politecnico di milano · politecnico di milano ... the semi-empirical model simulates both the pure...
TRANSCRIPT
POLITECNICO DI MILANO
Scuola di Ingegneria Industriale e dell’Informazione
Corso di Laurea Magistrale in Ingegneria Energetica
Experimental analysis and modeling of oil retention effects on
heat transfer and pressure drop during evaporation of low
GWP refrigerants in microchannel heat exchangers
Relatore: Prof. Luca Molinaroli
Correlatore: Prof. Lorenzo Cremaschi
Tesi di Laurea di:
Carlo Andres
Matr. 804830
Anno Accademico 2014/2015
i
Acknowledgements
I would like to express my sincere gratitude to my advisor Prof. Luca Molinaroli for the
useful comments, remarks and engagement through the learning process of this master
thesis.
Furthermore, I would like to thank Prof. Lorenzo Cremaschi for giving me the
opportunity to join his research group at Oklahoma State University.
I am also grateful to my colleagues for the support they gave me and the stimulating
discussions we had.
I would like to thank all my friends for having always been close to me.
Special and immeasurable thanks to my parents and sister, who always support me
during this period, and to Tecla, who sustained me in the last three years.
iii
Summary
In a vapor compression cycle the compressor needs oil employment in order to prevent
surface-to-surface contact, remove heat, provide sealing, keep out contaminants and
dispose of debris created by wear. Although most of the oil remains in the compressor, a
small amount, which ranges from 0.5 to 3 percent of the total refrigerant flow rate
circulating into the system, escapes the oil separator after the compressor and circulates
throughout the cycle. Furthermore, the lubricant can accumulate inside the heat
exchanger components, causing an insufficient oil return to the compressor, and then a
lack of lubrication. In heat exchangers, the oil addition effects are undesired yet
unavoidable: the presence of oil causes an increase of the pressure drop and a
penalization of the whole heat transfer process. Studies about oil return and oil transport
in suction lines are numerous in literature, while oil retention measurements in
evaporators used in air conditioning applications are rather sporadic.
Moreover, to meet the terms of 1987 Montreal Protocol, chlorofluorocarbons and
hydrochlorofluorocarbons have been gradually phased out and substituted with zero
ozone depletion potential (ODP) fluids. However, some of the replacement refrigerants,
such as R410A, present high global warming potential (GWP) values, that might still be
of concern from an environmental perspective in case of leakage or improper charge
management. Few studies about zero ozone depletion potential and low GWP
refrigerants are available in the literature, only with the aim of demonstrating good
performances in terms of COP and capacity, but there is no information on oil retention
and its effect on heat transfer and pressure drop characteristics.
For the reasons explained above, the objective of this work is to investigate the oil
retention and its effects on heat transfer and pressure drops of new low GWP
refrigerants in microchannel type evaporators and, consequently, to provide a
comparison between the alternative fluids and the well known R410A. Furthermore, in
order to understand deeper the phenomenon related to the lubricant addition, a semi-
empirical heat exchanger model is developed.
A well known and highly-employed fluid, R410A, and three new low GWP refrigerants,
DR5A, R32 and R1234yf, along with synthetic polyol ester oil, are tested using two
different louvered-fin aluminum microchannel evaporators. The microchannel
configuration is considered one of the best technology commercially available to
increase the energy efficiency and limit the environmental impact, through a reduction
in refrigerant inventories and in vapor compression components size. The experiments
are conducted at Oklahoma State University Psychrometric Chamber, which is able on
one hand to reproduce a wide range of outdoor climate conditions, and on the other
hand to control the amount of oil injected and measure the oil trapped into the heat
iv
exchangers, the heat transfer rates and the pressure drop. The test conditions, such as
refrigerant saturation temperature and degree of superheating, are selected based on
R410A typical applications in air-conditioning systems. The circulating oil mass
fraction (OMF), a representative of the ratio between the lubricant and refrigerant mass
flow rates, is investigated higher than the usual values encountered in vapor
compression cycles, in order to clearly understand the detrimental effects on heat
transfer, pressure drop and oil retention related to the lubricant addition.
The first objective of the experiments is providing a comparison between R410A and
the developmental refrigerant DR5A, with the aim of demonstrating their similar
behavior and the possible use of DR5A as replacement for R410A. The comparison is
performed keeping fixed the system parameters, such as the air side volumetric flow
rate and dry bulb temperature, and maintaining the same saturation temperature and
degree of superheating on the refrigerant side. An additional comparison series between
R410A and DR5A is conducted under the same mass flux condition to better understand
the oil retention phenomenon. The second target of the experiments is providing
comparison data among DR5A, and respectively R32 and R1234yf, which are the two
refrigerants the DR5A is composed of; the tests are performed under the same
refrigerant mass flow rate conditions, in order to compare the behavior of the two pure
fluids with the performance given by their mixture after oil addition. Keeping fixed the
refrigerant side parameters, such as the refrigerant mass flow rate, degree of
superheating and saturation temperature, it is possible to understand which refrigerant
among R32 and R1234yf presents a higher detrimental effect in terms of oil retention,
pressure drop increase and heat transfer capacity reduction.
The oil effects are evaluated through three different parameters. The first one is the Heat
Transfer Factor (HTF), defined as the ratio between the heat transfer capacity when oil
and refrigerant are flowing in the microchannel heat exchanger and the capacity when
only pure refrigerant is flowing. The second one is the Pressure Drop Factor (PDF),
which evaluates the ratio between the pressure drop measured with oil circulating into
the system and the pressure drop without lubricant addition. The last parameter is the
Oil Retention Volume normalized (ORVN), which represents the mass of lubricant
trapped in the microchannel evaporator divided by the oil density and then normalized
dividing by the internal volume of the heat exchanger.
The semi-empirical model simulates both the pure refrigerant and the refrigerant and oil
mixture behavior in the two microchannel heat exchangers tested. The simulation model
is based on the segment-by-segment approach: the entire evaporator is reduced to a set
of vertical parallel columns, each one representing a single microchannel tube.
Assuming uniform conditions both on the air and refrigerants sides, guaranteed by the
psychrometric outdoor chamber, and controlling the refrigerant conditions at the
evaporator inlet, it is possible to consider only one column in the calculations. The
single tube is then divided into 100 segments, each one solved iteratively until
v
convergence is reached. Providing the first segment conditions as input, the simulation
model calculates the output variables and passes them as input data for the succeeding
one, until the last element is solved. The algorithm uses the ε-NTU method to determine
the heat transfer capacity, while correlations developed for microchannel heat
exchangers are utilized in order to estimate the pressure drop. To analyze the oil
detrimental effects on heat transfer and pressure drop, the thermodynamic approach is
used as reference. Since the oil partial pressure in the vapor-phase is negligible due to
its higher boiling point than pure refrigerant, the lubricant is considered a non-volatile
component at the operating conditions used. Hence the oil is present only in the liquid
phase while the vapor phase is only made of pure refrigerant. The thermodynamic
approach peculiarity consists in treating the lubricant and refrigerant mixture as a
zeotropic mixture and then considering the mixture bubble temperature instead of the
pure refrigerant saturation temperature to describe the evaporating process. Lastly, the
oil retention at the inlet header and microchannel tubes is calculated through empirical
correlations depending on void fraction, while the lubricant trapped in the outlet header
is determined considering both the void fraction and the internal geometry effects.
The lubricant addition always penalizes the microchannel heat exchanger performances,
decreasing the evaporator capacity and increasing the pressure drop. The oil detrimental
effect is proportional to the oil mass fraction for all the refrigerants studied.
The experimental comparison between DR5A and R410A shows that DR5A can be
considered a good low GWP replacement for air conditioning applications. The DR5A
refrigerant seems operating properly at the conditions typical of the systems using
R410A as working fluid. The comparison series conducted keeping fixed the system
parameters, that are the air side volumetric flow rate and dry bulb temperature, illustrate
similar trends for both R410A and DR5A in terms of Heat Transfer Factor and Pressure
Drop Factor. Moreover, the behavior of the two different fluids is the same while
changing the refrigerant degree of superheating. Furthermore, R410A and DR5A
present analogous Oil Retention Volume normalized when the comparison between the
two fluids is conducted under the same mass flux conditions. The experimental
comparison between R32, R1234yf and DR5A refrigerants, conducted keeping fixed the
refrigerant side parameters, shows that R32 on one hand has the highest Heat Transfer
Factor penalization, while on the other hand presents the lowest amount of Oil
Retention Volume normalized.
The semi-empirical model is initially validated with pure refrigerant data. The
simulation results always show a good agreement with the experimental values, in terms
of capacity and pressure drop. Additionally, the simulation results with oil present
similar values compared to the experimental data. The semi-empirical model is able to
catch the capacity reduction and the pressure drop increase due to lubricant addition for
all the refrigerants analyzed and the different heat exchanger geometries used. The
simulation model addresses most of the oil retention to the horizontal outlet header,
vi
where the highest local oil concentration is reached. Two different trends are pointed
out in the predicted oil retention values: R410A results are in good agreement with the
experimental data, while DR5A, R32 and R1234yf simulation values are
underestimated. The main reason is identified in the different mass flux conditions
among the series. Low mass fluxes determine low vapor-phase velocity and shear stress,
so the vapor refrigerant is less effective in removing the lubricant droplets, resulting in
higher amount of oil retained.
vii
Riassunto Esteso
In un ciclo a compressione di vapore, il compressore necessita l’utilizzo di olio per
evitare il contatto fra le superfici rotanti, rimuovere il calore in eccesso ed eliminare i
contaminanti e i detriti causati dall’usura. Sebbene gran parte dell’olio lubrificante
rimanga all’interno del compressore, un piccolo quantitativo, che varia fra lo 0.5 e il 3
per cento della portata totale di refrigerante del sistema, riesce a scappare dal separatore
posto a valle del compressore e circola attraverso l’intero sistema. Inoltre, l’accumulo
d’olio all’interno degli scambiatori di calore può causare un insufficiente ritorno di olio
al compressore e una carenza di lubrificazione. Gli effetti correlati all’aggiunta di olio
lubrificante negli scambiatori di calore sono negativi ma al tempo stesso inevitabili: la
presenza di olio determina un incremento delle perdite di carico e una penalizzazione
dell’intero processo di scambio termico. Studi riguardo il ritorno di olio lubrificante al
compressore e il trasporto dell’olio nella linea di aspirazione del compressore sono
numerosi nella letteratura scientifica, mentre misure dirette della ritenzione dell’olio
lubrificante in evaporatori usati in applicazioni di condizionamento dell’aria sono
abbastanza sporadici.
Per rispettare i termini imposti dal Protocollo di Montreal del 1987, clorofluorocarburi e
idroclorofluorocarburi sono stati gradualmente sostituiti da fluidi con zero ozone
depletion potential (ODP). Alcuni fluidi refrigeranti individuati come possibili sostituti,
ad esempio l’R410A, hanno un alto global warming potential (GWP). Quest’ultimo
parametro può rappresentare un problema da un punto di vista ambientale nel caso di
perdite o impropria gestione del fluido refrigerante. Nella letteratura scientifica i pochi
studi riguardanti fluidi refrigeranti con ODP nullo e basso GWP hanno principalmente il
fine di dimostrare il raggiungimento di buone prestazioni in termini di COP e potenza
termica scambiata, ma non sono presenti informazioni riguardo la ritenzione dell’olio
lubrificante e i suoi effetti su scambio termico e perdite di carico.
Per i motivi introdotti precedentemente, l’obiettivo di questo lavoro è indagare la
ritenzione dell’olio lubrificante e i suoi effetti sullo scambio termico e sulle perdite di
carico in scambiatori di calore a micro canali. L’aspetto innovativo del lavoro consiste
nel fornire un confronto dettagliato fra il fluido R410A e i suoi possibili sostituti a basso
GWP. Inoltre, per comprendere più a fondo le conseguenze dovute all’aggiunta di olio
lubrificante, un modello semi-empirico dello scambiatore di calore è stato sviluppato.
Un fluido refrigerante molto impiegato nelle applicazioni di condizionamento dell’aria,
l’R410A, e tre nuovi fluidi con basso potenziale di riscaldamento globale, più olio
lubrificante sintetico a base di poliolestere, sono stati testati usando due diversi
scambiatori di calore a micro canali con alette perforate in alluminio. La geometria a
micro canali è considerata una delle migliori tecnologie attualmente in commercio,
viii
poiché permette un incremento dell’efficienza energetica e una limitazione dell’impatto
ambientale, grazie alla riduzione delle dimensione dei componenti del ciclo a
compressione di vapore e, conseguentemente, del carico di fluido refrigerante richiesto.
Gli esperimenti sono stati condotti utilizzando la camera psicrometrica presso il
Dipartimento di Ingegneria Meccanica e Aerospaziale dell’Oklahoma State University,
apparecchiatura capace di riprodurre un’ampia varietà di condizioni climatiche esterne e
adatta a misurare il quantitativo di olio lubrificante iniettato e trattenuto dallo
scambiatore di calore, la potenza termica scambiata e le perdite di carico. Le condizioni
sperimentali, come la temperatura di saturazione e il grado di surriscaldamento del
refrigerante, sono state scelte a partire dai valori tipici utilizzati nei sistemi di
condizionamento che utilizzano R410A come fluido di lavoro. La frazione massica
d’olio lubrificante studiata, rappresentativa del rapporto fra le portate massiche di olio e
refrigerante, è stata appositamente scelta più alta dei valori convenzionali incontrati nei
cicli a compressione di vapore, in maniera da comprendere al meglio gli effetti negativi
connessi all’aggiunta di olio lubrificante relativamente a scambio termico, perdite di
carico e ritenzione d’olio.
Il primo obiettivo della parte sperimentale è fornire un paragone fra R410A e il fluido
refrigerante di nuova generazione DR5A, atto a dimostrare il loro simile
comportamento e il possibile utilizzo del DR5A come sostituto per l’R410A. Il
paragone è stato realizzato tenendo fissi i parametri del sistema, ossia la portata
volumetrica e la temperatura di bulbo secco dell’aria e mantenendo la stessa
temperatura di saturazione e grado di surriscaldamento del refrigerante. Un’ulteriore
campagna sperimentale di confronto fra DR5A e R410A è stata condotta mantenendo
costante la portata massica di refrigerante, in modo tale da comprendere meglio il
fenomeno della ritenzione di olio lubrificante. Il secondo scopo della parte sperimentale
è fornire un paragone fra DR5A, R32 e R1234yf; questi ultimi sono i due fluidi puri che
compongono la miscela DR5A. Gli esperimenti sono realizzati mantenendo le stesse
condizioni di portata massica di refrigerante, con il fine di analizzare il comportamento
dopo l’iniezione di olio lubrificante sia per i due fluidi puri sia per la miscela. Tenendo
fissi i parametri caratterizzanti il lato refrigerante, ossia la portata massica, la
temperatura di saturazione e il grado di surriscaldamento del fluido refrigerante, è
possibile comprendere quale fra R32 e R1234yf presenti una maggior penalizzazione in
termini di ritenzione d’olio, incremento delle cadute di pressione e riduzione della
potenza termica scambiata.
Gli effetti dovuti alla presenza di olio lubrificante sono valutati tramite tre differenti
parametri. Il primo è l’Heat Transfer Factor (HTF), definito come il rapporto fra la
potenza termica scambiata quando la miscela di olio e refrigerante circola nello
scambiatore a micro canali e la potenza scambiata quando solo puro refrigerante circola
nello scambiatore. Il secondo parametro è il Pressure Drop Factor (PDF), che
rappresenta il rapporto fra perdite di carico misurate con olio circolante nel sistema e
perdite di carico in assenza di olio lubrificante. Il terzo e ultimo parametro è l’Oil
ix
Retention Volume normalized (ORVN), definito come il volume di olio lubrificante
ritenuto nell’evaporatore a micro canali, normalizzato rispetto al volume interno dello
scambiatore di calore.
Il modello semi-empirico è in grado di simulare il comportamento sia del refrigerante
puro sia della miscela composta da refrigerante e olio lubrificante nei due scambiatori di
calore a micro canali usati. Il modello numerico utilizza un approccio segmento-per-
segmento: l’intero evaporatore è ridotto a una serie di colonne verticali parallele,
ognuna delle quali rappresenta un singolo micro canale. Assumendo condizioni
uniformi sia dal lato aria sia da quello refrigerante, garantite rispettivamente tramite la
camera psicrometrica e il controllo delle condizioni del refrigerante in ingresso
all’evaporatore, è possibile effettuare i calcoli solo relativamente a una singola colonna.
Il singolo micro canale è suddiviso in 100 segmenti, ognuno dei quali viene risolto
iterativamente fino a convergenza. Fornendo le condizioni di input per il primo
segmento, il modello numerico è in grado di calcolare le variabili di output, che
successivamente vengono usate come input per il successivo segmento; questa
procedura si interrompe quando l’ultimo elemento viene risolto. La Potenza termica
scambiata è calcolata tramite il metodo ε-NTU, mentre le perdite di carico sono
determinate attraverso correlazioni sviluppate per scambiatori a micro canali. Per
comprendere gli effetti penalizzanti relativamente allo scambio termico e alle cadute di
pressione dovuti all’aggiunta di olio lubrificante, l’approccio termodinamico è usato
come linea guida. Dal momento che la pressione parziale dell’olio nella fase vapore è
trascurabile, a causa del suo alto punto di ebollizione rispetto al refrigerante puro, l’olio
lubrificante è trattato come un componente non volatile durante le normali condizioni
operative. In questo modo è possibile considerare la presenza dell’olio lubrificante solo
nella fase liquida, mentre la fase vapore composta solo da puro refrigerante. La
peculiarità dell’approccio termodinamico consiste nel trattare la miscela formata da
refrigerante e olio lubrificante come una miscela zeotropica e nel considerare quindi la
temperatura di bolla della miscela anziché la temperatura di saturazione del refrigerante
puro per descrivere il processo evaporativo. Infine, la ritenzione dell’olio lubrificante
nel serbatoio d’accumulo all’ingresso e nei micro canali è determinata attraverso
correlazioni empiriche dipendenti dalla frazione di vuoto, mentre l’olio intrappolato nel
serbatoio d’accumulo all’uscita è calcolato considerando sia la frazione di vuoto sia gli
effetti dovuti alla geometria interna.
L’aggiunta di olio lubrificante penalizza sempre le prestazioni dello scambiatore a
micro canali, causando una diminuzione della potenza termica scambiata e un
incremento delle cadute di pressione. L’effetto di penalizzazione dovuto alla presenza di
olio lubrificante è direttamente proporzionale alla frazione massica di olio per tutti i
refrigeranti analizzati in questo lavoro.
Il confronto sperimentale fra DR5A e R410A mostra come il fluido refrigerante DR5A
possa essere considerato un buon sostituto a basso potenziale di riscaldamento globale
x
nelle applicazioni di condizionamento dell’aria che usano R410A come fluido di lavoro.
Il confronto, condotto mantenendo fissi i parametri di sistema, ossia la portata
volumetrica e la temperatura di bulbo secco dell’aria, evidenzia un andamento simile in
termini di Heat Transfer Factor e Pressure Drop Factor sia per il refrigerante DR5A sia
per l’R410A. In aggiunta, il comportamento dei due diversi fluidi al variare del grado di
surriscaldamento risulta essere lo stesso. L’R410A e il DR5A presentano valori simili di
Oil Retention Volume normalized fissando le stesse condizioni di portata massica.
Infine, il confronto fra DR5A, R32 e R1234yf, condotto mantenendo costanti i
parametri caratteristici del lato refrigerante, mostra come il fluido refrigerante R32
presenti sia la maggior penalizzazione in termini di Heat Transfer Factor, sia il minor
valore di Oil retention Volume normalized.
Il modello semi-empirico è inizialmente validato usando i dati sperimentali relativi al
puro refrigerante. I risultati di simulazione sono sempre in buon accordo con i valori
sperimentali, per quanto riguarda la potenza scambiata e le cadute di pressione. Anche i
risultati delle simulazioni che considerano la presenza di olio lubrificante mostrano
valori simili a quelli dei dati sperimentali. Il modello semi-empirico è in grado di
riprodurre la riduzione della potenza termica scambiata e l’incremento delle perdite di
carico dovuti alla presenza di olio per tutti i refrigeranti analizzati e le geometrie di
scambiatore a micro canali considerate. Dall’analisi dei risultati di simulazione, il
serbatoio di uscita, dove vengono raggiunti i massimi valori di concentrazione locale di
olio, presenta la maggiore ritenzione di olio lubrificante. I risultati di simulazione
presentano due diversi andamenti nella previsione dei valori numerici di olio trattenuto
nel sistema: il fluido R410A mostra un buon accordo, mentre DR5A, R32 e R1234yf
mostrano una sottostima rispetto ai dati sperimentali. La principale ragione di questa
deviazione è rappresentata dalle diverse condizioni di portata massica fra le diverse
serie sperimentali: infatti basse portate massiche determinano basse velocità della fase
vapore e minori sforzi di taglio, quindi il refrigerante in fase vapore risulta essere meno
efficace nella rimozione delle goccioline di olio lubrificante depositate all’interno dello
scambiatore di calore, risultando in un aumento della ritenzione di olio.
xi
Contents
INTRODUCTION .......................................................................................................... 1
1 LITERATURE REVIEW ....................................................................................... 3
1.1 MICROCHANNEL HEAT EXCHANGERS .................................................................. 3
1.2 CURRENT LIMITS ON REFRIGERANT USE .............................................................. 4
1.3 REFRIGERANTS AND LUBRICANT TESTED ............................................................ 5
1.4 REFRIGERANT-OIL MIXTURES .............................................................................. 6
1.4.1 Different approaches in studying refrigerant-oil mixtures ......................... 7
1.5 REFRIGERANT-OIL MIXTURE FLOW CHARACTERISTICS STUDY ............................. 8
1.5.1 Oil influence on Pool Boiling ..................................................................... 8
1.5.2 Oil influence on flow boiling ...................................................................... 8
1.6 EFFECTS OF OIL ON REFRIGERATION COMPONENTS ............................................ 12
1.7 PREVIOUS MODELING WORKS ............................................................................ 14
2 THE EXPERIMENTAL APPARATUS .............................................................. 19
2.1 THE PSYCHROMETRIC CHAMBER ...................................................................... 19
2.2 THE MICROCHANNEL EVAPORATORS ................................................................ 21
2.3 THE AIR SAMPLING DEVICE ............................................................................... 23
2.4 THE REFRIGERANT AND OIL LOOPS .................................................................... 24
2.5 TEST PROCEDURE .............................................................................................. 28
2.6 DATA ANALYSIS ................................................................................................ 30
2.6.1 Heat transfer and pressure drop analysis ................................................. 32
2.6.2 Oil retention volume analysis ................................................................... 35
2.7 UNCERTAINTY ANALYSIS .................................................................................. 38
3 EXPERIMENTAL RESULTS ............................................................................. 43
3.1 DR5A AND R410A COMPARISON ...................................................................... 44
3.1.1 Comparison under same air dry bulb temperature conditions ................. 44
3.1.2 Comparison under same mass flux conditions ......................................... 47
3.1.3 Effect of different degree of superheating ................................................. 51
3.1.4 Effect of different geometry ....................................................................... 55
3.2 R32, R1234YF AND DR5A COMPARISON .......................................................... 58
4 SIMULATION CODE DESCRIPTION.............................................................. 63
4.1 PREVIOUS WORK ON THE SIMULATION CODE ..................................................... 63
4.2 MICROCHANNEL EVAPORATOR SIMULATION CODE............................................ 63
4.3 THE AIR SIDE CORRELATIONS ............................................................................ 67
4.4 THE REFRIGERANT SIDE CORRELATIONS ............................................................ 69
4.5 REFRIGERANT-OIL MIXTURE PROPERTIES CALCULATION ................................... 74
4.6 OIL RETENTION CALCULATION .......................................................................... 78
xii
5 SIMULATION RESULTS WITHOUT OIL ....................................................... 81
5.1 EVAPORATOR CAPACITY AND PRESSURE DROP VALIDATION .............................. 81
5.2 FURTHER SIMULATION RESULTS WITHOUT OIL ................................................... 85
6 SIMULATION RESULTS WITH OIL ............................................................... 89
6.1 EVAPORATOR CAPACITY, PRESSURE DROP AND OIL RETENTION ...................... 89
6.2 R410A RESULTS WITH OIL ................................................................................. 93
6.3 DR5A RESULTS WITH OIL .................................................................................. 98
7 CONCLUSIONS .................................................................................................. 105
7.1 CONCLUSIONS ABOUT THE EXPERIMENTAL WORK ........................................... 105
7.2 CONCLUSIONS ABOUT THE SIMULATION WORK ................................................ 107
7.3 FUTURE WORKS ............................................................................................... 108
NOMENCLATURE .................................................................................................... 109
REFERENCES ............................................................................................................ 113
xiii
List of figures
Fig. 2.1: Air conditioning loop and evaporator inside the psychrometric chamber ....... 20
Fig. 2.2: Microchannel evaporator A: (top) actual picture of the heat exchanger and
(bottom) scheme of the evaporator with refrigerant and air flow direction during the
tests ................................................................................................................................. 22
Fig. 2.3: Microchannel evaporator B: (top) scheme with refrigerant and air flow
directions during the tests; (bottom left) scheme of the evaporator with indication of the
main dimensions and (bottom right) actual picture of the heat exchanger ..................... 23
Fig. 2.4 Air sampling device in front of the microchannel evaporator ........................... 24
Fig. 2.5: Particular of the refrigerant loop ...................................................................... 26
Fig. 2.6: Schematic representation of refrigerant and oil loops ...................................... 27
Fig. 2.7: The Target Labview interface for oil retention tests ........................................ 31
Fig. 2.8: The Host Labview interface for oil retention tests ........................................... 31
Fig. 2.9: Schematic representation of fixed power electrical preheaters and
microchannel evaporator ................................................................................................. 33
Fig. 2.10: Particular of the two sightglasses placed at the evaporator outlet .................. 36
Fig. 2.11: trend as function of for inlet and outlet test ..................................... 37
Fig. 2.12: Total amount of lubricant mass injected during outlet test ............................ 38
Fig. 3.1: Experimental Heat Transfer Factor under different refrigerant side saturation
temperature conditions .................................................................................................... 45
Fig. 3.2: Experimental Pressure Drop Factor under different refrigerant side saturation
temperature conditions .................................................................................................... 46
Fig. 3.3: Experimental Oil Retention Volume Normalized under different refrigerant
side saturation temperature conditions ........................................................................... 47
Fig. 3.4: Experimental Heat Transfer Factor under same mass flux conditions ............. 48
Fig. 3.5: Experimental Pressure Drop Factor under same mass flux conditions ............ 49
Fig. 3.6: Experimental Oil Retention Volume Normalized under same mass flux
conditions ........................................................................................................................ 50
Fig. 3.7: Experimental Oil Retention Volume Normalized with two different mass flux
conditions ........................................................................................................................ 51
Fig. 3.8: Experimental Heat Transfer Factor with different degree of superheating ...... 53
xiv
Fig. 3.9: Experimental Pressure Drop Factor with different degree of superheating ...... 54
Fig. 3.10: Experimental Oil Retention Volume Normalized with different degree of
superheating .................................................................................................................... 55
Fig. 3.11: Experimental Heat Transfer Factor for Evaporator B .................................... 56
Fig. 3.12: Experimental Pressure Drop Factor for Evaporator B .................................... 57
Fig. 3.13: Experimental Oil Retention Volume Normalized for Evaporator B .............. 58
Fig. 3.14: R1234yf, DR5A and R32 experimental Heat Transfer Factor under same
refrigerant side conditions ............................................................................................... 59
Fig. 3.15: R1234yf, DR5A and R32 experimental Pressure Drop Factor under same
refrigerant side conditions ............................................................................................... 60
Fig. 3.16: R1234yf, DR5A and R32 experimental Oil Retention Volume normalized
under same refrigerant side conditions ............................................................................ 61
Fig. 4.1: Schematic representation of the segment by segment approach used in the
simulation program ......................................................................................................... 64
Fig. 5.1: Comparison between predicted and experimental capacity without oil ........... 82
Fig. 5.2: Comparison between predicted and experimental pressure drop without oil ... 83
Fig. 5.3: Particular of the comparison between predicted and experimental pressure drop
without oil ....................................................................................................................... 83
Fig. 5.4: Predicted Heat transfer coefficient trend without oil ........................................ 86
Fig. 5.5: Predicted Capacity trend without oil ................................................................ 87
Fig. 6.1: Comparison between predicted and experimental capacity with oil ................ 90
Fig. 6.2 Comparison between predicted and experimental pressure drop with oil ......... 91
Fig. 6.3: Particular of the comparison between predicted and experimental pressure drop
with oil ............................................................................................................................. 91
Fig. 6.4 Comparison between predicted and experimental oil retention mass ................ 92
Fig. 6.5: R410A simulation and experimental Heat Transfer Factor .............................. 94
Fig. 6.6: Effect of different oil concentrations on the simulation Heat Transfer
Coefficient ....................................................................................................................... 95
Fig. 6.7: R410A simulation and experimental Pressure Drop Factor ............................. 96
Fig. 6.8 R410A simulation and experimental Oil Retention Volume normalized .......... 97
Fig. 6.9: Refrigerant and lubricant masses trend inside a microchannel tube ................. 98
Fig. 6.10: DR5A simulation and experimental Heat Transfer Factor ........................... 100
Fig. 6.11: DR5A simulation and experimental Pressure Drop Factor .......................... 101
xv
Fig. 6.12: DR5A simulation and experimental Oil Retention Volume normalized ..... 102
Fig. 6.13: Dimensionless oil retention and mixture quality trends inside a microchannel
tube ................................................................................................................................ 103
Fig. 7.1: Particular of the Evaporator A horizontal outlet header ................................. 106
xvii
List of Tables
Table 2.1: Main parameters of the two evaporators ....................................................... 22
Table 2.2: Air side sensors specifications ....................................................................... 39
Table 2.3: Refrigerant and oil side sensors specifications .............................................. 40
Table 2.4: Calculated uncertainties of the main paramenters ......................................... 41
Table 3.1: Experimental test matrix ................................................................................ 43
Table 4.1: Index of parameters used in the simulation code for refrigerant side
correlations ...................................................................................................................... 71
Table 5.1: Pressure drop without oil analysis ................................................................. 85
Table 6.1: Simulation input and output parameters for R410A tests .............................. 93
Table 6.2: Simulation input and output parameters for DR5A tests ............................... 99
xix
Abstract
In a vapor compression cycle the compressor needs oil employment in order to prevent
surface-to-surface contact, remove heat, provide sealing, keep out contaminants and
dispose of debris created by wear. Although most of the oil remains in the compressor, a
small amount, which ranges from 0.5 to 3 percent of the total refrigerant flow rate
circulating into the system, escapes the oil separator after the compressor and circulates
throughout the cycle. Furthermore, the lubricant can accumulate inside the heat
exchanger components, causing an insufficient oil return to the compressor, and then a
lack of lubrication. In heat exchangers, the oil addition effects are undesired yet
unavoidable: the presence of oil causes an increase of the pressure drop and a
penalization of the whole heat transfer process.
The aim of this work is to analyze the oil retention and its effects on heat transfer rate
and pressure drop in microchannel type evaporators. The unique feature consists of
providing comparison data between a well-known and highly employed refrigerant,
such as R410A, and its possible low Global Warming Potential replacements, DR5A,
R32 and R1234yf.
The oil effects are evaluated comparing both the heat transfer capacity and the pressure
drop after oil addition with the ones of the corresponding pure refrigerant test having
the same total mass flow rate. This approach allows to address all the variations to the
lubricant replacing refrigerant inside the evaporator.
The extensive experiments and simulations demonstrate that the oil addition always
penalizes the microchannel heat exchanger performances over the wide range of
conditions tested. The oil retention, the heat transfer degradation and the pressure drop
increase are proportional to the oil mass fraction, representative of the ratio between the
lubricant and refrigerant mass flow rates entering the coil.
Keywords: microchannel evaporator, low GWP refrigerants, oil retention, capacity,
pressure drop
xxi
Abstract
In un ciclo a compressione di vapore, il compressore necessita l’utilizzo di olio per
evitare il contatto fra le superfici rotanti, rimuovere il calore in eccesso ed eliminare i
contaminanti e i detriti causati dall’usura. Sebbene gran parte dell’olio lubrificante
rimanga all’interno del compressore, un piccolo quantitativo, che varia fra lo 0.5 e il 3
per cento della portata totale di refrigerante del sistema, riesce a scappare dal separatore
posto a valle del compressore e circola attraverso l’intero sistema. Inoltre, l’accumulo
d’olio all’interno degli scambiatori di calore può causare un insufficiente ritorno di olio
al compressore e una carenza di lubrificazione. Gli effetti correlati all’aggiunta di olio
lubrificante negli scambiatori di calore sono negativi ma al tempo stesso inevitabili: la
presenza di olio determina un incremento delle perdite di carico e una penalizzazione
dell’intero processo di scambio termico.
L’obiettivo del presente lavoro è analizzare la ritenzione dell’olio lubrificante e i suoi
effetti in termini di scambio termico e perdite di carico in evaporatori a micro canali. Il
contributo originale consiste nel confrontare le prestazioni di un fluido refrigerante ben
noto e altamente impiegato come l’R410A e i suoi possibili sostituti a basso potenziale
di riscaldamento globale, DR5A, R32 e R1234yf.
Gli effetti dovuti alla presenza di olio lubrificante sono valutati confrontando i valori di
potenza scambiata e cadute di pressione con olio in circolo e quelli misurati nel caso di
refrigerante puro. Mantenendo costante la portata massica totale, è possibile indirizzare
tutte le variazione all’olio lubrificante che sostituisce il refrigerante puro.
Il lavoro sperimentale e di modellazione numerica dimostra che l’aggiunta di olio
penalizza sempre le prestazioni dello scambiatore di calore a micro canali in tutte le
condizioni di lavoro analizzate. La ritenzione dell’olio lubrificante, la penalizzazione
dello scambio termico e l’incremento delle perdite di carico sono direttamente
proporzionali alla frazione massica d’olio, parametro rappresentativo del rapporto fra le
portate massiche di olio e refrigerante in ingresso allo scambiatore di calore.
Parole chiave: evaporatore a micro canali, fluidi refrigeranti a basso potenziale di
riscaldamento globale, ritenzione dell’olio lubrificante, potenza termica scambiata,
cadute di pressione
Introduction
The present work provides an experimental and simulation study of lubricant effects on
heat transfer and pressure drop in microchannel heat exchanger evaporators. This thesis
focuses on the new environmentally friendly refrigerants and their performances with
oil, analyzing the behavior of three different new-generation refrigerants, R32, R1234yf
and DR5A, in comparison with a well-known and highly employed fluid, R410A, that is
the baseline of the study.
The presence of oil in refrigeration and air-conditioning vapor compression systems is
necessary for the proper operation of the compressor, which is the most important
component of the cycle. The oil has a fundamental role because it protects the
compressor mechanical moving elements with a thin lubricating film and increases the
durability of the overall structure thanks to its protection against wear. The lubricant can
also have secondary roles as limiting the noise, helping evacuation of chemical elements
and impurities that may be released in the system and sometimes it is used as a heat
transfer medium for cooling the compressor. The oil addition to the refrigerant
determines changes in flow configurations, thermodynamic equilibrium and
thermodynamic properties. Although the lubricant presence is essential, it tends to
decrease the heat transfer and increase the pressure drop of refrigerant in the two-phase
heat exchangers, in particular when the oil fraction is high, due to its large viscosity and
mass transfer resistance effect. Moreover, in some cases, the opposite heat transfer and
pressure drop trend has been observed at low-medium oil fraction around 2-3%. The
lubricant influence on refrigerant heat transfer and pressure drop is a difficult topic and
no consistent agreement has been reached yet in the scientific literature.
In a vapor compression cycle and starting from the compressor, the lubricating oil
typically comes in contact with the refrigerant as a mist, dissolves into the liquid
refrigerant in the condenser creating a mixture and then, after the expansion valve,
enters the evaporator. During the evaporation process, the oil remains in the liquid state
and leaves the evaporator as a liquid mist in the refrigerant vapor stream or it is trapped
in the evaporator tubes, especially in the high vapor quality zone. In refrigeration
systems the amount of oil is usually very low, around 0.5-3 wt.% of the total refrigerant
charge, but it has a very detrimental effect on the evaporator thermal capacity, since the
lubricant decreases the evaporating heat transfer coefficient, increases the two-phase
pressure drop, prevents all the refrigerant from evaporating and reduces the log mean
temperature difference. Another problem related to oil retained in heat exchanger tubes
regards the oil return to compressor: a consistent part of oil trapped or blocked in other
system components can affect the compressor reliability, causing a lack of lubrication.
2
The first part of the work consists in an experimental study of the lubricant addition
effect on a microchannel evaporator, through a heat transfer, pressure drop and oil
retention analysis. The results are presented as comparison between a fluid currently
used in refrigerating systems, such as R410A, and three possible new-generation low
Global Warming Potential replacements, R32, R1234yf and DR5A. The purpose of the
experiments is to determine if oil addition has similar effects on heat exchanger
performances while using different refrigerants and hence evaluate advantages and
disadvantages of the possible replacement fluids.
The second part of the work consists in a microchannel evaporator simulation model
development. Correlations and equations available from literature are implemented in
the Fortran heat exchanger numerical solver based on previous studies carried out at
Oklahoma State University. The simulation code has the aim of reproducing the
behavior of pure refrigerant or refrigerant and oil mixtures in microchannel evaporators,
in order to predict the heat transfer rate, pressure drop and oil retention. Moreover, since
the model is based on segment-by-segment method, it is possible to analyze the
thermodynamic properties in each point of the heat exchanger.
The thesis is organized as follows:
Chapter 1 deals with the literature review;
Chapter 2 presents the experimental apparatus description, the data analysis
procedure and the uncertainty analysis;
Chapter 3 contains the experimental results;
Chapter 4 treats about the simulation code description;
Chapter 5 shows the simulation results without oil in comparison with the
experimental capacity and pressure drop values;
Chapter 6 contains the simulation results with oil in comparison with the
capacity, pressure drop and oil retention experimental values;
Chapter 7 deals with the conclusions of the experimental and simulation works.
1 Literature review
In this chapter some important studies about microchannel heat exchanger geometry, a
description of the refrigerant tested and current limits in their utilization, works
regarding oil effect on refrigerant boiling and on vapor-compression system
components are presented. Lastly, some previous refrigerant and lubricant mixture
modeling works are introduced.
1.1 Microchannel heat exchangers
The purpose of increasing energy efficiencies and decreasing energy consumptions in
air-conditioning systems might be achieved by using microchannel heat exchangers. A
good overview of this technology is provided by Garimella [1] in a study about
innovations in energy efficient and environmentally friendly space-conditioning
systems. The microchannel heat exchanger is presented as a solution to increase the
energy efficiency and decrease the environmental impact, reducing the refrigerant
inventories and the size of the vapor-compression heat pump components. The
microchannel technology was initially developed in automobile air-conditioning
systems, where the microchannel tube and multilouver fin heat exchangers replaced
conventional evaporators and condensers (e.g. round tube, fin and tube heat
exchangers). Compared to round tubes, the microchannel heat exchanger configuration
presents a smaller frontal obstruction to air flow, that reduces the drag and fan power
and has a larger surface area per volume ratio, which results in compactness. The use of
microchannel heat exchangers in residential air conditioning systems permits size
reductions by a two-or-three factor compared to conventional heat exchangers. A more
compact geometry results in low refrigerant inventories: this fact implies a refrigerant
charge reduction up to one order of magnitude, which has a direct impact on the
reduction of global emissions. Garimella shows that the total material required for the
heat pump system based on two microchannel heat exchangers is only 36% of the round
tube system, and the refrigerant charge of the microchannel solution is 1.7 kg instead of
2 kg for the round tube system. Furthermore, a system using a microchannel
configuration provides equal heat capacity, but higher COP and smaller indoor and
outdoor heat exchanger frontal areas, with respect to conventional arrangements. On the
other hand, the microchannel solution needs a more careful preliminary design:
potential refrigerant maldistribution problems, which can be more severe in
microchannel evaporators than in larger round-tube evaporators, have to be minimized
to achieve the desired performance.
4
1.2 Current limits on refrigerant use
The Montreal Protocol [2] on substances that deplete the ozone layer, signed by all the
United Nations members in 1987, took on great importance in refrigerant development,
since it was designed to phase out the production of numerous substances that are
responsible for ozone depletion, such as halogenated hydrocarbons which contain either
chlorine or bromine. In the international Protocol, the Ozone Depletion Potential (ODP)
and the Global Warming Potential (GWP) are introduced as the two most important
parameters being used as a measure of how environmentally detrimental can be
refrigerants. The ODP of a chemical compound is the relative amount of degradation
that it can cause to the ozone layer, with trichlorofluoromethane (R-11 or CFC-11)
being fixed at an ODP of 1.0. ODP of a given substance is defined as the ratio of global
loss of ozone due to given substance over the global loss of ozone due to CFC-11 of the
same mass.
The Global Warming Potential is the measure of how much a given mass of greenhouse
gas is estimated to contribute to the global warming. It is a relative scale which
compares the gas in question to that of the same mass of carbon dioxide (whose GWP is
by definition 1). The GWP is calculated over a specific time interval, usually 20, 100 or
500 years.
It must be stated that the direct Global Warming Potential and the Ozone Depletion
Potential values are not the only predictor of environmental or climate change impact
for any refrigerant. Another important parameter, presented for the first time in the
report of the Montreal Protocol Technology and Economic Assessment Panel (1999), is
the Life Cycle Climate Performance (LCCP). The LCCP is a guideline referred to the
system overall environmental impact and represents a comprehensive metric to
calculate the equivalent mass of carbon dioxide released in the atmosphere
throughout its lifetime. The total is composed by two terms: the direct emissions
and the indirect emissions . The first term represents the
contributions due to refrigerant production and transportation, refrigerant leakage both
during system operation and at the end of useful life, accidents and maintenance. The
second term represents the emissions related to energy required for
manufacturing and recycling at the end of their useful life either system and refrigerant
and lifetime emissions due to electric energy consumption during system operation.
Therefore, through the LCCP value is possible to compare the overall environmental
impact in same technology systems, such as automobile air conditioning, residential and
commercial refrigeration and HVAC chillers.
5
In the interest of improved environmental sustainability, new generation refrigerants
have to maintain the quality of life and health benefits coming from air conditioning and
refrigeration, but in an energy efficient and environmentally sustainable way. The
replacement fluids must, on one hand, follow current fixed GWP, ODP and LCCP
limits and non-flammability values and, on the other hand, achieve the same
performances in refrigeration capacity and efficiency (COP) of the replaced fluids.
Thus, for every purpose and application, it is necessary to choose the fluid providing the
best balance of properties to meet together performance, environmental and safety
needs.
1.3 Refrigerants and lubricant tested
The refrigerants analyzed in this study are the most common HFC refrigerant, R410A,
which is treated as the baseline, and three environmentally friendly potential
replacements, DR5A, R1234yf and R32.
The HFC refrigerant R410A belongs to the first generation of working fluids intended
to replace CFCs and HCFCs. The R410A is a near-azeotropic mixture of 50 wt.% HFC-
32 and 50 wt.% HFC-125. The R410A critical temperature and pressure are respectively
and . It belongs to the safety group A1 (low toxicity and non-
flammability), its ODP is zero, but has a very high GWP (calculated for 100 years) of
2100. The letter “A” identifies the percentage of R32 and R125.
The first possible replacement fluid investigated is DR5A, a mixture of 68.9 wt.% R32
and 31.1wt.% R1234yf [3]. DR5A has chemical stability, it is not corrosive and it
belongs to A2L group. Its GWP is 460 and it is expected to have good compatibility
with POE oils. In Leck et al.’s work [4] about DR5A can be seen that the width of the
two phase dome is slightly higher than that of R410A. Through the PH diagram it is
also illustrated that the critical pressure of DR5A is close and a little bit higher than that
of R410A. Thanks to the higher critical temperature and the wider two-phase PH dome,
the loss in COP and capacity of DR5A is smaller than R410A while operating a cooling
cycle at hot climates. As observed in [4], DR5A can be considered a good R410A
replacement fluid. DR5A represents an example of refrigerant blend: adding HFCs as
R32 to R1234yf the volumetric capacity increases and the resulting blend works well in
air conditioning and heat pump systems. In any case the relevance of blended
refrigerants concerns the possibility of reaching energy efficiency and environmental
sustainability purpose when pure refrigerants cannot be used by themselves. Using
blended refrigerants it is possible to obtain high-performance mixtures and to reduce the
single pure component negative effects, as high GWP, flammability, high boiling point,
low volumetric capacity and temperature glide.
6
The second refrigerant investigated is R1234yf, a hydrofluoroolefin developed after the
strict regulations [5] [6] adopted in Europe to phase down the use of R134a in
automobile air-conditioning systems, beginning in the year 2011. Although R1234yf is
not considered a replacement for R410A, it is used in low GWP refrigerant blends, such
as DR5A. The R1234yf critical temperature and pressure are respectively 94.7 and
3.382 . It belongs to the mildly flammable and non-toxic group A2L, its ODP is
zero and its 100 years GWP is 4. The vapor pressures of HFO-1234yf and R-134a are
very similar and nearly overlay one another: at about 40°C the vapor pressures are
essentially the same, at lower temperatures the vapor pressure of HFO-1234yf is higher
than that of R-134a, and above 40°C the HFO-1234yf drops to less than that of R-134.
In [7], to demonstrate the compatibility and the thermal stability of HFO-1234yf with
typical materials used in refrigerating and air-conditioning systems and with most
common lubricants, tests at the standard conditions of 175°C for 14 days were
performed and no evidence of breakdown or reaction of the refrigerant with the metals
(aluminum, steel and copper) or the lubricants was seen.
The third replacement fluid tested is difluoromethane R32, an organic compound
composed of one carbon, two hydrogen and two fluorine atoms. R32 critical pressure
and temperature are respectively 5.38 and 78.4 . It belongs to the mildy
flammable and non-toxic group A2L, its ODP is zero and its GWP based on 100 year
time frame is 675.
The lubricant used in the present study is Emkarate RL 32-3MAF, which is an ISO VG
32 synthetic polyolester POE. This oil has additives less than 1% and a midpoint
viscosity of 0.032 at 40 . The POE ISO VG 32 is chosen by its interaction with
refrigerant molecules, since it does not react with refrigerant and it is completely
miscible with HFCs and HFOs.
1.4 Refrigerant-oil mixtures
A refrigerant-oil mixture behaves as a zeotropic mixture in which the refrigerant is
combined with miscible, lubricating oil. Some refrigerant-oil mixtures are miscible only
within a certain range of temperature or up to a certain oil mass fraction, so a part of the
refrigerant loop may pass outside the miscible range. The oil can be considered one
component in such a mixture, even though normally lubricating oil is a multi-
component mixture including various additives. A zeotropic mixture is a mixture that
never has the same vapor phase and liquid phase composition at the vapor–liquid
equilibrium state. Zeotropic mixture dew point and bubble point curves do not touch
each other over the entire composition range, with the exception of the pure
components.
7
1.4.1 Different approaches in studying refrigerant-oil mixtures
Three different approaches have been developed to study the influence of oil on
refrigerants boiling:
The oil “contamination” approach, which considers all properties of pure
refrigerant and oil as a contamination parameter. This approach does not lead to
proper results because it is not thermodynamically correct: it ignores the
influence of oil on the boiling point temperature, specific heat, latent heat,
viscosity, density and all the properties of the refrigerant-oil mixture.
The empirical approach, which consists in developing and validating models for
specific refrigerants with particular oils. Mermond et al. [8] identified the major
deficiency of this approach: every correlation requires a lot of experimental
measurements, with a large number of adjustable parameters, which limit the
application to a small range of experimental conditions. This approach can
describe only specific oil-refrigerant mixtures or blends with similar molecular
interactions.
The thermodynamic approach, which considers the refrigerant-oil mixture as a
real mixture. Even though this approach is difficult and necessitates lots of
correlations, a reduction in complexity is given since the vapor pressure of
lubricant is very little compared to the refrigerant one, therefore the oil affects
only the liquid phase and oil composition in the vapor phase can be considered
negligible.
The method to determine correctly properties of the refrigerant-oil mixture by using the
thermodynamic approach is described in Thome [9]. In his study he provides
relationships for the local bubble point temperature and for the temperature-enthalpy-
vapor quality curve. The thermodynamic approach consists in using the bubble
temperature instead of the pure refrigerant saturation temperature in the calculation of
the boiling heat transfer coefficient, since the refrigerant and oil mixture is treated as a
zeotropic mixture. Oil addition to refrigerant entails an increase in the bubble point
temperature and in the local saturation temperature at which evaporation takes place.
The local bubble point temperature should also be used, instead of the saturation
temperature of the pure refrigerant, to calculate the log-mean difference temperature.
An exhaustive discussion about the influence of the thermodynamic approach on this
study is given in section 4.5, regarding the modeling of the refrigerant and lubricant
mixtures.
8
1.5 Refrigerant-oil mixture flow characteristics study
1.5.1 Oil influence on Pool Boiling
The lubricant addition has a very negligible overall influence at concentrations below
3% in weight, while at concentration above 5% the oil contribution becomes substantial
in reducing the pool boiling heat transfer.
Some types of oil at concentration between 2% and 4% enhance pool boiling. A good
explanation of this phenomenon is given by Kedzierski [10], who states three possible
reasons for lubricant enhancing, relating the pool boiling mechanism to the variation of
bubble size and number:
1. The oil rich layer at the heated surface reduces the solid-liquid interaction, then
it entails a reduction in bubble size and an increase in bubble frequency;
2. The lubricant higher viscosity determines a thicker thermal boundary layer at the
heated surface, increasing the site density to activate the bubbles;
3. If the lubricant has a partial miscibility and the partial miscible refrigerant-oil
mixture boils close to the solution critical temperature, two liquid films can be
identified surrounding the bubble, where one is oil-rich and the other is
refrigerant-rich. The interface of the two films has a large pressure gradient,
which can move the superheated liquid to the bubble side, increasing the bubble
superheat and enhancing the nucleate boiling.
Kedzierski [11], using a fluorescent measurement technique, proves that the oil excess
layer decreases with increasing heat flux, since a greater heat flux activates larger
bubble site density and removes more lubricant. The removal rate is proportional to the
excess layer thickness and the bubble diameter. Kedzierski also proposes a refrigerant-
oil mixture pool boiling correlation in which the most important parameters are the
lubricant mass fraction, the difference between oil and refrigerant viscosity and the
difference between the refrigerant-oil mixture saturation temperature and the solution
critical temperature.
1.5.2 Oil influence on flow boiling
Many authors consider the pattern assumed during evaporating process a very important
parameter to describe the flow boiling characteristics, both in case of pure refrigerant
and refrigerant-oil mixtures.
One of the most noteworthy studies to understand the phenomenon marking out pure
fluids is the work by Collier and Thome [12], which considers a vertical tube heated
uniformly along its length at low heat flux, with subcooled liquid at the bottom inlet and
9
superheated vapor at the top outlet. Eight different flow patterns during the complete
evaporative process are identified. A single phase heat transfer mechanism occurs in the
subcooled region, followed by subcooled flow boiling when the wall temperature rises
above the saturation temperature. The nucleate boiling starts only in the superheated
boundary layer, with vapor bubbles condensing as they get in contact with the
subcooled core. Once the liquid reaches its saturation temperature, the flow pattern
becomes saturated nucleate boiling, in the form of bubbly flow regime and then slug
flow regime. Increasingly moving away from the inlet section, the flow pattern develops
into annular flow regime and annular flow with liquid entrainment in the vapor core. In
the annular region, the forced convection heat through the liquid film governs the heat
transfer. Subsequently, the annular liquid film dries out or is sheared from the wall by
the core vapor stream at the point referred as the onset of dryout region. Overtaken the
dryout, a mist flow pattern of entrained liquid droplets in the vapor current is identified,
which arises with a large increase in the wall temperature at constant heat flux. While
the two-phase temperature is rising above the saturation temperature, it is possible to
find droplets in the vapor stream until all the liquid evaporates and a single phase region
marked out by a single phase convective heat transfer mechanism starts. Collier and
Thome also propose a boiling map, depicting the heat transfer coefficient as a function
of vapor quality, in which higher heat transfer coefficients in wet regions than those in
the film boiling and liquid deficient region are observed.
Kattan et al. [13] propose a two-phase flow pattern model and a pure refrigerant map for
evaporation in horizontal tubes, based on flow pattern data for five different refrigerants
covering a wide range of mass velocities and vapor qualities. The work also shows an
equation for predicting the onset of dryout at the top of the tube, as a function of flow
parameters and mass flux. The model proposed by Kattan is subsequently improved by
Wojtan [14], who also investigates flow boiling in horizontal tubes. In the latter paper
the stratified-wavy region is analyzed with more accuracy and analytical forms for
transition from annular to dryout region and from dryout to mist flow zone are
proposed.
Another work about pure refrigerant flow pattern is Mishima and Hibiki’s [15]. Their
work is focused on typical regimes observed in an air-water two-phase flow inside
capillary vertical tubes with inner diameters in the range from 1 to 4 mm. Special flow
regimes peculiar of capillary flows are denoted. In bubbly flow, smaller bubbles form a
spiral train along the tube axis and larger bubbles, lined next to each other, form
intermittent bubble trains, without coalescing. In slug flow, slug bubbles, surrounded by
thin liquid film, are relatively long and have a significant spherical nose, effects of the
capillarity force. In churn flow, the long slug bubbles are deformed and loose the
spherical shape. Annular and annular-mist flows, do not present any appreciable
difference with respect to the flow patterns in large diameter tubes.
10
Considering a refrigerant-oil evaporating process, not all the over-described flow
patterns are observed. The effects of oil on physical properties have the potential to
modify the threshold of transition between different flow patterns. For example, in the
annular flow regime, the liquid film is prevented from drying out at the top of the tube
because the oil-rich liquid film bubble temperature rises towards that of the heated wall
and prevents the evaporation, delaying the dryout region. It should be underlined, as
mentioned in section 1.4.1, that the correct approach in determining refrigerant-oil
mixtures boiling heat transfer coefficient follows the thermodynamic approach; on the
contrary, some studies include combinations of real oil effects and others due to the
improper use of the pure refrigerant instead of , which properly characterizes
the mixture.
One of the best available work about oil influence on flow boiling is Shen and Groll’s
review [16]. The lubricant impacts on flow boiling through a flow pattern change: the
increased mixture viscosity and surface tension promote the early formation of annular
flow, especially in microfin tubes and with large mass fluxes. The foaming effect, if
present, increases the fluid volume, which is more effective to wet the heat transfer
surface. The lubricant influence on flow boiling is very dependent on the vapor quality.
At low and intermediate quality, when the flow is mainly stratified, the oil addition may
increase the wetted surface thanks to its high viscosity, surface tension or foaming; at
high vapor quality, for the same reasons, can be predominant a lubricant accumulation
effect. Whether the oil addition can improve overall evaporator performance or not, it
depends on the balance between the intensification of oil mass transfer resistance at
high quality and the increased foaming and wetted surface effect at low quality.
Manwell and Bergles [17] study the lubricant effect on flow boiling in two different
tube configurations analyzing the R-12/300-SUS oil mixture flow pattern both in
smooth and microfin tubes. Microfin tubes activate annular flow automatically although
the oil presence reduces this promotion effect and suppresses the foaming. In terms of
wetted surface and hence heat exchanged, the microfin tube shows advantages at high
and intermediate qualities, with high lubricant concentration and high heat flux. At
lower heat fluxes, the microfin tube better behavior over the smooth tube disappears
with the increase of mass flux.
Regarding the oil effect on thermodynamic properties, McMullan et al. [18] report on
the flow boiling heat transfer of R-12 mixed with three different lubricants. The
formation of annular flow is accelerated by higher mixture viscosity and increases oil
mass transfer resistance effect: a larger oil viscosity and surface tension decrease the
convective heat transfer both in the low quality and in the high quality region, but
change the stratified flow to annular flow in the first part of the evaporator. The
optimum lubricant concentration is determined by the trade-off between the reduction in
convective heat transfer and the improvement in wetted surface.
11
Cho and Tae [19] examine the effect of the mass fluxes. Higher mass fluxes make the
refrigerant-oil mixture more uniform, reducing the oil mass transfer resistance and
accumulation negative effects. Furthermore large mass fluxes activate the annular flow
pattern immediately.
Specific studies about lubricant effect on flow boiling heat transfer coefficient are
presented by Nidegger et al. [20] and Zurcher et al. [21] [22] [23]. The first work [20]
treats about evaporating tests for R-134a/oil in a microfin tube with an internal diameter
of 11.90 mm only for one mass flux value of 200 kg/m2s. The inlet saturation pressure
was 3.40 bar, corresponding to a saturation temperature of 4.4 °C for the pure
refrigerant. Adding oil always reduces the heat transfer at vapor qualities before the
peak in the heat transfer coefficient. After the peak, R134a/oil heat transfer data are
higher than the pure R-134a values, since the oil presence delays or eliminates the onset
of dryout region, depending on mixture properties. Lower heat transfer coefficients than
those of pure refrigerant are reported at high nominal oil mass fractions. Zurcher et al.
[21] [22] [23], in their flow pattern dependent methodology, suggest to predict the flow
boiling heat transfer coefficients of R134a/oil and R407C/oil by using the mixture
properties. In this way, well-predicted heat transfer degradation due to lubricant
presence is shown, but the results ignore the possible heat transfer enhancement caused
by a little quantity of oil. The flow pattern seen during refrigerant-oil flow boiling at
low qualities is stratified flow shifting to annular due to the increased mixture viscosity
and surface tension. As far as the flow maps are considered, the onset mass flux and
vapor quality of annular flow for refrigerant-oil mixtures are lower than those of pure
refrigerants. Thus, the model reveals the tendency of the lubricant presence to increase
the wetted surface, but does not predict properly heat transfer enhancement at low oil
concentrations and low vapor qualities since, during the transition from stratified to
annular flow, the increase in wetted surface is not large enough to compensate the
degradation in convective evaporation. Furthermore, an oil “hold-up” phenomenon at
high quality is reported in [22]. The hold-up is referred to a large amount of lubricant
trapped in the high vapor quality region. This accumulation is mainly present at low
mass fluxes in microfin tubes, while is much reduced at high mass fluxes in smooth
tubes. Lastly, Zurcher et al.’s studies suggest that the microfin tube takes advantage
over the smooth tube at high vapor quality and at high oil concentrations, since the
swirling effect of the microfin prevents the onset of dryout.
The oil effect on flow pattern and the oil film characteristics in refrigeration cycle
suction lines are analyzed by Fukuta et al. [24] and Sethi et al. [25]. Suction line studies
are significant especially for modeling the last part of the evaporator, in which
superheated vapor is present. The main aspect treated in the first work [24] is the oil
return during upward flows. The results, referred to air with 20 and 56 VG MO
lubricants, show that oil always flows upward even in case of low gas velocity. The
refrigerant gas core Reynolds number is observed to be the main parameter upon which
the flow pattern depends. Furthermore, a correlation between thickness, pressure
12
gradient and average velocity of the oil film and the Reynolds number of the refrigerant
gas core region is proposed.
Sethi et al.’s work is a quantitative comparison of oil retention and pressure drop
characteristics of R1234yf and R134a with POE oil inside horizontal, vertical and
inclined suction lines at a saturation temperature of 13 °C with 15 °C of superheat. The
effects of refrigerant mass flux, oil circulating ratio (OCR) and tube orientation on oil
retention and pressure drop are investigated. The oil retained is determined with a direct
measurement method, removing and weighing the test section. In their experiments,
OCR varies from 1 to 5%. R1234yf, in a same diameter suction line, has similar oil
retention, but 20 to 30% higher pressure drop than R134a, both in vertical and
horizontal suction lines. Furthermore, R1234yf oil retention increases as the mass flux
decreases or the OCR increases, due to its higher refrigerant vapor density compared to
R134a, which leads to lower refrigerant vapor velocity at the same mass flux. The oil
retention increases sharply in vertical lines as the mass flux is reduced below the point
of liquid film reversal, corresponding to the flow regime transition to churn flow. In
horizontal suction lines, oil retention is independent on flow regime transitions and
increases only at very low mass fluxes. Another interesting result observed is that
inclined suction lines retain more oil than the horizontal or vertical ones, with a
maximum value at an angle of inclination between 45° and 60°. Moreover, when
reducing the mass flux, for vertical lines the pressure drop reaches a minimum value
corresponding to the transition from annular flow to churn flow, while for horizontal
suction lines it decreases continuously. In both the orientations, increasing the OCR, the
pressure drop raises.
1.6 Effects of oil on refrigeration components
One important work about the oil effect on the components of a vapor compression
system is the one from Cremaschi [26]. Refrigerants R22 and R410A with miscible
lubricants, respectively blended white mineral oil BWMO and ISO VG 32 polyester oil,
are investigated over a wide range of OCR and refrigerant mass flow rates. The oil
retention in each component of the system is evaluated using a lubricant
injection/extraction methodology, which consists in determining the difference between
the amount of oil injected and the one extracted, once the steady state is reached.
Considering The R22/MO pair, if the OCR increases from 1 to 8 wt. %, the oil retention
in the suction line increases from 2% to 28% of the initial mass of oil charged into the
compressor. While only few percentage of the initial lubricant charge is retained in the
evaporator, the cumulative oil retention in liquid line, evaporator and suction line rises
up to 40%. Analyzing the cumulative oil retention mass in all the components at various
refrigerant mass flow rates and different OCR, the oil retention is observed strongly
13
dependent on the oil concentration in the mixture but not so much affected by the
refrigerant mass flow rate. However, at high refrigerant mass flow rates, lower oil
retention values are registered, except for the liquid line, where oil and refrigerant are
homogeneously mixed. When OCR increases from 1% to 8 wt. %, the pressure drop
raises from 10% to 40% in the suction line and from 2% to 15% in the evaporator, with
respect to the case without oil. The pressure drop mainly affects the suction line, due to
less mutual solubility between the refrigerant core and lubricant film and hence higher
values of liquid film viscosity. Similar trends with different values are observed for the
R410A/POE pair. Comparing the oil retention volume between the two different pairs at
similar flow rate, the R410A/POE mixture presents higher oil retention volume in the
suction line and evaporator, due to higher film viscosity.
Another interesting work is Cremaschi et al.’s [27] extensive experimental study of oil
retention in air conditioning system components, which is focused on the suction line,
the most critical component for oil retention due to highest liquid film viscosity and low
inertia force of the vapor refrigerant core. The ratio between inertia force and viscous
force represents the transport driving force for the refrigerant and oil mixture: the higher
the inertia force, the more easily the oil is carried over the system and the lower the oil
thickness is. The described simulation model calculates the force ratio and correlates it
to the oil film thickness. The refrigerant and oil combinations used as working fluids
are: R22/MO, R410A/MO and R410A/POE, R134a/POE and R134a/PAG. The oil
retention volume is shown proportional to the refrigerant-oil mixture OMF in each
component and higher mass fluxes reduce the amount of oil retained in the suction line.
Furthermore, a reduction of pipe diameter promotes the lubricant transportation, but at
the same time increases the frictional pressure drops along the pipe. Thus the suction
line pipe diameter has to be determined minimizing both the pressure drop and the oil
retention. Comparing different pipe orientations at the same refrigerant mass flux and
liquid film viscosity, gravity effects are observed to be important in vertical upward
flow, since they are able to increase the oil retention up to 50% more. In the vertical
upward flow suction line a critical phenomenon regarding liquid film flow reversal
appears: when the refrigerant mass flux is decreased below a threshold known as the
critical refrigerant mass flux, the liquid film starts becoming instable and reverses its
flow, increasing pressure drop due to the cross-sectional area restriction at the suction
line bottom caused by the oil accumulation. Moreover, higher oil retention values are
observed when the mixture viscosity ratio (the ratio of liquid film viscosity over
refrigerant vapor viscosity) increases and the degree of solubility and miscibility
between oil and refrigerant decreases. The lubricant retained, which increases pressure
drop and reduces heat transfer coefficient, also reduces the COP and the cooling
capacity of the system.
A study about lubricant effects on two-phase refrigerant distribution in microchannel
evaporator is made by Li and Hrnjak [28], using R134a as working fluid. Their work
improves the Tuo et al.’s [29] experimentally validated microchannel evaporator model,
14
which only considers pure refrigerant maldistribution. The lubricant viscosity and the
oil circulating rate (OCR) effects on refrigerant distribution are analyzed in a single tube,
a simplified parallel tubes and microchannel evaporator models. The results, obtained at
fixed mass flow rate and evaporator outlet pressure, show that, as lubricant viscosity
increases, refrigerant distribution worsens and there is not a univocal behavior after
increasing OCR. In fact, as OCR increases from 0.1% to 3%, the distribution becomes
worse, while distribution appears to be slightly more uniform when OCR increases from
3% to 10%.
1.7 Previous modeling works
An interesting segment by segment heat exchanger model is proposed by Jiang [30].
This approach is able to take into account within each tube on one hand, the non-
uniform air distribution across the heat exchanger with its consequences on heat transfer
coefficient and on the other hand, the two-phase regime and different pure refrigerant
flow patterns inside tubes. The air-to-refrigerant heat transfer and refrigerant pressure
drop are calculated for each individual segment. Inlet enthalpy, pressure and mass flow
rate on the refrigerant side, and inlet air temperature on the air side are required as input
parameters for predicting segment outlet conditions. The predicted conditions at the
outlet of the first segment are passed as input for the consecutive and so on until the
refrigerant circuitry is completed.
One remarkable work about modeling the overall vapor compression system and its
components and their response to an increase in the amount of circulating oil is the one
from Lottin et. al [31] [32]. Their refrigeration system simulation software, developed
for a refrigerant-oil mixture composed by R-410A and ISO 32 POE synthetic oil,
considers the effects of lubricant addition on the thermodynamic properties of the
mixture and the changes in refrigerant-oil physical properties. In the first part [31], the
liquid-vapor thermodynamic equilibrium (VLE) curves of the R410A-POE mixture are
extrapolated from the work of Henderson [33], which gives the saturation temperatures
of R125-oil and R32-oil mixtures, the two refrigerants R410A is composed of. Then a
polynomial function which takes into account the blending process to form R410A from
the two couples R32-oil and R125-oil is built. Furthermore, a method for calculating
refrigerant-oil mixture physical and thermodynamic properties is described. Assumption
for the compressor modeling are provided: on one hand, the possible change of phase of
the refrigerant dissolved in the oil between the suction and discharge does not affect the
volumetric efficiency; on the other hand, both the isentropic and the electro-mechanical
efficiency are functions of the amount of lubricant circulating into the compressor. In
order to verify the convergence, the model goes on running until the first law of
thermodynamic is respected (balance between the condenser, evaporator and
15
compressor powers exchanged) and when two successive iterations do not present any
significant variations. The pressure drops in the pipes are estimated with the
homogeneous model and the friction coefficient is evaluated through Blasius
correlations. The changes in the pressure due to gravity and momentum effects are
neglected. Comparing simulation and experimental data, the numerical model works
very well when the ratio between high and low pressure is low, while the discrepancy is
larger with higher pressure ratio. The difference between evaluated and measured power
at the compressor and at the evaporator never exceeds 10%. Increasing the quantity of
lubricant from 0 to 0.5%, the software catches both a small increment in evaporator and
condenser heat transfer rate and a small reduction in refrigeration and heat pump COP.
Oil effect is more detrimental when its amount reaches 5% in weight. At high oil mass
fractions, the remaining liquid at evaporator outlet is a mixture of oil and diluted
refrigerant that, while not evaporating does not produce anymore the required thermal
effect and causes an evaporator heat transfer to fall (up to 13%). With a large amount of
oil the model is able to account the increase in compressor power, caused by two
different physical phenomena: on one hand, the high heat capacity of the lubricant and
the energetic cost induced by its heating; on the other hand, the energy required by the
vaporization of refrigerant, since the quantity of refrigerant diminishes between the
compressor suction and discharge ports.
The second part of Lottin et al.’s [32] study is more focused on modeling the heat
exchangers response to the increase circulating oil. The evaporator considered has a
plate geometry and the model is based on splitting the heat exchanger into elementary
segments, where the heat transfer coefficients and thermodynamic and physical
properties of the fluids are treated as constant. The resulting best compromise between
accuracy, numerical stability, and computation time of the simulations is 500 segments
per channel. Three different correlations for the refrigerant side heat transfer coefficient
are used to investigate its evolution when liquid evaporates and to study its variations
when the amount of oil mixed with the refrigerant is varying:
Gungor and Winterton [34] correlation, which gives good results even if it is not
well adapted to plate heat exchangers since it is established for flows inside
smooth tubes;
Yan and Lin [35] correlation, derived from the evaporation of R134a in a plate
heat exchanger;
Bivens and Yokozeki [36] correlation, which represents the best adequacy
between the numerical and experimental results, with the device of use Yan and
Lin form for the liquid evaporative heat transfer coefficient.
All the results are presented as function of the normalized coordinate along the channels
and oil mass fraction. The predicted mean and local refrigerant side heat transfer
coefficients in the evaporator on one hand have similar trend but with very different
numerical values if calculated using Bivens and Yokozeki and Yan and Lin correlations,
16
on the other hand they have similar average but very different local values if calculated
using Bivens and Yokozeki and Gungor and Winterton. The model is able to catch the
enhancement of the average value of the heat transfer coefficient at OMF equal to 0.1%.
The refrigerant-oil mixture predicted temperature shows two delimited zone: the first
one with very weak temperature changes characteristic of evaporation and the second
one representing the overheating of pure refrigerant. Increasing OMF, the clear limit
between the two zones disappears. Moreover, the predicted quantity of lubricant which
remains in liquid phase during evaporation and the predicted vapor mass fraction in the
evaporator are presented. At the evaporator outlet, the liquid composition is quite
independent from OMF increase, while the quantity of liquid cannot be considered
negligible, in particular at low temperature and high oil mass fraction, when the
solubility of refrigerant becomes more important. Furthermore, the model provides a
pressure losses quantification: they are not significant with oil mass fractions below
0.5%, while become very important at higher OMF values: for example at 5% OMF, the
pressure drop was multiplied by 4 to 6.5 compared to its initial value without oil.
Implementing an infrared thermography based method to describe the quality
distribution in the inlet header, Li and Hrnjak [37] develop a microchannel heat
exchanger model, which takes into account both thermodynamic and transport
properties of refrigerant-oil mixture and lubricant impact on boiling heat transfer and
pressure drop. The lubricant effect on flow resistance is implicitly counted using
mixture properties in heat transfer and pressure drop correlations in each microchannel
tube. The finite volume approach used in modeling the entire heat exchanger and all the
assumptions are described in another work from Li and Hrnjak [38]. Since the
refrigerant-side capacity in the two-phase region is mainly the latent heat of the liquid
refrigerant, the infrared quantification method consists in building the relationship
between the liquid mass flow rate through each microchannel tube and the
corresponding air-side capacity in the two-phase region. The vapor mass flow rate is
then determined from the mass conservation equation of both phases. The refrigerant-oil
mixture model results are validated against experimental data and compared to two
different pure refrigerant models. The lubricant model is much better in predicting both
capacity and pressure drop, especially at OCR higher than 5%, while its superiority is
not obvious in superheat prediction. The infrared quantification model catches the
beneficial lubricant effects on distribution and predicts the capacity more precisely.
Another interesting study is described by Jin and Hrnjak [39]. The semi-empirical
model, developed with R134a-PAG46 oil and R1234yf-PAG46 oil, predicts refrigerant
and lubricant inventories in both heat exchangers of an automotive air conditioning
system. The evaporator has a fin and plate geometry, 4 passes and 19 plates with
horizontal headers, while the microchannel condenser is a two-pass design, with 31
channels in the first pass and 17 in the second. Each channel has a hydraulic diameter of
1 mm and the headers are vertical. The model divides the two heat exchangers into
small volume elements, whose number is determined in order to have any further
17
increment in discretized elements effecting on calculation less than 0.5%. Temperature,
pressure and mass flow rate and oil circulation ratio (OCR) measurements at the inlet of
the evaporator or condenser are the input parameters for the simulation; in each volume
temperature, pressure and mass inventory are calculated by applying heat transfer,
pressure drop and different void fraction correlations; the calculated temperature and
pressure at the element outlet are used as the inlet condition for the next one. The heat
transfer between air and refrigerant is modeled using the Effectiveness-Number of
Transfer Unit (ε-NTU) method. The correct thermodynamic approach is used to study
the refrigerant-oil mixture: the two components are analyzed as a zeotropic mixture,
whose temperature moves along the dew point line during condensation and along the
bubble point line during evaporation. The evaporator capacity is predicted within 10%
error and lubricant and refrigerant mass inventories are both predicted within 20% error
with all the different void fractions correlations. The condenser model estimates the
capacity and the refrigerant mass inventory with an average error less than 10% and
15% respectively, while the oil retention is much under estimated, especially at oil mass
fraction around 4%. In order to have a better prediction of the lubricant inventory, the
condenser model is modified accounting for the inlet vertical header that behaves as an
oil separator, due to the gravity and refrigerant shear stress and that the oil, separated
from the flow, starts to fill up the bottom channels. The lubricant does not change phase
in the condenser and is cooled down faster than the two-phase fluid. After dividing the
condenser into vapor and liquid channels and adding to the model the different possible
distribution in each type of channels, the simulation well predicts both the refrigerant
and the lubricant mass inventory.
2 The experimental apparatus
The facility was built by two previous students at Oklahoma State University, Pratik
Deokar and Ardiyansyah Yatim, with the purpose of testing the oil retention effects on
microchannel condensers. At the end of the condenser experimental series, the
apparatus was modified in order to obtain the same oil retention analysis on
microchannel evaporators. The description of the condenser test facility provided by
Deokar [40] is used as a reference.
2.1 The Psychrometric Chamber
The OSU Psychrometric Chamber is an environmental simulator consisting of two
similar adjacent air conditioned rooms allowing test conditions such as temperature,
humidity and air flow rate to be controlled over a wide range. It is able to maintain
conditions with the addition of internal thermal loading. One of the two rooms
artificially reproduces the outdoor climate conditions, while the other room simulates
the indoor environment and replicates indoor comfort conditions. Each room can
operate independently in order to run a number of experiments requiring controlled
ambient conditions. In this work only the outdoor room is used. The outdoor chamber
reproduces the outdoor weather having a design temperature range from -40 °C to
+54.5 °C. The uniform air flow condition and distribution are ensured through a supply
plenum under the perforated floor and a ceiling return plenum. Furthermore, the air flow
rate can be adjusted using variable speed blowers and electro-mechanical dampers. The
desired conditions are achieved by conditioning the air and then circulating it into the
room. A schematic cross section and the position of various components and
temperature, pressure and relative humidity transducers are reported in Fig. 2.1.
20
Fig. 2.1: Air conditioning loop and evaporator inside the psychrometric chamber
The cooling and dehumidification processes are realized by a set of coils, fed with a
mixture of ethylene glycol and water. An additional refrigerating chiller, connected
through a heat exchanger to the water and glycol loop, is provided in order to increase
the cooling and dehumidification required. The coils surface temperature is controlled
by variable speed pump, electronic mixing valves and electro-mechanical bypass
valves. After being cooled and dehumidified, the air passes over electric heating
resistances, which increase the air temperature up to the desired value. The outdoor
chamber also has one air flow measurement apparatus, the code tester, located between
the two conditioning loops. The code tester is equipped with air dampers, air diffusers,
elliptical flow nozzles and fans; the diffusers generate the air stream for static pressure
measurements, while the flow nozzles are used to determine the air flow rate up to
13600 m3/h. The variable frequency driven blower makes possible to have always the
same air flow rate by changing its speed. The volumetric air flow rate through
a single nozzle is determined according to ASHRAE Standard [41]:
(2.1)
21
(2.2)
Where the parameters used are defined as follows:
Coefficient of discharge for the nozzle
Exit area of the nozzle
Pressure difference across the nozzle
Adjusted specific volume of the air at the
nozzle
Specific volume of the air as measured
from wet and dry bulb temperatures
Pressure at the nozzle throat
Humidity ratio at the nozzle
2.2 The microchannel Evaporators
Two different microchannel evaporators are tested, called respectively evaporator A and
evaporator B. Both the microchannel heat exchangers are installed in a duct inside the
outdoor psychrometric room, in order both to control the condition of the air flowing
across the evaporator and to ensure the same air velocity profile on the entire slab. The
evaporator B is installed in the duct upstream the evaporator A. As the heat exchangers
are tested separately, a ball valve permits to isolate the heat exchanger not used. The
refrigerant liquid and vapor lines enter the room through its wall and are respectively
connected at the inlet and at the outlet of the evaporator. The oil line also passes across
the wall and two different valves permit to realize the oil injection at the inlet or at the
outlet of the microchannel heat exchanger. Four differential pressure transducers are
installed, each one with a distinct differential pressure range, with the aim of measuring
in the more accurate way the refrigerant side pressure drop between the inlet and the
outlet of the heat exchangers. The refrigerant side line is also equipped with absolute
pressure transducer and inline thermocouples placed both at the inlet and at the outlet of
the evaporator. To measure the outlet air temperature distribution, a grid of 20
thermocouples is placed 3 cm after the microchannel heat exchanger slab. The
geometric characteristics of the different evaporators are described in Table 2.1 and
schematic representations are given in Fig. 2.2 and Fig. 2.3.
22
Table 2.1: Main parameters of the two evaporators
Parameter Evaporator A Evaporator B
Coil length [mm] 884 501
Coil height [mm] 438 546
Coil depth [mm] 25.4 30.5
Number of pass 1 2
Number of tubes 98 50
Tube thickness [mm] 0.35 0.4
Fin height [mm] 7.44 7.62
Fin pitch [fin/mm] 0.787 0.63
Hydraulic diameter [mm] 1.36 0.87
Overall free flow area [mm2] 1234.8 1100
Material aluminium aluminium
Header diameter [mm] 31.75 33
Header to header length [mm] 924.5 536.575
Header to header height [mm] 508 574.7
Evaporator volume [dm3] 1.526 1.9
Fig. 2.2: Microchannel evaporator A: (top) actual picture of the heat exchanger and (bottom)
scheme of the evaporator with refrigerant and air flow direction during the tests
refrigerant flow direction
Refrigerant out
Refrigerant in
refrigerant flow direction
Refrigerant out
Refrigerant in
H
W
Air flow direction
23
Fig. 2.3: Microchannel evaporator B: (top) scheme with refrigerant and air flow directions
during the tests; (bottom left) scheme of the evaporator with indication of the main dimensions
and (bottom right) actual picture of the heat exchanger
2.3 The air sampling device
The air sampling devices, placed before and after the microchannel heat exchangers, as
illustrated in Fig. 2.1, are built according to ANSI/ASHRAE Standard 41.1 [42]. The
two constructions are very similar and composed by a horizontal PVC sampling tree
with vertical perforated branches, perpendicular to the air flow. The trees assist to
collect small samples of air over a large region, to mix and then transport them through
a flexible duct to the dry bulb and wet bulb temperature measuring RTDs. An inline
centrifugal fan overcomes the pressure drop in the flexible duct. The air sampling
device placed before the microchannel heat exchanger is shown in Fig. 2.4. It is also
possible to notice the blue air tunnel built in front of the evaporators to guide the flow
and have the same velocity profile on the entire surface of the microchannel heat
exchangers.
W
H
Refrigerant in
Refrigerant out
Refrigerant in Refrigerant out
Air flow direction
refr
iger
ant
flo
w
dir
ecti
on
Refrigerant in
Refrigerant out
24
Fig. 2.4 Air sampling device in front of the microchannel evaporator
2.4 The refrigerant and oil loops
The refrigerant and oil loops are the main components of the oil retention experimental
facility and its schematic is shown in Fig. 2.6. The refrigerant loop is equipped with a
gear pump GC-MC25.JVS manufactured by Micropump, which can supply fluid at a
rate of 1.82 liters per 1000 revolutions at a maximum differential pressure of 860 kPa.
Its rotational speed is depending on the alternating voltage frequency supplied by the
Variable Frequency Drive VFD. The VFD model VS1SP21-1B, configured for the
motor of the gear pump described above, is manufactured by Baldor Electric Company
and requires a 3-phase input of 230 Volts at 60 Hz. Lastly, the electric motor of the gear
pump, model CEM3545 manufactured by Baldor Reliance Super-E motors, has 0.75
kW nominal power and can rotate at 3450 rpm. After the gear pump, the refrigerant
flow encounters a filter dryer, model C-083-S-HH 3/8 manufactured by Parker Hannifin
Corp. Sporlan Division, which is able to remove moisture dirt, acid and sludge from the
liquid refrigerant. Finally, the flow enters a Coriolis mass flow meter (model CMF025
by Micro Motion Inc.), which measures the mass flow rate transferred by the refrigerant
gear pump.
After the Coriolis mass flow meter, the refrigerant encounters a series of electrical
heaters. The first one, with respect to the flow direction, is a variable power electrical
heater. The variable power is given by a variac voltage regulator, which can change the
25
supply voltage to provide up to 600 W. Refrigerant side temperature and pressure
sensors are placed just after the variable power electrical heater, in order to obtain a
reliable enthalpy measurement, since at that point subcooled liquid refrigerant is
flowing into the system. The heat provided by the first variable power heater to the
refrigerant is usually not enough to reach the saturation conditions required at the inlet
of the microchannel heat exchanger, so two additional fixed power electrical preheaters,
each one able to supply 300 W, are installed. The enthalpy at the inlet of the
microchannel evaporator is then calculated adding the heat provided by the preheaters
to the heat gain transferred from the surroundings to the refrigerant. The procedure is
explained in detail in section 2.6.1.
After the saturation conditions are reached, the refrigerant passes through the
microchannel evaporator. At the heat exchanger exit, one helical and one coalescent oil
separators, with the purpose of removing all the lubricant droplets entrained in the
refrigerant vapor bulk, compose the oil extraction system. The separators limit the
possible oil recirculation all over the loop. Both the helical separator, model # S-5188
by Henry Tecnologies Inc., and the coalescent one, model # 925R manufactured by
Temprite (able to detach up to 0.05 microns particles), do not have floating valve, to
avoid sticking problem of internal valves against the surface and lubricant drains. Since
the separators work with very favorable conditions, that is oil and vapor refrigerant
mixture at intermediate pressure around 850 kPa, the separation efficiency is accounted
to be unitary. The separators have three different vertical sightglasses, with the aim of
monitoring the internal lubricant level.
Once the refrigerant is purified from oil droplets, the vapor refrigerant stream
encounters a plate heat exchanger (model GB400L-14 by GEA), where it is cooled
down and turns into subcooled liquid refrigerant, required condition at the inlet of the
refrigerant gear pump. In the plate heat exchanger low temperature side R404A
refrigerant is flowing, which, in turn, is cooled down by a chiller. The chiller, model
CPCW-12LT/TC2-1-9X2 manufactured by Cooling Technology Inc., supplies up to 7
kW (2.0 tons) capacity with R404A leaving temperature of -31.67 °C. The secondary
coolant used in the chiller is Dynalene HC40.
All the liquid line from the outlet of the plate heat exchanger to the inlet of the
microchannel evaporator is made with copper tube 10.9 mm diameter size (3/8”), while
the vapor line from the outlet of the microchannel evaporator to the inlet of the plate
heat exchanger is realized with copper tube 16.7 mm diameter size (5/8”).
A particular of the refrigerant loop, depicting the gear pump, the Coriolis mass flow
meter, and the recovery machine with the manifold used to charge the refrigerant, is
shown in Fig. 2.5.
26
Fig. 2.5: Particular of the refrigerant loop
Before and after the oil injection, the lubricant loop is distinct and independent from the
refrigerant circuit. The main components in the oil loop are the variable speed gear
pump, model #1802R/56C manufactured by Micropump, controlled by VFD to supply
the desired mass flow rate, and the reservoir tank, model AOR-4 by Emerson Climate
Technologies, where the oil is stored. The oil loop is connected to the main refrigerant
loop through an injection port, consisting of a small diameter copper tube that links
perpendicularly to the refrigerant line. The small copper tube is equipped with one
needle valve at the inlet (near the oil loop), which regulates the injection through its
opening or closing, and two different ball valves placed at the outlet (near the
intersection with the main refrigerant loop), which permit to choose between inlet or
outlet injections. A certain amount of pure refrigerant is directly injected in the oil
reservoir to pressurize it and keep a pressure difference between the two loops around
140-210 kPa, value able to maintain the lubricant flow stable among all the injection
time. The small diameter copper tube configuration, counting six sharp elbows, is
realized in order to mix as best as possible the refrigerant and oil mixture before
entering the refrigerant main loop. Furthermore, the lubricant loop is equipped with a
Coriolis mass flow meter and an electrical tape heater surrounding the oil reservoir. The
mass flow meter helps in monitoring outright the exact mass flow rate of refrigerant-oil
mixture injected into the main refrigerant loop, while the electrical tape is used
periodically to boil out the refrigerant from the liquid oil and then avoid, increasing the
temperature and decreasing solubility, too high concentration of refrigerant dissolved.
The schematic representation of the refrigerant and oil loops is given in Fig. 2.6.
27
Fig. 2.6: Schematic representation of refrigerant and oil loops
28
2.5 Test procedure
Several different ways, whose effectiveness largely changes depending on the
refrigerant tested, are employed to reach the desired test conditions:
The first parameter investigated is the total refrigerant charge in the system.
Charging more refrigerant in the system mainly increases the pressure and then
reduces the superheated vapor temperature. Before turning on the system, it is
important to charge a proper amount of refrigerant, since even little quantity of
refrigerant can have a very detrimental effect on reaching the desired conditions;
The refrigerant mass flow rate is varied using the variable frequency drive of the
pump, which mutates the electric motor rotational speed and provides the round
per minute desired;
Another way to reach the conditions is changing the set point temperature of the
dynalene flowing in the chiller. The dynalene is cooled down in the chiller and
then heated through electric heaters to reach the set point temperature desired.
Low dynalene temperature ensures a satisfactory refrigerant subcooling, which
avoids pump cavitation. A reduction in the dynalene temperature is effective in
decreasing the liquid refrigerant temperature and pressure;
The parameters at the inlet of the microchannel heat exchanger are also
depending on the electric power entering the system through the inline variable
power electric preheater. Modifying the variac voltage is possible to keep the
refrigerant to the required conditions. Increasing the variac voltage and then the
electrical power, the temperature at the inlet of the microchannel and the
pressure of the overall system increase;
The last option analyzed is changing the outdoor room temperature. An increase
in the outdoor chamber set point temperature results in a growth of the capacity
exchanged between air and refrigerant, that means a raise in the outlet
temperature of the microchannel and in the overall pressure of the system.
Through the code tester blower is also possible to vary the air mass flow rate
across the heat exchanger.
Depending on the purpose of the experiments, not all the ways described above can be
performed. For example, to compare the performances of different refrigerants under
the same outdoor room air dry bulb temperature conditions, it is necessary not to vary
the air side parameters (temperature and mass flow rate of the air at the inlet of the heat
exchanger).
Along all the experimental series, the conditions at the microchannel heat exchanger
inlet are determined from the subcooled refrigerant conditions at the fixed power
29
preheater inlet through a heat balance. A direct measurement from the temperature and
pressure transducers at the evaporator inlet is avoided, since it can lead to not accurate
results in case of two-phase refrigerant flow. The refrigerant at the preheater inlet has at
least 5 °C of subcooling. Adding the heat gain received from the surroundings to the
fixed electrical power from the preheaters, the enthalpy and then the quality at the inlet
of the microchannel are determined. The refrigerant uniform distribution inside the heat
exchanger is guaranteed by the uniform air side conditions both before the evaporator,
where a duct in front of the heat exchanger guides the flow to have the same velocity
profile on all the surface area, and after the microchannel, where through a
thermocouple mesh is possible to check the symmetry of the air temperature.
Once the conditions are stable for at least 20 minutes, in order to exclude possible late
transients, the oil injection procedure starts. This process consists in injecting into the
refrigerant loop a certain amount of oil for a well defined time interval, depending on
the oil mass fraction chosen. The lubricant is forced into the refrigerant loop thanks to a
pressure difference. To increase the pressure difference between the two loops, a small
quantity of refrigerant is charged in the oil loop; the higher the oil mass fraction, the
higher is pressure difference required to get the lubricant injection. Through a Coriolis
mass flow meter is always possible to monitor the oil mass flow rate entering the
refrigerant loop and moreover, using a metring valve, it is possible to vary the flow. To
avoid excessive pressure swings between the oil and refrigerant loops, the oil mass flow
rate is increased up to the desired value only after opening the injection valves. The
injection lasts about 15 minutes, enough to reach again stable conditions; with low oil
mass fractions the duration is few minutes longer. Two sightglasses are installed on the
vapor side of the refrigerant circuit, with the aim of observing the fully developed flow
at those points of the loop. At the end of the injection, a sample of oil is taken from the
oil loop and a gravimetric measurement of the refrigerant solubility in the oil injected is
provided, in order to calculate the correct oil mass fraction circulating into the system.
The procedure to calculate the refrigerant solubility is the following:
The sampling cylinder is cleaned from oil droplets, vacuumed and weighted
(M0);
The sampling cylinder is filled with oil and refrigerant mixture taken from the
lubricant loop. Then is weighted again (M1);
The sampling cylinder is opened to the atmosphere, so the refrigerant boils and
comes out through an expansion valve. After two hours the cylinder, which
contains only oil, is weighted again (M2).
The procedure to calculate the oil retention and its effects on heat transfer and pressure
drop consists of inlet and outlet injection tests. In the first group of tests, the lubricant is
injected at the microchannel evaporator inlet, with the purpose of obtaining a global
performance overview in terms of Heat Transfer Factor (HTF) and Pressure Drop
Factor (PDF) of the heat exchanger when the working fluid is a refrigerant-oil mixture.
30
In the outlet injection tests, the lubricant is injected at the microchannel evaporator
outlet, in order to measure the amount of oil retained in the refrigerant line between the
heat exchanger exit and the oil separators, and obtain a more accurate value of the oil
retention volume representative only of the microchannel heat exchanger. The outlet
tests have the same temperature, pressure and mass flow rate conditions at the
evaporator outlet measured during the inlet tests.
Once the injection period is over, in order to remove the oil from the system and store it
in the two oil separators, flushing cycles are performed. A flushing cycle consists in
reducing the air flow rate decreasing the code tester fan speed and in increasing the
refrigerant pump rpm, to obtain two-phase refrigerant at the outlet of the heat exchanger,
since liquid refrigerant is more efficient than vapor in carrying over lubricant droplets.
Cleaning the evaporator from the oil is an essential operation to demonstrate that the air
capacity does not change within the experimental series with the same air and
refrigerant side conditions.
2.6 Data analysis
All the data are collected through sensors connected to National Instruments Data
Acquisition (NI-DAQ) system with Real Time Labview Graphic Software Interface.
The sensors measure temperatures, pressures and mass or volume flow rates of the air,
refrigerant and oil sides. The data read from sensors are displayed, plotted and recorded
every two seconds. An example of the graphic interface of Labview is given in Fig. 2.7
and Fig. 2.8
31
Fig. 2.7: The Target Labview interface for oil retention tests
Fig. 2.8: The Host Labview interface for oil retention tests
The Labview control and graphic interface for oil retention tests consists in two
different screens: the Target and the Host. The first one is employed to provide the
inputs for the air, refrigerant and oil loops, while the latter is used to plot the sensors
32
values as function of time, in order to identify steady state conditions and the perfect
timing to start the oil injection test.
Few assumptions in data reduction are accounted. For example, the refrigerant solubility
in the oil injected, mainly dependent on temperature and pressure, is taken as constant
for all the injection period. This hypothesis can be considered valid since the oil
reservoir used in the experiments is large enough to have only little and negligible
variation in both pressure and temperature. Another important assumption is evaluating
the oil separators efficiency as unitary, that means all the oil injected is collected in the
separators and no lubricant recirculation is present in the system. The sightglass placed
at the microchannel evaporator inlet, where oil is never observed during normal
operation, demonstrates the goodness of the last assumption. These two assumptions
permit to calculate easily the oil mass fraction, thus only dependent on the mass flow
rates read from the mass flow meters and solubility, through equations (2.3) to (2.6):
Where is the solubility, the subscript refers to the oil and refrigerant in the
injected mixture, while the subscript symbolizes the pure refrigerant mass flow rate
at the steady state conditions reached before the beginning of the injection test.
2.6.1 Heat transfer and pressure drop analysis
First of all, both the consistency of the air side capacity and the heat balance across the
heat exchanger are calculated for each experiment.
The consistency of the air side capacity is ensured comparing the values obtained
before and after flushing within each experimental series with same air and refrigerant
side conditions. Negligible variations are found.
(2.3)
(2.4)
(2.5)
(2.6)
33
The heat balance without lubricant, schematically represented in Fig. 2.9, is verified
through equations (2.7) to (2.10).
Fig. 2.9: Schematic representation of fixed power electrical preheaters and microchannel
evaporator
Where:
represents the air side capacity, calculated as the enthalpy variation
across the microchannel heat exchanger. The volumetric air flow rate passing
through the evaporator is determined from the pressure drop on the nozzle
placed in the code tester, following ASHRAE Standard as indicated in section
2.1;
is the heat gained by the refrigerant between the fixed power
electrical preheaters inlet and microchannel inlet, due to the higher temperature
of the surrounding ambient.
consists in the heat given by the fixed power electrical preheaters to
the refrigerant, calculated dividing the squared voltage of the supply by the
electrical resistance of the preheater;
is the refrigerant side capacity. is used instead of ,
since the pressure and temperature readings are more accurate in case of liquid
(2.7)
(2.8)
(2.9)
(2.10)
34
refrigerant with high degree of subcooling (5 °C) than liquid with low degree of
subcooling (1-2 °C) or two-phase refrigerant.
The error is always found to be negligible, which means the heat balance is
respected for all the experimental series performed.
During oil injection tests, the microchannel heat exchanger capacity is calculated
through the air side capacity, in order to be independent from the refrigerant-oil mixture
properties, which affect enthalpy calculation. It is possible to use the air side capacity
since the latent load is always within the 10% of the total load (sensible and latent). All
the tests are done in the driest conditions, reached running the ethylene glycol and water
loop at the lowest possible temperature and dehumidifying the air through silica gel.
Moreover, in order to ensure dry conditions on the heat exchanger surface, the air
volumetric flow rate is always maintained at its maximum value of 3300 m3/h during all
the experimental series. In this way, it is possible to obtain the inlet air dew point
temperature lower than the refrigerant saturation temperature and minimize the latent
load.
The heat transfer factor HTF is calculated based on the average heat transfer capacity
measured during tests with oil and the heat transfer capacity at the corresponding
operating conditions without oil, as in equation (2.11):
The pressure drop factor PDF accounts for the increase in pressure drop across the
microchannel heat exchanger due to the additional presence of lubricant. It is defined as
the ratio between the pressure drop during tests with oil and the corresponding operation
conditions without oil, as in (2.12):
The two tests, with and without lubricant, have on one hand the same refrigerant side
inlet pressure, temperature and total mass flow rate and on the other hand the same air
side inlet dry bulb temperature and volumetric flow rate.
While oil is injected, the pressure drop and the heat exchanged are calculated every time
step of two seconds. The HTF and PDF are then calculated at every time step and
averaged over the entire period after steady state conditions are reached again.
During all experimental series, the oil injection causes an increase in the total mass flow
rate circulating in the system, with respect to the case of pure refrigerant. For this reason,
baseline tests are run. The baseline test is performed with pure refrigerant at the same
(2.11)
(2.12)
35
pressure and temperature inlet conditions than the oil test, but with different mass flow
rates. The purpose is interpolating or extrapolating capacity and pressure drop values as
functions of refrigerant mass flow rate. Then, the heat transfer capacity and pressure
drop during the injection period are compared with the experimental data without
lubricant having the same equivalent total mass flow rate. Using baseline tests, it is
possible to address the increase in pressure drop and the loss in capacity only to the
lubricant replacing refrigerant.
2.6.2 Oil retention volume analysis
The oil retention volume is calculated dividing the oil retention mass
in the microchannel evaporator by the oil density , as shown in
equation (2.13):
The pure POE lubricant density is calculated from lubricant manufacturer data (when
available) or as a function of temperature using the equation (2.14) provided by Cavestri
[43]:
Another important parameter is the oil retention volume normalized , which
consists in the oil retention volume ORV divided by the internal volume of the
microchannel heat exchanger , as in (2.15):
Two different sightglasses, shown in Fig. 2.10, are placed in the suction line after the
heat exchanger outlet, in order to observe and note the time when the oil-refrigerant
mixture flow is fully developed.
(2.13)
(2.14)
(2.15)
36
Fig. 2.10: Particular of the two sightglasses placed at the evaporator outlet
Two important parameters in oil retention calculation are:
The time interval between the injection start and the observed time when the oil-
refrigerant mixture flow is fully developed at the sightglass. Since two
sightglasses are installed, is the time interval referred to sightglass #1 and
is the time interval with respect to sightglass #2. Using recorded data from
Labview, it is possible to integrate the oil mass flow rate during the injection
time interval and obtain the total amount of lubricant mass injected, both for
inlet and outlet injection tests. The upper limit of the mentioned integral can be
or , as shown in (2.16):
, defined as the difference .
Despite same refrigerant side conditions and oil mass fraction, the time interval between
injection start and observed time at both the sightglasses, and , is largely
(2.16)
37
different in case of inlet or outlet injection tests, that means the oil-refrigerant mixture
flow is affected by the injecting point. For this reason the procedure to calculate the oil
trapped in the evaporator internal volume is defined as follows:
When inlet tests are less than three, three outlet injection tests with a value of oil
mass fraction equal to 0.5, 1.5 and 3% are performed; when inlet tests are
more than three, the same number of outlet tests are performed. In both cases,
the purpose is determining the trend as function of .
The of the inlet test, is plotted as function of ;
An equivalent outlet , having the same of the inlet injection test, is then
calculated from the outlet test trend. In this way, the inlet and outlet tests have
the same , and approximately have the same refrigerant and oil mixture
velocity. An example of trend, as function of , both for inlet and outlet
tests and the equivalent outlet , is represented in Fig. 2.11:
Fig. 2.11: trend as function of for inlet and outlet test
The total amount of lubricant mass injected during the inlet and outlet tests is
directly calculated from Labview data, while the total amount of lubricant mass
injected during the equivalent outlet test, is estimated interpolating the data from
the outlet test trend. An example of total amount of lubricant mass injected
during the outlet test is presented in Fig. 2.12:
0
100
200
300
400
0 1 2 3 4 5
Δt
[s]
OMF [%]
Outlet tests
Inlet tests
Equivalent outlet OMF with the same Δt of the inlet test with OMF=1%
38
Fig. 2.12: Total amount of lubricant mass injected during outlet test
The resultant oil retention mass ORM in the microchannel heat exchanger
internal volume is then calculated as difference between the amount of oil of the
inlet test and the equivalent outlet test.
As indicated in Fig. 2.12, the total amount of lubricant mass injected can be calculated
with respect to sightglass #1 or sightglass #2. The procedure works properly in both
cases, since the oil retained values obtained with respect to the two different
sightglasses are always very similar.
2.7 Uncertainty analysis
According to the theory, every result obtained by measurements has to be connected to
an estimate of its associated error. Error analysis is the study and evaluation of the
measurement uncertainties: the three main reasons why it is so important are to allow
the scientists to estimate the size of uncertainties, to know where they are mainly
located and possibly to understand how to reduce them when necessary. Errors in any
experimental measurement are inevitable, but can be reduced to a minimum value
following the correct procedures to get the readings and taking particular care of each
instrument used. A good starting point is checking periodically the calibration of the
equipment, especially the air side sensors, more sensitive than the refrigerant side ones.
The calibration consists in removing the RTDs or thermocouples from their operating
positions and dipping them in a temperature bath filled with ethylene glycol and
0
30
60
90
120
150
0 1 2 3 4
Ma
ss i
nje
cte
d [
g]
OMF [%]
Mass injected @ sightglass #2
Mass injected @ sightglass #1
Total amount of lubricant mass
injected for the equivalent outlet test, with respect to sightglass #2
39
insulated from the surroundings. The temperature bath allows to set up a predetermined
range of temperatures. Comparing the RTDs or thermocouples readings from Labview
data acquisition system with the value shown by a high precision thermometer dip in the
same ethylene glycol and repeating the procedure for at least six different temperature
points, it is possible to determine the proper calibration line. Then, to check the
accuracy of the calibration, few verification tests were performed applying the same
procedure described above and the difference between the Labview readings and the
thermometer measurements was always found to be less than 0.05 °C.
The experimental apparatus employed in this work contains multiple sensors to measure
temperatures, absolute and differential pressures, mass flow rates, volumetric flow rates
and other properties of refrigerant, oil and air sides. A brief description of the
instrumentation used and the uncertainty of the output parameters is given in Table 2.2
and Table 2.3.
Table 2.2: Air side sensors specifications
Sensor Use Manufacturer Model Nominal
range
Accuracy
Resistance
Temperature
Detector
Air side dry and
wet bulb
measurements
Omega
Engineering Inc.
Pt 100 -40 °C to
+54 °C 0.1 °C
Relative
Humidity
Sensor
Air side relative
humidity
measurements
Omega
Engineering Inc.
HX71-MA 0% to
100% 3.5% from
RH = 15%
to RH =
85%
Air Flow
Nozzles
Volumetric air
flow
measurements
Helander Metal
Spinning
Company
Aluminium
elliptical
nozzle
252 to
3600
m3/h
0.04% of
volumetric
flow rate
40
Table 2.3: Refrigerant and oil side sensors specifications
Sensor Use Manufacturer Model Nominal
range
Accuracy
Inline
thermocouple
Refrigerant side
temperature
measurements
Omega
Engineering
Inc.
T-type
(copper and
constantan)
-40 °C to
+54 °C 0.3 °C
Absolute
pressure
transducer
Refrigerant side
pressure
measurements
Setra System
Inc.
C206 50 to 3450
kPa 4.5 kPa
High precision
pressure
transducer
Pressure
measurement at
microchannel
evaporator
outlet
Omega
Engineering
Inc.
DPGM049 0 to 1200
kPa 1 kPa
Differential
pressure
transducer
Differential
pressure
measurement
across
microchannel
evaporator
Validyne
Engineering
Diagram
typer P55D
0 to 50 kPa
(4 different
transducers)
0.25%
of the full
scale
Refrigerant
Mass Flow
Meter
Refrigerant
mass flow rate
measurements
Micro Motion
Inc.
Coriolis
mass flow
meter
2700C12
0 to 2180
kg/h 0.1% of
the
reading
Oil Mass
Flow Meter
Oil mass flow
rate
measurements
Micro Motion
Inc.
Coriolis
mass flow
meter
2700C12
0 to 108 kg/h 0.1% of
the
reading
Weighing
scale
Oil sample
weight
measurements
Arlyn Scales SAW-L 0 to 22 kg 2.2 g
In this study, many main parameters such as OMF, HTF, PDF and ORVN are not
measured directly, but are determined from other quantities through a functional
relation , as follow:
Where symbolizes the output estimate, result of the measurement;
represent the input estimates and the number of the input quantities. Hence the
uncertainty on the final result is estimated assuming reasonable uncertainty values of
the input parameters given by the manufacturers. The uncertainty analysis was
(2.17)
41
conducted according to the uncertainty propagation method suggested by Taylor and
Kuyatt [44], described in equation (2.18):
Where is the uncertainty of the measurement result and is the uncertainty
associated to each input parameters.
The averaged uncertainties of the most important calculated parameters are shown in
Table 2.4:
Table 2.4: Calculated uncertainties of the main paramenters
Parameter Uncertainty
OMF 0.11
HTF 0.06
PDF 0.036
ORVN 0.004
(2.18)
3 Experimental results
The matrix, which summarizes all the experimental tests, is provided in Table 3.1:
Table 3.1: Experimental test matrix
Refrigerant Evaporator Saturation
Temperature
Degree of
Superheating
Oil Mass
Fraction
Number of Oil
Tests
R410A A
3.9 °C 2.8 °C
8.3 °C
13.9 °C
0%
1%
3% 38 (16 inlet and
22 outlet tests)
R410A B
3.9 °C 2.8 °C
8.3 °C
13.9 °C
0%
1%
3%
DR5A A
3.9 °C
8.9 °C
2.8°C
8.3 °C
13.9 °C
8.3 °C
0%
0.5%
1%
3%
5% 39 (20 inlet and
19 outlet tests)
DR5A B
3.9 °C
8.3 °C
0%
1%
3%
R32 A
3.9 °C
8.3 °C
0%
0.5%
1%
3%
5%
9 (5 inlet and 4
outlet tests)
R1234yf A
3.9 °C
8.3 °C
0%
0.5%
1%
3%
5%
10 (6 inlet and 4
outlet tests)
The first objective of the experimental work is providing a comparison between a well-
known refrigerant used in air conditioning applications, R410A, and one of its possible
low Global Warming Potential replacement, DR5A. Most of the experimental series are
carried out at the same air dry bulb temperature and air volumetric flow rate, in order to
analyze the behavior of the different refrigerants, in terms of HTF, PDF and ORVN,
under the same air side conditions. On the other hand, keeping the same refrigerant
saturation temperature and degree of superheating and the same air dry bulb
temperature leads to fixed temperature difference between the air and the refrigerant
44
flowing inside the microchannel tubes, namely the heat transfer driving force.
Furthermore, the effects of different saturation temperatures, degrees of superheating,
mass flow rates geometries are treated. Few experimental series are carried out at the
same refrigerant mass flow rate, in order to better understand the mass flux effect on oil
retention. The oil mass fraction is varied from 0% to 5%, identified as the typical range
of oil circulating ratio in refrigerating systems. The data are divided in different sections,
to underline the aim of the comparison. The refrigerant DR5A presents a significant
temperature glide in the two phase dome. Hence, the saturation temperature in the
experimental test matrix refers to the saturated liquid temperature at the evaporator inlet,
while the superheating refers to the difference between the temperature of vapor at the
microchannel heat exchanger outlet and the temperature of saturated vapor at the
evaporator outlet pressure.
The second objective of the experimental work is to provide a comparison between R32,
R1234yf and DR5A, three low GWP refrigerants currently under investigation in the
air-conditioning and refrigeration industry. Since DR5A is composed by 68.9 wt.% R32
and 31.1 wt.% R1234yf, it is possible to analyze both the pure fluid and mixture
behavior. The experimental series are carried out at the same refrigerant side conditions,
while the air side parameters are varying. Keeping constant the saturation temperature,
degree of superheating and mass flow rate of the three refrigerants leads to understand
the different lubricant addition effect in terms of HTF, PDF and ORVN.
3.1 DR5A and R410A comparison
In the following section a comparison between R410A and its possible low GWP
replacement DR5A is presented.
3.1.1 Comparison under same air dry bulb temperature conditions
In this section DR5A and R410A comparison tests with same air dry bulb temperature,
air volumetric flow rate and constant degree of superheating are presented. The
experimental comparison is conducted with Evaporator A geometry at refrigerant
saturation temperature equal to 3.9 °C. Some tests are performed twice to obtain a
reliable result in terms of repeatability. In order to maintain the constant parameters
described above, the refrigerant mass flow rate and the refrigerant mass flux are varying.
In addition, the effect of two different saturation temperatures, 3.9 °C and 8.9 °C
respectively, is shown for DR5A experiments. In order to obtain the higher saturation
temperature condition, the air dry bulb temperature is increased, while the refrigerant
mass flow rate is maintained within a tolerance of 5 kg/hr.
45
The data collected in Fig. 3.1, Fig. 3.2 and Fig. 3.3 are classified as follows:
The blue solid triangles represent DR5A tests at saturation temperature equal to
3.9 °C, superheating of 8.3 °C and refrigerant mass flow rate of 95 kg/hr. For
this series the air dry bulb temperature is 15.8 °C and the air volumetric flow
rate is equal to 3300 m3/h;
The green hollow triangles symbolize DR5A tests at saturation temperature of
8.9 °C, superheating of 8.3 °C and refrigerant mass flow rate of 100 kg/hr. The
air dry bulb temperature is increased to 21.1 °C to reach the higher saturation
temperature condition, while the air volumetric flow rate is equal to 3300 m3/h;
The red solid squares refer to R410A experiments at saturation temperature of
3.9 °C, superheating of 8.3 °C and refrigerant mass flow rate of 160 kg/hr. The
air dry bulb temperature is 15.8 °C and the air volumetric flow rate is equal to
3300 m3/hr.
Fig. 3.1: Experimental Heat Transfer Factor under different refrigerant side saturation
temperature conditions
The Heat Transfer Factor in Fig. 3.1 shows a proportional decrease with the increase in
oil mass fraction. This trend is explained by the physical phenomenon induced by oil
presence: the oil film inside the microchannel heat exchanger tubes determines an
additional mass and heat transfer resistance, lowering the heat transferred from the air to
0.92
0.94
0.96
0.98
1
1.02
0 1 2 3 4 5
HT
F [
-]
OMF [%]
DR5A
DR5A high T sat
R410A
46
the refrigerant and oil mixture. The heat transfer factor is slightly higher for DR5A,
which means better performances in terms of coil capacity compared to R410A. The
HTF seems not affected by the saturation temperature increase.
Fig. 3.2: Experimental Pressure Drop Factor under different refrigerant side saturation
temperature conditions
Fig. 3.2 depicts the Pressure Drop Factor as function of the oil mass fraction circulating
in the microchannel evaporator. The oil addition causes an increase in the overall
pressure drop across the heat exchanger, with respect to the pure refrigerant case. On
one hand, the lubricant creates a film inside the tubes, which reduces the available cross
sectional area and determines an increase of refrigerant liquid and vapor velocities and
then a pressure drop raise. On the other hand, the oil mixed with refrigerant causes a
liquid phase viscosity increase and determines higher pressure drop measured across the
evaporator. The DR5A experiments show a slightly lower penalty factor, even if the
order of magnitude is the same, which suggests a similar behavior. A possible reason
for R410A higher PDF can be the different and higher mass flux conditions compared
to DR5A series, since higher mass flow rates are associated to higher pressure drops.
Among DR5A experiments, conducted with similar mass flow rates, the high saturation
temperature test presents lower PDF: a saturation temperature raise, which determines
lower liquid mixture density, results in a smaller Pressure Drop Factor.
0.8
1
1.2
1.4
1.6
1.8
0 1 2 3 4 5
PD
F [
-]
OMF [%]
DR5A
DR5A high T sat
R410A
47
Fig. 3.3: Experimental Oil Retention Volume Normalized under different refrigerant side
saturation temperature conditions
The Oil Retention Volume Normalized is the index which represents the amount of
lubricant retained in the microchannel heat exchanger tubes and headers. The oil
retention phenomenon affects mostly the inlet and outlet horizontal headers, where the
available cross sectional area is higher than the microchannel tubes one. In the inlet
header the oil retention is limited by the large quantity of liquid refrigerant, able to
carryover the lubricant. The oil retention is a very detrimental phenomenon in the outlet
header, where the high lubricant concentration liquid phase, due to its higher density
and viscosity with respect to the refrigerant vapor phase, settles on the surfaces,
increasing the internal volume occupied by oil. Fig. 3.3 points out how the amount of
lubricant trapped in the heat exchanger increases proportionally to the oil mass fraction
circulating in the system. The ORVN is higher for DR5A tests than R410A tests. The
main reason is the different mass flux conditions: indeed R410A has a larger mass flux,
that means an increase in the vapor shear stress and a smaller amount of oil retained in
the microchannel heat exchanger.
3.1.2 Comparison under same mass flux conditions
This section contains a comparison series between DR5A and R410A refrigerants,
which is conducted to understand the HTF, PDF and ORVN variations under the same
mass flux conditions. The mass flux represents the ratio between the mass flow rate
0
0.03
0.06
0.09
0.12
0.15
0.18
0 1 2 3 4 5
OR
VN
[-]
OMF [%]
DR5A
DR5A high T sat
R410A
48
flowing across the heat exchanger and the total flow area available. The total flow area
of both evaporators, listed in Table 2.1, is calculated as the cross sectional area of each
tube multiplied by the number of microchannel tubes.
Both DR5A and R410A tests are performed with Evaporator A geometry, at saturation
temperature equal to 3.9 °C, superheating of 8.3 °C, refrigerant mass flow rate of 95
kg/hr and air volumetric flow rate equal to 3300 m3/hr
In this case, the air dry bulb temperature is not constant between the experimental series,
but it is varied to obtain the same mass flux and refrigerant mass flow rate conditions.
The air dry bulb temperature is equal to 15.8 °C in DR5A experiments, while is 11.7 °C
in R410A ones. The main purpose of this comparison is maintaining the same
conditions on the refrigerant side, such as degree of superheating, mass flux and mass
flow rate, same geometry, same saturation temperature, in order to address any variation
to the different refrigerant used.
In Fig. 3.4, Fig. 3.5 and Fig. 3.6, solid blue triangles represent the first DR5A series,
while the solid red squares stand for the R410A experiments.
Fig. 3.4: Experimental Heat Transfer Factor under same mass flux conditions
The HTF has its usual trend and decreases with the increase in the circulating oil mass
fraction. At OMF higher than 1%, DR5A shows a different behavior compared to
R410A, since the Heat Transfer Factor decreases slightly. One possible reason can be
0.84
0.88
0.92
0.96
1
1.04
0 1 2 3 4 5
HT
F [
-]
OMF [%]
DR5A
R410A
49
the different value of phase-change enthalpy, which is higher for DR5A. The refrigerant
pressure and temperature conditions are the same among the two series, but the larger
latent specific heat of DR5A increases the length occupied by evaporating two-phase
flow inside the microchannel heat exchanger. Replacing single-phase refrigerant with
two-phase refrigerant results, in turn, in a heat transfer enhancement, which makes
DR5A less sensitive to the capacity penalization due to oil addition. Comparing DR5A
and R410A series, an increase of the evaporating length determines a larger portion of
the tubes being occupied by fluid at low vapor qualities. Especially at low vapor
qualities, the oil addition increases the wetted surface and improves the heat transfer,
thanks to to liquid mixture higher viscosity, as previously demonstrated in [16].
Fig. 3.5: Experimental Pressure Drop Factor under same mass flux conditions
Under the same mass flux conditions, the Pressure Drop Factor presents both for DR5A
and R410A a very similar trend, mostly at smaller OMFs. At larger oil mass fractions
the difference in PDF among the two series becomes higher, which means the higher
amount of oil circulating has a larger penalization in DR5A series. Even in this case, the
increase of evaporating length, which determines a larger amount of liquid mixture
carried around into the microchannel evaporator is identified as the most effective
parameter. Indeed the liquid phase has higher viscosity than vapor one at same
temperature conditions and causes an increase in pressure drop.
0.8
1
1.2
1.4
1.6
1.8
0 1 2 3 4 5
PD
F [
-]
OMF [%]
DR5A
R410A
50
Fig. 3.6: Experimental Oil Retention Volume Normalized under same mass flux conditions
The Oil Retention Volume Normalized for both DR5A and R410A experiments shows a
very similar behavior: it is strongly dependent on the oil mass fraction circulating, even
if the effects due to the different refrigerant used seem to be negligible. The same mass
flux conditions leads to underline an important aspect involving the amount of oil
retained inside the microchannel evaporator A. Due to the geometry, a sort of filling
process is encountered in the horizontal outlet header: the refrigerant-oil boiling mixture
enters the outlet header flowing upward from the microchannels and the small volumes
created by the tubes entering the header tend to be filled by the oil rich mixture because
of the gravity force. This aspect is very important especially in the low mass flux
experiments, when the vapor has lower velocities and it is not able to carry over
lubricant droplets.
Moreover, to better understand the mass flux effect on oil retention, an additional DR5A
series is performed using Evaporator A, at saturation temperature of 3.9 °C, degree of
superheating equal to 8.3 °C and higher refrigerant mass flow rate. The comparison
results between the DR5A higher and lower mass flux series, shown in Fig. 3.7, are
classified as follows:
The blue solid triangles represent DR5A at saturation temperature equal to
3.9 °C and superheating of 8.3 °C. For this series the refrigerant mass flow rate
is 95 kg/hr
0
0.03
0.06
0.09
0.12
0.15
0 1 2 3 4 5
OR
VN
[-]
OMF [%]
DR5A
R410A
51
The hollow red squares stand for experimental results for saturation temperature
of 3.9 °C, superheating of 8.3 °C and refrigerant mass flow rate equal to 135
kg/hr.
Fig. 3.7: Experimental Oil Retention Volume Normalized with two different mass flux
conditions
The Fig. 3.7 shows that the ORVN slightly reduces while increasing the refrigerant mass
flow rate. This behavior is also found in [27]. Since higher mass flow rates and mass
fluxes determine higher vapor core velocities, the inertia force becomes predominant on
the viscous one, resulting in a smaller amount of oil trapped in the microchannel heat
exchanger.
3.1.3 Effect of different degree of superheating
In this section a comparison between the refrigerants DR5A and R410A is performed
with the aim of carrying out the effects on HTF, PDF and ORVN of different degrees of
superheating, respectively 2.8 °C, 8.3 °C and 13.9 °C. Each DR5A series is compared
with the R410A one with the same air dry bulb temperature, equal refrigerant degree of
superheating and same refrigerant saturation temperature. All the experiments are
performed with Evaporator A geometry and air volumetric flow rate equal to 3300
m3/hr. The refrigerant mass flow rate and then the mass flux are varied to obtain the
0
0.03
0.06
0.09
0.12
0.15
0 1 2 3 4 5
OR
VN
[-]
OMF [%]
DR5A lower mass flux
DR5A higher mass flux
52
conditions described above. Also for this experimental series some tests are performed
twice.
The collected data in Fig. 3.8, Fig. 3.9 and Fig. 3.10 are classified as follows:
Red solid triangles symbolize DR5A tests run with degree of superheating equal
to 2.8 °C and at saturation temperature of 3.9 °C. The refrigerant mass flow rate
is 105 kg/hr and the air side dry bulb temperature is equal to 13.8 °C;
Blue solid triangles refer to DR5A tests with 8.3 °C of superheating and at
saturation temperature equal to 3.9 °C. The refrigerant mass flow rate is 95 kg/hr
and the air side dry bulb temperature is equal to 15.8 °C;
Green solid triangles stand for DR5A experiments with degree of superheating
equal to 13.9 °C and at saturation temperature of 3.9 °C. The refrigerant mass
flow rate is 85 kg/hr and the air side dry bulb temperature is equal to 19.5 °C;
Red hollow squares symbolize R410A tests run with degree of superheating
equal to 2.8 °C and at saturation temperature of 3.9 °C. The refrigerant mass
flow rate is 160 kg/hr and the air side dry bulb temperature is equal to 13.8 °C;
Blue hollow squares refer to R410A tests with 8.3 °C of superheating and at
saturation temperature equal to 3.9 °C. The refrigerant mass flow rate is 160
kg/hr and the air side dry bulb temperature is equal to 15.8 °C;
Green hollow squares stand for R410A experiments with degree of superheating
equal to 13.9 °C and at saturation temperature of 3.9 °C. The refrigerant mass
flow rate is 160 kg/hr and the air side dry bulb temperature is equal to 19.5 °C.
53
Fig. 3.8: Experimental Heat Transfer Factor with different degree of superheating
The Heat Transfer Factor, for all the degrees of superheating analyzed, is inversely
proportional to the oil mass fraction, that means a penalization in the heat transfer when
the amount of oil injected is increasing. The thermal phenomenon associated to the
decrease in superheating is a reduction in the refrigerant outlet enthalpy and then in the
coil capacity, for fixed refrigerant inlet pressure, temperature and mass flow rate
conditions. Indeed, the different degree of superheating is obtained varying the heat flux
provided by the air; in particular, the air dry bulb temperature is changed while the
volumetric flow rate is maintained equal to a constant value. The HTF plot shows how a
low degree of superheating, regardless of the refrigerant circulating, is more sensitive in
terms of heat transfer penalization than a high one.
0.92
0.94
0.96
0.98
1
1.02
0 1 2 3 4 5
HT
F [
-]
OMF [%]
DR5A SH 2.8°C
DR5A SH 8.3°C
DR5A SH 13.9°C
R410A SH 2.8°C
R410A SH 8.3°C
R410A SH 13.9°C
54
Fig. 3.9: Experimental Pressure Drop Factor with different degree of superheating
The Pressure Drop Factor increases with the increase in the oil mass fraction circulating
into the system. The plot above does not prove an evident reduction in case of different
superheating values, since the data present a good uniformity, especially for OMF lower
than 3% .
0.8
1
1.2
1.4
1.6
1.8
0 1 2 3 4 5
PD
F [
-]
OMF [%]
DR5A SH 2.8°C
DR5A SH 8.3°C
DR5A SH 13.9°C
R410A SH 2.8°C
R410A SH 8.3°C
R410A SH 13.9°C
55
Fig. 3.10: Experimental Oil Retention Volume Normalized with different degree of
superheating
The ORVN displays an important dependence on the different degree of superheating.
Low values of superheating determine a higher amount of refrigerant dissolved in the
oil rich mixture, caused by the increase of R410A solubility in POE oil at lower
temperatures, which leads to lower viscosity of the refrigerant-oil mixture and lower
amount of oil retained. Even if DR5A and POE lubricant solubility properties are not
available, the DR5A behavior seems close to the R410A one. In [45] a similar
substantial reduction in the amount of oil retained is observed after decreasing the
degree of superheating.
3.1.4 Effect of different geometry
This section contains a comparison between two experimental series carried out with a
different heat exchanger geometry, called Evaporator B, in order to generalize the
results and trends exposed in the previous sections in terms of Heat Transfer Factor,
Pressure Drop Factor and Oil Retention Volume normalized, and then obtain an overall
validation.
The DR5A and R410A series are performed at the same saturation temperature of
3.9 °C, equal superheating degree of 8.3 °C, constant air dry bulb temperature of
13.5 °C and air volumetric flow rate of 3300 m3/hr. In order to achieve the fixed
0
0.03
0.06
0.09
0.12
0.15
0.18
0 1 2 3 4 5
OR
VN
[-
]
OMF [%]
DR5A SH 2.8°C
DR5A SH 8.3°C
DR5A SH 13.9°C
R410A SH 2.8°C
R410A SH 8.3°C
R410A SH 13.9°C
56
conditions described above, the mass flow rates are varying between the two different
series. DR5A and R410A experiments are conducted with mass flow rate equal to 75
kg/hr and 160 kg/hr, respectively.
The solid blue triangles in Fig. 3.11, Fig. 3.12 and Fig. 3.13 represent DR5A
experimental data, while the solid red squares symbolized the R410A ones.
Fig. 3.11: Experimental Heat Transfer Factor for Evaporator B
The Heat Transfer Factor shows a decreasing trend with the increase in the oil
circulating into the system. The observed HTF presents a heat transfer penalization due
to oil addition, almost of the same order of magnitude both for DR5A and R410A. It is
clear that lubricant injection creates an additional heat transfer resistance, with respect
to the conditions without oil. In order to reach the required conditions on the refrigerant
and air sides, DR5A tests are run with very low mass fluxes. Low refrigerant mass flow
rates determine, in turn, higher oil retention in the microchannel tubes, since the vapor
velocities are not enough to remove effectively the lubricant settled in the outlet header.
The higher value of Oil Retention Volume normalized in DR5A test, showed in Fig.
3.13, is also affecting the Heat Transfer Factor, since the thicker oil film deteriorates the
capacity transferred from air to refrigerant.
0.94
0.96
0.98
1
1.02
0 0.5 1 1.5 2 2.5 3
HT
F [
-]
OMF [%]
R410A
DR5A
57
Fig. 3.12: Experimental Pressure Drop Factor for Evaporator B
The Pressure Drop Factor, even for the different geometry analyzed, displays the same
behavior: PDF raises with the increase in OMF. Indeed, the oil film inside the
microchannels determines a reduction in the cross sectional area, which, in turn, causes
an increase in refrigerant and oil velocities. Moreover, the refrigerant and lubricant
viscosity is higher than the pure refrigerant one. These are the two main parameters
responsible for higher pressure drop across the heat exchanger. While the lubricant
circulating into the system is increasing, this two phenomena become more significant.
0.8
1
1.2
1.4
1.6
0 0.5 1 1.5 2 2.5 3 3.5
PD
F [
-]
OMF [%]
R410A
DR5A
58
Fig. 3.13: Experimental Oil Retention Volume Normalized for Evaporator B
The Oil Retention Volume normalized raises while OMF is increasing. Two important
differences between Evaporator B and Evaporator A data are observed in Fig. 3.13: the
ORVN seems to have both lower absolute values and more linear variations with OMF.
Possible reasons are found only in the different geometry, since the test procedures are
the same during all the experimental series. In particular, Evaporator B consists in a
single slab and double pass geometry, with headers at the bottom, as shown in Fig. 2.3.
Thanks to outlet header position, no filling process is encountered and less lubricant
accumulation is observed. In this case, the gravity is a favorable effect, since it helps the
vapor phase refrigerant flowing downward in carrying over the oil and in determining
lower ORVN absolute values. Last, but not least, since the Oil Retention Volume
normalized is defined as the ratio between the oil retention mass and the evaporator
volume, another reason for Evaporator B lower ORVN values can be the larger
Evaporator B volume measured, , against Evaporator A volume, equal to
.
3.2 R32, R1234yf and DR5A comparison
In the following section a comparison between R32, R1234yf and DR5A refrigerants is
presented. All the experimental series are conducted with the same Evaporator A
geometry at the refrigerant saturation temperature of 3.9 °C, degree of superheating
0
0.03
0.06
0.09
0.12
0 0.5 1 1.5 2 2.5 3 3.5
OR
VN
[-]
OMF [%]
R410A
DR5A
59
equal to 8.3 °C and mass flow rate of 95 kg/hr. The air side volumetric flow rate is kept
constant at 3300 m3/hr, while the dry bulb temperature is varying to achieve the desired
conditions on the refrigerant side. In order to analyze the lubricant addition effect on
HTF, PDF and ORVN, a wide range of circulating oil mass fractions, equal to 0, 0.5, 1,
3 and 5% is considered.
The collected data in Fig. 3.14, Fig. 3.15 and Fig. 3.16 are classified as follows:
Red solid squares symbolize R1234yf tests conducted with degree of
superheating equal to 8.3 °C at refrigerant saturation temperature of 3.9 °C. The
air dry bulb temperature is equal to 13 °C;
Blue solid diamonds refer to DR5A tests with degree of superheating of 8.3 °C
at refrigerant saturation temperature equal to 3.9 °C. The air dry bulb
temperature is equal to 15.8 °C;
Green solid triangles stand for R32 experiments run with degree of superheating
equal to 8.3 °C at refrigerant saturation temperature of 3.9 °C. The air dry bulb
temperature is equal to 13.9 °C.
Fig. 3.14: R1234yf, DR5A and R32 experimental Heat Transfer Factor under same refrigerant
side conditions
0.92
0.94
0.96
0.98
1
1.02
0 1 2 3 4 5
HT
F [
-]
OMF [%]
R1234yf
DR5A
R32
60
The Heat Transfer Factor shows for all the refrigerants a reduction with the increase in
oil mass fraction. Although DR5A is mainly composed by R32 (68.9 wt. %), its HTF
seems closer to the one of R1234yf. A possible reason is observed analyzing the
different behavior of the three refrigerants after lubricant addition. The oil presence
always causes a reduction in the temperature at the microchannel heat exchanger outlet,
with respect to the pure refrigerant case. The temperature decrease is small for R1234yf
and DR5A experiments (0.5 °C), while for R32 experiments it is more significant,
particularly at low oil mass fractions (2.5 °C). As consequence, the R32 capacity
considerably decreases after oil addition and determines a consistent reduction in the
Heat Transfer Factor.
Fig. 3.15: R1234yf, DR5A and R32 experimental Pressure Drop Factor under same refrigerant
side conditions
The Pressure Drop Factor presents for all the three low GWP refrigerants analyzed the
same trend and similar numerical values. The oil addition always causes an increase in
the pressure drop measured across the microchannel evaporator, proportional to the
increase in the oil mass fraction circulating into the system.
0.8
1
1.2
1.4
1.6
1.8
0 1 2 3 4 5
PD
F [
-]
OMF [%]
R1234yf
DR5A
R32
61
Fig. 3.16: R1234yf, DR5A and R32 experimental Oil Retention Volume normalized under same
refrigerant side conditions
The Oil Retention Volume normalized shows a similar trend for all the refrigerants
tested, that is an increase with the increase in oil mass fraction. In Fig. 3.16 it is possible
to observe the filling process encountered in the low mass flux experiments conducted
with evaporator A. The small volumes created by the microchannel tubes protrusion
inside the outlet header, when filled by the oil rich liquid mixture determine an
exponential increase in the ORVN with respect to the no oil condition (OMF 0%),
especially at low oil mass fraction condition (OMF 0.5%). R32 refrigerant presents
lower Oil Retention Volume normalized compared to DR5A and R1234yf. The R32
temperature reduction at the microchannel outlet due to lubricant addition determines
higher solubility in POE oil, that leads to smaller viscosity of the liquid mixture and
lower amount of oil retained.
0
0.03
0.06
0.09
0.12
0.15
0.18
0 1 2 3 4 5
OR
VN
[-]
OMF [%]
R1234yf
DR5A
R32
4 Simulation code description
4.1 Previous work on the simulation code
The simulation model was initially developed from Ipseng Iu [46], who provided an air-
to-air heat pump program with an advanced heat exchanger algorithm. The aim of this
study was to fill the lack of empirical data regarding the ozone-safe HFC refrigerants,
such as R410A and R134a. Three new aspects, neglected by the previous models, were
considered. The first one consisted in modeling the entire heat exchanger circuitry, an
important tool universally ignored in simple models, but with a significant impact on
coil capacity and pressure drop. Indeed, the connections between the coil tubes were
demonstrated an important parameter for the system optimization. The second aspect
was the use of refrigerant mixture properties instead of the average saturation
temperature of the mixture components in the heat transfer calculation. The last
innovative aspect was considering the local air side heat transfer coefficient, in order to
account for the heat transfer coefficient variations from row to row. The final result was
the capability to simulate on one hand both pure refrigerant and refrigerant mixtures and
on the other hand complicated heat exchanger circuiting.
Each heat pump component is developed as an independent simulation program, so it is
possible to focus on and improve only one part at a time. The single component models
are integrated into a global program to simulate the overall system operation. In the
present thesis only the part of the code which allows to solve just the evaporator, is
used.
4.2 Microchannel evaporator simulation code
The first simulation work which presents a heat exchanger segment by segment
approach is carried out by Jiang [30]. This approach is able to account for the two-
dimensional non uniformity of air distribution across the heat exchanger and the change
of properties and heat transfer coefficients along the refrigerant flow direction. The
microchannel evaporator simulation model, developed at Oklahoma State University by
[47] and [48], is based on the segment-by-segment approach and uses the ε-NTU
method to solve a combined heat and mass transfer problem in each segment. The air-
to-refrigerant heat transfer and the refrigerant pressure drop are calculated for each
individual segment and the predicted conditions at the outlet are passed as input for the
adjacent segment until the refrigerant circuitry is completely solved. Each segment is
provided with the inlet refrigerant enthalpy, pressure and mass flow rate and the inlet air
64
dry bulb temperature, relative humidity and mass flow rate. The simulation program
handles various refrigerant and oil mixtures and different heat exchanger geometries.
The refrigerant side heat transfer flow boiling coefficient is calculated from Bertsch et
al.’s [49] correlation, developed for a wide range of mass fluxes and hydraulic
diameters, but for pure refrigerant. When oil is circulating into the system, the heat
transfer coefficient is calculated using the same correlation but adopting the mixture
properties instead of the pure refrigerant properties for describing the liquid phase of the
mixture.
In order to solve the entire heat exchanger, the microchannel evaporator is reduced to a
set of vertical columns connected in parallel. Each column represents a single tube of
the microchannel evaporator and is provided with a complete multi-folded louvered fin,
for calculating the heat transfer surfaces, the local air side heat transfer coefficient and
frictional pressure drop. The fins are split into two halves, one half on each side of the
tube, as shown in Fig. 4.1.
Fig. 4.1: Schematic representation of the segment by segment approach used in the simulation
program
The total heat transfer capacity of the microchannel heat exchanger is then calculated as
the sum of the individual column heat transfer capacities. The overall pressure drop is
the pressure drop of one microchannel tube, plus the pressure drop in the inlet and outlet
headers. The assumption made in the calculation of the overall capacity and pressure
drop are reasonable, since the uniform conditions both on the refrigerant and air side are
satisfied along all the tubes. The refrigerant mass flow rate uniformity is guaranteed by
controlling the inlet conditions to be near to saturated liquid during all the experimental
tests. Furthermore the manufacturer ensures that the inlet header geometry provides a
uniform filling among all the parallel microchannels with inlet qualities below 0.1. The
air side requirements are reached obtaining the temperature and relative humidity
uniformity and the same velocity profile on all the evaporator surface area.
65
The segment-by-segment modeling method is used to divide each column into small
elements along the refrigerant flow direction. The size of the elements is determined
considering the properties of air and refrigerant constant within the segment with a
negligible error. All the simulations presented in this study are run with 100 segments.
Each segment is solved iteratively until convergence and its output refrigerant
parameters are passed as input for the succeeding one, until the last element is reached.
The first segment conditions are equal to the evaporator inlet conditions and come from
the experimental data. The inlet and outlet headers are modeled as adiabatic segments
located at the beginning and at the end of the microchannel tubes. For each element, the
heat transfer capacity is calculated using the ε-NTU method, defined as:
Where:
symbolizes the heat transfer capacity of each segment or actual heat
transfer rate and can be determined from an energy balance on the refrigerant or
air side and can be expressed as:
is the coil effectiveness, defined as the ratio between the actual heat
transfer rate and the maximum possible heat transfer rate:
represents the smallest capacity rate between the refrigerant and
the air one:
, also defined as , is the maximum temperature
difference across the heat exchanger.
In order to solve the combined heat and mass transfer problem, the coil effectiveness is
calculated differently depending on the conditions on refrigerant and air sides, using
relations for cross-flow geometry heat exchangers from Kays and London [50]. For
single-phase refrigerant flow equations (4.5) and (4.6) are used; equation (4.5) if is
the refrigerant side capacity rate and equation (4.6) if is the air capacity rate
, respectively:
(4.1)
(4.2)
(4.3)
(4.4)
66
Both the equations account for refrigerant unmixed flow and air mixed flow. The
refrigerant flow is unmixed since it is forced through a particular direction and it is
prevented from moving in the transverse coordinate, while the air flow is mixed since
the fluid is free to move in the transverse direction, thanks to the louvered fins.
is the capacity ratio and is defined as:
is the Number of Transfer Units, a dimensionless parameter representative of
the heat transfer surface area of the coil (the larger , the larger is the heat
exchanger), expressed as:
For two-phase refrigerant flow, the capacity ratio is equal to 0, since the evaporating
fluid specific heat is infinite, so the ε-NTU equation becomes:
The overall heat transfer coefficient is calculated assuming negligible the
fouling resistances both on the inner and outer surfaces of the microchannel, as
expressed in (4.10):
The three terms in equation (4.10) represent the internal convective resistance on the
refrigerant side, the tube conductive resistance and the external convective resistance on
the air side.
(4.5)
(4.6)
(4.7)
(4.8)
(4.9)
(4.10)
67
The overall effectiveness for a finned surface, defined as the ratio between the total heat
transfer from the finned surface and the total heat transfer from the same surface if there
are no fins, is given by equation (4.11):
The fin effectiveness , calculated for a rectangular fin with length equal to half
of the distance between two consecutive microchannel tubes, is expressed as follows:
is a parameter depending on the air side heat transfer coefficient, the conductivity
of the material, the cross-sectional area and perimeter of the fin:
4.3 The air side correlations
As described in section 4.2, the simulation model requires three input parameters for the
air side calculation: the air mass flow rate, the inlet air dry bulb temperature and the
inlet air relative humidity. The air side heat transfer coefficient is calculated using a
correlation for louvered fin geometry in flat tubes from Chang and Wang [51]:
Where:
is the air mass flux calculated using the minimum flow area of the heat
exchanger column;
and are the specific heat and the Prandtl number of the air;
represents the j-factor, an empirical correlation derived from the experimental
data and defined as following:
(4.11)
(4.12)
(4.13)
(4.14)
(4.15)
68
The correlation (4.15), developed for louvered fin geometry, predicts the 90% of Chang
and Wang’s experimental database with an error smaller than 15%. The Reynolds
Number based on louver pitch of the current simulation work is always within the
validity range of the correlation, that is .
Another correlation from Chang and Wang [52] is used to calculate the Fanning
Frictional Factor and the air side pressure drop. The validity range of correlation
(4.16) is :
The parameters and are functions of the fin geometry and are summarized in
the equations from
(4.17) to (4.19):
Where:
is the louver pitch
is the louver angle
is the fin pitch
is the fin length
is the tube depth
is the louver length
is the tube pitch
(4.16)
(4.17)
(4.18)
(4.19)
69
is the fin thickness
is the hydraulic diameter of fin array
is the major diameter
For wet conditions, when the air crossing the heat exchanger is cooled below the dew
point, the approach used in the simulation code is similar to the one developed by
Harms et al. [53]. In the segments where condensation of the air moisture occurs, the
heat transfer capacity is estimated using a wet ε-NTU method, with enthalpy differences
instead of temperature differences:
The term represents the saturation air enthalpy with respect to the refrigerant
temperature. The equations (4.5), (4.6) and (4.9) are also valid for wet conditions,
provided that substituting with , defined as:
Finally, the overall heat transfer coefficient during wet condition is calculated through
(4.22):
Where symbolizes the specific heat of moist air and the
saturation specific heat. The parameter is defined as the derivative with respect to
temperature of the saturated air enthalpy evaluated at the refrigerant temperature.
4.4 The refrigerant side correlations
The simulation model requires as input four different parameters on the refrigerant side:
the inlet pressure, enthalpy, mass flow rate and the absolute oil mass fraction. The
refrigerant side heat transfer coefficient is calculated using the correlation from Bertsch
et al. [49], which presents a predictive method applicable over a wide range of
conditions. The innovation with respect to the previous studies it is considering nucleate
boiling and convective heat transfer terms affected by confinement of bubbles in small
and micro channels. The semi-empirical correlation, based on the formulation presented
by Chen [54], is independent from specific flow pattern and accounts for the effect of
small size channels. The experimental database used to validate the correlation contains
wetting and non-wetting fluids, hydraulic diameters from 0.16 to 2.92 (covering
micro and mini channels), mass fluxes from 20 to 3000 , heat fluxes from 0.4
(4.20)
(4.21)
(4.22)
70
to 115 , saturation temperatures between -194 to 97 , vapor qualities from 0
to 1, round and rectangular channels, single and multiple parallel channels, horizontal
and vertical orientations. The flow boiling heat transfer coefficient follows the
basic form of Chen correlation. The nucleate pool boiling term is calculated with
Cooper’s [55] correlation and the convective heat transfer coefficient is the
average of the single phase liquid and vapor ones ( and respectively),
with a linear dependence on the vapor quality .
The single-phase liquid and vapor convective heat transfer coefficients are estimated
with Hausen’s [56] correlation, which accounts for the laminar flow usually
encountered in microchannels, due to the low Reynolds Number and is given by (4.26):
The heat flux and mass flux dependences are addressed in nucleate boiling and
convective heat transfer expressions, while the effect of vapor quality and confinement
of bubbles dependence are counted in the suppression factor and in the enhancement
factor , defined as following:
On one hand, the phenomenon of suppression of bubbles, independent on channel
diameter, is mainly present at high vapor qualities, near to the dryout region; on the
other hand, the enhancement of convective heat transfer coefficient is mostly influenced
by the confinement of bubbles in small channels and has a weak dependence on vapor
quality. The enhancement factor reduces to for pure liquid and pure vapor phases,
while is greater than within the two-phase region.
All the parameters used in refrigerant side correlations are summarized in Table 4.1.
(4.23)
(4.24)
(4.25)
(4.26)
(4.27)
(4.28)
71
Table 4.1: Index of parameters used in the simulation code for refrigerant side correlations
Parameter Description Unit of
measure
Flow boiling heat transfer coefficient
Nucleate boiling heat transfer coefficient
Convective heat transfer coefficient,
referred to liquid, vapor or two-phase
Suppression factor
Enhancement factor
Prandtl number
Reynolds number
Surface roughness
Molecular mass of the fluid
Heat flux
Thermal conductivity
Hydraulic diameter
Length in flow direction
Surface tension
Gravitational acceleration
Density, referred to liquid or vapor phase
Confinement number
For the segments where single phase refrigerant is flowing, the heat transfer coefficient
is calculated using correlations based on Nusselt Number.
Gnielinski’s [57] correlation is employed with either liquid or vapor turbulent flow, and
it is defined as follows:
Where the Fanning Friction Factor is calculated through equation (4.30):
(4.29)
(4.30)
72
In order to characterize the laminar flow (Reynolds Number below 2300), a different
expression to calculate the Nusselt Number is used, adapted from Bergman [58] and
valid for constant heat flux:
The lubricant presence is considered introducing a correction factor , which
multiplies the Cooper’s correlation (4.24) and accounts for the liquid-phase mass
transfer effect on nucleate boiling contribution, as suggested by Thome [59]. The
correction factor takes into account the temperature glide typical of the multi-
component zeotropic mixtures evaporation, such as refrigerant-oil mixture.
Where:
is the boiling range, defined as the dew point temperature minus the
bubble point temperature of the mixture at its local liquid composition;
is the ideal heat transfer coefficient calculated using the Cooper
correlation (4.24);
is the mass transfer coefficient, whose value is , based on
comparison to numerous experimental pool boiling studies;
is the vaporization enthalpy.
The overall pressure drop occurring in the microchannel tubes is divided into three
main components: frictional, momentum and gravitational pressure drop.
The two-phase frictional pressure drop on the refrigerant side is predicted
using a correlation developed by Mishima and Hibiki [15] for flow inside capillary
vertical tubes with inner diameters in the range from 1 to 4 mm and Reynolds
Numbers in the range from 50 to 10000. Each phase is assumed to travel at its own
mean velocity, as described in Lockhart and Martinelli’s [60] approach. The two-
phase frictional pressure drop correlation is based on Chisholm’s equation [61], and
it is expressed as follows:
(4.31)
(4.32)
(4.33)
(4.34)
73
Where:
and are respectively the frictional pressure losses when either
single-phase liquid or vapor component flow in the tube;
is the two-phase friction multiplier. It multiplies the single-phase pressure
drop in order to obtain the two-phase pressure drop;
is the Lockhart-Martinelli’s parameter, defined as the square root of the ratio
between the single-phase liquid and single-phase vapor frictionl pressure drops.
The Chisholm’s parameter , which depends on the flow regime of each phase
(laminar or turbulent), is modified to consider the dependence on the hydraulic diameter
:
The momentum pressure drop is estimated according to Ragazzi’s [62] correlation,
considering the increase in flow momentum between the inlet and the outlet of each
segment:
Where is the mass flux flowing in each segment and is the void
fraction.
The void fraction is one of the most important parameters used to characterize the two-
phase flow and it is defined as the fraction of the flow-channel volume occupied by the
vapor phase or, alternatively, as the fraction of the channel cross-sectional area occupied
by the vapor phase. A void fraction model is necessary to calculate the volume occupied
by liquid and vapor phases in the main components of the heat exchanger, such as
headers and microchannel tubes. The correlation chosen is the one from Mandrusiak et
al. [63], since it is able to account for the high viscosity characterizing the liquid phase
in laminar flow:
(4.35)
(4.36)
(4.37)
(4.38)
74
Moreover, the simulation program accounts for the gravitational pressure drop in each
segment, with the following equation:
Where is the vertical length of each segment and the homogeneous density is
calculated through equation (4.40), as described in [12]:
The pressure drop across the residual two-phase components, such as the inlet and
outlet headers and liquid and suction lines, are determined through the following
equation suggested in [64]:
Where and are respectively the coefficient of local resistance in single-phase flow
and the component two-phase multiplier. The two-phase multiplier, in turn, can be
presented in a generalized form as:
Where is the experimental coefficient adjusted to pipe components, like bends,
ball valves, expansions and contractions, and is the ratio between single-phase liquid
and vapor pressure drop across the component.
4.5 Refrigerant-oil mixture properties calculation
When lubricant is injected in the refrigerant loop, the refrigerant-oil mixture has a
different behavior with respect to the pure refrigerant. The correlations used in the
simulation code account for this change. The thermodynamic approach, which considers
the refrigerant-oil mixture as a real mixture, is used as the baseline for the model
development [9] [39]. Since the vapor pressure of lubricant is very little compared to the
refrigerant one, the oil affects only the liquid phase and the oil composition in the vapor
phase can be considered negligible. Hence, only the liquid properties of the oil-
refrigerant mixture have to consider the modifications due to oil addition. The
parameter which accounts for the oil presence is the absolute oil mass fraction or oil
circulation rate, defined as the oil concentration when all the circulating fluid is in liquid
phase and form a homogeneous mixture:
(4.39)
(4.40)
(4.41)
(4.42)
75
Similarly to the expression of the two-phase refrigerant quality , the quality of a
refrigerant oil mixture can be calculated as:
The local oil mass fraction depends on the refrigerant-oil mixture quality and on
the absolute oil mass fraction. In the operating conditions range, the oil is assumed to be
miscible with refrigerant. As phase change occurs, the oil concentration becomes
enriched during the evaporation process. Based on mass conservation, from (4.43) and
(4.44), the local oil mass fraction is expressed as:
(4.45)
When the liquid phase is composed only by oil, is equal to . Therefore, the equation
(4.45) provides an important limit concerning the refrigerant-oil mixtures, since the
maximum refrigerant-oil mixture quality reachable, defined as is .
Once is reached, the refrigerant and oil enter the superheated region even if
part of the mixture is still in liquid phase. Hence, in the last part of the evaporator, the
mixture temperature can increase without increasing the quality, with strong effects on
heat transfer, pressure drop and oil retention.
The thermodynamic approach consists in using the bubble point temperature instead of
the pure refrigerant saturation temperature in the boiling heat transfer coefficient
calculations. Indeed oil addition to refrigerant entails an increase in the bubble point
temperature and in the local saturation temperature at which evaporation takes place. In
order to calculate the bubble point temperature, an empirical correlation firstly proposed
by Takaishi and Oguchi [65] and then improved by Thome [66] is used:
(4.46)
Where is the saturation pressure and is the bubble point
temperature. A and B are parameters depending on the local oil mass fraction ,
calculated as:
(4.47)
(4.43)
(4.44)
76
(4.48)
The values of the empirical parameters to and to have a negligible
dependence on the oil-refrigerant pair and are considered as constant values, while
and are very sensitive to the refrigerant used.
The oil addition makes the mixture behaving as a zeotropic mixture. Due to the increase
in the local saturation temperature at which evaporation takes place, the temperature
difference between the air and the mixture flowing into the microchannel tubes
decreases with respect to increase local oil mass fraction, determining a drop in the heat
transferred. The assumption at the base of the enthalpy calculation is that the
thermodynamic equilibrium exists throughout the refrigerant evaporation, as follows:
(4.49)
The equation (4.49) is referred to @ ( , so to for completely saturated liquid
and it is composed by three contributions:
- The first term is the latent heat to the fraction of liquid vaporized, ; is
the local vapor quality and is the latent heat of vaporization of the pure
refrigerant;
- The second term refers to the sensible heat to the fraction of the fluid in the
liquid phase ; is the specific heat of the liquid phase
refrigerant-oil mixture;
- The third term is the sensible heat to the fraction of fluid in the vapor phase
; is the specific heat of the pure refrigerant vapor.
The expression (4.49) reduces to the first term if oil is absent, because there is no
increase in the bubble point temperature due to oil addition.
The liquid specific heat of the refrigerant and oil mixture is calculated from a
correlation by Liley and Gambill [67]. The method, based on weighting the single
component values by their mass fractions, predicts as function of the local oil
mass fraction and the bubble point temperature, under the assumption of ideal mixing
and negligible mixing heats:
(4.50)
Moreover, another empirical correlation for predicting liquid specific heats of petroleum
oils, proposed in [67], is used when the manufacterers’ lubricant data are not available:
77
(4.51)
Where is the temperature of the oil and must be in the range s
is the liquid specific gravity at and has to be in the range .
Other important physical properties are affected by the lubricant addition: liquid
mixture density, viscosity, thermal conductivity and surface tension.
The density of the refrigerant and oil liquid mixture can be approximated to that of the
ideal mixture, using a correlation given by Jensen and Jackman [68]:
(4.52)
The oil viscosity is 2-3 order of magnitude higher than the refrigerant one. The mixture
viscosity increases during the evaporating process. At low qualities, the
high amount of refrigerant dissolved mitigates the overall liquid viscosity. At high
quality, when the liquid phase is mainly composed by non-evaporated lubricant, the
mixture viscosity reaches its maximum value. The increase in liquid mixture viscosity
has detrimental effects both on heat transfer coefficient and pressure drop. The most
important study on this topic is the one from Yokozeki [69], who formulated the
following equation to estimate the liquid viscosity of refrigerant and oil mixtures:
Where is called the effective weight fraction average of pure compounds, is the
molecular weight, is the mole fraction, is the number of compounds and is an
adjustable parameter determined from experimental data. Setting provides
accurate results for most refrigerant and lubricant pairs.
The thermal conductivity of the refrigerant and oil liquid mixture
is
calculated using the method by Filippov [70]:
(4.53)
(4.54)
(4.55)
78
The surface tension is the elastic tendency of liquids which makes them acquire the
least surface area possible. Refrigerant-oil mixture surface tension increases with the
increase of oil concentration, as carried out in previous studies [71], and influences the
wetting behavior. Higher surface tension increases the tube wetting with respect to the
pure refrigerant flow. The liquid mixture surface tension
is determined
according to Jensen and Jackman [68]:
4.6 Oil retention calculation
The overall oil retained in the evaporator is determined as the sum of the lubricant
trapped in both headers and inside the microchannel tubes. The heat exchanger internal
geometry and the behavior of the liquid mixture along the components are considered
the two main parameters affecting the oil retention. The internal geometry, especially in
the outlet header, constitutes a critical feature, since the microchannel tubes, while
entering the cylindrical header, creates internal small volumes, where the oil rich liquid
mixture accumulates and stops. The lubricant, due to its small vapor pressure, is always
in the liquid phase and create a liquid mixture with the refrigerant dissolved in it. The
local oil mass fraction increases with the vapor quality, that means the maximum
value of oil concentration is reached in the final part of the microchannel tubes and in
the outlet header. The equation used to describe the oil retention in the inlet header and
microchannel tubes is:
Where is the geometrical volume of the inlet header or the volume of
each segment composing the microchannel tube. The equation (4.57) provides the
lubricant retained calculated from the mass amount of liquid mixture.
The equation used for describing the behavior in the outlet horizontal header is the
following:
Where the term represents the volume of the valleys created
by the microchannel tubes entering the cylindrical header.
(4.56)
(4.57)
(4.58)
79
The two different equations describing the oil retention phenomenon in the
microchannel heat exchanger depend on the local oil mass fraction. In this way, it is
possible to reproduce both the capability of the liquid phase refrigerant to carry over
most of the lubricant droplets settled in the inner header and the incapability of the
vapor phase refrigerant to remove the oil trapped in the outlet header. Furthermore, the
second term of equation (4.58) reproduces the large lubricant accumulation in the outlet
header, where the high oil concentration liquid mixture, due to the large viscosity and
density values, is retained between the geometrical valleys.
The dividing factor in (4.58) is estimated by the header internal geometry proportions.
5 Simulation results without oil
In this chapter, the simulation results are compared with the experimental values, in
order to provide a reasonable and reliable validation. The code needs as input the
refrigerant table properties to calculate the exact conditions both at the inlet and outlet
of each segment. RefProp software is used to overcome this requirement. The saturated,
subcooled and superheated refrigerant properties are imported directly from RefProp to
a VBA excel spreadsheet and then generating an .idf file (type of format the Fortran
code is able to read) is generated.
RefProp, acronym for reference fluid properties, is a scientific software developed by
the National Institute of Standards and Technology (NIST), with the aim of providing
thermodynamic and transport property tables for the most common refrigerant fluids
and mixtures at different temperatures and pressures.
R410A, R32 and R1234yf are disclosed refrigerants, so their property tables in RefProp
database are very accurate. On the contrary, DR5A is a developmental refrigerant and
significant problems are found in modeling its table property. Since an accurate mixture
model is not available, few assumptions are made. The saturation thermodynamic and
transport properties available from the manufacturer are imported in a spreadsheet; the
missing subcooled and superheated properties, are determined through RefProp,
modeling the mixture composed by 68.9 wt.% R32 and 31.1 wt.% R1234yf refrigerants,
with the most suitable mixing rule. These assumptions are considered reasonable, since
the heat exchanger is mainly occupied by two-phase evaporating refrigerant.
5.1 Evaporator capacity and pressure drop validation
In this section a comparison between simulation and experimental results in terms of
overall heat exchanger capacity and pressure drop is provided. All the Labview
recorded data are averaged over a specific time interval and then used as simulation
input values for both refrigerant and air sides. The method used to determine the
refrigerant inlet enthalpy and pressure, which are two simulation input parameters, is
the same used during the experimental tests. The refrigerant inlet enthalpy is calculated
adding the heat supplied by the fixed power electrical preheaters and the heat gained
from the surroundings to the subcooled liquid refrigerant enthalpy measured at the
preheaters inlet. The refrigerant inlet pressure is determined adding the pressure
difference across the evaporator measured by the differential pressure transducer to the
82
reading from the high precision absolute pressure transducer placed at the microchannel
outlet.
Both in case of pure refrigerant and oil tests, the refrigerant mass flow rate used as
simulation input is the average value coming from the mass flow meter placed in the
refrigerant loop.
The data in Fig. 5.1, Fig. 5.2 and Fig. 5.3 are classified only based on the fluid and
evaporator used, regardless of the various testing conditions, as follows:
Blue hollow diamonds represent R410A tests with Evaporator A;
Orange solid triangles symbolize R410A experiments using Evaporator B;
Green solid diamonds refer to DR5A tests with Evaporator A;
Purple hollow triangles stay for DR5A experiments performed with Evaporator
B;
Black solid circles represent R32 tests conducted with Evaporator A;
Red hollow circles symbolize R1234yf experiments with evaporator A.
Fig. 5.1: Comparison between predicted and experimental capacity without oil
The simulation capacity is always in good agreement with the experimental value, in
particular the 96% of the data are predicted within an error smaller than 20%. The plot
above shows three main region, representing a capacity range from 3.8 to 10.5 ,
0
2
4
6
8
10
12
0 2 4 6 8 10 12
Sim
ula
tio
n C
ap
aci
ty [
kW
]
Experimental Capacity [kW]
R410A Evap A
R410A Evap B
DR5A Evap A
DR5A Evap B
R32 Evap A
R1234yf Evap A
83
representative of the different air side and refrigerant side conditions used during the
experimental tests. Although two different geometries and four distinct refrigerants are
used, the simulation program seems able to properly account for the wide range of
conditions analyzed.
Fig. 5.2: Comparison between predicted and experimental pressure drop without oil
Fig. 5.3: Particular of the comparison between predicted and experimental pressure drop
without oil
0
20
40
60
80
100
0 20 40 60 80 100
Sim
ula
tio
n P
ress
ure
Dro
p [
kP
a]
Experimental Pressure Drop [kPa]
R410A Evap A
R410A Evap B
DR5A Evap A
DR5A Evap B
R32 Evap A
R1234yf Evap A
Zoom in Fig. 6.3
0
5
10
15
20
25
0 5 10 15 20 25
Sim
ula
tio
n P
ress
ure
Dro
p [
kP
a]
Experimental Pressure Drop [kPa]
R410A Evap A
R410A Evap B
DR5A Evap A
DR5A Evap B
R32 Evap A
R1234yf Evap A
84
The simulation pressure drop shows a good agreement with the experimental results: the
92% of the data are predicted within an error smaller than 30%, while the maximum
error is 37%. Overall, the model is able to satisfactorily predict the pressure drop for
different refrigerants and large range of mass flow rates.
The contributions to the total pressure drop are many, such as the pressure drop in the
liquid line, inlet header, distributor, microchannel tubes and outlet header. The most
important pressure drop occurs in the distributor. The distributor is a perforated pipe
installed in the inlet header, with the purpose of directing the flow into the parallel
microchannel tubes and distributing the refrigerant homogeneously along the header
length. The holes in the distributor are modelled as single one-dimensional orifices and
an equivalent diameter is assigned to them. The pressure drop across the distributor is
calculated through equation (5.1), as suggested by [72]:
Where is the squared ratio between the cross sectional area of the distributor pipe
and the orifice equivalent diameter and is the contraction factor of the flow, equal to
0.8. No data from the manufacturer are available about the distributor hole equivalent
diameter for both the heat exchangers, so part of the modelling work consists in
determining the correct value. The equivalent orifice diameter is found to be 3.4 mm for
Evaporator A distributor and 2.95 mm for Evaporator B distributor.
A brief analysis, performed to better understand which are the main parameters
affecting the pressure drop calculation, shows that the simulation model is very
sensitive to input quality variations, which, in turn, depend on enthalpy variations. In
particular, when subcooled refrigerant is flowing at the inlet, the single-phase pressure
drop correlation under predicts the simulation result, especially the liquid line, inlet
header and distributor pressure drop terms. The difference between experimental and
simulation values reduces when two-phase refrigerant is flowing at the microchannel
heat exchanger inlet, which is the typical situation found in real applications.
Considering the uncertainties of the pressure transducers and thermocouples readings at
the fixed power electrical preheaters inlet, the accuracy of the voltage and current
measurements and a reasonable error in estimation, it is possible to calculate the
microchannel inlet enthalpy value uncertainty.
As an example, the R410A series with saturation temperature equal to 3.9 °C,
superheating degree of 8.3 °C and Evaporator A is presented. The calculated enthalpy
value at the microchannel inlet is . The results of two different
simulations, run with all the input parameters fixed except for the inlet refrigerant
(5.1)
85
enthalpy, equal to and respectively, are summarized in Table
5.1:
Table 5.1: Pressure drop without oil analysis
Oil mass fraction [%] 0 0
Total mass flow rate [g/s] 0.459 0.459
Evaporator inlet P [kPa] 899.22 899.22
Evaporator inlet h [kJ/kg] 209.5 206.5
Evaporator inlet quality [%] 2.3 0.8
Evaporator outlet P [kPa] 877.398 882.074
Evaporator outlet h [kJ/kg] 433.72 432.585
Evaporator outlet quality [%] 100 100
Refrigerant outlet degree of superheating [°C] 10.5 9.37
Evaporator capacity [kW] 10.09 10.174
ΔP total [kPa] 21.822 17.146
ΔP liquid line and inlet header [kPa] 3.546 1.815
ΔP distributor [kPa] 15.442 12.455
ΔP microchannel tubes [kPa] 1.491 1.55
ΔP outlet header [kPa] 0.512 0.505
ΔP suction line [kPa] 0.831 0.821
A small variation in refrigerant inlet enthalpy is able to change the inlet quality of the
two-phase refrigerant flow, affecting the pressure drop predicted value. Smaller inlet
qualities mainly reduce the pressure drop in the liquid line, inlet header and distributor.
The other contributions, such as the pressure drop in the microchannel tubes, outlet
header and suction line do not show any appreciable change over the refrigerant inlet
quality range analyzed.
Moreover, both the evaporator capacity and refrigerant outlet degree of superheating
present small variations while changing the inlet quality.
5.2 Further simulation results without oil
Besides overall capacity and pressure drop values, the simulation model provides more
specific information, such as the local Heat Transfer Coefficient and Capacity at each
one of the 100 segments that compose a microchannel tube. Fig. 5.4 and Fig. 5.5 show
the Heat Transfer Coefficient and Capacity, respectively.
86
Fig. 5.4: Predicted Heat transfer coefficient trend without oil
Fig. 5.4 shows the Heat Transfer Coefficient (HTC) at each one of the segments of the
microchannel tube.
At the inlet segments the quality is very low and the predominant heat transfer
mechanism is the nucleate boiling. As stated in [16], the nucleate boiling mechanism
proportionally decreases with increasing vapor quality. On the other hand, the
convective mechanism increases with vapor quality up to a maximum and then
drastically drops. The Heat Transfer Coefficient trend accounts for both the main
mechanisms and presents a monotonic reduction towards the onset of dryout at high
vapor qualities. The HTC reaches a constant value when all the liquid refrigerant is
evaporated and the flow becomes single-phase vapor.
0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
0 20 40 60 80 100
Hea
t T
ran
sfer
Co
effi
cien
t [k
W/(
m2K
)]
Segment
87
Fig. 5.5: Predicted Capacity trend without oil
Fig. 5.5 depicts the exchanged capacity at each one of the segments of one single
microchannel tube. The transferred capacity from the air to the refrigerant flowing in
the microchannel tubes depends both on the refrigerant Heat Transfer Coefficient and
on the temperature difference between air and refrigerant. The simulation model seems
to catch both the aspects: in the first segments, where the vapor quality is low, the
capacity is enhanced by the temperature difference between the fluids, reaches a
maximum value and then drops in the final part of the evaporator, when both the
refrigerant HTC and the temperature difference achieve their minimum.
0
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
0 20 40 60 80 100
Ca
pa
city
[W
]
Segment
6 Simulation results with oil
In the following chapter, the simulation results with oil are compared with the
experimental values, in order to understand how the model captures the oil effects on
predicted capacity, pressure drop and oil retention. The chapter is divided in two main
sections. In the first part, a global overview in terms of overall heat exchanger capacity,
pressure drop and oil retention is provided including all the refrigerant studied. In the
second part, a deeper analysis on experimental and simulation Heat Transfer Factor,
Pressure Drop Factor and Oil Retention Volume normalized for refrigerant R410A and
DR5A is given. Only R410A and DR5A refrigerants are extensively presented in order
to show how the simulation model behaves in two opposite sets of conditions: on one
hand a well-known and highly employed fluid, R410A, and on the other hand a new low
GWP refrigerant, DR5A.
6.1 Evaporator Capacity, Pressure Drop and Oil Retention
In this section a comparison between simulation and experimental results in terms of
overall heat exchanger capacity, pressure drop and oil retention mass is provided. In
Fig. 6.1, Fig. 6.2, Fig. 6.4 and Fig. 6.4, the collected data are classified only based on
the fluid and evaporator used, regardless of the various testing conditions, as follows:
Blue hollow diamonds represent R410A tests with Evaporator A;
Orange solid triangles symbolize R410A experiments using Evaporator B;
Green solid diamonds refer to DR5A tests with Evaporator A
Purple hollow triangles stay for DR5A experiments performed with Evaporator
B.
Black solid circles represent R32 tests with Evaporator A;
Red hollow circles symbolize R1234yf experiments with evaporator A.
90
Fig. 6.1: Comparison between predicted and experimental capacity with oil
The simulation capacity is always in good agreement with the experimental value, in
particular the 97% of the data are predicted within an error smaller than 20%.
According to the results, the model seems to consider properly the capacity reduction
caused by oil addition over a wide range of refrigerant side conditions, such as different
mass flow rates, saturation temperatures, geometries, degrees of superheating and
various fluids. The maximum capacity values are reached for the R410A tests, since in
these series a higher mass flux is used.
0
2
4
6
8
10
12
0 2 4 6 8 10 12
Sim
ula
tio
n C
ap
aci
ty [
kW
]
Experimental Capacity [kW]
R410A Evap A
R410A Evap B
DR5A Evap A
DR5A Evap B
R32 Evap A
R1234yf Evap A
91
Fig. 6.2 Comparison between predicted and experimental pressure drop with oil
Fig. 6.3: Particular of the comparison between predicted and experimental pressure drop with
oil
0
20
40
60
80
100
0 20 40 60 80 100
Sim
ula
tio
n P
ress
ure
Dro
p [
kP
a]
Experimental Pressure Drop [kPa]
R410A Evap A
R410A Evap B
DR5A Evap A
DR5A Evap B
R32 Evap A
R1234yf Evap AZoom in Fig. 7.3
0
5
10
15
20
25
30
35
40
0 5 10 15 20 25 30 35 40
Sim
ula
tio
n P
ress
ure
Dro
p [
kP
a]
Experimental Pressure Drop [kPa]
R410A Evap A
R410A Evap B
DR5A Evap A
DR5A Evap B
R32 Evap A
R1234yf Evap A
92
The predicted pressure drop in Fig. 6.2 and Fig. 6.3 shows a good agreement with the
experimental results. The 87.5% of the data are predicted within an error smaller than
30%. The simulation catches the pressure drop raise caused by oil addition, regardless
of the refrigerant and evaporator tested. R410A series conducted with Evaporator B
present the highest pressure drop, both predicted and experimentally measured, due to
the complex microchannel heat exchanger internal geometry and the high value of
refrigerant mass flow rate.
Fig. 6.4 Comparison between predicted and experimental oil retention mass
Fig. 6.4 points out two different trends in the simulation results: the oil retention mass in
R410A series is well predicted, while for DR5A, R32 and R1234yf results
underestimated. The reason is the different mass flux conditions used during the
experimental tests: indeed, the R410A series present the highest refrigerant mass flow
rate. The model underestimates the reduction in vapor-phase shear stress caused by
lower mass fluxes, which, in turn, determines higher oil retention values.
0
50
100
150
200
0 50 100 150 200
Sim
ula
tio
n O
R [
g]
Experimental OR [g]
R410A Evap A
R410A Evap B
DR5A Evap A
DR5A Evap B
R32 Evap A
R1234yf Evap A
93
6.2 R410A results with oil
In this section, the R410A series run with evaporator A, saturation temperature of 3.9
°C, mass flow rate of 163.3 kg/hr and degree of superheating equal to 8.3 °C is
presented.
The simulation HTF and PDF are calculated as the ratio between the predicted capacity
or pressure drop with oil and the corresponding values of the pure refrigerant baseline
with the same total mass flow rate, in order to reproduce the procedure used for the
experimental results. The simulation ORVN is estimated dividing the predicted oil
retention mass ORM by the lubricant density and then normalizing with respect to the
Evaporator A internal volume. In Table 6.1 are summarized the main parameters of the
simulation input and output. The procedure used to generate the following table is
keeping constant the total refrigerant mass flow rate entering the evaporator and
changing the oil mass fraction. In this way, it is possible to compare three different tests
with the same exact conditions at the microchannel inlet and observe the lubricant
addition effects on the outlet parameters.
Table 6.1: Simulation input and output parameters for R410A tests
Oil mass fraction [%] 0 0.9 2.55
Total mass flow rate [g/s] 0.459 0.459 0.459
Evaporator inlet P [kPa] 893.59 893.59 893.59
Evaporator inlet T [°C] 2.981 2.987 2.988
Evaporator inlet h [kJ/kg] 209.32 209.32 209.32
Evaporator inlet quality [%] 0.022 0.022 0.022
Evaporator outlet P [kPa] 877.069 876.192 875.447
Evaporator outlet T [°C] 13.535 13.371 13.218
Evaporator outlet h [kJ/kg] 433.77 433.619 433.475
Evaporator outlet quality [%] 100 99.1 97.45
Evaporator capacity [kW] 10.1 10.093 10.087
ΔP total [kPa] 16.521 17.398 18.143
ΔP liquid line and inlet header [kPa] 3.627 3.595 3.584
ΔP distributor [kPa] 10.053 9.942 9.913
ΔP microchannel tubes [kPa] 1.497 1.987 2.488
ΔP outlet header [kPa] 0.512 0.586 0.616
ΔP suction line [kPa] 0.832 1.288 1.542
Total oil retained [g] 0 65.553 86.272
Oil retained inlet header [g] 0 3.96 7.361
Oil retained channels [g] 0 5.677 16.046
Oil retained outlet header [g] 0 55.916 62.865
94
Fig. 6.5 shows the comparison between predicted and experimental Heat Transfer
Factor for the R410A series considered. Red solid squares represent the simulation
results, while blue solid diamonds symbolize the experimental ones.
Fig. 6.5: R410A simulation and experimental Heat Transfer Factor
The model is able to predict both the experimental Heat Transfer Factor trend and value.
The simulation accounts for the reduction in the overall capacity due to the oil replacing
refrigerant, also underlined by the evaporator outlet temperature and enthalpy decrease.
Since oil normal boiling points are about 300 °C or higher, the lubricant is considered a
non-volatile component in the refrigerant and oil mixture. Hence, from the inlet to the
outlet of the microchannel heat exchanger, the oil increases its temperature without
boiling and prevents a certain amount of refrigerant from evaporating. Increasing the oil
mass fraction circulating into the evaporator, the temperature reduction in the
microchannel heat exchanger is amplified and the amount of non-boiling liquid
refrigerant is higher, with important consequences in the capacity decrease, as shown in
Table 6.1.
The oil addition also affects the Heat Transfer Coefficient (HTC), whose decrease
contributes to the overall coil capacity reduction.
In Fig. 6.6 the effect of different oil concentrations on the refrigerant side Heat Transfer
Coefficient is shown. The solid blue diamonds represent the HTC of the pure
refrigerant, the solid green diamonds symbolize the HTC with circulating oil mass
0.85
0.9
0.95
1
1.05
0 1 2 3
HT
F [
-]
OMF [%]
Experimental data
Simulation data
95
fraction equal to 1% and, lastly, the solid red diamonds stand for the HTC when the
OMF is 3%.
Fig. 6.6: Effect of different oil concentrations on the simulation Heat Transfer Coefficient
The simulation Heat Transfer Coefficient proportionally decreases after increasing the
oil mass fraction. The lubricant presence determines an additional mass transfer
resistance, which affects both the nucleate and convective boiling terms, and results in
the two-phase flow boiling HTC reduction.
The comparison between predicted and experimental Pressure Drop Factor is presented
in Fig. 6.7. The red solid squares represent predicted results, while blue solid diamonds
symbolize the experimental ones.
0
1
2
3
4
5
0 20 40 60 80 100
HT
C [
kW
/(m
2K
)]
Segment
NO OIL
OMF 1%
OMF 3%
96
Fig. 6.7: R410A simulation and experimental Pressure Drop Factor
The predicted PDF has the same trend, but slightly lower values with respect to the
experimental results. From Table 6.1, it is possible to analyze the pressure drop in each
heat exchanger section. The oil addition has different consequences whether the quality
is high or low. On one hand, in the first part of the evaporator, where a big amount of
liquid refrigerant is flowing and local oil concentrations have moderate values, the
change in liquid mixture physical properties is very small and negligible. On the other
hand, at the outlet of the evaporator, where superheated vapor and rich oil liquid
mixture are present, the high local oil concentration makes the liquid mixture very
viscous and the pressure drop increases.
The simulation (red solid squares) and experimental (blue solid diamonds) results in
term of Oil Retention Volume normalized are plotted in Fig. 6.8.
0.8
1
1.2
1.4
1.6
0 1 2 3
PD
F [
-]
OMF [%]
Experimental data
Simulation data
97
Fig. 6.8 R410A simulation and experimental Oil Retention Volume normalized
The model well predicts both the trend and the values of oil retention in the
microchannel evaporator. The correlation used to calculate the amount of oil retained is
quite unaffected by the OMF increase, since it reproduces the filling process
encountered in the experiments, caused by the outlet header internal geometry. As a
consequence, the ORVN raises exponentially between the no oil condition (OMF 0%)
and the low oil mass fraction condition (OMF 1%), and then increases proportionally
further increasing the OMF. Moreover, the Oil Retention Volume normalized results
little overestimated for OMF equal to 1%.
From Table 6.1 it is possible to understand the oil retention contributions of the
different heat exchanger sections. The component more affected by the oil retention
phenomenon is the outlet header, where the highest value of local oil concentration is
reached. The other sections, such as the inlet header and the microchannel tubes, present
a smaller amount of oil retained.
The refrigerant mass and the oil retention mass trend at each one of the 100 segments
that compose a microchannel tube is shown in Fig. 6.9. The blue solid diamonds
symbolize the vapor-phase refrigerant mass, while the red solid squares stand for the
liquid-phase refrigerant mass. The total amount of refrigerant inventory is represented
by the solid green triangles. Lastly, the purple solid circles refer to the mass of oil
retained in each segment.
0
0.02
0.04
0.06
0.08
0 1 2 3
OR
VN
[-]
OMF [%]
Experimental data
Simulation data
98
Fig. 6.9: Refrigerant and lubricant masses trend inside a microchannel tube
At the microchannel entrance, the refrigerant liquid mass reaches its maximum and then
decreases continuously up to the complete evaporation. The amount of vapor, calculated
through the refrigerant mass balance in each segment, has the opposite trend and raises
while approaching the exit. Lastly, the lubricant retained increases considerably only at
high vapor qualities, when the local oil concentration reaches appreciable values.
6.3 DR5A results with oil
In this section, the DR5A series run with evaporator A, saturation temperature of 3.9
°C, mass flow rate of 90 kg/hr and degree of superheating equal to 8.3 °C is analyzed.
The simulation Heat Transfer Factor, Pressure Drop Factor and Oil Retention Volume
normalized are calculated using the same procedure illustrated in the previous R410A
section. Moreover, in Table 6.2, generated keeping constant the total refrigerant mass
flow rate at the evaporator inlet and changing the oil mass fraction, the simulation
inputs and outputs are listed.
0
4
8
12
16
0 20 40 60 80 100
Ma
ss [
g]
Segment
Vapor mass
Total refrigerant mass
Oil retained mass
Liquid mass
99
Table 6.2: Simulation input and output parameters for DR5A tests
Oil mass fraction [%] 0 0.97 2.86
Total mass flow rate [g/s] 0.255 0.255 0.255
Evaporator inlet P [kPa] 886.49 886.49 886.49
Evaporator inlet T [°C] 4.386 4.399 4.401
Evaporator inlet h [kJ/kg] 218.96 218.96 218.96
Evaporator inlet quality [%] 0.047 0.047 0.047
Evaporator outlet P [kPa] 877.802 877.688 877.393
Evaporator outlet T [°C] 11.373 11.34 11.328
Evaporator outlet h [kJ/kg] 468.723 468.453 468.45
Evaporator outlet quality [%] 100 99.03 97.14
Evaporator capacity [kW] 6.244 6.24 6.237
ΔP total [kPa] 8.688 8.802 9.097
ΔP liquid line and inlet header [kPa] 2.559 2.44 2.423
ΔP distributor [kPa] 4.857 4.624 4.583
ΔP microchannel tubes [kPa] 0.8 1.02 1.247
ΔP outlet header [kPa] 0.176 0.212 0.223
ΔP suction line [kPa] 0.296 0.506 0.621
Total oil retained [g] 0 96.626 130.821
Oil retained inlet header [g] 0 3.876 7.589
Oil retained channels [g] 0 6.245 18.358
Oil retained outlet header [g] 0 86.505 104.874
In Fig. 6.10, the comparison between predicted and experimental Heat Transfer Factor
for the DR5A series is presented. Blue solid diamonds represent the simulation results,
while red solid squares symbolize the simulation ones.
100
Fig. 6.10: DR5A simulation and experimental Heat Transfer Factor
The predicted HTF values catches the experimental trend value. The capacity reduction
due to oil presence, with respect to the pure refrigerant simulation results, is underlined
by the enthalpy and temperature reduction at the microchannel heat exchanger outlet.
The oil and refrigerant mixture behaves like a zeotropic mixture, so the temperature at
which evaporation takes place changes and increases from the microchannel inlet to the
outlet. The raise in the refrigerant and lubricant mixture along the evaporator length
reduces the temperature difference between the air and the refrigerant side and
decreases the heat exchanger capacity.
Fig. 6.11 shows the comparison between predicted and experimental Pressure Drop
Factor. Red solid squares stand for the simulation results, while blue solid diamonds
refer to the experimental ones.
0.85
0.9
0.95
1
1.05
0 1 2 3
HT
F [
-]
OMF [%]
Experimental data
Simulation data
101
Fig. 6.11: DR5A simulation and experimental Pressure Drop Factor
The simulation PDF is able to reproduce the experimental results. The predicted and
experimental DR5A Pressure Drop Factor rapidly increases between the no oil
condition and the low oil mass fraction condition (OMF equal to 1%), while it is quite
unaffected by a further increase in the circulating oil mass fraction. The simulation code
properly accounts for both the lubricant detrimental effects with the different OMFs
tested.
The comparison between experimental and predicted Oil Retention Volume normalized
is plotted in Fig. 6.12. The blue solid diamonds and the red solid squares depict the
experimental and simulation results respectively.
0.8
1
1.2
1.4
1.6
0 1 2 3
PD
F [
-]
OMF [%]
Experimental data
Simulation data
102
Fig. 6.12: DR5A simulation and experimental Oil Retention Volume normalized
The simulation results on one hand under predict the amount of oil retained inside the
microchannel evaporator and on the other hand capture the trend found in the
experimental data, that is an increase in the ORVN with respect to the circulating oil
mass fraction.
The DR5A series is run with low refrigerant mass flow rate and then low mass flux.
Low mass fluxes determine low vapor-phase velocity and shear stress, so the vapor
refrigerant is less effective in removing the lubricant droplets, resulting in higher
amount of oil retained. The different numerical values between experimental and
predicted ORVN shows how the simulation underestimates the effect of different
refrigerant mass flow rates values on oil retention calculation.
The simulation normalized oil retention and the mixture quality trends in the
microchannel tubes are plotted in Fig. 6.13. The normalized oil retention is a
dimensionless parameter representing the ratio between the local oil retention and the
maximum oil retention in the circuitry.
0
0.03
0.06
0.09
0.12
0.15
0 1 2 3
OR
VN
[-]
OMF [%]
Experimental data
Simulation data
103
Fig. 6.13: Dimensionless oil retention and mixture quality trends inside a microchannel tube
In the first part of the evaporator, almost all the refrigerant is in the liquid phase and the
mixture qualities are very low. Although along the heat exchanger the refrigerant
evaporates and the mixture quality monotonically increases, the normalized oil retention
has initially a decreasing trend and only after the segment #60 starts to raise, reaching
the maximum value at the outlet. The simulation results point out how in the first
segments the reduction in liquid volume is predominant over the increase in local oil
concentration, with a consequent reduction in the normalized oil retention values, while
in the last segment the behavior is the opposite.
0
0.2
0.4
0.6
0.8
1
1.2
0 20 40 60 80 100No
rma
lize
d O
R [
-] &
Mix
ture
qu
ali
ty [
-]
Segment
Normalized oil retained
Mixture quality
7 Conclusions
In this chapter the conclusions both about the experimental and simulation works are
presented.
7.1 Conclusions about the experimental work
The oil presence always penalizes the microchannel heat exchanger performances, both
in terms of heat transfer degradation and pressure drop increase.
The oil addition decreases the evaporator capacity and the reduction is proportional to
the circulating oil mass fraction. The lubricant and refrigerant mixture behaves as a
zeotropic mixture and the evaporating process is described by the bubble point
temperature instead of the pure refrigerant saturation temperature. The bubble point
temperature increases along the evaporator length and reaches its maximum value at the
microchannel heat exchanger outlet. During the evaporating process the lubricant
increases its temperature without boiling and prevents a certain amount of refrigerant
from evaporating. Furthermore, the oil presence determines an additional heat transfer
resistance, which results in lower evaporator outlet temperature, degree of superheating,
and heat transferred with respect to the pure refrigerant case.
The parameters which mainly affect oil retention is the refrigerant mass flow rate
circulating into the system and the refrigerant degree of superheating. On one hand,
higher refrigerant mass flow rates determine higher capability of the refrigerant vapor to
carryover the oil droplets settled in the microchannel internal volume, resulting in lower
amount of lubricant retained. On the other hand, low degrees of superheating cause a
larger amount of refrigerant dissolved in the liquid mixture, determining low viscosity,
that, in turn, entails to lower amount of lubricant retained.
The comparison between R410A and DR5A demonstrates better DR5A performances in
terms of Heat Transfer Factor. Over the wide range of conditions analyzed, the DR5A
refrigerant always presents higher HTF than R410A, which means lower capacity
penalization due to oil addition. The maximum capacity reduction observed in the
experimental data, with respect to the pure refrigerant case, is about 15% for R410A at
OMF equal to 5%.
The non-evaporating refrigerant, which determines two-phase flow inside the
microchannel headers and tubes, adds up to the lubricant and creates a mixture, whose
viscosity is higher than the pure refrigerant one; these phenomena make the pressure
106
drop across the heat exchanger always increasing after oil addition. Moreover, the
lubricant creates a film inside the tubes, which reduces the available flow cross
sectional area and increases liquid and vapor velocities, contributing to the overall
pressure drop increase. The pressure drop raise, with respect to the pure refrigerant case,
is proportional to the oil mass fraction circulating into the system.
The Oil Retention Volume normalized seems to be strongly dependent on the oil mass
fraction; the OMF increase always leads to larger amount of oil trapped in the
microchannel heat exchanger. In tests conducted with Evaporator A, the ORVN shows
an exponential raise at small OMF values and then a proportional increase with higher
OMFs, especially at high degrees of superheating and low mass fluxes. This behavior
can be explained by the internal geometry of the Evaporator A horizontal outlet header.
In Fig. 7.1 it is possible to observe the small pockets created by the tubes entering the
header, which are filled by the oil rich liquid mixture because of the gravity force.
Fig. 7.1: Particular of the Evaporator A horizontal outlet header
In Evaporator B experimental data lower lubricant accumulation are observed.
Furthermore, low refrigerant mass flow rates lead to large amount of oil retained,
regardless of the evaporator geometry and fluid tested. The maximum ORVN value is
about 0.15 when OMF is equal to 5%, both for DR5A and R410A, that means the oil
occupies the 15% of the evaporator internal volume.
The experimental tests show that DR5A can be considered as a good low GWP
replacement for R410A. No particular difference is observed in terms of ORVN and
PDF between the two fluids, but the HTF shows slightly better performances in terms of
capacity after oil addition. Moreover, DR5A has a larger phase-change enthalpy than
107
R410A, which is a positive feature from an environmental perspective, thus it is
possible to obtain the same R410A capacity with a lower refrigerant charge.
The comparison between R1234yf, R32 and DR5A, performed under the same
refrigerant side conditions, shows similar behavior between all the fluids in terms of
HTF, PDF and ORVN. R32 experimental series presents lower HTF and smaller ORVN.
The possible explanation for this trend is the larger reduction in the refrigerant degree of
superheat observed at the microchannel outlet after oil addition. As consequences, the
lower outlet temperature on one hand worsens the R32 capacity with respect to the pure
refrigerant case and on the other hand determines a larger amount of refrigerant
dissolved in the liquid mixture, which leads to smaller viscosity and lower amount of oil
retained in the microchannel heat exchanger.
7.2 Conclusions about the simulation work
The simulation capacity and pressure drop without oil are always in good agreement
with the experimental values. Considering two different evaporator geometries, four
refrigerants, and a wide range of conditions both on the refrigerant and air sides, the
96% of the capacity data are predicted within an error smaller than 20%, while the 92%
of the pressure drop is predicted within an error smaller than 30%.
The simulation pressure drop calculation is very sensitive to the inlet enthalpy and
quality. In particular, the pressure drop is underestimated when the input parameters
correspond to subcooled liquid refrigerant flow. Since most of the heat exchanged is
latent, the capacity prediction is only slightly affected by the inlet enthalpy and quality
variations.
The simulation capacity and pressure drop with oil show good results compared to the
experimental data. The evaporator model is able to account for the capacity reduction
and pressure drop increase due to oil addition. As found in the experiments, the
variations are proportional to the circulating oil mass fraction. The model catches both
the heat transfer coefficient reduction and the decrease in temperature and enthalpy at
the microchannel heat exchanger outlet due to the non-evaporating liquid refrigerant
mixed with oil. Furthermore, the simulation pressure drop addresses most of the
pressure drop variations in the last part of the evaporator (microchannel tubes and outlet
header), where the high local lubricant concentration makes the liquid mixture very
viscous.
The simulation oil retention has two different trends; R410A data are well predicted,
while the DR5A, R32 and R1234yf results are underestimated. The model addresses
most of the oil retention to the horizontal outlet header, where the highest value of local
oil concentration and the maximum liquid viscosity are reached. The amount of
108
lubricant retained presents a sudden raise between no oil condition and low OMF
condition, while with further increasing the OMF, it increases slightly. This trend well
reproduces the filling phenomenon due to the small volumes created by the
microchannel tube protrusions entering the header. The code well predicts the oil
retention in case of high mass flux, while underestimates it for low mass fluxes.
7.3 Future works
In this work the entire heat exchanger, that is the inlet header, the microchannel tubes
and the outlet header all together, is considered. This approach on one hand provides
accurate results in terms of heat transfer rate, pressure drop and overall oil retention, on
the other hand presents limitations, such as inability of measuring local heat transfer
coefficient and observing flow regime.
An interesting improvement would be to extend the results to new heat exchangers with
different orientations and flow configurations, such as vertical inlet and outlet headers
and horizontal tubes.
Further analysis could be conducted to understand the internal geometry effects on oil
retention, especially in the inlet and outlet headers, which strongly affect the oil trapped
inside the microchannel heat exchangers. Oil retention studies in headers are
recommended in future works.
The method used to calculate the experimental oil retention shows for the new low
GWP refrigerants the same trends encountered in the literature for well known fluids. A
further but time-consuming improvement could be developing a new procedure to
provide a direct oil retention measurement. In [73], an example of oil retention direct
measurement is presented. Once the system stabilizes after lubricant injection, closing
simultaneously the valves on both ends of the evaporator, allows isolating the oil
amount present into the heat exchanger. Finally, it is possible to determine the oil
retention mass removing and weighing the test section.
Another future work is testing low GWP refrigerants with new generation oils, such as
nanolubricants, in order to understand if it is possible to reduce the detrimental effects
related to oil addition.
In the present work the refrigerant is always considered well distributed among the
microchannels. The uniformity is guaranteed controlling the refrigerant conditions at the
evaporator inlet. Studies about the maldistribution effects on heat transfer rate, pressure
drop and oil retention, as well as about the parameters that mainly affect non-uniform
distribution are recommended in future works.
Nomenclature
Symbols
A Area [m2]
Area ratio between distributor tube and equivalent hole diameter [-]
BWMO Blended white mineral oil
C Capacity rate [J/(Ks)]
Experimental coefficient adjusted to pipe components
Coefficient of discharge of the nozzle [-]
Capacity ratio [-]
CFC Chlorofluorocarbon
Chisholm’s parameter [-]
Co Confinement number [-]
COP Coefficient of performance [-]
Specific heat [J/(kgK)]
Hydraulic diameter [mm]
Major diameter [mm]
err Error
Fanning fiction factor [-]
F Convective boiling enhancement factor [-]
Cooper’s correlation correction factor
Fin length [mm]
Fin pitch [mm]
Fin thickness [mm]
g Gravitational acceleration [m/s2]
G Mass flux [kg/(m2s)]
GWP Global warming potential
Enthalpy [kJ/(kgK)]
HCFC Hydrochlorofluorocarbon
HFC Hydrofluorocarbon
HFO Hydrofluoroolefin
HTC Heat transfer coefficient [W/m2K)]
HTF Heat transfer factor [-]
j-factor [-]
k Thermal conductivity [W/(mK)]
Contraction factor of the flow
110
L Length [mm]
Louver length [mm]
Louver pitch [mm]
Mass flow rate [kg/s]
Fin parameter [m-1
]
M Molecular mass [kg/kmol]
MO
Mineral oil
NTU
Number of transfer units [-]
Nu Nusselt number
OCR
Oil circulating ratio [-]
ODP Ozone Depletion Potential
OMF
Oil mass fraction [-]
ORM
Oil retention mass [g]
ORV
Oil retention volume [dm3]
ORVN
Oil retention volume normalized [-]
Pressure [kPa]
Pressure difference across the nozze
Perimeter
PAG Polyalkylene glycol oil
Pressure drop factor
POE
Polylester oil
Pr Prandtl number
q’’ Heat flux [W/m2]
Capacity [kW]
Surface roughness [μm]
Re Reynolds number
rpm Rounds per minute
RTD
Resistance temperature detector
s Liquid specific gravity [-]
S
Solubility [g/g]
Nucleate boiling suppression factor [-]
t
Time [s]
T
Temperature [K]
Tube depth [mm]
Tube pitch [mm]
thk Thickness [mm]
u Uncertainty
U Overall heat transfer coefficient [W/(m2K)]
Specific volume [m3/kg]
111
Adjusted specific volume [m3/kg]
V
Volume [dm3]
Volumetric flow rate [CFM]
VG Viscosity grades
VLE
Vapor-liquid equilibrium
VF Void fraction [-]
VFD
Variable frequency drive
w Local oil mass fraction [-]
Electrical work [kW]
x Vapor quality [-]
Lockhart-Martinelli parameter [-]
Greek symbols
α
Heat transfer coeffcient [W/(m2K)]
Ideal heat transfer coefficient [W/(m2K)]
Two-phase multiplier of the component
Mass transfer coefficient [m/s]
Δ Delta, difference
ε
Heat transfer effectiveness [-]
η Effectiveness for fin or finned surface
θ Louver angle [-]
ϑ Liquid phase only/vapor phase only pressure drop ratio
μ Viscosity [mm2/s]
ξ Effective weight fraction average of pure compounds [-]
Coefficient of local resistance in single-phase flow
ρ Density [kg/m3]
σ Surface tension [N/m]
Two-phase fiction multiplier
ψ Mole fraction [-]
ω Absolute humidity [gvap/gdry air]
Subscripts
air
Air side
bp Boiling range
bub
Bubble point
c Cross-sectional
conv Convective
112
FB Flow boiling
gain
Heat gained
grav gravitational
HB
Heat balance
in
Inlet
inj
Injected
liq or l Liquid phase
lv Latent or vaporization
max Maximum
min Minimum
mix Refrigerant and oil mixture
most Most air
MCHX
Microchannel heat exchanger
n
Nozzle
NB Nucleate boiling
oil
Lubricant/oil
out
Outlet
preh
Electrical preheater
ref
Refrigerant side
r,s Saturation with respect to refrigerant temperature
s Surface
sat Saturation
segm Segment
ss Steady state conditions
tp Two-phase
vap or v Vapor phase
wet Wet conditions
References
1. Garimella, S., Innovations in energy efficient and environmentally friendly
space-conditioning systems. Energy, 2003. 28(15): p. 1593-1614.
2. Protocol, M., Protocol on substances that deplete the ozone layer.,'.
International Legal Materials, 1987. 26: p. 1550-61.
3. Smith, J., Heat Transfer and Pressure Drop of New LGWP Refrigerants and
Lubricant Mixtures in a 9.5 mm Micro-Finned Tube Evaporator. 2015
ASHRAE Annual Meeting, Atlanta, GA, USA.
4. Leck, T.J., et al., Novel Reduced GWP Refrigerant Compositions for Stationary
Air Conditioning. International Refrigeration and Air Conditioning Conference,
Purdue, IN, 2014.
5. Committee, E., Regulation (EC) No 842: 2006 of the European parliament and
of the council of 17 May 2006 on certain fluorinated greenhouse gases. Official
Journal of the European Union, 2006. 161: p. 1-8.
6. Union, E., Directive 2006/40/EC of the European Parliament and of the Council
of 17 May 2006 relating to emissions from air-conditioning systems in motor
vehicles and Amending Council Directive 70/156/EEC. Off J Eur Union, 2006.
1.
7. Leck, T.J. Evaluation of HFO-1234yf as a Potential Replacement for R-134a in
Refrigeration Applications. in 3rd IIR Conference on Thermophysical Properties
and Transfer Processes of Refrigerants Boulder, CO. 2009.
8. Mermond, Y., M. Feidt, and C. Marvillet, Thermodynamic and physical
properties of mixtures of refrigerants and oils. International Journal of
Refrigeration, 1999. 22(7): p. 569-579.
9. Thome, J.R., Comprehensive thermodynamic approach to modeling refrigerant-
lubricating oil mixtures. HVAC&R Research, 1995. 1(2): p. 110-125.
10. Kedzierski, M. Refrigerant/lubricant mixture boiling heat transfer research at
NIST. in Sixteenth National Convention of Mechanical Engineers and All India
Seminar on Future Trends in Mechanical Engineering, Research and
Development, Dept. of Mech. & Ind. Eng., UOR, Roorkee. 2000.
11. Kedzierski, M., Use of fluorescence to measure the lubricant excess surface
density during pool boiling. International journal of refrigeration, 2002. 25(8): p.
1110-1122.
12. Collier, J.G. and J.R. Thome, Convective boiling and condensation. 1994:
Oxford University Press.
13. Kattan, N., J. Thome, and D. Favrat, Flow boiling in horizontal tubes: Part 1—
Development of a diabatic two-phase flow pattern map. Journal of Heat
Transfer, 1998. 120(1): p. 140-147.
114
14. Wojtan, L., T. Ursenbacher, and J.R. Thome, Investigation of flow boiling in
horizontal tubes: Part I—A new diabatic two-phase flow pattern map.
International Journal of Heat and Mass Transfer, 2005. 48(14): p. 2955-2969.
15. Mishima, K. and T. Hibiki, Some characteristics of air-water two-phase flow in
small diameter vertical tubes. International Journal of Multiphase Flow, 1996.
22(4): p. 703-712.
16. Shen, B. and E.A. Groll, Review Article: A Critical Review of the Influence of
Lubricants on the Heat Transfer and Pressure Drop of Refrigerants, Part 1:
Lubricant Influence on Pool and Flow Boiling. HVAC&R Research, 2005.
11(3): p. 341-359.
17. Manwell, S. and A. Bergles, Gas-liquid flow patterns in refrigerant-oil mixtures.
ASHRAE Trans, 1990. 96(2): p. 456-464.
18. McMullan, J., et al., The influence of oil viscosity and refrigerant quality on
evaporator performance. International journal of energy research, 1992. 16(7):
p. 567-581.
19. Cho, K. and S.-J. Tae, Evaporation heat transfer for R-22 and R-407C
refrigerant–oil mixture in a microfin tube with a U-bend. International journal of
refrigeration, 2000. 23(3): p. 219-231.
20. Nidegger, E., J.R. Thome, and D. Favrat, Flow boiling and pressure drop
measurements for R-134a/oil mixtures Part 1: evaporation in a microfin tube.
HVAC&R Research, 1997. 3(1): p. 38-53.
21. Zürcher, O., J.R. Thome, and D. Favrat, Flow boiling and pressure drop
measurements for R-134a/oil mixtures Part 2: Evaporation in a plain tube.
HVAC&R Research, 1997. 3(1): p. 54-64.
22. Zürcher, O., J.R. Thome, and D. Favrat, In-tube flow boiling of R-407C and R-
407C/Oil Mixtures Part I: Microfin tube. Hvac&R Research, 1998. 4(4): p. 347-
372.
23. Zürcher, O., J.R. Thome, and D. Favrat, In-tube flow boiling of R-407C and R-
407C/Oil Mixtures Part II: Plain tube results and predictions. Hvac&R
Research, 1998. 4(4): p. 373-399.
24. Fukuta, M., et al., Flow characteristics of oil film in suction line of refrigeration
cycle. International Refrigeration and Air Conditioning Conference, Purdue, IN,
2000.
25. Sethi, A. and P. Hrnjak, Oil retention and pressure drop of R1234yf and R134a
with POE ISO 32 in suction lines. HVAC&R Research, 2014. 20(6): p. 703-720.
26. Cremaschi, L., Y.H. Hwang, and R. Radermacher, Investigation of oil retention
in residential heat pumps. 2004.
27. Cremaschi, L., Y. Hwang, and R. Radermacher, Experimental investigation of
oil retention in air conditioning systems. International Journal of Refrigeration,
2005. 28(7): p. 1018-1028.
28. Li, H. and P. Hrnjak, Effect of lubricant on two-phase refrigerant distribution in
microchannel evaporator. 2013, SAE Technical Paper.
115
29. Tuo, H., A. Bielskus, and P. Hrnjak, Experimentally validated model of
refrigerant distribution in a parallel microchannel evaporator. 2012, SAE
Technical Paper.
30. Jiang, H., Development of a simulation and optimization tool for heat exchanger
design. University of Maryland, 2003.
31. Lottin, O., P. Guillemet, and J.-M. Lebreton, Effects of synthetic oil in a
compression refrigeration system using R410A. Part I: modelling of the whole
system and analysis of its response to an increase in the amount of circulating
oil. International Journal of Refrigeration, 2003. 26(7): p. 772-782.
32. Lottin, O., P. Guillemet, and J.-M. Lebreton, Effects of synthetic oil in a
compression refrigeration system using R410A. Part II: quality of heat transfer
and pressure losses within the heat exchangers. International journal of
refrigeration, 2003. 26(7): p. 783-794.
33. Henderson, D., Solubility, Viscosity and Density of Refrigerant/Lubricant
Mixtures. 1994, Final Tech. Rept. DOE/CE/23810-34, Spauschus Assoc.
34. Gungor, K. and R. Winterton, Simplified general correlation for saturated flow
boiling and comparisons of correlations with data. Chemical engineering
research & design, 1987. 65(2): p. 148-156.
35. Yan, Y.-Y. and T.-F. Lin, Evaporation heat transfer and pressure drop of
refrigerant R-134a in a plate heat exchanger. Journal of Heat Transfer, 1999.
121(1): p. 118-127.
36. Bivens, D. and A. Yokozeki, Heat transfer coefficients and transport properties
for alternative refrigerants. International Refrigeration and Air Conditioning
Conference, Purdue, IN, 1994.
37. Li, H. and P.S. Hrnjak, An Experimentally Validated Model for Microchannel
Heat Exchanger Incorporating Lubricant Effect. International Refrigeration and
Air Conditioning Conference, Purdue, IN, 2014.
38. Li, H. and P. Hrnjak, Quantification of liquid refrigerant distribution in parallel
flow microchannel heat exchanger using infrared thermography. Applied
Thermal Engineering, 2015. 78(0): p. 410-418.
39. Jin, S. and P. Hrnjak, An Experimentally Validated Model for Predicting
Refrigerant and Lubricant Inventory in MAC Heat Exchangers. 2014, SAE
Technical Paper.
40. Deokar, P.S., Development of an experimental methodology for measurement of
oil retention and its effect on the microchannel heat exchanger. 2013:
OKLAHOMA STATE UNIVERSITY.
41. ASHRAE, ANSI/ASHRAE Standard 116-1995: Methods of Testing for Rating
Seasonal Efficiency of Unitary Air Conditioners and Heat Pumps. 1995.
42. ASHRAE, ANSI/ASHRAE Standard 41.1-1986: Standard Method for
Temperature Measurement. 1986.
116
43. Cavestri, R.C. and W.R. Schafer, Measurement of solubility, viscosity, and
density of R-410A refrigerant/lubricant mixtures. 2000, Imagination Resources,
Inc., Dublin, OH (US).
44. Taylor, B.N., Guidelines for Evaluating and Expressing the Uncertainty of NIST
Measurement Results (rev. 2009: DIANE Publishing.
45. Zoellick, K.F. and P. Hrnjak, Oil retention and pressure drop in horizontal and
vertical suction lines with R410A/POE. 2010.
46. Iu, I., et al., Applying the effectiveness-NTU method to elemental heat exchanger
models. ASHRAE Transactions, 2007. 113(1): p. 504-513.
47. Bigi, A.A., L. Cremaschi, and D.E. Fisher, Modeling of Lubricant Effects in a
Microchannel Type Condenser. International Refrigeration and Air Conditioning
Conference, Purdue, IN, 2014.
48. Dell'Orto, S., Experimental analysis and modeling of oil effects on a
microchannel evaporator. Politecnico di Milano, 2014.
49. Bertsch, S.S., E.A. Groll, and S.V. Garimella, A composite heat transfer
correlation for saturated flow boiling in small channels. International Journal of
Heat and Mass Transfer, 2009. 52(7): p. 2110-2118.
50. Kays, W.M. and A.L. London, Compact heat exchangers. McGraw-Hill, New
York, NY, 1984.
51. Chang, Y.-J. and C.-C. Wang, A generalized heat transfer correlation for Iouver
fin geometry. International Journal of heat and mass transfer, 1997. 40(3): p.
533-544.
52. Chang, Y.-J., et al., A generalized friction correlation for louver fin geometry.
International Journal of Heat and Mass Transfer, 2000. 43(12): p. 2237-2243.
53. Harms, T.M., E.A. Groll, and J.E. Braun, Accurate charge inventory modeling
for unitary air conditioners. HVAC&R Research, 2003. 9(1): p. 55-78.
54. Chen, J.C., Correlation for boiling heat transfer to saturated fluids in convective
flow. Industrial & Engineering Chemistry Process Design and Development,
1966. 5(3): p. 322-329.
55. Cooper, M., Using Reduced Properties. Advances in heat transfer, 1984. 16: p.
157.
56. Hausen, H., Darstellung des Warmeuberganges in Rohren durch
verallgemeinerte Potenzbeziehungen. Z. VDI Beih. Verfahrenstech, 1943. 4: p.
91-98.
57. Gnielinski, V., New equations for heat and mass-transfer in turbulent pipe and
channel flow. International chemical engineering, 1976. 16(2): p. 359-368.
58. Bergman, T.L., F.P. Incropera, and A.S. Lavine, Fundamentals of heat and mass
transfer. 2011: John Wiley & Sons.
59. Thome, J.R., Prediction of the mixture effect on boiling in vertical thermosyphon
reboilers. Heat Transfer Engineering, 1989. 10(2): p. 29-38.
117
60. Lockhart, R. and R. Martinelli, Proposed correlation of data for isothermal two-
phase, two-component flow in pipes. Chem. Eng. Prog, 1949. 45(1): p. 39-48.
61. Chisholm, D., A theoretical basis for the Lockhart-Martinelli correlation for
two-phase flow. International Journal of Heat and Mass Transfer, 1967. 10(12):
p. 1767-1778.
62. Ragazzi, F., Modular-based computer simulation of an air-cooled condenser.
1991, Air Conditioning and Refrigeration Center. College of Engineering.
University of Illinois at Urbana-Champaign.
63. Mandrusiak, G. and V. Carey, Pressure drop characteristics of two-phase flow
in a vertical channel with offset strip fins. Experimental Thermal and Fluid
Science, 1988. 1(1): p. 41-50.
64. Paliwoda, A., Generalized method of pressure drop calculation across pipe
components containing two-phase flow of refrigerants. International journal of
refrigeration, 1992. 15(2): p. 119-125.
65. Takaishi, Y. and K. Oguchi. Measurements of vapor pressures of R22/oil
solution. in 18th International Congress of Refrigeration. 1987.
66. Thome, J.R., Engineering data book III. Wolverine Tube Inc, 2004.
67. Liley, P. and W. Gambill, Physical and Chemical Data, in Chemical
Engineering Handbook. 1973, New York: Mc Graw-Hill.
68. Jensen, M. and D. Jackman, Prediction of nucleate pool boiling heat transfer
coefficients of refrigerant-oil mixtures. Journal of heat transfer, 1984. 106(1): p.
184-190.
69. Yokozeki, A., Solubility and viscosity of refrigerant-oil mixtures. International
Refrigeration and Air Conditioning Conference, Purdue, IN, 1994.
70. Filippov, L., Liquid thermal conductivity research at Moscow University.
International Journal of Heat and Mass Transfer, 1968. 11(2): p. 331-345.
71. Lin, L., H. Peng, and G. Ding, Influence of oil concentration on wetting
behavior during evaporation of refrigerant-oil mixture on copper surface.
International Journal of Refrigeration, 2015.
72. Citrini, D. and G. Noseda, Idraulica Casa Editrice Ambrosiana. 1987, Milano.
73. Crompton, J.A., T.A. Newell, and J.C. Chato, Experimental measurement and
modeling of oil holdup. 2004, Air Conditioning and Refrigeration Center.
College of Engineering. University of Illinois at Urbana-Champaign.