This paper presents an overview of the reasons for charge reduction in air conditioning and refrigeration systems and discusses strategies for charge reduction: in compressors (oil), vessels, pipes, and heat exchangers. The focus is on heat exchangers, microchannel in particular. In addition to a trivial reduction of internal volume as a strategy for charge reduction, the effect of mass flux on void fraction and needed manipulation of circuiting is presented. A framework and example of comparison between refrigerants based on their potential for low condenser charge is provided.


This paper presents the issues related to refrigerant charge reduction in following parts:

  1. Reasons for charge reduction in general
  2. Strategies, specifically in heat exchanger, microchannel in particular
  3. Evaluation of potential for charge reduction in various refrigerants
  4. Examples of extremely low charged ammonia chillers


Reduction of refrigerant charge is attractive for any refrigerant for several reasons:

  • Low charge may expand the possibilities for some very good refrigerants (based on high cycle efficiency, high heat transfer/pressure drop performance, etc.) in areas and locations where either these fluids are totally restricted or only allowed in limited quantities due to flammability or toxicity issues. This is particularly true for fluids like ammonia and hydrocarbons which in some applications and locations are already accepted below certain levels (typically 150 kg for NH3 and 150 or 50 g for HCs).
  • System efficiency could be increased in cycling when refrigerant migration in off periods is significant.
  • Direct environmental effects (limited importance for ammonia or hydrocarbons) are reduced: ozone depleting issues if any, global warming issues, etc… especially in the case of catastrophic leaks.
  • Reduction of the first costs for refrigerant and lubricant expenses (not very important for ammonia or hydrocarbons since they are very inexpensive).


Reducing charge beyond some point could result in problems in operation such as:

  • If receiver is inadequate, then small leaks will result in insufficient charge thus reducing performance.
  • If charge reduction is achieved by reducing the hydraulic diameter of pipes too severely, then pressure drops and flashing may affect efficiency and perhaps capacity.
  • Reduced reliability of pumps (if present) due to possible vapor entrainment and cavitation.


  1. The first and almost trivial options for charge reduction is to add a loop with another fluid in the refrigeration system and in that way reduce the primary refrigerant charge. These options could be either cascades or secondary refrigerant loops which actually turn a refrigeration system into a chiller with either single or two phase secondary refrigerants. These options were, and are, very popular for their simplicity. Refrigerant in the primary (or only) system is in each component: compressor, vessel(s), liquid, two phase and suction lines, and finally, heat exchangers.
  2. Reduction of liquid quantities in vessels can be achieved by reducing their volume and liquid levels and sometimes orientation.
  3. Reducing refrigerant charge in compressor by a) reducing the internal volume b) reducing the quantity of lubricant and c) reducing refrigerant solubility (if reasonable). Needless to say, reduction of the volume on the high side yields greater results than on the suction due to density difference. A part of lubricant charge in the compressor is a function of refrigerant charge in the system to allow for oil retention in the system.
  4. Reduction of the refrigerant charge in pipes is typically achieved by reducing internal volume (diameter and possibly even length). Reduction of diameter results in increased pressure drops. Pressure drops in liquid lines do not affect the system performance but can cause nondesirable flashing which affects the operation of flow controllers. Pressure drops in suction lines affect system efficiency and therefore may not be advisable.
  5. The most attractive and important strategy is to reduce the charge in heat exchangers and that will be our main focus. Reduction of refrigerant charge in heat exchangers is always related to a reduction in internal volume. One must be careful to balance the adverse effects regarding either increased pressure drop or reduced heat transferred.


Almost every refrigerant inventory reduction strategy is related to reduction of internal volume. We should not forget that the internal volume of heat exchangers is typically determined in the desing phase based on the heat transfer area needed for the main purpose: heat transfer, pipe size and circuiting based on the target pressure drop. It is a great advantage that reduction of the volume is proportional to the channel (pipe) diameter squared while the heat transfer area is a linear function of the diameter. Typically consequence of the diameter reductoion is pressure increase that can me mitigated by increasing number of parallel circuits. Nevertheless, reduction of the internal volume is just one, relatively trivial, path to reaching the objective.

In the cases of spray evaporators, significant charge reductions were achieved compared to shell and tube even for the same volume.

To provide sufficient capacity with in-tube flow heat exchangers, it is necessary to provide a refrigerant flow that is inversely related to the latent heat of vaporization (hfg). Therefore, refrigerants with larger values latent heats will have lower mass flow rates for the same capacity and all else being equal, consequence is lower pressure drop compared to those that require higher flow rate. These lower pressure drops may then allow for smaller diameter tubes to be utilized. We should not forget that ammonia has very high latent heat of vaporization compared to other refrigerats but also very light vapor that increases in tube velocity for the same mass flux.

Heat transfer effects of the charge reduction strategies are related to refrigerant side heat transfer. Higher heat transfer coefficients mean that smaller surfaces are required for a given duty. The effect of refrigerant heat transfer coefficient is greater for heat exchangers where the refrigerant side is a relatively large component of the total heat transfer resistance. As an example, charge is easier to be minimized in heat exchangers like water cooled condensers or chiller evaporators where water, rather than air, is heated or cooled. This is because a greater heat transfer area is required for a given capacity in the case of a heat exchanger with air due to its lower overall heat transfer coefficient.

In addition, those refrigerants that allow flow regimes that are better for heat transfer (annular or intermittent in evaporation or mist in condensation) or geometries that stimulate such flow regimes (i.e. microfins) will provide better opportunities for change reduction. Whether these opportunities will be realized is a function the adequacy of the design.

As already said, every reduction of internal diameter, with everything else being the same will increase pressure drop. The option to reduce the pressure drop in the same diameter tube with the same fluid is to reduce the length of the tube or decrease flow through the tube. This drives heat exchanger design towards parallel flows of small diameter tubes with the shortest possible circuits. That indicates a parallel heat exchanger design with single pass as an asymptote.

Nevertheless, the reality is more complex than this straight forward direction. The main goal should be to optimize system COP for a given mass flow by working on the balance between heat transfer and pressure drop. A fine discussion of this subject was given by Cavallini, 2011.

When calculating the refrigerant charge in a heat exchanger, engineers typically use void fraction, c (c=Vapor volume/(Vapor+Liquid Volumes)) as a main variable. Figure 1 presents the relationship between R134a and void fraction in a microchannel tube (Nino et. al. 2002).

So, the mass of liquid Ml can be calculated as: 

and total mass is calculated as:

Void fraction is a misnomer. There are no voids in any heat exchanger. Better name would be “vapor volume fraction”. Even more, we are interested not in vapor but instead on liquid because liquid carries the mass of refrigerant which we are trying to reduce:

It is very important to have in mind that the void fraction is a function of mass flux (this fact even some void fraction correlations neglect). This relationship is a very important tool when designing to reduce refrigerant charge. The trend is the same as discussed above: reduction in diameter increases mass flux and increases void fraction, consequently reducing charge. This trend was shown in the graph shown in Figure 1.

There is a practical difficulty when following the strategy described above. When a higher capacity of a heat exchanger is desired, the number of parallel passes increases; consequently the headers where channel flows will merge are becoming larger. Flow regimes in headers are less well described and much less predictable, and typically void fractions are lower in header than in tubes.

Probably the best known low charge heat exchangers based on reduction of internal volume are of the microchannel type. They offer extremely low charge (less than 10 g/kW even for air on the other side), unless the headers are designed incorrectly. Hrnjak and Litch (2001) reported charges of 18 g/kW for a 15 kW air cooled microchannel ammonia chiller with a condenser charge of 6 g/kW. Traeger and Hrnjak (2005) reported an R290 system with 8-10 g of charge in a 1kW serpentine microchannel evaporator. There are some microchannel heat exchanger designs with water as the fluid on the other side (see Palm, 2009).

Other effective low charge heat exchangers are the plate type. They are better known in chiller applications and are typically made of stainless steel which is either brazed with copper or nickel or else are gasketed. Similar variations are becoming more common in plate and shell designs. Plate heat exchangers with air on the other side are predominantly used in automotive a/c systems as evaporators. Typical refrigerant charges in larger heat exchangers (Pearson, 2003) are 1kg/kW, 0.5kg/kW, and 0.25kg/kW for shell and tube, plate, and gravity fed plate heat exchangers, respectively.

Z. Ayub, (1996) reports on low charges (54 g/kW) in large spray evaporators for ammonia (4MW).


Even though void fraction has been studied extensively [see Zivi, (1964), Butterworth, (1975), Newell, (1999), Adams et al., (2003) & (2006), etc.] mass flux effects were not always considered. Typically, for a given fluid and local quality, increase of the mass or heat flux affect flow regimes in a way that increases void fraction thus reducing charge.

Nino, Hrnjak, and Newell, (2002) showed the effect of mass flux on the void fraction for R134a in microchannel tubes (Figure 1), while Adams, Hrnjak, and Newell, (2003) shed some light on void fraction for ammonia (shown in Figure 2 and 3), carbon dioxide, and R245fa in microchannels. Figures 2 and 3 present void fraction as a function of quality for three mass fluxes, along with correlation predictions from the homogenous and Nino et al., (2002) models. It is obvious, despite some scattering of experimental data that higher mass fluxes result in higher void fractions (reduced refrigerant charges). Photos in Figures 4, 5, and 6 [from Niño, Hrnjak, & Newell, (2002)] provide visual evidence of the same trend, in transparent microchannels, for R134a, R410A, and air-water (whose fluid flow resembles ammonia). The same conclusion holds for tubes of larger diameter even not presented here.

It is clear from experimental data and the photographs that increase in mass flux reduces the charge (increase void fraction).


As discussed previously, it is clear that low charge is beneficial for every refrigerant in every application.

Comparisons of refrigerants based on their efficiency are common. The basis for efficiency comparison is cycle analysis (typically Rankine or EvansPerkins) using thermophysical properties. Cycle analysis involves significant simplification that could be even misleading in some cases. It does not include realities of the heat exchanger (heat transfer, pressure drops, power for air or water movers, etc.). That approach is based on thermodynamics and is accepted as a first approximation in the comparative analysis.

It would be interesting to know whether refrigerants can be compared based on their potential for charge reduction. We will try to present here the concept to compare refrigerants based on their potential for charge reduction in heat exchangers and illustrate it with an example. The concept is in principle based on assumption that refrigerant which is less sensitive to reduction of internal volume and has less (high void fraction) and lighter (lower density) liquid has greater potential for charge reduction. The sensitivity to reduction of internal volume is defined by reduction of COP due to pressure drop. Pressure drop is a function mostly of mass flow rate determined by (available) portion of latent heat, specific volume of vapor, and viscosity of liquid.


As said above, the most logical starting point for reducing charge in heat exchangers begins with the reduction of internal volume. A result of reducing the pipe diameter is increased pressure drop. The fair comparison of refrigerants should be based not on equal pressure drop itself but on the effect pressure drop has on efficiency (COP) while maintaining the same capacity.

The heat transfer on both sides (not only on refrigerant that is mostly affected when varying design options for charge reduction) is important and affects the result. That is the reason why maintaining identical air side is important. To maintain identical outside (air-side) conditions for each refrigerant, the heat exchanger type selected is the microchannel serpentine (two circuits) shown in the Figure 7 and with dimensions given in Table 1. Selection of microchannel type is due to its relevance today and for its anticipated acceptance by more manufacturers. In addition, high heat transfer coefficients and fin enhancement place this type of heat exchanger between conventional round tube plate fin types and water cooled or cooling types, providing a more universal basis for conclusion. A serpentine design was selected to avoid uncertainties in predicting the charge in headers for various refrigerants. The serpentine design does not reduce the generality of the conclusions or even limit the increase in mass flux. Figure 7: Baseline serpentine condenser design

For the example presented here, the internal diameter of the channel was the only variable in the model used to adjust refrigerant pressure drop. In the experiments it was easier to vary the number of active channels and the model was adjusted to facilitate validation.

To estimate the condenser charge required for different refrigerants, the internal volumes (diameters) of identical baseline heat exchangers are shrunk to create a pressure drop which causes an equal (here selected 1%) reduction in COP. Certainly, 1% is arbitrary value, but it is equal for each refrigerant. One could question if equal penalty is fair, Wujek, (2012). More about this subject may be found in Hrnjak (2009) and Padilla and Hrnjak (2012).

The adopted model takes into account the effect of the ratio between the refrigerant and air side heat transfer areas when iterating to find the solution.

A fair comparison of minimum refrigerant charges requires maintaining the system capacity while modeling heat exchangers with the same face area, exterior tube dimensions, identical fins, and same operating conditions on the air side (velocity, temperature, humidity). Additionally, the effect of the condenser on the rest of the refrigerant side of the system should be the same. Here it is defined as a 1% difference between COPs of the system with a real condenser and one without pressure drop on the refrigerant side. The condenser channel diameter is modified to generate the same degradation of COP due to refrigerant side pressure drop compared to an ideal condenser (without pressure drop) while maintaining the same evaporator capacity. A similar option would be to vary the number of active channels while maintaining the channel diameter and the outer dimensions of the flat tube. This approach has been tested experimentally in Hoehne and Hrnjak, (2004).

Fig. 8 compares the “ideal” baseline cycle, in solid line, to the “real” cycle, with pressure drop, shown with the dashed line. The pressure drop in the “real” condenser is set to cause a 1% reduction in COP compared to “ideal” cycle. Isenthalpic expansion and isentropic compression are assumed in the ideal cycle for all fluids. The Zivi void fraction model was used because it is refrigerant independent but consequently the effect of mass flux was not reflected. These assumptions do not affect the generality of conclusions. Cycle operating conditions in this example are: Tevaporation=0o C, Tairin=20o C, superheat 5 K. Based on the evaporator and condenser models, the necessary charge to achieve 1 kW cooling capacity for various refrigerants is evident from the Fig. 9. Additionally, the “real” hydraulic diameter of the condenser to minimize charge with only 1% degradation in COP can be found. Ammonia requires the lowest charge but not the smallest tube size. Isobutane requires a much larger tube diameter.

R717 and R744 show the best potential for low charge systems but for different reasons. R717 is known to have a very high pressure drop for a given mass flux because of very low vapor density (see second to last column in Table 2) which causes higher velocity for a given mass flow rate in comparison to other fluids. However, due to its very high latent heat (hfg = 1167 kJ/kg), the mass flow needed for the same capacity is significantly lower for R717 than for any other fluid. The sensitivity of ammonia to pressure drop is neither exceptionally low nor high in comparison to the other fluids shown in Table 2 (column 5, ∆P). Since vapor density is very low for ammonia, the total mass is the lowest for a given void fraction. Light vapor is helpful in building a low charge system.

Carbon dioxide (R744) has different characteristics than ammonia. R744 has a low sensitivity to pressure drop, which means that high pressure drop will not result in high temperature drop. The sensitivity to pressure drop is given in Table 2, Column 5. The potential for building a low charge R744 system comes from having low pressure drop (due to dense vapor) and very low sensitivity to pressure drop. The small channel sizes, dense vapor, and low sensitivity to pressure drop indicated that microchannel heat exchangers are ideal for R744. Because R744 has a dense vapor, it will have more refrigerant mass at a given void fraction and internal volume. The largest hydraulic diameter is required for R600a (Isobutane) mostly because of the lighter vapor (second to ammonia). The combination of both light vapor and liquid keeps the charge reasonably low even with the largest diameter.

The two refrigerants that require the highest charge are R1234yf and R134a even though their individual thermophysical properties are balanced. High density of liquid combined with high sensitivity to pressure drop requires significantly greater charge than other fluids presented here.


Recent advances in manufacturing technologies of microchannel tubes and heat exchangers resulted in expansion of some important mass production markets and consequently opened opportunity for further reduction of costs. That situation generates the possibilities for application of microchannel heat exchangers in areas with traditionally lower production volumes, ammonia being one of them.

Litch and Hrnjak [9] presented data for an ammonia chiller with an air cooled microchannel condenser. This resulted in the lowest specific charge air-cooled chiller for ammonia reported in the literature so far.

Two aluminum condensers were evaluated: one with a single serpentine tube and the other with a parallel tube arrangement between headers having 24 tubes in the first pass and 14 in the second. Each tube has 19 triangular ports of equal dimensions with a hydraulic diameter less than 1 mm. The fins are multi-louvered. The serpentine condenser has a single tube that passes 16 times through multilouvered fins. There are five enhanced square ports in the tube. Additional details of these condensers may be found in Litch & Hrnjak [9].

Overall heat transfer performance and charge measurements were taken for each condenser and the system as a whole. The microchannel heat exchanger with parallel flow performed better in every respect. Overall, condenser performance was quantified in terms of U values for different air flow rates, superheating and subcooling conditions and is presented in Figure 10. Refrigerant inventory measurements of the condenser were taken at different operating conditions.

Refrigerant inventory measurements are compared to model results using different void fraction model predictions. All void fraction correlations perform similarly in helping to predict total charge. The use of Newell’s correlation (Newell et al. [10]) for serpentine condenser yields the smallest average error of 9.3%, with a maximum of 15.7%. With the Butterworth [5] and Zivi [14] correlations, the average and maximum errors are 10.1/22.8% and 12.3/24.9%, respectively. The slight over prediction results in a simulated subcooled region that is larger than the actual region, inflating predicted charge. Data by Adams, Hrnjak and Newell [1] fit well in the prediction. These results are presented in Figure 11.

Obviously, predictions for a serpentine condenser are much more accurate than for microchannel when using the same correlation and experimental data. That clearly indicated a significant inaccuracy in the charge prediction in headers (see Figure 12.). Another insight from Figure 7. (serpentine) is that liquid subcooling is a large contributor to total charge. The relative predicted charge contributions from the refrigerant phase zones for the data point with the highest liquid subcooling tested are 0.5% in superheated zone, 29.2% in the two-phase region, and 70.3% in subcooling region. From the data point with the lowest liquid subcooling, the contributions are 0.5%, 60.1%, and 39.4% in subcooling. Even though the subcooling region is only 26% of the total tube length, it comprises 70% of the total charge. Thus it is advantageous to reduce subcooling not only for increased heat transfer, but to reduce refrigerant charge.

From the experimental data taken, the microchannel parallel flow condenser appears to outperform the macrochannel serpentine condenser. The overall heat transfer coefficient for a given face velocity is 60-80% higher than for the serpentine condenser; and the charge is an average of 53% less. The microchannel condenser has a smaller volume for approximately the same face area. Also, it has less charge and better heat transfer than the serpentine and typical condensers.


Probably the most important recent development is the new Mycom hermetic compressor that is used for both refrigeration and heat pumping. The wrap is specifically designed for use with ammonia. Nominal capacity in cooling (at -5oC/50oC) is 45 kW while in heat pumping is 47 kW. Motor is IPM type with aluminum windings. There are two models: for low and high temperature. The weight of the hermetic version is about 100 kg. This compressor is equipped with an oil pump. The ammonia charge of the unit is 6 kg (see Table 3 and Figures 13 and 14).

Development of microchannel condensers for ammonia has moved from the Air Conditioning and Refrigeration Center at the University of Illinois to Creative Thermal Solutions (CTS), a high tech company that specializes in research and development of novel refrigeration and air conditioning approaches. Figure 15 presents a photo of a condensing unit with microchannel heat exchanger used in an experimental facility for evaluation of ammonia evaporators, while Figure 16 shows a unit from Figure 9, CTS instrumented for implementation of MC condensers. The MC condensers used improved performance with 87% face area of original round tube condenser, with only 19% of core volume and just 7% or original weight and 27% or original refrigerant volume – charge. Just few weeks ago M. Tomooka of Mycom presented these results in the paper at IIAR meeting in Orlando, FL “Application of Microchannel heat exchangers to compact ammonia systems”

Another very good example is presented by Cecchinato and others [23] who described the main features of the newly designed prototype, including:

  • refrigerating capacity of 120 kW
  • open inverter-driven screw compressor with nominal volumetric flow rate equal to 118 m3/h
  • evaporation and condensation temperatures of 2°C and 50°C respectively
  • temperature of the secondary refrigerant (water) at the evaporator outlet was set at 7°C and at the evaporator inlet at 12°C
  • plate heat exchanger evaporator with 52 plates having high chevron angle with overall dimensions equal to 618×191 mm

The chiller use low internal volume heat exchangers and the direct expansion evaporator providing the charge of 10.0 kg of ammonia. Experimental results showed COP of 5.0 to 2.7 at ambient temperatures from 10 to 40°C. The authors estimated potential for a charge reduction of 20% if microchannel condenser would be used.


At this point of heat exchanger development, the lowest charges have been achieved by using a microchannel approach and will be presented later in more detail. Nevertheless, microchannel technology is not the only way to reduce charge. Very good results have been achieved by using plate evaporators or condensers with water, or other fluids, on the other side (brazed, gasketed, cassette, welded shell and plate, etc…). The automotive industry has developed plate evaporators for air cooling, but the application is still limited to mobile air conditioning (aero, automobiles, off-road vehicles etc…). Spray evaporators are also known for their low charge. It should not be forgotten that in microchannel heat exchangers significant liquid quantity is retained in the headers.

Typical values for refrigerant inventory in larger heat exchangers as given by Pearson [13] are shown in Table 4. Ayub reports about low charges in spray evaporators and recent improvements, shown in Table 5.

Pearson reports that “optimal charge” chiller had 100g/kW charge. The optimal value had an unspecified additional charge for leakage and operation.

Litch and Hrnjak presented data for some small ammonia systems with published charges in Table 6.

B. Palm in the summary of a decade of charge reduction at KTH presented a small ammonia system (a laboratory setup simulating a domestic water to water heat pump) as a part of Sherhpa project. Their largest challenge was to get the oil back to the compressor in the direct expansion system so they used miscible oil and a heat exchanger with narrow channels. The same special aluminum heat exchangers were used as condenser and evaporator. Plate heat exchangers were also tested and performed well as condensers but not as evaporators due to problems with oil return. The system with an open compressor had 9kW capacity with 100g of charge (an amazing 11 g/kW, half of the charge of ILKA MAFA 100.2-11K45).


This paper presented reasons for charge reduction and strategies for the actions:

  1. Introduction of another (secondary) fluid
  2. In the compressor:
    • Reducing the internal volume
    • Reducing quantity in lubricant
    • Reducing solubility to reduce refrigerant absorption.
  3. In vessels by reducing volume and liquid levels.
  4. In pipes by reducing internal volume (diameter and possibly even length).
  5. In heat exchangers by reducing tube diameter, length and most importantly balance mass flux and design

Because it is the most attractive, the last strategy is elaborated in detail. Obviously, to reduce the charge internal volume needs to be reduced, but the most important is to take in consideration effects of heat transfer balance and mass flux on void fraction.

Special attention is given to explore and design a methodology for fair way of comparing refrigerants based on their potential to be used in low charge condensers.

This paper also presented a case for a small, low charge, air cooled ammonia chiller using microchannel condensers and hermetic compressor with miscible oil. Microchannel air cooled condensers along with DX plate or similar evaporators provided basis for low charge.

In addition, the external volume of the chiller could be reduced because the external volume of a microchannel design is small. Besides being compact, microchannel heat exchangers are also made lightweight, from aluminum. Thanks for the technology developed in the automobile industry these exchangers are relatively inexpensive.

Expanding the use of aluminum beyond MC condenser it is possible to reduce the weight and the cost even further, making the chillers cost competitive to conventional systems.

Hermetic compressor with miscible oil provides a low leak and a low maintenance environment and in every respect similarity to conventional chillers.

Since ammonia is one of few refrigerants that have vapor lighter than air location of the chiller should be on the roof. Assuming unobstructed release, even in the worst case scenario of catastrophic leak refrigerant vapor cannot increase its concentration in specific zones beyond LFL (lower flammability level) or toxic concentrations values. That represents great improvement in safety and puts ammonia below the radar of regulations.

All said above leads to an excellent opportunity for the ammonia as a refrigerant in urban areas: very low charged, hermetic chiller placed on the roof with unobstructed vapor release.


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