Effect of Lubricant on R410A Horizontal

Flow Boiling

Nitin N. Sawant

Mark A. Kedzierski

J. Steven Brown

NISTIR 7456
2
3

NISTIR 7456

Effect of Lubricant on R410A Horizontal

Flow Boiling

Nitin N. Sawant

Mark A. Kedzierski

J. Steven Brown

U.S DEPARTMENT OF COMMERCE

National Institute of Standard and Technology

Building Environment Division

Building and Fire Research Laboratory

Gaithersburg, MD 20899-8631

September 2007

U.S. Department of Commerce

Carlos M. Gutierrez, Secretary

National Institute of Standards and Technology

James M. Turner, Acting Director
4

Effect of Lubricant on R410A Horizontal Flow Boiling 

N. N. Sawant 

The Catholic University of America 

Washington, DC 20064 

M. A. Kedzierski 

National Institute of Standards and Technology 

Gaithersburg, MD 20899 

J. S. Brown 

The Catholic University of America 

Washington, DC 20064 

ABSTRACT 

This paper experimentally determines the effect of lubricant on R410A horizontal flow 

boiling. A baseline was established by measuring the convective boiling heat transfer 

performance of pure R410A. In addition, the heat transfer performance of an 

R410A/lubricant (99.6/0.4) mixture was measured and compared to that of the lubricantfree refrigerant. The measured convective boiling Nusselt numbers for the lubricant-free 

R410A and R410A/lubricant mixture were compared to an existing correlation. The 

correlation was shown to predict 71 % and 83 % of the lubricant-free R410A and the 

R410A/lubricant Nusselt numbers to within ± 20 %, respectively. The mean degradation 

in the heat transfer performance with the addition of lubricant was approximately 6.7 %. 

A measurement technique for observing the fluorescence of the lubricant in the annular 

flow was developed. Further improvements in the technique are necessary before it can 

be used to visualize the distribution of lubricant within the flow. 

Keywords: Flow boiling, micro-fin, R410A, refrigerant/lubricant mixture 
5

TABLE OF CONTENTS 

ABSTRACT........................................................................................................................ 4 

TABLE OF CONTENTS.................................................................................................... 5 

LIST OF FIGURES ............................................................................................................ 5 

LIST OF TABLES.............................................................................................................. 5 

INTRODUCTION .............................................................................................................. 6 

EXPERIMENTAL APPARATUS...................................................................................... 7 

MEASUREMENTS............................................................................................................ 8 

Heat Transfer Coefficient ............................................................................................. 10 

RESULTS ......................................................................................................................... 12 

Comparison With Correlation....................................................................................... 12 

LUBRICANT FLUORESCENCE.................................................................................... 14 

Spectrofluorometer ....................................................................................................... 14 

Fluorescence Setup ....................................................................................................... 15 

CONCLUSIONS............................................................................................................... 16 

ACKNOWLEDGEMENTS.............................................................................................. 16 

NOMENCLATURE ......................................................................................................... 18 

Subscripts...................................................................................................................... 19 

REFERENCES ................................................................................................................. 20 

LIST OF FIGURES 

Fig. 1 Schematic of test rig .............................................................................................. 24 

Fig. 2 Cross section of test section................................................................................... 25 

Fig. 3 Test section detail ................................................................................................. 26

Fig. 4 Comparison of measurements with REFPROP 7.1 (2006) ................................... 27 

Fig. 5 Difference between oil-free R410A and R410A/POE saturated liquid equilibrium 

equations ................................................................................................................... 28 

Fig. 6 Heat flux distribution for pure R410A and R410A/POE mixture......................... 29 

Fig. 7 Comparison of the heat transfer coefficient for pure R410A and R410A/POE 

mixture ...................................................................................................................... 30

Fig. 8 Comparison of counterflow data with existing correlation .................................... 31 

Fig. 9 Apparent heat transfer degradation due to oil ........................................................ 32 

Fig. 11 Fluorescence spectra of R410A/POE mixture...................................................... 35 

Fig. 12 Mercury Spectra (Excitation = 300 nm)............................................................... 36 

Fig. 13 Lubricant concentration observation test section ................................................ 37 

LIST OF TABLES 

Table 1 Median estimated 95 % relative expanded uncertainties for measurements...22 

Table 2 Calibration check of spectrofluorometer against Mercury lamp.....................23 
6

INTRODUCTION 

Many new evaporators and condensers in the residential air-conditioning industry employ 

enhanced micro-fin tubes rather than smooth tubes. In fact, the micro-fin tube dominates 

unitary equipment design because it provides the highest heat transfer with the lowest 

pressure drop among the commercially available internal enhancements (Webb, 1994). 

Because of the recent changeover to the production of environmentally friendlier 

refrigerants without ozone depleting chlorine, R410A has received significant interest 

from researchers. For example, Kim and Shin (2005) compared fluid-heated flow boiling 

heat transfer performance of R22 to that of R410A in smooth and micro-fin tubes. They 

investigated the heat transfer coefficient for the micro-fin tubes as a function of 

refrigerant quality. When compared to R22 at the same test conditions, the evaporating 

heat transfer coefficients for R410A were 97 % to 129 % of R22. 

Because most compressors are lubricated, the effect of lubricant on the boiling heat 

transfer coefficient of refrigerants has been extensively studied. For example, Cremaschi 

et al. (2005) conducted measurements on oil retention characteristics in the condenser, 

evaporator, liquid and suction lines of air-conditioning systems employing refrigerants 

R22, R134a, and R410A along with miscible and non-miscible lubricants. The oil 

retention was shown to depend on several parameters such as oil mass fraction, vapor 

refrigerant mass flux, mixture viscosity ratio, and orientation of the pipe. At an oil mass 

fraction of 5 %, the poorly miscible and soluble R410A/Mineral Oil combination retained 

oil in the suction line by approximately 31 % more by mass than that of the miscible 

R410A/Polyolester Oil combination. In addition, Lottin et al. (2003) used a theoretical 

model to explain the effect synthetic oil would have on the performance of a R410A 

compression refrigeration system. In the evaporator, optimum performance was observed 

with a 0.1 % oil mass fraction. The motivation for this study is the fact that there is no 

significant documentation of the effect of lubricant on the convective boiling heat transfer 

of R410A. 

More measurements of local evaporation heat transfer are needed by the refrigeration 

industry for accurate sizing of micro-fin tube evaporators. One of the objectives of this 

study, therefore, is to provide a source of information for the boiling of R410A with and 

without the addition of a polyolester (POE) lubricant. A second purpose is to visually 

investigate horizontal nucleate flow boiling characteristics of R410A with a high-speed 

camera. A further motivation for this study is to describe the development of a 

fluorescence measurement technique to explore the influence of heat flux on the lubricant 

surface density. A mixture of napthenic mineral oil and POE lubricant (1 % and 99 % by 

mass, respectively) was chosen for its favorable fluorescence characteristics and to 

demonstrate a new measurement technique. The POE employed in this study had a 

nominal kinematic viscosity of 68 μm

2

/s at 297.8 K. 
7

EXPERIMENTAL APPARATUS 

The evaporative heat transfer of R410A was studied in a concentric tube-in-tube heat 

transfer test apparatus. A schematic of the test apparatus is shown in Fig. 1. The test 

setup consisted of an evaporator (test section), condenser, subcooler, preheater, and 

magnetic gear pump. A magnetically coupled gear pump delivered R410A to the 

entrance of the test section in a subcooled state. Another magnetically coupled gear 

pump supplied a steady flow of distilled water to the annulus of the test section. 

The refrigerant flow rate and pressure were fixed at the inlet to the test section. 

Redundant refrigerant-side flow rate measurements were made with a Coriolis flowmeter 

and a turbine flowmeter. The inlet water temperature was held constant for each test with 

a water-chilled heat exchanger and variable electric heaters. The fixed water flow rate 

and the inlet water temperature were used to establish the overall refrigerant quality 

change in the test section. The water temperature change, the tube wall temperature, the 

refrigerant temperatures, pressures, and pressure drops were measured at several axial 

locations along the test section. These measurements were used to calculate the local 

heat-transfer coefficient for the micro-fin tube. 

The test section consisted of a pair of 3.34 m long, horizontal tubes connected by a Ubend. A fixed test pressure was maintained by balancing the refrigerant duty between the 

subcooler, the test section, and the evaporator. Figure 2 shows a cross section of the test 

section with a detail of the micro-fin tube geometry. In the test section, the refrigerant 

flowed inside a micro-fin tube, while distilled water flowed either in counterflow or in 

parallel flow to the refrigerant in the annulus that surrounded the micro-fin tube. The 

annulus gap was 2.2 mm, and the micro-fin tube wall thickness was 0.3 mm. The microfin tube has 60, 0.2 mm high fins with an 18o helix angle. For this geometry, the cross 

sectional flow area was 60.8 mm2

 giving an equivalent smooth diameter (De) of 8.8 mm. 

The root diameter of the micro-fin tube was 8.91 mm. The inside-surface area per unit 

length of the tube was estimated to be 44.6 mm. The hydraulic diameter (Dh) of the 

micro-fin tube was estimated to be 5.45 mm. The ratio of the inner surface area of the 

micro-fin tube to the surface area of a smooth tube of the same De was 1.6. The fins 

rifled down the axis of the tube at a helix angle of 18o with respect to the tube axis. 

Figure 3 provides a detailed description of the test section. The annulus was constructed 

by connecting a series of tubes with 14 pairs of stainless steel flanges. This construction 

permitted the measurement of both the outer micro-fin wall temperature and the water 

temperature change as discussed in the following two paragraphs. The design also 

avoided abrupt discontinuities such as unheated portions of the test section and tube-wall 

“fins” between thermopile ends. 

Figure 3 shows that thermocouple wires (labeled LT) pass between 12 of the gasketed 

flange pairs to measure the refrigerant-tube wall temperature at ten locations on the top, 

side, and bottom of the tube wall. These locations were separated by 0.6 m on average, 

and they were located near the intersection of the shell flanges. In addition to these, 

thermocouples were also mounted next to the pressure taps near the middle of each test 

section length. The thermocouple junction was soldered to the outside surface and was 
8

sanded to a thickness of 0.5 mm. The leads were strapped to a thin non-electricallyconducting epoxy layer on the wall for a distance of 14.3 mm before they passed between 

a pair of the shell flanges. The wall temperature was corrected for a heat-flux dependent 

fin effect. The correction was typically 0.05 K. Figure 3 also shows that a chain of 

thermopiles was used to measure the water temperature change between each flange 

location. Each thermopile consisted of 10 thermocouples in series, with ten junctions at 

each end evenly spaced around the circumference of the annulus. Because the upstream 

junctions of one thermopile and the downstream junctions of another tend to enter the 

annulus at the same axial location (except at the water inlet and outlet), the junctions of 

the adjacent piles were alternated around the circumference. A series of Teflon1

 half 

rings attached to the inner refrigerant tube centered the tube in the annulus. The halfrings were circumferentially baffled to mix the water flow. Mixing was further ensured 

by a high water Reynolds number (Kattan et al., 1995). 

As shown in Fig. 3, six refrigerant pressure taps (labeled P) along the test section allowed 

the measurement of the upstream absolute pressure and five pressure drops along the test 

section. Two sets of two water pressure taps were used to measure the water pressure 

drop along each tube. Also, a sheathed thermocouple measured the refrigerant 

temperature at each end of the two refrigerant tubes, with the junction of each centered 

radially. Only the thermocouple at the inlet of the first tube was used in the calculations. 

The entire test section was wrapped with 5 cm of foam insulation to minimize heat 

transfer between the water and the ambient. 

MEASUREMENTS 

The system was charged with a commercially available R410A. Its composition was 

maintained by thorough evacuation of the test rig before charging and only charging 

liquid from the commercial cylinder. 

Flow boiling heat transfer measurements were obtained for three different average 

refrigerant pressures of 1.034 MPa, 1.068 MPa, and 1.103 MPa. The refrigerant mass 

flow rate was varied to achieve mass fluxes in the range from 100 kg/m2

s to 400 kg/m2s. 

The average refrigerant saturation temperature, Tr, was varied between 7.2 oC and 10 oC 

with approximately 1 K subcooling at the test section inlet. 

A data acquisition system and a tailored software program were used to record all the 

conditions in a data file. All saturated refrigerant properties, with the exception of the 

equilibrium mixture temperature, were evaluated at the measured saturation pressure 

using version 7.1 of REFPROP (Lemmon et al., 2006) mixture property routines. 

 

1

 Certain trade names and company products are mentioned in the text or identified in an illustration in 

order to adequately specify the experimental procedure and equipment used. In no case does such an 

identification imply recommendation or endorsement by the National Institute of Standards and 

Technology, nor does it imply that the products are necessarily the best available for the purpose. 
9

Equilibrium Refrigerant Temperature

Considering that the test fluid was a commercially prepared mixture, the pressure and the 

temperature characteristics could vary with composition and purity. Depending on the 

manufacturing process, stray components such as R115 may be present in the mixture 

causing a deviation from the "standard" or desired mixture properties of R410A. For 

example, the measured saturation temperature of the R410A test fluid used in this study 

differed from that obtained from REFPROP 7.1 by approximately 1 K for the pressure 

range tested. To alleviate this discrepancy, vapor equilibrium measurements were made 

for the test fluid of this study for temperatures between 280 K and 300 K and pressures 

between 1.0 MPa and 1.2 MPa. 

A precisely controlled temperature liquid bath was used with a glass standard platinum 

resistance thermometer calibrated to within ± 0.005 K. A constant volume vessel 

instrumented with a pressure transducer with an uncertainty of ± 0.01 kPa was charged 

with the test fluid and fully immersed in the bath. The mass quality of the charge was 

calculated from the known volume and the charged mass. The densities of the vapor and 

the liquid were obtained from version 7.1 of REFPROP (Lemmon et al., 2006) at the 

measured temperature. The saturation temperature (Tr), refrigerant pressure (Pr) and 

thermodynamic mass quality (xq) were correlated for R410A without lubricant as: 

01 r 2 ( ) r

1 ln A A P Axq T

=+ + (1) 

where Tr is in K and Pr is given in kPa. The constants are A0 = 0.658452 x 10-2 K-1

; A1 = -

0.434741 x 10-3 K-1

; A2 = -0.129204 x 10-5 K-1 with the temperature residuals being 

between –0.005 K and 0.01 K. 

Figure 4 shows that eq. (1) represents the measured temperature to within ± 0.01 K for 

pressures between 1 MPa and 1.2 MPa. Equation 1, the locally measured pressure, and 

the calculated thermodynamic quality were used to calculate the saturated temperature for 

all of the measurements reported in this study. 

The constants in eq. (1) were modified to predict the saturated conditions of 

R410A\lubricant mixtures following the procedure given by Thome (1995): 

( ) 2

r

1

r

ln

1 q

A Pb x

A

T a

− +

= (2) 

The a and b are fourth degree polynomials in the local lubricant mass fraction in the 

refrigerant liquid (wl): 

234

ll l l

234

ll l l

2300.2K +182.5K 724.2K 3868.0K 5268.9K

15.146 - 0.722 2.391 13.779 17.066

a ww w w

b ww w w

=− − + −

= +− +

 (3) 
10

where the local lubricant mass fraction was obtained from the quality and the all-liquid 

(xq = 0) lubricant mass fraction (wb) as: 

l

b

1

1

1 q

w

x

w

= − −

 (4) 

All of the coefficients of the a and b polynomials with the exception of the constant terms 

were taken from Thome (1995). The constant terms of the polynomials were adjusted to 

reproduce the oil-free R410A expression given in eq. (1) when wl = 0. 

Figure 5 illustrates the magnitude of the effect that 0.4 % mass fraction of lubricant has 

on the saturation temperature of R410A for a saturation pressure of 1100 kPa and various 

qualities. In general, the lubricant increases the saturation temperature by less than 

0.025 K for qualities less than 0.5. 

Heat Transfer Coefficient

Table 1 shows the expanded measurement uncertainty (U) of the various measurements 

along with the range of each parameter in this study. The U was estimated with the law 

of propagation of uncertainty. All expanded measurement uncertainties are reported at 

the 95 % confidence level. The estimates shown in Table 1 are median values of U for 

the correlated data. 

The convective boiling heat transfer coefficient based on the actual inner surface area 

(h2Φ) was calculated as: 

2

" Φ =

w r −

q h

T T (5) 

where the measured wall temperatures (Tw) were fitted to their axial positions to reduce 

the uncertainty in the measurement. The measured wall temperatures were fitted to: 

0

2 T = + z A z w 2 A A1 + (6) 

The water temperature (Tf) was determined from the measured temperature change 

obtained from each thermopile and the inlet water temperature measurement. The water 

temperature was regressed to the axial location of the thermopiles along the z-coordinate 

(see Fig. 3). The water temperatures were fitted to: 

2 3 T B Bz Bz Bz f 01 2 3 =+ + + (7) 
11

The water temperature fits, the measured water mass flow rate ( f m ), and the properties of 

the water were used to calculate the local heat flux (q”) to the micro-fin tube based on the 

actual inner surface area: 

d d

d d f

f f f

p f

mTP

qc v

p z z

⎛ ⎞ ′′ = + ⎜ ⎟ ⎝ ⎠

 (8) 

where p is the wetted perimeter of the inside of the micro-fin tube. The specific heat (cpf) 

and the specific volume (vf) of the water were calculated locally as a function of the water 

temperature. The local axial water temperature gradient (dTf/dz) was calculated from the 

derivative of eq. (7). The water pressure gradient (dPf/dz) was linearly interpolated 

between the pressure taps to the location of the wall thermocouples. The pressure 

gradient term was typically less than 3 % of the temperature gradient term. 

Figure 6 shows an example of the local heat flux for lubricant-free R410A and the 

R410/lubricant mixture as calculated from eq. (8) versus thermodynamic quality for 

counterflow and a fixed exit test section pressure of 1.1 MPa. For a given quality, the 

heat flux for the lubricant-free R410A is roughly 3 kW/m2

 larger than that of the 

R410A/POE (99.6/0.4) mixture, but has approximately the same nearly constant rate of 

increase with respect to quality. 

The enthalpy of the refrigerant at the inlet of the test section was calculated from its 

equilibrium refrigerant temperature and measured pressure. The subsequent change in 

the refrigerant enthalpy along the test section was calculated from the local heat flux and 

the measured refrigerant mass flow rate. The refrigerant pressures were measured at six 

pressure taps along the test section. The pressure was linearly interpolated between the 

taps. The test refrigerant entered the test section nearly as saturated liquid having 

approximately 1 K of subcooling. The average saturation temperature, Tr, at the inlet 

varied between 7.2 o

C and 10 oC. 

The local Nusselt number (Nu) was calculated using the hydraulic diameter (Dh) and the 

heat transfer coefficient based on the actual inner surface area of the tube (h2Φ) as: 

 2 h

l

Nu

h D

k

φ = (9) 

where kl is the thermal conductivity of the liquid. 

The hydraulic diameter was measured with a polar planimeter from a scaled drawing of 

the tube cross section, but it can be approximated for other tube geometries with fin 

parameters by expanding on the expression that was given for Dh in Kedzierski and 

Goncalves (1999): 
12

 ( )

( )

2

r fb c

h

f

4 cos 2 cos

2

cos / 2

α π α

β

− = =

+

A D Nte

D

N S e

s

 (10) 

Figure 2 shows the fin parameters that are used in eq. (10) where S is the perimeter of one 

fin and channel taken perpendicular to the axis of the fin, s is the spacing between the 

fins, β is the fin-tip angle, e is the fin height, α is the twist angle of the fins, tb is the 

thickness of the fin at its base, Nf is the total number of fins, and Dr is the diameter of the 

tube at the fin root, i.e., fin base. The hydraulic diameter of the present tube geometry of 

this study as estimated from eq. (10) is 5.2 mm, while that obtained from the planimeter 

and used in the data reduction was 5.45 mm. 

The internal surface area of the fin per unit length (Ai/L) can be estimated from: 

 ( )

i

f

2

cos cos cos / 2 α αβ

⎛ ⎞

= + ⎜ ⎟

⎝ ⎠

As e N

L (11) 

The Ai/L estimated from eq. (11) is 46.8 mm, while that obtained from the planimeter and 

used in the data reduction was 44.6 mm. 

Figure 7 shows a comparison of the R410A heat transfer coefficient versus quality to that 

for the R410A/POE (99.6/0.4) mixture for a mass flux of approximately 200 kg/m2

. In 

general, the measured boiling heat transfer coefficients are nearly constant with respect to 

quality over the range tested. Thinning of the liquid film on the wall for larger qualities 

is expected to induce an increase in the heat transfer coefficient. It is speculated that the 

actual heat transfer increase is not evident because the increase is within the uncertainty 

of the heat transfer measurement. On average, the refrigerant/lubricant mixture heat 

transfer coefficient is approximately 300 Wm-2K-1 less than the pure R410A heat transfer 

coefficient, which represents approximately a 6 % degradation in the heat transfer due to 

the lubricant for these particular conditions. 

DISCUSSION 

Comparison With Correlation

Hamilton et al. (2005) correlated micro-fin-tube-convective boiling Nusselt numbers for 

four pure refrigerants: R22, R32, R125, and R134a; and four refrigerant mixtures: R410B 

(R32/R125, 45 %/55 % mass), R32/R134a (27 %/73 % and 30 %/70 % mass) and R407C 

(R32/R125/R134a, 25 %/23 %/52 % mass) to a single expression consisting of a product 

of dimensionless properties: 2

1 4 3 56

C

0.3 C C r r C CC

10 r

c c

Nu 482.18Re Pr Bo ( log ) 1.1 P P M

P P

⎛ ⎞ = − ⎜ ⎟ ⎝ ⎠

 (12) 

 where 

C1 = 0.51 xq

C2 = 5.57 xq – 5.21 xq

2
13

C3 = 0.54 – 1.56 xq + 1.42 xq

2

C4 = -0.81 + 12.56 xq – 11.00 xq

2

C5 = 0.25 – 0.035 xq

2

C6 = 1/Ts ((TLV – TMV){279.8(xv - xl) – 4298(Td – Tb)/Ts}) 

The limits of applicability for the correlation are: 

70 kg/m2

s < Gr < 370 kg/m2s and 0 < xq < 0.7 

The correlation was shown by Hamilton et al. (2005) to predict some of the existing data 

from the literature to within ± 20 %. Cooper (1984) suggested that the fluid properties 

that govern nucleate pool boiling can be well represented by a product of the reduced 

pressure (Pr/Pc), the acentric factor (-log10(Pr/Pc)), and other dimensionless variables to 

various powers. The all-liquid Reynolds number (Re), the Boiling number (Bo), the 

liquid Prandtl number (Pr), the reduced pressure (Pr/Pc), and the quality (xq) were all 

evaluated locally at the saturation temperature. Mr is the molar mass of the refrigerant in 

g/mol. 

The all-liquid Reynolds number and the Nusselt number are based on the hydraulic 

diameter. Also, the Nusselt number is based on the actual inner surface area of the tube. 

The Td and Tb are the dew point temperature and the bubble point temperature of the 

mixture, respectively, evaluated at the local saturation pressure and overall composition. 

The TLV and the TMV are the saturation temperatures of the least volatile (pure) 

component and the most volatile component, respectively, evaluated at the saturation 

pressure of the mixture. The mass fraction of the vapor (xv) and that of the liquid (xl) are 

evaluated at the saturation pressure and the local thermodynamic quality, while the 

overall composition is the all-liquid or all-vapor value. The constant C6 is zero for pure 

refrigerants and was approximately 0.15 for much of the R410A data. 

Figure 8 provides a comparison of the Nusselt numbers predicted with eq. (12) for the 

micro-fin tube to those measured in this study for lubricant-free R410A and the 

R410A/POE mixture. Approximately 71 % of the measured, lubricant-free, data fall 

within ± 20 % of the predicted data. The lubricant-free measurements that were outside 

the ± 20 % region of the predicted data were for conditions that were very close to those 

of the measurements that were well predicted. No justifiable reason was found to 

identify these points as outliers. Figure 8 also shows that eq. (12) predicts 83 % of the 

R410A/POE Nusselt numbers to within ± 20 %. The R410A/POE predictions were made 

using the R410A/POE mixture liquid viscosity and liquid density evaluated at the local 

lubricant mass fraction to calculate the Re and Pr in eq. (12). Use of the lubricant 

properties in the correlation caused between a 0.3 % and a 0.8 % reduction in the Nusselt 

number as compared to results using the lubricant-free R410A properties alone. 

Figure 9 shows the heat transfer degradation due to the lubricant (ΔNud) as a function of the 

thermodynamic quality for all of the measured R410A/POE data. The ΔNud was calculated 

from the following definition: 

 
14

ΔNu 100(Nu - Nu ) / Nu d pm p = (10) 

where Nup is the Nusselt number obtained from the lubricant-free correlation (eq. (12)) 

using the lubricant-free R410A properties, while Num is the Nusselt number obtained from 

the measured refrigerant/lubricant mixture heat transfer data. The figure shows that the 

degradation in heat transfer due to the addition of 0.4 % mass fraction is between –20 % and 

+42 % exhibiting no apparent relationship with respect to quality. The average degradation 

due to the presence of the lubricant was approximately 6.7 %, which is within the 

uncertainty of both the measurements and the predictions. 

LUBRICANT FLUORESCENCE 

Several measurement techniques exist to measure the bulk lubricant concentration in a 

refrigerant (Navarro de Andrade et al., 1999). These techniques rely on the measurement 

of various properties of the lubricant: density, light absorption, viscosity, and speed of 

sound. For example, Navarro de Andrade et al. (1999), Meyer and Saiz Jabardo (1994), 

and Baustian et al. (1988a) used temperature and speed of sound measurements of a 

liquid refrigerant/lubricant mixture to determine the bulk concentration of the mixture. 

Baustian et al. (1988b) used a bypass viscometer to correlate the mixture viscosity to the 

flowing lubricant concentration. Suzuki et al. (1993) used the light absorption properties 

of lubricant to quantify the fraction of lubricant in the refrigerant. Finally, Baustian et al. 

(1988c) and Bayani et al. (1995) demonstrated the use of a flowmeter densimeter to 

measure the concentration of lubricant in the liquid line of a refrigeration system. All of 

these methods are more suited for the measurement of the bulk lubricant concentration of 

a non-boiling fluid rather than of the localized lubricant distribution in a two-phase 

annular flow, which cannot be measured using a global measurement technique. 

With the purpose of confirming the existence of the lubricant excess layer on a boiling 

heat transfer surface, Kedzierski et al. (1998) proposed an in-situ measurement technique 

that relied on the fluorescent properties of the lubricant. The proposed method made use 

of the relative fluorescent properties of the lubricant and the refrigerant. Ideally, the 

lubricant should strongly fluoresce while the refrigerant should not fluoresce. In this 

way, the amount of lubricant in the mixture can be calculated from the intensity of the 

fluorescence emission. Kedzierski et al. (1998) demonstrated that the fluorescence 

intensity of a lubricant/methanol film on an adiabatic surface could be measured and that 

the intensity of the fluorescence increased with lubricant concentration. The present 

study has developed a method for exciting the lubricant in the annular flow so that its 

location can be documented and studied in order to contribute data for the modeling of 

refrigerant/lubricant flow boiling. 

 

Spectrofluorometer

Figure 10 shows a simplified schematic of the right angle spectrofluorometer consisting 

of a xenon light source, an excitation and an emission monochromator, and an emission 

photomultiplier tube (detector). The light source was focused into the excitation 

monochromator by a collimating lens. The monochromator was set to emit light at a 

wavelength of 320 nm. Because the intensity of the xenon lamp varies with wavelength, 

a corrected excitation module was used to compensate for the variation. The light from 
15

the excitation monochromator passed through a 2.5 nm slit before it entered the sample 

chamber. All of the slit widths were 2.5 nm to limit the bandwidth of the wavelength. 

The spectrofluorometer was designed to accept 45 mm x 10 mm x 10 mm fluorescent 

samples or cuvettes filled with fluorescent material. The emission monochromator was 

used to select the optimum wavelength to measure the fluorescent intensity. The 

intensity of the emission was measured with a photomultiplier tube and accompanying 

electronics. 

Preliminary tests with the pure POE lubricant in a cuvette showed that the fluorescence 

intensity of the lubricant was not sufficient for measurement in the test rig. 

Consequently, the POE lubricant was mixed with a mineral oil (MO) that was known to 

have a strong fluorescent intensity. Several mixtures of the MO and the POE (MO/POE 

(20 %/80 % by mass), MO/POE (10 %/90 % by mass), and MO/POE (1 %/99 % by 

mass)) were attempted in order to check for sufficient fluorescence intensity. All of the 

mixtures fluoresced well. Consequently, the mixture with the least amount of MO was 

selected for this study. With the addition of 1 % mass MO, the intensity of the mixture 

(MO/POE (1 %/99 % by mass)) was more than ten times larger than that of the pure 

POE. 

Figure 11 shows the analysis of the emission and excitation spectra of MO/POE 

(1 %/99 % by mass) in a cuvette. The test sample was placed directly in the sample 

chamber of the right angle spectrofluorometer. The excitation wavelength that produced 

the maximum fluorescence emission was iteratively found by scanning through both the 

excitation and the emission wavelengths. The excitation and the emission wavelengths 

for the lubricant that produced the largest intensities were located at 320 nm and 380 nm, 

respectively. 

The emission and excitation wavelength measurements were verified with a mercury 

standard and a “crossover-peak” from the excitation. The emission wavelength 

measurement obtained from the spectrofluorometer was checked against a mercury vapor 

light. Figure 12 and Table 2 show a comparison of the published values of the peak 

wavelengths for mercury (Reader et al., 1980) to those obtained from the 

spectrofluorometer. The absolute difference between the measured and published 

wavelengths was approximately ± 2.5 nm, which was the width of the slits in the 

spectrofluorometer. 

The excitation wavelength measurement obtained from the spectrofluorometer was 

checked with a “crossover peak” from the excitation. In other words, the excitation 

monochromator was set to a specific wavelength with no specimen in the sample 

chamber. Under these conditions, the emission intensity should peak at the excitation 

wavelength. The wavelength of the emission peaked at the excitation wavelength to 

within the resolution of the digital display (± 1 nm) for the wavelengths that were tested. 

Fluorescence Setup

Figure 13 shows a schematic of the apparatus that was used to record the presence of the 

lubricant in the refrigerant/lubricant mixture as it passed through a quartz tube at the exit 
16

of the two-phase flow heat transfer test section. An aluminum chamber was designed to 

enclose the quart tube, hold the excitation source, and support the digital camera. The 

camera was orientated perpendicular to the axis of the incident light to minimize 

interference from the excitation source. The enclosure was fitted at the bottom and at the 

front side to accommodate the 380 nm and the 320 nm interference filters, which 

correspond to the optimal excitation and emission intensities as shown in Fig. 13. The 

inside of the box was painted black to reduce reflection from the inner surfaces. The base 

of the box was designed to ensure the light source, a long-wave ultra violet lamp, would 

fit into the casing. The enclosure was carefully fitted over the quartz tube to prevent light 

leakage from the enclosure to the room. 

Observations were made of both the fluorescent and the excitation light incident to an 

annular flow of R410A/POE in the quartz tube at a wavelength of approximately 380 nm. 

By design, only the fluorescent light induced by the POE should pass to the camera. In 

order to have enough light to excite the camera, observations were made with the 

emission filter removed from the box. As a result, it was impossible to distinguish 

between the fluorescent and the excitation light. Future improvements to the design of 

the setup would possibly include a greater wattage excitation lamp and/or wider bandpass 

interference filters to permit more light to the camera, and/or a lens to focus the light 

source into the test tube. 

CONCLUSIONS 

Local convective boiling measurements for lubricant-free R410A and a R410A/POE 

lubricant mixture (99.6/0.4) in a fluid heated micro-fin tube were presented. The 

measured convective boiling Nusselt numbers were compared with an existing 

correlation. Approximately 71 % of the measured, lubricant-free, data fall within ± 20 % 

of the predicted data. Similarly, 83 % of the measured R410A/POE Nusselt numbers 

were predicted to within ± 20 %. 

In general, the measured boiling heat-transfer coefficients exhibited little variation with 

increasing qualities. An average degradation of 6.7 % was observed in the heat transfer 

coefficient of the R410A/POE mixture when compared with the lubricant-free R410A. 

The relative heat transfer of the R410A/POE mixture was between –20 % and +42 % of 

that of the lubricant-free R410A. 

A fluorescence setup was designed to observe fluorescence for identification of and 

location of the lubricant in the convective flow. The optimum emission and excitation 

wavelengths were identified using the spectrofluorometer. A source lamp with a greater 

intensity, a lens to focus the source onto the tube, and interference filters with wider band 

widths would all likely improve the observation of the fluorescence intensity. 

ACKNOWLEDGEMENTS

This work was funded by NIST. The authors thank the following NIST personnel for 

their constructive criticism of the first draft of the manuscript: Dr. P. Domanski, Dr. W. 

V. Payne, and Dr. M. Kim. Thanks also go to Wolverine Tube, Inc., for supplying the 

Turbo-A, micro-fin tube for the test section. The RL68H (POE) lubricant that was 
17

donated by Dr. S. Randles of ICI is much appreciated. The mineral oil that was mixed 

with the POE to obtain a significant fluorescence (York-C) was donated by Mr. K. 

Starner and Dr. M. Naduvath of York and is much appreciated. 
18

NOMENCLATURE 

a fourth degree polynomial in wl, eq. (2) (K) 

Ac cross sectional flow area inside tube (m2) 

An coefficients given in eq. (3) 

Ai actual inner surface area of tube (m2) 

Ar ratio of actual inner surface area of micro-fin tube to that of a smooth tube of 

inner diameter Dr 

b fourth degree polynomial in wl, eq. (2) 

Bn coefficients given in eq. (4) 

Bo local boiling number, q”/Grifg

cp specific heat (J/kg·K) 

Cn coefficients given in eq. 8 

D tube diameter (m) 

De equivalent inner diameter of smooth tube, (4Ac/Π)

0.5 (m) 

Dh hydraulic diameter of micro-fin tube (m) 

e fin height (m) 

Eh heat-transfer enhancement ratio 

G total mass velocity (kg/m2·s) 

h2φ local two-phase heat-transfer coefficient (W/m2·K) 

ifg latent heat of vaporization (J/kg) 

k refrigerant thermal conductivity (W/m·K) 

L tube length (m) 

Nu local Nusselt number based on Dh 

Nf number of fins 

m mass flow rate (kg/s) 

Mr molar mass (g/mol) 

p wetted perimeter (m) 

P local fluid pressure (Pa) 

Pr liquid refrigerant Prandtl number 

q” local heat flux based on Ai (W/m2) 

Re all liquid, refrigerant Reynolds number based on Dh

s spacing between the fins 

S perimeter of one fin and channel (m) 

Sv non-dimensional refrigerant specific volume 

tb thickness of the fin at its base (m) 

tw thickness of the tube wall (m) 

T temperature (K) 

Tb bubble point temperature of mixture (K) 

Td dew point temperature of mixture (K) 

U expanded relative uncertainty 

w lubricant mass fraction 

xq thermodynamic mass quality 

z axial distance (m) 
19

Subscripts

b bulk condition, fin base, bubble point 

c critical condition 

d degradation, dew point 

f water 

i inlet, inner 

l local 

LV least volatile 

m measured 

MV most volatile 

p plain or smooth tube, predicted 

r refrigerant 

s saturated state 

w heat transfer surface 

v vapor 
20

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alternatives: Part 1 – In tube flow-boiling tests,” ASHRAE Transactions, Vol. 101, Pt. 1, 

pp. 491 – 508. 

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Measuring the Concentration of Lubricant on a Boiling Heat Transfer Surface,” NISTIR 

6110, U.S. Department of Commerce, Washington, D.C. 

Kedzierski, M. A., and Goncalves, J. M., 1999, “Horizontal Convective Condensation of 

Alternative Refrigerants Within a Micro-Fin Tube,” Journal of Enhanced Heat 

Transfer, Vol. 6, No. 2-4, pp. 161-178. 

Hamilton, L. J., Kedzierski, M. A., and Kaul, M. P., 2005, “Horizontal Convective 

Boiling of Refrigerants and Refrigerant Mixtures within a Micro-Fin Tube,” NISTIR 

7243, U.S. Department of Commerce, Washington, D.C.

Kim, M., and Shin, J., 2005, “Evaporating Heat Transfer of R22 and R410A in 

Horizontal Smooth and Microfin Tubes,” Int. J. Refrig., Vol. 28, pp. 940-948. 

Lemmon, E. W., Huber, M. L, and McLinden, M. O., 2006, NIST reference fluid 

thermodynamic and transport properties-REFPROP. NIST standard reference database 

23-Version 7.1. 
21

Lottin, O., Guillemet, P., and Lebreton, J.-M., 2003, “Effects of Synthetic Oil in a 

Compression Refrigeration System Using R410A. Part II: Quality of Heat Transfer and 

Pressure Losses Within the Heat Exchangers,” Int. J. Refrig., Vol. 28, pp. 1018-1028. 

Meyer, J. J., and Saiz Jabardo, J. M., 1994, "An Ultrasonic Device for Measuring the Oil 

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Concentration in liquid refrigerants: in situ measurement," Int. J. of Refrigeration, Vol. 

22, No. 6, pp. 499-508. 

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22

 

Table 1 Median estimated 95 % relative expanded uncertainties and range for 

measurements 

Parameter Minimum Maximum U % 

Gr [kg/m2·s] 57 552 2.0 

Tr [K] 293.0 323.0 0.1 (0.3 K) 

P [kPa] 600 2000 1.5 

Tw [K] 288.0 318.0 0.1 (0.25 K) 

mf  [kg/s] 0.0150 0.0450 2.0 

Tf [K] 278.0 313.0 0.1 

Pf [kPa] 200 110 1.0 

q" [kW/m2] 1.5 43.8 5.1 

dTf/dz [K/m] 0.009 0.45 5.2 

Nu 78 410 16.4 

Re 5300 16500 4.0 

xq 0.001 .72 8.0 

ΔTs [K] 0.8 6.0 15.2 (0.44 K) 
23

Table 2 Calibration check of spectrofluorometer against Mercury lamp 

Published* 

Wavelength (nm) 

Measured Wavelength 

(nm) 

Absolute Difference 

(%) 

Relative Difference 

(%) 

312.567 310 2.567 0.8 

365.015 362 3.015 0.83 

404.656 402 2.656 0.66 

435.833 434 1.833 0.42 

546.074 545 1.074 0.2 

576.96 578 -1.04 -0.18 

* Reader et al. (1980) 
24

Fig. 1 Schematic of test rig 
25

Fig. 2 Cross section of test section 
26

Fig. 3 Test section detail 
27

Fig. 4 Comparison of measurements with REFPROP 7.1 (2006) 
28

Fig. 5 Difference between oil-free R410A and R410A/POE saturated liquid equilibrium 

equations 
29

Fig. 6 Heat flux distribution for pure R410A and R410A/POE mixture 
30

Fig. 7 Comparison of the heat transfer coefficient for pure R410A and R410A/POE 

mixture 
31

Fig. 8 Comparison of counterflow data with existing correlation 
32

Fig. 9 Apparent heat transfer degradation due to oil 
33

Fig. 10 Schematic of right angle spectrofluorometer 

 

35

Fig. 11 Fluorescence spectra of R410A/POE mixture 
36

Fig. 12 Mercury Spectra (Excitation = 300 nm) 
37

Fig. 13 Lubricant concentration observation test section