through the plates toward two channels were condensed liquid flows.

48 Table 2.6 Size and weight of plate exchangers

2 Types of Compression Heat Pumps and Their Main Components Length (mm)

Width (mm)

Weight (kg)

194 306 613 …

33–101 34–298 62–470 …

1.3–2.8 2.5–16.6 15.3–84.8 …

Some features of BPHE for air/water heat pumps from HYDAC International are reported herein just to give some order of magnitude of the main parameters [5]: Some operating data: “Operating data Plate material Stainless steel 1.4401 (AISI 316). Braze material Copper (standard), Nickel. Pressures Copper brazed: max. 30 bar (test pressure 45 bar). Nickel braze: max. 10 bar. Use nickel-brazed plate heat exchangers with corrosive fluids: e.g., ammonia, sulphides and sulphates, deionizer or dematerialized water and other fluids on request. Temperature range up to +200 °C (freezing point and boiling point must be taken into consideration). Contamination: the quantity of particles in suspension should be less than 10 mg/l…” (Table 2.6). Capacities ranging from 0.7 to 186 kW are, for example, available for water cooled BPHE condensers and from 0.7 to 141 kW for evaporator, both using R410A, and from 0.7 to 176 kW and from 0.7 to 141 kW respectively for condensers and evaporators using R134a. Tube in tube heat exchangers are also employed, constituted by two coaxial tubes with the inner one corrugated, as shown in Fig. 2.32, to increase the heat transfer area and to promote turbulence. Both heat transfer coefficient and self-cleaning capability increase in so doing. Different heat transfer capacities are available. These heat exchangers are often spiral windings shaped to reduce the occupied room. Still to give some order of magnitude, few data related to this type of heat exchangers are supplied. For example Packless Industries [6] provides heat exchanger capacities in the range 1.76 kW (1/2 ton12)–105.5 kW(30 tons). The outer and inner tubes are, in general, respectively made by steel and copper. The volume of these spirally wound exchanger is evaluated as the one of a parallelepiped (like a box containing them) with sides a and b and height c. For lower capacities they may be a = 25 cm, b = 17 cm and c ranging from 8 to 10 cm. For higher capacities a = 64 cm, b = 51 cm, c = 56 cm or a = 90 cm, b = 31 cm, c = 31 cm. There are also the so called “trombone” heat exchangers, just shaped as

12

Ton (ton of refrigeration) is an Anglo-Saxon unit: 1 ton = 3.517 kW.

2.5 Evaporator and Condenser

49 liquid

helical corrugated heat transfer suface.

spiral condenser (Packless Industries) http://www.packless.com/products/condenser-heatpump-coils-ton.htm

Fig. 2.32 Tube in tube heat exchanger with corrugated surface. On the bottom a spiral condenser (coil)

cooling water exit

vapor inlet

vapor

liquid A liquid layer condenses on tubes and falls down.

cooling water inlet

condensate exit

Fig. 2.33 Shell and tube condenser

the musical instrument. The sides of the basis are larger than before, but c is in the order of 5 or 6 cm. Shell and tube heat exchangers are employed as well. Figure 2.33 shows the scheme of a flooded condenser. Cooling water flows in tubes and refrigerant in the shell. Vapor enters the shell and condenses, in contact with cold tubes,. As an example Alfa Laval, [6], supplies heat exchangers for R407C and R134a, cooled by

50

2 Types of Compression Heat Pumps and Their Main Components

Table 2.7 Evaporation data for R134a Mass flow rate (kg/m2s)

Temperature (°C)

Heat transfer coefficient (kW/m2K)

Vapor quality

200–900 400 400 400

−17.8 −18–5 / –

5–15 8.5–9.0 6.0–9.0 9.0

– – 0.05–0.2 >0.2

water coming from cooling towers, wells, rivers and lakes as well as from industrial processes, with condensing power between 60 and 1680 kW. Heat exchangers are sized for nominal requirements. So that, for instance, vapor exiting the compressor is taken to subcooled liquid to the expansion valve. When refrigerant flow rate is lower than the nominal value, with the same cooling water flow rate, the region of subcooled liquid refrigerant widens. As above mentioned, a liquid receiver is placed at condenser exit, aimed at collecting refrigerant in the case of maintenance and to allow for flow rate modulation, so that only liquid could flow into the lamination valve. Another tank can be located at evaporator exit (wet evaporators). As stressed before in dry evaporators only superheated vapor flows out of the heat exchangers. Thus, there is no danger for the compressor, but we lose heat transfer efficiency as variable temperature region of refrigerant exist in the final part of the evaporator. As a remedy to this a flooded evaporator is used, where the change of phase ends with no superheating. A gravity vapor separator is then placed before compressor suction. Once more we can mention some technical data for the related heat transfer coefficient ranging from 2 to 7 kW/m2K for a corrugated plate evaporator. Similar value for condensing micro finned tube with tube diameters up to 9.40 mm [7]. The following data (Table 2.7) for R134a can be roughly obtained by diagrams [8]. Air cooled heat exchangers are basically composed by finned tube bundles. Refrigerant flows in the tubes, often made of copper, and air is blown by fans. Air flow rate can be varied both stepwise (usually three steps) or continuously, by an inverter, according to the requested load. A typical basic scheme is shown Fig. 2.34. This is the most commonly adopted type for low capacity heat exchangers. Much more complex devices are employed for larger machines, anyhow working with the same operation principle. Whatever the size, these exchangers may undergo frost formation.

2.5.1

The Effect of Outside Air Humidity and Frosting

The external heat exchanger is generally sized referring to summer conditions, roughly with a temperature difference of 12–15 °C between the flowing refrigerant

2.5 Evaporator and Condenser

51

fins variable speed fan

air flow

tubes Fig. 2.34 Heat exchange coil

and the outdoor air. Therefore if the air temperature is 35 °C, condensation takes place at around 50 °C, and only sensible heat is exchanged. In winter air flowing to the fins, with an inlet temperature Ti and specific heat cpA, transfers a heat q (J/kg), equal to the difference between the inlet enthalpy, hi, and the outlet one, hu. Then we have: q ¼ hi  hu ¼ cpA ðTi  Tu Þ þ r 0 ðxi  xu Þ 0 Tu ¼ Ti  ½qr cðxpAi xu Þ se xi ¼ xu Tu ¼ Ti  cqpA From the above equations we infer, Ti and q being the same, the outlet air temperature Tu is lower for a pure sensible heat exchange (xu = xi) than for a transformation with a latent heat exchange (xu 6¼ xi). Figure 2.35 shows, for a given Ti = 5 °C and q = 10 kJ/kg, three transformations starting from different inlet relative humidities (RH): 90% (blue circles), 50% (red circles) and 40% (green circles). In the usual operating conditions we can affirm that no condensation occurs for a relative humidity lower than 50%, with such air temperature (5 °C). Example 2.3 To go into more details consider the case with an outdoor air temperature Ti = 5 °C. By extracting the numerical values from the graph of Fig. 3.32a, we can roughly say that xi = 2.8 g/kg and a dew temperature Td = −4.6 °C at a relative humidity RH = 50% and xi = 4.8 g/kg and Td = 3.0 °C with RH = 90%.

52

2 Types of Compression Heat Pumps and Their Main Components

Fig. 2.35 Transformations on an evaporating coil with different values of outdoor air humidity. Symbols triangle RH = 50%, ellipse RH = 90%, rhombus RH = 40%

10 8

5 0

6

x (g/kg)

i’(RH=90%) 4

u’ u

i(RH=50%)

u’’

i”(RH=40%)

2 0

-5

0

5

10

Temperature (°C)

Thus we get (cp,A = 1.0 kJ/kgK and r = 2500 kJ/kg) RH = 90%; hi’ = 17.0 kJ/kg RH = 50%; hi = 13.0 kJ/kg RH = 90%; hu’ = 7.0 kJ/kg RH = 50%; hu = 3.0 kJ/kg. With RH = 90% we respectively have a sensible and latent heat exchange equal to 2.0 and 8.0 kJ/kg, while at RH = 50% the heat exchange is essentially sensible.

The evaporation temperature is normally assumed 4 °C lower the outflowing cooling air so that the relative humidity improves COP at temperatures above that of defrost cycles’ start up(e.g., just 5 °C we referred to in the previous example). At lower temperatures the air humidity causes more demanding defrost cycles, decreasing the performance of a heat pump as much as higher its value is. Therefore a COP trend versus outside air temperature similar to the one sketched in Fig. 2.36 has to be expected. The red and the blue curves respectively refer to RH = 50% and RH = 90%, while the brown curve refers to an intermediate RH value. As it is clear from the above discussion, the knowledge of outlet air temperature is basic to know how far we are from frosting. The formula below can be used to this aim: Tu ¼ Ti  0:8

Pt  Pc V

2.5 Evaporator and Condenser

53

COP

Fig. 2.36 COP versus outdoor temperature with different relative humidities (RH)

RH>90%

RH≤50%

4

5

Te (°C)

With the following meaning of symbols: • Pt thermal power kW, according to Eurovent,13 • Pc compressor(s) power in kW, according to Eurovent. Eurovent certifies the total absorbed power PA; to obtain the compressor power we need to subtract the power of fan(s). • V volumetric cooling air flow rate in m3/s.

Example 2.4 Let us refer to a heat pump with the data below: Pt = 5.28 kW; PA = 1.64 k W; V = 2350 l/s; Fan power 0.12 kW. Consequently the cooling air temperature difference (outlet minus inlet temperatures) is 1.28 K. The normal reference conditions are: air temperature 7 °C and RH = 87%. With the given values Tu = 5.72 °C and the dew temperature can be evaluated as 4.6 °C.

In winter ice can freeze over, both on tube-fins and on tubes themselves, owing to outside air relative humidity and low temperature. This phenomenon takes place with an outdoor temperature even higher than 5 or 6 °C and a humidity exceeding 60%. At the very beginning a thin ice layer forms. At this stage the formed ice is a good thermal conductor, increases the heat exchange area, and lowers the flow 13 Eurovent is the Europe’s Industry Association for Indoor Climate (HVAC), Process Cooling, and Food Cold Chain Technologies. Its associates (more than a thousand companies) belong to Europe, the Middle East and Africa.

54

2 Types of Compression Heat Pumps and Their Main Components

Table 2.8 Defrost typical data Outdoor temperature (°C)

Relative humidity (%)

Duration (minutes)

0

70 80 90 100 70 80 90 100

220 100 50 30 220 100 50 30

5

section (gaps between fins) increasing air velocity, then, it enhances heat transfer rate. The additional ice layers forming afterwards are porous and contain air. Therefore they are insulating and deteriorate the heat exchange. Consequently evaporator efficiency increases at first, but dimishes afterwards. The insulating ice, thus, reduces heat pump performances. It is fundamental to take action at the right time and for a proper period to remove all the grown ice. The best way to set the time when starting defrost is to provide the heat pump with sensors detecting: air temperature, its flow rate trough the finned tubes and, at the same time, the pressure of the refrigerant. This way, defrost starts at the right moment and lasts just the suitable time, as defrost cycles effects on heat pump performances are far from be negligible. Just to give an idea, some data are provided, concerning the time interval between two consecutive defrosts, depending on outdoor temperature and relative humidity in Table 2.8. The table reports just indicative values of this time intervals, and the actual ones should be set up on a case-by-case basis. The defrosting technique may consist either in an electric resistance which switches on when fin temperature approaches 0 °C14 or, more commonly, thermal cycle reversal. This means that the unit switches over to the cooling mode and the outdoor coil (evaporator) becomes the hot condenser. In doing so, some discomfort to users is caused. The process takes place according to the following stages: • switch off of the outdoor coil fan, through a dedicated relay. • Cycle reversing valve switching to the cooling mode. • Switch on of an auxiliary heating source for the indoor environment, if available. In any case an amount of ice that reduces the cooling air flow rate more than 50% of the nominal value is not acceptable, as it might impair the compressor. On one hand equipment safety would suggest frequent defrosts, but economy and machine viability require performing few defrosting cycles.

Or at a certain level of obstruction of the fin gap, due to frost formation.

14

2.5 Evaporator and Condenser

55

Several aspects have to be accounted for to optimize defrost start up. • A first control can be performed on air pressure within the coil, (differential pressure between inlet and outlet). When this value exceeds a given set point value, the process starts. This type of systems reacts to low pressure difference, so that they might be activated by a wind burst. Therefore a time delay has to be introduced to verify the permanency of such a pressure drop increase, before starting the defrost. Even debris and leaves can cause an improper system action. • It is also possible to refer to temperature differences. This method is based on the fact that the usual temperature difference between outdoor and evaporator temperatures varies in the range 3–9 °C. As ice builds up, this difference increases. Defrost starts when a set threshold is exceeded. • As both the afore said methods revealed not to be always reliable, and also somehow costly, at first many manufacturers decided to use a timer in residential applications. So the process started at given time intervals. This is a very simple method and was the most widespread at least as long as electronics was introduced. The timer was coupled to a thermostat measuring air temperature at coil exit generally set at 3 °C. If air is cooler than this for a given period, say 30 min, defrost starts, otherwise it does not. Defrost ends when evaporator temperature achieves a preset value by manufacturers. This method, timer + thermostat, is the most used in residential buildings also because of its low cost. A further control based on the pressure difference mentioned above may be added so that defrost startup is also influenced by this, when the related increase is about 100 Pa. An additional method consists in injecting superheated vapor from the compressor into the evaporator through a dedicated defrosting valve. This is aimed at preventing the indoor environment from being cooled, even if some power is anyway subtracted from it. Other solutions use tanks where thermal energy can be stored and then released to coils, as those using melting salts or ethylene glycol exchanging heat with the condenser de-superheating stage. In general defrost can affect energy consumption by more than 10%, depending on the adopted solution. In case of absorption heat pumps some hot fluid (e.g., ammonia), coming from the generator, can be diverted to the outdoor coil, without any cycle reversal. Such a reversal does not even occur in endothermic engine driven heat pumps. Before ending this paragraph, some features of this type of coils are provided, see Table 2.9.

2.6

Economizer and Vapor Injection

A way to save energy in the case of a large temperature difference between thermal sources, i.e., large pressure ratio, is injecting vapor into the compressor at an intermediate pressure.

56

2 Types of Compression Heat Pumps and Their Main Components

Table 2.9 Heat pump coils (R410A) for residential use Nominal cooling power (1) (min/max) kW Nominal power input (1) E.E.R. (1) W/W E.S.E.E.R. W/W Nominal cooling power (2) (min/max) kW Nominal power input (3) E.E.R. (2) W/W Nominal heating power (3) (min/max) kW Nominal power input (3) C.O.P. (3) W/W Nominal heating power (4) (min/max) kW Nominal power input (4) C.O.P. (4) W/W 1.

4.13 6.49 8.20 10.51 (1.80/5.00) (3.00/8.20) (3.70/10.0) (4.0/13.10) kW 1.33 2.08 2.65 3.39 3.11 3.12 3.10 3.10 3.43 3.49 3.41 3.48 5.72 8.93 12.36 14.00 (2.30/0.20) (3.70/0.90) (4.60/13.20) (6.00/16.00) kW 1.44 2.27 2.98 3.64 3.98 3.93 4.15 3.85 5.48 8.43 11.81 13.38 (2.10/0.80) (3.50/9.30) (4.40/12.60) (5.60/14.80) kW 1.65 2.55 3.45 4.13 3.32 3.30 3.42 3.24 5.77 9.06 12.40 14.16 (2.40/6.50) (4.00/10.0) (4.70/13.40) (6.30/16.40) kW 1.39 2.21 2.95 3.45 4.15 4.11 4.21 4.15 Cooling: outdoor air temperature 35 °C; inlet/outlet water temperature 12/7 °C 2. Cooling: outdoor air temperature 35 °C; inlet/outlet water temperature 23/18 °CC 3. Heating: outdoor air temperature 7 °C d.b. 6 °C w.b.; inlet/outlet water temperature 40/45 °C 4. Heating: outdoor air temperature 7 °C d.b. 6 °C w.b.; inlet/ outlet water temperature 30/35 °C Note Data declared according to UNI EN 14511:2011. The performance data shown in the table refer to units without options and/or accessories and could be subject to change. Attention: for antifreeze unit version, for lowest ambient temperature 5 °C, you must add a suitable quantity of antifreeze additives

In Fig. 2.37, the classical cycle is drawn in the pressure enthalpy plane, with dotted lines (points marked by capital letters), while that with continuous lines (point marked by numbers) represents a cycle with vapor injection. At the end of condensation (4!5), the refrigerant is sent to a first expansion valve (5!6), at its exit vapor15 is separated from liquid in a separator, at pressure p7, and forwarded to the compressor (6!3) at the beginning of the second stage of compression. Liquid goes to a second expansion valve to enter the evaporator.

Dry vapor from separator mixes with vapor coming from the first compression stage, point 3 in the figure.

15

2.6 Economizer and Vapor Injection

57

p

4

mC = mE + mi

5

(C)

(B) mi

7

6

6

(D)

mE

8

1

3

3

2

vapor

(A) 6

h

liquid

7

Fig. 2.37 Cycle with vapor injection

The following relations hold: QC ¼ mC ðh5  h4 Þ

Q E ¼ m E ð h8  h1 Þ L1;3

mC ¼ mE þ mi ¼ mE ðh1  h2 Þ

L3;4 ¼ mC ðh3  h4 Þ COP ¼

QC ð h5  h4 Þ ¼ L1;3 þ L3;4 ðh3  h4 Þ þ mmCE ðh1  h2 Þ mi mE h3 ¼ h6 þ h2 mC mC

where subscripts C, E and i respectively indicate: condenser, evaporator and injection. Vapor injection produces an intermediate cooling that lowers the work of compression. The reduction of compression work can be easily evaluated by comparing the single compression stage cycle with the one obtained by vapor injection. Fluid exiting from the condenser can also undergo a double expansion, a first one between the whole cycle pressure difference (between condenser and evaporator) and a second one between the condenser pressure and an intermediate value, as sketched in Fig. 2.38. A given amount of refrigerant (primary fluid) flows to a first lamination valve (path 5-7) after passing through a heat exchanger, economizer, and goes to evaporator. In this heat exchanger primary fluid transfers heat to another amount of fluid flowing (secondary fluid) along path 5-6-3’. This secondary fluid is laminated to an intermediate pressure in a second valve and reaches the

58

2 Types of Compression Heat Pumps and Their Main Components economizer 4

5 condenser

vapor injection

3’

3 compressor

6 expansion valves evaporator

7

1

Fig. 2.38 Cycle with economizer

p

5

(C)

mC = mE + m i

4

mi

3

(B)

2

6 (D) 7

mE

1 (A)

h Fig. 2.39 Cycle with economizer in the pressure-enthalpy plane

compressor as superheated vapor. This way, primary fluid is further subcooled and secondary fluid heats up, then mixing with vapor of the first compression stage (point 3). Figure 2.39 shows the related cycle in the pressure-enthalpy plane. Still to make an example: for a 8 kW heating power and 6 kW cooling power heat pump, using R407C, with condensation temperature and pressure of 50 °C and 22 bar and evaporation at −7 °C and 4 bar Copeland [9] provides a scroll compressor (Model ZH09KVE-TFD) with a total flow rate of 29.7 g/s and a vapor injection flow rate of 9.70 g/s, at an intermediate pressure of 5.97 bar. Vapor

2.6 Economizer and Vapor Injection

59

injection is also employed in screw compressor with a COP claimed increase around 20%. This technology is also implemented for sanitary water supply at about 50 °C, even with outdoor temperatures below 0 °C. In addition scroll compressors exist on the market which can bear liquid, thus allowing for saturated vapor injection, named wet vapor injection, instead of dry or superheated vapor. In so doing the operation range of compressors can be widen, but it is generally fixed a top value of time duration of such a type of injection (e.g., 2000 h). The use of wet vapor injection is aimed at limiting the discharge temperature, so that it does not exceed a safety value, say around 140 °C. The effect of vapor injection on compressor operating range has been already shown in Fig. 2.25.

2.7

The Four Way Reversing Valve

The scheme of a cycle reversing valve for reversible heat pumps is drawn in Figs. 2.40 and 2.41. Four ports are placed on it. On top we have the port where compressor discharge fluid flows in, high pressure port. Of the three ports placed on

Winter

compressor discharge

s

from evaporator to condenser to compressor suction

indoor Fig. 2.40 Reversing valve configuration in winter operation mode

outdoor

60

2 Types of Compression Heat Pumps and Their Main Components

Summer

compressor discharge

s

from evaporator

to condenser to compressor suction

indoor

outdoor

Fig. 2.41 Reversing valve configuration in summer operation mode

the lower side the central one send fluid to the compressor suction. Cycle inversion is obtained by a slide S that puts into contact these ports by twos, moving right and left. Its movement is caused by the refrigerant itself flowing through dedicated capillary tubes. This flow is controlled by a valve activated by an electric coil. If the valve coil is fed, the winter mode of operation is active (Fig. 2.40), while, when it is not, the summer mode takes place. This is done both for seasonal operation change and for defrosting.

2.8

Engine Driven Heat Pumps (GHP)

In this type of heat pumps, usually addressed as GHP (Gas Heat Pumps), the compressor is driven by a gas engine, instead of the more commonly used electric motor. Beyond the mechanical work delivered to the compressor these machines recover the engine exhaust heat according two different ways: direct and indirect heat recovering.

2.8 Engine Driven Heat Pumps (GHP) QFC

TC 3

61 QMC

TC 2

condenser

heat exchanger B heat exchanger A QMC engine

QM

L evaporator

4 TF

ηPQM

1

QM

QF T

2

T QFC

3

QMC

TC

ηPQM

TC QF

TF 4

1

s

s

Fig. 2.42 Engine driven heat pump (GHP)

The direct heat recovering uses the engine cooling water either for indoor environment heating or for sanitary water production. Figure 2.42 shows a typical scheme with a diesel engine. The used symbols have the following meaning: TC TF TM QFC QF QM QMC

Hot heat source temperature Cold source temperature Equivalent temperature of “engine source” Amount of heat released by heat pump to cold source Heat exchanged with the cold heat source Heat supplied to engine by combustion process, TM Heat supplied to hot source by regenerator

The system, heat pump plus engine, interacts with three thermal sources. In fact, at its boundary, it only exchanges heat, while work L is an internal mechanical exchange between engine and compressor. This work is related to the supplied combustion heat, through the engine thermodynamic efficiency ηM. In other words: jLj ¼ jQFC j  QF ¼ gM QM If eR is the efficiency of exhaust heat recovery system, i.e., the amount, (1 − ηM) QM, of heat that can be recovered by cooling the engine we obtain that

62

2 Types of Compression Heat Pumps and Their Main Components

TM QM

ideal engine

QMC

TC ideal heat pump

QFC

QF TF

Fig. 2.43 Basic scheme of a GHP with direct heat recovery system

QMC = eR(1 − ηM)QM, while the heat released to outdoor environment is (1 − eR) (1 − ηM)QM. Thus the heat exchanged with the hot source is: Qc ¼ QFC þ QMC ¼ QFC þ eR ð1  gM ÞQM where the second term on the right hand side of the above equation can be used to produce hot sanitary water, in summer. Therefore QC is the “useful” heat we can obtain from an engine driven heat pump in winter. By applying the first and second Principle of Thermodynamic for ideal conditions16 (Carnot cycle), we obtain: 8
TF) and refrigerant follows the cycle 1234 instead of the initial 1’234’, increasing the heat exchanged by the evaporator and lowering the compression power. In a fully reversible condition (referring to Fig. 2.44) we have: QM ¼ L þ jQMF j Qev ¼ QF þ QMF where QMF is the power recovered by cooling the engine (>0), QF the one obtained by the cold source (>0) and Qev the power supplied to the evaporator (0), QF the g from the cold source(>0) and Qev the supplied to the evaporator ( Te*), the heat pump capacity is not adequate to supply the required power on the left of the balance point. A backup system must be used either replacing the heat pump (say a condensation boiler) or adding further power to that of the heat pump. In the former case the heat pump stops working at the balance point, B,2 and the backup generator starts operating according to the dotted line in Fig. 4.1b providing the full load required. In this case we speak about “alternating operation”. 1

If we size the heat pump at the maximum required power, it is advisable to use an additional storage tank to avoid too frequent transients. The storage inertial volume is usually calculated as the difference between an “adequate volume”, in the range 8–13 L per kW (for an effective design it is always the case to deal with the heat pump manufacturer) and the volume of the hydraulic facility. In the case of large water contents (e.g., radiant floors) this additional volume may be not necessary. 2 The International Standard (in Italy UNI-TS-11300/4, 2012) defines the balance (or bivalence) point as the point where a heat pump stops operating with a load factor equal to 1.

4.1 Full Load and Partial Load Operation, the Balance Point

91

In the latter case an additional power source takes action only supplying the difference between the user’s power requirement and the power given by the heat pump, down to temperature Tp (Fig. 4.1c). This is the “parallel operation”. It is necessary to analyze the trend of the outdoor temperature of the location we are interested in. We need to know how many times each value occurs during the period we are referring to. For instance we could learn that 0 °C occurs for 40 h, 4 °C for 300 h and so on, in winter. To better clarify this, the following example may be used. Example 4.1 Let us suppose that a given place has an outdoor reference temperature equal to Te* (in Italy commonly included in the range 0, −5 °C and anyway obtainable through the National Standards) and a reference load Q* in correspondence of the above temperature. In addition let us suppose this temperature occur for a percent p* of the heating season. If P is the heating period in hours, a temperature Te lower or equal to Te* (Te
Te*) takes place during a time interval p*P. We refer to Fig. 4.2 where the cumulative curve p = f(Te) is represented. Temperature Te is on the abscissa and its percent of occurrence, p, in ordinate. For instance at a given pk correspond temperatures Te
Te,k. We can divide the seasonal temperature variation TH in n temperature sub-intervals T = TH/n. If at time tk Te = Te,k, the outdoor temperature will

Fig. 4.2 Cumulative curve of the occurrence of different temperatures versus their percent of occurrence

92

4 Operating Conditions

reach the value Te,k+1 = Te,k + T at a time tk+1, with tk+1 = tk + pk. Therefore the probability of having an outdoor temperature in the range Te,k  Te,k+1 is given by pk+1 − pk = pk/P.3 The power to be supplied to the user is supposed to be proportional to the difference between the indoor temperature Ti and the outdoor one. Thus the energy to be supplied is proportional to (pk+1 − pk)P(Ti − Tm,k), where Tm,k is the average outdoor temperature occurring in the afore mentioned period calculated as follows: Tm;k ¼

1 pk

2t

tkZþ pk

Zþ pk

Ztk

k

Te ðtÞdt ¼ tk

14 pk

Te dt t0

3 Te dt5

t0

and t k ¼ pk P tk þ pk ¼ pk P þ ðpk þ 1  pk ÞP ¼ pk þ 1 P dt ¼ Pdp 2 3 2 3 pZ pZ k þ 1P k þ 1P Zpk P Zpk P P6 1 7 6 7 Tm;k ¼ 4 Te dp Te dp5 ¼ Te dp Te dp5 4 pk pk þ 1  pk p0 P

p0 P

p0 P

p0 P

where the quantity within brackets corresponds to the shaded area in Fig. 4.2. If the function f(Te), can be linearized in the interval we get: Tm;k ¼

Tk þ Tk þ 1 T ¼ Tk þ 2 2

Of course to do this it is appropriate to choose a rather small T, usually 1 °C as said before. The energy to be supplied is therefore proportional to (black   P area in the figure): ðpk þ 1  pk ÞP Ti  Tk  T2 If Tk = T0 + kT this energy is proportional to:   T ðpk þ 1  pk ÞP Ti  T0  kT  2 n     X 1 ¼ ðpk þ 1  pk ÞP Ti  T0  k þ T 2 n

X

3

As a reference time interval, the already quoted Standard (11300-4/2012) refers to the month or a shorter period, named bin, for evaluating the employed energy. Bin is the time interval where the outdoor temperature changes by 1 °C (H=1 °C and pk to be determined consequently). In this example we only consider one day as seasonal period, for the sake of simplicity.

outdoor temperature (°C)

4.1 Full Load and Partial Load Operation, the Balance Point

hour of the day

Fig. 4.3 Trend of outdoor temperature of the reference day

Just to make an example we could suppose to have an outdoor temperature during the daily heating period as the one given in Fig. 4.3. We could refer to the typical day of a month, as usually provided by the National Standards, and to consider a fan-coil heating system of an office with a working time ranging from 9 a.m. to 5 p.m. The office has a floor area of 400 m2, an overall equivalent heat transfer coefficient UeqS = 300 W/K a ventilation hourly flow rate equal to 1.5 V (V is the office volume) plus the contribution of thermal bridges. In conclusion the total heat exchanged is: Q ¼ 345  ðTi  Te Þ þ 402  ðTi  Te Þ ¼ 702  ðTi  Te ÞW

Requested power (W)

This trend is represented in Fig. 4.4, where the shaded area corresponds to the required working time. The related cumulative curve is shown in Fig. 4.5 for a heating time of 9 h (to account for further activity to be performed in the office) and with T = 1 °C. 14000 12000 10000 8000 6000 4000 2000 0 0

1

2

3

4

5

6

7

8

9 10 11 12 13 14 15 16 17 18 19

outdoor temperature (°C)

Fig. 4.4 User’s load curve

93

94

4 Operating Conditions

Fig. 4.5 Cumulative curve of outdoor temperature occurrence. Ti-Te is also reported to give an idea of the power supplied

The reference indoor temperature (also in the graph) has been assumed 18 °C to account for some energy gains. Doing this (considering energy gains) it is actually appropriate to evaluate the value of the energy required, but it is not precautionary to evaluate the power to be installed. Besides Fig. 4.6 gives the instantaneous power to be provided to the user.

8000 7000 6000 5000

W 4000 3000 2000 1000 0 9

10

11

12

13

14

hour of the day

Fig. 4.6 Instantaneous power to supply to the user

15

16

17

18

4.1 Full Load and Partial Load Operation, the Balance Point

95

Table 4.1 Reference temperatures to which manufacturers have to refer to provide performance data for heating use or mixed use heating and sanitary water production (Tp,c hot sink temperature in °C) Temp. (°C) cold source Air −7 Water

2

5 Ground −5 0

Tp,c air heating

Tp,c water heating

Tp,c sanitary water

7

12

20

35

45

55

45

55

10

15

20

35

45

55

45

55

5

10

20

35

45

55

45

55

Of course to have a comprehensive evaluation of heat pumps performance the behaviour of COP4 it is necessary to examine its operation at partial load. To this purpose manufacturers should provide several kind of information concerning: • Type of heat pump, also specifying whether it is reversible (heating and cooling) or not. • Intended use. e.g., only heating (cooling), heating and cooling, mixed heating and sanitary water, only sanitary water. • Type of operation: on-off, modulating. • Heat pump thermal sources. • Furthermore manufacturers must provide the full load performances in correspondence of different sources temperatures (EN 14825 and TS 11300). The Standards give such conditions, reported in Tables 4.1 and 4.2 for sanitary water production. • Performances at a climatic part load ratio, PLR, other than 1. PLR is supposed to go to zero at an outdoor temperature equal to 16 °C and is defined as:

PLR ¼

Te  16 Tep  16

where Te and Tpe respectively are the actual outdoor temperature and the reference outdoor temperature, employed in the system design. For example for Te = 8 °C and TPe = 0 °C, PLR = 0.5. Moreover a “second principle efficiency”, η″, is defined as the ratio between COP in the actual conditions and the ideal one at the same source temperatures.

4

(and/or the significant parameter(s) to be considered for the type of heat pump we are using and the working period we are referring to in the actual use, as stressed in the next paragraph).

96

4 Operating Conditions

Table 4.2 Reference time interval Dtk for energy evaluation Cold source

Hot sink Air(1)

Outdoor air

Monthly bins Monthly bins Month

Water at constant temperature

Water at variable temperature

Monthly bins

Monthly bins

Indoor air (recovery) temperature Monthly bins Monthly bins depending on climatic conditions Indoor air (recovery) temperature Month Month independent of climatic conditions Perturbed soil/rock by climatic Month Month Month conditions Unperturbed soil/rock by climatic Month Month Month conditions Sea, river and lake water Month Month Month Waste water and sewage from Month Month Month technological processes Urban wastewater Month Month Month 1. A constant set point temperature is assumed 2. Constant or variable temperature is referred to the heating fluid in the heat generator during the considered time interval. For example we refer to constant temperature once the user is fed by a mixing valve, and to variable temperature when the user is directly fed by a variable temperature generator (it commonly occurs in the heating period) Note monthly bins refer to outdoor air temperature

  T c  Tf g ¼ COP  Tc 00

where, as usual, Tc and Tf are the temperatures of the hot and cold sources and Tc is the hot source temperature in K. Once data by manufacturers are available heat pumps performances can be calculated in a reliable manner at different load factors, CR (ratio between the actual power and the nominal one). An example of data provided by the manufacturer is given below for an air/water heat pump with an external design temperature equal to −10 °C and a correspondent supplied power of 5.7 kW, employed for radiant floor. Inlet temperature to radiant floor Tin = 35 °C Te (°C) PLR Full load supplied (kW) at the given source temperatures CR Power required by user (kW)

−7 0.88 5.05 1 5.0

2 0.54 6.22 0.49 3.0

7 0.35 7.3 0.27 1.98

12 0.15 8.18 0.11 0.88 (continued)

4.1 Full Load and Partial Load Operation, the Balance Point

97

(continued) Inlet temperature to radiant floor Tin = 35 °C COP COP(CR = 1) Correction factor f = COP(CR = 1)/COP

3.14 3.14 0.97

5.39 3.91 1.38

6.68 4.51 1.48

4.37 5.38 0.81

Power and COP(CR = 1) at full load for various Tin.

−7 2 7 12

Full load (kW) Tin = 45 Tin = 35

Tin = 55

COP(CR = 1) Tin = 35 Tin = 45

Tin = 55

5.05 6.22 7.30 8.18

4.62 5.44 6.37 7.42

3.14 3.86 4.59 5.38

2.16 2.35 2.76 3.40

4.91 5.80 6.80 7.94

2.54 2.97 3.47 4.09

Power and COP(CR = 1) at full load for production of sanitary water.

7 15 20 35

Full load (kW) Tin = 55 °C

COP(CR = 1)

6.37 7.90 8.54 8.54

2.76 3.59 3.85 3.81

Parameter η″ (second principle efficiency) is calculated in correspondence of the available data (sources temperatures and COP) and evaluated by linear interpolation for intermediate conditions. If the above data are not provided, Standard EN14825/2012 gives the following equation to estimate COP. Heat pumps air/air, antifreeze/air, water/air: COP ¼ COPðCR ¼ 1Þ  ½1  Cd ð1  CRÞ If data are not known Cd = 0.25 (ground/air included) Heat pumps air/water, antifreeze/water, water/water: COP ¼ COPðCR ¼ 1Þ 

CR ½Cc CR þ ð1  Cc Þ

If data is not known Cc = 0.9 (ground/water included). For heat pump with inverter no correction is applied in the range of CR from 1 to 0.5. η″ is calculated at the conditions referred to the available data and kept constant in the remaining range. While addressing the interested reader to the above quoted standard for better details, we refer to Example 4.1 to make an additional one for the sake of clarity.

98

4 Operating Conditions

Example 4.2 Let us refer to an air/water heat pump, i.e., the heat sources are respectively outdoor air and water of the heating system. The heat pump has been installed several years ago and there are no manufacturer data available other than the values of the nominal power and COP and the related source temperatures. Consider the following values: Nominal power 7.5 kW. Nominal COP 3.2. Outdoor air reference temperature 2 °C. Water reference temperature 45 °C. For a more accurate estimation we should have data at the outdoor temperatures −7, 2, 7, 12 °C. As we have not, we can follow the procedure below. We first evaluate the ideal COP at the reference temperatures: COPid ¼

1 ¼ 7:40 2 þ 273:16 1  45 þ 273:16

Besides, by changing the outdoor air temperature we, for instance, obtain the following set of values: T(°C) COPid

7 8.37

8 8.60

9 8.84

10 9.09

11 9.36

12 9.64

13 9.94

14 10.26

With reference to nominal conditions the efficiency η″ is: g00cos t ¼

COPcos t 3:2 ¼ 0:43 ¼ 7:4 COPid

The dependence of COP versus CR is given by (Cc = 0.9): COPðCRÞ ¼ COPðCR ¼ 1Þ 

CR 0:9CR þ 0:1

where CR is the load factor. As we do not have further data we suppose η″ to keep constant whatever the load factor is. COP(CR = 1) at the different outdoor temperatures is: T(°C) COP(CR = 1)

7 3.60

8 3.70

9 3.80

10 3.91

11 4.02

12 4.15

13 4.28

14 4.41

On the basis of the user’s energy demand, the power required is of 8 kW at the office opening, then 0.5 kW higher than the one available by the heat pump. On the other hand the cumulative curve outlines that the outdoor temperature keeps below 8 °C for around half a hour and can be estimated,

4.1 Full Load and Partial Load Operation, the Balance Point

99

for this period, around 7.5 °C. The consequent energy, during this period, required is 4 kWh, while the energy supplied by the heat pump at full load in the same period (7.5  0.5 = 3.75 kWh) is practically sufficient. Anyway, to be sure, we can also use an integration electric resistance. It is now clear that the balance point was considered at an outdoor temperature of 8 °C, at the planning stage. At the end of the working time this temperature is 11 °C and its top value 13 °C. In correspondence of this latter value the load factor achieves its minimum CR = 0.48. So it varies in the range 1–0.48. Its hourly trend can be obtained and is shown in Fig. 4.7. The trend of COP is additionally drawn in Fig. 4.8. In this same figure COP (CR = 1) is reported, that is the value of 1,00 0,90 0,80 0,70

CR

0,60 0,50 0,40 0,30 0,20 0,10 0,00 9

10

11

12

13

14

15

16

17

18

hour of the day

Fig. 4.7 Trend of load factor

COP(CR=1)

COP

5,00 4,00 3,00 2,00 1,00 0,00 9

10

11

12

13

14

15

16

17

18

hour of the day

Fig. 4.8 Trends of COP. The curve designated by COP (square symbols) is related to a constant power heat pump working on an on-off basis. The curve designated by COP (CR = 1) (rhomb symbols) beside corresponds to a continuously modulating heat pump

100

4 Operating Conditions

COP the heat pump would have working at the full load each hour. In this case there is not a big difference between the two curves because the load factor is not lower than 0.5. For lower CRs this difference is much more pronounced.

Furthermore we show the trends of f = COP/COP (CR = 1) for air/air and air/water heat pumps respectively in Figs. 4.9 and 4.10. The on-off operation is indicated by a continuous line, while symbols represent modulating operation starting from 30% of the full load. In Example 4.3 we also report the behaviour of a modulating heat pump. For absorption heat pumps directly fed by a fuel (direct gas fired machines) the effect of CR is evaluated as follows. As better specified in the next paragraph the coefficient of performance is now named GUE (Gas Utilization Efficiency), because these heat pumps directly use a primary source (fuel) instead of electricity. The ideal GUE and η″ are expressed as;   Tc Tgen  Tf GUEid ¼ Tgen Tc  Tf GUE g00 ¼ GUEid Subscript “gen” means generator.

on/off

modulante modulating

1,1

COP/COP(CR=1)

1 0,9 0,8 0,7 0,6 0,5 0

0,1

0,2

0,3

0,4

0,5

0,6

CR

Fig. 4.9 f = COP/COP(CR) versus CR for an air/air heat pump

0,7

0,8

0,9

1

4.1 Full Load and Partial Load Operation, the Balance Point on/off

101

modulante finotoal30% 30% Modulating

1,1

COP/COP(CR=1)

1 0,9 0,8 0,7 0,6 0,5 0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

0,9

1

CR

Fig. 4.10 f = COP/COP(CR) versus CR for an air/water heat pump

The relation between GUE and its value at CR = 1 is: GUEðCRÞ ¼ Cd GUEðCR ¼ 1Þ Cd ¼ f ðCRÞ In this case too a distinction is made between on-off and modulating operations, adopting two different trends for Cd, as shown in Fig. 4.11. In fact it is possible to adjust the temperature of the water supplied to users on the basis of an appropriate load curve as can be commonly done in modern boilers. It is clear from the figure

on-off

modulating

1,1 1

Cd

0,9 0,8 0,7 0,6 0

0,1

0,2

0,3

0,4

0,5

CR

Fig. 4.11 Cd versus CR for an absorption heat pump

0,6

0,7

0,8

0,9

1

102

4 Operating Conditions

how direct gas fired heat pumps are poorly influenced from the type of operation; for CR ranging from 0.4 to 1 Cd varies from 0.9 to 1. It is the case to remark that there is also the possibility (in addition to the use of a backup system) of splitting the full power of a heat pump into two or more units in cascade. Furthermore we recall that classes of energy quality (from A to G) exist also for heat pumps, as for other domestic appliances. For example, air/air split systems to be included in class A have to meet the requirements: COP > 3.60 and EER > 3.20. In the end we want to stress that the above treated procedure has some obvious limitations: • defrost is not accounted for, so that COP is generally overestimated for air cooled machines. • The presence of inertial water tanks that could allow for a better use also of on-off heat pumps is not taken into consideration.

Example 4.3 The following data are provided for a modulating heat pump with reference outdoor design temperature, Te = −10 °C. Hot sink temperature 35 °C Te PLR Full load power CR User’s power requirement COP1 = COP part load COP2 = COP full load fcop = COP1/COP2 η″ COPideal

−10 – – >1 5.7 – – – – –

−7 0.88 5.05 1 5 3.14 3.14 0.97 0.43 7.34

2 0.54 6.22 0.49 3 5.39 3.91 1.38 0.42 9.34

7 0.35 7.30 0.27 1.98 6.68 4.51 1.48 0.41 11.01

12 0.15 8.18 0.11 0.88 4.37 5.38 0.81 0.40 13.40

The balance point is at Te = −7 °C. The curves of the requested power (dotted curve) and of the power supplied by the heat pump (continuous line) are reported in Fig. 4.12. If the heat pump were not modulating the trends of COP were as in Fig. 4.13. Otherwise, if modulating, it keeps COP = COP (CR = 1) within the whole modulating range, shaded area in Fig. 4.12. In this case manufacturers provide COP at partial load (COP1 in table) corresponding to the modulation range. The deviation with respect to COPmax (COP2 in table) is given by a correction factor fCOP, of course depending on the type of machine. It is as if the second principle efficiency, η″, became η* = η″/fCOP in the modulation zone. The trends of ideal COP, COPid, COP1 and COP2 are shown in Fig. 4.14.

4.1 Full Load and Partial Load Operation, the Balance Point Power supplied by HP

103

Requested power

9 8

5,38 kW

4,51

7

3,91

6

COP=3,14

5 4

5,39 4,37

3

6,68

2

-12 -11 -10 -9 -8

-7

-6

-5

1 Modulation zone 0 -4 -3 -2 -1 0 1 2 3

4

5

6

7

8

9

10 11 12

Te (°C)

Fig. 4.12 Trends of power demand by user (rhomb symbols), of power supplied by HP (square symbols) and of power supplied by HP in the modulating region (no symbols and shaded area)

COP non modulating HP

COPmax

6 5,5 5

COP

4,5 4 3,5

CR=0,49

3

CR=0,27

CR=1

CR=0,11

2,5 2

-8

-7

-6

-5

-4

-3

-2

-1

0

1

2

3

4

5

6

7

8

9

10 11

12

Te (°C)

Fig. 4.13 Trend of COP for non-modulating HP

Figure 4.15 reports the trends of η″ and η* for 8 examined modulating heat pumps together with their average values (circles for η* and rhombs for η″), just to give an idea of the trends.

13

104

4 Operating Conditions COP ideal

COP1 (partial load)

COP2 (full load)

14,00 12,00 10,00

COP

8,00 6,00 4,00 2,00 0,00

-9

-8

-7

-6

-5

-4

-3

-2

-1

0

1

2

3

4

5

6

7

8

9

10 11 12 13 14

Te (°C)

Fig. 4.14 Trends of COPs

Fig. 4.15 Trends of η″ and η* for 8 modulating heat pumps. The solid curves refer to their average values (rhombs for η″ and circles for η*)

4.2

Comparison Among the Different Types of Heat Pump

We want to remark how the evaluation of the real performances of a heat pump needs to take into consideration several aspects, in many respects already treated in this text. Thus the energy used by heat pumps auxiliaries has to be taken into account as well as the effect of sources temperature on volumetric compressors. In fact the latter affect the specific volume of vapor modifying the value of mass flow rate at a given volumetric flow rate. This occurs in particular for evaporators, where the specific volume increases with decreasing the evaporation temperature and the mass

4.2 Comparison Among the Different Types of Heat Pump

105

flow rate reduces. In addition the air humidity has to be accounted for as it affects the heat exchangers’ performance and the defrost cycles. An appropriate parameter to compare the performances of the different types of heat pumps has been defined and named P.E.R., Primary Energy Ratio. This is the ratio between the useful energy, QUT, delivered to users and the employed primary energy, QEP. For boilers this ratio is larger than unit only for condensation boilers (referring to the lower calorific value), where it reaches the value of 1.11. For heat pumps the following considerations hold. Electric heat pumps (EHP)—The primary energy is supplied by the electric power stations and delivered to heat pumps through the electric power lines system. Therefore it is influenced by the average equivalent efficiency of the system of available power stations, ηCE, and by the equivalent efficiency of the power lines system, eTR, thus, we have: power to compressor L ¼ gMC  gCE  eTR QEP ¼ gMC Eel QEP total efficiency ofelectric system Eel ¼ gCE  eTR PEREHP ¼

QUT QC ¼ gMC Eel  COPEHP ¼ gMC Eel  QEP L

The term ηMC is the efficiency of the set of devices installed between the points where the electric power is delivered to the work L the compressor delivers to the fluid. It includes the efficiencies of the electric motor, of an eventual mechanical coupling (motor-compressor), of the inverter and possible other auxiliaries. Example 4.4 Suppose ηMC = 0.9, the trend of PER versus COP for different values of Eel is reported in Fig. 4.16. This same figure shows the case when COP corresponds to a PER larger than unit. The figure also reports the values of PER for a traditional boiler with an efficiency of 0.9 and for a condensation boiler with an efficiency of 1.11.

Gas heat pumps (GHP)—By assuming QM = QEP it follows that: PERGHP ¼

QUT ¼ COPGHP QEP

106

4 Operating Conditions 2,5

Eel=0,37 Eel=0.37 Eel=0.40 Eel=0,40 Eel=0.46 Eel=0,46 caldaia rend.=0,9 boiler eff. =0.9 cond. boiler eff.=1.11 caldaia cond. rend.max

2

PER

1,5

1

0,5

0 2

2,2

2,4

2,6

2,8

3

3,2

3,4

3,6

3,8

4

4,2

4,4

4,6

4,8

5

COP

Fig. 4.16 PER versus COP for electric heat pumps

And comparing this PERGHP with PER of electric heat pumps, it is easy to see that: PERGHP [ PEREHP

if

COPGHP [ gMC Eel COPEHP

An additional parameter (already introduced) similar to PER, but referred to methane is the Gas Utilization Efficiency, GUE, that can be directly applied to gas—fuelled absorption heat pumps. A particular attention must be paid to the seasonal needs of the users both for a correct sizing of the adopted heat pump and for properly evaluating its performance during its working period/s. To do this we must refer to two parameters named SCOP (Seasonal Coefficient of Performance, in winter) and SEER (Seasonal Energy Efficiency Ratio, in summer). They have the same meaning of COP and EER, but instead of referring to power, they refer to the energy supplied and consumed during a season (winter or summer). The key point is the temperature of the sources, namely of the outdoor one, once the indoor temperature is kept constant. The external source temperature varies depending on its nature and location. It undergoes smaller changes if the source is water or ground, while the larger changes take place for air. Therefore the standards (refer to 11300-4/2012 and the references quoted therein) give the indications summarized in Table 4.2. The values of such time intervals depend on time changes of sources temperature as clearly reported in Table 4.2.

4.2 Comparison Among the Different Types of Heat Pump

107

Month is considered when sources temperatures are stable (e.g., water or ground as outdoor source and radiant floor or fan coil for the indoor one). The reference sources’ temperatures are the indoor set point temperature and the average outdoor temperature of the examined month. Monthly bins are taken into account for larger temperature variations (e.g., air as outdoor source and/or the indoor temperature is varied to adapt to local climate). Bin is a temperature interval, DTk, and the related bin time interval, Dtj, is the period during which temperature varies within DTk. Say we refer to a temperature Tk = 8 °C and DTk = 1 °C (temperature varies within 7.5 °C and 8.5 °C), Dtk is the time during which temperature keeps within 7.5 and 8.5 °C. Actually we need to reconstruct the trend of the temperature of the month through a normal distribution curve from the local climatic data. On this basis we can associate a bin time interval, Dtk, to any temperature bin DTk. The dedicated standard (11300-4/2012) provides a detailed procedure to perform these calculations and the reader is referred to this for any further detail. In any case to evaluate the heat pump performances the energy consumed by auxiliaries, e.g., circulation pumps and fans, must be taken into account. For example in the case of air cooled coils the energy employed by fans must be calculated, as well as the energy consumed by water pumped in and out of heat exchangers.

4.3

Further Features of Heat Pumps Operation

Even if the effects of irreversibilities and of sources temperatures have already been stressed at the beginning of this text we are going to treat them in a deeper detail and with a more oriented approach to applications. EUW and EUS commonly designate the useful energy delivered in winter and summer, i.e., respectively, the enthalpy difference at condenser and evaporator. Of course they depend at first on the phase change bell shaped curve of the employed refrigerant. With reference to Fig. 4.17: 0 EUW ¼ h2  h3 o EUW ¼ h20  h3 ¼ EUI þ h20  h2 EUS ¼ h1  h4

Subscript ‘marks the effect of a lower isentropic efficiency of the compressor.

108

4 Operating Conditions

Fig. 4.17 EUW and EUS. Dotted lines represent a worse compressor isentropic efficiency

p EU W EU W

3

4

2

2’

1

EU S

h

4.3.1

Change of Compressor’s Isentropic Efficiency

According to Fig. 4.17, a first effect of change of the isentropic efficiency is exerted on the value of EUW and, as a consequence, on COP and EER. If the isentropic efficiency gets poorer the heat exchanged with the user increases in winter, while it keeps the same in summer with the following effects on COP and EER (subscript ‘refer to a lower isentropic efficiency): COP0 ¼

h h

0 COP þ h202 h12 EUW EUW þ ðh20  h2 Þ ¼ ¼ h2 h20  h1 h2  h1 þ ðh20  h2 Þ 1 þ hh20h 2

1

h1  h4 EER ¼ EER ¼ h2 h2  h1 þ ðh20  h2 Þ 1 þ hh20h 0

2

1

If we suppose transformation 1–2 to be ideal, h2′ − h1 = (h2 − h1)/qc, we get: 0 EUW EUW þ ðh20  h1 þ h1  h2 Þ ¼ qc ¼ qc COP þ ð1  qc Þ ð h2  h1 Þ ð h2  h1 Þ   COP  COP0 1 ¼ ð 1  qc Þ 1  COP COP q ðh1  h4 Þ ¼ qc EER EER0 ¼ c ðh2  h1 Þ EER  EER0 ¼ 1  qc EER

COP0 ¼ qc

The percent changes of COP and EER are shown in Fig. 4.18. The effect of deterioration of isentropic efficiency is larger on EER than on COP.

4.3 Further Features of Heat Pumps Operation

109

Percent variation of COP and EER

35 EER

30

COP=3

25

COP=4

20

COP=5

15 10 5 0 0,7

0,75

0,8

0,85

0,9

0,95

1

-5

Isentropic efficiency

Fig. 4.18 Percent changes of COP and EER versus isentropic efficiency

4.3.2

Change of Source Temperature

A source temperature increase on the condenser side, with the same compressor isentropic efficiency, causes a decrease of EUW, while the absorbed power (by compressor) rises up. By contrast, if the isentropic efficiency reduces, EUW may keep unchanged or even increase, see Fig. 4.19a. Incidentally we recall that the isentropic efficiency of a scroll compressor generally decreases with increasing the compression ratio, while it can either increase or decrease for a screw compressor. In addition the increase of EUW is more marked for R410A than for R134a, due to the different saturation curves. A reduction of the evaporation temperature takes to a decrease of EUS (Fig. 4.19b). As already said, at the same value of volumetric flow rate, an increase of compression ratio due to lowering of evaporator temperature causes a reduction of volumetric efficiency and of density of the sucked vapor. Consequently the mass flow rate flowing in the heat pump decreases. Such a reduction can be easily evaluated for R134a (for example), referring to the following Table 4.3.

4.3.3

The Buffer Tank

The introduction of a thermal capacity within a hydraulic loop is generally aimed at smoothing the effects of transients. For example in the case of a heat pump sized for the full load requested by the user and/or on-off regulation it is appropriate to install a buffer tank, possibly used

110

4 Operating Conditions

Fig. 4.19 Temperature increase on the condenser side (a) and temperature decrease at evaporator (b)

p

(a)

2’

3’ 3

2

1

4

h p

(b)

2

3

1

4 4’

1’

h

Table 4.3 Properties of saturated R134a

T (°C)

Pressure (kPa)

Density (kg/m3)

−10 0 10 20 30 40 50 60 70 80 90

200.60 292.93 414.92 572.25 771.02 1017.61 1319.00 1682.76 2117.34 2632.97 3242.87

Liquid 1325.3 1293.3 1259.8 1224.4 1186.7 1146.1 1101.8 1052.5 995.9 927.8 837.3

Vapor 10.044 14.435 20.236 27.791 37.540 50.072 66.225 87.287 115.442 155.01 217.162

2’’

4.3 Further Features of Heat Pumps Operation

111

Table 4.4 Some proxies of effective volume (by Rossato Group) DT (°C)

4.0

5.0

Minimum 7.2

Optimum 12.7

Min. 5.7

6.0 Opt. 10.1

Min. 4.8

Opt. 8.4

also for sanitary water production, to reduce the number of transients and any swings. The volume of this tank is related to the type of plant. We have to refer to an “effective volume” which must allow the heat pump for operating at least during a minimum time requested by compressor (it cannot stop working abruptly). Thus it is necessary to ensure the availability of an effective volume, Veff, according to the following formula: Veff ¼

P  60tmin  3  m qcp DT

where: P tmin q cp DT

Heat pump power Minimum working time Water density Specific heat of water Evaporator temperature difference

kW minutes kg/m3 kJ/kgK K

The minimum time is commonly in the order of 2–5 min. Some reference values are reported in Table 4.4. As a rule of thumb an average plant volume of 15 L per installed kW can be considered. More in detail the following values can be adopted: • • • • •

Ventilation systems (e.g., fan coils): 8 l/kW. Steel radiators: 11 l/kW. Cast iron radiators: 14 l/kW. Floor heating: 23 l/kW. Central heating of large buildings: 20 l/kW.

Chapter 5

The Refrigerants

Abstract This chapter is dedicated to the most common fluids used in the current heat pumps’ technology. The main general features and parameters of these refrigerants are illustrated together with the nomenclature adopted for their classification. In particular the properties of some of the most commonly used fluids (both organic and “natural” as carbon dioxide and ammonia) are reported both as diagrams and tables.

The history of synthetic refrigerants (pure) starts in the thirties of 1900 in the United States. At the beginning they were produced from hydrocarbons like ethane, C2H6, and methane, CH4, by halogenation, i.e., by substituting atoms of hydrogen with atoms of chlorine and fluorine. In this case the halogenation was complete, that is all the atoms of hydrogen were substituted. These products were called chlorofluorocarbons, CFC. Only in 1974 their negative effects were made clear. In fact, thanks to the strong and stable bond between chlorine and fluorine, CFCs can last tens of years and accumulate in the stratosphere increasing the greenhouse effect and the depletion of the ozone layer (due to chlorine). Therefore their production was shut down in 1996 and a partial halogenation process was employed to mitigate the consequences of the use of refrigerants. The hydro-chlorofluorocarbons, HCFC, were so obtained, with a lighter impact than CFC. Let us shortly see the meaning of the abbreviations which characterize these fluids. Prefix R means refrigerant and the following numbers designate their chemical composition, from left: • the first digit gives the number of atoms of carbon, C, minus 1 (0 means 1 atom of C, 1 means 2 atoms of C); • the second digit gives the number of hydrogen, H, atoms plus 1 (2 means 1 atom of H, 3 means 2 atoms of H); • the third digit gives the number of atoms of fluorine, F; • a letter (a, b, c) following the last digit, indicates the isomer they refer to.

© Springer International Publishing AG 2018 W. Grassi, Heat Pumps, Green Energy and Technology, DOI 10.1007/978-3-319-62199-9_5

113

114

5 The Refrigerants

C

R 134

F

H

R 134a

Fig. 5.1 Molecule of R134 and its isomer a (different position of hydrogen and fluorine atoms)

R134a (tetrafluoroethane CH2FCF3) is a classical example. In fact it contains two atoms of carbon, two atoms of hydrogen and three atoms of fluorine and refers to the isomer a of Fig. 5.1. This refrigerant has no atoms of chlorine and belongs to the group of the hydro fluorocarbons (HFC), which do not contain chlorine to avoid damages to ozone. It has replaced R12 and R114 in heat pumps. Hydro fluorocarbons have not effect on atmospheric ozone but, unfortunately, they act as greenhouse gases and they have lower performances (as refrigerants) than CFCs and HCFCs. Besides HFCs, that are pure fluids, mixtures of the previously described fluids marked by series number 400 and 500 are used. The former fluids (series 400) are non-azeotropic and the latter (series 500) are azeotropic mixtures. The mixtures properties depend on the type of component fluids and on their concentrations. Generally the transition from liquid to vapor at constant pressure occurs with a temperature increase, named glide. In the case of zeotropic mixtures both vapor and liquid compositions change during the process, until their original composition is restored at the end of the phase change. Figure 5.2 shows a phase change process at constant pressure for these mixtures. The two curves correspond to the saturated vapor (upper line) and the saturated liquid line (lower line). Let us take a mixture of two components A and B, with boiling temperatures TA* and TB* (TB* > TA*) at a given pressure. At the beginning its composition corresponds to point F with temperature T1. The liquid content (mass of liquid divided by the mass of the mixture) is equal to l1 and the vapor content v1′. Vapor is richer in A (lower boiling point) and liquid in B. The same happens in an intermediate condition (2-2′, temperature T2) even if with different concentrations than before. At temperature T3 the mixture reaches l3′ and v3. This behavior of zeotropic mixtures may constitute an inconvenient either in the case of the charge or in the case of a leak of fluid, as it causes the loss of the more volatile component, thus changing the mixture composition. Fluids of this kind are, for example, R410A and R407C, which has a glide of 7 °C at atmospheric pressure and is not recommended for application where the working cycle is often inverted. R407C has been widely employed as the substitute of R22 in the field of air

5 The Refrigerants

115

T Vapor (v)

TB* 3

3’

T3

2’

2 1

T1

1’

T2

'

TA* Liquid (l)

l3

0% 100%

l2

F (l1,v3)

v2’

v1’

A B

100% 0%

Mixture composition Fig. 5.2 Phase change (liquid–vapor) of a two component mixture

conditioning. Thereafter R410A was preferred, that is a quasi-azeotropic mixture with a glide of 0.5 °C at atmospheric pressure. It is worth remarking that the replacement of a fluid with another one must be carefully evaluated accounting for the thermo physical properties, chemical properties (circuit’s corrosion), safety etc. Some mixtures can also behave as pure fluids for some values, or ranges of values of its composition, so that, at constant pressure, the phase change from liquid to vapor occurs with a constant composition and temperature (no glide). They are named azeotropic or quasi-azeotropic if both the above mentioned quantities keep almost constant. Corresponding to this it may happen that the phase change temperature is either less (positive azeotrope) or higher (negative azeotrope) than the ones of its component. Figure 5.3 shows the two cases. For example R507 is suggested as a substitute of R22 and R502 for low temperature applications (below 0 °C). Furthermore class 600 is dedicated to organic compounds like butane, propane and isobutane and class 700 to inorganic compounds like ammonia (R717) and carbon dioxide (R744). Flammability and toxicity are respectively marked by letters A and B with an increasing scale from 1 to 3 (A1 less flammable than A3 and B1 less toxic than B3). Two indexes, ODP (Ozone Depletion Potential) and GWP (Greenhouse Warming Potential) are used to classify the environmental effects with regard to ozone depletion and greenhouse gases.

116

5 The Refrigerants T

T

Vapor (v) l+v

Vapor (v) TB

TA l+v

Liquid (l) (a1)

Liquid (l) 0% 100%

TA TB

100%

A B

0%

(a2) 0% 100%

A B

Mixture composition

100% 0%

Mixture composition

(a1) Positive azeotrope

(a2) Negative azeotrope

Fig. 5.3 Liquid–vapor phase change of an azeotropic mixture

ODP is referred to R11, that is a no longer used fluid and to which the value 1 of this index is attributed. ODP varies from 0 to 1. Attributing value 1 to a refrigerant means that 1 kg of this refrigerant depletes a quantity of ozone equal to that depleted by 1 kg of R11. ODP = 0 means that there is no effect on ozone. GWP refers to the potential contribution to greenhouse effect with respect to carbon dioxide. It is generally related to the action performed by 1 kg of carbon dioxide during hundred years. If a fluid is denoted by an index GWP100 = 1000, it means that 1 kg of this fluid causes the same greenhouse effect of 1000 kg of carbon dioxide in 100 years. Unfortunately several fluids with ODP = 0 have high values of GWP100. Values of the above parameters for some refrigerants are reported in Table 5.1, as an example. It has to be emphasized that the “natural” fluid ammonia (R717) has very good values of both indexes. The properties of refrigerants can be generally evaluated on the basis of the following features: • Vapor pressure—it has to be high enough at the actual evaporation temperature to limit size and costs of compressors, without taking the condenser pressure too close to the critical point. Table 5.1 Effect of some fluids on ozone and on greenhouse phenomenon Fluid

ODP

GWP100

Average life in atmosphere (years)

R11 R22 R134a R290 (propane) R407C (R32; R125; R134a) R410A (R32; R125) R717

1 0.05 0 0 0

3800 8500 1300 20 1500

45 12 14 3 (6; 33; 14)

0

1700

(6; 33)

0

0, being supplied to the machine by the environment). Furthermore if TR > TE, as in winter, |QR| > |QE|. This means that in winter we have to subtract heat from the external environment and that its value is lower (thanks to the addition of mechanical work) than the one provided to the room. Conversely, in summer, the amount of heat delivered to the environment is higher than that extracted from the room. In real conditions, irreversible, the second principle needs to be formulated in a different manner. With the same meaning of symbols and indicating with Sg (Sg > 0) the effect of irreversibilities (entropy production): QR QE TR þ þ Sg ¼ 0 ) QR ¼
QE
Sg T R TR TE TE This means that, with the same QE, it is possible to give a lower value of power to the room or, with the same power, QR, supplied to the room, a higher value of power must be subtracted from the environment. Similar conclusions can be drawn in summer. We already said that irreversibilities can originate either inside the cycle than outside it. Let us shortly deal with the former types. The generalized differential form of Bernoulli’s theorem reads: dl ¼
vdp
de
dla where v, dp, dl, de, dla respectively represent the specific volume, the pressure difference, the work exchanged during an infinitesimal transformation, the change of macroscopic energy (kinetic and potential) and the work lost due to friction (dla > 0). Once the variation of macroscopic energy is negligible the specific work exchanged by a mechanical component (e.g., pump, compressor, and fan) to let fluid flow, can be written: dl ¼
vdp
dla Refer to an inverse cycle where the following components are present: • a compressor takes the fluid from pressure p1 to pressure p2; • two heat exchangers (for example: one providing heat to the room and the other extracting heat from the outside); • an expansion valve.

170

8 Additional Thermodynamic Remarks

Work can only be exchanged by the compressor. Therefore: Compressor It is the only component that can exchange work with fluid. It is dedicated to compress the fluid Z
vdp compression

Z

l¼


vdp
laðcompressorÞ compression

to compensate its internal losses plus those at its inlet and outlet, la (compressor), and the total pressure losses of the circuit. Heat exchanger dedicated to the inside Z 0 ¼ lexch:1 ¼
vdp
laðexch:1Þ exch:1

Z

vdp ¼
laðexch:1Þ exch:1

Pressure decreases along the heat exchanger, due to friction. In the case of phase change this implies a decrease of temperature. Expansion valve (adiabatic) Z 0¼l¼
vdp
laðvalveÞ Z

valve

vdp ¼
laðvalveÞ valve

This valve is aimed at making the pressure decrease from the value of the condenser to the one of the evaporator. This pressure decrease is obtained by an intrinsically irreversible process due to friction. A part from the case of a capillary tube or a simple throttling (like in domestic refrigerator), this valve regulates the flow rate by both an on-off and a modulating operation. Both in heat pumps and in refrigeration this valve controls the fluid flow rate in order to prevent the inflow of liquid into the compressor. Heat exchanger with the external environment. Z 0¼l¼
vdp
laðexch:2Þ Z

exch:2

vdp ¼
laðexch:2Þ exch:2

8.2 First and Second Principles of Thermodynamics

171

In conclusion, the specific work (J/kg) to be supplied by the compressor is: Z vdp
la l¼
compressor

la ¼laðcompressorÞ þ laðexch:1Þ þ laðvalveÞ þ laðexch:2Þ If the friction effects on the heat exchangers can be neglected: Z l¼


vdp
laðcompressorÞ
laðvalveÞ compression

Figure 8.1 shows two curves relating the pressure difference to flow rate, one characteristic (dotted line) of the fluid loop and the other (continuous line) of the compressor. If the pressure drop in the loop increases (for example due to throttling) with the same rotation speed of the compressor, the work point moves from A to B. To restore the same flow rate as in point A the rotation speed has to be increased to reach point C, with a pressure drop increase. On the other hand, if we want to keep the same Dp, we move to point D, with a flow rate increase. The corresponding mechanical power for an adiabatic compressor is L = ml = −mDh, where m is the mass flow rate and Dh (>0) the difference of enthalpy between the compressor discharge and suction. If the compressor is not adiabatic and exchanges a thermal power Q ( x1) is: qEV ¼ h2
h1 ¼ ðx2
x1 Þr [ 0

1

Ratio between the mass of vapor and the mass of mixture, i.e., the vapor mass in 1 kg of mixture. Recall that, at constant pressure, the temperature keeps constant as well, during a change of phase.

2

174

T

8 Additional Thermodynamic Remarks isobar

Critical isotherm

p

Critical isotherm

G.

G.

Critical point

Critical point V.

L.

S.V.

V.

L.

isotherm

isobar

L. Liquid

S.V.

G. Gas

s

V. Vapor

S.V. Saturated vapor

h

Fig. 8.3 Two phase region in the temperature-entropy and in the pressure enthalpy planes

Therefore heat has to be supplied to the fluid. In the case of condensation (x2 < x1) it is: qCOND ¼ h2
h1 ¼ ðx2
x1 Þr\0 Heat is supplied by the fluid. During both these heat exchanges fluid flows in the ducts of the dedicated heat exchangers progressively change its quality and decrease its pressure. Figure 8.4 Water to be cooled Evaporating refrigerant

Annular motion: Super heated liquid in contact vapor with wall and vapor “core”. Some two – phase flow regimes during refrigerant evaporation

Cold refrigerant, low quality

Large vapor mass originate.

Fig. 8.4 Sketch of some two–phase flow regimes in an evaporator

8.3 Phase Change of Pure Substances

175

Table 8.1 Indicative values of overall heat transfer coefficients Heat exchange

Typical value (W/m2K)

Range

Gas/gas at normal pressure 20 5–50 Gas/gas at high pressione 200 50–500 Gas/liquid 50 10–100 Liquid/liquid in tubular heat exchanger 1000 200–2000 Liquid/liquid in plate heat exchanger 2500 500–5000 Gas cooled condenser 50 10–100 Liquid cooled condenser 3000 500–6000 Gas heated evaporator 50 10–100 Liquid heated evaporator 5000 500–10,000 Note—global heat transfer coefficients ranging from 2000 to 7000 W/m2K, with R410A, have been obtained for a corrugated flat plate evaporator

qualitatively shows the different two–phase regimes which may occur in an annular heat exchanger during an evaporation. The refrigerant enters with a low quality and the water to be cooled at a given temperature. Water cools down to the necessary temperature for space cooling. At the same time the refrigerant quality increases up to 1, then giving origin to superheated vapor in dry evaporators. Qualitatively we might say that the reverse occurs in condensation. For further details on the subject the reader is referred to [1–3]. Table 8.1 reports some characteristic values of the overall heat transfer coefficient for different types of heat exchange.

References 1. Butterworth D. & Hewitt G.F., Two-phase flow and heat transfer, Harwell Series, Oxford University Press, Oxford (UK), 1979. 2. Collier J.G., Thome J.R., Convective Boiling and Condensation, Clarendon Press, Oxford, 1994. 3. Hewitt G.F. executive editor, Heat Exchanger Design Handbook 1998, Part 3 Thermal and Hydraulic Design of Heat Exchangers, Begell House Inc. New York.