Diesel fuel injection systems—Lucas Diesel Systems

 Diesel fuel injection systems—Lucas Diesel Systems

 

1 Compression ignition combustion processes

 Introduction of fuel In indirect injection engines, the fuel is introduced via pintle nozzles into an anti-chamber to the space above the piston, which is called a 'prechamber' or a 'swirl chamber', depending upon the intensity of rotary air motion in this chamber. A swirl chamber layout is shown below in Figure 11.1 together with a much magnified section through a pintle nozzle. In IDI engines, much of the energy to mix the fuel with the air comes from the air motion, including swirl before ignition and combustion- driven mixing in the swirl-chamber throat and above the piston after ignition.

When fuel pressure is applied to the differential area between the guide and the seat, the needle lifts when the force developed exceeds the preload in the spring that holds the needle valve closed. If the pressure is applied progressively, an 'obturator' in the nozzle hole restricts flow during the first part of the valve. This controls the initial rate of injection to reduce noise. At the bottom of the needle is a 'pintle' which forms the spray into a hollow cone Indirect injection (IDI) engines are being replaced progressively by direct injection (DI) engines, from the larger sizes downwards, and all European truck and larger engines have been direct injection diesel engines for some time. In most of these combustion systems, more of the energy to mix the fuel with the air comes from the momentum imparted to the fuel as it leaves the nozzle. In small high speed direct injection engines for van and passenger car applications, the combustion chamber has a high swirl re-entrant bowl (also known as having a 'squish lip') to promote turbulence and hence faster mixing towards the end of combustion. In the intermediate medium truck (1 litre/cylinder) size, both quiescent and swirl-assisted combustion have their champions. The quiescent chamber requires more nozzle holes as shown in Figure 11.2.

 The arrangement of the hole-type injector in a swirl-assisted 2-valve HSDI combustion chamber is shown in Figure 11.3. A cross section of the complete injector is shown in Figure 11.4. The piston is close to the top of its travel, and autoignition can occur, between 20° crank angle before top dead centre (btdc) to 40° after top dead centre (atdc), approximately, with 18:1 compression ratio. At 2400 rev/min this represents only 4.2 ms  and in the case of passenger car engines rotating at up to 4500 rev/min, this is only 2.2 ms.

 

 Sprays

 

The literature on spray formation is very extensive and covers over 20 years of intensive developmen of the spray combustion  processes. During this period the fuel injection equipment has changed considerably. Early papers reflect the performance of injection systems which provided pressures across the nozzle holes that rarely exceeded 500 bars. Some of the injected fuel may be spread along the wall by its own momentum and that of the air swirl. The careful observations by many academic and industrial researchers provided empirical relationships  and insight into the break-up of liquid sprays into ligaments, droplets and daughter droplets. Several alternative models of the various combustion processes were developed.

If the injector nozzle holes are reduced in diameter, and the fuel pressure across them is increased to obtain the same penetration in the compressed air charge, several workers found that the soot and paniculate matter in the exhaust were reduced

. In North America, high injection pressures have been common for many years in injection systems for truck engines which have been equipped with shallow quiescent combustion chambers. More recent work to apply mixing models to guide further reduction in soot generation has shown that these chambers respond well to further improvements in atomization.

. In addition, the large surface area of the fuel droplets expedites mixing and evaporation. In the 199Os injection pressures in European fuel injection systems rose to over 1000 bars at the nozzle and the upward trend continues through 2000 bars in the late 1990s to reduce the mass of particulate matter in the exhaust.

A typical diesel injection spray leaves each nozzle hole with a narrow included angle, and develops a head vortex where spray momentum is transferred to the compressed air. Each successive element of fuel seems to pass through the head vortex of the previous element, to form a new head vortex further across the bowl until the combustion chamber wall is reached, or the injection is terminated. The spray entrains air as it moves through the air in the combustion chamber6. The air entrainment and mixing models, some of which have been extensively validated against experimental results, show that when fuel is introduced into the combustion chamber as finely atomized sprays, the air entrainment increases providing a mixture which is closer to stoichiometric near the centre of each spray. The generation of diesel sprays by the injector nozzle has been studied with large-scale models. Above a critical pressure  ratio, cavitation in the nozzle hole occurs which finely divides the fuel before it leaves the hole. In consequence the spray angle is larger and more air is entrained into the spray.

 U Ingnition

Compression ignition combustion occurs only after a delay of approximately 0.0002 to 0.002 seconds after the start of injection.

This delay occurs because:

• The fuel has to travel into the combustion chamber.

• It has to mix with the air sufficiently to form a near-stoichiometric mixture.

• The fuel has to evaporate by taking heat from the compressed air.

• The mixture has to heat up to the auto-ignition temperature (Figure 11.5).

• Certain chemical reactions must take place in which unstable hydrocarbon-oxygenate species form which will ignite spontaneously.

 Pre-mixed burn

Fuel which has been injected into the combustion chamber leaves the nozzle at 150 to 500 m/s, so the 20 to 30 mm radius of an HSDI combustion chamber can be traversed by a liquid jet in 0.04 to 0.2 ms, and somewhat longer by the evaporating droplets in a finely atomized spray tip. A significant proportion of the fuel injected during the ignition delay period will have mixed with the air in the combustion chamber when the first element ignites spontaneously (autoignition). Thus virtually all of the fuel which is injected in the ignition delay period (less between one and two crank degrees depending upon the injection pressure and the engine speed) is consumed in the premixed burn, unless the injection is specifically configured to avoid this (for example the M-combustion system16.

The premixed burn provides the rapid initial heat release which is typical of unrefined direct injection combustion systems, and causes a rapid increase in cylinder pressure that is the origin of combustion noise and in extreme cases piston failures. Fuel injection system developments such as pilot injection (pilot)   and initial rate controls (IRC) have evolved to control the initial rate of heat release .

 Diffusion burn

Fuel that is injected after ignition meets very different temperature and pressure conditions to that injected before ignition. The conditions may exceed the critical temperature and pressure so that the fuel flashes into vapour as it gains heat. The temperature in the combustion chamber is sufficient to pyrolize fuel if insufficient oxygen is available to burn it, especially near the centre-line of a spray. The rate of consumption of the fuel governed mainly by the rate at which it is mixed with sufficient oxygen and the exhaut emissions reflect this dependence. When the piston moves down from top dead centre, the vortices generated by the movement of gas force the burning mixture over the lip of the combustion chamber (reverse squish). If the lip is provided with a sharp edge, the turbulence caused will promote even more intimate mixing between fuel and air and will reduce much of the soot formed previously. This is very beneficial for high speed direct injection engines; but calculation of the remaining soot is not easy.

Formation of nitric oxide by lean combustion

Nitrogen and oxygen will combine together to form nitric oxide when heated to temperatures above 1500 K, primarily by the Zeldovich mechanism, although several other chemical reactions have been mentioned in the literature. Under steady state conditions, the rate of formation of nitric oxide increase rapidly with temperature as shown in Figure 11.6. The abscissa of Figure 11.6 is the equivalence ratio of the mixture which is unity when the mixture is stoichiometric. When the equivalence ratio is greater than 1, very little nitric oxide is formed, as the fuel consumes all the oxygen available at the flame front. As the mixture becomes leaner, increasing proportions of nitric oxide are formed in the mixture. If the temperature is increased at an equivalence ratio of say 0.8, the initial step from 1500 K to 2000 K results in only 0.3% increase in nitric oxide formation as a result of a new equilibrium between O2, N2 and NO. The next 300 K leads to a similar 0.3% increase in nitric oxide formation. However, the increase of 300 K between the top two curves yields over 0.5% increase in the equilibrium value of nitric oxide.

 Two courses of action are available to reduce nitric oxide formation:

(1) Reduction of the volume of lean mixture in the combustion process.

(2) Reduction of the peak local temperatures of lean burn combustion.

From these two basic options, a number of alternative treatments have emerged which are effective in reducing nitric oxide formation:

(1) Retarding the injection timing, to reduce the peak cycle temperature. (However the thermodynamic efficiency, and hence the fuel efficiency suffers.)

(2) Recirculating some exhaust gas in controlled proportions to dilute the oxygen available at part load conditions and slow the chemical reaction rate. If the recirculated exhaust gas is cooled, the temperature of combustion is reduced.

(3) Injecting water through the same spray nozzle as the fuel (the evaporation of the water reduces the peak cycle temperature).

(4) Reducing the oxide of nitrogen with catalytic combustion of extra fuel or ammonia injected into the exhaust system. The fuel injection system and electronic control unit are central to all these treatments; either directly, or indirectly since even the EGR and aftertreatments require a precise measurement of the fuelling to be effective without making other emissions worse. 

Unburned hydrocarbons

A compression ignition engine emits far less unburned hydrocarbons at normal operating temperatures than intake- and port-injected spark ignition engines, simply because less fuel comes into contact with the film of lubricating oil in compression ignition engines. However, there are about six independent sources of unburned hydrocarbons in a diesel engine18. Of these, four are directly controlled by the fuel injection equipment:

(1) If fuel is injected at high velocity into the combustion chamber, before 20° crank btdc, the air motion mixes the fuel so effectively before ignition so that a mixture forms that becomes too lean to auto-ignite and too lean to sustain a flame. This is the 'lean-limit source' described by Greevesetal.

(2) Hydrocarbons trapped in the sac and holes downstream of the needle seat join the burning mixture late in the cycle, when air motion draws this fuel out19. Injector nozzles with smaller sacs or with valve-covers-orifice (VCO) confi- gurations reduce this volume as shown in Figure 11.7.

(3) Fuel that is injected late in the engine cycle, will find little oxygen in which to burn. As the piston descends, cylinder temperatures drop below the auto-ignition temperature. 

(4) Fuel that is sprayed onto the combustion chamber walls, can conribute to unburned hydrocarbon emissions where the quantity of fuel on the wall exceeds the capacity of the air motion to evaporate it at some particular operating conditions or temperatures. Excess smoke may arise from such conditions also.

Origins of noise in diesel combustion processes 

The rapid consumption of most of the fuel injected during the ignition delay period in the pre-mixed burn release heat very rapidly, and the cylinder pressure rises almost instantaneously. This imparts a large and steep-fronted force pulse to the sructure which excites most of the mechanical resonances and causes the surfaces of the engine to vibrate20. Acoustic radiation from the vibrating surfaces completes the transmission to the ear. The sound generated by traditional diesel combustion has a characteristic 'knock'; however the structure responds in the same way to mechanical impacts as the pistons move in their bores and in the timing drive. Subjectively the noise sounds very similar to combustion knock if it has a mechanical origin. Much of the literature suggests that combustion noise depends almost entirely upon the peak rate of pressure rise caused in turn by an initial peak in the rate of heat release. For individual engines, quite good relationships exist between peak rate of rise in cylinder pressure and the noise that originates directly in the combustion processes. However, when such relationships for several different engines are compared, large discrepancies appear

The shape of the cylinder pressure curve can be related to the Fourier Analysis (or spectrum) of the cylinder pressure quite simply:

(1) The compression ratio and turbocharger boost ratio directly influence the level of components up to 500 Hz (as well as indirectly influencing the peak rate of pressure rise via the compression temperature and hence the ignition delay).

(2) The peak cylinder pressure influences the average level of low frequency spectrum components up to between 500 and 900 Hz.

(3) The peak rate of pressure rise influences the components between 500/900 Hz and 3/5 kHz depending upon the engine speed, turbocharger boost (if any) and the rate of injection diagram.

(4) The second derivative of pressure with respect to time can influence very high frequency components (above 5 kHz). The first option to reduce combustion noise is to reduce the ignition delay and hence the quantity of fuel that contributes to the premixed burn. Increasing the compression ratio, heating the intake air, turbocharging or supercharging, fumigating and reducing heat transfer into the piston and cylinder head may be deployed to this end.

Pilot injection reduces the fuel injected during the ignition delay, and hence the heat released from the premixed burn.

Control of the initial rate of injection (IRC) and 'boot-shaped' injection rate diagrams are effective for the same reasons, but only if the control extend over most of the ignition delay; hence pilot is more appropriate for cool combustion conditions encountered in urban traffic and cold start/cold idle. Models of fuel injection systems which reduce combustion noise have been used to explore the design freedoms in such systems.

Exhaust gas recirculation will reduce the rate of the premixed burn and hence the peak rate of rise in the cylinder pressure. The modulation of the initial rate of fuel injection, either by pilot or IRC or combinations of these has become an area where the fuel injection equipment manufacturers can add value in terms of refinement in passenger car and public service vehicle applications.

 Particulate emissions

Particulate emissions from compression ignition engines are mostly microscopic pieces of carbon, bound together with unburned hydrocarbons into particles which range in size from a few hundredths of a micron to over ten microns. Sulphate particles and particles formed from any metals in the fuel may add to the fine particulate emissions. The fuel injection equipment has a controlling influence upon the generation of soot-based particulates via the air entrainment in the sprays. If more oxygen can be introduced into the centre of the sprays by increasing atomization and spray velocity and hence air entrainment, less soot particles form in the centre of the sprays. This is the basis of soot and particulate reduction by reducing the nozzle hole diameter and increasing injection pressure. If the combustion system is designed to work with appreciable wall wetting, then the injection equipment is required to control the proportion of the fuel that is deposited upon the wall.

Soot reduction during subsequent combustion

As the piston of a HSDI engine descends, the burning mass of fuel and air is forced over the rim of the chamber, which generates considerable turbulence and mixes the remaining unburned hydrocarbons with the air above the piston crown. Incandescent particles of carbon are bought into contact with the remaining oxygen and up to 95% of them are burned completely. The detail of the complex motion involving swirl and squish giving rise to a moving toroid of air, into which the fuel is injected; followed by the expansion and spilling of this burning mass of air and fuel over a lip into the space above the piston as the piston starts to descend is difficult to model accurately. Further- more, as the soot in the exhaust is the difference between that generated during the earlier parts of the diffusion burn, and consumed in almost equal amounts by later diffusion burn processes, its computation involves the subtraction of two large quantities to predict a small difference. Research with a variety of mechanical fuel injection equipment devices indicates that the main injection must terminate as abruptly as is mechanically feasible, as well as meeting certain injection pressure criteria in order to minimize soot generation.

 Traditional jerk pump

The traditional jerk pump was developed to give the prime example of a variable-delivery, positive displacement hydraulic

pump. Figure 11.8 shows the components of a traditional pump- pipe-nozzle system.

The chamber above the plunger within the barrel, in which high pressure is generated by upward movement of the plunger, is connected by a central drilling to a helical or angled groove cut into the side of the plunger. There are two drillings in the barrel wall, one to allow fuel to flow into the chamber when the plunger is withdrawn, which is known as the 'filling port' and a second drilling which is known as the 'spill port'. In operation, the plunger is withdrawn down the barrel, opening the filling port, allowing fuel to enter the chamber above the plunger from a low pressure supply. As the camshaft rotates and the cam bears upon the roller of the tappet, the plunger is driven upwards until it seals the filling port. The fuel trapped above it is pressurized rapidly, and the pressure opens the delivery valve. When the delivery valve opens, the rapid motion of the plunger creates a high pressure hydraulic wave in the injector pipe. The high pressure hydraulic wave travels along the pipe until it meets the seat of the needle valve in the injector. While the needle valve is closed, the wave is reflected, causing an additional pressure at the seat of the needle valve from the combined pressure of the incident and reflected waves. When the total hydraulic pressure on the differential area around the valve seat exceeds the nozzle opening pressure (NOP), the needle valve in the injector rises off its seat. The needle opens rapidly when the whole area of the needle is exposed to this pressure. Subsequent injection ensures that the valve remains open, held at its lift stop, despite any temporary fall in pressure caused by the needle displacement. Inside the injection pump, the plunger continues to travel upward until the helical groove in the plunger engages with the spill port. When this occurs, fuel flows through the central drilling to the helical groove and out through the spill port causing the fuel pressure in the chamber to collapse. This terminates the effective stroke of the pump and generates an expansion wave in the injector pipe. When this expansion wave reaches the drilling within the injector, it causes the pressure under the needle at the seat to collapse below the nozzle closing pressure. (NCP is the pressure which acts on the whole needle area to provide a force equal to that of the spring.) The spring above the needle then begins to accelerate the needle towards its seat with a force that is equal to the spring force minus the force due to any remaining pressure acting upon the bottom of the needle. The needle displacement is a modified sinusoidal movement as the spring relaxes and drives the needle onto its seat. As the expansion wave passes through the delivery valve, its spring begins to close the valve and the collar on the valve (Figure 11.9} enters the valve guide, thus separating fuel within the injection pump from fuel within the pumping chamber. As the delivery valve continues to close, the displacement of the collar multiplied by the area of the guide 'unloads' the system downstream to control the residual pressure within the pipe and hence reduces the risk of secondary injections (injections after the main injection which cannot burn completely as they appear near the end of the combustion process). Jerk pump operation is reliable and repeatable. The plunger and barrel are made to tight manufacturing tolerances with a very small clearance between them to contain the high injection pressures inside the pumping chamber. The rate of spill controls the collapse of line pressure and hence the rate of the expansion wave that terminates injection. This can be improved by machining the spill port to be oval or even spark eroding it to a parallelogram section to match the helix angle, in order to increase the rate of area increase as the port is opened.

 These pumps are used on large truck engines with separate control of the timing of the start and end of injection. The cam rate, plunger diameter and nozzle characteristics are chosen to control soot generation by injecting at high pressures through small nozzle holes with an injection rate which increases as the injection timing is retarded.

 

CLASSIFICATION OF AIR CONDITIONING SYSTEMS ACCORDING TO CONSTRUCTION AND OPERATING CHARACTERISTICS

 CLASSIFICATION OF AIR CONDITIONING SYSTEMS ACCORDING TO CONSTRUCTION AND OPERATING CHARACTERISTICS 

Air conditioning systems can also be classified according to their construction and operating characteristics as follows.

Individual Room Air Conditioning Systems

Individual room, or simply individual air conditioning systems employ a single, self-contained

room air conditioner, a packaged terminal, a separated indoor-outdoor split unit, or a heat pump. A heat pump extracts heat from a heat source and rejects heat to air or water at a higher temperature for heating. Unlike other systems, these systems normally use a totally independent unit or units in each room. Individual air conditioning systems can be classified into two categories:

● Room air conditioner (window-mounted)

● Packaged terminal air conditioner (PTAC), installed in a sleeve through the outside wall

The major components in a factory-assembled and ready-for-use room air conditioner include the following: An evaporator fan pressurizes and supplies the conditioned air to the space. In tube-

and-fin coil, the refrigerant evaporates, expands directly inside the tubes, and absorbs the heat energy from the ambient air during the cooling season; it is called a direct expansion (DX) coil. When the hot refrigerant releases heat energy to the conditioned space during the heating season, it acts as a heat pump. An air filter removes airborne particulates. A compressor compresses the refrigerant

from a lower evaporating pressure to a higher condensing pressure. A condenser liquefies refrigerant from hot gas to liquid and rejects heat through a coil and a condenser fan. A temperature control

system senses the space air temperature (sensor) and starts or stops the compressor to control its cooling and heating capacity through a thermostat .

The difference between a room air conditioner and a room heat pump, and a packaged terminal air conditioner and a packaged terminal heat pump, is that a four-way reversing valve is added to all

room heat pumps. Sometimes room air conditioners are separated into two split units: an outdoor condensing unit with compressor and condenser, and an indoor air handler in order to have the air

handler in a more advantageous location and to reduce the compressor noise indoors.

Individual air conditioning systems are characterized by the use of a DX coil for a single room.

This is the simplest and most direct way of cooling the air. Most of the individual systems do not employ connecting ductwork. Outdoor air is introduced through an opening or through a small air

damper. Individual systems are usually used only for the perimeter zone of the building. 

Evaporative-Cooling Air Conditioning Systems

Evaporative-cooling air conditioning systems use the cooling effect of the evaporation of liquid water to cool an airstream directly or indirectly. It could be a factory-assembled packaged unit or a

field-built system. When an evaporative cooler provides only a portion of the cooling effect, then it becomes a component of a central hydronic or a packaged unit system.

An evaporative-cooling system consists of an intake chamber, filter(s), supply fan, direct-contact or indirect-contact heat exchanger, exhaust fan, water sprays, recirculating water pump, and water

sump. Evaporative-cooling systems are characterized by low energy use compared with refrigeration cooling. They produce cool and humid air and are widely used in southwest arid areas in the

United States

Desiccant-Based Air Conditioning Systems

A desiccant-based air conditioning system is a system in which latent cooling is performed by desiccant dehumidification and sensible cooling by evaporative cooling or refrigeration. Thus, a

considerable part of expensive vapor compression refrigeration is replaced by inexpensive evaporative cooling. A desiccant-based air conditioning system is usually a hybrid system of dehumidifica-

tion, evaporative cooling, refrigeration, and regeneration of desiccant.

There are two airstreams in a desiccant-based air conditioning system: a process airstream and a regenerative airstream. Process air can be all outdoor air or a mixture of outdoor and recirculatingair. Process air is also conditioned air supplied directly to the conditioned space or enclosed manufacturing process, or to the air-handlin  unit (AHU), packaged unit (PU), or terminal for further treatment. Regenerative airstream is a high-temperature airstream used to reactivate the desiccant.

A desiccant-based air conditioned system consists of the following components: rotary desiccant dehumidifiers, heat pipe heat exchangers, direct or indirect evaporative coolers, DX coils and vapor compression unit or water cooling coils and chillers, fans, pumps, filters, controls, ducts, and piping. 

Thermal Storage Air Conditioning Systems

In a thermal storage air conditioning system or simply thermal storage system, the electricity-driven refrigeration compressors are operated during off-peak hours. Stored chilled water or stored ice in tanks is used to provide cooling in buildings during peak hours when high electric demand charges and electric energy rates are in effect. A thermal storage system reduces high electric demand for HVAC&R and partially or fully shifts the high electric energy rates from peak hours to off-peak hours. 

A thermal storage air conditioning system is always a central air conditioning system using chilled water as the cooling medium. In addition to the air, water, and refrigeration control systems, there are chilled-water tanks or ice storage tanks, storage circulating pumps, and controls.

Clean-Room Air Conditioning Systems

Clean-room or clean-space air conditioning systems serve spaces where there is a need for critical control of particulates, temperature, relative humidity, ventilation, noise, vibration, and space pressurization. In a clean-space air conditioning system, the quality of indoor environmental control directly affects the quality of the products produced in the clean space.

A clean-space air conditioning system consists of a recirculating air unit and a makeup air unit—both include dampers, prefilters, coils, fans, high-efficiency particulate air (HEPA) filters, ductwork, piping work, pumps, refrigeration systems, and related controls except for a humidifier in the makeup unit.

Space Conditioning Air Conditioning Systems

Space conditioning air conditioning systems are also called space air conditioning systems. They have cooling, dehumidification, heating, and filtration performed predominately by fan coils, water-source heat pumps, or other devices within or above the conditioned space, or very near it. A fan coil consists of a small fan and a coil. A water-source heat pump usually consists of a fan, a finned coil to condition the air, and a water coil to reject heat to a water loop during cooling, or to extract heat from the same water loop during heating. Single or multiple fan coils are always used to serve a single conditioned room. Usually, a small console water-source heat pump is used for each control zone in the perimeter zone of a building, and a large water-source heat pump may serve several rooms with ducts in the core of the building.

Space air conditioning systems normally have only short supply ducts within the conditioned space, and there are no return ducts except the large core water-source heat pumps. The pressure drop required for the recirculation of conditioned space air is often equal to or less than 0.6 in. water column (WC) (150 Pa). Most of the energy needed to transport return and recirculating air is saved in a space air conditioning system, compared to a unitary packaged or a central hydronic air conditioning system. Space air conditioning systems are usually employed with a dedicated (separate) outdoor ventilation air system to provide outdoor air for the occupants in the conditioned space.

Space air conditioning systems often have comparatively higher noise level and need more periodic maintenance inside the conditioned space.

 Unitary Packaged Air Conditioning Systems

Unitary packaged air conditioning systems can be called, in brief, packaged air conditioning systems or packaged systems. These systems employ either a single, self-contained packaged unit or two split units. A single packaged unit contains fans, filters, DX coils, compressors, condensers, and other accessories. In the split system, the indoor air handler comprises controls and the air system, containing mainly fans, filters, and DX coils; and the outdoor condensing unit is the refrigeration system, composed of compressors and condensers. Rooftop packaged systems are most widely

used.

Packaged air conditioning systems can be used to serve either a single room or multiple rooms.

A supply duct is often installed for the distribution of conditioned air, and a DX coil is used to cool it. Other components can be added to these systems for operation of a heat pump system; i.e., a centralized system is used to reject heat during the cooling season and to condense heat for heatingduring the heating season. Sometimes perimeter baseboard heaters or unit heaters are added as a part of a unitary packaged system to provide heating required in the perimeter zone.

Packaged air conditioning systems that employ large unitary packaged units are central systems by nature because of the centralized air distributing ductwork or centralized heat rejection systems.

Packaged air conditioning systems are characterized by the use of integrated, factory-assembled, and ready-to-use packaged units as the primary equipment as well as DX coils for cooling, compared to chilled water in central hydronic air conditioning systems. Modern large rooftop packaged units have many complicated components and controls which can perform similar functions to the central hydronic systems in many applications.

 

المواصفات الواجب توافرها بمجمعات وخطوط أنابيب سحب الطلمبات

المواصفات الواجب توافرها بمجمعات وخطوط أنابيب سحب الطلمبات

فى مجمعات وخطوط سحب الطلمبات يراعى تقليل الدوامات والجيوب الغازية لتقليل فاقدد الضغط وبالتالى تقليل 

وخصوصا العوامل المؤثرة  على حدوث ظاهرة التكهف Cavitation فى الطلمبات ذات الضغط المنخفض

Pumps Head-Low( التى تسحب السائل من خزان والتى يطلق علي أحيانا  الطلمبات المناولة

ولذلك توجد بع المواصفات المقررة والواجب تحقيقها بمجمعات وخطوط سحب الطلمبات وهي: 

1 -يراعى إختيار قطر خط السحب الرئيسى )خط السحب الرئيسى المجمع الذى يصب فيه فروع أنابيب السحب من الخزانات والممتد من أمام منطقة المستودعات حتى ترنش الخطوط أمام عنبر الطلمبات( بحيث تكون سرعة سريان السائل داخله حوالى 9,0 متر/ثانية أى يكون قطر خط السحب الرئيسى )بوصة( يساوى:

((di = √(Q/(1.824×0.9

 Q: مجموع معدلات الطلمبات المناولة التى تسحب فى آن واحد من خط السحب العمومى ) 3متر/ ساعة

* يراعى أن يكون منسوب الراسم العلوى لخط السحب الرئيسى أقل من منسوب قاع الخزان

2 -يراعى إختيار قطر خط سحب الطلمبة )خط السحب من فلانشة سحب الطلمبة حتى خط السحب الرئيسى 

بترنش الخطوط أمام عنبر الطلمبات( بحيث ألا تزيد سرعة السائل داخله عن حوالى 5,1 متر/ثانية

أى يكون قطر خط سحب الطلمبة )بوصة( يساوى

((di = √(Q/(1.824×1.5

 Q : معدل سحب الطلمبة الواحدة من خط السحب الرئيسى 3 متر/ ساعة(

3 -إذاكانت سرعة السائل عند فلانشة سحب الطلمبة لا تزيد عن 5,1 متر/ثانية  يفضل أن يكون قطر خط سحب الطلمبة يساوى قطر فلانشة سحب الطلمبة 

* يجب عدم تخفيض قطر خط السحب من الطلمبة حتى الوصول إلى خط السحب الرئيسى بترنش الخطوط 

أمام عنبر الطلمبات 

4 -إذا كانت سرعة السائل عند فلانشة سحب الطلمبة أكبر من 5,1 متر/ثانية يفضل  أن يكون قطر خط سحب الطلمبة أكبر من قطر فلانشة سحب الطلمبة بالقيمة التى تحافظ على سرعة سريان السائل بخط سحب الطلمبة فى حدود 5,1 متر/ثانية وينبغى إستخدام مساليب الامركزية Reducer Eccentric كما يتضح من الشكل التالى و ذلك لعدم تكون جيوب غازية بالجزء أعلى المسلوب 

* يراعى ألا يزيد الفرق بدين قطر خط سدحب الطلمبة وقطر فلانشة سحب الطلمبة عن 4 بوصة و ذلك لتقليل 

الدوامات والجيوب الغازية عند فلانشة سحب الطلمبة

اذا كان مصدر السحب Supply of Source أعلى الطلمبة يتم تركيب مسلوب الامركزى بحيث يكون الجانب المستقيم )العدل( من المسلوب لاسفل (FOB (Bottom On Flat كما يتضح من الشكل التالى

اذا كان مصدر السحب Supply of Source أسغل الطلمبة يتم تركيب مسلوب الامركزى بحيث يكون 

الجانب المستقيم )العدل( من المسلوب لاعلى (FOT (Top On Flat كما يتضح من الشكل التالى

 - يفضل تركيب قيعان من نوع ) نصف قطر التقوس الطويل long Radius Elbow  والذى يساوى ١.٥

قطر الخط( وذلك بخط سحب الطلمبة 

6 -إذا كان السحب فى الاتجاه الرأسى Top Suction   يراعى أن يكون الطول المستقيم لخط السحب Straight  Run( من فلانشة سحب الطلمبة حتى بداية القوع90  long Radius Elbow  لا يقل عن )3 أمثال قطر خط السحب( ما يتضح من الشكل التالى

فى حالة تركيب قوع من نوع ) نص قطر التقوس القصير   short Radius Elbow  والذى يساوى قطر الخط( أو تيه بخط السحب يراعى أن يكون الطول المستقيم لخط السحب  Straight Run قبل فلانشة سحب الطلمبة الايقل عن )8 أمثال قطر خط السحب(

8 -فى الطلمبات المزدوجة السحب Pumps Suction Double يراعى عدم تركيب قدوع أفقى )فى مستوى 

مطابق أو موازى لعمود الطلمبة( بفلانشة سحب الطلمبة كما يتضح من الشكل التالى

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.

• 
  

 

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.