Refrigeration Machines

 REFRIGERATION BASICS

 

Vapor Compression Refrigeration Cycle

The term refrigeration, as part of a building HVAC system, generally refers to avapor compression system wherein a chemical substance alternately changes from liquid to gas (evaporating, thereby absorbing heat and providing a cooling effect) and from gas to liquid (condensing, thereby releasing heat). This “cycle” actually consists of four steps:

1. Compression: Low-pressure refrigerant gas is compressed, thus raising its pressure by expending mechanical energy. There is a corresponding increase in temperature along with the increased pressure.

2. Condensation: The high-pressure, high-temperature gas is cooled by outdoor air or water that serves as a “heat sink” and condenses to a liquid form at high pressure.

3. Expansion: The high-pressure liquid flows through an orifice in the expansion valve, thus reducing the pressure. A small portion of the liquid “flashes” to gas due to the pressure reduction.

4. Evaporation: The low-pressure liquid absorbs heat from indoor air or water and evaporates to a gas or vapor form. The low-pressure vapor flows to the compressor and the process repeats.

 As shown in Figure 1.1, the vapor compression refrigeration system consists of four components that perform the four steps of the refrigeration cycle. The compressor raises the pressure of the initially low-pressure refrigerant gas. The condenser is a heat exchanger that cools the high-pressure gas so that it changes phase to liquid. The expansion valve controls the pressure ratio, and thus flow rate, between the high- and low-pressure regions of the system. The evaporator is a heat exchanger that heats the low pressure liquid, causing it to change phase from liquid to vapor (gas).

Thermodynamically, the most common representation of the basic refrigeration cycle is made utilizing a pressure enthalpy chart as shown in Figure 1.2. For each refrigerant, the phase-change line represents the conditions of pressure and total heat content (enthalpy) at which it changes from liquid to gas and vice versa. Thus each of the steps of the vapor compression cycle can easily be plotted to demonstrate the actual thermodynamic processes at work, as shown in Figure 1.3. Point 1 represents the conditions entering the compressor. Compression of the gas raises its pressure from P1 to P2. Thus the “work” that is done by the compressor adds heat to the refrigerant, raising its temperature and slightly increasing its heat content. Point 2 represents the condition of the refrigerant leaving the compressor and entering the condenser. In the condenser, the gas is cooled, reducing its enthalpy from h2 to h3. Point 3 to point 4 represents the pressure reduction that occurs in the expansion process. Due to a small percentage of the liquid evaporating because of the pressure reduction, the temperature and enthalpy of the remaining liquid is also reduced slightly. Point 4, then, represents the condition entering the evaporator. Point 4 to point 1 represents the heat gain by the liquid, increasing its enthalpy from h4 to h1, completed by the phase change from liquid to gas at point 1.

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FIGURE 1.1. Basic components of the vapor compression refrigeration system.

FIGURE 1.2. Basic refrigerant pressure –enthalpy relationship.

FIGURE 1.3. Ideal refrigeration cycle imposed over pressure –enthalpy chart.

For any refrigerant whose properties are known, a pressure-enthalpy chart

can be constructed and the performance of a vapor compression cycle analyzed

by establishing the high and low pressures for the system. (Note that Figure 1.3

represents an “ideal” cycle and in actual practice there are various departures

dictated by second-law inefficiencies.)

Refrigerants

Any substance that absorbs heat may be termed a refrigerant. Secondary

refrigerants, such as water or brine, absorb heat but do not undergo a phase

change in the process. Primary refrigerants, then, are those substances that possess the chemical, physical, and thermodynamic properties that permit their efficient use in the typical vapor compression cycle.

In the vapor compression cycle, a refrigerant must satisfy several (and

sometimes conflicting) requirements:

1. The refrigerant must be chemically stable in both the liquid and vapor state.

2. Refrigerants for HVAC applications must be nonflammable and have low toxicity.

3. Finally, the thermodynamic properties of the refrigerant must meet the temperature and pressure ranges required for the application. Early refrigerants, developed in the 1920s and 1930s, used in HVAC applications were predominately chemical compounds made up of chlorofluorocarbons (CFCs) such as R-11, R-12, and R-503. While stable and efficient in the range of temperatures and pressures required for HVAC use, any escaped refrigerant gas was found to be long-lived in the atmosphere. In the lower atmosphere, the CFC molecules absorb infrared radiation and, thus, contribute to atmospheric warming. Once in the upper atmosphere, the CFC molecule breaks down to release chlorine that destroys ozone and, consequently, damages the atmospheric ozone layer that protects the earth from excess UV radiation.

The manufacture of CFC refrigerants in the United States and most other industrialized nations was eliminated by international agreement in 1996. While there is still refrigeration equipment in use utilizing CFC refrigerants, no new equipment using these refrigerants is now available in the United States. Researchers found that by modifying the chemical compound of CFCs by substituting a hydrogen atom for one or more of the chlorine or fluorine atoms resulted in a significant reduction in the life of the molecule and, thus, reduced the negative environmental impact it may have. These new compounds, called hydrochlorofluorocarbons (HCFCs), are currently used in HVAC refrigeration

systems as R-22 and R-123. While HCFCs have reduced the potential environmental damage by refrigerants released into the atmosphere, the potential for damage has not been totally eliminated. Again, under international agreement, this class of refrigerants is slated for phaseout for new equipment installations by 2010– 2020, with total halt to manufacturing and importing mandated by 2030, as summarized in Table 1.1. A third class of refrigerants, called hydrofluorocarbons (HFCs), are not regulated by international treaty and are considered, at least for the interim, to be the most environmentally benign compounds and are now widely used in HVAC refrigeration systems. HVAC refrigeration equipment is currently undergoing transition in the use of refrigerants. R-22 has been the workhorse for positive displacement compressor systems in HVAC applications. The leading replacements for R-22 in new equipment are R-410A and, to a lesser extent, R-407C, both of which are blends of HFC compounds. R-134A, an HFC refrigerant, is utilized in positive-pressure compressor systems. At least one manufacturer continues to offer negative-pressure centrifugal compressor water chillers using R-123 (an HCFC), but it is anticipated that, by 2010, the manufacture of new negative pressure chillers using HCFCs will be terminated. These same manufacturers already market a positive pressure centrifugal compressor systems using R-134A (though one manufacturer currently limits sales to outside of the United States since many countries,

particularly in Europe, have accelerated the phaseout of HCFCs).

Based on the average 20– 25 year life for a water chiller (see Chap. 8) and the HCFC refrigerant phaseout schedule summarized in Table 1.1, owners should avoid purchasing any new chiller using R-22. After 2005, owners should avoid purchasing new chillers using R-123. ASHRAE Standard 34-1989 classifies refrigerants according to their toxicity (A = nontoxic and B = evidence of toxicity identified) and flammability (1 =no flame progation, 2 = low flammability, and 3 = high flammability). Thus, all refrigerants fall within one of the six “safety groups,” A1, A2, A3, B1, B2, or B3. For HVAC refrigeration systems, only A1 refrigerants should be considered. Table 1.2 lists the safety group classifications for refrigerants commonly used in HVAC applications.

 CHILLED WATER FOR HVAC APPLICATIONS

The basic vapor compression cycle, when applied directly to the job of building cooling, is referred to as a direct expansion (DX) refrigeration system. This reference comes from the fact that the building indoor air that is to be cooled passes “directly” over the refrigerant evaporator without a secondary refrigerant being utilized. While these cooling systems are widely use in residential, commercial, and industrial applications, they have application, capacity, and performance limitations that reduce their use in larger, more complex HVAC applications. For these applications, the use of chilled water systems is dictated. Typical applications for chilled water systems include large buildings (offices, laboratories, etc.) or multibuilding campuses where it is desirable to provide cooling from a central facility. As shown in Figure 1.4, the typical water-cooled HVAC system has three heat transfer loops: Loop 1 Cold air is distributed by one or more air-handling units to the spaces within the building. Sensible heat gains, including heat from temperature-driven transmission through the building envelope; direct solar radiation through windows; infiltration; and internal heat from people, lights, and equipment, are “absorbed” by the cold air, raising its temperature. Latent heat gains, moisture added to the space by air infiltration, people, and equipment, is also absorbed by the cold air, raising its specific humidity. The resulting space temperature and humidity condition is an exact balance between the sensible and latent heat gains and capability of the entering cold air to absorb those heat gains. Loop 2 The distributed air is returned to the air handling unit, mixed with the required quantity of outdoor air for ventilation, and then directed over the cooling coil where chilled water is used to extract heat from the air, reducing both its temperature and moisture content so it can be distributed once again to the space. As the chilled water passes through the cooling coil in counter flow to the air, the heat extraction process results in increased water temperature. The chilled water temperature leaving the cooling coil (chilled water return) will be 8– 168F warmer than the entering water temperature

 (chilled water supply) at design load. This temperature difference (range ) establishes the flow requirement via the relationship shown in Eq. 1.1.

Fchw  =Q/(500 × Range)

where Fchw = chilled water flow rate (gpm), Q = total cooling system

load (Btu/hr), Range = chilled water temperature rise (F), 500 = conversion factor (Btu min/gal F hr) (1 Btu/lb F × 8.34 lb/gal × 60min/hr).

The warmer return chilled water enters the water chiller where it is cooled to the desired chilled water supply temperature by transferring the heat extracted from the building spaces to a primary refrigerant. This process, obviously, is not “free” since the compressor must do work on the

refrigerant for cooling to occur and, thus, must consume energy in the process. Since most chillers are refrigerant-cooled, the compressor energy, in the form of heat, is added to the building heat and both must be rejected through the condenser.

Loop 3 The amount of heat that is added by the compressor depends on the efficiency of the compressor. This heat of compression must then be added to the heat load on the chilled water loop to establish the amount of heat that must be rejected by the condenser to a heat sink, typically the outdoor air.

 Determining Chilled Water Supply Temperature

The first step in evaluating a chilled water system is to determine the required

chilled water supply temperature. For any HVAC system to provide simultaneous control of space temperature and humidity, the supply air temperature must be low enough to simultaneously satisfy both the sensible and latent cooling loads imposed.

Sensible cooling is the term used to describe the process of decreasing the

temperature of air without changing the moisture content of the air. However, if moisture is added to the room by the occupants, infiltrated outdoor air, internal processes, etc., the supply air must be cooled below its dew point to remove this excess moisture by condensation. The amount of heat removed with the change in moisture content is called latent cooling. The sum of the two represents the total cooling load imposed by a building space on the chilled water cooling coil.

The required temperature of the supply air is dictated by two factors:

1. The desired space temperature and humidity setpoint and

2. The sensible heat ratio (SHR) defined by dividing the sensible cooling load by the total cooling load.

 On a psychrometric chart, the desired space conditions represents one end point of a line connecting the cooling coil supply air conditions and the space conditions. The slope of this line is defined by the SHR. An SHR of 1.0 indicates that the line has no slope since there is no latent cooling. The typical SHR in comfort HVAC applications will range from about 0.85 in spaces with a large number of people to approximately 0.95 for the typical office. The intersection between this “room” line and the saturation line on the psychrometric chart represents the required apparatus dewpoint (ADP) temperature for the cooling coil. However, since no cooling coil is 100% efficient, the air leaving the coil will not be at a saturated condition, but will have a temperature about 1 –2 F above the ADP temperature. While coil efficiencies as high as 98% can be obtained, the economical approach is to select a coil for about 95% efficiency, which typically results in the supply air wet bulb temperature being about 1F lower than the supply air dry bulb temperature.

Based on these typical coil conditions, the required supply air temperature can determined by plotting the room conditions point and a line having a slope equal to the SHR passing through the room point, determining the ADP temperature intersection point, and then selecting a supply air condition on this line based on a 95% coil efficiency. Table 1.3 summarizes the results of this analysis for several different typical HVAC room design conditions and SHRs. For a chilled water cooling coil, approach is defined as the temperature difference between the entering chilled water and the leaving (supply) air. While this approach can range as low as 3 F to as high as 10 F, the typical value for HVAC applications is approximately 7 F. Therefore, to define the required chilled water supply temperature, it is only necessary to subtract 7 F from the supply air dry bulb temperature determined from Table 1.3.

 

Establishing the Temperature Range

Once the required chilled water supply temperature is determined, the desired

temperature range must be established.

From Eq. 1.1, the required chilled water flow rate is dictated by the imposed

cooling load and the selected temperature range. The larger the range, the lower the flow rate and, thus, the less energy consumed for transport of chilled water through the system. However, if the range is too large, chilled water coils and other heat exchangers in the system require increased heat transfer surface and, in

some cases, the ability to satisfy latent cooling loads is reduced. Historically, a 108F range has been used for chilled water systems, resulting in a required flow rate of 2.4 gpm/ton of imposed cooling load. For smaller systems with relatively short piping runs, this range and flow rate are acceptable. However, as systems get larger and piping runs get longer, the use of higher ranges will reduce pumping energy requirements. Also, lower flow rates can also result in economies in piping installation costs since smaller sized piping may be

used. At a 12 F range, the flow rate is reduced to 2.0 gpm/ton and, at a 14 F

range, to 1.7 gpm/ton. For very large campus systems, a range as great as 16 F (1.5 gpm/ton) to 20 F (1.2 gpm/ton) may be used.

VAPOR COMPRESSION CYCLE CHILLERS

As introduced in Section 1.1, a secondary refrigerant is a substance that does not change phase as it absorbs heat. The most common secondary refrigerant is water and chilled water is used extensively in larger commercial, institutional, and industrial facilities to make cooling available over a large area without introducing a plethora of individual compressor systems. Chilled water has the advantage that fully modulating control can be applied and, thus, closer temperature tolerances can be maintained under almost any load condition. For very low temperature applications, such as ice rinks, an antifreeze component, most often ethylene or propylene glycol, is mixed with the water and the term brine (left over from the days when salt was used as antifreeze) is used to describe the secondary refrigerant. In the HVAC industry, the refrigeration machine that produces chilled water is generally referred to as a chiller and consists of the compressor(s), evaporator, and condenser, all packaged as a single unit. The condensing medium may be water or outdoor air. The evaporator, called the cooler, consists of a shell-and-tube heat exchanger with refrigerant in the shell and water in the tubes. Coolers are designed for 3 –11 fps water velocities when the chilled water flow rate is selected for a 10 –20F range.

 For air-cooled chillers, the condenser consists of an air-to-refrigerant heat exchanger and fans to provide the proper flow rate of outdoor air to transfer the heat rejected by the refrigerant. For water-cooled chillers, the condenser is a second shell-and-tube heat exchanger with refrigerant in the shell and condenser water in the tubes. Condenser water is typically supplied at 70 – 85 F and the flow rate is selected for a 10 –15 F range. A cooling tower is typically utilized to provide condenser water cooling, but other cool water sources such as wells, ponds, and so on, can be used. 

Positive Displacement Compressors

Water chillers up to about 100 tons capacity typically utilize one or more positive displacement type reciprocating compressors. The reciprocating compressor uses pistons in cylinders to compress the refrigerant gas. Basically, it works much like a 2-cycle engine except that the compressor consumes shaft energy rather than producing it. Refrigerant gas enters the cylinder through an intake valve on the downward stroke of the piston. The intake valve closes as the piston starts its upward compression stroke, and when the pressure is high enough to overcome the spring resistance, the discharge valve opens and the gas leaves the cylinder. The discharge valve closes as the piston reaches top-dead-center and the cycle repeats itself as the piston starts down with another intake stroke. The pistons are connected to an offset lobed crankshaft via connecting rods. The compressor motor rotates the crankshaft, and this rotational motion is transformed to a reciprocating motion for the pistons. Control of the reciprocating compressor refrigeration system is fairly simple. At the compressor, a head-pressure controller senses the compressor discharge pressure and opens the unloaders on the compressor if this pressure rises above the setpoint. The unloader is a simple valve that relieves refrigerant gas from the high-pressure discharge side of the compressor into the low-pressure suction side, thus effectively raising the inlet pressure and reducing the net pressure difference that is required of the compressor. The high-pressure setpoint is based on the condensing requirements and is normally a pressure corresponding to approximately 1058F for the refrigerant, (R-22 or R-410A). A temperature sensor located on the suction line leaving the evaporator modulates the expansion valve to maintain the setpoint. Thus, as the load on the evaporator changes, the flow rate through the expansion valve is changed correspondingly. The expansion valve sensor will detect an increased temperature (i.e., superheat) if the flow rate is too low and a decreased temperature (i.e., subcooling) if the flow rate is too high. This temperature setpoint is typically 40 F for comfort applications. Reciprocating water chillers larger than about 20 tons capacity are almost always multiple-compressor units. In the selection of a multiple compressor    chiller, it is important that the compressors have independent refrigerant circuits so that in the event of one compressor failing, the remaining one(s) can continue to operate. Some lower-cost units will have all the compressors operating in parallel on one refrigerant circuit.

 Rotary Compressors

For larger capacities (100 tons to over 10,000 tons), rotary compressor water chillers are utilized. There are two types of rotary compressors applied: positive displacement rotary screw compressors and centrifugal compressors. Figure 1.6 illustrates the rotary helical screw compressor operation. Screw compressors utilize double mating helically grooved rotors with “male” lobes and “female” flutes or gullies within a stationary housing. Compression is obtained by direct volume reduction with pure rotary motion. As the rotors begin to unmesh, a void is created on both the male and the female sides, allowing refrigerant gas to flow into the compressor. Further rotation starts the meshing of another male lobe with a female flute, reducing the occupied volume, and compressing the trapped gas. At a point determined by the design volume ratio, the discharge port is uncovered and the gas is released to the condenser. Capacity control of screw compressors is typically accomplished by opening and closing a slide valve on the compressor suction to throttle the flow rate of refrigerant gas into the compressor. Speed control can also be used to control capacity. The design of a centrifugal compressor for refrigeration duty originated with Dr. Willis Carrier just after World War I. The centrifugal compressor raises the pressure of the gas by increasing its kinetic energy. The kinetic energy is converted to static pressure when the refrigerant gas leaves the compressor and expands into the condenser. Figure 1.5 illustrates a typical centrifugal water chiller configuration. The compressor and motor are sealed within a single casing and refrigerant gas is utilized to cool the motor windings during operation. Low-pressure gas flows from the cooler to the compressor. The gas flow rate is controlled by a set of preswirl inlet vanes that regulate the refrigerant gas flow rate to the compressor in response to the cooling load imposed on the chiller. Normally, the output of the chiller is fully variable within the range 15– 100% of full-load capacity. The high-pressure gas is released into the condenser, where water absorbs the heat and the gas changes phase to liquid. The liquid, in turn, flows into the cooler, where it is evaporated, cooling the chilled water. Centrifugal compressor chillers using R-134A or R-22 are defined as positive-pressure machines, while those using R-123 are negative-pressure machines, based on the evaporator pressure condition. At standard ARI rating conditions and using R-134A, the evaporator pressure is 36.6 psig and the condenser pressure is 118.3 psig, yielding a total pressure increase or lift provided by the compressor of 81.7 psig. However, for R-123, these pressure conditions are 25:81 psig in the evaporator and 6.10 psig in the condenser, yielding a total lift of 11.91 psig. Mass flow rates for both refrigerants are essentially the same at approximately 3 lb/min ton. However, due to the significantly higher density of R-134A, its volumetric flow rate (cfm/ton), which defines impeller size, is over five times smaller than R-123 volumetric flow rate. Compressors using R-123 typically use large diameter impellers (approximately 40 in. diameter) and direct-coupled motors that (at 60 Hz) turn at 3600 rpm. Compressors using R-134A typically use much smaller impellers (about 5 in. diameter) that are coupled to the motor through a gearbox or speed increaser and can operate at speeds approaching 30,000 rpm. The large wheel diameters required by R-123 puts a design constraint on the compressor and, to reduce the diameter, they typically utilize two or three impellers in series or stages to produce an equivalent pressure increase.

In practice, the flow paths from the outlet of one stage to the inlet of the next introduce pressure losses that reduce efficiency to some degree. Since the evaporator in positive-pressure chillers is maintained at a

pressure well above atmospheric, any leaks in the refrigeration system will result in a loss of refrigerant and the effect of any leaks is quickly evidenced by low refrigerant levels in the chiller. However, any leaks associated with a negative- pressure machine result in the introduction of atmospheric air (consisting of noncondensable gases and water vapor) into the chiller. Noncondensable gases create two problems:

1. The compressor does work when compressing the noncondensable

gases, but they offer no refrigerating effect.

2. Noncondensable gases can “blanket” evaporator and condenser tubes,

lowering heat exchanger effectiveness.

Noncondensable gases can lower the efficiency of the chiller by as much as 14%

at full load.

Moisture introduced with atmospheric air is a contaminant that can allow

the formation of acids within the chiller that can cause serious damage to motor

windings (of hermetic motors) and bearings.

To remove potential noncondensable gases and moisture from negative-

pressure chillers, these chillers are furnished with purge units. While purge units

are very efficient at separating and venting noncondensable gases and moisture

from the refrigerant, they are not 100% efficient and some refrigerant is vented to

the atmosphere each time the purge unit operates. Additionally, to reduce the

potential for leaks when chillers are off, the evaporator should be provided with

an external heater to raise the refrigerant pressure to above atmospheric.

The energy requirement for a rotary compressor chiller at peak load is a

function of (1) the required leaving chilled water temperature, and (2) the

temperature of the available condenser water. As the leaving chilled water

temperature is reduced, the energy requirement to the compressor increases, as

summarized in Table 1.4. Similarly, as the condenser water temperature

increases, the compressor requires more energy (see Chap. 10). Thus, the

designer can minimize the cooling energy input by utilizing a rotary compressor

chiller selected to operate with the highest possible leaving chilled water

temperature and the lowest possible condenser water temperature.

Electric-Drive Chillers

Chiller efficiency is typically defined in terms of its coefficient of performance

(COP). The COP is the ratio of output Btus divided by the input Btus. If the nominal rating of the chiller is 1 ton of refrigeration capacity, equivalent to 12,000 Btu/hr output, and the input energy is equivalent to 1 kW, or 3,413 Btu/hr, the resulting COP is 12,000/3,413 or 3.52. Air-cooled reciprocating compressor water chillers have a peak load power requirement of 1.0–1.3 kW/ton, depending on capacity and ambient air temperature. Thus, the peak load COP for these units will range from 3.52 to 2.70. Typical rotary compressor water chillers with water-cooled condensing have a peak load power requirement of 0.5–0.7 kW/ton, resulting in a COP of 7.0–5.0. The energy consumption by a rotary compressor chiller decreases as the imposed cooling load is reduced, as shown in Figure 1.7. These chillers operate efficiently at between approximately 30 and 90% load and most efficiently

between 40 and 80% load. Under these conditions, the gas flow rate is reduced,

yet the full heat exchange surface of the cooler and condenser are still available, resulting in higher heat transfer efficiency. Below about 30% load, the refrigerant gas flow rate is reduced to the point where (1) heat pickup from the motor and (2) mechanical inefficiencies have

stabilized input energy requirements.

The vast majority of electric-drive rotary compressor water chillers utilize a single compressor. However, if the imposed cooling load profile indicates there will

 be significant chiller usage at or below 30% of peak load, it may be advantageous to use a dual compressor chiller or multiple single compressor chillers. The dual compressor chiller typically uses two compressors, each sized

for 50% of the peak load. At 50 –100% of design load, both compressors

operate. However, if the imposed load drops below 50% of the design value,

one compressor is stopped and the remaining compressor is used to satisfy the imposed load. This configuration has the advantage of reducing the inefficient operating point to 15% of full load (50% of 30%), reducing significantly the operating energy penalties that would result from a single compressor operation.

Negative-pressure chillers are typically somewhat more efficient than positive-pressure chillers. A peak load rating of 0.5 kW/ton or less is available for

negative-pressure chillers, while positive-pressure chiller ratings below

0.55 kW/ton are difficult to obtain.

Positive-pressure chillers tend to be smaller and lighter than negative-

pressure chillers, which can result in smaller chiller rooms and lighter structures. Negative-pressure chillers generally have a higher first cost than positive- pressure machines. When purchasing a chiller, owners must decide if the improved efficiencies

of negative pressure chillers are worth the additional first cost, the environmental impact of releasing refrigerants, the higher toxicity of R 123, and the impact of the phaseout of HCFC refrigerants.

Engine-Drive Chillers

Natural gas and propane fueled spark ignition engines have been applied to rotary compressor systems. The full-load cooling COP’s for engine-driven chillers are approximately 1.0 for reciprocating compressors, 1.3 –1.9 for screw compressors, and 1.9 for centrifugal compressors. These low COP’s can be improved if the engine water jacket heat and exhaust heat can be recovered to heat service hot water or for other uses. Engine-drive chillers have been around for many years, but their application, most typically utilizing natural gas for fuel, has been limited by a number of factors:

1. Higher first cost.

2. Air quality regulations.

3. Much higher maintenance requirements.

4. Short engine life.

5. Noise.

6. Larger physical size.

7. Lack of integration between engine and refrigeration subsystems.

 Since the mid-1980s, manufacturers have worked very hard to reduce these negatives with more compact designs, emissions control systems, noise abatement measures, basic engine improvements, and development of overall systems controls using microprocessors. However, the maintenance requirements for engine-drive chillers remains high, adding about $0.02/ton hr to the chiller operating cost. Currently, the engines used for chillers are either spark-ignition engines based on automotive blocks, heads, and moving components (below about 400 ton capacity) or spark- ignition engines using diesel blocks and moving components (for larger chillers). While the automotive-derivative engines are advertised to have a 20,000 hr useful life, the real life may be much shorter, requiring an engine replacement every 2 years or so. The diesel-derivative engines require an overhaul every 10 – 12,000 hr (equivalent to a diesel truck traveling 500,000 miles at 50 mph).Newer engines use lean burn technology to improve combustion and reduce CO and NOX emissions. By adding catalytic converters to the exhaust and additional emissions controls, natural gas fired engine drive chillers can meet stringent California air quality regulations. Gas engine-drive chillers remain more expensive than electric-drive units and they have higher overall operating costs, including maintenance costs, (see Table 1.5). However, engine-drive chillers may be used during peak cooling load periods to reduce seasonal peak electrical demand charges 

 

 

Condensing Medium

The heat collected by the water chiller, along with the excess compressor heat, must be rejected to a heat sink. Directly or indirectly, ambient atmospheric air is typically used as this heat sink. For air-cooled chillers, the condenser consists of a refrigerant-to-air coil and one or more fans to circulate outdoor air over the coil. The performance of the condenser is dependent on the airflow rate and the air’s dry bulb temperature. Air-cooled condenser air flow rates range from 600 to 1200 cfm/ton with a 10 –30F approach between the ambient dry bulb temperature and the refrigerant condensing temperature. For R-22 in a typical HVAC application, the condensing temperature is about 1058F. Thus, the ambient air temperature must be no greater than 958F. As the ambient air temperature increases, the condensing temperature increases and net cooling capacity decreases by about 2% for each 5F increase in condensing temperature. Water-cooled chillers typically use a cooling tower to reject condenser heat to the atmosphere and Chaps. 9 –17 of this text address this topic in detail. At the chiller, with 85F condenser water supplied from the cooling tower, condensing temperatures are reduced to 94 –9F, reducing the lift required of the compressor and significantly improving the chiller COP compared to air-cooled machines. Table 1.5 illustrates the relative efficiency and operating cost for the various types of electric-drive chillers with both air- and water-cooled condensing.

 ABSORPTION CHILLERS

Approximately 92% of refrigeration systems utilized for HVAC applications in the United States are electric-drive vapor compression cycle systems. However, in some areas, principally in large cities and at some universities and hospital complexes, large steam distribution systems are available. In years past, this steam was often cheaper than electricity and was used to provide cooling, utilizing the absorption refrigeration cycle. This is generally no longer the case and little or no new absorption cooling is utilized except where a waste heatsource is available, such as with cogeneration or some industrial processes, or where the use of absorption cooling during peak cooling load periods may allow areduction in seasonal electric demand charges . The absorption refrigeration cycle is relatively old technology. The concept dates from the late 1700s and the first absorption refrigeration machine was built in the 1850s. However, by World War I, the technology and use of reciprocating compressors had advanced to the point where interest in and development of absorption cooling essentially stagnated until the 1950s. During this period, the two-stage absorption refrigeration machine was developed in the United States, while the direct-fired concept was perfected in Japan and other Pacific-rim countries. The fundamental “single stage” absorption cycle is represented in Figure 1.8. The evaporator consists of a heat exchanger, held at low pressure, with a separate refrigerant (typically, water) pump. The pump sprays the refrigerant over the tubes containing the chilled water, absorbs heat from the water, and evaporates as a low-pressure gas. The low-pressure gas flows to the absorber due to the pressure differential. The absorber is at a lower pressure than the evaporator is because the concentrated absorbent solution (typically lithium bromide) exerts a molecular attraction for the refrigerant. The absorbent solution is sprayed into contact with the refrigerant vapor. Condensing of the refrigerant occurs because the heat is absorbed by absorbent. The absorbent, then, is cooled by condenser water. The absorbent now consists of a dilute solution, due to its having absorbed water vapor refrigerant. The dilute solution is pumped to the concentrator, where heat is applied to re-evaporate the refrigerant. The concentrated solution of absorbent is then returned to the absorber. The refrigerant vapor goes to the condenser, where it is condensed by the condenser water. To improve efficiency, a heat exchanger is used to preheat the dilute solution, with the heat contained in the concentrated solution of the absorbent. Leaks allow air to enter the refrigerant system, introducing noncondensable gases. These gases must be removed, or purged, to prevent pressure in the absorber increasing to the point where refrigerant flow from the evaporator will stop. The solution in the bottom of the absorber is relatively quiet and these gases tend to collect at this point. They can be removed through the use of a vacuum pump, typically called a purge pump, Absorption chillers are defined as indirect-fired or direct-fired and may besingle-stage or two-stage (and research on a three-stage chiller is currently underway), as follows:

1. The indirect-fired single-stage machine uses low- to medium-pressure steam (5 – 40 psig) to provide the heat for the absorption process. This type of chiller requires approximately 18,500 Btu/hr per ton of cooling effect, resulting in a chiller COP of about 0.67.

2. The indirect-fired two-stage chiller utilizes high-pressure steam (atleast 100 psig) or high temperature hot water (4008F or higher) and requires approximately 12,000 Btu/hr per ton of cooling effect, resulting in a chiller COP of 1.0.

3. The direct-fired chiller, as its name implies, does not use steam but utilizes a natural gas and/or fuel oil burner system to provide heat. These chillers are two-stage machines with a resulting in an overall COP of 1.0– 1.1. For the indirect-fired units, the overall COP must be reduced to account for the losses in the steam production in the boilers. With a typical boiler firing efficiency of 80 – 85%, this reduces the overall COP for the single stage system to approximately 0.54 and to approximately 0.80 for the two-stage system. Because absorption cooling has a COP of only 0.54– 1.1, it competes poorly with electric-drive rotary compressor chillers, as shown in Table 1.5. Other factors that must be considered for absorption chillers include the following:

1. Absorption chillers require approximately 50% more floor area than theequivalent electric-drive (vapor compression cycle) chiller. Additionally, due to their height, mechanical equipment rooms must be 6 –10 ft taller than rooms housing electric-drive chillers. Finally, because theliquid solution is contained in long, shallow trays within an absorption chiller, the floor must be as close to absolutely level as possible.

2. Absorption chillers will weight at least twice as much the equivalent electric-drive chiller.

3. Due to their size, absorption chillers are sometimes shipped in several sections, requiring field welding for final assembly.

4. While most electric chillers are shipped from the factory with their refrigerant charge installed, the refrigerant and absorbent (including additives) must be field installed in absorption chillers.

5. While noise and vibration are real concerns for electric-drive chillers, absorption chillers (unless direct-fired) are quiet and essentially vibration-free.

6. Due to the potential for crystallization of the lithium bromide in the chiller if it becomes too cool, the condenser water temperature must be kept above 75 –80F.

7. An emergency power source may be required if lengthy power outages are common. Without power and heat input, the chiller begins to cool and the lithium bromide solution may crystallize. However, because an absorption chiller has a very small electrical load requirement (usually less than 10 kW), a dedicated back-up generator is not a major element.

8. The heat rejection rate from the condenser is 20 –50% greater than for the equivalent electric-drive chiller, requiring higher condenser water flow rates and larger cooling towers and condenser water pumps.

9. Finally, an indirect-fired absorption chiller will be at least 50% more expensive to purchase than the equivalent electric-drive chiller. Direct- fired absorption chillers will cost almost twice as much as an electric machine, and have the added costs associated with providing combustion air and venting (stack).

 Direct-fired absorption cycle chillers should be carefully evaluated anytime an engine-drive vapor compression cycle chiller is being considered. Even though the energy cost for the absorption chiller may be higher, the increased maintenance costs associated with engine-drive systems may make the absorption chiller more cost effective.

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