Some float switches (Figure 7) operate from movement of a magnetic armature located in the field of a permanent magnet. Others use solid-state circuits in which a variable signal is generated by liquid contact with a probe that replaces the float; this method is adapted to remote-controlled applications and is preferred for ultralow-temperature applications.

The float switch can maintain or indicate the level of a liquid, operate an alarm, control pump operation, or perform other functions. A float switch, solenoid liquid valve, and hand expansion valve combination can control refrigerant level on the high- or low-pressure side of the refrigeration system in the same way that high- or low-side float valves are used. The hand expansion valve, located in the refrigerant liquid line immediately downstream of the solenoid valve, is initially adjusted to provide a refrigerant flow rate at maximum load to keep the solenoid liquid valve in the open position 80 to 90% of the time; it need not be adjusted thereafter. From the outlet side of the hand expansion valve, refrigerant passes through a line and enters either the evaporator or the surge drum.

When the float switch is used for low-pressure level control, precautions must be taken to provide a calm liquid level that falls in response to increased evaporator load and rises with decreased evaporator load. The same recommendations for insulation of the body and liquid leg of the low-pressure float valve apply to the float switch when it is used for refrigerant-level control on the low-pressure side of the refrigeration system. To avoid floodback, controls should be wired to prevent the solenoid liquid valve from opening when the solenoid suction valve closes or the compressor stops.

The basic principle is that the electrical capacitance of a vertical conducting rod, centered within a vertical conducting cylinder, varies approximately in proportion to the liquid level in the enclosure. The capability to accomplish this depends on the significant difference between the dielectric constants of the liquid and the vapor above the liquid surface.

Capacitance probes are available in a variety of configurations, using a full range of refrigerants. Active lengths vary from 6 in. to 13 ft; output signals vary from 0 to 5 or 1 to 6 V, 4 to 20 mA, or digital readout. Operating temperatures range from −100 to 150°F. Both internal and external vessel mountings are available.

Figure 10 shows a schematic cross section of a typical thermostatic expansion valve, with the principal components identified. The following pressures and their equivalent forces govern thermostatic expansion valve operation:

At any constant operating condition, these pressures (forces) are balanced and P₁ = P₂ + P₃.

An additional force, which is small and not considered fundamental, arises from the pressure differential across the valve port. To a degree, it can affect thermostatic expansion valve operation. For the configuration shown in Figure 10, the force resulting from port imbalance is the product of pressure drop across the port and the area of the port; it is an opening force in this configuration. In other designs, depending on the direction of flow through the valve, port imbalance may result in a closing force.

Figure 10. Typical Thermostatic Expansion Valve

The principal effect of port imbalance is on the stability of valve control. As with any modulating control, if the ratio of the diaphragm area to the port is kept large, the unbalanced port effect is minor. However, if this ratio is small or if system operating conditions require, a balanced port valve can be used. Figure 11 shows a typical balanced port design.

Figure 12 shows an evaporator operating with R-410A at a saturation temperature of 40°F (118 psig). Liquid refrigerant enters the expansion valve, is reduced in pressure and temperature at the valve port, and enters the evaporator at point A as a mixture of saturated liquid and vapor. As flow continues through the evaporator, more of the refrigerant is evaporated. Assuming there is no pressure drop, the refrigerant temperature remains at 40°F until the liquid is entirely evaporated at point B. From this point, additional heat absorption increases the temperature and superheats the refrigerant gas, while the pressure remains constant at 118 psig, until, at point C (the evaporator’s outlet), the refrigerant gas temperature is 50°F. At this point, the superheat is 10°F (50 – 40°F).

Figure 11. Typical Balanced Port Thermostatic Expansion Valve

An increased heat load on the evaporator increases the temperature of refrigerant gas leaving the evaporator. The bulb of the valve senses this increase, and the thermostatic charge pressure P₁ increases and causes the valve to open wider. The increased flow results in a higher evaporator pressure P₂, and a balanced control point is again established. Conversely, decreased heat load on the evaporator decreases the temperature of refrigerant gas leaving the evaporator and causes the thermostatic expansion valve to start closing.

Figure 12. Thermostatic Expansion Valve Controlling Flow of Liquid R-410A Entering Evaporator, Assuming R-410A Charge in Bulb

The new control point, after an increase in valve opening, is at a slightly higher operating superheat because of the spring rate of the diaphragm and superheat spring. Conversely, decreased load results in an operating superheat slightly lower than the original control point.

These superheat changes in response to load changes are shown by the gradient curve of Figure 13. Superheat at no load, distance 0-A, is called static superheat and ensures sufficient spring force to keep the valve closed during system shutdown. An increase in valve capacity or load is approximately proportional to superheat until the valve is fully open. Opening superheat, represented by the distance A-B, is the superheat increase required to open the valve to match the load; operating superheat is the sum of static and opening superheats.

The factory superheat setting (static superheat setting) of thermostatic expansion valves is made when the valve starts to open. (See ASHRAE Standard 17 for methods of setting the thermostatic expansion valve.) Valve manufacturers establish capacity ratings on the basis of opening superheat, typically from 4 to 8°F, depending on valve design, size, and application. Full-open capacities usually exceed rated capacities by 10 to 40% to allow a reserve, represented by the distance B-C in Figure 13, for manufacturing tolerances and application contingencies.

A valve should not be selected on the basis of its reserve capacity, which is available only at higher operating superheat. The added superheat may have an adverse effect on performance. Because valve gradients used for rating purposes are selected to produce optimum modulation for a given valve design, manufacturers’ recommendations should be followed.

Figure 13. Typical Gradient Curve for Thermostatic Expansion Valves

Thermostatic expansion valve capacities are normally published for various evaporator temperatures and valve pressure drops. (See AHRI Standard 750 and ASHRAE Standard 17 for testing and rating methods.) Nominal capacities apply at 40°F evaporator temperature. Capacities are reduced at lower evaporator temperatures, because of the changed refrigerant pressure/temperature relationship at lower temperatures. For example, if R-410A is used, the change in saturated pressure between 40 and 45°F is 11.8 psi, whereas between −20 and –15°F the change is 4.9 psi. Although the valve responds to pressure changes, published capacities are based on superheat change. Thus, the valve opening and, consequently, valve capacity are less for a given superheat change at lower evaporator temperatures.

Pressure drop across the valve port is always the net pressure drop available at the valve, rather than the difference between compressor discharge and compressor suction pressures. Allowances must be made for the following:

Variations in valve capacity related to changes in system conditions are approximately proportional to the following relationship:

(1)

where

Thermostatic expansion valve capacity is dependent on vapor-free liquid entering the valve. If there is flash gas in the entering liquid, valve capacity is reduced substantially because

Flashing of liquid refrigerant may be caused by pressure drop in the liquid line, filter-drier, vertical lift, or a combination of these. If refrigerant subcooling at the valve inlet is not adequate to prevent flash gas from forming, additional subcooling means must be provided.

There are several principal types of thermostatic charges, each with certain advantages and limitations.

Figure 14. Pressure/Temperature Relationship of R-134a Gas Charge in Thermostatic Element

Gas Charge. Conventional gas charges are limited liquid charges that use the same refrigerant in the thermostatic element as in the refrigeration system. The amount of charge is such that, at a predetermined temperature, all of the liquid has vaporized and any temperature increase above this point results in practically no increase in element pressure. Figure 14 shows the pressure/temperature relationship of an R-134a gas charge in the thermostatic element. Because of the characteristic pressure-limiting feature of its thermostatic element, the gas-charged valve can provide compressor motor overload protection on some systems by limiting the maximum operating suction pressure (MOP). It also helps prevent floodback (return of refrigerant liquid to the compressor through the suction line) on starting. Increasing the superheat setting lowers the MOP; decreasing the superheat setting raises the MOP because the superheat spring and evaporator pressure balance the element pressure through the diaphragm.

Gas-charged valves must be carefully applied to avoid loss of control from the bulb. The small amount of charge in the bulb condenses at the coldest point. If the diaphragm chamber or connecting tube becomes colder than the bulb, the valve will throttle or close at a superheat different than the evaporator-leaving superheat, as detailed in the section on Application.

Liquid Charge. Straight liquid charges use the same refrigerant in the thermostatic element and refrigeration system. The volumes of the bulb, bulb tubing, and diaphragm chamber are proportioned so that the bulb contains some liquid under all temperatures. Therefore, the bulb always controls valve operation, even with a colder diaphragm chamber or bulb tubing.

The straight liquid charge (Figure 15) results in increased operating superheat as evaporator temperature decreases. This usually limits use of the straight liquid charge to moderately high evaporator temperatures. The valve setting required for a reasonable operating superheat at a low evaporator temperature may cause floodback during cooling at normal ambient temperatures.

Liquid Cross Charge. Liquid cross charges use a volatile liquid that is different from the refrigerant in the system and can be mixed with a noncondensable gas in the thermostatic element. Cross charges have flatter pressure/temperature curves than the system refrigerants with which they are used. Consequently, their superheat characteristics differ considerably from those of straight liquid or gas charges.

Cross charges in the commercial temperature range generally have superheat characteristics that are nearly constant or that deviate only moderately through the evaporator temperature range. This charge, also shown in Figure 15, is generally used in the evaporator temperature range of 40 to 0°F or slightly below.

Figure 15. Typical Superheat Characteristics of Common Thermostatic Charges

For evaporator temperatures substantially below 0°F, a more extreme cross charge may be used. At high evaporator temperatures, the valve controls at a high superheat. As the evaporator temperature falls to the normal operating range, the operating superheat also falls to normal. This prevents floodback on starting, reduces load on the compressor motor at start-up, and allows rapid pulldown of suction pressure. To avoid floodback, valves with this type of charge must be set for the optimum operating superheat at the lowest evaporator temperature expected.

Gas Cross Charge. Gas cross charges combine features of the gas charge and liquid cross charge. They use a limited amount of liquid, thereby providing a maximum operating pressure. The liquid used in the charge is different from the refrigerant in the system, and is often mixed with a noncondensable gas, such as dry air or nitrogen. This type of charge provides superheat characteristics similar to those of liquid cross charges (low temperature). Consequently, it provides both the superheat characteristics of a cross charge, and the maximum operating pressure of a gas charge (Figure 15).

Adsorption Charge. Typical adsorption charges depend on the property of an adsorbent, such as silica gel or activated carbon, that is used in the thermostatic element bulb to adsorb and desorb a gas such as carbon dioxide, with accompanying changes in temperature. The amount of adsorption or desorption changes the pressure in the thermostatic element. Because adsorption charges respond primarily to the temperature of the adsorbent material, they are the charges least affected by the ambient temperature surrounding the bulb, bulb tubing, and diaphragm chamber. The comparatively slow thermal response of the adsorbent results in a charge characterized by its stability. Superheat characteristics can be varied by using different charge fluids, adsorbents, and/or charge pressures. The pressure-limiting feature of the gas or gas cross charges is not available with the adsorption charge.

Internal Equalizer. When the refrigerant pressure drop through an evaporator is relatively low (e.g., equivalent to 2°F change in saturation temperature), a thermostatic expansion valve that has an internal equalizer may be used. Internal equalization describes valve outlet pressure transmitted through an internal passage to the underside of the diaphragm (see Figure 10).

Pressure drop in many evaporators is greater than the 2°F equivalent. When a refrigerant distributor is used, pressure drop across the distributor causes pressure at the expansion valve outlet to be considerably higher than that at the evaporator outlet. As a result, an internally equalized valve controls at an abnormally high superheat. Under these conditions, the evaporator does not perform efficiently because it is starved for refrigerant. Furthermore, the distributor pressure drop is not constant, but varies with refrigerant flow rate and therefore cannot be compensated for by adjusting the superheat setting of the valve.

External Equalizer. Because evaporator and/or refrigerant distributor pressure drop causes poor system performance with an internally equalized valve, a valve that has an external equalizer is used. Instead of the internal communicating passage shown in Figure 10, an external connection to the underside of the diaphragm is provided. The external equalizer line is connected either to the suction line, as shown in Figure 16, or into the evaporator at a point downstream from the major pressure drop.

Figure 16. Bulb Location for Thermostatic Expansion Valve

Pilot-operated thermostatic expansion valves are used on large systems in which the required capacity per valve exceeds the range of direct-operated valves. The pilot-operated valve consists of a piston-type pilot-operated regulator, which is used as the main expansion valve, and a low-capacity thermostatic expansion valve, which serves as an external pilot valve. The small pilot thermostatic expansion valve supplies pressure to the piston chamber or, depending on the regulator design, bleeds pressure from the piston chamber in response to a change in the operating superheat. Pilot operation allows the use of a characterized port in the main expansion valve to provide good modulation over a wide load range. Therefore, a very carefully applied pilot-operated valve can perform well on refrigerating systems that have some form of compressor capacity reduction, such as cylinder unloading. Figure 17 shows such a valve applied to a large-capacity direct-expansion chiller.

Figure 17. Pilot-Operated Thermostatic Expansion Valve Controlling Liquid Refrigerant Flow to Direct-Expansion Chiller

The auxiliary pilot controls should be sized to handle only the pilot circuit flow. For example, in Figure 17 a small solenoid valve in the pilot circuit, installed ahead of the thermostatic expansion valve, converts the pilot-operated valve into a stop valve when the solenoid valve is closed.

Equalization Features. When the compressor stops, a thermostatic expansion valve usually moves to the closed position. This movement sustains the difference in refrigerant pressures in the evaporator and the condenser. Low-starting-torque motors require that these pressures be equalized to reduce the torque needed to restart the compressor. One way to provide pressure equalization is to add, parallel to the main valve port, a small fixed auxiliary passageway, such as a slot or drilled hole in the valve seat or valve pin. This opening allows limited fluid flow through the valve, even when the valve is closed, and allows the system pressures to equalize on the off cycle. The size of a fixed auxiliary passageway must be limited so its flow capacity is not greater than the smallest flow that must be controlled in normal system operation. For a drilled hole, the hole’s diameter should be bigger than the maximum allowable particle size circulating in the system, to prevent permanent obstructions. Slots in the seat may be a more robust solution, because any particle obstructing the slot would be flushed when the expansion valve opens fully.

Another, more complex control is available for systems requiring shorter equalizing times than can be achieved with the fixed auxiliary passageway. This control incorporates an auxiliary valve port, which bypasses the primary port and is opened by the element diaphragm as it moves toward and beyond the position at which the primary valve port is closed. Flow capacity of an auxiliary valve port can be considerably larger than that of the fixed auxiliary passageway, so pressures can equalize more rapidly.

Flooded System. Thermostatic expansion valves are seldom applied to flooded evaporators because superheat is necessary for proper valve control; only a few degrees of suction vapor superheat in a flooded evaporator incurs a substantial loss in system capacity. If the bulb is installed downstream from a liquid-to-suction heat exchanger, a thermostatic expansion valve can be made to operate at this point on a higher superheat. Valve control is likely to be poor because of the variable rate of heat exchange as flow rates change (see the section on Application).

Some expansion valves have modified thermostatic elements in which electric heat is supplied to the bulb. The bulb is inserted in direct contact with refrigerant liquid in a low-side accumulator. The contact of cold refrigerant liquid with the bulb overrides the artificial heat source and throttles the expansion valve. As liquid falls away from the bulb, the valve feed increases again. Although similar in construction to a thermostatic expansion valve, it is essentially a modulating liquid-level control valve.

Desuperheating Valves. Thermostatic expansion valves with special thermostatic charges are used to reduce gas temperatures (superheat) on various air-conditioning and refrigeration systems. Suction gas in a single-stage system can be desuperheated by injecting liquid directly into the suction line. This cooling may be required with or without discharge gas bypass used for compressor capacity control. The line upstream of the valve bulb must be long enough so the injected liquid refrigerant can mix adequately with the gas being desuperheated. On compound compression systems, an expansion valve may be used to inject liquid directly into the interstage line upstream of the expansion valve bulb location to provide intercooling.

Hunting is alternate overfeeding and starving of the refrigerant feed to the evaporator. It produces sustained cyclic changes in the pressure and temperature of refrigerant gas leaving the evaporator. Extreme hunting reduces refrigeration system capacity because mean evaporator pressure and temperature are lowered and compressor capacity is reduced. If overfeeding of the expansion valve causes intermittent flooding of liquid into the suction line, the compressor may be damaged.

Although hunting is commonly attributed to the thermostatic expansion valve, it is seldom solely responsible. One reason for hunting is that all evaporators have a time lag. When the bulb signals for a change in refrigerant flow, the refrigerant must traverse the entire evaporator before a new signal reaches the bulb. This lag or time lapse may cause continuous overshooting of the valve both opening and closing. In addition, the thermostatic element, because of its mass, has a time lag that may be in phase with the evaporator lag and amplify the original overshooting.

It is possible to alter the thermostatic element’s response rate by either using thermal ballast or changing the mass or heat capacity of the bulb, thereby damping or even eliminating hunting. A change in valve gradient may produce a similar result.

Slug flow or percolation in the evaporator can also cause hunting. Under these conditions, liquid refrigerant moves in waves (slugs) that fill a portion of the evaporator tube and erupt into the suction line. These unevaporated slugs chill the bulb and temporarily reduce the feed of the valve, resulting in intermittent starving of the evaporator.

On multiple-circuit evaporators, a lightly loaded or overfed circuit also floods into the suction line, chills the bulb, and throttles the valve. Again, the effect is intermittent; when the valve feed is reduced, flooding ceases and the valve reopens.

Hunting can be minimized or avoided in the following ways:

Bulb Location. Most installation requirements are met by strapping the bulb to the suction line to obtain good thermal contact between them. Normally, the bulb is attached to a horizontal line upstream of the external equalizer connection (if used) at a 3 or 9 o’clock position as close to the evaporator as possible. The bulb is not normally placed near or after suction-line traps, but some designers test and prove locations that differ from these recommendations. A good moisture-resistant insulation over the bulb and suction line diminishes the adverse effect of varying ambient temperatures at the bulb location.

Occasionally, the bulb of the thermostatic expansion valve is installed downstream from a liquid-suction heat exchanger to compensate for a capacity shortage caused by an undersized evaporator. Although this procedure seems to be a simple method of maximizing evaporator capacity, installing the bulb downstream of the heat exchanger is undesirable from a control standpoint. As the valve modulates, the liquid flow rate through the heat exchanger changes, causing the rate of heat transfer to the suction vapor to change. An exaggerated valve response follows, resulting in hunting. There may be a bulb location downstream from the heat exchanger that reduces hunting considerably. However, the danger of floodback to the compressor normally overshadows the need to attempt this method.

Certain installations require increased bulb sensitivity as a protection against floodback. The bulb, if located properly in a well in the suction line, has a rapid response because of its direct contact with the refrigerant stream. Bulb sensitivity can be increased by using a bulb smaller than is normally supplied. However, using a smaller bulb is limited to gas-charged valves. Good piping practice also affects expansion valve performance.

Figure 18 shows the proper piping arrangement when the suction line runs above the evaporator. A lubricant trap that is as short as possible is located downstream from the bulb. The vertical riser(s) must be sized to produce a refrigerant velocity that ensures continuous return of lubricant to the compressor. The terminal end of the riser(s) enters the horizontal run at the top of the suction line; this avoids interference from overfeeding any other expansion valve or any drainback during the off cycle.

If circulated with lubricant-miscible refrigerant, a heavy concentration of lubricant elevates the refrigerant’s boiling temperature. The response of the thermostatic charge of the expansion valve is related to the saturation pressure and temperature of pure refrigerant. In an operating system, the false pressure/temperature signals of lubricant-rich refrigerants cause floodback or operating superheats considerably lower than indicated, and quite often cause erratic valve operation. To keep lubricant concentration at an acceptable level, either the lubricant pumping rate of the compressor must be reduced or an effective lubricant separator must be used.

Figure 18. Bulb Location When Suction Main Is Above Evaporator

The external equalizer line is ordinarily connected at the evaporator outlet, as shown in Figure 18. It may also be connected at the evaporator inlet or at any other point in the evaporator downstream of the major pressure drop. On evaporators with long refrigerant circuits that have inherent lag, hunting may be minimized by changing the connection point of the external equalizer line.

In application, the various parts of the valve’s thermostatic element are simultaneously exposed to different thermal influences from the surrounding ambient air and the refrigerant system. In some situations, cold refrigerant exiting the valve dominates and cools the thermostatic element to below the bulb temperature. When this occurs with a gas-charged or gas cross-charged valve, the charge condenses at the coldest point in the element and control of refrigerant feed moves from the bulb to the thermostatic element (diaphragm chamber). Pressure applied to the top of the diaphragm diminishes to saturation pressure at the cold point. Extreme starving of the evaporator, progressing to complete cessation of refrigerant flow, is characteristic. For this reason, gas-charged or gas cross-charged valves should be applied only to multicircuited evaporators that use refrigerant distributors. The distributor typically provides sufficient pressure drop to maintain a saturation temperature at the valve outlet well above the temperature at the bulb location.

Internally equalized gas-charged or gas cross-charged valves should only be considered in very carefully selected applications where the risk of loss of control can be minimized. Some gas cross-charge formulations may be slightly less susceptible to the described loss of control than are straight gas charges, but they are far from immune. Gas-charged and gas cross-charged valves with specially constructed thermostatic power elements that positively isolate the charge fluids in the temperature-sensing element (bulb) have been applied in situations where there was high risk of control loss and the pressure-limiting feature of a gas-charged valve was required.

Gas-charged bulbless valves, frequently called block valves (Figure 19), are practically immune to loss of control because the thermostatic element (diaphragm chamber) is located at the evaporator outlet. The valve is constructed so that the temperature-sensing function of the remote bulb is integrated into the thermostatic element by purposely confining all of the charge fluid to this chamber.

Figure 19. Typical Block Valve

Liquid, liquid cross-charged, and adsorption-charged valves are not susceptible to the same type of loss of control that gas-charged or gas cross-charged valves are. However, exposure to extreme ambient temperature environments causes shifting of operating superheats. The degree of superheat shift depends on the severity of the thermal exposure. High ambient temperatures surrounding thermally sensitive parts of the valve typically lower operating superheats, and vice versa. Gas-charged and gas cross-charged valves, including bulbless or block valves, respond to high ambient exposure similarly but starve the evaporator when exposed to ambient temperatures below evaporator outlet refrigerant temperatures.

Direct-operated evaporator-pressure-regulating valves are relatively simple, as shown in Figure 25. The inlet pressure acts on the bottom of the seat disk and opposes the spring. The outlet pressure acts on the bottom of the bellows and the top of the seat disc. Because the effective areas of the bellows and the port are equal, these two forces cancel each other, and the valve responds to inlet pressure only. When inlet pressure rises above the equivalent pressure exerted by the spring force, the valve begins to open. When inlet pressure falls, the spring moves the valve in the closing direction. In operation, the valve assumes an intermediate throttling position that balances the refrigerant flow rate with evaporator load.

Because both spring and bellows must be compressed through the entire opening valve stroke, a significant change in inlet pressure is required to open the valve to its rated capacity. Inlet pressure changes of 5 to 10 psi or more, depending on design, are typically required to move direct-operated evaporator pressure regulators from their closed position to their rated flow capacity. Therefore, these valves have relatively high gradients and the system may experience significant changes in regulated evaporator pressure when large load changes occur.

Figure 25. Direct-Operated Evaporator Pressure Regulator

Pilot-operated evaporator-pressure-regulating valves are either self powered or driven by high pressure. The self-powered regulator (Figure 26) starts to open when the inlet pressure approaches the equivalent pressure setting of the diaphragm spring. The diaphragm lifts to allow inlet pressure to flow through the pilot port, which increases the pressure above the piston. This increase moves the piston down, causing the main valve to open. Flow through the opening valve relieves evaporator pressure into the suction line. As evaporator pressure diminishes, the diaphragm throttles flow through the pilot port, a bleed hole in the piston relieves pressure above the piston to the low-pressure outlet side of the main valve, and the main spring moves the valve in the closing direction. Balanced flow rates through the pilot port and piston bleed hole establish a stable piston pressure that balances against the main spring. The main valve assumes an intermediate throttling position that allows the refrigerant flow rate required to satisfy the evaporator load. Pilot-operated regulators have relatively low gradients and are capable of precise pressure regulation in evaporators that experience large load changes. Typically, pressure loss of up to 2 psi may be needed to move the valve to its fully open position.

Figure 26. Pilot-Operated Evaporator Pressure Regulator (Self-Powered)

Suction stop service can be provided with this style regulator by adding a pilot solenoid valve in the equalizer flow passage to prevent inlet pressure from reaching the underside of the pressure pilot diaphragm regardless of inlet pressure. Suction stop service is often required to facilitate and control evaporator defrost.

High-pressure-driven pilot-operated regulating valves are of a normally open design and require high-pressure liquid or gas to provide a closing force. One advantage of this design over self-powered regulators is that it does not require any suction-pressure drop across the valve or large inlet pressure change to operate. When valve inlet pressure increases above set point (Figure 27), the diaphragm moves up against the spring, allowing the pilot valve pin spring to move the pilot valve pin, pin carrier, and push rods (not shown) up toward closing the pilot valve port. Gas or liquid from the high-pressure side of the system is throttled by the pilot valve, and pressure in the top of the piston chamber bleeds to the valve’s downstream side through a bleed orifice. As pressure on top of the piston diminishes, the main body spring moves the valve piston in the opening direction.

As inlet pressure diminishes, increased flow of high-pressure liquid or gas through the pilot valve drives the piston down toward a closed position.

Figure 27. Pilot-Operated Evaporator Pressure Regulator (High-Pressure-Driven)

A solenoid valve may be used to the drive the piston to the closed position for suction stop service, either by closing the bleed orifice or by supplying high pressure directly to the top of the piston chamber. Note that, in the latter arrangement, a continuous but very small flow of liquid or gas from the system high side is discharged into the suction line downstream of the regulator while the valve is closed. In some applications, this bleed may enhance compressor cooling and lubricant return.

Selection of evaporator-pressure-regulating valves is based on the flow capacity required to satisfy the load imposed on the evaporator being regulated and the pressure drop available across the regulator. For example, if an evaporator is to be regulated to a pressure of 30 psig and the regulator discharges into a suction line that normally operates or is maintained at 20 psig, the regulator should be selected to satisfy the evaporator load at a 10 psi pressure loss across the valve. To select direct-operated regulators, consider the high gradient of this design; ensure that the variation of inlet pressure that occurs with load changes is acceptable for the application. For example, a direct-operated regulator set at high-load operating conditions to protect against chiller freeze-up may allow evaporator pressure to drop into the freeze-up danger zone at low loads because of the large reduction in inlet pressure needed to throttle the valve to near-closed stroke. Externally powered regulators should be selected to satisfy the flow requirements imposed by the evaporator load at pressure drops compatible with the application.

Grossly oversized regulating valves are very susceptible to unstable operation, which may in turn upset the stability of other controls in the system, significantly degrade system performance, and risk damage to other system components.

Evaporator pressure regulators are used on air-cooling evaporators to control frosting or prevent excessive dehumidification. They are also used on water and brine chillers to protect against freezing under low-load conditions.

When multiple evaporators are connected to a common suction line, as shown in Figure 28, evaporator pressure regulators may be installed to control evaporator pressure in each individual unit or in a group of units operating at the same pressure. The regulators maintain the desired saturation temperature in evaporators serving the high- and medium-temperature loads; those for low-temperature loads may be directly connected to the suction main. In these systems, the compressor(s) are loaded, unloaded, and cycled to maintain suction main pressure as the combined evaporator loads vary.

The pilot-operated self-powered evaporator pressure regulator, with internal pilot passage, receives its source of pressure to both power the valve and sense the controlled pressure at the regulator inlet connection. A regulator with an external pilot connection allows a choice of remote pressure source for both controlled pressure sensing and driving the valve. The external pilot connection can also facilitate use of remote pressure and solenoid pilot valves. Figure 28 shows a pilot solenoid valve installed in the external pilot line. This arrangement allows the regulator to function as a suction stop valve as well as an evaporator pressure regulator. The suction stop feature is particularly useful on a flooded evaporator to prevent the refrigerant from leaving the evaporator when the load is satisfied and the evaporator is deactivated. The stop feature is also effective during evaporator defrost cycles, especially with gas defrost systems. When regulator inlet pressure is unstable to the point of upsetting regulator performance, the external pilot connection may be used to facilitate use of volume chambers and other non-flow-restricting damping means to smooth the pilot pressure source before it enters the regulator pilot connection.

Figure 28. Evaporator Pressure Regulators in Multiple System

A remote pressure pilot installed in the external pilot line can be located to facilitate manual adjustment of the pressure setting when the main regulator must be installed in an inaccessible location.

Multiple pilots, including temperature-actuated pilots, pressure pilots, and solenoid pilot stop valves, may be connected in various parallel-series arrangements to the external pilot connection, thus allowing the main valve to function in different modes and pressure settings, depending on which pilot is selected to control. The controlling pilot is then selected by activating the appropriate solenoid stop valve. Pressure pilots may also be adapted to accept connection to pneumatic control systems, allowing automatic resetting of the pressure pilot as part of a much more comprehensive control strategy.

Although evaporator pressure regulation is the most common use of upstream pressure regulators, they are also used in a variety of other refrigeration system applications. Upstream pressure regulators may be adapted for internal pressure relief, air-cooled condenser pressure regulation during low ambient operation, and liquid receiver pressure regulation.

Figure 29 shows a cross section of a typical constant-pressure expansion valve. The valve has both an adjustable opening spring, which exerts its force on top of the diaphragm in an opening direction, and a spring beneath the diaphragm, which exerts its force in a closing direction. Evaporator pressure is admitted beneath the diaphragm, through either an internal or external equalizer passage, and the combined forces of evaporator pressure and closing spring counterbalance the adjustable opening spring force.

During normal system operation, a small increase in evaporator pressure pushes the diaphragm up against the adjustable opening spring, allowing the closing spring to move the pin in a closing direction. This restricts refrigerant flow and limits evaporator pressure. When evaporator pressure drops below the valve setting (a decrease in load), the opening spring moves the valve pin in an opening direction. As a result, refrigerant flow increases and raises the evaporator pressure, bringing the three primary forces in the valve back into balance.

Figure 29. Constant-Pressure Expansion Valve

Because constant-pressure expansion valves respond to evaporator load changes inversely, their primary application is in systems that have nearly constant evaporator loading.

The constant-pressure expansion valve should be selected to provide the required liquid refrigerant flow capacity at the expected pressure drop across the valve, and should have an adjustable pressure range that includes the required design evaporator (valve outlet) pressure. The system designer should decide whether off-cycle pressure equalization is required.

Constant-pressure expansion valves overfeed the evaporator as load diminishes, and underfeed as load increases. Their primary function is to balance liquid flow rate with compressor capacity at constant evaporator pressure/temperature as loading varies, protecting against product freezing at low loads and compressor motor overload when evaporator loading increases. Because the valve responds inversely to evaporator load variations, other means to protect the compressor against liquid floodback at low loads and overheating at high loads (e.g., suction-line accumulators, enhanced compressor cooling or liquid injection devices) are needed in systems that experience significant load variation.

Constant-pressure expansion valves are best suited to simple single-compressor/single-evaporator systems when constant evaporator temperature is important and significant load variation does not occur. They are commonly used in drink dispensers, food dispensers, drinking fountains, ice cream freezers, and self-contained room air conditioners. They are typically direct-operated devices; however, they may be pilot-operated for applications requiring very large capacity. They are also used to regulate hot gas in discharge bypass capacity reduction arrangements, as described in the section on Discharge Bypass Valves.

Constant-pressure expansion valves close in response to the abrupt increase in evaporator pressure when the compressor cycles off, preventing flow during the off cycle. A small, fixed auxiliary passageway, described in the section on Thermostatic Expansion Valves, can also be built into constant-pressure expansion valves to provide off-cycle pressure equalization for use with low-starting-torque compressor motors.

Direct-acting suction-pressure regulators respond to their own outlet or downstream pressure. They are relatively simple devices, as shown in Figure 30. The valve outlet pressure acts on the bottom of the disk and exerts a closing force, which is opposed by the adjustable spring force. The inlet pressure acts on the underside of the bellows and the top of the seat disk. Because the effective areas of the bellows and port are equal, these two forces cancel each other and the valve responds to outlet pressure only. When outlet pressure falls below the equivalent force exerted by the spring, the seat disk moves in an opening direction to maintain outlet pressure. If outlet pressure rises, the seat disk moves in a closing direction and throttles the refrigerant flow to limit the downstream pressure. The proper pressure setting for a specific system is one that is low enough to protect the compressor from overload without unnecessarily compromising system capacity. Because both spring and bellows must be compressed through the entire closing valve stroke, a significant change in outlet pressure is required to close the valve to its minimum capacity. Outlet pressure changes of 5 to 10 psi or more, depending on design, are typically required to move direct-operated downstream pressure regulators from open position to near closed position. Therefore, these valves have relatively high gradients, and regulated suction pressure may change significantly when large changes in load occur.

Figure 30. Direct-Acting Suction-Pressure Regulator

Pilot-operated suction-pressure regulators are available for larger systems and applications requiring more precise pressure regulation over wide load and inlet pressure variations. Their design is significantly more complex, because of pilot operation. The method of operation is similar to that described for pilot-operated evaporator pressure regulators. However, in downstream pressure regulators, the pilot is reverse acting and functions similarly to the constant-pressure expansion valve. Suction stop service can also be provided with this type of regulator.

The suction-pressure regulator should be selected to provide the required flow capacity at a low pressure loss to minimize system capacity penalty. However, take care to avoid oversizing, which can lead to unstable regulator operation. The significant change in outlet pressure required to stroke direct-operated regulators should also be considered during selection.

Figure 31. Condenser Pressure Regulation (Two-Valve Arrangement)

Suction-pressure-regulating valves are primarily used to prevent compressor motor overload caused by excessive suction pressure related to warm evaporator start-up, defrost termination, or intermittent high-load operation. These regulating valves are typically designed to operate at normal refrigeration system low-side pressures and temperatures. However, similar-type downstream pressure regulators may be modified to include suitable seat materials and high-gradient springs for application in system high-side pressure and temperature conditions. For example, they may be used in a variety of schemes to maintain necessary operating pressures in air-cooled condensers during cold-weather operation. Additionally, modified regulators are used to bypass compressor discharge gas in refrigeration system capacity reduction schemes, as mentioned in the application sections in the Constant-Pressure Expansion Valves and Discharge Bypass Valves sections.

The first valve in the two-valve arrangement shown in Figure 31 is typically an upstream pressure regulator that is constructed and operates similarly to the evaporator-pressure-regulating valves shown in Figures 25 to 27. Pilot-operated regulators are typically used to meet the flow capacity requirements of large systems. They may have special features that make them suitable for high-pressure and high-temperature operating conditions. This control may be installed at either the condenser inlet or outlet; the outlet is usually preferred because a smaller valve can satisfy the system’s flow capacity requirements. It throttles when the condenser outlet or compressor discharge pressure falls as a result of cold-weather operating conditions.

The second valve in the two-valve arrangement is installed in a condenser bypass line. It may be a downstream pressure regulator similar to the suction-pressure regulator in Figure 30, or a differential pressure regulator as in Figure 31. The differential regulator is often preferred for simplicity; however, the minimum opening differential pressure must be greater than the pressure drop across the condenser at full-load summer operating conditions. As the first valve throttles in response to falling compressor discharge or condenser outlet pressure, the second valve opens and allows hot gas to bypass the condenser to mix with and warm the cold liquid entering the receiver, thereby maintaining adequate high-side saturation pressure.

Special-purpose three-way pressure-regulating valves similar to that shown in Figure 32 are also used for condenser pressure regulation. The valve in Figure 32 is a direct-acting pressure regulator with a second inlet that performs the hot-gas bypass function, eliminating the need for the second valve in the two-valve system described previously. The three-way valve simultaneously throttles liquid flow from the condenser and bypasses hot compressor discharge gas to the valve outlet, where it mixes with and warms liquid entering the receiver, thereby maintaining adequate high-side saturated liquid pressure.

In the three-way valve, the lower side of a flexible metal diaphragm is exposed to system high-side pressure, while the upper side is exposed to a noncondensable gas charge (usually dry nitrogen or air). A pushrod links the diaphragm to the valve poppet, which seats on either the upper or lower port and throttles either discharge gas or liquid from the condenser, respectively. Valves that respond to condenser pressure are frequently used; however, valve designs that respond to receiver pressure are available.

Figure 32. Three-Way Condenser-Pressure-Regulating Valve

During system start-up in extremely cold weather, the poppet may be tight against the lower seat, stopping all liquid flow from the condenser and bypassing discharge gas into the receiver until adequate system high-side pressure is generated. During stable operation in cold weather, the poppet modulates at an intermediate position, with liquid flow from the condensing coil mixing with compressor discharge gas in the valve outlet and flowing to the receiver. With higher condensing pressure during warm weather, the poppet seats tightly against the upper port, allowing free flow of liquid from condenser to receiver and preventing compressor discharge gas from bypassing the condenser.

The three-way condenser-pressure-regulating valve set point is usually not field adjustable. The pressure setting is established by the pressure of the gas charge placed in the dome above the diaphragm during manufacture. Some designs allow field selection between two factory-predetermined set points.

Systems using pressure regulators for air-cooled condenser pressure control during cold-weather operation require careful design. Condenser pressure is maintained by partially filling the condenser with liquid refrigerant, reducing the effective condensing surface available. The condenser is flooded with liquid refrigerant to the point of balancing condenser capacity at low ambient with condenser loading. The system must have adequate refrigerant charge and receiver capacity to maintain a liquid seal at the expansion valve inlet while allowing sufficient liquid for head pressure control to accumulate in the condenser. If the system cycles off or otherwise becomes idle during cold weather, the receiver must be kept warm during the off time so that adequate liquid pressure is available at start-up. When receivers are exposed to low temperatures, it may be necessary to provide receiver heaters and insulation to ensure start-up capability. A check valve at the receiver inlet may be advisable.

Because these systems necessarily contain abnormally large refrigerant charges, careful consideration must be given to controlling refrigerant migration during system idle times under adverse ambient temperatures.

A typical mechanical discharge bypass valve has the same basic configuration as the constant-pressure expansion valve in Figure 29. Construction materials for the discharge bypass valves are suitable for application at high pressure and temperature.

The equivalent pressure from the adjustable spring is balanced across the diaphragm against system evaporator pressure. When evaporator pressure falls below the valve setting, the spring strokes the valve member in the opening direction.

Discharge bypass valves are rated on the basis of allowable evaporator temperature change from closed position to rated opening. This is 6°F for most air-conditioning applications, although capacity multipliers for other changes are available to make the appropriate valve selection. Because several basic system factors are involved in appropriate selection, it is important to know the type of refrigerant, minimum allowable evaporating temperature at the reduced-load condition, compressor capacity at minimum allowable evaporating temperature, minimum evaporator load at which the system is to be operated, and condensing temperature when minimum load exists.

Refrigeration systems experience load variations to some degree throughout the year and may use discharge bypass valves to balance compressor capacity with system load. Depending on the system size and type of compressor used, this valve type may be used in place of compressor cylinder unloaders or to handle the unloading requirements below the last step of cylinder unloading. Depending on the system components and configuration, hot gas may be introduced (1) directly into the suction line or (2) into the evaporator inlet through a specially designed refrigerant distributor, where it mixes with cold refrigerant from the system thermostatic expansion valve to control suction gas temperature before the gas reaches the compressor. If the hot gas is introduced directly into the suction line, be careful to prevent the compressor from overheating (see Chapter 1 for recommendations).

The location of the discharge bypass valve and necessary accessories depend on the type of system (see Chapter 1). Valve manufacturers’ literature gives proper locations and piping recommendations for their products.

In variable-displacement compressors (widely used in automotive air-conditioning systems), a discharge bypass valve (also called a compressor-control valve) is located between the discharge and crankcase of the compressor. This bypass valve regulates pressure in the crankcase and maintains balance between the suction, discharge, and crankcase pressures, thereby establishing the angular position of the wobble or swash plate (for piston compressors), which results in regulating the compressor’s displacement. The compressor-control valve can be pneumatically driven (bellows) or electronically driven (direct solenoid or solenoid acting on bellows).

A high-side float valve controls the mass flow of refrigerant liquid entering the evaporator so it equals the rate at which the refrigerant gas is pumped from the evaporator by the compressor. Figure 33 shows a cross section of a typical valve. The refrigerant liquid flows from the condenser into the high-side float valve body, where it raises the float and moves the valve pin in an opening direction, allowing the liquid to pass through the valve port, expand, and flow into the evaporator or low-pressure receiver. Most of the system refrigerant charge is contained in the evaporator or low-pressure receiver at all times.

For acceptable performance, the high-side float valve is selected for the refrigerant and a rated capacity neither too large nor too small. The orifice is sized for the maximum required capacity with the minimum pressure drop across the valve. The valve operated by the float may be a pin-and-port construction (Figure 33), a butterfly valve, a balanced double-ported valve, or a sliding gate or spool valve. The internal bypass vent tube allows installation of the high-side float valve near the evaporator and above the condenser without danger of the float valve becoming gas bound. Some large-capacity valves use a high-side float valve for pilot operation of a diaphragm or piston spring-loaded expansion valve. This arrangement can improve modulation over a wide range of load and pressure-drop conditions.

Figure 33. High-Side Float Valve

A refrigeration system in which a high-side float valve is typically used may be a simple single evaporator/compressor/condenser system or have a low-pressure liquid receiver with multiple evaporators and compressors. The high-pressure receiver or liquid sump at the condenser outlet can be quite small. A full-sized high-pressure receiver may be required for pumping out flooded evaporator(s) and/or low-pressure receivers. The amount of refrigerant charge is critical with a high-side float valve in simple single evaporator/compressor/condenser systems. An excessive charge causes floodback, whereas insufficient charge reduces system capacity.

The low-side float valve performs the same function as the high-side float valve, but it is connected to the low-pressure side of the system. When the evaporator or low-pressure receiver liquid level drops, the float opens the valve. Liquid refrigerant then flows from the liquid line through the valve port and directly into the evaporator or surge drum. In another design, the refrigerant flows through the valve port, passes through a remote feed line, and enters the evaporator through a separate connection. (A typical direct-feed valve construction is shown in Figure 34.) The low-side float system is a flooded system.

Low-side float valves are selected in the same way as the high-side float valves discussed previously.

Figure 34. Low-Side Float Valve

In these systems, the refrigerant charge is not critical. The low-side float valve can be used with multiple evaporators such that some evaporators may be controlled by other low-side float valves and some by thermostatic expansion valves.

The float valve is mounted either directly in the evaporator or surge drum or in an external chamber connected to the evaporator or surge chamber by equalizing lines (i.e., a gas line at the top and a liquid line at the bottom). In the externally mounted type, the float valve is separated from the float chamber by a gland that maintains a calm level of liquid in the float chamber for steady actuation of the valve.

In evaporators with high boiling rates or restricted liquid and gas passages, the boiling action of the liquid raises the refrigerant level during operation. When the compressor stops or the solenoid suction valve closes, boiling of the liquid refrigerant ceases, and the refrigerant level in the evaporator drops. Under these conditions, the high-pressure liquid line supplying the low-side float valve should be shut off by a solenoid valve to prevent overfilling the evaporator. Otherwise, excess refrigerant will enter the evaporator on the off cycle, which can cause floodback when the compressor starts or the suction solenoid valve opens.

When a low-side float valve is used, ensure that the float is in a calm liquid level that falls properly in response to increased evaporator load and rises with decreased evaporator load. In low-temperature systems particularly, it is important that the equalizer lines between the evaporator and either the float chamber or surge drum be generously sized to eliminate any reverse response of the refrigerant liquid level near the float. Where the low-side float valve is located in a nonrefrigerated room, the equalizing liquid and gas lines and float chamber must be well insulated to provide a calm liquid level for the float.

Several valve designs can be reliably operated with a relatively low-powered solenoid, which helps to minimize energy consumption. The major force that must be overcome by the solenoid operator to open a normally closed valve or close a normally open valve is related to the port area multiplied by the pressure differential across the valve when it is closed. The maximum pressure differential that a specific solenoid valve will reliably open against is called the maximum operating pressure differential (MOPD). To provide acceptable MOPD for refrigeration applications with reasonably powered solenoids, it is necessary to design valves with small ports. Small-ported valves can satisfy low-flow-capacity requirements and pilot duty applications. Thus, direct-acting solenoid valves are limited to low-capacity or low-MOPD applications. Large-capacity and high-MOPD applications use pilot-operated valves similar to those in Figures 36 and 37.

Figure 36. Normally Closed Pilot-Operated Solenoid Valve with Direct-Lift Feature

Solenoid valves are divided into two basic categories related to flow capacity and MOPD:

Figure 37. Normally Closed Pilot-Operated Solenoid Valve with Hammer-Blow and Mechanically Linked Piston-Pin Plunger

These basic categories are each divided into several subcategories:

Figure 35 shows a direct-acting normally closed two-way valve with a long-stroke solenoid that uses a “lost-motion hammer-blow” mechanism to hit the valve pin and overcome the pressure differential force holding the valve closed. This design, though providing some additional opening capability, is less reliable with respect to MOPD repeatability and has a shorter cyclic life because of the high-momentum impacts between plunger, pin, and top plug. It also relies entirely on gravity to move the valve in the closing direction when the solenoid is deenergized. This reliance limits installation to an upright position with the magnetic assembly on top.

Figure 38. Four-Way Refrigerant-Reversing Valve Used in Heat Pumps (Shown in Heating Mode)

Figure 39. Four-Way Refrigerant-Reversing Valve (Shown in Cooling Mode)

A common design of direct-acting solenoid is the direct-lift type such as the solenoid operator portion of the pilot-operated valve in Figure 36. The short stroke minimizes mechanical damage related to impact of the plunger with the top plug and minimizes inrush current in alternating current systems. The closing spring allows installation in any position. Some solenoids of this short-stroke design can survive extended periods of time in failure mode (i.e., when the coil is energized but the valve fails to open because of excessive pressure, low voltage, or other reasons) without significant damage to the coil.

Figure 36 also shows a pilot-operated valve that uses a semi-flexible diaphragm to operate the main valve. When the solenoid is energized, the pilot port (shown centered over the main port) opens and allows pressure above the diaphragm to diminish as it discharges to low pressure at the valve outlet. The higher valve inlet pressure on the underside of the diaphragm surrounding the main port causes the diaphragm to move up, opening the main port to full flow. The ratio of diaphragm area to port area as well as the relative flow rates of bleed hole D and the pilot port are carefully balanced in the design. This ensures that adequate opening force develops to meet the MOPD requirements of the valve specification. When the solenoid is deenergized, the plunger moves down, closing the pilot port, and inlet pressure flowing through bleed hole D allows pressure above the diaphragm to equalize with valve inlet pressure. The entire top of the diaphragm is exposed to inlet pressure; on the bottom side, only the annular area surrounding the port is exposed to inlet pressure; the center portion is exposed to outlet pressure. The net effect of the pressure and area differences is a force that pushes the diaphragm down to close the main port. The spring in the top of the plunger causes the plunger to follow the falling pilot port, keeping it sealed and providing additional downward push on the diaphragm to help close the main port. In the valve of Figure 36, the pilot operator is centrally located directly over the main port. In this configuration, the diaphragm stroke is limited. When larger diaphragm strokes are required for greater flow capacity, the pilot operator and pilot port are relocated to a point beyond the perimeter of the diaphragm over the main valve outlet connection. Separate flow passages in the valve body are provided to conduct pilot port flow from the top of the diaphragm to the valve outlet. This arrangement is usually called an “offset pilot” and retains the benefits of a short-stroke solenoid operator while allowing a diaphragm stroke commensurate with full flow through the main port.

The pilot-operated valve shown in Figure 37 operates according to the same principles as the diaphragm valve, but uses a piston instead of a diaphragm. The carefully controlled annular clearance between piston and inside bore of the valve body is often used to perform the function of bleed hole D (shown in Figure 36), eliminating the need for a separate bleed hole in the piston. The valve in Figure 37 uses a long-stroke hammer-blow pilot operator, which allows long main valve strokes to accommodate large flow capacity requirements without offsetting the pilot. The centered pilot allows mechanically linking the main piston to the plunger assembly to help hold the valve wide open at near-zero pressure drops in low-temperature suction-line applications.

Figure 38 shows a four-way valve, piloted by a four-way direct acting valve shown in the deenergized (heating) position. The reversing valve can be installed to allow for either heating or cooling when energized, depending on the heat pump designer’s preference; this setup cannot be changed after the system is in operation. While the heat pump compressor operates, the high-pressure refrigerant vapor is pumped into reversing valve tube D. A small amount of the flow is carried by capillary tube K to the direct-acting pilot valve. This pilot valve sends this small flow into chamber E of the reversing valve through capillary tube L. The pilot valve also connects chamber F to reversing valve tube S attached to the compressor suction port by way of capillary tubes M and N. The flows applied to chambers E and F build a pressure difference that moves and holds the reversing valve main slide C in its required position. While in position, slide C allows flow from tube D to go along flow path H and out of the reversing valve through tube B. This tube is connected to the indoor heat exchanger. Refrigerant vapor returning from the outdoor heat exchanger enters tube A of the reversing valve and is directed by slide C along flow path J and out the reversing valve tube S. Tube S is connected to the compressor suction port.

Figure 39 shows the reversing valve in its energized (cooling) position. To reach and maintain this position, the capillary tube flows are now from tube D through capillary tubes K and M to chamber F. There is also a small flow from chamber E going to the compressor suction through capillary tubes L and N. As the reversing valve shifts, refrigerant flow from the compressor follows internal path P going out through tube A and flow to suction follows path Q and out tube S. The outdoor heat exchanger is now the condenser, and the indoor heat exchanger is the evaporator. The system has reversed to cooling mode.

Solenoid valves are generally vulnerable to particles in the refrigerant stream and should be protected by a strainer or filter-drier.

Attitude-sensitive valves must be carefully oriented and properly supported to ensure reliable operation. Take care to avoid overheating sensitive valve parts when installation involves soldering, brazing, or welding.

Electrical service provided to solenoid operators deserves careful attention. Most solenoid valve performance failures are related to improper or inadequate provision of electric power to the solenoid. Undervoltage when attempting to open seriously compromises MOPD and causes a failure to open. Continued application of power to an alternating current (AC) solenoid coil installed on a valve that is unable to open overheats the coil and may lead to premature coil failure, even at undervoltage. Although direct current (DC) solenoids may tolerate a little more voltage variation, overvoltage leads to overheating, even when the valve successfully opens, and shortens coil life; undervoltage reduces the MOPD.

The probability of experiencing undervoltage at the moment of opening with AC systems increases when a control transformer of limited capacity supplies power to the solenoid. This type of transformer is commonly used to supply power to low-voltage control systems using class 2 wiring. The situation is aggravated when more than one device served by the same transformer is energized simultaneously.

Condensing water regulators modulate the quantity of water passing through a water-cooled refrigerant condenser in response to the condensing pressure. They are available for use with most refrigerants, including ammonia (R-717). Most manufacturers stress that these valves are designed for use only as operating devices. Where system closure, improper flow, or loss of pressure caused by valve failure can result in personal injury, damage, or loss of property, separate safety devices must be added to the system.

These devices are used on vapor-cycle refrigeration systems to maintain satisfactory condensing pressure. The regulator automatically modulates to correct for both variations in temperature or pressure of the water supply and variations in the quantity of refrigerant gas being pumped into the condenser.

Operation. The condensing water regulator consists of a valve and an actuator linked together, as shown in Figure 40. The actuator consists of a metallic bellows and adjustable spring combination connected to the system high side.

After a compressor starts, the compressor discharge pressure begins to rise. When the opening pressure setting of the regulator spring is reached, the bellows moves to open the valve disk gradually. The regulator continues to open as condenser pressure rises, until water flow balances the required heat rejection. At this point, the condenser pressure is stabilized. When the compressor stops, the continuing water flow through the regulator causes the condenser pressure to drop gradually, and the regulator becomes fully closed when the opening pressure setting of the regulator is reached.

Figure 40. Two-Way Condensing Water Regulator

Selection. The regulator should be selected from the manufacturer’s data on the basis of maximum required flow, minimum available pressure drop, water temperature, and system operating conditions to avoid hunting or internal erosion caused by high pressure drops through an oversized valve because it operates only partially open for most of its duty cycle. Also, depending on the specific refrigerant being used, special components may be required (e.g., stainless steel rather than brass bellows for ammonia).

The water flow required depends on condenser performance, temperature of available water, quantity of heat that must be rejected to the water, and desired operating condenser pressure. For a given opening of the valve, which corresponds to a given pressure rise above the regulator opening point, the flow rate handled by a given size of water regulator is a function of the available water-pressure drop across the valve seat.

Application. Because there are two types of control action available (direct- or reverse-acting), these regulators can be used for various applications (e.g., water-cooled condensers, bypass service on refrigeration systems, ice machines, heat pump systems that control water temperature). For equipment with a large water flow requirement, a small regulator is used as a pilot valve for a diaphragm main valve.

Manufacturers of these types of devices have technical literature to help in applying their products to specific systems.

These regulators are similar to two-way regulators, but they have an additional port, which opens to bypass water around the condenser as the port controlling water flow to the condenser closes (Figure 41). Thus, flow through the cooling tower decking or sprays and the circulating pump is maintained, although the water supply to individual or multiple condensers is modulated for control.

Operation. The three-way regulator operates similar to the two-way regulator. Low refrigerant head pressure, which may result from low cooling-tower water temperature, decreases the refrigeration system’s cooling ability rapidly. The three-way regulator senses the compressor head pressure and allows cooling water to flow to the condenser, bypass the condenser, or flow to both the condenser and bypass line to provide correct refrigerant head pressures.

Figure 41. Three-Way Condensing Water Regulator

The regulator allows water flow to the tower through the bypass line, even though the condenser does not require cooling. This provides an adequate head of water at the tower at all times so the tower can operate efficiently with minimum maintenance on nozzles and wetting surfaces.

Selection. Selection considerations are the same as for two-way regulators, including the cautions about oversizing.

Application. Pressure-actuated three-way regulators are used for condensing units cooled by atmospheric or forced-draft cooling towers requiring individual condenser pressure control. They may be used on single or multiple condenser piping arrangements to the tower to provide the most economical and efficient use.

These regulators must be supplemented by other means if cooling towers are to be operated in freezing weather. An indoor sump is usually required, and a temperature-actuated three-way water control valve diverts condenser leaving water directly to the sump when the water becomes too cold.

Strainers are not generally required with water regulators.

Although precision metal seats may be manufactured nearly bubbletight, they are not economical for refrigerant check valves. Seats made of synthetic elastomers provide excellent sealing at medium and high temperatures, but may leak at low temperatures because of their lack of resilience. Because high temperatures deteriorate most elastomers suitable for refrigerants, plastic materials have become more widely used, despite being susceptible to damage by large pieces of foreign matter.

In compressor discharge lines, check valves are used to prevent flow from the condenser to the compressor during the off cycle or to prevent flow from an operating compressor to an idle compressor. Although a 2 to 6 psi pressure drop is tolerable, the check valve must resist pulsations caused by the compressor and the temperature of discharge gas. Also, the valve must be bubbletight to prevent liquid refrigerant from accumulating at the compressor discharge valves or in the crankcase.

In liquid lines, a check valve prevents reverse flow through the unused expansion device on a heat pump or prevents back-up into the low-pressure liquid line of a recirculating system during defrosting. Although a 2 to 6 psi pressure drop is usually acceptable, the check-valve seat must be bubbletight.

In the suction line of a low-temperature evaporator, a check valve may be used to prevent transfer of refrigerant vapor to a lower-temperature evaporator on the same suction main. In this case, the pressure drop must be less than 2 psi, the valve seating must be reasonably tight, and the check valve must be reliable at low temperatures.

In hot-gas defrost lines, check valves may be used in the branch hot-gas lines connecting the individual evaporators to prevent crossfeed of refrigerant during the cooling cycle when defrost is not taking place. In addition, check valves are used in the hot-gas line between the hot-gas heating coil in the drain pan and the evaporator to prevent pan coil sweating during the refrigeration cycle. Tolerable pressure drop is typically 2 to 6 psi, seating must be nearly bubbletight, and seat materials must withstand high temperatures.

Be careful when sizing check valves, because oversized check valves may chatter or pulsate.

These valves are most commonly a pop-type design, which open abruptly when the inlet pressure exceeds the outlet pressure by the valve setting pressure (Figure 42). Seat configuration is such that once lift begins, the resulting increased active seat area causes the valve seat to pop wide open against the force of the setting spring. Because the flow rate is measured at a pressure of 10% above the setting, the valve must open within this 10% increase in pressure.

Figure 42. Pop-Type Safety Relief Valves

This relief valve operates on a fixed pressure differential from inlet to outlet. Because the valve is affected by back pressure, a rupture disk must not be installed at the valve outlet.

Relief valve seats are made of metal, plastic, lead alloy, or synthetic elastomers. Elastomers are commonly used because they have greater resilience and, consequently, reseat more tightly than other materials. Some valves that have lead-alloy seats have an emergency manual reseating stem that allows reforming the seating surface by tapping the stem lightly with a hammer. Advantages of the pop-type relief valve are simplicity of design, low initial cost, and high discharge capacity.

Capacities of pressure-relief valves are determined by test in accordance with the provisions of the ASME Boiler and Pressure Vessel Code. Relief valves approved by the National Board of Boiler and Pressure Vessel Inspectors are stamped with the applicable code symbol(s). (Consult the Boiler and Pressure Vessel Code for specific text and marking details.) In addition, the pressure setting and capacity are stamped on the valve.

When relief valves are used on pressure vessels of 10 ft³ internal gross volume or more, a relief system consisting of a three-way valve and two relief valves in parallel is required.

Pressure Setting. The maximum pressure setting for a relief device is limited by the design working pressure of the vessel to be protected. Pressure vessels normally have a safety factor of 5. Therefore, the minimum bursting pressure is five times the rated design working pressure (design pressure). The relief device must have enough discharge capacity to prevent pressure in the vessel from rising more than 10% above its design pressure. Because the capacity of a relief device is measured at 10% above its stamped setting, the setting cannot exceed the design pressure of the vessel.

To prevent loss of refrigerant through pressure-relief devices during normal operating conditions, the relief setting must be substantially higher than the system operating pressure. For rupture members, the setting should be 50% above a static system pressure and 100% above a maximum pulsating pressure. Failure to provide this margin of safety causes fatigue of the frangible member and rupture well below the stamped setting.

For relief valves, the setting should be 25% above maximum system pressure. This provides sufficient spring force on the valve seat to maintain a tight seal and still allow for setting tolerances and other factors that cause settings to vary. Although relief valves are set at the factory to be close to the stamped setting, the variation may be as much as 10% after the valves have been stored or placed in service.

Discharge Piping. The size of the discharge pipe from the pressure-relief device or fusible plug must not be less than the size of the pressure-relief device or fusible plug outlet. The maximum length of discharge piping can be obtained using the table or formula provided in section 9.7.9 of ASHRAE Standard 15-2016.

Selection and Installation. When selecting and installing a relief device,

Functional relief valves are usually of the diaphragm type: system pressure acts on a diaphragm that lifts the valve disk from the seat (Figure 43). The other side of the diaphragm is exposed to both the adjusting spring and atmospheric pressure. The ratio of effective diaphragm area to seat area is high, so outlet pressure has little effect on the operating point of the valve.

Because the diaphragm’s lift is not great, the diaphragm valve is frequently built as the pilot or servo of a larger piston-operated main valve to provide both sensitivity and high flow capacity. Construction and performance are similar to the previously described pilot-operated evaporator pressure regulator, except that diaphragm valves are constructed for higher pressures. Thus, they are suitable for use as defrost relief from evaporator to suction pressure, as large-capacity relief from a pressure vessel to the low side, or as a liquid refrigerant pump relief from pump discharge to the accumulator to prevent excessive pump pressures when some evaporators are valved closed.

Figure 43. Diaphragm Relief Valve

Fusible plugs and rupture disks (Figure 44) provide similar safety relief. The former contains a fusible member that melts at a predetermined temperature corresponding to the safe saturation pressure of the refrigerant, but is limited in application to pressure vessels with internal gross volumes of 3 ft³ or less and internal diameters of 6 in. or less. The rupture member contains a preformed disk designed to burst at a predetermined pressure. These devices may be used as stand-alone devices or installed at the inlet to a safety relief valve.

When these devices are installed in series with a safety relief valve, the chamber created by the two valves must have a pressure gage or other suitable indicator. A rupture disk will not burst at its design pressure if back pressure builds up in the chamber.

Figure 44. Safety Relief Devices

The rated relieving capacity of a relief valve alone must be multiplied by 0.9 when it is installed in series with a rupture disk (unless the relief valve has been rated in combination with the rupture disk).

Discharge Capacity. The minimum required discharge capacity of the pressure-relief device or fusible plug for each pressure vessel is determined by the following formula, specified by ASHRAE Standard 15:

(2)

where

Equation (2) determines the required relief capacity for a pressure vessel containing liquid refrigerant. See ASHRAE Standard 15 for other relief device requirements, including relief of overpressure caused by compressor flow rate capacity.

Separators are usually given capacity ratings for several refrigerants at several suction and condensing temperatures. Another method rates capacity in terms of compressor displacement volume. Some separators also show a marked reduction in separation efficiency at some stated minimum capacity. Because compressor capacity increases when suction pressure rises or condensing pressure drops, system capacity at its lowest compression ratio should be the criterion for selecting the separator.

Figure 45. Discharge-Line Lubricant Separator

A capillary tube passes liquid much more readily than vapor because of the latter’s increased volume; as a result, it is a practical metering device. When a capillary tube is sized to allow the desired flow of refrigerant, a liquid seal develops at its inlet. If the system becomes unbalanced, some vapor (uncondensed refrigerant) enters the capillary tube. This vapor reduces the mass flow of refrigerant considerably, which increases condenser pressure and causes subcooling at the condenser exit and capillary tube inlet. The result is increased mass flow of refrigerant through the capillary tube. If properly sized for the application, the capillary tube compensates automatically for load and system variations and gives acceptable performance over a limited range of operating conditions.

A common flow condition is to have subcooled liquid at the entrance to the capillary tube. Bolstad and Jordan (1948) described the flow behavior from temperature and pressure measurements along the tube (Figure 46) as follows:

Figure 46. Pressure and Temperature Distribution along Typical Capillary Tube (Bolstad and Jordan 1948)

Li et al. (1990) and Mikol (1963) showed that the first vapor bubble is not generated at the point where the liquid pressure reaches the saturation pressure (point 2 on Figure 46), but rather the refrigerant remains in the liquid phase for some limited length past point 2, reaching a pressure below the saturation pressure. This delayed evaporation, often referred to as a metastable or superheated liquid condition, must be accounted for in analytical modeling of the capillary tube, or the mass flow rate of refrigerant will be underestimated (Kuehl and Goldschmidt 1991; Wolf et al. 1995).

The rate of refrigerant flow through a capillary tube always increases with an increase in inlet pressure. Flow rate also increases with a decrease in external outlet pressure down to a certain critical value, below which flow does not change (choked flow). Figure 46 shows a case in which outlet pressure inside the capillary tube has reached the critical value (point 3), which is higher than the external pressure (point 4). This condition is typical for normal operation. The point at which the first gas bubble appears is called the bubble point. The preceding portion of capillary tube is called the liquid length, and that following is called the two-phase length.

A capillary tube must be compatible with other components. In general, once the compressor and heat exchangers have been selected to meet the required design conditions, capillary tube size and system charge are determined. However, detailed design considerations may be different for different applications (e.g., domestic refrigerator, window air conditioner, residential heat pump).

Capillary tube size and system charge together are used to determine subcooling and superheat for a given design. Performance at off-design conditions should also be checked. Capillary tube systems are generally much more sensitive to the amount of refrigerant charge than expansion valve systems.

The high-pressure side must be designed for use with a capillary tube. To prevent rupture in case the capillary tube becomes blocked, the high-side volume should be sufficient to contain the entire refrigerant charge. A sufficient refrigerant storage volume (e.g., additional condenser tubes) may also be needed to protect against excessive discharge pressures during high-load conditions.

Pressure equalization during the off period is another concern in designing the high side. When the compressor stops, refrigerant continues to pass through the capillary tube from the high side to the low side until pressures are equal. If liquid is trapped in the high side, it will evaporate during the off cycle, pass through the capillary tube to the low side as a warm gas, condense, and add latent heat to the evaporator. Therefore, good liquid drainage to the capillary tube during this equalization interval should be provided. Liquid trapping may also increase the time for the pressure to equalize after the compressor stops. If this interval is too long, pressures may not be sufficiently equalized to allow low-starting-torque motor compressors to start when the thermostat calls for cooling.

The maximum quantity of refrigerant is in the evaporator during the off cycle and the minimum during the running cycle. Suction piping should be arranged to reduce the adverse effects of the variable-charge distribution. A suitable suction-line accumulator is sometimes needed.

Selection of a capillary tube depends on the application and anticipated range of operating conditions. One approach to the problem involves the concept of capacity balance. A refrigeration system operates at capacity balance when the capillary tube’s resistance is sufficient to maintain a liquid seal at its entrance without excess liquid accumulating in the high side (Figure 47). Only one such capacity balance point exists for any given compressor discharge pressure. A curve through the capacity balance points for a range of compressor discharge and suction pressures (as in Figure 48) is called the capacity balance characteristic of the system. Ambient temperatures for a typical air-cooled system are shown in Figure 48. A given set of compressor discharge and suction pressures associated with condenser and evaporator pressure drops establish the capillary tube inlet and outlet pressures.

Figure 47. Effect of Capillary Tube Selection on Refrigerant Distribution

Figure 48. Capacity Balance Characteristic of Capillary System

The capacity balance characteristic curve for any combination of compressor and capillary tube may be determined experimentally by the arrangement shown in Figure 49. Although Figure 49 shows the capillary tube suction-line heat exchanger application, a similar test setup without heat exchange would be used for adiabatic capillary tube systems. This test arrangement makes it possible to vary suction and discharge pressures independently until capacity balance is obtained. The desired suction pressure may be obtained by regulating heat input to the low side, usually by electric heaters. The desired discharge pressure may be obtained by a suitably controlled water-cooled condenser. A liquid indicator is located at the entrance to the capillary tube. The usual test procedure is to hold high-side pressure constant and, with gas bubbling through the sight glass, slowly increase suction pressure until a liquid seal forms at the capillary tube entrance. Repeating this procedure at various discharge pressures determines the capacity balance characteristic curve similar to that shown in Figure 48. This equipment may also be used as a calorimeter to determine refrigerating system capacity.

Figure 49. Test Setup for Determining Capacity Balance Characteristic of Compressor, Capillary, and Heat Exchanger

Always investigate whether initial capillary tube selection and charge are optimum for the unit, even in simple applications such as a room air-conditioning unit. The refrigerant charge in the unit can be varied using a small refrigerant bottle (valved off and sitting on a scale) connected to the circuit. The interconnecting line must be flexible and arranged so that it is filled with vapor instead of liquid. The charge is brought in or removed from the unit by heating or cooling the bottle.

The only test for varying capillary tube restriction is to remove the tube, install a new capillary tube, and determine the optimum charge, as outlined previously. Occasionally, pinching the capillary tube is used to determine whether increased resistance is needed.

The refrigeration system should be operated through its expected range of conditions to determine power requirements and cooling capacity for any given selection and charge combination.

Processing and Inspection. To prevent mechanical clogging by foreign particles, a strainer should be installed ahead of the capillary tube. Also, all parts of the system must be evacuated adequately to eliminate water vapor and noncondensable gases, which may cause clogging by freezing and/or corrosion. The lubricant should be free from wax separation at the minimum operating temperature.

The interior surface of the capillary tube should be smooth and uniform in diameter. Although plug-drawn copper is more common, wire-drawn or sunk tubes are also available. Life tests should be conducted at low evaporator temperatures and high condensing temperatures to check the possibility of corrosion and plugging. Material specifications for seamless copper tube are given in ASTM Standard B75, and for hard-drawn copper tubes in ASTM Standard B360.

Establish a procedure to ensure uniform flow capacities, within reasonable tolerances, for all capillary tubes used in product manufacture. The following procedure is a good example:

Considerations. In selecting a capillary tube for a specific application, practical considerations influence the length. For example, the minimum length is determined by geometric considerations such as the physical distance between the high and low sides and the length of capillary tube required for optimum heat exchange. It may also be dictated by exit velocity, noise, and the possibility of plugging with foreign materials. Maximum length may be determined primarily by cost. It is fortunate, therefore, that flow characteristics of a capillary tube can be adjusted independently by varying either its bore or its length. Thus, it is feasible to select the most convenient length independently and then (within certain limits) select a bore to give the desired flow. An alternative procedure is to select a standard bore and then adjust the length, as required.

Figure 50. Mass Flow Rate of R-134a Through Capillary Tube (Capillary tube ID is 0.034 in. and length is 130 in.)

Figure 51. Flow Rate Correction Factor ϕ for R-134a

ASTM Standard B360 lists standard diameters and wall thicknesses for capillary tubes. Many nonstandard tubes are also used, resulting in nonuniform interior surfaces and variations in flow.

Wolf and Pate (2002) developed refrigerant-specific rating charts to predict refrigerant flow rates through capillary-tube/suction-line heat exchangers. The methodology involves determining four quantities (three for quality inlet conditions) from a series of curves and multiplying the results together to obtain the flow rate prediction. The calculation procedure is very similar to the adiabatic capillary tube selection process, with the exception of the additional flow rate factors. Figures 58 to 62 are the rating charts for pure R-134a through capillary-tube/suction-line heat exchangers. Figure 58 plots refrigerant flow rate as a function of capillary tube inlet pressure and inlet condition for a reference heat exchanger configuration (D_c = 0.034 in., L_c = 130 in., and L_hx = 60 in.). The thermodynamic state of the refrigerant vapor at the suction-line inlet to the heat exchanger has also been fixed (24 psia and 20°F superheat). Figure 59 plots flow rate correction factor ϕ₁ as a function of capillary tube diameter and length for subcooled inlet conditions at the capillary tube inlet. Figure 60 plots flow rate correction factor ϕ₂ as a function of suction-line inlet pressure and superheat level at the inlet to the suction line for subcooled inlet conditions at the capillary tube inlet. Figure 61 plots flow rate correction factor ϕ₃ as a function of heat exchange length for subcooled inlet conditions at the capillary tube inlet. Figure 62 is the capillary tube geometry correction factor for quality inlet conditions at the capillary tube inlet, and Figure 63 is the suction-line condition correction factor for quality inlet conditions at the capillary tube inlet. There is no ϕ₃ correction factor for heat exchange length for quality inlet conditions. In general, heat exchange length delays the onset of vaporization of the refrigerant inside the capillary tube. However, without recondensation in the capillary tube as a result of the heat exchange effect, there is no vaporization for two-phase conditions at the capillary tube inlet.

Figure 60. Suction-Line Condition Correction Factor for R-134a Subcooled Inlet Conditions

Figure 61. Heat Exchange Length Correction Factor for R-134a Subcooled Inlet Conditions

Figure 62. Capillary Tube Geometry Correction Factor for R-134a Quality Inlet Conditions

To predict refrigerant mass flow rate for other conditions and geometries, multiply the necessary correction factors by the reference flow rate from Figure 58. Additional design variables examined by Wolf and Pate (2002) were the adiabatic upstream entrance length and inner diameter of the suction-line tubing. For adiabatic entrance lengths from 6 to 24 in., there was no observed effect. There was also no effect observed that could be attributed to the inner diameter of the suction line. Because of their lack of effect, these two design variables are not included when predicting refrigerant flow rate.

Figure 63. Suction-Line Condition Correction Factor for R-134a Quality Inlet Conditions

Based on tests with R-134a, R-22, R-410A, and R-600a, Wolf and Pate (2002) developed a general method for predicting refrigerant mass flow rate through a capillary-tube/suction-line heat exchanger. In this method, the Buckingham π theorem was applied to the physical factors and fluid properties that affect capillary tube flow rate through a capillary-tube/suction-line heat exchanger. Physical factors included capillary tube diameter, capillary tube length, suction-line diameter, heat exchange length, adiabatic entrance length of the capillary tube, capillary tube inlet pressure and condition, and suction-line inlet pressure and condition. Fluid properties included specific volume, viscosity, specific heat, and latent heat of vaporization. The result of this analysis was a group of 15 dimensionless π parameters, shown in Table 4.

All fluid properties for the respective π terms are evaluated at the capillary tube inlet temperature. Separate π₇ terms are presented for subcooled and quality inlet conditions to the capillary tube. The π terms in Table 4 that do not appear in Equations (5) and (6) were statistically insignificant and were not included in the final correlations.

Regression analysis of refrigerant flow rate data for subcooled inlet conditions produced the following equation:

(5)

where π₇ = ΔT_scC_pfc D_c²/μ_fc² ν_fc² and 2°F < ΔT_sc < 30°F.

Equation (6) is valid for quality inlet conditions and contains the quality inlet term π₇ = 1 – x.

(6)

where x = quality (decimal) from 0.02 < x < 0.10.

Wolf and Pate (2002) compared these equations to R-152a test results as well as published R-134a, R-152a, and R-12 data. Excellent agreement (±10%) was observed in nearly all cases. The only exception was the R-134a results by Dirik et al. (1994). Equation (5) overpredicted these results by an average of 45%. However, the adiabatic entrance length used was approximately 134 in., indicating that very long adiabatic entrance lengths may decrease refrigerant flow rates.

For further information on other refrigerants, please see the Bibliography.

Short-tube restrictors are widely used in residential air conditioners and heat pumps. They offer low cost, high reliability, ease of inspection and replacement, and potential elimination of check valves in the design of a heat pump. Because of their pressure-equalizing characteristics, short-tube restrictors allow the use of a low-starting-torque compressor motor.

Short-tube restrictors, as used in residential systems, are typically 3/8 to 1/2 in. in length, with a length-to-diameter (L/D) ratio greater than 3 and less than 20. Short-tube restrictors are also called plug orifices or orifices, although the latter is reserved for restrictors with an L/D ratio less than 3. Capillary tubes have an L/D ratio much greater than 20.

An orifice tube, a type of short-tube restrictor, is commonly used in automotive air conditioners. Its L/D ratio falls between that of a capillary tube and a short-tube restrictor. Most automotive applications use orifice tubes with L/D ratios between 21 and 35 and inside diameters from 0.04 to 0.08 in. An orifice tube allows the evaporator to operate in a flooded condition, which may improve performance on some systems. To prevent liquid from flooding the compressor, an accumulator/dehydrator is installed to separate liquid from vapor and to allow a small amount of lubricant to return to the compressor. However, the accumulator/dehydrator does cause a pressure drop penalty on the suction side.

Figure 64. Schematic of Movable Short-Tube Restrictor

Figure 65. R-22 Pressure Profile at Various Downstream Pressures with Constant Upstream Conditions: L = 0.5 in., D = 0.053 in., Subcooling 25°F (Adapted from Aaron and Domanski 1990)

There are two basic designs for short-tube restrictors: stationary and movable. Movable short-tube restrictors consist of a piston that moves within its housing (Figure 64). A movable short-tube restrictor restricts refrigerant flow in one direction. In the opposite direction, the refrigerant pushes the restrictor off its seat, opening a larger area for the flow. The stationary design is used in units that cool only; movable short-tube restrictors are used in heat pumps that require different flow restrictions for cooling and heating modes. Two movable short-tube restrictors, installed in series and faced in opposite directions, eliminate the need for check valves, which would be needed for capillary tubes and thermostatic expansion valves.

The refrigerant mass flow rate through a short-tube restrictor depends strongly on upstream subcooling and upstream pressure. For a given inlet pressure, inlet subcooling, and downstream pressure below the saturation pressure corresponding to the inlet temperature, flow has a very weak dependence on the downstream pressure, indicating a nearly choked flow. This flow dependence is shown in Figure 65, which presents R-22 test data obtained on a 0.5 in. long, laboratory-made short-tube restrictor at three different downstream pressures and the same upstream pressure.

Figure 66. R-22 Mass Flow Rate Versus Condenser Pressure for Reference Short Tube: L = 0.5 in., D = 0.053 in., Sharp-Edged (Adapted from Aaron and Domanski 1990)

A significant drop in downstream pressure from approximately 170 psia to 70 psia produces a smaller increase in the mass flow rate than does a modest change of downstream pressure from 190 psia to 170 psia. Pressure drops only slightly along the length of the short-tube restrictor. The large pressure drop at the entrance is caused by rapid fluid acceleration and inlet losses. The large pressure drop in the exit plane, typical for heat pump operating conditions and represented in Figure 65 by the bottom pressure line, indicates that choked flow nearly occurred.

Among geometric parameters, the short-tube restrictor diameter has the strongest influence on mass flow rate. Chamfering the inlet of the short-tube restrictor may increase the mass flow rate by as much as 25%, depending on the L/D ratio and chamfer depth. Chamfering the exit causes no appreciable change in mass flow rate.

Although refrigerant flow inside a short tube is different from flow inside a capillary tube, choked flow is common for both, making both types of tubes suitable as metering devices. Systems equipped with short-tube restrictors, as with capillary tubes, must be precisely charged with the proper amount of refrigerant. Inherently, a short-tube restrictor does not control as well over a wide range of operating conditions as does a thermostatic expansion valve. However, its performance is generally good in a properly charged system.

Figures 66 to 69, from Aaron and Domanski (1990), can be used for preliminary evaluation of the mass flow rate of R-22 at a given inlet pressure and subcooling in air-conditioning and heat pump applications (i.e., applications in which downstream pressure is below the saturation pressure of refrigerant at the inlet). The method requires determination of mass flow rate for the reference short tube from Figure 66 and modifying the reference flow rate with multipliers that account for the short-tube geometry according to the equation

(7)

where

Figure 67. Correction Factor for Short-Tube Geometry (R-22) (Adapted from Aaron and Domanski 1990)

Figure 68. Correction Factor for L/D Versus Subcooling (R-22) (Adapted from Aaron and Domanski 1990)

Aaron and Domanski (1990) also provide a more accurate correlation than the graphical method. Neglecting downstream pressure on the graphs may introduce an error in the prediction as compared to the correlation results; however, this discrepancy should not exceed 3% because of the choked-flow condition at the tube exit. Note: The lines for 0 and 40°F subcooling in Figure 66 were obtained by extrapolation beyond the test data and may carry a large error.

Additional research has been done to add several other refrigerants’ capacity ratings to this section. However, only data for R-410A and R-407c have been published in ASHRAE Transactions (Payne and O’Neal 1998, 1999).

Figure 69. Correction Factor for Inlet Chamfering (R-22) (Adapted from Aaron and Domanski 1990)