EPA/600/A-96/115
Equipment Configurations for Use of Ammonia in Supermarket Applications
T. McDowell, J. Mitchell, and S, Klein
Solar Energy Laboratory
University of Wisconsin at Madison
Madison, Wisconsin 53706-1687
ABSTRACT
International agreements have legislated the phase-out of many of the refrigerants currently
being used in the world, including R502 and R12 which are commonly used to provide the
cooling for the refrigerated cases in supermarkets. R22 and ammonia (R717) are candidate
replacement refrigerants having appropriate thermodynamic properties and less environmental
effects. This paper identifies the optimal design for ammonia - secondary fluid systems and
compares their performance to that of R22 systems. Both R22 and ammonia have high
discharge temperatures leaving the compressor, necessitating staged compression. Three
methods of staging the compression were compared for both refrigerants. Six secondary
fluids were evaluated for use with ammonia in the supermarket system. The overall system
performance of the ammonia - secondary fluid refrigeration system is governed by a large set
of design parameters. The influence that these parameters have on the overall system
performance was studied in a systematic manner. From this parametric study, design rules
leading to optimum ammonia - secondary fluid systems were developed. The performance of a
well-designed ammonia - secondary fluid system was found to be only about 4% lower than
that of an R22 system.
INTRODUCTION
The growing concern about the environment has led to international agreements to eliminate
substances that cause ozone depletion or global warming, including many of the refrigerants
currently being used. The refrigerants of most concern are the fully halogenated
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chlorofluoroearbons (CFCs) and the nori-fully halogenated chlorofluorocarbons (HCFCs).
Most supermarket refrigeration systems currently utilize either refrigerant R12 or R502, both
of which are scheduled to be phased out. For both R12 and R502, the near-term replacement
refrigerant in supermarket applications is the refrigerant R22. However, R22 is an HCFC and
according to current plans the production of HCFCs will be phased out before the year 2030.
Another possible replacement refrigerant for supermarket applications is ammonia (refrigerant
R717).
The advantages of ammonia as a refrigerant are discussed by Stoecker (Stoecker 1989).
Ammonia is cheaper and has higher cycle efficiencies, higher heat transfer coefficients, and a
higher critical temperature than most CFCs and HCFCs. Because ammonia has a pungent
odor, it is easy to detect leaks. Water vapor remains in solution with ammonia and will not
separate and freeze as it can with other refrigerants. However, water contamination can still
cause chemical changes and should be avoided. Oils and ammonia are virtually insoluble in
each other, while oils and hydrocarbon liquids are mutually soluble. Oil that collects in the
ammonia system will have to be drained off at an inactive point in the system and returned to
the compressor. A major advantage is that ammonia has zero ozone depletion. When released
into the atmosphere, it reacts with water in the air to form ammonium hydroxide and is quickly
removed from the atmosphere. Ammonia also has a zero Global Wanning Potential (GWP).
There are drawbacks to the use of ammonia as a refrigerant (Stoecker 1989). The behavior
of ammonia with oils can also be considered a disadvantage, because an oil separator is
required. Ammonia is not compatible with copper and copper bearing alloys, so steel and
aluminum must be used as the construction materials. Hermetically sealed compressors cannot
be used with ammonia because the ammonia destroys the copper wiring in the motors. Open
compressors must be used instead. The temperature of ammonia leaving the compressors of
refrigeration systems is very high and steps, such as cooling the heads of the compressors or
staged compression with intercooling, need to be taken to reduce the temperature (R22 has the
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3
same problem). Ammonia is weakly flammable at concentrations of 16 to 25% by volume in
air. However, the major drawback is the low concentrations at which it is considered toxic.
Special precautions must be taken to prevent the build-up of dangerous concentrations in
occupied areas.
A refrigeration system utilizing R22 as the refrigerant consists of a condenser and
compressor rack in the mechanical room and distribution pipes that transport the refrigerant to
the refrigerated cases where evaporation occurs. A diagram of the system is shown in Figure
1.
EXPANSION
VALVE
COMPRESSOR
EVAPORATOR
CONDENSER
REFRIGERATED CASE
Figure 1: Diagram of R22 refrigeration system
AMMONIA CYCLE
EXPANSION
VALVE
COMPRESSOR
SECONDARY FLUID LOOP
PUMP
HEAT EXCHANGER
HEAT EXCHANGER
CONDENSER
REFRIGERATED CASE
Figure 2: Diagram of an ammonia - secondary fluid refrigeration system
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A system using ammonia would require a different configuration. Because of its toxicity,
it would be a potential health risk to circulate ammonia throughout the supermarket. The
ammonia would have to be confined to a well-ventilated mechanical room. To provide the
cooling to the refrigerated cases, ammonia would be used to cool a secondary heat transfer
fluid which would then be used in a heat exchanger to cool the air in the refrigerated case. The
warm secondary fluid would return to the evaporator of the ammonia system where it would be
cooled. A diagram of such a system is shown in Figure 2,
MODEL DEVELOPMENT
The performance of conventional refrigeration systems is well known. In contrast, the
performance of ammonia systems with secondary loops has not been studied. A computer
model of the ammonia with a secondary fluid refrigeration system was written utilizing an
equation solving program (Klein and Alvarado 1993) to determine the performance and to
develop design rules. The methods used to model the different system components are
discussed here.
Condenser
The condenser model is representative of an air-cooled condenser and uses the
effectiveness-NTU method to solve for the heat transfer (Chapman 1984). The total heat
transfer in an actual process includes both the desuperheating, the condensing, and possibly
the subcooling of the refrigerant. The major heat flow is due to condensation, and the
mechanism equation used in this condenser model assumes that the refrigerant is isothermal at
the condensing pressure as recommended by Stoecker and Jones (Stoecker and Jones 1982).
The effectiveness-NTU equation for an isothermal phase change is used to calculate the heat
transfer and the change in enthalpy of the ammonia:
£ = 1 - exp(-NTU) (1)
-------
Q — £ ri\-ur Cp;ijr (Tref - Tainh) (2)
Compressor
A polytropic model based on actual physical dimensions is used to model the reciprocating
compressor (Threlkeld 1970 and Chlumsky 1965). The polytropic exponent (n) of the
refrigerant determines the relationship between the entering and exiting states. The power used
by the compressor is calculated from the flow rate of the refrigerant, the enthalpy change in the
refrigerant, and the polytropic efficiency of the compressor. The polytropic process is defined
by
Pinvin" = Poutvoutn (3)
The compressor work is given in terms of the polytropic efficiency by
«/ _ mrefAhref
w comp - ~ (4)
11 polytropic
Evaporator
The evaporator model is for a flooded, shell and tube heat exchanger with the secondary
fluid flowing through the tubes and ammonia in the shell. The secondary fluid enters the tube
bundle of the heat exchanger, and is cooled by the ammonia and then recirculated to the
refrigerated case. The ammonia enters as a mixture of liquid and vapor after leaving the
expansion valve. All expansion valves are assumed to be isenthalpic. As the ammonia cools
the secondary fluid, it evaporates and the saturated ammonia vapor flows to the compressor
where it is compressed. The effectiveness of the heat exchanger is calculated from the heat
exchanger geometry and the heat transfer coefficients on the inside and outside of the pipes.
On the inside of the pipes, the flow is considered to be developing hydrodynamically and
developed thermally. For laminar flow, the Hausen correlation is used to calculate the Nusselt
number (Chapman 1984):
0.()668p)ReDPr
Nud = 3.66 + -L- (5)
1 +0.4
l^jRepPr
2/3
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6
For turbulent flow, an equation developed by Nusselt accounting for the entry length effects is
used (Chapman 19X4):
NuD = O.O36ReDa8Pr1/301/lg (6)
On the outside of the pipes, ammonia is evaporated in a pool boiling process. To calculate the
heat transfer coefficient, a correlation developed by Rohsenow for pool boiling is used
(Rohsenow 1952):
h = (iihf.
fi(Pl- pv)
a
l/2r
Cp,
(TwjiI 1" Tjaturated)^
(7)
CjifhfgtPr,)1-7
Once the heat transfer coefficients are calculated, the overall heat transfer coefficient and the
UA product are determined. The heat transfer and temperature changes are calculated using the
effectiveness-NTU method (Chapman 1984).
Refrigerated Case
The refrigerated case is modeled as a heat exchanger which cools the air circulating in the
refrigerated case with a secondary fluid flowing through a tube bundle oriented cross-flow to
the air stream. Standard heat exchanger modeling techniques are used to determine the overall
heat transfer coefficient. The secondary fluid flow through the pipes in the heat exchanger is
assumed to be developing hydrodynamically and developed thermally, and the Hausen
correlation for laminar flow and the Nusselt correlation for turbulent flow are again used. Air
is circulated on the outside of the pipes, resulting in heat transfer by forced convection over
horizontal pipes. It is assumed that the pipes do not have extended surfaces. Churchill and
Bernstein developed a correlation equation for this geometry (Chapman 1984):
Nud = 0.3 +
0.62ReD1/2Pr1/3
[l +(0.4/Pr)2/3]1/4
Rep
2,82 x 10-
5/8
4/5
The effectiveness-NTU method is used to determine the heat transfer and temperature changes.
The effectiveness of the case heat exchanger is calculated using the formula for cross-flow heat
exchangers with both fluids unmixed (Incropera and Dewitt 1985):
e = 1 - exp
U(NTU)('-22 jexp[-Cr(NTU)°-78] - 1J
C^f /
(9)
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7
The heat transfer in the case heat exchanger is the load met by the refrigerated case. Either the
temperature change of the secondary fluid can be provided to calculate the case load, or the
required load can be specified to determine the necessary temperature change.
Piping Thermal Losses
The secondary fluid piping system between the two heat exchangers involves both thermal
losses and pumping requirements. In order to determine the thermal losses to the environment,
it is necessary to calculate the overall heat transfer coefficient of the piping system. The flow is
assumed to be fully developed both hydrodynamically and thermally and to be turbulent. The
heat transfer coefficient on the inside of the pipes is calculated using the Dittus-Boelter
correlation (Incropera and Dewitt 1985):
The pipes are exposed to the air of the supermarket, and a constant heat transfer coefficient
of 6 W/m2-C (1.05 BTU/hr-ft^-R) based on convection and radiation is used for the outside of
the pipes (Chapman 1984). The thickness of the insulation on the pipes is a design parameter.
Heat transfer resistance due to conduction through the pipe walls is neglected. Since the heat
transfer properties are dependent on the bulk temperature of the secondary fluid and the
temperatures depend on the heat transfer from the pipes to the environment, both an energy
balance and a heat transfer rate equation are necessary to determine the inlet and outlet
temperatures and the heat transfer. The heat transfer rate is based on the log mean temperature
difference of the secondary fluid and air temperatures.
Piping Head Losses
The first step in determining the pump work is to calculate the head losses in the pipes.
The head losses arise from the friction losses and the minor losses due to bends and valves.
The friction losses are calculated using the friction factor (f) from the Moody diagram (White
1986):
Nud = 0.023ReD°-8Pr0-4
(10)
(H)
-------
X
The head losses from the minor losses due to bends and valves in the piping system depend on
the number and type of the bends and valves. Each bend and valve is assigned an equivalent
loss coefficient (Keq) and then the total equivalent loss coefficient is the sum of all of the
equivalent loss coefficients (White 19X6). For the calculations completed in this study it was
assumed that each pipe had a sharp entrance and exit. This equates to a Keq of 8 per pipe for
each pipe in the heat exchangers and for the supply and return pipes:
The total head loss is the sum of the losses due to friction and the minor losses. Once the total
head losses are determined, the pressure drop and the pump work are calculated. The pump
work is calculated accounting for the pressure drop in the distribution lines, the heat exchanger
in the refrigerated case, and the heat exchanger with ammonia. Motor and mechanical
inefficiencies are not included in the pump work.
The higher cycle efficiencies of ammonia are offset by the additional pump work required
to circulate the secondary fluid. The pump work is added to the compressor work to calculate
the system coefficient of performance (COP):
Compressor Staging
The practical use of ammonia necessitates a means of controlling the compressor discharge
temperatures. The temperature leaving the compressor can be reduced by staging the
compression and making use of intercooling between the stages. Three methods of staging the
compression were compared for use with ammonia and R22 (McDowell 1993).
The first method is known as basic staged compression (Gosney 1982). This method
involves extracting some of the refrigerant leaving the condenser at an intermediate pressure
and mixing it with the refrigerant leaving the first compressor at the same intermediate
(12)
Refrigeration Load
(13)
^compressor + WpUinp
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9
pressure. The advantage of basic staging is that the refrigerant is desuperheated to saturated
conditions between the two compressors, causing the compressed gas to exit the second stage
of compression at a lower temperature. The lower temperature also leads to a higher
volumetric efficiency in the second compressor. The gas entering the second compressor has a
smaller specific volume than it would if no desuperheating took place, allowing a smaller
compressor to be used in the second stage. However, a higher mass flow rate of refrigerant is
needed to provide both the refrigeration and the intercooling.
The second method, known as staged compression and evaporation, differs from basic
staging because all of the refrigerant leaving the condenser is expanded at an intermediate
pressure (Gosney 19X2). The liquid refrigerant separated out at the intermediate pressure is
expanded again for use in the evaporator, while the vapor is used to mix with and cool the
vapor leaving the first compressor. This type of staging produces the same desuperheating
advantage as for the staged compression method but it is less effective than staged compression
since the vapor is not cooled to as low a temperature. However, an advantage of the staged
compression and evaporation method is that the refrigerant is expanded twice, so the enthalpy
difference across the evaporator is greater and less mass flow of refrigerant is needed to meet
the refrigeration load, reducing the size of the first compressor stage.
CONDENSER
COMPRE SSOR
FLASH TANK
E\A PORATOR
COMPRE SSOR
Figure 3: Refrigeration cycle with staged compression and flash tank
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The third method utilizes a flash tank between the condenser and the evaporator and
between the two stages of compression (Stoecker and Jones 1982), as shown in Figure 3.
This method is similar to the staged compression and evaporation method in that the refrigerant
leaving the condenser is expanded at an intermediate pressure and some of the resulting liquid
is used in the evaporator. However, it differs in that the vapor entering the second compressor
is saturated. The liquid and vapor from the expansion of the refrigerant leaving the condenser
enter a flash tank where some of the liquid is extracted and sent to the evaporator. The
refrigerant leaving the evaporator is compressed to the intermediate pressure in the low
pressure compressor and bubbled through the remaining liquid and vapor in the flash tank.
The resulting saturated vapor in the flash tank is removed and compressed in the second
compressor to the condensing pressure. A higher mass flow rate through the condenser is
needed than in the staged compression and evaporation method to provide the refrigerant for
intercooling. However, the refrigerant is desuperheated to the saturation point, resulting in
increased volumetric efficiency and decreased size for the second stage compressor. The
expansion is staged and has the same refrigeration capacity advantage as the staged
compression and evaporation method. An optimal intermediate pressure, discussed in the
section on Design Rules, gives the highest performance.
The models of the different staging methods with R22 and ammonia as the refrigerant were
compared for evaporator temperatures of 244 and 267 K (434 and 481°R) and a refrigeration
load of 52.8 kW (15 tons). The COPs for the different systems are shown in Table 1. With
R22 as the refrigerant, staged compression and evaporation has the highest performance. R22
has a smaller superheating horn than ammonia and does not benefit as much from
desuperheating and intercooling. Thus the higher flow rate needed in staged compression with
a flash tank penalizes the performance of R22 more than the advantage of intercooling. With
ammonia as the refrigerant, staged compression with a flash tank yields the highest
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performance. Staged compression with a flash tank is used with ammonia, and staged
compression and evaporation is used with R22 in the rest of this study.
Method
Evaporating Temperature
244 K (434 °R)
267 K (481 °R)
COP R22
COP NH3
COP R22
COP NH3
Single stage of compression
1.44
1.52
2.50
2.90
Staged compression
1.43
1.61
2.49
3.03
Staged compression and evaporation
1.69
1.69
2.76
3.09
Staged compression with flash tank
1.68
1.84
2.74
3.25
Table 1: Compression staging performance comparison
SECONDARY FLUTD SELECTION
The secondary fluids evaluated in this study are propylene glycol, ethylene glycol, mineral
oil, ethanol, propane, and a silicone-based heat transfer fluid. Propylene glycol-water
solutions are used in applications where oral toxicity is a concern, such as applications with
drinking water or food processing. Ethylene glycol is less viscous than propylene glycol, and
it generally provides greater heat transfer and better low temperature performance. It is,
however, moderately orally toxic and should be used with caution where accidental contact
with food can occur. A low temperature mineral oil fluid, polyalphaolefin, is a non-toxic
substance that meets the Food and Drug Administration (FDA) regulation for use as a synthetic
white mineral oil for non-food articles in contact with food. The low temperature silicone-
based heat transfer medium is a specially formulated silicone polymer, dimethyl polysiloxane.
Ethanol (ethyl alcohol) is both flammable and explosive. Propane can also be used as a
secondary fluid, although it is highly flammable and explosive. It is necessary to ensure that
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12
the propane pressure is high enough that the propane remains in liquid form throughout the
system.
Correlations were developed to relate the properties of the different fluids to temperature
and concentration (when applicable) (McDowell 1993). These correlations were then used in
the refrigeration system model, and the overall system performance was calculated for each
fluid. Figure 4 shows the overall system COP (Eq. 3) versus the temperature difference across
the refrigerated case heat exchanger. At the smaller temperature differences, a larger mass flow
rate of secondary fluid is needed to meet the refrigeration load and the pump work is higher.
At higher temperature differences, the temperature of the ammonia in the ammonia - secondary
fluid heat exchanger needs to be lower and the compressor work is higher. With a refrigerated
case temperature of 267 K (4X1 °R), propane has the highest performance. Propylene glycol
and ethylene glycol have the next highest performance. Ethanol has a performance almost as
high as propylene glycol and ethylene glycol. The silicone based fluid and the mineral oil have
the lowest performance.
2.4
2.2
2.0
cx
O
u
propylene glyco'
- - ethylene glycol
- • - mineral oil
ethanol
- - - silicone fluid
propane
2
4
3
6
5
7
8
Temperature Difference Across Case Heat Exchanger (K)
Figure 4: Performance comparison of the secondary fluids
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Figure 4 shows that the choice of propane as the secondary fluid will yield the highest
performance by about 10%. However, propane is both flammable and explosive, while
propylene glycol and ethylene glycol are both non-flammable. Propylene glycol is non-toxic,
and ethylene glycol is orally toxic. In a supermarket, safety is a concern and propylene glycol
would likely be the best choice of the six secondary fluids examined here.
COMPARISON OF R22 AND AMMONIA WITH SECONDARY FLUID SYSTEMS
Ammonia - secondary fluid refrigeration systems will be practical only if their performance
is comparable to the performance of the R22 systems that they will replace. The model used to
evaluate the performance of the ammonia - secondary fluid systems was written to include the
pumping and thermal losses associated with the heat exchangers and the distribution of the
secondary fluid throughout the supermarket. To compare the performance to that of a R22
system, it was necessary to develop a model of a R22 system that included a heat exchanger in
the refrigerated case and pumping costs.
The equations used to calculate the thermal losses and the pressure drops in the supply and
return Ones for the refrigerated case in the R22 system model are identical to those used in the
ammonia - secondary fluid system model. The refrigerated case is assumed to have R22
circulated in a tube bundle oriented cross-flow to the air stream. The heat transfer coefficients
of the air flow on the outside of the pipes are calculated in the same manner as in the ammonia -
secondary fluid model. The correlation equations become confounded by the forced
convection during evaporation. An assumption was made that the inside heat transfer
coefficient would be at least an order of magnitude greater than the heat transfer coefficient on
the outside of the pipes due to the phase change of the R22 on the inside of the pipes. The
effectiveness of the heat exchanger is determined using the effectiveness - NTU equation for
heat exchangers with one stream changing phase.
-------
The performance of the R22 system (using staged compression and evaporation) and the
ammonia with propylene glycol system (using staged compression with a flash tank) were
compared at refrigerated case temperatures of 267 K (481 °R) and 244 K (434 °R). The results
are shown in Table 2. The system COP for the R22 system is around 4% higher than the
system COP for the ammonia with propylene glycol system at 267 K (481 °R) and 10% higher
at 244 K (434 °R). This difference between the two systems could possibly be slightly
reduced with improvements in the heat exchanger design used in the ammonia - propylene
glycol system.
Refrigerant
COP (267 K) (481 °R)
COP (244 K) (434 °R)
R22
2.84
1.40
Ammonia with propylene glycol
2.72
1.25
Table 2: Performance comparison of R22 system and ammonia with propylene glycol system
DESIGN RULES
Intermediate Pressure
The selection of the operating pressure ratio of the staged compression in the ammonia
refrigeration cycle is important in providing the most intercooling and refrigeration capacity
increase, resulting in the highest COP. An analysis of the influence of the exponent in the
pressure ratio equation on the overall system performance shows that the highest performance
occurs when the exponent (X) is between 0.5 and 0.6 (McDowell 1993).
rinter
Pevap
Pcond
Pevap.
(14)
This result agrees with the estimate that the maximum performance will occur at the geometric
mean of the condensing and evaporating pressures (Stoecker and Jones 1982).
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Temperature Difference across Case Heal Exchanger
The temperature difference across the refrigerated case heat exchanger is a design
consideration. As the temperature change of the secondary fluid through the case heat
exchanger increases, colder ammonia temperatures are required in the ammonia - secondary
fluid heat exchanger. The colder the ammonia, the more compressor work is needed. If the
temperature difference decreases, the ammonia temperature can be higher and the compressor
work is reduced, but the pump work needed to circulate the secondary fluid increases. The
highest system performance will occur at a temperature difference that balances the compressor
and pump work. The relative influence of the ratio of the compressor work and the pump
work was calculated for four heat exchanger and piping system combinations as shown in
Figure 5, where "num" stands for the number of pipes in the heat exchanger and "radius" is the
inside radius of the pipes. The highest overall system performance occurs when the ratio of
pump work to compressor work is between 0.01 and 0.03.
Q-*
o
u
2.4
2.3
2.2
2.1
/
/
- num = 300; radius = 0.01 m
- - num = 300; radius = 0.005 m
- - num = 500; radius = 0.01m
¦ - num = 500; radius = 0.005 m
_L_
0.01 0.02 0.03
Pump Work/Compressor Work
0.04
Figure 5: System performance as a function of compressor work - pump work ratio
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Relative Heat Exchanger Sizes
The supermarket refrigerated case system with ammonia and a secondary fluid utilizes two
heat exchangers to provide the cooling. The first is between the ammonia and the secondary
coolant, and the second is in the refrigerated case. As the mass flow rate - specific heat ratio of
the secondary fluid stream increases, the effectiveness of the refrigerated case heat exchanger
increases and the effectiveness of the ammonia - secondary fluid heat exchanger decreases,
leading to an optimization problem involving the overall loss coefficients (UA) of the two heat
exchangers.
area =
200 m2;
radius = 0.01 m
— ~ " area =
200 m2;
radius = 0.005 m
* area =
150 m2;
radius = 0.01 m
- • - • - area =
150 m2;
radius = 0.005 m
2.5
2.4
0-
O
o
2.3
2.2
2.1
0.2 0.3 0.4 0.5 0.6 0.7
Ratio of Heat Exchanger Values
Figure 6: System performance as a function of UA value ratio
-
— — — —
~~ — — -
i
1
-
-
• 1
' !
¦ !
/ . __.
-
-
h "
i ' i ' i
rill
I i i I
i i i i
i i i i
The system performance was calculated for four total heat transfer areas and pipe diameter
combinations. The total heat transfer area is the combined heat transfer area of the refrigerated
-------
case heat exchanger and the ammonia - secondary fluid heat exchanger. The results of the
comparison as a function of the ratio of the UA value of the refrigerated case heat exchanger to
the UA value of the ammonia - secondary fluid heat exchanger are shown in Figure 6, where
"area" is the total heat transfer area and "radius" is the inside radius of the pipes in the heat
exchanger. All of the plots show a maximum performance at UA ratios between 0.4 and 0.5,
so for optimal performance the heat exchangers should be sized in such a manner that the UA
value for the ammonia - secondary fluid heat exchanger is 0.4 to 0.5 times the UA value of the
refrigerated case heat exchanger.
Piping Diameters and Lengths
The length, diameter, and insulation thickness of the secondary fluid supply and return
pipes are important in the design because they influence the pump work and thermal losses. It
is assumed that the same length, diameter, and amount of insulation are used for both pipes.
Maps of the performance of an ammonia - propylene glycol system at a refrigerated case
temperature of 267 K (481 °R) as a function of pipe length, diameter, and insulation thickness
were developed. The ranges of the parameters were pipe length from 10 to 80 m (32,8 to
262.5 ft), pipe diameter from 0.05 to 0.30 m (0.164 to 0.9X4 ft), and insulation thickness
from 0.01 to 0.03 m (0.0328 to 0.0984 ft). The influence of each individual parameter is
different than the influence when all three parameters are taken together. The other parameters
in the model were held constant at their base values, and the temperature difference across the
case heat exchanger was set to 3 K (5.4 °R), which is the optimal ratio for the base values.
The maximum system COP in this comparison range was calculated at a pipe length of 10 m
(32.8 ft), a pipe diameter of 0.10 m (0.328 ft), and 0.03 m (0.0984 ft) of insulation.
To develop the performance maps, the combinations of the pipe length, pipe diameter, and
amount of insulation that yielded system COPs that were 97.5, 95.0, 92.5, 90, and 80% of the
maximum were determined and plotted. The performance maps are shown here in three parts:
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18
Figure 7 shows the imp with 0.01 in (0.0328 ft) of insulation, Figure 8 with 0.02 m (0.0656
ft) of insulation, and Figure 9 with 0.03 m (0.0984 ft) of insulation.
0.30
0.25
g" 0.20
u.
u
£
£ 0.15
0.10
0.05
< 80.0%
«o«
mmm
> 80.0%
> 91.5%
> 95 m
40 50
Length (m)
Figure 7: Performance map with 0.01 m (0.0328 ft) of insulation
The maps provide an easy way to estimate an optimal design. Assume, for example, that
the supermarket requires 40 m (131.2 ft) of pipe between the two heat exchangers. Using
0.01 m (0.0328 ft) of insulation, the system can attain a COP between 95 and 92.5% of the
maximum COP by using pipe diameters between 0.05 and 0.12 m (0.164 and 0.394 ft). With
0.02 m (0.656 ft) of insulation, performance between 95 and 92.5% of the maximum COP can
be attained with pipe diameters between 0.05 and 0.21 m (0.164 and 0.689 ft). With 0.03 m
(0,984 ft) of insulation, performance between 95 and 92.5% can be attained with pipe
diameters between 0.05 and 0.29 in (0.164 and 0.951 ft).
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19
p 0.20
&
"C
mm
~
0.05 i
Q 0.15
40 50
Length (m)
Figure 8: Performance map with 0.02 m (0.0656 ft) of insulation
The maps can also be used to determine the amount of insulation needed to attain a
specified performance level with a specific pipe length and diameter. If a pipe length of 25 m
(X2.0 ft) and a diameter of 0.17 m (0.558 ft) is to be used, 0.01 m (0.0328 ft) of insulation
will give performance between 95 and 92.5% of the maximum, 0.02 m (0.0656 ft) of
insulation between 97.5 and 95%, and 0.03 m (0.0984 ft) of insulation within 97.5% of
maximum.
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20
jB
v
r-
C
5
().3(
0.25
0.20
0.15
0.1(
0.05
10 20 30
40 50
Length (m)
60
Figure 9: Performance map with 0.03 m (0.0984 ft) of insulation
CONCLUSIONS
The search for replacements for chlorinated refrigerants has focused on finding new
refrigerants and mixtures. Ammonia is a proven refrigerant that has been in use for many
years. When coupled with a secondary heat transfer fluid, ammonia can be used in
applications where its toxicity would be a concern if it is used directly. This study shows that
a well designed supermarket system that uses ammonia with propylene glycol can have a
performance that is within 4 to 10% of the performance of the R22 systems currently being
utilized. This design includes an ammonia system that utilizes staged compression with a flash
tank to provide desuperheating and increased refrigeration capacity. The pressure ratio used
for the staging provides an exponent for the pressure ratio equation (Eq. 4) between 0.5 and
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21
0.6. The pump work is around 0.02 times the compressor work, and the heat exchangers are
sized so that the ratio of the UA values is around 0.5. The secondary fluid piping system is
selected using the performance maps (Figures 7-9) to achieve the highest performance
possible. The calculations indicate that the ammonia with secondary fluid system is a possible
replacement for the R22 system, and further research into improvements for the system and
heat exchangers could improve the performance of the ammonia with secondary fluid system.
ACKNOWLEDGMENTS
This research was made possible through Environmental Protection Agency (EPA)
Cooperative Agreement CR 820631-01-0.
NOMENCLATURE
Symbol Definition
COP coefficient of performance
Cp specific heat
Cr mass flow rate-specific heat product ratio for heat exchangers
Csf empirical constant for pool boiling
D diameter
f friction factor
g acceleration of gravity
h enthalpy; heat transfer coefficient
hfg heat of fusion of refrigerant
k thermal conductivity
Keq equivalent length for minor losses
L length
rh mass flow rate
n polytropic exponent
NTU number of transfer units
Nu Nusselt number
P pressure
Pr Prandtl number
O heat transfer
Re Reynolds number
T temperature
UA loss coefficient
v specific volume: velocity
W work
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22
X
pressure ratio exponent
A
difference
£
heat exchanger effectiveness; pipe roughness
"Hpolytropic
polytropic efficiency of compressor
a
surface tension
P
density
Subscripts
Symbol
Definition
air
air properties
amb
ambient conditions
eomp
compressor
cond
condenser
D
diameter
evap
evaporator
in
into compressor
inter
intermediate pressure
1
liquid state
out
out of compressor
ref
refrigerant
saturated
saturated conditions
v
vapor state
wall
surface between ammonia and secondary fluid in heat exchanger
REFERENCES
Chapman, A. J,, Heat Transfer, Mac VI ill an Publishing Company, New York, 1984,
Chlumsky, V., Reciprocating and Rotary Compressors, SNTL—Publishers of Technical
Literature, Prague, 1965.
Gosney, W, B., Principles of Refrigeration, Cambridge University Press, Cambridge, 1982.
Incropera, F. P., and Dewitt, D. P., Introduction to Heat Transfer, John Wiley & Sons,
New York, 19X5.
Klein, S. A., and Alvarado, F. L., EES: Engineering Equation Solver, F-chart Software,
Middleton, Wl, 1993.
McDowell, T. P., "Investigation of Ammonia and Equipment Configurations for Supermarket
Applications," M.S. Thesis, University of Wisconsin - Madison, 1993.
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23
Rohsenow, W, M., "A Method of Correlating Heat Transfer Data for Surface Boiling
Liquids," Transactions of the ASME, vol. 74, 1952, p. 969.
Stoecker, W. F„ Opportunities for Ammonia Refrigeration, Heating/Piping/Air
Conditioning, September 1989, pp. 93-1 OH.
Stoecker. W. F., and Jone.s, J. W,, Refrigeration and Air Conditioning, McGraw-Hill Inc.,
New York, 1982.
Threlkeld. J. L., Thermal Environmental Engineering. 2nd edition, Prentice-Hall Inc.,
New Jersey, 1970.
White, F. M., Fluid Mechanics, McGraw-Hill Book Company, New York, 1986.
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„ TECHNICAL REPORT DATA
IX La jr IIO/S (Please read Instructions on the reverse before completing)
1. REPORT NO. 2.
EPA/600/A-96/115
3. RECIPiE
4. TITLE AND SUBTITLE
Equipment Configurations for Use of Ammonia in
Supermarket Applications
5. REPORT OATE
6. PERFORMING ORGANIZATION CODE
7.AUTHOR(s)^mo(.j1y McDowell, John W. Mitchell, and
Sanford A. Klein
8. PERFORMING ORGANIZATION REPORT NO.
9. PERFORMING ORGANIZATION NAME AND ADDRESS
Solar Energy Laboratory
University of Wisconsin—Madison
1500 Johnson Drive
Madison. Wisconsin 53706-1687
10. PROGRAM ELEMENT NO.
11. CONTRACT/GRANT NO-
CR820631-01-0
12. SPONSORING AGENCY NAME ANO ADDRESS
EPA, Office of Research and Development
Air and Energy Engineering Research Laboratory
Research Triangle Park, NC 27711
13. TYPE OF REPORT AND PERIOD COVERED
Published paper: 9/92-5/94
14. SPONSORING AGENCY CODE
EPA/600/13
15. supplementary NOTES AEERL project officer is Evelyn Baskin, Mail Drop 62B, 919/541-
2429. Presented at ASHRAE Meeting, Chicago, IL, 1/28-2/1/95.
is. abstract paper discusses equipment configurations for the use of ammonia in
supermarket refrigeration applications. International agreements have legislated
the phaseout of many refrigerants currently being used in the world, including R502
and R12 which are commonly used to provide the cooling for refrigerated cases in
supermarkets. R22 and ammonia (R717) are two of the refrigerants that have been
proposed as replacements. This paper identifies the optimal design for ammonia-
secondary fluid systems and compares their performance to that of R22 systems.
Both R22 and ammonia have high discharge temperatures leaving the compressor,
necessitating staged compression. Three methods of staging the compression were
compared for both refrigerants. Six secondary fluids were evaluated for use with
ammonia in a supermarket system. Overall system performance of the ammonia-
secondary fluid refrigeration system is governed by a large set of design parameters
The influence of these parameters on overall system performance was studied sys-
tematically. From this parametric study, design rules were developed, leading to
optimum ammonia-secondary fluid systems. The performance of a well-designed
ammonia-secondary fluid system was found to be only about 4% lower than that of an
R22 system.
17. KEY WORDS AND DOCUMENT ANALYSIS
a. DESCRIPTORS
b.IDENTIFIERS/OPEN ENDED TERMS
c. cosati Field/Group
Pollution
Ammonia
Refrigerators
Refrigerants
Substitutes
Pollution Prevention
Stationary Sources
Supermarkets
13 B
07 B
13 A
14G
18. DISTRIBUTION STATEMENT
Release to Public
19. SECURITY CLASS (This Report)
Unclassified
21, NO. OF PAGES
20. SECURITY CLASS (Thispage)
Unclassified
22. PRICE
EPA Form 2220-1 (9-73)
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