Proceedings of International Forum for Green Asia 2019
5th International Exchange and Innovation Conference on Engineering & Sciences (IEICES 2019)

GA05 / P17
An Exergy Analysis of R410a as a ‘Drop-In’ Refrigerant for a Mechanical Vapor
Compression Chiller
Colombatantirige Uthpala Amoda Perera1*, Nobuo Takata1, Takahiko Miyazaki1,2, Yukihiro Higashi2, Kyaw Thu1
1
Interdisciplinary Graduate School of Engineering Sciences (IGSES), Kyushu University, Japan
2
International Institute for Carbon Neutral Energy Research (WPI-I2CNER), Kyushu University, Japan
Email: uthpala.perera@live.com
Abstract: The Heating, Ventilation, Air-Conditioning and Refrigeration (HVAC&R) industry, is contributing to the
Global Warming crisis. This crisis prompted a global initiative such as the Kigali Amendment, of which one key target
is to reduce the use and eventually phase out Hydrofluorocarbons (HFCs), with high Global Warming Potentials(GWPs). This target has prompted the exploration of low GWP refrigerants to replace these HFCs. Some researchhence focuses on using low GWP refrigerants and their mixtures as ‘drop-in’ replacements in existing systems,
allowing easier integration and a more cost effective solution. Hence, a baseline analysis is carried out on a
Mechanical Vapor Compression Chiller, according to AHRI Standards 551/591. R410a is chosen as the 'drop-in'
refrigerant due to its abundance in the global market and its expected phase out due to its high GWP. 1st law analysis
and an exergetic approach is chosen to provide further insight into the irreversibilities.
Keywords: Mechanical Vapor Compression Chiller; R410a; Exergy; Low GWP; HFC; Drop-In Refrigerant
1. INTRODUCTION
The frequency of climatic events affecting humans
across all parts of the world have been on the rise[1].
These events are said to be caused by the increased
influence humans have on the climate which is evident
through various phenomena such as the Global
Warming Crisis[2]. This particular issue is caused by
the increased release of greenhouse gases into out
atmosphere which eventually cause an excessive
greenhouse effect, leading to heightened global
temperatures[3]. Though the beginnings of this crisis ispossible attributed to the industrial revolution, in thepresent day various industries collectively contribute tothis issue. One of which is the Heating, Ventilation, Air
Conditioning and Refrigeration (HVAC&R) industry[4].
The HVAC&R industry contributes to this Global
Warming crisis through both indirect emissions ofCarbon dioxide during power consumption and the
direct emission of greenhouse gases, which are used as
refrigerants. These have collectively resulted in arounda 20% contribution to the global emissions ofgreenhouse gases, which gives the HVAC&R industry avital role to play in solving these issues[5].
Since the HVAC&R industry has a global influence, aninternational effort to move the industry towards asustainable path has been promoted through countlessnational policies and international treaties such as the
Kyoto Protocol[6], Montreal Protocol[7], Kigali
Amendment[8], Paris Accord[9] and Japan and EU
Fluorinated Gas (F-Gas) policies[10], [11]. In order toachieve many of the targets proposed under theseadministrative
procedures, the Research andDevelopment (R&D) sections of the industry have had
to take the lead. This is especially true of the research
work carried out at various research institutions which
provide the fundamental pathways for the industry toprogress. Many of these works have been focused oncontributing to environmental issues caused by theindustry and include work which focuses on replacing
high Global Warming Potential (GWP) refrigerants withalternative low GWP ones such as the work by Jakobs et.
al. which utilized non-azeotropic mixtures and their
specific temperature glides in order to provide better
heat transfer characteristics[12], or the work focused on
the reemergence of natural refrigerants such as Carbon
dioxide[13] by Lorentzen et. al. and even the use ofwater, ammonia and air[14] as proposed by Riffat et. al.
Another method which has been used is termed as‘drop-in’ replacements which allows the integration of
more suitable refrigerants in to pre-existing systems toreplace high GWP refrigerants. This ‘drop-in’
replacement method offers a more cost effective way oftackling the need to replace refrigerants with high
GWPs in existing systems and is hence highly sought
off by both the consumers and producers. Researchpertaining to this area has already been initiated and isbeing carried out of various t various types of systems
ranging from domestic refrigerators[15] in the case ofRangel-Hernandez et. al and even on chillers aspresented by Fukuda et. al., who focused on using
ternary mixtures to replace R410a[16]. However, thelatter work did not focus on a particular international
standard when testing the chiller performance which
creates a gap in research for fair and uniform
comparison with alternative refrigerants and different
chillers.
In order to fill this gap, the present work was proposed
where the performance of a water cooled chiller
operating under AHRI standards 551/591 (SI)[17] wasevaluated. The refrigerant which was chosen for theseexperiments was R410a due to several reasons such as
its high GWP of 2088, the potential phase out of use
under the Kigali Amendment as well as Japan’s F-Gas
policies as well as the extensive use within the current
global market ranging from Japan[10], [18] to USA[19]and to the EU[20]. The following work will present this‘drop-in’ experiments as well as the analysis of datawith respect to both the first law and second law ofthermodynamics. This allows the investigation of theirreversibilities or exergy destruction occurring within-9-the chiller. It is hoped that this analysis can be used as abaseline for comparison with alternative refrigerants
which are to be tested using the same setup as ‘drop-in’
low GWP refrigerants.

conditions for varied cooling loads between 1.0 – 1.6
kW at various charge amounts between 0.67 – 0.80 kg.
The experiments are started at the lowest charge
amounts, where the range of cooling capacities aretested, after which a small amount of refrigerant is
added to the circuit and the process repeated. In order tocharge the refrigerant, the expansion valve must be fully
closed, and the refrigerant is charged into the preevaporator and evaporator section. With the use of theload available from the constant temperature baths, therefrigerant can now be vaporized, which allows theoperation of the compressor. Care must be taken to
ensure that only vaporized refrigerant enters thecompressor, otherwise any liquid present could damage
the compressor unit. The cooling capacities are alsoachieved using the variable speed of the compressor
which allows the operation from full load conditions topart load conditions. LabView software is used for thecollection of data which is then processed using
Mathematica. Extended uncertainties are also calculated
based on Type A uncertainties which deal withstatistical analysis of the raw data and Type B
uncertainties which included instrument andmeasurement uncertainties.

2. METHOD
This section will highlight the experiment procedure aswell as provide a brief outline of the AHRI standard
551/591 (SI).
2.1 Experimental Setup

Fig. 1. Schematic of Vapor Compression Chiller System

2.3 AHRI Standard 551/591 (SI)
Table 1. A summary of the AHRI standard for water cooled
chiller experiments

Flow Rate
Ls-1kW-1

Leaving
Temp. °C

Note
2
1: Rated water flow is determined by the watertemperatures at the rated capacity. Then normalized
flow rate shown, per unit evaporator capacity, is forreference only at Standard Rating Conditions.
2: Rated water flow is determined by the watertemperatures at the rated capacity and rated efficiency.
12.00

Fig. 1 shows the experimental setup, which includes therefrigerant circuit shown in pink, the chilled water
circuit shown in green and the cooling water circuit
shown in red. The chiller itself consists of four main
components. The compressor, evaporator, condenser
and expansion valve. A charging point is located justafter the expansion valve in order to charge different
amounts of refrigerant, whilst a sampling port is locatedin between the condenser and expansion valve, forsampling refrigerant mixtures to test theirconcentrations. For the experimental purposes thecooling load and heating load are provided using two
constant temperature baths, which allow the control ofboth the temperature of water flowing to the heatexchangers as well as the flow rates. There are
Temperature (T) and Pressure (P) sensors located inboth the refrigerant and water circuits as shown in Fig.
1. Together with the different mass or volume flow
meters, the required energy and exergy calculations canbe made using the measured data. For the refrigeration
circuit, the key points of measurement are shown
utilizing a Pressure-Enthalpy (p-h) diagram (See Fig. 2)

7.00

0.0478Tower (Water
Conditions)Entering
Temp. °C

Entering
Temp. °C

Flow Rate
Ls-1kW-1

All CoolingLeaving
Temp. °C

Operating
Category

Fig. 2. Pressure – Enthalpy Diagram for the Chiller Cycle

Evaporator Conditions

1

30.00

35.00

A summary of the AHRI standard 551/591[17] isprovided in Table 1, which dictate the requiredconditions for the chilled and cooling watertemperatures at full load conditions.
3. DATA ANALYSISThe following assumptions are included for the analysis
and discussion section:• Steady state conditions are achieved• The Black-box model is incorporated for eachsubsystem (where only the input and output
for each such subsystem is considered)• The overall circuits remain as closed systems,
but each subsystem or component is
considered as an open system with mass
flow and heat exchange• Kinetic and potential energies are neglected

2.2 Experimental Procedure
For the experimental procedure, it is required tomeasure the afore mentioned properties at steady state- 10 -

kg, and its COP behavior with varied cooling amounts
will be explored. This will be followed with the constantcooling capacity of 1.4 kW and the COP behavior withrespect to changing refrigerant charge amounts.
At 0.70 kg, when the cooling capacity is increased from1.0 – 1.4 kW, an increase in COP is observed. The
reasons for this behavior would be the effects of boththe cooling capacity as well as the compressor power
input as it is visible through Eq. 2. However, both thecooling effect as well as compressor power input rise inthis case, mainly due to an increase in refrigerant andwater mass flow rates, which are required to achieve
large cooling loads. However, the required power input
to the compressor to achieve a unit cooling load is smallenough to produce a net positive effect on the COP,
which would suggest that the heat exchangers arecontinuously performing well. This is possible mainly
due to the availability of free heat transfer area, which
will be occupied with the increased amount ofrefrigerant available at larger cooling loads, which
hence allow a better heat transfer process. However, theamount of area available is limited, which signifies thedrop in COP which is illustrated in Fig. 3 when
progressing above the optimum cooling capacity of 1.4
kW to the larger capacity of 1.6 kW.
Moving on to the scenario in which the optimum
cooling capacity is maintained, Fig. 4 represents thechanges in COP with different R410a charge amounts.

As an initial step to evaluate the nature of the data an
Energy Balance is carried out according to Eq. 1.

EB =QEvap + WComp - QCond
QCond

(1)

The EB values for all scenarios fell below 0.038,
which highlights both the stability of the results and theaccuracy of the measurements. As the initial parameterfor determining the performance of the chiller, theCoefficient of Performance (COP) is determined. The
COP is based on recommended equation provided by
the AHRI standard which is shown in Eq. 2, which
involves the chilled water parameters for calculating thecooling effect at the evaporator QEvap, whilst the direct
power measurements are used to measure the power
consumption by the compressor WComp.

COP =

QEvap
WComp

=

VCH rCH cP ,CH DTEvap
Pin,Comp(2)

The errors for the COP all fall between the range of ±0.057-0.060 which indicates percentage errors ranging
in between 0.99 - 1.09 %.

Fig. 1. The COP versus Cooling Capacity for all Charge
Amounts

Fig. 2. The COP for all charge amounts at constant 1.4 kWof cooling

Based on Fig. 3 there are two key observations which
can be made. Firstly, the COP shows a peak likebehavior at 1.4 kW of cooling. Secondly, the highest
COP is observed for the charge amount of 0.70 kg. The
initial observation of COP peaking, helps determine theoptimum capacity of this particular system which occurs
at around 88% of the full load. Interestingly this peak at
1.4 kW of cooling is observed irrespective of theamount of charge, which reiterates that the chiller has an
optimum capacity of 1.4 kW. Moving onto the second
observation, 0.70 kg shows the highest performance
which occurs at 1.4 kW, whilst it also maintain thehighest COP values together with several other charge
amounts at lower cooling loads, whilst it exhibits thebest COP at full load conditions of 1.6 kW. These
factors help determine that the optimum charge amountof R410a for this system is 0.70 kg. Now that the
optimum cooling capacity of the chiller and the
optimum charge amount of R410a has been determined,further analysis is presented for each of these specificconditions. Firstly; the constant charge amount of 0.70

The optimum charge amount of 0.70 kg shows the best
COP whilst at any charge amount above or below this, adrop in COP is seen. This is possible due toovercharging and undercharging phenomena. In the caseof overcharging, the presence of any additional
refrigerant is met with limited heat exchanger heattransfer area, which inhibits the heat exchange process
and causes subcooled liquid buildup at the end of thecondenser and even slight increases in pressure lift andhence larger power inputs to the compressor. The
opposite, where the system is undercharged, there is alack of refrigerant to produce the required cooling andheat rejection effect, which also causes the drop in COP
as the compressor requires more power input to moverefrigerant through to the condenser.
With the COP analyzed, it is possible to move on to thesecond law analysis where the exergy destruction ĖD is
calculated. The exergy destruction is defined withrespect to the dead state conditions of T0 = 298.15 K
and P0 = 101.35 kPa and represents the potential ofwork which has been lost according to Eq. 3 and Eq. 4.
- 11 -

the phase change occurring at higher cooling loads,
would result in larger entropy changes causing an
increase in the ĖD as well.

n
dV ö
æ T ö
æ
E! D = å ç1 - 0 ÷Q! j + çW! CV + p0 CV ÷ + å mi e fi - å mo e fo
T ø
dt ø iè
j =1 è
o(3)

e fj = (h j - h0 ) - T0 ( s j - s0 )(4)

The total exergy destruction ĖD, is presented in Fig. 5
for 0.67, 0.70 and 0.80 kg, which are lowest, optimum
and largest charge amounts.

Fig. 4. The Component Exergy Destruction for 0.70 kg for
Cooling Capacities of (a) 1.0; (b) 1.2; (c) 1.4; (d) 1.6 kW

Fig. 3. The Total Exergy Destruction for 0.67, 0.70 and 0.80
kg of Refrigerant Charge at 1.0 – 1.6 kW of Cooling

Looking at the optimum capacity of 1.4 kW with varied
charge amounts of smallest, optimum and largest, Fig. 7
presents the ĖD.

Using Fig. 5, it was possible to notice that the total
exergy destroyed increased with larger cooling
capacities. This is attributed to the increasing mass flow
rates which are required for handling the increased heattransfer duty at the evaporator and condenser. The
larger mass flow rates will also cause larger
irreversibilities occurring within each component due
increased friction losses and possible eddy current
generations. These losses combined together, with theincreasing power input to the compressor which directly
results in exergy destroyed, an increasing ĖD trend withlarger cooling capacities is justifiable. Similar to theanalysis of the COP, both the optimum charge amountand optimum cooling capacity are isolated andanalyzed.
Fig. 6 shows the component ĖD at the optimum
charge amount of 0.70 kg, for the 4 cooling capacities
between 1.0 – 1.6 kW. It is noticed that the compressor
creates the largest exergy destruction, which is thenfollowed by the 2 condenser and evaporator. The
expansion valve seems to exhibit the smallest
contribution in almost all scenarios. The possible
reasons for the compressor being the major contributor
is due to the increasing power consumption created via
the larger pressure lift required for cycle operation,whilst the two heat exchangers would be affected by theincreased mass flow rates which again contribute to thelarger losses within these components, and since thecondenser is required to handle a larger heat load incomparison to the evaporator, it is also expected tocontribute more towards the ĖD. The expansion process
which is considered to be isenthalpic, would only resultin ĖD due to entropy changes when moving from thesubcooled region at the end of the condenser to the 2phase region at the evaporator inlet. Since there is aslight increase in condenser pressure with cooling loads,

Fig. 5. The Component Exergy Destruction at 1.4 kW for theCharge Amounts of (a) 0.67; (b) 0.70; (c) 0.80 kg

Fig. 7 shows the component ĖD for 0.67, 0.70 and 0.80
kg R410a charge amounts at 1.4 kW of cooling
capacity. In this scenario, the trends are slightly
different from the previous case. The compressor shows
an increasing ĖD similar to the section above, which
could be a result of liquid refrigerant accumulation at
the exit of the condenser, causing a hindrance in theheat transfer process. This results in the slight increasein condenser pressure as well. Following suit, it would
be expected for the compressor ĖD to increase as well
due a positive trend in electrical power consumption.
However, this is not visibly clear possibly due to an
influence in the lower mass flow rates through thecompressor system which create an unclear pattern inthe ĖD of the compressor. The evaporator on the other
hand shows a decreasing contribution towards the ĖD at- 12 -

larger cooling loads, which is an accumulated effect of
the reduced mass flow rates at the expansion valvecausing less irreversibilities, coupled with a slight dropin evaporator temperature, moving the component
temperature slightly away from the dead state. The
expansion valve also follows a negative trend in ĖD aslarger cooling loads are approached, which help justify
the lower irreversibility occurrences with lower mass
flow rates, since this particular process is isenthalpic
and dependent only on the refrigerant mass flowing
across combined with its entropy change.
Finally, the exergy efficiency is also looked at andanalyzed using Fig. 8. Eq. 5 provides the method of
calculating the efficiency which allows the comparison
to an ideal Carnot cycle.

be 1.4 kW. The energy balance was maintained at below
0.04 for all test conditions which signify the steady state
conditions and accuracy of measurements. The optimum
charge amount suggested that undercharging andovercharging the system negatively affects the COP,
whilst the exergy destruction analysis provided thefollowing findings; the compressor produced the mostexergy destruction and was followed by the condenser,
evaporator and finally the expansion valve; the mass
flow rates of refrigerant and water affected the exergy
destructions as well. The exergy efficiency reiterated thefindings on optimum cooling capacity and optimum
charge amount due to its similar behavior to the COP.
The presented analysis can be utilized for further
investigation of potential low GWP alternative ‘drop-in’
refrigerants on this MVC Chiller to provide a baseline
for comparison purposes.

æT ö
QEvap ç1 - 0 ÷
ç T ÷
æ T
ö
Evap ø
è= COPç 0 - 1÷h Ex =
çT÷h Elec ´ WComp
è Evap ø

5. ACKNOWLEDGEMENTS
The author would like to express their gratitude to Mr.
Takezato and Mr. Senba for their contributions throughexperiments and is also grateful to the Green Asia
Program for their funding.

(5)

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Fig. 6 The Exergy Efficiency versus Cooling Capacity for
All Charge Amounts

The trends in ηEx with an average error of ±1.16%, show
a similar behavior to that of COP, where the peakoccurs at 1.4 kW irrespective of the R410a charge
amount. It is possible to analyze this behavior for 0.67,
0.70 and 0.80 kg for the cooling capacity ranges of 1.0 1.4 kW and lastly from 1.4 - 1.6 kW. ηEx efficiency
shows a positive trend across all charge amounts up
until the optimum cooling capacity of 1.4 kW. The main
parameters affecting this increase are the COP and TEvap
as seen in Eq. 5. As it was shown before, the COP
increased as a result of increased mass flow rates
occupying the readily available heat transfer areas,
whilst the Evaporator temperatures also showed a slightdecrease with cooling capacity. These two factors
contributed to the increase in ηEx and the optimum
charge amount of 0.70 kg showed the highest exergy
efficiency at the optimum load of 1.4 kW. The ηEx thendrops as it reaches 1.6 kW mainly due to the drop in
COP, which dominates the reduction in TEvap.
4. CONCLUSION
The ‘drop-in’ investigations of R410a on a Mechanical
Vapor Compression (MVC) Chiller were carried out
under AHRI Standard 551/591 (SI) and analyzed
utilizing the first and second laws of thermodynamics.
Through the experimental data analysis, the optimum
charge amount was determined to be 0.70 kg whilst the
optimum cooling capacity for the system was found to- 13 -

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[12]

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