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Refrigerant Flow Through Flexible Short-Tube Orifices
a
b
c
d
Y. C. Kim , D. L. O'Neal , W. V. Payne & M. Farzad
a
Department of Mechanical Engineering , Korea University , Seoul
b
c
Department of Mechanical Engineering , Texas A&M University , College Station
Building and Fire Research Laboratory , National Institute of Standards and Technology ,
Gaithersburg, Maryland
d
Carrier Corporation , Syracuse, New York
Published online: 28 Feb 2011.
To cite this article: Y. C. Kim , D. L. O'Neal , W. V. Payne & M. Farzad (2002) Refrigerant Flow Through Flexible Short-Tube
Orifices, HVAC&R Research, 8:2, 179-190
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VOL. 8, NO. 2
HVAC&R RESEARCH
APRIL 2002
Refrigerant Flow Through
Flexible Short-Tube Orifices
Y.C. Kim, Ph.D.
Member ASHRAE
D.L. O’Neal, Ph.D., P.E.
Fellow ASHRAE
W.V. Payne, Ph.D.
M. Farzad, Ph.D.
Member ASHRAE
Flexible short-tube orifices were designed to decrease their diameter as the pressure differential
across them increases under high outdoor temperatures. A series of tests for an R-22/lubricant
mixture (mass fraction of oil 1.2%) were performed with two flexible short-tube orifices to
develop flow data over a range of typical air conditioner operating conditions. One short tube
had a modulus of elasticity of 7063 kPa and the other a value of 9860 kPa. Both short tubes had
identical lengths (14.5 mm), entrance diameters (2.06 mm), and exit diameters (2.46 mm). The
tests included both single- and two-phase flow conditions at the inlet of the flexible short tube.
Upstream pressures were varied from 1179 kPa to 2144 kPa, which corresponded to saturated
condensing temperatures of 29.4 to 54.4°C. Experimental results were presented as a function
of pressure, subcooling/quality, evaporating pressure, and modulus of elasticity. Mass flow
rates were compared with those of a rigid short tube. The flow rate through the flexible short-
tube orifices was strongly dependent on the condensing pressure, subcooling/quality, and modu-
lus of elasticity (which, in effect, changed diameter at different upstream pressures) of the short-
tube material. However, the flow rate showed little dependence on the evaporating pressure.
An empirical flow model was developed using the experimental data. This flow model was
then combined with an air-conditioning system simulation model. Results from the model indi-
cated that the flexible short tube provided approximately 2 to 3% higher capacity than the fixed
expansion devices at temperatures lower than 25°C and higher than 45°C. However, the flow
control with the flexible short-tube orifice was not as good as with a thermal expansion valve.
INTRODUCTION
The expansion device controls refrigerant flow through the system by restricting the flow of
refrigerant pumped by the compressor. The performance of an expansion device not only helps
determine the capacity of the refrigeration system but can also affect system reliability.
The most commonly used expansion devices in small air conditioners are capillary tubes, short
tubes, and thermostatic expansion valves (TXVs). Both capillary tubes and short tubes are con-
stant-flow-area expansion devices, and control the flow rate based on the flow conditions entering
the expansion devices such as condensing pressure and the amount of subcooling/quality. TXVs
control refrigerant flow rate by varying the flow area as a response to the amount of superheat at
the exit of the evaporator. Because of their low cost and easy installation, capillary tubes or short
tubes are often preferred in lower-cost air conditioners.
Yongchan Kim is an associate professor in the Department of Mechanical Engineering at Korea University, Seoul. Den-
nis L. O’Neal is a professor in the Department of Mechanical Engineering at Texas A&M University, College Station.
W. Vance Payne is a mechanical engineer at the Building and Fire Research Laboratory, National Institute of Standards
and Technology, Gaithersburg, Maryland. Mohsen Farzad is senior staff engineer at Carrier Corporation, Syracuse, New
York.
179

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HVAC&R RESEARCH
Constant-flow-area expansion devices have performance limitations. For an air conditioner,
as the outdoor temperature increases, the pressure differential between condenser and evapora-
tor increases. The increase in pressure differential decreases the pumping capacity of the com-
pressor, but increases the capacity of a constant-flow-area expansion device. The result of this
imbalance is that more refrigerant is directed to the evaporator, resulting in possible flooding,
and under extreme conditions, two-phase (liquid and vapor) flow enters the suction of the com-
pressors. This condition could reduce the life of the compressor. One alternative to the constant-
area expansion device has been the thermal expansion valve. It adjusts refrigerant flow (within
its operating range) so that a fixed superheat is maintained at the exit of the evaporator. The flex-
ible short-tube orifice was invented to provide flow control similar to that provided by a TXV,
but at a cost nearer that of conventional fixed-area expansion devices (Drucker and Cann 1991;
Drucker 1992; Drucker and Abbot 1993). Designing a system with a flexible short tube requires
knowledge of how the flow is affected by system variables such as evaporator and condenser
pressure and upstream subcooling/quality.
The flexible short-tube orifice has the same shape as a rigid short tube, but it is made from
materials that deform as pressure is applied. Flexible short-tubes orifices are made from elasto-
meric materials and are designed so that the flow area decreases as the tube is compressed by
increasing pressure upstream of the short tube (Drucker and Cann 1991; Drucker 1992). Previ-
ously, no work had been conducted on the flow of R-22 through flexible short tubes. Most of the
previous work on short tubes was focused on flow through rigid short tubes (Mei 1982; Krakow
and Lin 1988; Aaron and Domanski 1990; Kim and O’Neal 1994).
This paper presents experimental results for flow of an R-22/lubricant (mass fraction of 1.2%)
through flexible short-tube orifices that have two different moduli of elasticity. The experimen-
tal results were used to develop an empirical flow model that was incorporated into an air condi-
tioner simulation model. The performance of an air conditioner with the flexible short tube was
compared to the same system with a capillary tube, short tube, and TXV.
EXPERIMENTAL SETUP
A schematic diagram of the experimental setup is shown in Figure 1. The test loop was
designed to allow easy control of each operating parameter, such as upstream subcooling or
quality, upstream pressure, and downstream pressure. It also allowed for changing of the oil
concentration by injection of oil into the system. The test rig consisted of three major flow
loops: (1) a refrigerant flow loop containing a detachable test section, (2) a hot-water flow loop
used for the evaporation heat exchanger and (3) a chilled water-glycol flow loop used for the
condensation heat exchanger.
A diaphragm liquid pump with a variable-speed motor was used to provide a wide range of
refrigerant mass flow rates. The pressure entering the test section (upstream or condenser pres-
sure) was controlled by adjusting the speed of the refrigerant pump. The refrigerant subcooling or
quality entering the test section was set by a water-heated heat exchanger (evaporation heat
exchanger) and a heat tape. For single-phase conditions at the inlet of the test section, most of the
energy transfer to the refrigerant was supplied by the evaporation heat exchanger. A heat tape with
adjustable output from 0 to 0.9 kW was used to provide precise control of upstream subcooling.
For two-phase flow conditions at the inlet of the test section, the flow from the pump was heated
by the evaporation heat exchanger to 1 K of subcooling, and a heat tape was used to reheat the
refrigerant to the desired inlet quality. The quality of the refrigerant flow entering the test section
was calculated from the enthalpy and the measured pressure at the inlet of the test section.
After all upstream conditions were established, the flow entered the test section. The pressure
and temperature were measured upstream and downstream of the short tube. Two-phase refrig-
erant exiting the test section was condensed and subcooled in the water/glycol-cooled heat

VOLUME 8, NUMBER 2, APRIL 2002
181
Figure 1. Schematic Diagram of Flexible Short-Tube Test Setup
Figure 2. Schematic of Flexible Short-Tube Test Section
Table 1. Characteristics of Flexible Short-Tube Orifice Used in Test
Tube No.
Uncontracted Diameter, mm
D = 2.06
Modulus of Elasticity, kPa
1
7063
1
D = 2.46
2
2
D = 2.06
9860
1
D = 2.46
2
exchanger (condensation heat exchanger) so that the refrigerant pump had only liquid at its suc-
tion side. A liquid receiver was used before the refrigerant pump to ensure only liquid entered
the pump. The pressure at the exit of the test section (downstream or evaporator pressure) was
controlled by adjusting the temperature and flow rate of chilled water/glycol entering the heat
exchanger.
The flexible short-tube orifice test section was located between the heat tape and condensa-
tion heat exchanger. The current testing used flexible orifices with a length of 14.5 mm and no
entrance chamfering. The short tubes were made with flexible material with different moduli of
elasticity. Figure 2 shows the schematic of the flexible short-tube orifice test section. The flexi-
ble short tubes used in this investigation are listed in Table 1.

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HVAC&R RESEARCH
Table 2. Uncertainties of Experimental Parameters
Parameters
Temperature
Pressure
Standard Uncertainty
±0.4 K
±6.9 kPa
Mass flow rate
Power input
Quality (estimated)
±0.5% of reading
±0.5% of rated output
±2.5% of calculated value
The lubricant was injected into the suction side of the refrigerant pump using an air cylinder
in a batch process. The amount of the lubricant injected was calculated from the rod displace-
ment and the diameter of the cylinder. Oil concentration was determined by sampling. The pro-
cedure for sampling and calculating the oil concentration was based on ASHRAE (1984).
A series of experiments for each short-tube orifice were run to investigate the influence of the
operating parameters on the mass flow rate though the flexible orifices. Experiments were per-
formed with an oil concentration of approximately 1.2%. Conditions were chosen to cover a
wide range of operating conditions for a short-tube expansion device found in a typical residen-
tial air conditioner.
The experimental variables controlled included (1) upstream subcooling or quality, (2) up-
stream pressure, and (3) downstream pressure. Operating pressures for the flexible short tubes
tested were selected based on condensing saturation temperatures. In the present study, nominal
condensing saturation temperatures of 29.4, 37.8, 46.1, and 54.4°C were selected. The corre-
sponding upstream saturation pressures were 1179, 1455, 1779, and 2144 kPa. Downstream pres-
sures were selected based on evaporating temperatures of –1.1, 4.4, and 10.0°C. Upstream
conditions were varied by altering the degree of subcooling from 16.7 to 0 K and altering the
quality from 0 to 5%. Data were taken only at steady state. Several criteria had to be met to ensure
the reliability of the test data. The limits of controlling parameters were set as follows: upstream
pressure, ±7 kPa; downstream pressure, 14 kPa; subcooling, 0.3 Κ; and quality, ±0.3%.
Temperatures, pressures, flow rate, and power input were monitored in the test loop using a
computer and data acquisition system. The temperature at each measuring station was monitored
using a thermocouple with a standard uncertainty of ±0.4 K. The pressure at each measuring sta-
tion was measured using a pressure transducer that was calibrated with a standard deadweight
tester. The estimated standard uncertainty of the pressure measurements was ±0.2% of full scale
(
3447 kPa). The primary device used to measure mass flow rate was a Coriolis effect mass flow
meter. The estimated standard uncertainty of the flow meter was ±0.5% of full scale
10.9 kg/min). A turbine flow meter was used as a secondary meter for mass flow rate. The
(
estimated standard uncertainty of the turbine meter was ±1.0% of full scale (0.2 L/s). The two
flow measurements showed good agreement each other within 2%, which is very reasonable tol-
erance to accept the data as valid in the present experiments. A voltage transformer and a watt
transducer were used to measure the power input into the heat tape. Estimated experimental
standard uncertainty for the power was ±0.5% full scale (1.5 kW). Table 2 summarizes the stan-
dard uncertainties of experimental parameters and inlet quality.
EXPERIMENTAL RESULTS
The flexible short-tube orifices of different moduli were tested under the flow conditions
encountered by air conditioners. Two different moduli of elasticity flexible orifices were tested.
As upstream pressure was increased, mass flow rate through the flexible short tube increased.
The approximately choked conditions seen with rigid short tubes were evident for the flexible
short tubes. Once the approximately choked conditions were established, any further decrease in

VOLUME 8, NUMBER 2, APRIL 2002
183
downstream pressure produced minor increases within 5% in mass flow rate. In general, the
flow trends for the flexible short tube were consistent with those of the rigid short tubes, being
different only in the magnitude and response to changes in upstream pressure and subcooling.
The experimental results with the flexible short tube were compared with those predicted
from an empirical model for a rigid short tube (Kim and O’Neal 1994). The flexible and rigid
short tubes were of slightly different shapes. The flexible short tubes were tapered whereas the
rigid short tubes had a constant cross-sectional diameter along their length. The small-diameter
side of the flexible short tube was chosen as the diameter for comparison with a fixed short
tube.
Generally, the flexible short tubes showed higher flow rates compared to a rigid short tube
having the same entrance diameter (D = 2.06 mm). However, as the condensing pressure
1
increased from 1179 kPa to 2144 kPa, the slope of the flow rate with pressure of a flexible short
tube with a modulus of 7063 kPa decreased gradually. Thus, for higher condensing pressures the
refrigerant flow rate of the flexible short tube was almost the same as that for the rigid short
tube. The flow trends would suggest that the pressure around the flexible tube forced the flexible
short tube to contract, which would decrease the flow through it.
If the flow rate through the flexible short tube-orifice at the design point was set to the same
value as that of the rigid short tube, the flexible short tube would have a lower flow rate than the
rigid short tube at higher ambient temperature above the design point. However, for lower con-
densing pressure (due to lower ambient temperature), the flexible short tube would have a higher
flow rate than the rigid short tube. When a heat pump is operating in the cooling mode with a con-
densing temperature above the design point, it is desirable to have lower flow rate compared to a
capillary tube to balance the lower flow pumped by the compressor. As the condensing tempera-
ture decreases, it is desirable to have higher flow rate through the expansion device due to higher
compressor pumping rate at the lower condensing pressure.
Effect of Condensing Pressure
Figure 3 shows the effects of upstream pressure on refrigerant mass flow rate through a rigid
short tube and flexible short tubes with two different moduli of elasticity. The mass flow rates
for the rigid short tube were predicted values using a semiempirical model developed in an ear-
lier study (Kim 1993).
Figure 3 shows that the 9860 kPa modulus short tube produced higher flow rates for all
upstream pressures than the 7063 kPa modulus short tube. The 7063 kPa and 9860 kPa modulus
short-tube orifices produced mass flow rates of 320 kg/h and 339 kg/h, respectively, at the low-
est upstream pressure. As the upstream pressure increased to 2144 kPa, the 7063 kPa and
9860 kPa modulus short-tube orifices produced mass flow rates of 375 kg/h and 399 kg/h,
respectively.
It was hypothesized that the exterior of the flexible short tube was exposed to upstream condi-
tions because an axial compression of the flexible short tube provided a small gap between the
fixture and the front side of the short tube. However, there was no flow passing through the exte-
rior of the flexible short tube due to sealing at the rear. The combination of low pressure on the
interior of the flexible short tube with high pressure on the exterior could result in a constricted
inside diameter, which would be smaller than the diameter measured prior to installation. It was
also hypothesized that, once deformed, the entrance of the flexible short tubes was not sharp-
edged like the rigid short tube. This would result in higher refrigerant flow rates as compared to
the rigid short tube (Kim 1993). The degree of constriction for the flexible short tube would
depend on the modulus of elasticity of the short-tube material. The higher-modulus material
would constrict less than the lower-modulus material for a given upstream pressure. The lower-
modulus short tube showed a smaller increase in flow as the upstream pressure was increased.

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HVAC&R RESEARCH
Figure 3. Effect of Upstream (Condensing) Pressure and Modulus of
Elasticity at Subcooling of 16.7 K
This behavior would be expected to result from the increased restriction with increased pressure
provided by the more flexible, lower-modulus short tube.
An aspect of the design of the flexible short-tube orifice that would affect mass flow rate was
the tapered cross-section of the interior. The progression from a small diameter at the entrance to
a larger diameter at the exit would influence the effectiveness of constriction in changing the
mass flow rate through the short tube. The greater the difference in the two diameters, the larger
the diameter could be after the short tube was compressed. During the reinstallation of the test
section, it was also noticed that compression of the short tube installed in the fixture could result
in a bulging or increase in the diameter of the short tube. This could produce an increase in mass
flow rate. The holding fixture would need to be redesigned to ensure reliability by reducing the
compression of the flexible short tube.
Effects of Subcooling/Quality
Figure 4 shows the effects of upstream subcooling on mass flow rate for the 7063 kPa modu-
lus short-tube orifice. The line for upstream pressure of 1179 kPa in the figure corresponds to a
flow model developed for rigid short tubes with the same inlet diameter as the flexible short
tube. The general trends seen for the rigid short-tube orifice were also seen for the flexible short
tube: an increase in subcooling resulted in an increase in mass flow rate. The one difference was
the larger decrease in flow rate seen by the flexible short tubes as upstream subcooling was
reduced. Rigid short-tube orifices tended to have a fractional flow rate drop of approximately
20%, whereas the flexible tube orifices showed decreases of 25 to 30% in going from subcool-
ing of 16.7 to 0 K. At the upstream pressure of 1179 kPa and subcooling of 16.7 K, the mass
flow rate peaked at 321 kg/h. As the upstream subcooling was decreased to 0 K, the mass flow
rate decreased by 35% to 208 kg/h. The transition from single-phase at the short tube entrance to
two-phase conditions caused a sharp decrease in mass flow rate. Mass flow rate decreased from
saturated levels by 25% to 155 kg/h. The same trends occurred for the other upstream pressures
and subcoolings with a maximum mass flow rate of 366 kg/h occurring at a subcooling of
16.7 K and 2144 kPa.

VOLUME 8, NUMBER 2, APRIL 2002
185
Figure 4. Comparison of Flow Rate of Flexible Short Tube to that of Rigid Short Tube
Figure 5. Effect of Downstream (Evaporating) Pressure on Flow Rate for Flexible
Short Tube with Modulus of 7063 kPa
Effect of Evaporating Pressure
Figure 5 shows the effects of downstream pressure on mass flow rate through a flexible short
tube with a modulus of 7063 kPa. Approximately choked-flow conditions existed for all pres-
sures tested with flow rates decreasing by a maximum of 4%. The figure shows that at higher
upstream pressures, there was less tendency to increase flow rate as the downstream pressure

1
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HVAC&R RESEARCH
was lowered. For an upstream pressure of 1779 kPa, the drop in mass flow rate with an increase
in downstream pressure was less than 1%.
Mass Flow Model
A modified single-phase flow model has been successfully used in correlating two-phase flow
data through short-tube orifices (Kim 1993). Data are typically correlated with a “modified”
downstream pressure and a series of constants. The present flow model was derived from the
single-phase orifice equation with adequate modifications to capture the flow characteristics
through flexible short tubes. The single-phase orifice equation used for short tubes can be
derived from equations of continuity and energy with the given assumptions (ASME 1971). The
single-phase orifice equation for a single-component, single-phase substance is given as
2
γ ρ(P – P
)
down
·
c
up
m = CA ---- -- ------ -- ----- -- ------ -- ------ -- ----- -- --
(1)
s
4
1
– β
4
Several simplifications were made to Equation 1 to establish a flow model. The term (1 – β ) is
small compared with unity, and the term Pdown was replaced by P , the preflashing pressure, to
f
satisfy the pressure conditions at the downstream flow conditions. The adjusted downstream
pressure P covered the assumptions of incompressible flow and choked flow conditions. The
f
orifice constant C was set equal to unity, and P was correlated with the experimental data. The
f
resulting flow model is given by
·
m = CA 2γ ρ(P – Pf)
(2)
(3)
s
c
up
The adjusted downstream model ID given by
b
b
b
6
3
4
Pf = Psat[b + b (PRA )(SUB ) + b (EVAP )]
1
2
5
Based on the measured data, the adjusted downstream pressure P was correlated with a
f
normalized form of inlet subcooling, upstream pressure, downstream pressure, and flexible
short-tube diameter. The liquid saturation pressure Psat was used as a reference value for Pf ,
because flashing occurred when the pressure was near P . A correlation between the adjusted
sat
downstream pressure of P and normalized parameters was determined using a nonlinear regres-
f
sion technique along with experimental data. All coefficients included in the flow model are
given in Table 3. Due to limited data (two moduli of elasticity were tested in the present study),
the present flow model was developed separately for each flexible short tube.
Table 3. Coefficients for Flow Model
Modulus of Elasticity
Coefficients
7063 kPa
0.9187
9860 kPa
1.0434
B1
B2
B3
B4
B5
B6
10.4091
–0.0500
1.1787
7.4048
0.0191
1.0421
–0.0276
0.9241
–0.1683
–0.7797

VOLUME 8, NUMBER 2, APRIL 2002
187
Table 4. Limitations on Application of Flow Model
Parameters
Minimum
14.5 mm
2.06 mm
2.46 mm
7063 kPa
1176 kPa
481 kPa
0 K
Maximum
14.5 mm
2.06 mm
2.46 mm
9860 kPa
2149 kPa
Psat
L
D1
D2
Modulus
Pup
Pdown
Subcooling
Oil concentration
16.7 K
1.2%
1.2%
Using Equations (2) and (3) and the coefficients in Table 3, the mass flow rate can be estimated
for a given set of operating conditions and with either modulus of elasticity. When applying the
empirical flow model, the application of the model has a limited range due to the limited test range
of the experimental data (Table 4). Approximately 95% of the measured data were within ±5% of
the model's prediction. The maximum difference between the measured data and the model’s pre-
diction was within ±8%.
Flexible Short-Tube Orifice Performance Simulation
A nominal 5.3 kW capacity room air conditioner with a coefficient of performance (COP) of
2.34 at 35°C was simulated with the flexible short tube and three other expansion devices (capil-
lary tube, rigid short tube, and thermal expansion device). Performance estimates were made for
outdoor temperatures ranging from 20 to 46°C. The proprietary simulation model had several
key features that included
•
•
•
•
Tube-by-tube model of heat exchangers
Refrigerant inventory estimation
Compressor map or first principles compressor model
Models for four types of expansion valves (TXV, capillary tube, rigid short tubes, and flexible
short tubes)
The empirical equations describing flow for the flexible short tubes in the previous section were
used to model these types of short tubes. The condensate from the evaporator was sprayed over
the condenser coil using the condenser fan. The basic physical descriptions of the systems are
provided in Table 5. The unit had a reciprocating compressor with a charge of 680 g of R-22
refrigerant. Each expansion device was sized to give the same refrigerant mass flow rate, evapo-
rator superheat, capacity, and COP at 35°C outdoor air temperature.
The results of the simulation are provided in Figures 6 through 9. For the mass flow rate,
capacity, and COP, the results were normalized to 35°C outdoor temperature. The original intent
Table 5. Evaporator and Condenser Used in Window Air Conditioner
Item
Evaporator
Condenser
Rows
2
7 mm enhanced
630
2
7 mm enhanced
709
Tubes
Fin density (fins/m)
Face area (m2)
0.155
0.245
3
Airflow (m /min)
13.0
29.8

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HVAC&R RESEARCH
Figure 6. Normalized Cooling COP for Window Air Conditioner as Function of
Outdoor Temperature for Four Expansion Devices
Figure 7. Refrigerant Superheat/Quality Entering Compressor
of the design of the flexible short tubes was that they would provide control better than a fixed-
area mass flow expansion device. Figure 6 shows that the flexible short tube provides better
flow control than either the rigid short tube or capillary tube, but not as good as the TXV. At
outdoor temperatures above 45°C, the refrigerant leaving the evaporator became saturated for
both the rigid short tube and the capillary tube (Figure 7). In addition, a dirty condenser, which
is very common at outdoor temperatures above 45°C, can cause the system to operate at these
extreme conditions. As shown in Figure 7, the flexible short-tube orifice can prevent liquid
floodback at high outdoor temperatures. For outdoor temperatures below 25°C, the flexible
short tube provided 2 to 4 Κ less superheat than the fixed expansion devices.

VOLUME 8, NUMBER 2, APRIL 2002
189
Figure 8. Estimated Normalized Cooling Capacity for Window Air Conditioner as
Function of Outdoor Temperature for Four Expansion Devices
The flexible short-tube orifice provided for 2 to 3% higher capacity relative to the capacity of
the fixed expansion device system at lower (<25°C) outdoor temperatures and higher (>45°C)
outdoor temperatures (Figure 8). Capacities with the TXV-controlled system were about 1%
higher relative to the flexible short-tube-controlled system. Normalized COPs followed a similar
trend except the percentage improvements were larger.
CONCLUSIONS
The general trends, such as mass flow rate versus upstream subcooling/quality and pressure,
observed in the flexible short tubes were consistent with results for rigid short tubes (Aaron and
Domanski 1990; Kim and O’Neal 1994). For the flexible short-tube orifices, the slope of mass
flow rate versus upstream pressure was dependent on the modulus of elasticity of the short-tube
material. The higher upstream pressures forced the flexible short tube to change shape and
restrict flow. This flow restriction provided better control of evaporator outlet superheat than a
rigid short tube at high outdoor temperatures when the condenser pressure was high. Short tubes
with two different moduli of elasticity were evaluated. The data indicated that the more flexible
short-tube orifice (modulus of 7063 kPa) provided better flow control than the stiffer short-tube
orifice (modulus of 9860 kPa).
The flexible short-tube orifice provided better control of an air-conditioning system than con-
ventional fixed short tubes or capillary tubes, but there are a number of important engineering
issues that must be addressed. These include optimum flexibility of the short tube, long-term
operating performance and reliability compared to conventional rigid short tubes, and material
compatibility with current and future refrigerants and lubricants. Research will need to be con-
ducted to evaluate the optimum flexibility that would provide the best control characteristics
without risking a total collapse of the short tube.
Long-term performance and reliability of flexible short tubes should be evaluated. Although
this study demonstrated their performance benefits in the laboratory, it is not the same as a long-
term evaluation in the field. Rigid short tubes are typically made of a metal such as brass. The
flexible short tubes used in this study were made from elastomeric materials that provided the

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HVAC&R RESEARCH
flexibility. Long-term tests would need to be conducted to determine whether these materials
maintain their resiliency and do not erode after years of use.
Material compatibility of flexible short-tube orifices relates to long-term performance. With a
wide variety of refrigerants and lubricants currently in use, short-tube materials that work for
one refrigerant may not work for another. Implementing flexible short-tube orifices will require
a comprehensive study of material compatibility.
NOMENCLATURE
As
entrance cross-sectional area of a flexible
Pf
adjusted downstream pressure, kPa
upstream liquid saturation pressure, kPa
upstream pressure, kPa
short-tube orifice, m2
Psat
Pup
PRA
b
C
D1
D2
empirical coefficient
orifice coefficient
flexible short-tube entrance diameter, mm
flexible short-tube exit diameter, mm
ratio of upstream pressure to critical
pressure, P /Pc
up
SUB
Tc
Tsat
Tup
ρ
β
normalized subcooling, (Tsat – T )/Tc
up
EVAP normalized downstream pressure,
Pc – Pdown)/Pc
dimensional constant, 1.296 × 10
R-22 critical temperature, 369.17 K
liquid saturation temperature, K
temperature of upstream fluid, K
density, kg/m3
ratios of orifice throat diameter to
upstream tube diameter
(
1
0
γc
2
2
[
s ·N/(h ·kN)]
·
m
mass flow rate, kg/h
critical pressure, 4977.4 kPa
Pc
Pdown downstream pressure, kPa
REFERENCES
Aaron, D.A. and P.A. Domanski. 1990. Experimentation, Analysis, and Correlation of Refrigerant-22 Flow
Through Short-Tube Restrictors. ASHRAE Transactions 96 (1):729-742.
ASHRAE. 1984. Standard Method for Measurement of Proportion of Oil in Liquid Refrigerant. ANSI/
ASHRAE Standard 41.4-1984. Atlanta: American Society of Heating, Refrigerating, and Air-Condi-
tioning Engineers.
ASME. 1971. Fluid Meters—Their Theory and Application, 6th ed. New York: American Society of
Mechanical Engineers.
Drucker, A.S. 1992. Variable Area Refrigerant Expansion Device Having a Flexible Orifice for Heating
Mode of a Heat Pump. US Patent 5134860.
Drucker, A.S. and A.D. Abbot. 1993. Variable Area Refrigerant Expansion Device Having a Flexible Ori-
fice. US Patent 5214939.
Drucker, A.S. and P.L. Cann. 1991. Variable Area Refrigerant Expansion Device Having a Flexible Ori-
fice. US Patent 5031416.
Kim, Y. 1993. Two-Phase Flow of HCFC-22 and HFC-134a Through Short-tube Orifices. Ph.D. disserta-
tion, Texas A&M University.
Kim, Y. and D.L. O’Neal 1994. Two-Phase Flow of Refrigerant-22 Through Short-Tube Orifices. ASH-
RAE Transactions 100(1):323-334.
Krakow, K.I. and S. Lin. 1988. Refrigerant Flow Through Orifices. ASHRAE Transactions 94(1):484-506.
Mei, V.C. 1982. Short-Tube Refrigerant Restrictors. ASHRAE Transactions 88(2):157-168.