32nd Annual Meeting International Institute of Ammonia...

Technical Papers32nd Annual Meeting

International Institute of Ammonia Refrigeration

March 14–17, 2010

2010 Industrial Refrigeration Conference & ExhibitionManchester Grand Hyatt

San Diego, California

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ACKNOWLEDGEMENT

The success of the 32nd Annual Meeting of the International Institute of Ammonia Refrigeration is due to the quality of the technical papers in this volume and the labor of its authors. IIAR expresses its deep appreciation to the authors, reviewers and editors for their contributions to the ammonia refrigeration industry.

Board of Directors, International Institute of Ammonia Refrigeration

ABOUT THIS VOLUME

IIAR Technical Papers are subjected to rigorous technical peer review.

The views expressed in the papers in this volume are those of the authors, not the International Institute of Ammonia Refrigeration. They are not official positions of the Institute and are not officially endorsed

International Institute of Ammonia Refrigeration 1001 North Fairfax Street

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2010 Industrial Refrigeration Conference & Exhibition Manchester Grand Hyatt

San Diego, California


© IIAR 2010 1

Abstract

There is increasing interest in operating ammonia evaporators at reduced liquid overfeed rates (less than n=2), down to and including dry expansion (DX). Benefits include: reduced or eliminated liquid pumping power, dry or near dry suction lines, reduced liquid and suction line sizes, greatly reduced charge inventory of ammonia evaporators (as much as 30X reduction for dry expansion compared to pumped ammonia), reduced first cost, etc. A small amount of water (typically 2–6% by weight) is found in most industrial ammonia systems. This water is always concentrated in the low pressure side of the system and has the negative thermodynamic effect of increasing the bubble point of the ammonia liquid along the length of the evaporator circuits. This shift in the bubble point is normally neglected by evaporator manufacturers and designers since the shift is small (only 1–2°F) for pumped ammonia systems. However, as the liquid overfeed rate is reduced below n=2, and particularly in the case of dry expansion, this shift in the bubble point toward the end of the evaporator circuits may easily exceed the initial temperature difference (Entering Air Temperature – Evaporating Temperature), severely reducing evaporator thermal performance. Additionally, the shift in bubble point can “confuse” dry expansion control valves, further reducing performance. These thermodynamic effects of water in ammonia have not been fully recognized or well understood in the past and are the subject of this paper. The author also reviews and recommends methods for managing and removal of water from these systems.

Technical Paper #8

Thermodynamic Effects of Water in Ammonia on Evaporator Performance

Bruce I. Nelson, P.E.Colmac Coil Manufacturing, Inc.

Colville, WA


Technical Paper #8 © IIAR 2010 3


Introduction

There is increasing interest in operating ammonia evaporators at reduced liquid overfeed rates (less than n = 2), down to and including dry expansion (DX), due to the following benefits: reduced or eliminated liquid pumping power, dry or near dry suction lines, reduced liquid and suction line sizes, greatly reduced charge inventory in ammonia evaporators (as much as 30X reduction for dry expansion compared to pumped ammonia), reduced first cost, etc.

In spite of the advantages stated above, a long-standing rule of thumb exists in the industrial refrigeration fraternity regarding the application of dry expansion with ammonia: i.e., “Just don’t do it!” We are not really sure why ammonia doesn’t work well in DX systems, we just know it doesn’t. Getting to the bottom of one of the issues which makes the application of DX (and low overfeed ratios) with ammonia challenging is the topic of this paper.

Background

Ammonia is highly soluble in water due to the polarity of NH3 molecules and their ability to form (very strong) hydrogen bonds. This high solubility makes ammonia-water a good working fluid pair in absorption refrigeration machines, taking advantage of the large vapor pressure differences between the ammonia vapor and weak solution. This same behavior, on the other hand, makes water removal from ammonia refrigeration systems somewhat challenging.

It has been shown that small amounts of water (typically 2–6% by weight) can be found in most industrial ammonia systems (Nielsen 1998). The negative effects of water on ammonia refrigeration system performance are well known and widely recognized (Cotter 2007, Ficker, IIAR), however, the thermodynamic effects of small


4 © IIAR 2010 Technical Paper #8

2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California

amounts of water on ammonia evaporator performance are not so well known or recognized.

The water in an ammonia refrigeration system will always be concentrated in the low pressure side of the system and this has the negative thermodynamic effect of increasing the bubble point (i.e., temperature at which boiling begins) of the ammonia liquid along the length of the evaporator circuits. This shift in the bubble point is normally neglected by designers of pumped ammonia evaporators since it is relatively small (in the range of 1–2°F) for standard overfeed ratios of n = 3 to 4. However, as the liquid overfeed rate is reduced (n < 2), this shift in the bubble point toward the end of the evaporator circuits can exceed the initial temperature difference (Entering Air Temperature – Evaporating Temperature), severely reducing evaporator thermal performance. In the case of DX evaporators, the expansion valve controls the flow of refrigerant to maintain a constant leaving refrigerant temperature (i.e., the superheated gas temperature). Whenever water is present in the ammonia, however, the expansion valve temperature sensor interprets the increase in bubble point as superheat and actually allows the liquid ammonia-water solution to exit the evaporator and enter the suction line.

Considering ammonia-water as a binary zeotropic refrigerant mixture is a helpful approach in analyzing and quantifying the effects of small amounts of water in ammonia on evaporator performance. By now almost all of us are familiar with the zeotropic refrigerant blends, those having the refrigerant designation “R4XX.” These refrigerant mixtures, such as R410a, R404a, etc., exhibit temperature “glide” during evaporation (and condensing). This “glide,” or shift in the bubble point in the evaporator is due to a distillation process whereby the more volatile component evaporates first, leaving the remaining liquid with a different composition (and higher bubble point). Think of the evaporator as being like a moonshiner’s still…

Ammonia-water could be compared to another “R400-series” zeotropic refrigerant blend, only with a temperature glide of around 212°F – (–40°F) = 252°F!




The presence of water in ammonia negatively affects heat transfer in evaporators in two ways:

1. Temperature glide. The glide in evaporators is “up,” increasing from the circuit entrance to exit. This has the effect of reducing the overall mean temperature difference and cooling capacity compared to pure ammonia. In some cases, this shift in bubble point may actually create a “pinch-point” at the exit of the evaporator circuit. Pinch-point is a term used by heat exchanger designers to describe an undesirable condition wherein the difference between the two fluid temperatures (in our case between the ammonia and air temperatures) at some position within the heat exchanger becomes very small.

2. Heat Transfer Coefficient. Investigators have found that nucleate pool boiling (the boiling phenomenon which occurs within a pool of liquid) heat transfer coefficients for zeotropic mixtures are significantly lower than the ideal coefficients calculated for the pure components themselves (Taboas 2006). Some of the likely reasons for this trend include:

a. A local increase in the liquid temperature near the tube wall due to a preferential evaporation of the more volatile component,

b. A mass transfer resistance to the more volatile component moving toward the bubble interface,

c. A higher energy required for generating a bubble nucleus for mixtures,

d. A decrease in the nucleation sites, and

e. A nonlinear variation of mixture properties with composition.

No theoretical or empirical model currently exists which accounts for all these effects in zeotropic mixtures. Further, most research done to date on multi-component mixture boiling phenomena has been limited to nucleate pool boiling rather than convective boiling (also called “flow boiling,” “convective boiling” refers to the boiling phenomenon within a moving stream of liquid




or liquid-vapor; Thome 2006). It is convective boiling at low heat flux and low mass flux with ammonia and ammonia-water that is of primary interest to industrial evaporator designers. This remains a challenge for researchers studying such matters.

The total flow boiling heat transfer coefficient in tubes is widely assumed to be the sum of pool boiling and convective boiling coefficients (Chen 1963). In the case of air cooling evaporators, the convective boiling coefficient is the dominant component (rather than pool boiling). It is the author’s opinion that a reduction in the total flow boiling heat transfer coefficient likely exists when water is present in ammonia. However, since (a) no correlations exist to properly quantify this phenomenon, and (b) the purpose of this analysis is to examine the thermodynamic effect of shifting bubble point on evaporator performance, the boiling heat transfer coefficients used will be those for pure ammonia.

Phase Equilibrium for Ammonia-Water

In order to correctly characterize the change in composition of the ammonia-water solution through the evaporator circuit with its accompanying shift in bubble point, a phase equilibrium diagram is used. This diagram plots dew point and bubble point temperatures at varying compositions, from pure ammonia on one side of the diagram to pure water on the other side.

Detailed discussions of phase equilibria for refrigerant mixtures and methods for calculation of dew point (i.e., the temperature at which condensing begins) and bubble point (i.e., the temperature at which boiling begins) temperatures are complex and outside the scope of this paper. The interested reader is referred to Collier and Thome (Collier 1994) for an introduction to this topic.




Fortunately, software programs such as REFPROP 8 (NIST 2007) can be used to predict thermodynamic properties of ammonia-water with accuracy suitable to the purposes of this paper. REFPROP 8 was used to generate the phase equilibrium diagram shown in Figure 1.

Figure 1 shows dew point and bubble point temperatures for ammonia-water at four (4) different pressures. The horizontal axis indicates composition as mass fraction of water in the ammonia-water mixture. Pure ammonia appears to the left (mass fraction = 0), and pure water to the right side of the diagram (mass fraction = 1).

The following observations can be made when looking at Figure 1:

1. Temperature glide is very large. This results in an increase in the bubble point temperature of 10°F (5.6C) – a typical TD for many refrigeration evaporators – at a relatively low composition, i.e., between only 0.2 to 0.24.

2. The dew point lines are nearly vertical on the left side of the diagram. For the range of compositions typical for evaporators (initial composition up to approx. 0.25) this means the vapor leaving the evaporator will be very nearly pure ammonia, i.e., all of the water will be concentrated in the liquid refrigerant leaving the evaporator. This is due primarily to the very low partial pressure of water relative to ammonia in the temperature range of interest.

When using the diagram to analyze evaporator performance, the area of interest will be to the lower left on the bubble point curve (at compositions between 0 and 0.3). The ammonia-water mixture (liquid) enters the evaporator circuit at the initial composition on the left side of the diagram. As the refrigerant mixture boils, the water stays behind resulting in an increase in composition in the remaining liquid (i.e., the liquid becomes more “water-rich”). The boiling point of the liquid will follow the bubble point curve rising and moving to the right until the final composition at the coil exit is reached. The temperature of the coexisting vapor will be the same as that of the liquid, i.e., rising through the evaporator circuit.




The corresponding composition of the vapor is found by following a horizontal line starting at the bubble point line moving to the left and intersecting the dew point line, which in our case falls on the nearly vertical part of the dew point curve to the left of the diagram (again, justifying our pure ammonia vapor assumption).

The composition of the liquid can be related to the coexisting vapor mass fraction (vapor quality) in the evaporator as follows:

y =

y' (1)

(1 – x)

This relationship between initial composition, vapor quality, and resulting composition is illustrated in Figure 2. Note that composition in Figure 2 is actually shown as % water concentration.

Thermodynamic Properties of Ammonia-Water

In addition to calculating bubble point temperatures (Figure 1), REFPROP 8 was used to calculate latent heat of vaporization at various pressures as shown in Figure 3. It makes sense that the latent heat of vaporization would change with shifting composition, and in fact, it does. The change in latent heat of vaporization is significant (and quite non-linear) and is taken into account in the model calculations.

The increase in bubble point temperature with increasing composition is shown in Figure 4. Since evaporator designers normally think in terms of increasing vapor mass fraction (quality) through the evaporator circuit, the increase in the bubble point for various initial water concentrations versus vapor quality (based on the relationship from Equation 1) are shown in Figures 5 through 8. These figures can be used to determine the mass fraction of refrigerant leaving the evaporator as liquid depending on how the refrigerant flow is controlled.




For example, a DX evaporator is operating at a suction pressure of 18.3 psia (1.3 bar) with ammonia initially containing 5% water and a superheat setting of 10°F. Figure 6 shows that when the bubble point temperature has shifted 10°F the vapor quality will be 0.80. This means that 20% of the refrigerant mass leaving the evaporator will unavoidably be liquid. Figure 2 then shows that for a leaving vapor quality of 0.8 and initial water concentration of 5%, the water concentration in the liquid leaving the evaporator will be approx. 25%. Care must be taken to provide sufficient suction accumulator capacity to capture and manage this liquid before it can reach the compressor and cause damage!

Evaporator Model

In order to characterize and quantify the thermodynamic effects of water in an ammonia evaporator, a computer model was created. The model takes a segmented finite volume approach using NTU-effectiveness to calculate heat transferred for each of the finite volumes in cross flow. As heat transfer for each finite volume is calculated, a new vapor quality and resulting liquid composition for the next volume is calculated. Adjustments to bubble point temperature and latent heat of vaporization are made based on the new liquid composition as the calculation progresses from one finite volume to the next.

The following simplifying assumptions are used:

1. Heat transfer on the fin side is sensible only. This is a common assumption among evaporator manufacturers and is justified in most cases by the typically high Sensible Heat Ratios seen with refrigeration evaporators, particularly at freezer temperatures.

2. Boiling heat transfer coefficient is constant over the length of the evaporator circuit. Very few correlations exist which offer a prediction of convective boiling heat transfer coefficients with ammonia at the very low heat and




mass flux rates found in ammonia evaporators operating with dry or near dry expansion. The correlation of Zurcher et al (Zurcher 1999) was used with properties for pure ammonia at an averaged vapor quality. Heat transfer coefficients for the DX example in this paper were calculated based on the typically very low mass fluxes (approx. 15–25 kg/m2·s) which result in stratified-wavy flow patterns (Nelson 1998). Calculated heat transfer coefficients with pure ammonia for the DX example were in the range of 140–190 Btu/h·ft2 F (790–1050 W/m2·K). While correlations for convective boiling heat transfer coefficients with pure ammonia can at least be found, correlations for ammonia-water mixtures just plain don’t exist. Taboas et al presented very interesting results for nucleate pool boiling of ammonia-water as a binary mixture (Taboas 2007). Their work confirmed the work of other researchers in showing a significant reduction in the pool boiling heat transfer coefficient (as much as 50%) at all mixture concentrations except very nearly pure ammonia and very nearly pure water. However, as explained above, for purposes of this model and analysis flow boiling heat transfer coefficients for pure ammonia will be used.

3. Shift in the bubble point (decrease) due to two-phase pressure drop is linear over the length of the evaporator circuit. The depression of the bubble point over the circuit length due to frictional pressure drop is typically kept small by coil designers (0.5–1°F total). In the model this change in bubble point due to pressure drop is subtracted from the calculated bubble point based on composition to find the current refrigerant temperature for a finite volume.

4. Enthalpy changes in the refrigerant vapor and liquid phases due to shifting bubble point (sensible heat transfer) are negligible over the water concentrations of interest. Thome (Collier 1994) suggests that a temperature-enthalpy curve be used to calculate the change in specific enthalpy with change in the bubble point to account for the sensible heat imparted to the vapor and liquid refrigerant. For the range of mixture compositions being




modeled, the sensible heating component of the enthalpy change is only on the order of 1% of the total and so has been ignored.

It is believed that these assumptions are justified bearing in mind that the purpose of the model is to isolate the thermodynamic effects of shifting bubble point over the entire evaporator circuit length for various types of circuiting arrangements.

Using the segmented model approach also makes it possible to compare the thermodynamic effects of water on different circuiting arrangements; cross flow, counter flow, and parallel flow.

Results

The evaporator model was used to predict evaporator performance and thermodynamic effects of water in ammonia for two typical cases:

1. Constant Refrigerant Flow Rate. Occurs with pumped ammonia evaporators being fed by a fixed displacement liquid pump.

2. Constant Leaving Refrigerant Temperature. Occurs with dry expansion (DX) evaporators operating with expansion valves controlling refrigerant flow based on superheat.

Constant Refrigerant Flow Rate

Evaporators used in pumped ammonia systems can be circuited such that the refrigerant and air flow relative to one another in one of three ways:• Cross flow• Counter flow• Parallel flow




All three of these arrangements are commonly used in the industry and offer manufacturers and operators various benefits. Orientation of headers and connections along with arrangement of finite volumes used in the model are shown in Figures 9 through 11. It has been suggested that parallel flow circuiting produces the best thermal performance (Nelson 1990, Aljuwayhel et al 2007) along with other benefits such as more uniform frost distribution when compared to counter or cross flow circuiting.

The model was used to calculate thermal performance at four different refrigerant pressures with varying initial water concentrations. Coil geometry and operating conditions used in the calculations were as follows:

Tubes: 7/8" OD x 0.065" 3003 Aluminum

Tube Pitch: 60 mm x 60 mm inline

Rows Deep: 8

Fins: 0.016" 1100 Aluminum

Fin Spacing: 3 per inch (8.5 mm)

Air Face Velocity: 700 ft/min (3.6 m/s)

Initial Temperature Difference (TD): 10°F (5.6°C)

Refrigerant mass flow rate was calculated for each pressure based on the pure ammonia (0% water) case and then held constant.

Normally, the degradation in ammonia evaporator performance due to the presence of water is either (a) ignored, or (b) assumed to be proportional to the reduction




in coil TD resulting from the increase in the boiling point at the initial water concentration. This ideal capacity (reduction) ratio is calculated as follows:

qqº

ideal

= (Tair ent – Tbub) (2)

(Tair ent – Tsat)

The ideal and thermodynamic capacity ratios were calculated with the model and are shown in Table 1 based on a saturation pressure of 18.3 psia (1.3 bar).

Figures 13 through 15 show temperature profiles generated by the model at these conditions with 0% water (pure ammonia) for Cross Flow, Counter Flow, and Parallel Flow circuiting. Figures 16 through 18 show temperature profiles for initial water concentration of 5%.

As expected, the model indicates a very slight capacity increase for parallel flow circuiting with pure ammonia over counter flow and cross flow circuiting, which agrees with Nelson (1990) and Aljuwayhel et al (2007). This apparent performance advantage for parallel flow circuiting disappears, however, for the n = 1 case with only small amounts of water present. As the overfeed ratio increases to n = 2, parallel flow circuiting appears to maintain a slight performance advantage over counter flow circuiting even at higher water concentrations (up to 5%). As the water concentration increases to 10%, however, counter flow circuiting begins to outperform cross flow and parallel circuiting, even at the higher overfeed ratios (n = 2).

In all cases as the water concentration exceeds 1%, the degradation in capacity is more severe than predicted by the ideal capacity ratio calculation.

In general the model predicts the following for pumped ammonia:

1. As the overfeed ratio is reduced, the penalty to performance due to the presence of water becomes more severe.




2. The penalty to performance due to the presence of water is always more severe than predicted by an ideal capacity ratio calculation whenever n < 1.5.

3. The penalty to performance due to the presence of water is always more severe than predicted by an ideal capacity ratio calculation whenever initial water concentration is greater than 1%.

In other words, it’s worse than you thought.

Constant Leaving Refrigerant Temperature

In this case, refrigerant temperature leaving the evaporator is held constant by the expansion valve controller. When water is present, the controller “thinks” the refrigerant has been completely evaporated and superheated when, in fact, some fraction of the refrigerant is leaving the coil as liquid. The water concentration at the evaporator exit will correspond to a bubble point temperature equal to the superheat setting of the expansion valve controller. In the case of a 10°F (5.6°C) superheat setting, the composition of the liquid at the exit of the evaporator will be between 0.2 to 0.24 (Figure 4).

The vapor mass fraction of the refrigerant leaving the coil is related to the initial and final composition of the liquid according to Equation 1, which can be rearranged as follows:

xlvg

=

1 –

y' (3)

ylvg

The mass flow rate of refrigerant leaving the coil as liquid can then calculated as follows:

mliq = m (1 – xlvg) (4)




For the DX calculations only counter flow circuiting was considered. This arrangement is shown in Figure 12.

Thermal performance was calculated at four different refrigerant pressures with varying initial water concentrations. Coil geometry and operating conditions used in the calculations were as follows:

Tubes: 7/8" OD x 0.065" 3003 Aluminum

Tube Pitch: 60 mm x 60 mm inline

Rows Deep: 8

Fins: 0.016" 1100 Aluminum

Fin Spacing: 3 per inch (8.5 mm)

Air Face Velocity: 700 ft/min (3.6 m/s)

Initial Temperature Difference (TD): 12°F (6.7°C)

Expansion Valve Superheat Setting: 10°F (5.6°C)

Entering Vapor Mass Fraction (quality): 0.15

Leaving refrigerant temperature was held constant and refrigerant mass flow rate allowed to vary to maintain constant entering vapor mass fraction (quality).

Calculated ideal and thermodynamic capacity ratios are shown in Table 2. Figures 19 through 22 show typical temperature profiles at various initial water concentrations for a saturation pressure of 18.3 psia (1.3 bar).




As with the pumped ammonia calculations at n = 1, the DX calculations indicate thermodynamic capacity ratios which are significantly lower than the ideal capacity ratios. As water concentration increases, the difference between the ideal capacity ratio and the thermodynamic capacity ratio becomes larger. In other words, using an ideal capacity ratio assumption will result in significantly underestimating the performance penalty that results from water concentrations greater than about 1%.

It is also apparent from the data in Table 2 that performance penalties become more severe as saturation pressure increases. So, water removal is critically important at all temperatures when operating DX and low overfeed ratio ammonia systems.

Water Removal

From the preceding discussion, it is apparent that proper operation of direct expansion and low overfeed (n < 2) ratio systems depends upon effective removal of water from the ammonia. Regular monitoring and measurement of water is particularly important with systems operating at suction pressures below one atmosphere due to the increased risk of undetected ingress of air (with its accompanying water vapor).

Unlike halocarbon refrigerants, water cannot be removed from ammonia by desiccant type driers. The only effective means of removing water from an ammonia refrigeration system is by capturing the ammonia-water liquid as it leaves the evaporator, then heating to drive off enough ammonia so that the water-rich liquid can be drained and disposed of. Fortunately, a number of ammonia removal devices are commercially available which can easily be installed in new systems or retrofitted to existing systems.

The phase equilibrium diagram (Figure 1) is useful in determining the maximum temperature to which ammonia-water liquid can be heated without generating significant amounts of water vapor. The “knee” to the left of the dew point curve(s)




in Figure 1 indicates the temperature at which water vapor will begin to leave the ammonia-water solution for a given pressure. At pressures typical of freezer operations (10.4–18.3 psia) the dew point curves of Figure 1 imply that heating the captured ammonia-water solution much above 120°F (49°C) will generate increasing amounts of water vapor, which will recombine with the ammonia if vented back to the suction side of the system. The maximum possible composition (mass fraction water) in the ammonia-water liquid heated to a given temperature can be read directly from the bubble point line at the operating pressure. For example if ammonia-water at 10.4 psia (0.7 bar) were heated to 120°F (49°C), the composition of the liquid will be 0.85 water (mass fraction). From the dew point curve for this same example, composition of water in the coexisting vapor would be less than 0.05.

Disposal of the concentrated ammonia-water solution can be challenging depending on limitations placed on allowable ammonia concentrations and pH by local environmental jurisdictions. Unfortunately, only very small amounts of ammonia in water result in a pH in excess of 10.0. The volume of water required to dilute an ammonia-water solution which has been concentrated by heating (i.e., water compositions of 0.75 to 0.85) to a pH acceptable for disposal, therefore, may be impractically large (Stoecker 1998). The operator may then have to rely on neutralizing the concentrated ammonia-water solution chemically prior to disposal.

Conclusions

A computer model has been written to examine and quantify the thermodynamic effects of small amounts of water in ammonia on evaporator performance. Based on the results and discussion above, the following conclusions can be drawn:

1. Water in ammonia behaves as a zeotropic refrigerant mixture.

2. Composition shift occurs in the ammonia-water liquid resulting in increasing bubble point temperature over the length of evaporator circuit.




3. The performance penalty due to this shift in bubble point temperature is significantly greater than the penalty normally associated with the increase in bubble point at the initial water concentration (the ideal capacity ratio).

4. The performance penalty increases as the overfeed ratio is reduced (n < 2). Conversely, the performance penalty due to water becomes smaller, i.e., is ‘masked,’ as overfeed ratio increases (n > 2).

5. The performance penalty due to water increases as operating pressure (suction temperature) increases.

6. Whenever water is present in ammonia, it becomes impossible for the refrigerant to become superheated at the exit of a DX evaporator circuit. Consequently, some fraction of the refrigerant will always leave the evaporator as liquid.

7. Provisions should always be made to capture and remove water from ammonia refrigeration systems, especially for operation with DX and low overfeed ratios (n < 2).

8. Based on the phase equilibrium diagram for ammonia-water, the maximum water mass fraction in the liquid (composition) that can practically be reached by simple heating is approximately 0.75 to 0.85. This implies that further treatment of the concentrated ammonia-water liquid prior to discharge to the sewer is likely to be needed.




Nomenclature

m = total mass flow rate of refrigerant

mliq = mass flow rate of liquid refrigerant leaving the evaporator

n = liquid overfeed ratio

q = evaporator cooling capacity with water present in ammonia

qº = evaporator cooling capacity with pure ammonia

Tair ent = entering air dry bulb temperature

Tbub = bubble point temperature

Tsat = saturation temperature at the evaporator outlet pressure

x = vapor mass fraction (quality)

xlvg = vapor mass fraction (quality) leaving the evaporator

y = composition (mass fraction water) of refrigerant

y' = initial composition (mass fraction water) of refrigerant




Bibliography

Aljuwayhel N.F., Reindl D.T. and Klein S.A. (2007), “Comparison of parallel and counterflow circuiting in an industrial evaporator under frosting conditions,” International Journal of Refrigeration, Vol. 30, No. 8, pp. 1347-1357.

Chen J.C. (1963), “A Correlation for Boiling Heat Transfer to Saturated Fluids in Convective Flow.,” American Society of Mechanical Engineers (ASME) Paper 63-HT-34.

Collier J.G, and Thome J.R. (1994), “Convective Boiling and Condensation, Third Edition,” Oxford University Press, New York, Chapter 12:Multi-component Boiling and Condensation.

Cotter D. et al (2007), “Contaminants in Ammonia Refrigeration Systems,” International Institute of Refrigeration (IIR), Conference Proceedings, Ohrid, Macedonia.

Ficker R., “Water Contamination and Water Removal in Industrial Ammonia Refrigeration Systems.” Hansen Technologies Corporation, Technical Bulletin

IIAR Bulletin No. 108, “Water Contamination in Ammonia Refrigeration Systems,” International Institute of Ammonia Refrigeration (IIAR), Arlington, VA.

Nelson B.I. (1990), “Design of Evaporators for Liquid Overfeed Systems,” ASHRAE Transactions, Vol, 96, no. 1, pp. 1309-1315.

Nelson B.I. (1998), “Designing Air Coolers for Direct Expansion with Ammonia,” International Institute of Ammonia Refrigeration (IIAR), 20th Annual Meeting, Colo. Springs, CO, pp. 137-155.




Nielsen P.S. (1998), “Effects of Water Contamination in Ammonia Plants,” International Institute of Ammonia Refrigeration (IIAR), 20th Annual Meeting, Colo. Springs, CO, pp 89-103.

NIST (2007), Lemmon E.W., Huber M.L. and McLinden M.O., Standard Reference Database 23 Version 8.0 (REFPROP), National Institute of Standards and Technology (NIST), Gaithersburg, MD.

Stoecker W.F. (1998), Industrial Refrigeration Handbook, McGraw-Hill, Chap. 13, pp. 461-463

Taboas F. et al (2007), “Pool boiling of ammonia/water and its pure components: Comparison of experimental data in the literature with the predictions of standard correlations,” International Journal of Refrigeration, 30(2007), pp. 778-788.

Thome J. (2006), Engineering Data Book III, Chapter 15: “Thermodynamics of Refrigerant Mixtures and Refrigerant-Oil Mixtures,” Wolverine Tube, Inc.

Zurcher O., Thome J.R. and Favrat D. (1999), “Evaporation of Ammonia in a Smooth Horizontal Tube: Heat Transfer Measurements and Predictions,” Journal of Heat Transfer, Vol. 121, February, pp. 89-101.




Figure 1: Phase Equilibrium Diagram (Ammonia-Water) Temperature vs Composition

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Figure 3: Latent Heat of Vaporization vs Composition Ammonia-Water

Figure 4: Increase in Bubble Point Temperature vs Composition Ammonia-Water

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*+,%&-.&'"+'B$008&'C4"+9'/&36&%-9$%&D'!

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*+,%&-.&'"+'B$008&'C4"+9'/&36&%-9$%&D'!




Figure 9: Cross Flow Circuiting, Pumped Ammonia

!"#$%&'()'*%+,,'!-+.'*"%/$"0"1#2'3$45&6'744+1"8'

Figure 10: Counter Flow Circuiting, Pumped Ammonia

!"#$%&'()*'+,$-.&%'!/,0'+"%1$"."-#2'3$45&6'744,-"8'

Figure 11: Parallel Flow Circuiting, Pumped Ammonia

!"#$%&'(()'*+%+,,&,'!,-.'/"%0$"1"2#3'*$45&6'744-2"+'




Figure 12: Counter Flow Circuiting, DX Ammonia

!"#$%&'()*'+,$-.&%'!/,0'+"%1$"."-#2'34'566,-"7'

Figure 13: Cross Flow Temperature Profile Pumped Ammonia, n = 1, 0% Water

!"#$%&'()*'+%,--'!.,/'0&12&%34$%&'5%,6".&

5$12&7'811,9"3:'9;(:'<='>34&%

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$ $# "# +# *# )# (# '# &# %# $##

?&#1&94

0&12&%34$%&:'!

,-./

012,343




Figure 14: Counter Flow Temperature Profile Pumped Ammonia, n = 1, 0% Water

Figure 15: Parallel Flow Temperature Profile Pumped Ammonia, n = 1, 0% Water

!"#$%&'()*'+,$-.&%'!/,0'1&23&%4.$%&'5%,6"/&

5$23&7'822,-"49'-:(9';<'=4.&%

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0&12&%,3$%&8'!

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,01

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Figure 16: Cross Flow Temperature Profile Pumped Ammonia, n = 1, 5% Water

Figure 17: Counter Flow Temperature Profile Pumped Ammonia, n = 1, 5% Water

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5$12&7'811,9"3:'9;(:'<='>34&%

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Figure 18: Parallel Flow Temperature Profile Pumped Ammonia, n = 1, 5% Water

Figure 19: Temperature Profile DX Ammonia, Pure Ammonia, 18.3 psia (1.3 bar)

!"#$%&'()*'+,%,--&-'!-./'0&12&%,3$%&'+%.4"-&

+$12&5'611.7",8'79(8':;'<,3&%

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@1A

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0




Figure 20: Temperature Profile DX Ammonia, 1% Water, 18.3 psia (1.3 bar)


!"#$%&'()*'+&,-&%./$%&'0%12"3&

45'6,,17".8'9:';./&%8'9<=>'-?".'@9=>'A.%B

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/




Table 1: Thermodynamic Capacity Ratio vs Initial Water Content Pumped Ammonia, 10F TD, 18.28 psia

Initial

Water

Content, %

Tbub,

F

Ideal

Capacity

Ratio

Cross Flow

Thermodynamic

Capacity Ratio

Counter Flow

Thermodynamic

Capacity Ratio

Parallel Flow

Thermodynamic

Capacity Ratio

n=1 n=1.5 n=2 n=1 n=1.5 n=2 n=1 n=1.5 n=2

0 –20.0 1.00 1.00 1.00 1.01 1.00 1.00 1.00 1.01 1.01 1.01

1 –19.6 0.96 0.92 0.96 0.97 0.93 0.96 0.96 0.90 0.96 0.97

5 –18.5 0.85 0.69 0.76 0.78 0.71 0.77 0.79 0.66 0.74 0.78

10 –16.1 0.61 0.45 0.51 0.54 0.47 0.52 0.55 0.43 0.49 0.53

Table 2: Thermodynamic Capacity Ratio vs Initial Water Content DX Ammonia, 12F TD

Refrigerant Pressure

48.2 psia (3.3 bar) 30.4 psia (2.1 bar) 18.3 psia (1.3 bar) 10.4 psia (0.7 bar)

Initial

Water

Content,

%

Tbub,

F

Ideal

Capacity

Ratio

Thermo

Capacity

Ratio

Tbub,

F

Ideal

Capacity

Ratio

Thermo

Capacity

Ratio

Tbub,

F

Ideal

Capacity

Ratio

Thermo

Capacity

Ratio

Tbub,

F

Ideal

Capacity

Ratio

Thermo

Capacity

Ratio

0 20.0 1.00 1.00 0.0 1.00 1.00 –20.0 1.00 1.00 –40.0 1.00 1.00

1 20.5 0.96 0.91 0.4 0.97 0.93 –19.6 0.97 0.96 –39.7 0.97 0.96

5 22.3 0.81 0.61 2.1 0.83 0.64 –18.5 0.88 0.67 –38.3 0.86 0.68

10 24.7 0.61 0.38 4.3 0.64 0.41 –16.1 0.68 0.44 –36.5 0.71 0.46




Notes:


32nd Annual Meeting International Institute of Ammonia...

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