Refrigeration is the process of moving heat from one location to another by use of refrigerant in a closed cycle. Oil management; gas and liquid separation; subcooling, superheating, desuperheating, and piping of refrigerant liquid, gas, and two-phase flow are all part of refrigeration. Applications include air conditioning, commercial refrigeration, and industrial refrigeration. This chapter focuses on systems that use halocarbons (halogenated hydrocarbons) as refrigerants. The most commonly used halogen refrigerants are chlorine (Cl) and fluorine (F).
Halocarbon refrigerants are classified into four groups: chloro-fluorocarbons (CFCs), which contain carbon, chlorine, and fluorine; hydrochlorofluorocarbons (HCFCs), which consist of carbon, hydrogen, chlorine, and fluorine; hydrofluorocarbons (HFCs), which contain carbon, hydrogen, and fluorine; and hydrofluoroolefins (HFOs), which are HFC refrigerants derived from an alkene (olefin; i.e., an unsaturated compound having at least one carbon-to-carbon double bond). Examples of these refrigerants can be found in Chapter 29 of the 2017 ASHRAE Handbook—Fundamentals.
Desired characteristics of a halocarbon refrigeration system may include
Year-round operation, regardless of outdoor ambient conditions
Possible wide load variations (0 to 100% capacity) during short periods without serious disruption of the required temperature levels
Frost control for continuous-performance applications
Oil management for different refrigerants under varying load and temperature conditions
A wide choice of heat exchange methods (e.g., dry expansion, liquid overfeed, or flooded feed of the refrigerants) and use of secondary coolants such as salt brine, alcohol, glycol, and carbon dioxide.
System efficiency, maintainability, and operating simplicity
Operating pressures and pressure ratios that might require multistaging, cascading, etc.
A successful refrigeration system depends on good piping design and an understanding of the required accessories. This chapter covers the fundamentals of piping and system design as well as guidance on new design considerations in light of increasing regulation on halocarbon refrigeration systems. Hydrocarbon refrigerant pipe friction data can be found in petroleum industry handbooks. Use the refrigerant properties and information in Chapters 3, 29, and 30 of the 2017 ASHRAE Handbook—Fundamentals to calculate friction losses.
For information on refrigeration load, see Chapter 24. For R-502 information, refer to the 1998 ASHRAE Handbook—Refrigeration.
Development of halocarbon refrigerants dates back to the 1920s. The main refrigerants used then were ammonia (R-717), chloromethane (R-40), and sulfur dioxide (R-764), all of which have some degree of toxicity and/or flammability. These first-generation refrigerants were an impediment to Frigidaire’s plans to expand into refrigeration and air conditioning, so Frigidaire and DuPont collaborated to develop safer refrigerants. In 1928, Thomas Midgley, Jr., of Frigidaire and his colleagues developed the first commercially available CFC refrigerant, dichlorodifluoromethane (R-12) (Giunta 2006). Chlorinated halocarbon refrigerants represent the second generation of refrigerants (Calm 2008).
Concern about the use of halocarbon refrigerants began with a 1974 paper by two University of California professors, Frank Rowland and Mario Molina, in which they highlighted the damage chlorine could cause to the ozone layer in the stratosphere. This publication eventually led to the Montreal Protocol Agreement in 1987 and its subsequent revisions, which restricted the production and use of chlorinated halocarbon (CFC and HCFC) refrigerants. All CFC refrigerant production was phased out in the United States at the beginning of 1996. Replacement HFC, third-generation refrigerants were developed following these restrictions (Calm 2008).
Although HFC refrigerants do not contain chlorine and thus have no effect on stratospheric ozone, they have come under heavy scrutiny because of their global warming potential (GWP): like CFCs and HCFCs, they are greenhouse gases, and can trap radiant energy (IPCC 1990). In October 2016, in Kigali, Rwanda, the 1987 Montreal Protocol Agreement was revised to also include regulation of HFC refrigerants as controlled substances. This Kigali Agreement marks a commitment from a significant portion of the world to deal with the global warming consequences of HFC gases. As phasedown begins, interest in the future cost and availability of these refrigerants is likely to increase.
Indeed, portions of the United States and Europe already had HFC regulations that predated the Kigali Agreement. The latest fluorinated greenhouse gas (F-gas) regulation in Europe adopted in 2014 (revised from the initial adoption in 2006) aims to reduce HFC refrigerant sales to one-fifth of 2014 levels by 2030. Some HFCs have already been banned where suitable alternatives are widely available, and all systems require specific maintenance checks, servicing, and refrigerant reclamation when the system is decommissioned. In the United States, California’s Global Warming Solutions Act (Assembly Bill 32; www.arb.ca.gov/cc/ab32/ab32.htm) went into effect in 2011; this bill’s early adoption measures began regulating HFC refrigerants to reduce the environmental consequences of greenhouse gases. These early adoption measures were designed as the prelude to a proposed HFC phaseout, and include required service practices; leak inspection; charge monitoring and record keeping; system retrofit and retirement plans; and refrigerant distributor, wholesaler, and reclaimer prohibitions.
HFO refrigerants have significantly lower GWP values than HFCs, and are being developed and promoted as alternatives to HFC refrigerants. However, HFOs are classed as mildly flammable, which is an obvious barrier to adoption. Safety measures must be fully developed and widely adopted for common use of mildly flammable refrigerants to be feasible. For example, in the United States, entities such as ASHRAE, the U.S. Environmental Protection Agency (EPA), and Underwriters Laboratories (UL) will need to reach a coordinated agreement to allow broad use of these fourth-generation refrigerants before local and state codes will be in a position to allow their use.
HFC refrigeration systems are still widely used and will continue to be used during the transition to natural or other reduced-GWP refrigerants, so many owners, engineers, and manufacturers seek to reduce charge and build tighter systems to reduce the total system charge on site and ensure that less refrigerant is released into the atmosphere. Table 1 in Chapter 3 lists commonly used refrigerants and their corresponding GWP values.
Also, using indirect and cascade systems to reduce the total amount of refrigerant has become increasingly popular. These systems also reduce the possibility for leakage because large amounts of interconnecting piping between the compressors and the heat load are replaced mainly with glycol or CO2 piping. (See Chapter 9 for more information on refrigerant containment, recovery, recycling, and reclamation.)
ASHRAE Standard 15 and ASME Standard B31.5 should be used as guides for safe practice because they are the basis of most municipal and state codes. However, some ordinances require heavier piping and other features. The designer should know the specific requirements of the installation site. Only A106 Grade A or B or A53 Grade A or B should be considered for steel refrigerant piping.
The rated internal working pressure for Type L copper tubing decreases with (1) increasing metal operating temperature, (2) increasing tubing size (OD), and (3) increasing temperature of joining method. Hot methods used to join drawn pipe (e.g., brazing, welding) produce joints as strong as surrounding pipe, but reduce the strength of the heated pipe material to that of annealed material. Particular attention should be paid when specifying copper in conjunction with newer, high-pressure refrigerants (e.g., R-404A, R-507A, R-410A, R-407C) because some of these refrigerants can achieve operating pressures as high as 500 psia and operating temperatures as high as 300°F at a typical saturated condensing condition of 130°F.
Concentration calculations, based on the amount of refrigerant in the system and the volume of the space where it is installed, are needed to identify what safety features are required by the appropriate codes. Whenever allowable concentration limits of the refrigerant may be exceeded in occupied spaces, additional safety measures (e.g., leak detection, alarming, ventilation, automatic shutoff controls) are typically required. Note that, because halocarbon refrigerants are heavier than air, leak detection sensors should be placed at lower elevations in the space (typically 12 in. from the floor).
3. BASIC PIPING PRINCIPLES
The design and operation of refrigerant piping systems should (1) ensure proper refrigerant feed to evaporators, (2) provide practical refrigerant line sizes without excessive pressure drop, (3) prevent excessive amounts of lubricating oil from being trapped in any part of the system, (4) protect the compressor at all times from loss of lubricating oil, (5) prevent liquid refrigerant or oil slugs from entering the compressor during operating and idle time, and (6) maintain a clean and dry system.
Refrigerant Line Velocities
Economics, pressure drop, noise, and oil entrainment establish feasible design velocities in refrigerant lines (Table 1). Higher gas velocities are sometimes found in relatively short suction lines on comfort air-conditioning or other applications where the operating time is only 2000 to 4000 h per year and where low initial cost of the system may be more significant than low operating cost. Industrial or commercial refrigeration applications, where equipment runs almost continuously, should be designed with low refrigerant velocities for most efficient compressor performance and low equipment operating costs. An owning and operating cost analysis will reveal the best choice of line sizes. (See Chapter 37 of the 2019 ASHRAE Handbook—HVAC Applications for information on owning and operating costs.) Liquid drain lines from condensers to receivers should be sized for 100 fpm or less to ensure positive gravity flow without incurring back-up of liquid flow. Where calculated velocities exceed 100 fpm or where liquid may trap in the drain line, preventing a reverse flow of vapor from the receiver to the condenser, pressure equalization lines should be installed from the receiver to the condenser drain header. Liquid lines from receiver to evaporator should be sized to maintain velocities below 300 fpm, thus minimizing or preventing liquid hammer when solenoids or other electrically operated valves are used.
Refrigerant flow rates for R-22 and R-134a are indicated in Figures 1 and 2. To obtain total system flow rate, select the proper rate value and multiply by system capacity. Enter curves using saturated refrigerant temperature at the evaporator outlet and actual liquid temperature entering the liquid feed device (including subcooling in condensers and liquid-suction interchanger, if used).
Because Figures 1 and 2 are based on a saturated evaporator temperature, they may indicate slightly higher refrigerant flow rates than are actually in effect when suction vapor is superheated above the conditions mentioned. Refrigerant flow rates may be reduced approximately 3% for each 10°F increase in superheat in the evaporator.
Suction-line superheating downstream of the evaporator from line heat gain from external sources should not be used to reduce evaluated mass flow, because it increases volumetric flow rate and line velocity per unit of evaporator capacity, but not mass flow rate. It should be considered when evaluating suction-line size for satisfactory oil return up risers.
Suction gas superheating from use of a liquid-suction heat exchanger has an effect on oil return similar to that of suction-line superheating. The liquid cooling that results from the heat exchange reduces mass flow rate per unit of refrigeration. This can be seen in Figures 1 and 2 because the reduced temperature of the liquid supplied to the evaporator feed valve has been taken into account
Superheat caused by heat in a space not intended to be cooled is always detrimental because the volumetric flow rate increases with no compensating gain in refrigerating effect.
4. REFRIGERANT LINE SIZING
When sizing refrigerant lines, designers must consider not only the effects of velocity and pressure drop in the pipe on system performance, but also system cost and safety. Although smaller pipes may be cheaper, they inflict higher operating costs for the life of the system because of excessive pressure drop. However, there are diminishing efficiency benefits when moving to larger pipe sizes, and it is necessary to strike a balance.
When considering safety, remember that rated working pressures for any pipe material decrease as pipe diameters increase. Pipes should be carefully selected, ensuring that internal system pressures will not exceed the pipe’s rated working pressure while the system is in operation or at standstill. It is also important to understand that any brazed copper piping will be weakened by the annealing that occurs during brazing. Typically, two separate working pressures are published for copper: one for annealed copper and one for drawn copper. Drawn copper working pressures should only be used if and when pipes are fitted together without brazing (i.e., when mechanical fittings are used).
Pressure Drop Considerations
Suction- and discharge-line pressure drops cause loss of compressor capacity and increased power usage. Excessive liquid-line pressure drops can cause liquid refrigerant to flash, resulting in faulty expansion valve operation. Refrigeration systems are designed so that friction pressure losses do not exceed a pressure differential equivalent to a corresponding change in the saturation boiling temperature. The primary measure for determining pressure drops is a given change in saturation temperature.
Table 2 shows the approximate effect of refrigerant pressure drop on an R-22 system operating at a 40°F saturated evaporator temperature with a 100°F saturated condensing temperature.
Pressure drop calculations are determined as normal pressure loss associated with a change in saturation temperature of the refrigerant. Typically, the refrigeration system is sized for pressure losses of 2°F or less for each segment of the discharge, suction, and liquid lines.
Liquid Lines. Pressure drop should not be so large as to cause gas formation in the liquid line, insufficient liquid pressure at the liquid feed device, or both. Systems are normally designed so that pressure drop in the liquid line from friction is not greater than that corresponding to about a 1 to 2°F change in saturation temperature. See Tables 3 to 9 for liquid-line sizing information.
Liquid subcooling is the only method of overcoming liquid line pressure loss to guarantee liquid at the expansion device in the evaporator. If subcooling is insufficient, flashing occurs in the liquid line and degrades system efficiency.
Friction pressure drops in the liquid line are caused by accessories such as solenoid valves, filter-driers, and hand valves, as well as by the actual pipe and fittings between the receiver outlet and the refrigerant feed device at the evaporator.
Liquid-line risers are a source of pressure loss and add to the total loss of the liquid line. Loss caused by risers is approximately 0.5 psi per foot of liquid lift. Total loss is the sum of all friction losses plus pressure loss from liquid risers.
Example 1 shows the process of determining liquid-line size and checking for total subcooling required.
Example 1.
An R-22 refrigeration system using copper pipe operates at 40°F evaporator and 105°F condensing. Capacity is 5 tons, and the liquid line is 100 ft equivalent length with a riser of 20 ft. Determine the liquid-line size and total required subcooling.
Solution: From Table 3, the size of the liquid line at 1°F drop is 5/8 in. OD. Use the equation in Note 3 of Table 3 to compute actual temperature drop. At 5 tons,
Refrigeration systems that have no liquid risers and have the evaporator below the condenser/receiver benefit from a gain in pressure caused by liquid weight and can tolerate larger friction losses without flashing. Regardless of the liquid-line routing when flashing occurs, overall efficiency is reduced, and the system may malfunction.
The velocity of liquid leaving a partially filled vessel (e.g., receiver, shell-and-tube condenser) is limited by the height of the liquid above the point at which the liquid line leaves the vessel, whether or not the liquid at the surface is subcooled. Because liquid in the vessel has a very low (or zero) velocity, the velocity V in the liquid line (usually at the vena contracta) is V2 = 2gh, where h is the liquid height in the vessel. Gas pressure does not add to the velocity unless gas is flowing in the same direction. As a result, both gas and liquid flow through the line, limiting the rate of liquid flow. If this factor is not considered, excess operating charges in receivers and flooding of shell-and-tube condensers may result.
No specific data are available to precisely size a line leaving a vessel. If the height of liquid above the vena contracta produces the desired velocity, liquid leaves the vessel at the expected rate. Thus, if the level in the vessel falls to one pipe diameter above the bottom of the vessel from which the liquid line leaves, the capacity of copper lines for R-22 at 3 lb/min per ton of refrigeration is approximately as follows:
The whole liquid line need not be as large as the leaving connection. After the vena contracta, the velocity is about 40% less. If the line continues down from the receiver, the value of h increases. For a 200 ton capacity with R-22, the line from the bottom of the receiver should be about 3 1/8 in. After a drop of 1 ft, a reduction to 2 5/8 in. is satisfactory.
Suction Lines. Suction lines are more critical than liquid and discharge lines from a design and construction standpoint. Refrigerant lines should be sized to (1) provide a minimum pressure drop at full load, (2) return oil from the evaporator to the compressor under minimum load conditions, and (3) prevent oil from draining from an active evaporator into an idle one. A pressure drop in the suction line reduces a system’s capacity because it forces the compressor to operate at a lower suction pressure to maintain a desired evaporating temperature in the coil. The suction line is normally sized to have a pressure drop from friction no greater than the equivalent of about a 2°F change in saturation temperature. See Tables 3 to 15 for suction line sizing information.
At suction temperatures lower than 40°F, the pressure drop equivalent to a given temperature change decreases. For example, at −40°F suction with R-22, the pressure drop equivalent to a 2°F change in saturation temperature is about 0.8 psi. Therefore, low-temperature lines must be sized for a very low pressure drop, or higher equivalent temperature losses, with resultant loss in equipment capacity, must be accepted. For very low pressure drops, any suction or hot-gas risers must be sized properly to ensure oil entrainment up the riser so that oil is always returned to the compressor.
Where pipe size must be reduced to provide sufficient gas velocity to entrain oil up vertical risers at partial loads, greater pressure drops are imposed at full load. These can usually be compensated for by oversizing the horizontal and down run lines and components.
Discharge Lines. Pressure loss in hot-gas lines increases the required compressor power per unit of refrigeration and decreases compressor capacity. Table 2 shows power losses for an R-22 system at 40°F evaporator and 100°F condensing temperature. Pressure drop is minimized by generously sizing lines for low friction losses, but still maintaining refrigerant line velocities to entrain and carry oil along at all loading conditions. Pressure drop is normally designed not to exceed the equivalent of a 2°F change in saturation temperature. Recommended sizing tables are based on a 1°F change in saturation temperature per 100 ft.
Location and Arrangement of Piping
Refrigerant lines should be as short and direct as possible to minimize tubing and refrigerant requirements and pressure drops. Plan piping for a minimum number of joints using as few elbows and other fittings as possible, but provide sufficient flexibility to absorb compressor vibration and stresses caused by thermal expansion and contraction.
Arrange refrigerant piping so that normal inspection and servicing of the compressor and other equipment is not hindered. Do not obstruct the view of the oil-level sight glass or run piping so that it interferes with removing compressor cylinder heads, end bells, access plates, or any internal parts. Suction-line piping to the compressor should be arranged so that it will not interfere with removal of the compressor for servicing.
Provide adequate clearance between pipe and adjacent walls and hangers or between pipes for insulation installation. Use sleeves that are sized to allow installation of both pipe and insulation through floors, walls, or ceilings. Set these sleeves before pouring concrete or erecting brickwork.
Run piping so that it does not interfere with passages or obstruct headroom, windows, and doors. Refer to ASHRAE Standard 15 and other governing local codes for restrictions that may apply.
Protection Against Damage to Piping
Protection against damage is necessary, particularly for small lines, which have a false appearance of strength. Where traffic is heavy, provide protection against impact from carelessly handled hand trucks, overhanging loads, ladders, and fork trucks.
All piping joints and fittings should be thoroughly leak tested before insulation is sealed. Suction lines should be insulated to prevent sweating and heat gain. Insulation covering lines on which moisture can condense or lines subjected to outdoor conditions must be vapor sealed to prevent any moisture travel through the insulation or condensation in the insulation. Many commercially available types are provided with an integral waterproof jacket for this purpose. Although the liquid line ordinarily does not require insulation, suction and liquid lines can be insulated as a unit on installations where the two lines are clamped together. When it passes through a warmer area, the liquid line should be insulated to minimize heat gain. Hot-gas discharge lines usually are not insulated; however, they should be insulated if necessary to prevent injury from high-temperature surfaces, or if the heat dissipated is objectionable (e.g., in systems that use heat reclaim). In this case, discharge lines upstream of the heat reclaim heat exchanger should be insulated. Downstream lines (between the heat reclaim heat exchanger and condenser) do not need to be insulated unless necessary to prevent the refrigerant from condensing prematurely. Also, indoor hot-gas discharge line insulation does not need a tight vapor seal because moisture condensation is not an issue.
All joints and fittings should be covered, but it is not advisable to do so until the system has been thoroughly leak tested. See Chapter 10 for additional information.
Vibration and Noise in Piping
Vibration transmitted through or generated in refrigerant piping and the resulting objectionable noise can be eliminated or minimized by proper piping design and support.
Two undesirable effects of vibration of refrigerant piping are (1) physical damage to the piping, which can break brazed joints and, consequently, lose charge; and (2) transmission of noise through the piping itself and through building construction that may come into direct contact with the piping.
In refrigeration applications, piping vibration can be caused by rigid connection of the refrigerant piping to a reciprocating compressor. Vibration effects are evident in all lines directly connected to the compressor or condensing unit. It is thus impossible to eliminate vibration in piping; it is only possible to mitigate its effects.
Flexible metal hose is sometimes used to absorb vibration transmission along smaller pipe sizes. For maximum effectiveness, it should be installed parallel to the crankshaft. In some cases, two isolators may be required, one in the horizontal line and the other in the vertical line at the compressor. A rigid brace on the end of the flexible hose away from the compressor is required to prevent vibration of the hot-gas line beyond the hose.
Flexible metal hose is not as efficient in absorbing vibration on larger pipes because it is not actually flexible unless the ratio of length to diameter is relatively great. In practice, the length is often limited, so flexibility is reduced in larger sizes. This problem is best solved by using flexible piping and isolation hangers where the piping is secured to the structure.
When piping passes through walls, through floors, or inside furring, it must not touch any part of the building and must be supported only by the hangers (provided to avoid transmitting vibration to the building); this eliminates the possibility of walls or ceilings acting as sounding boards or diaphragms. When piping is erected where access is difficult after installation, it should be supported by isolation hangers.
Vibration and noise from a piping system can also be caused by gas pulsations from the compressor operation or from turbulence in the gas, which increases at high velocities. It is usually more apparent in the discharge line than in other parts of the system.
When gas pulsations caused by the compressor create vibration and noise, they have a characteristic frequency that is a function of the number of gas discharges by the compressor on each revolution. This frequency is not necessarily equal to the number of cylinders, because on some compressors two pistons operate together. It is also varied by the angular displacement of the cylinders, such as in V-type compressors. Noise resulting from gas pulsations is usually objectionable only when the piping system amplifies the pulsation by resonance. On single-compressor systems, resonance can be reduced by changing the size or length of the resonating line or by installing a properly sized hot-gas muffler in the discharge line immediately after the compressor discharge valve. On a paralleled compressor system, a harmonic frequency from the different speeds of multiple compressors may be apparent. This noise can sometimes be reduced by installing mufflers.
When noise is caused by turbulence and isolating the line is not effective enough, installing a larger-diameter pipe to reduce gas velocity is sometimes helpful. Also, changing to a line of heavier wall or from copper to steel to change the pipe natural frequency may help.
Refrigerant Line Capacity Tables
Tables 3 to 9 show line capacities in tons of refrigeration for R-22, R-134a, R-404A, R-507A, R-410A, and R-407C. Capacities in the tables are based on the refrigerant flow that develops a friction loss, per 100 ft of equivalent pipe length, corresponding to a 2°F change in the saturation temperature (Δt) in the suction line, and a 1°F change in the discharge line. The capacities shown for liquid lines are for pressure losses corresponding to 1 and 5°F change in saturation temperature and also for velocity corresponding to 100 fpm. Tables 10 to 15 show capacities for the same refrigerants based on reduced suction line pressure loss corresponding to 1.0 and 0.5°F per 100 ft equivalent length of pipe. These tables may be used when designing system piping to minimize suction line pressure drop.
The refrigerant line sizing capacity tables are based on the Darcy-Weisbach relation and friction factors as computed by the Colebrook function (Colebrook 1938, 1939). Tubing roughness height is 0.000005 ft for copper and 0.00015 ft for steel pipe. Viscosity extrapolations and adjustments for pressures other than 1 atm were based on correlation techniques as presented by Keating and Matula (1969). Discharge gas superheat was 80°F for R-134a and 105°F for R-22.
The refrigerant cycle for determining capacity is based on saturated gas leaving the evaporator. The calculations neglect the presence of oil and assume nonpulsating flow.
For additional charts and discussion of line sizing refer to Atwood (1990), Timm (1991), and Wile (1977).
Equivalent Lengths of Valves and Fittings
Refrigerant line capacity tables are based on unit pressure drop per 100 ft length of straight pipe, or per combination of straight pipe, fittings, and valves with friction drop equivalent to a 100 ft length of straight pipe.
Generally, pressure drop through valves and fittings is determined by establishing the equivalent straight length of pipe of the same size with the same friction drop. Line sizing tables can then be used directly. Tables 16 to 18 give equivalent lengths of straight pipe for various fittings and valves, based on nominal pipe sizes.
The following example shows the use of various tables and charts to size refrigerant lines.
Example 2.
Determine the line size and pressure drop equivalent (in degrees) for the suction line of a 30 ton R-22 system, operating at 40°F suction and 100°F condensing temperatures. Suction line is copper tubing, with 50 ft of straight pipe and six long-radius elbows.
Solution: Add 50% to the straight length of pipe to establish a trial equivalent length. Trial equivalent length is 50 × 1.5 = 75 ft. From Table 3 (for 40°F suction, 105°F condensing), 33.1 tons capacity in 2 1/8 in. OD results in a 2°F loss per 100 ft equivalent length. Referring to Note 4, Table 3, capacity at 40°F evaporator and 100°F condensing temperature is 1.03 × 33.1 = 34.1 ton. This trial size is used to evaluate actual equivalent length.
Oil Management in Refrigerant Lines
Oil Circulation. All compressors lose some lubricating oil during normal operation. Because oil inevitably leaves the compressor with the discharge gas, systems using halocarbon refrigerants must return this oil at the same rate at which it leaves (Cooper 1971).
Oil that leaves the compressor or oil separator reaches the condenser and dissolves in the liquid refrigerant, enabling it to pass readily through the liquid line to the evaporator. In the evaporator, the refrigerant evaporates, and the liquid phase becomes enriched in oil. The concentration of refrigerant in the oil depends on the evaporator temperature and types of refrigerant and oil used. The viscosity of the oil/refrigerant solution is determined by the system parameters. Oil separated in the evaporator is returned to the compressor by gravity or by drag forces of the returning gas. Oil’s effect on pressure drop is large, increasing the pressure drop by as much as a factor of 10 (Alofs et al. 1990).
One of the most difficult problems in low-temperature refrigeration systems using halocarbon refrigerants is returning lubrication oil from the evaporator to the compressors. Except for most centrifugal compressors and rarely used nonlubricated compressors, refrigerant continuously carries oil into the discharge line from the compressor. Most of this oil can be removed from the stream by an oil separator and returned to the compressor. Coalescing oil separators are far better than separators using only mist pads or baffles; however, they are not 100% effective. Oil that finds its way into the system must be managed.
Oil mixes well with halocarbon refrigerants at higher temperatures. As temperature decreases, miscibility is reduced, and some oil separates to form an oil-rich layer near the top of the liquid level in a flooded evaporator. If the temperature is very low, the oil becomes a gummy mass that prevents refrigerant controls from functioning, blocks flow passages, and fouls heat transfer surfaces. Proper oil management is often key to a properly functioning system.
In general, direct-expansion and liquid overfeed system evaporators have fewer oil return problems than do flooded system evaporators because refrigerant flows continuously at velocities high enough to sweep oil from the evaporator. Low-temperature systems using hot-gas defrost can also be designed to sweep oil out of the circuit each time the system defrosts. This reduces the possibility of oil coating the evaporator surface and hindering heat transfer.
Flooded evaporators can promote oil contamination of the evaporator charge because they may only return dry refrigerant vapor back to the system. Skimming systems must sample the oil-rich layer floating in the drum, a heat source must distill the refrigerant, and the oil must be returned to the compressor. Because flooded halocarbon systems can be elaborate, some designers avoid them.
System Capacity Reduction. Using automatic capacity control on compressors requires careful analysis and design. The compressor can load and unload as it modulates with system load requirements through a considerable range of capacity. A single compressor can unload down to 25% of full-load capacity, and multiple compressors connected in parallel can unload to a system capacity of 12.5% or lower. System piping must be designed to return oil at the lowest loading, yet not impose excessive pressure drops in the piping and equipment at full load.
Oil Return up Suction Risers. Many refrigeration piping systems contain a suction riser because the evaporator is at a lower level than the compressor. Oil circulating in the system can return up gas risers only by being transported by returning gas or by auxiliary means such as a trap and pump. The minimum conditions for oil transport correlate with buoyancy forces (i.e., density difference between liquid and vapor, and momentum flux of vapor) (Jacobs et al. 1976).
The principal criteria determining the transport of oil are gas velocity, gas density, and pipe inside diameter. Density of the oil/ refrigerant mixture plays a somewhat lesser role because it is almost constant over a wide range. In addition, at temperatures somewhat lower than −40°F, oil viscosity may be significant. Greater gas velocities are required as temperature drops and the gas becomes less dense. Higher velocities are also necessary if the pipe diameter increases. Table 19 translates these criteria to minimum refrigeration capacity requirements for oil transport. Suction risers must be sized for minimum system capacity. Oil must be returned to the compressor at the operating condition corresponding to the minimum displacement and minimum suction temperature at which the compressor will operate. When suction or evaporator pressure regulators are used, suction risers must be sized for actual gas conditions in the riser.
For a single compressor with capacity control, the minimum capacity is the lowest capacity at which the unit can operate. For multiple compressors with capacity control, the minimum capacity is the lowest at which the last operating compressor can run.
Riser Sizing. The following example demonstrates the use of Table 19 in establishing maximum riser sizes for satisfactory oil transport down to minimum partial loading.
Example 3.
Determine the maximum size suction riser that will transport oil at minimum loading, using R-22 with a 40 ton compressor with capacity in steps of 25, 50, 75, and 100%. Assume the minimum system loading is 10 tons at 40°F suction and 105°F condensing temperatures with 15°F superheat.
Solution: From Table 19, a 2 1/8 in. OD pipe at 40°F suction and 90°F liquid temperature has a minimum capacity of 7.5 tons. When corrected to 105°F liquid temperature using the chart at the bottom of Table 19, minimum capacity becomes 7.2 tons. Therefore, 2 1/8 in. OD pipe is suitable.
Based on Table 19, the next smaller line size should be used for marginal suction risers. When vertical riser sizes are reduced to provide satisfactory minimum gas velocities, pressure drop at full load increases considerably; horizontal lines should be sized to keep total pressure drop within practical limits. As long as horizontal lines are level or pitched in the direction of the compressor, oil can be transported with normal design velocities.
Because most compressors have multiple capacity-reduction features, gas velocities required to return oil up through vertical suction risers under all load conditions are difficult to maintain. When the suction riser is sized to allow oil return at the minimum operating capacity of the system, pressure drop in this portion of the line may be too great when operating at full load. If a correctly sized suction riser imposes too great a pressure drop at full load, a double suction riser should be used.
Oil Return up Suction Risers: Multistage Systems. Oil movement in the suction lines of multistage systems requires the same design approach as that for single-stage systems. For oil to flow up along a pipe wall, a certain minimum drag of gas flow is required. Drag can be represented by the friction gradient. The following sizing data may be used for ensuring oil return up vertical suction lines for refrigerants other than those listed in Tables 19 and 20. The line size selected should provide a pressure drop equal to or greater than that shown in the chart.
Double Suction Risers. Figure 3 shows two methods of double suction riser construction. Oil return in this arrangement is accomplished at minimum loads, but it does not cause excessive pressure drops at full load. Sizing and operation of a double suction riser are as follows:
Riser A is sized to return oil at minimum load possible.
Riser B is sized for satisfactory pressure drop through both risers at full load. The usual method is to size riser B so that the combined cross-sectional area of A and B is equal to or slightly greater than the cross-sectional area of a single pipe sized for acceptable pressure drop at full load without regard for oil return at minimum load. The combined cross-sectional area, however, should not be greater than the cross-sectional area of a single pipe that would return oil in an upflow riser under maximum load.
A trap is introduced between the two risers, as shown in both methods. During part-load operation, gas velocity is not sufficient to return oil through both risers, and the trap gradually fills up with oil until riser B is sealed off. The gas then travels up riser A only with enough velocity to carry oil along with it back into the horizontal suction main.
The trap’s oil-holding capacity is limited by close-coupling the fittings at the bottom of the risers. If this is not done, the trap can accumulate enough oil during part-load operation to lower the compressor crankcase oil level. Note in Figure 3 that riser lines A and B form an inverted loop and enter the horizontal suction line from the top. This prevents oil drainage into the risers, which may be idle during part-load operation. The same purpose can be served by running risers horizontally into the main, provided that the main is larger in diameter than either riser.
Often, double suction risers are essential on low-temperature systems that can tolerate very little pressure drop. Any system using these risers should include a suction trap (accumulator) and a means of returning oil gradually.
For systems operating at higher suction temperatures, such as for comfort air conditioning, single suction risers can be sized for oil return at minimum load. Where single compressors are used with capacity control, minimum capacity is usually 25 or 33% of maximum displacement. With this low ratio, pressure drop in single suction risers designed for oil return at minimum load is rarely serious at full load.
When multiple compressors are used, one or more may shut down while another continues to operate, and the maximum-to-minimum ratio becomes much larger. This may make a double suction riser necessary.
The remaining suction line portions are sized to allow a practical pressure drop between the evaporators and compressors because oil is carried along in horizontal lines at relatively low gas velocities. It is good practice to give some pitch to these lines toward the compressor. Avoid traps, but when that is impossible, the risers from them are treated the same as those leading from the evaporators.
Preventing Oil Trapping in Idle Evaporators. Suction lines should be designed so that oil from an active evaporator does not drain into an idle one. Figure 4A shows multiple evaporators on different floor levels with the compressor above. Each suction line is brought upward and looped into the top of the common suction line to prevent oil from draining into inactive coils.
Figure 4B shows multiple evaporators stacked on the same level, with the compressor above. Oil cannot drain into the lowest evaporator because the common suction line drops below the outlet of the lowest evaporator before entering the suction riser.
Figure 4C shows multiple evaporators on the same level, with the compressor located below. The suction line from each evaporator drops down into the common suction line so that oil cannot drain into an idle evaporator. An alternative arrangement is shown in Figure 4D for cases where the compressor is above the evaporators.
Figure 5 shows typical piping for evaporators above and below a common suction line. All horizontal runs should be level or pitched toward the compressor to ensure oil return.
Traps shown in the suction lines after the evaporator suction outlet are recommended by thermal expansion valve manufacturers to prevent erratic operation of the thermal expansion valve. Expansion valve bulbs are located on the suction lines between the evaporator and these traps. The traps serve as drains and help prevent liquid from accumulating under the expansion valve bulbs during compressor off cycles. They are useful only where straight runs or risers are encountered in the suction line leaving the evaporator outlet.
5. PIPING AT MULTIPLE COMPRESSORS
Multiple compressors operating in parallel must be carefully piped to ensure proper operation.
Suction piping should be designed so that all compressors run at the same suction pressure and oil is returned in equal proportions. All suction lines should be brought into a common suction header to return oil to each crankcase as uniformly as possible. Depending on the type and size of compressors, oil may be returned by designing the piping in one or more of the following schemes:
Oil returned with the suction gas to each compressor
Oil contained with a suction trap (accumulator) and returned to the compressors through a controlled means
Oil trapped in a discharge line separator and returned to the compressors through a controlled means (see the section on Discharge Piping)
The suction header is a means of distributing suction gas equally to each compressor. Header design can freely pass the suction gas and oil mixture or provide a suction trap for the oil. The header should be run above the level of the compressor suction inlets so oil can drain into the compressors by gravity.
Figure 6 shows a pyramidal or yoke-type suction header to maximize pressure and flow equalization at each of three compressor suction inlets piped in parallel. This type of construction is recommended for applications of three or more compressors in parallel. For two compressors in parallel, a single feed between the two compressor takeoffs is acceptable. Although not as good for equalizing flow and pressure drops to all compressors, one alternative is to have the suction line from evaporators enter at one end of the header instead of using the yoke arrangement. The suction header may have to be enlarged to minimize pressure drop and flow turbulence.
Suction headers designed to freely pass the gas/oil mixture should have branch suction lines to compressors connected to the side of the header. Return mains from the evaporators should not be connected into the suction header to form crosses with the branch suction lines to the compressors. The header should be full size based on the largest mass flow of the suction line returning to the compressors. Takeoffs to the compressors should either be the same size as the suction header or be constructed so that oil will not trap in the suction header. Branch suction lines to the compressors should not be reduced until the vertical drop is reached.
Suction traps are recommended wherever (1) parallel compressors, (2) flooded evaporators, (3) double suction risers, (4) long suction lines, (5) multiple expansion valves, (6) hot-gas defrost, (7) reverse-cycle operation, or (8) suction-pressure regulators are used.
Depending on system size, the suction header may be designed to function as a suction trap. The suction header should be large enough to provide a low-velocity region in the header to allow suction gas and oil to separate. See the section on Low-Pressure Receiver Sizing in Chapter 4 to find recommended velocities for separation. Suction gas flow for individual compressors should be taken off the top of the suction header. Oil can be returned to the compressor directly or through a vessel equipped with a heater to boil off refrigerant and then allow oil to drain to the compressors or other devices used to feed oil to the compressors.
The suction trap must be sized for effective gas and liquid separation. Adequate liquid volume and a means of disposing of it must be provided. A liquid transfer pump or heater may be used. Chapter 4 has further information on separation and liquid transfer pumps.
An oil receiver equipped with a heater effectively evaporates liquid refrigerant accumulated in the suction trap. It also ensures that each compressor receives its share of oil. Either crankcase float valves or external float switches and solenoid valves can be used to control the oil flow to each compressor.
A gravity-feed oil receiver should be elevated to overcome the pressure drop between it and the crankcase. The oil receiver should be sized so that a malfunction of the oil control mechanism cannot overfill an idle compressor.
Figure 7 shows a recommended hookup of multiple compressors, suction trap (accumulator), oil receiver, and discharge line oil separators. The oil receiver also provides a reserve supply of oil for compressors where oil in the system outside the compressor varies with system loading. The heater mechanism should always be submerged.
The piping arrangement in Figure 6 is suggested for discharge piping. The piping must be arranged to prevent refrigerant liquid and oil from draining back into the heads of idle compressors. A check valve in the discharge line may be necessary to prevent refrigerant and oil from entering the compressor heads by migration. It is recommended that, after leaving the compressor head, the piping be routed to a lower elevation so that a trap is formed to allow drainback of refrigerant and oil from the discharge line when flow rates are reduced or the compressors are off. If an oil separator is used in the discharge line, it may suffice as the trap for drainback for the discharge line.
Avoid using a bullheaded tee at the junction of two compressor branches and the main discharge header: this configuration causes increased turbulence, increased pressure drop, and possible hammering in the line.
When an oil separator is used on multiple-compressor arrangements, oil must be piped to return to the compressors. This can be done in various ways, depending on the oil management system design. Oil may be returned to an oil receiver that is the supply for control devices feeding oil back to the compressors.
Interconnecting Crankcases
When two or more compressors are interconnected, a method must be provided to equalize the crankcases. Some compressor designs do not operate correctly with simple equalization of the crankcases. For these systems, it may be necessary to design a positive oil float control system for each compressor crankcase. A typical system allows oil to collect in a receiver that, in turn, supplies oil to a device that meters it back into the compressor crankcase to maintain a proper oil level (Figure 7).
Compressor systems that can be equalized should be placed on foundations so that all oil equalizer tapping locations are exactly level. If crankcase floats (as in Figure 7) are not used, an oil equalization line should connect all crankcases to maintain uniform oil levels. The oil equalizer may be run level with the tapping, or, for convenient access to compressors, it may be run at the floor (Figure 8). It should never be run at a level higher than that of the tapping.
For the oil equalizer line to work properly, equalize the crankcase pressures by installing a gas equalizer line above the oil level. This line may be run to provide head room (Figure 8) or run level with tapping on the compressors. It should be piped so that oil or liquid refrigerant will not be trapped.
Both lines should be the same size as the tapping on the largest compressor and should be valved so that any one machine can be taken out for repair. The piping should be arranged to absorb vibration.
6. PIPING AT VARIOUS SYSTEM COMPONENTS
For a description of flooded fluid coolers, see Chapter 42 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment.
Shell-and-tube flooded coolers designed to minimize liquid entrainment in the suction gas require a continuous liquid bleed line (Figure 9) installed at some point in the cooler shell below the liquid level to remove trapped oil. This continuous bleed of refrigerant liquid and oil prevents the oil concentration in the cooler from getting too high. The location of the liquid bleed connection on the shell depends on the refrigerant and oil used. For refrigerants that are highly miscible with the oil, the connection can be anywhere below the liquid level.
Refrigerant 22 can have a separate oil-rich phase floating on a refrigerant-rich layer. This becomes more pronounced as evaporating temperature drops. When R-22 is used with mineral oil, the bleed line is usually taken off the shell just slightly below the liquid level, or there may be more than one valved bleed connection at slightly different levels so that the optimum point can be selected during operation. With alkyl benzene lubricants, oil/refrigerant miscibility may be high enough that the oil bleed connection can be anywhere below the liquid level. The solubility charts in Chapter 12 give specific information.
Where the flooded cooler design requires an external surge drum to separate liquid carryover from suction gas off the tube bundle, the richest oil concentration may or may not be in the cooler. In some cases, the surge drum has the highest concentration of oil. Here, the refrigerant and oil bleed connection is taken from the surge drum. The refrigerant and oil bleed from the cooler by gravity. The bleed sometimes drains into the suction line so oil can be returned to the compressor with the suction gas after the accompanying liquid refrigerant is vaporized in a liquid-suction heat interchanger. A better method is to drain the refrigerant/oil bleed into a heated receiver that boils refrigerant off to the suction line and drains oil back to the compressor.
For further information on refrigerant feed devices, see Chapter 11. The pilot-operated low-side float control (Figure 9) is sometimes selected for flooded systems using halocarbon refrigerants. Except for small capacities, direct-acting low-side float valves are impractical for these refrigerants. The displacer float controlling a pneumatic valve works well for low-side liquid level control; it allows the cooler level to be adjusted within the instrument without disturbing the piping.
High-side float valves are practical only in single-evaporator systems, because distribution problems result when multiple evaporators are used.
Float chambers should be located as near the liquid connection on the cooler as possible because a long length of liquid line, even if insulated, can pick up room heat and give an artificial liquid level in the float chamber. Equalizer lines to the float chamber must be amply sized to minimize the effect of heat transmission. The float chamber and its equalizing lines must be insulated.
Each flooded cooler system must have a way of keeping oil concentration in the evaporator low, both to minimize the bleedoff needed to keep oil concentration in the cooler low and to reduce system losses from large stills. A highly efficient discharge gas/oil separator can be used for this purpose.
At low temperatures, periodic warm-up of the evaporator allows recovery of oil accumulation in the chiller. If continuous operation is required, dual chillers may be needed to deoil an oil-laden evaporator, or an oil-free compressor may be used.
Direct-Expansion Fluid Chillers
For details on these chillers, see Chapter 43 in the 2016 ASHRAE Handbook—HVAC Systems and Equipment. Figure 10 shows typical piping connections for a multicircuit direct-expansion (DX) chiller. Each circuit contains its own thermostatic expansion and solenoid valves. One solenoid valve can be wired to close at reduced system capacity. The thermostatic expansion valve bulbs should be located between the cooler and the liquid-suction interchanger, if used. Locating the bulb downstream from the interchanger can cause excessive cycling of the thermostatic expansion valve because the flow of high-pressure liquid through the interchanger ceases when the thermostatic expansion valve closes; consequently, no heat is available from the high-pressure liquid, and the cooler must starve itself to obtain the superheat necessary to open the valve. When the valve does open, excessive superheat causes it to overfeed until the bulb senses liquid downstream from the interchanger. Therefore, the remote bulb should be positioned between the cooler and the interchanger.
Figure 11 shows a typical piping arrangement that has been successful in packaged water chillers with DX coolers. With this arrangement, automatic recycling pumpdown is needed on the lag compressor to prevent leakage through compressor valves, allowing migration to the cold evaporator circuit. It also prevents liquid from slugging the compressor at start-up.
On larger systems, the limited size of thermostatic expansion valves may require use of a pilot-operated liquid valve controlled by a small thermostatic expansion valve (Figure 12). The equalizing connection and bulb of the pilot thermostatic expansion valve should be treated as a direct-acting thermal expansion valve. A small solenoid valve in the pilot line shuts off the high side from the low during shutdown. However, the main liquid valve does not open and close instantaneously.
Direct-Expansion Air Coils
For further information on these coils, see Chapter 23 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment. The most common ways of arranging DX coils are shown in Figures 13 and 14. The method shown in Figure 14 provides the superheat needed to operate the thermostatic expansion valve and is effective for heat transfer because leaving air contacts the coldest evaporator surface. This arrangement is advantageous on low-temperature applications, where the coil pressure drop represents an appreciable change in evaporating temperature.
Direct-expansion air coils can be located in any position as long as proper refrigerant distribution and continuous oil removal facilities are provided.
Figure 13 shows top-feed, free-draining piping with a vertical up-airflow coil. In Figure 14, which shows a horizontal-airflow coil, suction is taken off the bottom header connection, providing free oil draining. Many coils are supplied with connections at each end of the suction header so that a free-draining connection can be used regardless of which side of the coil is up; the other end is then capped.
In Figure 15, a refrigerant upfeed coil is used with a vertical downflow air arrangement. Here, the coil design must provide sufficient gas velocity to entrain oil at lowest loadings and to carry it into the suction line.
Pumpdown compressor control is desirable on all systems using downfeed or upfeed evaporators, to protect the compressor against a liquid slugback in cases where liquid can accumulate in the suction header and/or the coil on system off cycles. Pumpdown compressor control is described in the section on Keeping Liquid from Crankcase During Off Cycles.
Thermostatic expansion valve operation and application are described in Chapter 11. Thermostatic expansion valves should be sized carefully to avoid undersizing at full load and oversizing at partial load. The refrigerant pressure drops through the system (distributor, coil, condenser, and refrigerant lines, including liquid lifts) must be properly evaluated to determine the correct pressure drop available across the valve on which to base the selection. Variations in condensing pressure greatly affect the pressure available across the valve, and hence its capacity.
Oversized thermostatic expansion valves result in cycling that alternates flooding and starving the coil. This occurs because the valve attempts to throttle at a capacity below its capability, which causes periodic flooding of the liquid back to the compressor and wide temperature variations in the air leaving the coil. Reduced compressor capacity further aggravates this problem. Systems having multiple coils can use solenoid valves located in the liquid line feeding each evaporator or group of evaporators to close them off individually as compressor capacity is reduced.
For information on defrosting, see Chapter 14.
Flooded evaporators may be desirable when a small temperature differential is required between the refrigerant and the medium being cooled. A small temperature differential is advantageous in low-temperature applications.
In a flooded evaporator, the coil is kept full of refrigerant when cooling is required. The refrigerant level is generally controlled through a high- or low-side float control. Figure 16 represents a typical arrangement showing a low-side float control, oil return line, and heat interchanger.
Circulation of refrigerant through the evaporator depends on gravity and a thermosiphon effect. A mixture of liquid refrigerant and vapor returns to the surge tank, and the vapor flows into the suction line. A baffle installed in the surge tank helps prevent foam and liquid from entering the suction line. A liquid refrigerant circulating pump (Figure 17) provides a more positive way of obtaining a high circulation rate.
Taking the suction line off the top of the surge tank causes difficulties if no special provisions are made for oil return. For this reason, the oil return lines in Figure 16 should be installed. These lines are connected near the bottom of the float chamber and also just below the liquid level in the surge tank (where an oil-rich liquid refrigerant exists). They extend to a lower point on the suction line to allow gravity flow. Included in this oil return line is (1) a solenoid valve that is open only while the compressor is running and (2) a metering valve that is adjusted to allow a constant but small-volume return to the suction line. A liquid-line sight glass may be installed downstream from the metering valve to serve as a convenient check on liquid being returned.
Oil can be returned satisfactorily by taking a bleed of refrigerant and oil from the pump discharge (Figure 17) and feeding it to the heated oil receiver. If a low-side float is used, a jet ejector can be used to remove oil from the quiescent float chamber.
7. DISCHARGE (HOT-GAS) LINES
Hot-gas lines should be designed to
Avoid trapping oil at part-load operation
Prevent condensed refrigerant and oil in the line from draining back to the head of the compressor
Have carefully selected connections from a common line to multiple compressors
Avoid developing excessive noise or vibration from hot-gas pulsations, compressor vibration, or both
Oil Transport up Risers at Normal Loads. Although a low pressure drop is desired, oversized hot-gas lines can reduce gas velocities to a point where the refrigerant will not transport oil. Therefore, when using multiple compressors with capacity control, hot-gas risers must transport oil at all possible loadings.
Minimum Gas Velocities for Oil Transport in Risers. Minimum capacities for oil entrainment in hot-gas line risers are shown in Table 20. On multiple-compressor installations, the lowest possible system loading should be calculated and a riser size selected to give at least the minimum capacity indicated in the table for successful oil transport.
In some installations with multiple compressors and with capacity control, a vertical hot-gas line, sized to transport oil at minimum load, has excessive pressure drop at maximum load. When this problem exists, either a double riser or a single riser with an oil separator can be used.
Double Hot-Gas Risers. A double hot-gas riser can be used the same way it is used in a suction line. Figure 18 shows the double riser principle applied to a hot-gas line. Its operating principle and sizing technique are described in the section on Double Suction Risers.
Single Riser and Oil Separator. As an alternative, an oil separator in the discharge line just before the riser allows sizing the riser for a low pressure drop. Any oil draining back down the riser accumulates in the oil separator. With large multiple compressors, separator capacity may dictate the use of individual units for each compressor located between the discharge line and the main discharge header. Horizontal lines should be level or pitched downward in the direction of gas flow to facilitate travel of oil through the system and back to the compressor.
Piping to Prevent Liquid and Oil from Draining to Compressor Head. Whenever the condenser is located above the compressor, the hot-gas line should be trapped near the compressor before rising to the condenser, especially if the hot-gas riser is long. This minimizes the possibility of refrigerant, condensed in the line during off cycles, draining back to the head of the compressor. Also, any oil traveling up the pipe wall will not drain back to the compressor head.
The loop in the hot-gas line (Figure 19) serves as a reservoir and traps liquid resulting from condensation in the line during shutdown, thus preventing gravity drainage of liquid and oil back to the compressor head. A small high-pressure float drainer should be installed at the bottom of the trap to drain any significant amount of refrigerant condensate to a low-side component such as a suction accumulator or low-pressure receiver. This float prevents excessive build-up of liquid in the trap and possible liquid hammer when the compressor is restarted.
For multiple-compressor arrangements, each discharge line should have a check valve to prevent gas from active compressors from condensing on heads of idle compressors.
For single-compressor applications, a tightly closing check valve should be installed in the hot-gas line of the compressor whenever the condenser and the receiver ambient temperature are higher than that of the compressor. The check valve prevents refrigerant from boiling off in the condenser or receiver and condensing on the compressor heads during off cycles.
This check valve should be a piston type, which closes by gravity when the compressor stops running. A spring-loaded check may incur chatter (vibration), particularly on slow-speed reciprocating compressors.
For compressors equipped with water-cooled oil coolers, a water solenoid and water-regulating valve should be installed in the water line so that the regulating valve maintains adequate cooling during operation, and the solenoid stops flow during the off cycle to prevent localized condensing of the refrigerant.
Hot-Gas (Discharge) Mufflers. Mufflers can be installed in hot-gas lines to dampen discharge gas pulsations, reducing vibration and noise. Mufflers should be installed in a horizontal or downflow portion of the hot-gas line immediately after it leaves the compressor.
Because gas velocity through the muffler is substantially lower than that through the hot-gas line, the muffler may form an oil trap. The muffler should be installed to allow oil to flow through it and not be trapped.
8. DEFROST GAS SUPPLY LINES
Sizing refrigeration lines to supply defrost gas to one or more evaporators is not an exact science. The parameters associated with sizing the defrost gas line are related to allowable pressure drop and refrigerant flow rate during defrost.
Engineers use an estimated two times the evaporator load for effective refrigerant flow rate to determine line sizing requirements. Pressure drop is not as critical during the defrost cycle, and many engineers use velocity as the criterion for determining line size. The effective condensing temperature and average temperature of the gas must be determined. The velocity determined at saturated conditions gives a conservative line size.
Controlled testing (Stoecker 1984) showed that, in small coils with R-22, the defrost flow rate tends to be higher as the condensing temperature increases. The flow rate is on the order of two to three times the normal evaporator flow rate, which supports the estimated two times used by practicing engineers.
9. HEAT EXCHANGERS AND VESSELS
Refrigerant receivers are vessels used to store excess refrigerant circulated throughout the system. Their purpose is to
Provide pumpdown storage capacity when another part of the system must be serviced or the system must be shut down for an extended time. In some water-cooled condenser systems, the condenser also serves as a receiver if the total refrigerant charge does not exceed its storage capacity.
Handle the excess refrigerant charge needed by air-cooled condensers that require flooding to maintain minimum condensing pressures (see the section on Head Pressure Control for Refrigerant Condensers).
Receive refrigerant draining from the condenser, to allow the condenser to maintain its usable surface area for condensing.
Accommodate a fluctuating charge in the low side on systems where the operating charge in the evaporator varies for different loading conditions. When an evaporator is fed with a thermal expansion valve, hand expansion valve, or low-pressure float, the operating charge in the evaporator varies considerably depending on the loading. During low load, the evaporator requires a larger charge because boiling is not as intense. When load increases, the operating charge in the evaporator decreases, and the receiver must store excess refrigerant.
Connections for Through-Type Receiver. When a through-type receiver is used, liquid must always flow from condenser to receiver. Pressure in the receiver must be lower than that in the condenser outlet. The receiver and its associated piping provide free flow of liquid from the condenser to the receiver by equalizing pressures between the two so that the receiver cannot build up a higher pressure than the condenser.
If a vent is not used, piping between condenser and receiver (condensate line) is sized so that liquid flows in one direction and gas flows in the opposite direction. Sizing the condensate line for 100 fpm liquid velocity is usually adequate to attain this flow. Piping should slope at least 0.25 in/ft and eliminate any natural liquid traps. Figure 20 shows this configuration.
Piping between the condenser and the receiver can be equipped with a separate vent (equalizer) line to allow receiver and condenser pressures to equalize. This external vent line can be piped either with or without a check valve in the vent line (see Figures 22 and 23). If there is no check valve, prevent discharge gas from discharging directly into the vent line; this should prevent a gas velocity pressure component from being introduced on top of the liquid in the receiver. When the piping configuration is unknown, install a check valve in the vent with flow in the direction of the condenser. The check valve should be selected for minimum opening pressure (i.e., approximately 0.5 psi). When determining condensate drop leg height, allowance must be made to overcome both the pressure drop across this check valve and the refrigerant pressure drop through the condenser. This ensures that there will be no liquid back-up into an operating condenser on a multiple-condenser application when one or more of the condensers is idle. The condensate line should be sized so that velocity does not exceed 150 fpm.
The vent line flow is from receiver to condenser when receiver temperature is higher than condensing temperature. Flow is from condenser to receiver when air temperature around the receiver is below condensing temperature. Flow rate depends on this temperature difference as well as on the receiver surface area. Vent size can be calculated from this flow rate.
Connections for Surge-Type Receiver. The purpose of a surge-type receiver is to allow liquid to flow to the expansion valve without exposure to refrigerant in the receiver, so that it can remain subcooled. The receiver volume is available for liquid that is to be removed from the system. Figure 21 shows an example of connections for a surge-type receiver. Height h must be adequate for a liquid pressure at least as large as the pressure loss through the condenser, liquid line, and vent line at the maximum temperature difference between the receiver ambient and the condensing temperature. Condenser pressure drop at the greatest expected heat rejection should be obtained from the manufacturer. The minimum value of h can then be calculated to determine whether the available height will allow the surge-type receiver.
Multiple Condensers. Two or more condensers connected in series or in parallel can be used in a single refrigeration system. If connected in series, the pressure losses through each condenser must be added. Condensers are more often arranged in parallel. Pressure loss through any one of the parallel circuits is always equal to that through any of the others, even if it results in filling much of one circuit with liquid while gas passes through another.
Figure 22 shows a basic arrangement for parallel condensers with a through-type receiver. Condensate drop legs must be long enough to allow liquid levels in them to adjust to equalize pressure losses between condensers at all operating conditions. Drop legs should be 6 to 12 in. higher than calculated to ensure that liquid outlets drain freely. This height provides a liquid pressure to offset the largest condenser pressure loss. The liquid seal prevents gas blow-by between condensers.
Large single condensers with multiple coil circuits should be piped as though the independent circuits were parallel condensers. For example, if the left condenser in Figure 22 has 2 psi more pressure drop than the right condenser, the liquid level on the left is about 4 ft higher than that on the right. If the condensate lines do not have enough vertical height for this level difference, liquid will back up into the condenser until pressure drop is the same through both circuits. Enough surface may be covered to reduce condenser capacity significantly.
Condensate drop legs should be sized based on 150 fpm velocity. The main condensate lines should be based on 100 fpm. Depending on prevailing local and/or national safety codes, a relief device may have to be installed in the discharge piping.
Figure 23 shows a piping arrangement for parallel condensers with a surge-type receiver. When the system is operating at reduced load, flow paths through the circuits may not be symmetrical. Small pressure differences are not unusual; therefore, the liquid line junction should be about 2 or 3 ft below the bottom of the condensers. The exact amount can be calculated from pressure loss through each path at all possible operating conditions.
When condensers are water cooled, a single automatic water valve for the condensers in one refrigeration system should be used. Individual valves for each condenser in a single system cannot maintain the same pressure and corresponding pressure drops.
With evaporative condensers (Figure 24), pressure loss may be high. If parallel condensers are alike and all are operated, the differences may be small, and condenser outlets need not be more than 2 or 3 ft above the liquid line junction. If fans on one condenser are not operated while the fans on another condenser are, then the liquid level in the one condenser must be high enough to compensate for the pressure drop through the operating condenser.
When the available level difference between condenser outlets and the liquid-line junction is sufficient, the receiver may be vented to the condenser inlets (Figure 25). In this case, the surge-type receiver can be used. The level difference must then be at least equal to the greatest loss through any condenser circuit plus the greatest vent line loss when the receiver ambient is greater than the condensing temperature.
Refrigerant pressure drop through air-cooled condensers must be obtained from the supplier for the particular unit at the specified load. If refrigerant pressure drop is low enough and the arrangement is practical, parallel condensers can be connected to allow for capacity reduction to zero on one condenser without causing liquid back-up in active condensers (Figure 26). Multiple condensers with high pressure drops can be connected as shown in Figure 26, provided that (1) the ambient at the receiver is equal to or lower than the inlet air temperature to the condenser; (2) capacity control affects all units equally; (3) all units operate when one operates, unless valved off at both inlet and outlet; and (4) all units are of equal size.
A single condenser with any pressure drop can be connected to a receiver without an equalizer and without trapping height if the condenser outlet and the line from it to the receiver can be sized for sewer flow without a trap or restriction, using a maximum velocity of 100 fpm. A single condenser can also be connected with an equalizer line to the hot-gas inlet if the vertical drop leg is sufficient to balance refrigerant pressure drop through the condenser and liquid line to the receiver.
If unit sizes are unequal, additional liquid height H, equivalent to the difference in full-load pressure drop, is required. Usually, condensers of equal size are used in parallel applications.
If the receiver cannot be located in an ambient temperature below the inlet air temperature for all operating conditions, sufficient extra height of drop leg H is required to overcome the equivalent differences in saturation pressure of the receiver and the condenser. Subcooling by the liquid leg tends to condense vapor in the receiver to reach a balance between rate of condensation, at an intermediate saturation pressure, and heat gain from ambient to the receiver. A relatively large liquid leg is required to balance a small temperature difference; therefore, this method is probably limited to marginal cases. Liquid leaving the receiver is nonetheless saturated, and any subcooling to prevent flashing in the liquid line must be obtained downstream of the receiver. If the temperature of the receiver ambient is above the condensing pressure only at part-load conditions, it may be acceptable to back liquid into the condensing surface, sacrificing the operating economy of lower part-load head pressure for a lower liquid leg requirement. The receiver must be adequately sized to contain a minimum of the backed-up liquid so that the condenser can be fully drained when full load is required. If a low-ambient control system of backing liquid into the condenser is used, consult the system supplier for proper piping.
10. REFRIGERATION ACCESSORIES
Liquid-Suction Heat Exchangers
Generally, liquid-suction heat exchangers subcool liquid refrigerant and superheat suction gas. They are used for one or more of the following functions:
Increasing efficiency of the refrigeration cycle. Efficiency of the thermodynamic cycle of certain halocarbon refrigerants can be increased when the suction gas is superheated by removing heat from the liquid. This increased efficiency must be evaluated against the effect of pressure drop through the suction side of the exchanger, which forces the compressor to operate at a lower suction pressure. Liquid-suction heat exchangers are most beneficial at low suction temperatures. The increase in cycle efficiency for systems operating in the air-conditioning range (down to about 30°F evaporating temperature) usually does not justify their use. The heat exchanger can be located wherever convenient.
Subcooling liquid refrigerant to prevent flash gas at the expansion valve. The heat exchanger should be located near the condenser or receiver to achieve subcooling before pressure drop occurs.
Evaporating small amounts of expected liquid refrigerant returning from evaporators in certain applications. Many heat pumps incorporating reversals of the refrigerant cycle include a suction-line accumulator and liquid-suction heat exchanger arrangement to trap liquid floodbacks and vaporize them slowly between cycle reversals.
If an evaporator design makes a deliberate slight overfeed of refrigerant necessary, either to improve evaporator performance or to return oil out of the evaporator, a liquid-suction heat exchanger is needed to evaporate the refrigerant.
A flooded water cooler usually incorporates an oil-rich liquid bleed from the shell into the suction line for returning oil. The liquid-suction heat exchanger boils liquid refrigerant out of the mixture in the suction line. Exchangers used for this purpose should be placed in a horizontal run near the evaporator. Several types of liquid-suction heat exchangers are used.
Liquid and Suction Line Soldered Together. The simplest form of heat exchanger is obtained by strapping or soldering the suction and liquid lines together to obtain counterflow and then insulating the lines as a unit. To maximize capacity, the liquid line should always be on the bottom of the suction line, because liquid in a suction line runs along the bottom (Figure 27). This arrangement is limited by the amount of suction line available.
Shell-and-Coil or Shell-and-Tube Heat Exchangers (Figure 28). These units are usually installed so that the suction outlet drains the shell. When the units are used to evaporate liquid refrigerant returning in the suction line, the free-draining arrangement is not recommended. Liquid refrigerant can run along the bottom of the heat exchanger shell, having little contact with the warm liquid coil, and drain into the compressor. By installing the heat exchanger at a slight angle to the horizontal (Figure 29) with gas entering at the bottom and leaving at the top, any liquid returning in the line is trapped in the shell and held in contact with the warm liquid coil, where most of it is vaporized. An oil return line, with a metering valve and solenoid valve (open only when the compressor is running), is required to return oil that collects in the trapped shell.
Concentric Tube-in-Tube Heat Exchangers. The tube-in-tube heat exchanger is not as efficient as the shell-and-finned-coil type. It is, however, quite suitable for cleaning up small amounts of excessive liquid refrigerant returning in the suction line. Figure 30 shows typical construction with available pipe and fittings.
Plate Heat Exchangers. Plate heat exchangers provide high-efficiency heat transfer. They are very compact, have low pressure drop, and are lightweight devices. They are good for use as liquid subcoolers.
For air-conditioning applications, heat exchangers are recommended for liquid subcooling or for clearing up excess liquid in the suction line. For refrigeration applications, heat exchangers are recommended to increase cycle efficiency, as well as for liquid subcooling and removing small amounts of excess liquid in the suction line. Excessive superheating of the suction gas should be avoided.
To take full advantage of the two-stage system, the refrigerant liquid should be cooled to near the interstage temperature to reduce the amount of flash gas handled by the low-stage compressor. The net result is a reduction in total system power requirements. The amount of gain from cooling to near interstage conditions varies among refrigerants.
Figure 31 shows an open or flash-type cooler. This is the simplest and least costly type, which has the advantage of cooling liquid to the saturation temperature of the interstage pressure. One disadvantage is that the pressure of cooled liquid is reduced to interstage pressure, leaving less pressure available for liquid transport. Although the liquid temperature is reduced, the pressure drops correspondingly, and the expansion device controlling flow to the cooler must be large enough to pass all the liquid refrigerant flow. Failure of this valve could allow a large flow of liquid to the upper-stage compressor suction, which could seriously damage the compressor.
Liquid from a flash cooler is saturated, and liquid from a cascade condenser usually has little subcooling. In both cases, the liquid temperature is usually lower than the temperature of the surroundings. Thus, it is important to avoid heat input and pressure losses that would cause flash gas to form in the liquid line to the expansion device or to recirculating pumps. Cold liquid lines should be insulated, because expansion devices are usually designed to feed liquid, not vapor.
Figure 32 shows the closed or heat exchanger type of subcooler. It should have sufficient heat transfer surface to transfer heat from the liquid to the evaporating refrigerant with a small final temperature difference. Pressure drop should be small, so that full pressure is available for feeding liquid to the expansion device at the low-temperature evaporator. The subcooler liquid control valve should be sized to supply only the quantity of refrigerant required for the subcooling. This prevents a tremendous quantity of liquid from flowing to the upper-stage suction in the event of a valve failure.
Discharge Line Oil Separators
Oil is always in circulation in systems using halocarbon refrigerants. Refrigerant piping is designed to ensure that this oil passes through the entire system and returns to the compressor as fast as it leaves. Although well-designed piping systems can handle the oil in most cases, a discharge-line oil separator can have certain advantages in some applications (see Chapter 11), such as
In systems where it is impossible to prevent substantial absorption of refrigerant in the crankcase oil during shutdown periods. When the compressor starts up with a violent foaming action, oil is thrown out at an accelerated rate, and the separator immediately returns a large portion of this oil to the crankcase. Normally, the system should be designed with pumpdown control or crankcase heaters to minimize liquid absorption in the crankcase.
In systems using flooded evaporators, where refrigerant bleedoff is necessary to remove oil from the evaporator. Oil separators reduce the amount of bleedoff from the flooded cooler needed for operation.
In direct-expansion systems using coils or tube bundles that require bottom feed for good liquid distribution and where refrigerant carryover from the top of the evaporator is essential for proper oil removal.
In low-temperature systems, where it is advantageous to have as little oil as possible going through the low side.
In screw-type compressor systems, where an oil separator is necessary for proper operation. The oil separator is usually supplied with the compressor unit assembly directly from the compressor manufacturer.
In multiple compressors operating in parallel. The oil separator can be an integral part of the total system oil management system.
In applying oil separators in refrigeration systems, the following potential hazards must be considered:
Oil separators are not 100% efficient, and they do not eliminate the need to design the complete system for oil return to the compressor.
Oil separators tend to condense out liquid refrigerant during compressor off cycles and on compressor start-up. This is true if the condenser is in a warm location, such as on a roof. During the off cycle, the oil separator cools down and acts as a condenser for refrigerant that evaporates in warmer parts of the system. A cool oil separator may condense discharge gas and, on compressor start-up, automatically drain it into the compressor crankcase. To minimize this possibility, the drain connection from the oil separator can be connected into the suction line. This line should be equipped with a shutoff valve, a fine filter, hand throttling and solenoid valves, and a sight glass. The throttling valve should be adjusted so that flow through this line is only a little greater than would normally be expected to return oil through the suction line.
The float valve is a mechanical device that may stick open or closed. If it sticks open, hot gas will continuously bypass to the compressor crankcase. If the valve sticks closed, no oil is returned to the compressor. To minimize this problem, the separator can be supplied without an internal float valve. A separate external float trap can then be located in the oil drain line from the separator preceded by a filter. Shutoff valves should isolate the filter and trap. The filter and traps are also easy to service without stopping the system.
The discharge line pipe size into and out of the oil separator should be the full size determined for the discharge line. For separators that have internal oil float mechanisms, allow enough room to remove the oil float assembly for servicing.
Depending on system design, the oil return line from the separator may feed to one of the following locations:
Directly to the compressor crankcase
Directly into the suction line ahead of the compressor
Into an oil reservoir or device used to collect oil, used for a specifically designed oil management system
When a solenoid valve is used in the oil return line, the valve should be wired so that it is open when the compressor is running. To minimize entrance of condensed refrigerant from the low side, a thermostat may be installed and wired to control the solenoid in the oil return line from the separator. The thermostat sensing element should be located on the oil separator shell below the oil level and set high enough so that the solenoid valve will not open until the separator temperature is higher than the condensing temperature. A superheat-controlled expansion valve can perform the same function. If a discharge line check valve is used, it should be downstream of the oil separator.
Surge Drums or Accumulators
A surge drum is required on the suction side of almost all flooded evaporators to prevent liquid slopover to the compressor. Exceptions include shell-and-tube coolers and similar shell-type evaporators, which provide ample surge space above the liquid level or contain eliminators to separate gas and liquid. A horizontal surge drum is sometimes used where headroom is limited.
The drum can be designed with baffles or eliminators to separate liquid from the suction gas. More often, sufficient separation space is allowed above the liquid level for this purpose. Usually, the design is vertical, with a separation height above the liquid level of 24 to 30 in. and with the shell diameter sized to keep suction gas velocity low enough to allow liquid droplets to separate. Because these vessels are also oil traps, it is necessary to provide oil bleed.
Although separators may be fabricated with length-to-diameter (L/D) ratios of 1/1 up to 10/1, the lowest-cost separators are usually for L/D ratios between 3/1 and 5/1.
Compressor Floodback Protection
Certain systems periodically flood the compressor with excessive amounts of liquid refrigerant. When periodic floodback through the suction line cannot be controlled, the compressor must be protected against it.
The most satisfactory method appears to be a trap arrangement that catches liquid floodback and (1) meters it slowly into the suction line, where the floodback is cleared up with a liquid-suction heat interchanger; (2) evaporates the liquid 100% in the trap itself by using a liquid coil or electric heater, and then automatically returns oil to the suction line; or (3) returns it to the receiver or to one of the evaporators. Figure 29 shows an arrangement that handles moderate liquid floodback, disposing of liquid by a combination of boiling off in the exchanger and limited bleedoff into the suction line. This device, however, does not have sufficient trapping volume for most heat pump applications or hot-gas defrost systems using reversal of the refrigerant cycle.
For heavier floodback, a larger volume is required in the trap. The arrangement shown in Figure 33 has been used successfully in reverse-cycle heat pump applications using halocarbon refrigerants. It consists of a suction-line accumulator with enough volume to hold the maximum expected floodback and a large enough diameter to separate liquid from suction gas. Trapped liquid is slowly bled off through a properly sized and controlled drain line into the suction line, where it is boiled off in a liquid-suction heat exchanger between cycle reversals.
With the alternative arrangement shown, the liquid/oil mixture is heated to evaporate the refrigerant, and the remaining oil is drained into the crankcase or suction line.
Refrigerant Driers and Moisture Indicators
The effect of moisture in refrigeration systems is discussed in Chapters 6 and 7. Using a permanent refrigerant drier is recommended on all systems and with all refrigerants. It is especially important on low-temperature systems to prevent ice from forming at expansion devices. A full-flow drier is always recommended in hermetic compressor systems to keep the system dry and prevent decomposition products from getting into the evaporator in the event of a motor burnout.
Replaceable-element filter-driers are preferred for large systems because the drying element can be replaced without breaking any refrigerant connections. The drier is usually located in the liquid line near the liquid receiver. It may be mounted horizontally or vertically with the flange at the bottom, but it should never be mounted vertically with the flange on top because any loose material would then fall into the line when the drying element was removed.
A three-valve bypass is usually used, as shown in Figure 34, to provide a way to isolate the drier for servicing. The refrigerant charging connection should be located between the receiver outlet valve and liquid-line drier so that all refrigerant added to the system passes through the drier.
Reliable moisture indicators can be installed in refrigerant liquid lines to provide a positive indication of when the drier cartridge should be replaced.
Strainers should be used in both liquid and suction lines to protect automatic valves and the compressor from foreign material, such as pipe welding scale, rust, and metal chips. The strainer should be mounted in a horizontal line, oriented so that the screen can be replaced without loose particles falling into the system.
A liquid-line strainer should be installed before each automatic valve to prevent particles from lodging on the valve seats. Where multiple expansion valves with internal strainers are used at one location, a single main liquid-line strainer will protect all of these. The liquid-line strainer can be located anywhere in the line between the condenser (or receiver) and the automatic valves, preferably near the valves for maximum protection. Strainers should trap the particle size that could affect valve operation. With pilot-operated valves, a very fine strainer should be installed in the pilot line ahead of the valve.
Filter-driers dry the refrigerant and filter out particles far smaller than those trapped by mesh strainers. No other strainer is needed in the liquid line if a good filter-drier is used.
Refrigeration compressors are usually equipped with a built-in suction strainer, which is adequate for the usual system with copper piping. The suction line should be piped at the compressor so that the built-in strainer is accessible for servicing.
Both liquid- and suction-line strainers should be adequately sized to ensure sufficient foreign material storage capacity without excessive pressure drop. In steel piping systems, an external suction-line strainer is recommended in addition to the compressor strainer.
Every refrigeration system should have a way to check for sufficient refrigerant charge. Common devices used are liquid-line sight glass, mechanical or electronic indicators, and an external gage glass with equalizing connections and shutoff valves. A properly installed sight glass shows bubbling when the charge is insufficient.
Liquid indicators should be located in the liquid line as close as possible to the receiver outlet, or to the condenser outlet if no receiver is used (Figure 35). The sight glass is best installed in a vertical section of line, far enough downstream from any valve that the resulting disturbance does not appear in the glass. If the sight glass is installed too far away from the receiver, the line pressure drop may be sufficient to cause flashing and bubbles in the glass, even if the charge is sufficient for a liquid seal at the receiver outlet.
When sight glasses are installed near the evaporator, often no amount of system overcharging will give a solid liquid condition at the sight glass because of pressure drop in the liquid line or lift. Subcooling is required here. An additional sight glass near the evaporator may be needed to check the refrigerant condition at that point.
Sight glasses should be installed full size in the main liquid line. In very large liquid lines, this may not be possible; the glass can then be installed in a bypass or saddle mount that is arranged so that any gas in the liquid line will tend to move to it. A sight glass with double ports (for back lighting) and seal caps, which provide added protection against leakage, is preferred. Moisture-liquid indicators large enough to be installed directly in the liquid line serve the dual purpose of liquid-line sight glass and moisture indicator.
Oil receivers serve as reservoirs for replenishing crankcase oil pumped by the compressors and provide the means to remove refrigerant dissolved in the oil. They are selected for systems having any of the following components:
Flooded or semiflooded evaporators with large refrigerant charges
Two or more compressors operated in parallel
Long suction and discharge lines
Double suction line risers
A typical hookup is shown in Figure 33. Outlets are arranged to prevent oil from draining below the heater level to avoid heater burnout and to prevent scale and dirt from being returned to the compressor.
Noncondensable gas separation using a purge unit is useful on most large refrigeration systems where suction pressure may fall below atmospheric pressure (see Figure 30 of Chapter 2).
11. HEAD PRESSURE CONTROL FOR REFRIGERANT CONDENSERS
For more information on head pressure control, see Chapter 39 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment.
With water-cooled condensers, head pressure controls are used both to maintain condensing pressure and to conserve water. On cooling tower applications, they are used only where it is necessary to maintain condensing temperatures.
Condenser-Water-Regulating Valves
The shutoff pressure of the valve must be set slightly higher than the saturation pressure of the refrigerant at the highest ambient temperature expected when the system is not in operation. This ensures that the valve will not pass water during off cycles. These valves are usually sized to pass the design quantity of water at about a 25 to 30 psi difference between design condensing pressure and valve shutoff pressure. Chapter 11 has further information.
In cooling tower applications, a simple bypass with a manual or automatic valve responsive to head pressure change can also be used to maintain condensing pressure. Figure 36 shows an automatic three-way valve arrangement. The valve divides water flow between the condenser and the bypass line to maintain the desired condensing pressure. This maintains a balanced flow of water on the tower and pump.
Among the methods used for condensing pressure control with evaporative condensers are (1) cycling the spray pump motor; (2) cycling both fan and spray pump motors; (3) throttling the spray water; (4) bypassing air around duct and dampers; (5) throttling air via dampers, on either inlet or discharge; and (6) combinations of these methods. For further information, see Chapter 39 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment.
In water pump cycling, a pressure control at the gas inlet starts and stops the pump in response to head pressure changes. The pump sprays water over the condenser coils. As head pressure drops, the pump stops and the unit becomes an air-cooled condenser.
Constant pressure is difficult to maintain with coils of prime surface tubing because as soon as the pump stops, the pressure goes up and the pump starts again. This occurs because these coils have insufficient capacity when operating as an air-cooled condenser. The problem is not as acute with extended-surface coils. Short-cycling results in excessive deposits of mineral and scale on the tubes, decreasing the life of the water pump.
One method of controlling head pressure is using cycle fans and pumps. This minimizes water-side scaling. In colder climates, an indoor water sump with a remote spray pump(s) is required. The fan cycling sequence is as follows:
Upon dropping head pressure
Upon rising head pressure
Damper control (Figure 37) may be incorporated in systems requiring more constant head pressures (e.g., some systems using thermostatic expansion valves). One drawback of dampers is formation of ice on dampers and linkages.
Figure 38 incorporates an air bypass arrangement for controlling head pressure. A modulating motor, acting in response to a modulating pressure control, positions dampers so that the mixture of recirculated and cold inlet air maintains the desired pressure. In extremely cold weather, most of the air is recirculated.
Methods for condensing pressure control with air-cooled condensers include (1) cycling fan motor, (2) air throttling or bypassing, (3) coil flooding, and (4) fan motor speed control. The first two methods are described in the section on Evaporative Condensers.
The third method holds condensing pressure up by backing liquid refrigerant up in the coil to cut down on effective condensing surface. When head pressure drops below the setting of the modulating control valve, it opens, allowing discharge gas to enter the liquid drain line. This restricts liquid refrigerant drainage and causes the condenser to flood enough to maintain the condenser and receiver pressure at the control valve setting. A pressure difference must be available across the valve to open it. Although the condenser imposes sufficient pressure drop at full load, pressure drop may practically disappear at partial loading. Therefore, a positive restriction must be placed parallel with the condenser and the control valve. Systems using this type of control require extra refrigerant charge.
In multiple-fan air-cooled condensers, it is common to cycle fans off down to one fan and then to apply air throttling to that section or modulate the fan motor speed. Consult the manufacturer before using this method, because not all condensers are properly circuited for it.
Using ambient temperature change (rather than condensing pressure) to modulate air-cooled condenser capacity prevents rapid cycling of condenser capacity. A disadvantage of this method is that the condensing pressure is not closely controlled.
The methods for low-ambient, condensing pressure control for microchannel condensers are essentially the same as those used for standard air-cooled condensers. However, because most microchannel condensers are made up of many individual heat exchangers, there is an opportunity to mechanically isolate portions of the condenser to reduce the usable surface area. This type of control scheme can be used instead of holding back excess refrigerant to flood portions of the condenser.
12. KEEPING LIQUID FROM CRANKCASE DURING OFF CYCLES
Control of reciprocating compressors should prevent excessive accumulation of liquid refrigerant in the crankcase during off cycles. Any one of the following control methods accomplishes this.
Automatic Pumpdown Control (Direct-Expansion Air-Cooling Systems)
The most effective way to keep liquid out of the crankcase during system shutdown is to operate the compressor on automatic pumpdown control. The recommended arrangement involves the following devices and provisions:
A liquid-line solenoid valve in the main liquid line or in the branch to each evaporator.
Compressor operation through a low-pressure cutout providing for pumpdown whenever this device closes, regardless of whether the balance of the system is operating.
Electrical interlock of the liquid solenoid valve with the evaporator fan, so refrigerant flow stops when the fan is out of operation.
Electrical interlock of refrigerant solenoid valve with safety devices (e.g., high-pressure cutout, oil safety switch, and motor overloads), so that the refrigerant solenoid valve closes when the compressor stops.
Low-pressure control settings such that the cut-in point corresponds to a saturated refrigerant temperature lower than any expected compressor ambient air temperature. If the cut-in setting is any higher, liquid refrigerant can accumulate and condense in the crankcase at a pressure corresponding to the ambient temperature. Then, crankcase pressure would not rise high enough to reach the cut-in point, and effective automatic pumpdown would not be obtained.
Crankcase Oil Heater (Direct-Expansion Systems)
A crankcase oil heater with or without single (nonrecycling) pumpout at the end of each operating cycle does not keep liquid refrigerant out of the crankcase as effectively as automatic pumpdown control, but many compressors equalize too quickly after stopping automatic pumpdown control. Crankcase oil heaters maintain the crankcase oil at a temperature higher than that of other parts of the system, minimizing absorption of the refrigerant by the oil.
Operation with the single pumpout arrangement is as follows. Whenever the temperature control device opens the circuit, or the manual control switch is opened for shutdown purposes, the crankcase heater is energized, and the compressor keeps running until it cuts off on the low-pressure switch. Because the crankcase heater remains energized during the complete off cycle, it is important that a continuous live circuit be available to the heater during the off time. The compressor cannot start again until the temperature control device or manual control switch closes, regardless of the position of the low-pressure switch.
This control method requires
A liquid-line solenoid valve in the main liquid line or in the branch to each evaporator
Use of a relay or the maintained contact of the compressor motor auxiliary switch to obtain a single pumpout operation before stopping the compressor
A relay or auxiliary starter contact to energize the crankcase heater during the compressor off cycle and deenergize it during the compressor on cycle
Electrical interlock of the refrigerant solenoid valve with the evaporator fan, so that refrigerant flow is stopped when the fan is out of operation
Electrical interlock of refrigerant solenoid valve with safety devices (e.g., high-pressure cutout, oil safety switch, and motor overloads), so that the refrigerant flow valve closes when the compressor stops
Control for Direct-Expansion Water Chillers
Automatic pumpdown control is undesirable for direct-expansion water chillers because freezing is possible if excessive cycling occurs. A crankcase heater is the best solution, with a solenoid valve in the liquid line that closes when the compressor stops.
Effect of Short Operating Cycle
With reciprocating compressors, oil leaves the crankcase at an accelerated rate immediately after starting. Therefore, each start should be followed by a long enough operating period to allow the oil level to recover. Controllers used for compressors should not produce short-cycling of the compressor. Refer to the compressor manufacturer’s literature for guidelines on maximum or minimum cycles for a specified period.
13. HOT-GAS BYPASS ARRANGEMENTS
Most large reciprocating compressors are equipped with unloaders that allow the compressor to start with most of its cylinders unloaded. However, it may be necessary to further unload the compressor to (1) reduce starting torque requirements so that the compressor can be started both with low-starting-torque prime movers and on low-current taps of reduced voltage starters and (2) allow capacity control down to 0% load conditions without stopping the compressor.
Full (100%) Unloading for Starting
Starting the compressor without load can be done with a manual or automatic valve in a bypass line between the hot-gas and suction lines at the compressor.
To prevent overheating, this valve is open only during the starting period and closed after the compressor is up to full speed and full voltage is applied to the motor terminals.
In the control sequence, the unloading bypass valve is energized on demand of the control calling for compressor operation, equalizing pressures across the compressor. After an adequate delay, a timing relay closes a pair of normally open contacts to start the compressor. After a further time delay, a pair of normally closed timing relay contacts opens, deenergizing the bypass valve.
Full (100%) Unloading for Capacity Control
Where full unloading is required for capacity control, hot-gas bypass arrangements can be used in ways that will not overheat the compressor. In using these arrangements, hot gas should not be bypassed until after the last unloading step.
Hot-gas bypass should (1) give acceptable regulation throughout the range of loads, (2) not cause excessive superheating of the suction gas, (3) not cause any refrigerant overfeed to the compressor, and (4) maintain an oil return to the compressor.
Hot-gas bypass for capacity control is an artificial loading device that maintains a minimum evaporating pressure during continuous compressor operation, regardless of evaporator load. This is usually done by an automatic or manual pressure-reducing valve that establishes a constant pressure on the downstream side.
Four common methods of using hot-gas bypass are shown in Figure 39. Figure 39A shows the simplest type; it will dangerously overheat the compressor if used for protracted periods of time. Figure 39B shows the use of hot-gas bypass to the exit of the evaporator. The expansion valve bulb should be placed at least 5 ft downstream from the bypass point of entrance, and preferably further, to ensure good mixing.
In Figure 39D, the hot-gas bypass enters after the evaporator thermostatic expansion valve bulb. Another thermostatic expansion valve supplies liquid directly to the bypass line for desuperheating. It is always important to install the hot-gas bypass far enough back in the system to maintain sufficient gas velocities in suction risers and other components to ensure oil return at any evaporator loading.
Figure 39C shows the most satisfactory hot-gas bypass arrangement. Here, the bypass is connected into the low side between the expansion valve and entrance to the evaporator. If a distributor is used, gas enters between the expansion valve and distributor. Refrigerant distributors are commercially available with side inlet connections that can be used for hot-gas bypass duty to a certain extent. Pressure drop through the distributor tubes must be evaluated to determine how much gas can be bypassed. This arrangement provides good oil return.
Solenoid valves should be placed before the constant-pressure bypass valve and before the thermal expansion valve used for liquid injection desuperheating, so that these devices cannot function until they are required.
Control valves for hot gas should be close to the main discharge line because the line preceding the valve usually fills with liquid when closed.
The hot-gas bypass line should be sized so that its pressure loss is only a small percentage of the pressure drop across the valve. Usually, it is the same size as the valve connections. When sizing the valve, consult a control valve manufacturer to determine the minimum compressor capacity that must be offset, refrigerant used, condensing pressure, and suction pressure.
When unloading (Figure 39C), head pressure control requirements increase considerably because the only heat delivered to the condenser is that caused by the motor power delivered to the compressor. Discharge pressure should be kept high enough that the hot-gas bypass valve can deliver gas at the required rate. The condenser head pressure control must be capable of meeting this condition.
14. MINIMIZING REFRIGERANT CHARGE IN COMMERCIAL SYSTEMS
Preventing refrigerant leaks is the most effective way to reduce halocarbons’ environmental effects. However, if a leak does occur, the consequences are reduced if the system charge has been minimized. There are many ways to reduce charge, but most require significant system modifications; consequently, charge reduction is usually performed during system remodeling or replacement.
One of the best opportunities to reduce refrigerant charge exists in the distribution piping that feeds liquid to the evaporator from the receiver and returns the suction gas to the compressor. Systems serving numerous evaporators across a facility (e.g., in supermarkets) rely on a network of distribution piping, which can contain a large portion of the entire system charge. For systems that use single circuiting, in which each evaporator (or small group of adjacent evaporators) has its own liquid and suction line piped back to the compressor, charge can be significantly reduced by zoning the loads. For loads operating at similar evaporator pressures, one suction and liquid line can run from the machinery room and branch out closer to the load to feed multiple evaporators (loop piping). Expansion, solenoid, and evaporator pressure regulating valves must be next to the heat load in these systems, but benefits beyond reduced charge include cheaper installation cost and less physical space required to run the lines. Note that using hot-gas defrost with this type of piping scheme is typically not preferred, because it requires a third branched line that must also be field installed.
In the liquid feed lines, subcooling the liquid can further reduce charge. Because subcooling increases the refrigerant’s quality, the required mass flow rate is reduced, thus allowing use of smaller liquid lines (and thus smaller-volume refrigerant charges) with comparable velocities and pressure drops. Subcooling is typically chosen for its energy benefits and is also often used to protect liquid from flashing before it reaches the expansion valve, so the fact that the refrigerant charge can be reduced is often considered a secondary benefit.
The other factor affecting the amount of refrigerant in the distribution piping is the equipment location. Minimizing the distance between the receiver and the evaporators also reduces the refrigerant charge in the liquid piping. For this reason, some users install compressor systems throughout their facilities instead of centralizing them in a compressor room. Distributed systems typically use quieter scroll compressors, along with special noise-reducing enclosures to allow installations in more exposed and occupied areas.
Replacing or retrofitting a direct system to an indirect (or secondary) system is another way to reduce refrigerant charge in distribution piping. This method requires a much more dramatic change to the system, but it is probably the most effective because it can restrict the halocarbon refrigerant to a compact unit composed of a compressor, condenser, and evaporator. The secondary fluid can then be pumped through air-cooling heat exchangers at the load. In this type of system, only a few evaporators are required and the distribution piping is eliminated, so the chance of refrigerant leaks is dramatically reduced.
Opportunities to reduce charge also exist on the high-pressure side of the system between the compressor and the receiver. In comparison to standard air-cooled condensers, systems that use water-cooled condensers operate with a lower charge. If a condensing water source is available, a flat-plate condenser can be mounted near the compressors and used to reject heat from the high-pressure side of the system to the water loop. Typically, bypass lines, variable-speed pumps, and/or flow-restricting valves are used to maintain minimum condensing pressures in water-cooled condensers. Because condenser flooding is no longer required, refrigerant charge can be reduced.
Microchannel condensers also have lower charges than standard air-cooled condensers but may require long runs of liquid piping in installations with indoor compressors. In systems that require flooding, microchannel condensers allow for reduced refrigerant charge because of their smaller internal volume. Alternatively, in low ambient conditions, in conjunction with fan controls, entire banks of some microchannel condensers can be isolated using solenoid valves if the outlet piping is correctly trapped; this approach provides the same benefit as condenser flooding, but requires less refrigerant.
15. REFRIGERANT RETROFITTING
Because of the halocarbon phaseout, many users are retrofitting existing systems to use newer, more acceptable refrigerants (e.g., converting from the HCFC R-22 to the HFC R-407A). Such conversions require planning and preparation.
The most glaring concern is the effect of the new refrigerant on system capacity. Not only should the capacity of the compressor(s) be considered, but also the capacity of every other component in the system (condensers, evaporators, valves, etc.). Equipment and component manufacturers often can provide the needed derating factors to adjust capacities appropriately. Before any work begins, it is a good idea to record how the system is performing: data such as high- and low-side pressures and temperatures suggest how the system should operate after the retrofit. Any available energy data should also be recorded, so the system’s efficiency can be compared to expectations.
Because pressure-temperature relationships will be different for the new refrigerant, the contractor must be prepared to adjust all the pressure controls and/or modify controller set points throughout the system. Thermal expansion valves (TXV) require attention in any retrofit. At the very least, the superheats need to be adjusted; often, the temperature-sensing bulbs and nozzles must be changed out. The designer should consult with the valve manufacturer to decide what action should be taken, and whether the entire TXV should be replaced.
Changing out the system’s lubricating oil is also often required during a retrofit. Mineral oils and alkylbenzene oils are often replaced with POE oils to maintain oil miscibility with the new refrigerant. It is always important to follow a thorough change-out procedure to ensure that all traces of the existing oil are removed from the system. A typical procedure includes (among other tasks) draining the existing oil; changing out liquid driers, suction filters, and oil filters; and recharging the system with the new oil. The draining and recharging steps may need to be repeated more than once to achieve the desired purity for the new oil. Traditionally, 95% or higher purity is required.
Elastomeric gasket and seal materials in the system will also react differently to new refrigerants and oils. Swell characteristics of different elastomers can be referenced from Table 9 in Chapter 29 of the 2017 ASHRAE Handbook—Fundamentals; however, testing is necessary to know exactly how gaskets and seals will react to mixtures of different refrigerants and oils and what factors other than swell may come into play, such as the overall integrity and functionality of the material. For this reason, it is common practice to change out all elastomeric gaskets and seals as part of the retrofit procedure.
After the system is up and running with the new refrigerant and oil, system performance can be evaluated to determine whether it is performing as expected. Refrigerant and oil levels should also be monitored until the correct levels are achieved, and filters should be changed until the system is clean. Finally, it is always crucial to make the appropriate signage and labeling modifications to prevent anyone from topping off the system with the old refrigerant or oil.
It is not uncommon to retrofit existing systems from single-component (azeotropic) refrigerants to blended, zeotropic refrigerants. (See Chapter 2 of the 2017 ASHRAE Handbook—Fundamentals for more information on zeotropic refrigerants.) In these scenarios, the designer and contractor need to be aware of how the refrigerant’s temperature glide behaves throughout the system and know how to properly use the bubble, mean, and dew-point temperatures at the evaporator and condenser to accurately calculate subcooling and superheating. Beyond this, the designer must know what temperatures to use to properly size equipment.
When a zeotropic refrigerant starts to condense in the condenser, it does so at a constant pressure (ignoring pressure drop) at the refrigerant’s dew-point temperature. As the refrigerant continues to condense, its temperature drops until it reaches the bubble-point temperature, at which point it is fully condensed. The liquid can then be subcooled. Conversely, when the refrigerant starts to boil in the evaporator, it starts at the bubble-point temperature and is not fully evaporated until it reaches the dew-point temperature. The gas can then be superheated. So, when calculating subcooling at the condenser exit, the bubble-point temperature must represent the saturation point; when calculating superheating at the evaporator exit, the dew-point temperature must represent the saturation point. Refrigerant manufacturers publish pressure-temperature charts that allow the bubble, mean, and dew-point temperatures to be easily referenced given a specific pressure.
This temperature change behavior during the phase-change process is known as the refrigerant’s temperature glide and is caused by the varying boiling points of the constituent refrigerants within the mixture. Blended refrigerants essentially separate (fractionate) during phase changes, so leaky condensers and evaporators create concern: refrigerant composition changes can occur in the system, leading to unpredictable system operation. For this reason, it is necessary to only charge systems with refrigerant in the liquid state unless the entire cylinder will be immediately used. Furthermore, if a leak occurs and the system is repaired, the refrigerant composition should be checked for significant changes before topping off the system.
When calculating temperature differences to check the rated capacity of existing condensers and evaporators, the mean temperatures should be used along with any derating factors provided by the manufacturer. When checking the capacity of existing compressors, however, using mean temperatures yields a slightly smaller capacity than they actually have because ANSI/AHRI Standard 540-2004 requires, when rating compressor capacities, that dew-point temperatures be used as the reference temperatures at the corresponding evaporating and condensing pressures. The challenge, however, exists in accurately determining the dew-point temperatures. Simply adding half of the glide to the mean temperature may not be accurate: it is difficult to determine what the actual mean temperature really must be for effective evaporator or condenser operation. Because most published capacity data for heat exchangers are based on temperature that is assumed to be constant during phase change, supplemental derating factors must be used.