Dinc Ibrahim. Refrigeration systems and applications 2th edition

Подождите немного. Документ загружается.

Refrigeration System Components 121

Figure 3.14 A centrifugal compressor (Courtesy of York International ).

refrigeration systems with low-pressure ratios and operate with adiabatic compression efﬁciencies

of up to 80%. Evaporator temperatures may reach −100

◦

Packaged water-cooled centrifugal compressors are available in sizes ranging from 85 to over

5000 tons. Larger sizes, typically 1200 to 1500 tons and larger, are shipped in subassemblies. Smaller

sizes are shipped as a factory-assembled package (Figure 3.14).

Centrifugal compressors use one or more rotating impellers to increase the refrigerant vapor

pressure from the evaporator enough to make it condense in the condenser. Unlike the positive

displacement, reciprocating, scroll or screw compressors, the centrifugal compressor uses the com-

bination of rotational speed (rpm) and tip speed to produce this pressure difference. The refrigerant

vapors from the chiller evaporator are commonly prerotated using variable inlet guide vanes. The

consequent swirling action provides extended part-load capacity and improved efﬁciency. The

vapors then enter the centrifugal compressor along the axis of rotation.

The vapor passageways in the centrifugal compressor are bounded by vanes extending from the

compressor hub, which may be shrouded for ﬂow-path efﬁciency. The combination of rotational

speed and wheel diameter combines to create the tip speed necessary to accelerate the refrigerant

vapor to the high-pressure discharge where they move on to the condenser. Due to their very high

vapor-ﬂow capacity characteristics, centrifugal compressors dominate the 200-ton level, where they

are the least costly and most efﬁcient cooling compressor design. Centrifugal forces are most

commonly driven by electric motors, but can also be driven by steam turbines and gas engines.

Depending on the manufacturer’s design, centrifugal compressors used in the packages may be one,

two, or three stages and use a semihermetic or an open motor with shaft seal.

Figure 3.15 shows a new type of centrifugal compressor which has recently been developed

by York International, providing completely oil-free compression using magnetic bearings, particu-

larly for large-tonnage refrigeration and gas compression applications. Tested extensively using

S2M magnetic bearing experience and technology, the magnetic bearing option eliminates all

the negatives of purchasing and maintaining a lubrication system. In fact, the magnetic bearings

enhance centrifugal compressor efﬁciency and operation. These compressors are available in all

sizes and options.

3.5.5.2 Turbo Compressors

In refrigeration technology, turbo compressors usually denote centrifugal compressors, but their

efﬁciencies are low. In this type of compressor, the discharge pressure is limited by the maximum

122 Refrigeration Systems and Applications

Figure 3.15 Cutaway view of a centrifugal compressor using magnetic bearings (Courtesy of York Interna-

tional).

permitted tip speed. A set of impellers is arranged for high-compression pressures. These

compressors have found applications in air conditioning and water chilling systems where high

suction volumes at high suction pressures are required.

3.5.6 Energy and Exergy Analyses of Compressors

Compressors are used to increase the pressure of a ﬂuid. In the case of a refrigeration cycle, it is

used to compress the refrigerant. Compressors operate continuously and the compression process

can be modeled as a steady-ﬂow process. Then, referring to Figure 3.16, conservation of mass

principle requires that

˙m

=˙m

−→ ρ

= ρ

−→

(3.1)

where ˙m is the mass ﬂow rate (kg/s), ρ is density (kg/m

), A is cross-sectional area (m

), V is

velocity (m/s), v is speciﬁc volume (m

/kg), and

V is the volume ﬂow rate (m

/s).

Compressor

Figure 3.16 The schematic of a compressor considered for mass and energy analysis.

Refrigeration System Components 123

A compressor involves power input

andenergyenteringandleavingbytheﬂuidstream.

The steady-ﬂow energy balance can be written as (with negligible kinetic and potential energies)

out

+˙mh

=˙mh

(3.2)

=˙m(h

− h

)

where h is enthalpy (kJ/kg). Compressors are normally not insulated and there can be heat transfer

between the ﬂuid being compressed and the surrounding air. Depending on the temperature of the

refrigerant across the compression process and the temperature of the surrounding air, the net heat

transfer could be from the compressor or to the compressor. However, the magnitude of this heat

transfer is small and it is usually neglected. Assuming that there is a net heat transfer from the

compressor, the energy balance equation becomes

+˙mh

out

+˙mh

(3.3)

−

out

=˙m(h

− h

)

Considering again an adiabatic compressor with a steady-ﬂow compression process, an entropy

balance may be written as

−

out

gen

= 

sys

= 0

gen

out

−

(3.4)

gen

=˙ms

−˙ms

=˙m(s

− s

)

Then the exergy destruction during the compression process becomes

dest

= T

gen

=˙mT

− s

) (3.5)

The exergy destruction can also be determined by writing an exergy balance on the compressor:

−

out

−

dest

= 0

dest

−

out

dest

−

− 

(3.6)

−˙m

[

− h

− T

− s

)

]

−

rev

=˙m(h

− h

) −˙m

[

− h

− T

− s

)

]

=˙mT

− s

)

where the reversible work input to the compressor is

rev

−

=˙m

[

− h

− T

− s

)

]

(3.7)

The exergy efﬁciency of the compressor may be expressed as the ratio of the reversible work to

the actual work.

Comp,ex

rev

= 1 −

dest

(3.8)

124 Refrigeration Systems and Applications

3.5.7 Compressor Capacity and Performance

All compressors are rated in terms of how much ﬂow they produce at a given ratio of outlet to inlet

pressure (compression ratio). This ﬂow is obviously a function of compressor size (e.g., the number

of cylinders and volume displacement for reciprocating compressors) and operating speed (rpm).

Compression ratio is deﬁned by the discharge pressure divided by the suction pressure (both in

absolute pressure, Pa or kPa).

The limits of clearance volumes and valve pressure differentials force some of the compressor’s

ﬂow volume capability to be lost as useful compression. This is referred to as volumetric efﬁciency.

For example, at a compression ratio of 3 to 1, 82% of the volume of the compressor is useful.

Thus, if the refrigeration effect required 10 cfm of vapor ﬂow from the evaporator, the compressor

would have to produce 10/0.82 or 12.2 cfm of ﬂow.

3.5.7.1 Compression Ratio

The compression ratio is deﬁned as the ratio of discharge pressure to suction pressure at saturated

conditions, expressed in absolute terms, for example, Pa or kPa.

CR =

(3.9)

where CR is compression ratio; P

is saturated discharge pressure, kPa; and P

is saturated suction

pressure, kPa.

The performance of a compressor is inﬂuenced by numerous parameters including the following:

• compressor speed,

• suction pressure and temperature,

• discharge pressure and temperature, and

• type of refrigerant and its ﬂow rate.

3.5.7.2 Compressor Efﬁciency

In practice, ARI Standard 500-2000 deﬁnes the compressor efﬁciency as the ratio of isentropic work

to the actual measured input power. This is also called isentropic efﬁciency or adiabatic efﬁciency.

Referring to Figure 3.16, the compressor isentropic efﬁciency becomes

Comp,isen

isen

act

˙m(h

− h

)

˙m(h

− h

)

− h

(3.10)

where ˙m is the mass ﬂow rate of refrigerant, kg/s; h

is speciﬁc enthalpy of refrigerant vapor

at discharge pressure at constant entropy (s

= s

), kJ/kg; h

and h

are speciﬁc enthalpies of

refrigerant at the inlet and exit of the compressor, kJ/kg;

isen

and

act

are isentropic and actual

powers, kW.

Note that the other compressor efﬁciency, the volumetric efﬁciency, can be approximately rep-

resented in terms of the ratio of the clearance volume to the displacement volume (R), and the

refrigerant-speciﬁc volumes at the compressor inlet (suction) and exit (discharge) (v

and v

), as

given below:

Comp,vol

= 1 − R



− 1



(3.11)

Also, note that the refrigeration capacity can be deﬁned in terms of the compressor volumetric

displacement rate (

V ,m

/s), compressor volumetric efﬁciency (η

comp,vol

), density of the refrigerant

Refrigeration System Components 125

at the compressor inlet (ρ

, kg/m

), and speciﬁc enthalpies of the refrigerant at the inlet (state 4)

and exit (state 1) of the evaporator. It is then written as

Vη

Comp,vol

− h

) (3.12)

Further details on the practical performance evaluations and ratings of compressors, and deﬁni-

tions of compressor related items, are given extensively in ARI (2000).

Although there are a number of issues that affect the compressor efﬁciency, the most signiﬁcant

one is the temperature lift (or compression ratio). To a lesser extent, the suction temperature,

lubrication, and cooling also play an important role. Therefore, the following solutions to increase

the efﬁciency of the compressor become crucial (DETR, 1999):

• Minimization of temperature lift. The compressor is most efﬁcient when the condensing pres-

sure is low and the evaporating pressure is high, leading to the minimum temperature lift and

compression ratio. The effect of operating conditions is illustrated by the compressor data example

in Figure 3.17. In conjunction with this, a good system design should ensure that the condensing

pressure is as low as possible and the evaporating temperature is as high as possible. Designing

Power input, kW

−25

−15

−5

−10

−20

Evaporating temperature, °C

−25

−15

−5

−10

−20

Evaporating temperature, °C

60 °C

55 °C

60 °C

55 °C

50 °C

45 °C

35 °C

40 °C

Condensing temperature

Capacity rating, kW

(a)

(b)

Figure 3.17 Compressor performance proﬁles at different evaporator and condenser temperatures (DETR,

1999).

126 Refrigeration Systems and Applications

a system with a small condenser and evaporator to save capital cost is always false economy.

Using a larger evaporator and condenser often means that a smaller compressor can be used and

that it always reduces running costs. The additional beneﬁt is that the compressor will be more

reliable because it does not have to work as hard and operates with lower discharge temperatures.

• Lowering suction temperature. The lower the suction gas temperature the higher the capacity

with no effect on power input. The discharge temperature will also be lower, thus increasing

reliability. Suction line insulation is essential.

• Effective lubrication and cooling. The compressor must be lubricated and efﬁciently cooled.

Insufﬁcient lubrication increases bearing friction and temperature and reduces compressor efﬁ-

ciency, often resulting in failure.

Example 3.1

Refrigerant-134a enters the compressor of a refrigeration cycle at 160 kPa and −10

◦

C with a ﬂow

rate of 0.25 m

/min and leaves at 900 kPa and 60

◦

C. The ratio of the clearance volume to the

displacement volume is 0.05. Determine (a) the volumetric efﬁciency of the compressor, (b) the

power input to the compressor, (c) the isentropic efﬁciency of the compressor, and (d) the rate of

exergy destruction and the exergy efﬁciency of the compressor. Take T

= 25

◦

Solution

(a) The properties of refrigerant at the inlet and exit states of the compressor are obtained from

R-134a tables (Tables B.3, B.4, and B.5):

= 160 kPa

=−10

◦



= 245.77 kJ/kg

= 0.9598 kJ/kg · K

= 0.1269 m

/kg

= 900 kPa

= 60

◦



= 295.13 kJ/kg

= 0.9976 kJ/kg · K

= 0.02615 m

/kg

= 900 kPa

= s

= 0.9598 kJ/kg · K



= 282.76 kJ/kg

Comp,vol

= 1 − R



− 1



= 1 − 0.05



0.1269

0.02615

− 1



= 0.807 = 80.7%

(b) The mass ﬂow rate of the refrigerant and the actual power input are

˙m =

(0.25/60) m

0.1269 m

/kg

= 0.03285 kg/s

act

=˙m(h

− h

) = (0.03285 kg/s)(295.13 − 245.77) kJ/kg = 1.621 kW

isen

=˙m(h

− h

) = (0.03285 kg/s)(282.76 − 245.77) kJ/kg = 1.215 kW

Comp,isen

isen

act

1.215 kW

1.621 kW

= 0.749 = 74.9%

Refrigeration System Components 127

(d) The exergy destruction is

dest

=˙mT

− s

) = (0.03285 kg/s)(298 K)(0.9976 − 0.9598) kJ/kg · K = 0.370 kW

The reversible power and the exergy efﬁciency are

rev

act

−

dest

= 1.621 − 0.370 = 1.251 kW

Comp,ex

rev

act

1.251 kW

1.621 kW

= 0.772 = 77.2%

3.5.7.3 Compressor Capacity Control

A capacity controlled refrigeration unit is a unit in which the compression ability of the compres-

sor can be controlled to reduce or increase refrigerant mass ﬂow rate. The concept of compressor

ﬂow modulation achieves improved performance in two ways. First, by using efﬁcient compressor

capacity reduction to prevent the increase in mass ﬂow rate of refrigerant at high ambient temper-

atures, the coefﬁcient of performance (COP) at higher ambient temperatures can be signiﬁcantly

increased. Reliability will also be increased because of reduced load on the compressor. The sec-

ond improvement in performance is realized by a change in system sizing strategy. Conventional

heat pumps are sized for the cooling load so that comfortable air conditioning is obtained. With

compressor capacity control, the heat pump can be sized for a greater heating capacity, thereby

having a lower balance point and eliminating some of the auxiliary heating. Then, via the capacity

control which is inherent in the concept, the capacity of the unit during cooling can be controlled

to achieve proper comfort control.

One method of system capacity control frequently in use today is hot gas bypass. Hot gas bypass,

where discharge gas from the compressor is vented back to the suction side of the compressor, is an

easy retroﬁt to most systems, but is disastrous from an energy savings viewpoint because capacity is

reduced without reducing compressor work, and is probably best avoided. Other possible capacity

control methods fall into essentially three categories:

• Speed control. Speed control can be done either continuously or stepwise. Continuously variable

speed control is one of the most efﬁcient methods of capacity control, and it offers good control

down to about 50% of rated speed of normal compressors. More than 50% speed reduction

is unacceptable because of lubrication requirements of the compressors. Continuously variable

speed control is also an expensive process, though not necessarily prohibitively expensive, for it

might be possible to replace some of the conventional starting controls with the motor controls

and hence reduce the cost increment. Stepwise speed control, as achieved, for example, by

using multipoled electric motors and switching the number of active poles, is another viable

alternative. It might be possible to achieve satisfactory improvements in performance by using

a ﬁnite number of stepped changes to vary compressor capacity. Step control is less costly than

continuously variable speed control, but is also limited to 50% of rated compressor speed because

of lubrication requirements. Also, step changes in load on the compressor could put high stresses

on compressor components.

• Clearance volume control. This requires substantial amounts of additional clearance volume

to achieve the amount of ﬂow reduction desirable. For example, to reduce the mass ﬂow rate

by 50%, the clearance volume must be equal to about half of the displacement volume, adding

substantially to the bulk of the compressor. Moreover, the large amount of residual mass causes

unacceptably high discharge temperatures with large amounts of ﬂow reduction. For this reason,

clearance volume control is considerably less attractive than some other types of control.

128 Refrigeration Systems and Applications

• Valve control. Suction valve unloading, a compressor capacity control method often used in

large air conditioning and refrigeration systems to reduce cooling capacity when load decreases,

can achieve some energy savings but has a number of drawbacks. In unloading, the suction

valve of one or more cylinders is held open so that gas is pumped into and out of the cylinder

through the valve without being compressed. Substantial losses can occur because of this repeated

throttling through the suction valve. In addition, stepwise cylinder unloading causes uneven

stresses on the crankshaft and provides inadequate, if not totally unacceptable, control in smaller

compressors. The method is, however, relatively inexpensive. Two newer methods of compressor

ﬂow regulation via valve control are late suction-valve closing and early suction-valve closing.

Late suction-valve closing again incurs the throttling loss by pumping gas back out of the

suction valve for part of the stroke. Late valve closing, however, gives more acceptable, smoother

control than complete valve unloading. At present, however, the method is limited to a maximum

of 50% capacity reduction and to large low-speed compressors. Early suction-valve closing

eliminates losses due to throttling gas back out of the suction valves. Instead, the suction valve,

or a secondary valve just upstream of the suction valve, is closed prematurely on the intake

stroke, limiting the amount of gas taken in. The gas inside the cylinder is expanded and then

recompressed, resulting in much lower losses. Continuously variable capacity control over a wide

range is possible with the early valve closing approach. The early suction-valve closing approach

requires the most development of the capacity control methods discussed above, but it also holds

promise for being one of the most efﬁcient and inexpensive approaches.

3.5.7.4 Capacity Control for Varying Loads to Provide Better Efﬁciency

There are several ways to meet varying loads, each with different efﬁciency, as summarized below

(DETR, 1999):

• Case 1. Single large compressor. This cannot meet variable load and results in wasted capacity

and lower efﬁciency when at part load.

• Case 2. Single large compressor with inbuilt capacity control. This is a good option to meet

variable load as long as load stays above 50%.

• Case 3. Three small compressors (two with same capacity and one with capacity control). This

allows fairly close matching to demand.

• Case 4. Three small compressors with different capacities. This is a good option to meet variable

load. The aim is to mix and match to varying load with sequence control.

• Case 5. Three compressors with parallel control. This is often used, but is not always recom-

mended due to nonlinear input power with capacity turn-down. For example, at 180% capacity

(i.e., 3 at 60%), it requires ∼240% power due to inefﬁciencies, which brings an additional input

of about 60%.

• Case 6. Three compressors (two are on and one is off). In this case, one compressor is used

at 100%, and one is used to trim to exact demand (e.g., 80% in the above case), giving 180%

capacity with 188% power (22% saving over the above case).

In the selection of one of the above cases, two main criteria are power demand and budget.

Note that the load proﬁle must be available to select the best compressor option. Different

options should be compared at the most common operating conditions as well as throughout

the load range. The efﬁciency of the different options varies enormously and there is no

hard and fast rule to selecting the best solution. Switching a compressor off to reduce the

system capacity is the most efﬁcient method of meeting a reduced load. The efﬁciency of a

compressor operating on inbuilt capacity control is always lower than when it operates at full

load. The efﬁciency of the different methods of capacity control varies. In general, any method

Refrigeration System Components 129

which recirculates compressed gas back into the suction of a compressor is very poor. When

considering compressors with capacity control, we should compare the options accurately.

Compressors are often oversized for an application because so many safety factors are used when

calculating the load. This should be avoided as oversized compressors often operate with a lower

power factor.

Regardless of the conﬁguration option selected to meet a load, the control of the compressors is

important. The control strategy should be designed to

• select the most efﬁcient mix of compressors to meet the load,

• avoid operation on inbuilt capacity control when possible, and

• avoid operation at low suction pressures when possible.

Selecting compressors of different sizes and designing a good control strategy to cycle them to

accurately match the most common loads is often the most efﬁcient option.

3.6 Condensers

There are several condensers to be considered when making a selection for installation. They are air-

cooled, water-cooled, shell and tube, shell and coil, tube within a tube, and evaporative condensers.

Each type of condenser has its own unique application. Some determining factors include the size

and the weight of the unit, weather conditions, location (city or rural), availability of electricity,

and availability of water.

A wide variety of condenser conﬁgurations are employed in the process industry. Selection of

condenser type is not easy and depends on the following criteria:

• condenser heat capacity,

• condensing temperature and pressure,

• the ﬂow rates of refrigerant and coolant,

• design temperature for water and/or air,

• operation period, and

• climatic conditions.

Condensers utilized in the refrigeration industry are commonly of three types, as follows:

• water-cooled condensers,

• air-cooled condensers, and

• evaporative condensers.

Common types of water- and air-cooled refrigerant condensers for commercial refrigeration

use are

• shell and tube, blow-through, horizontal airﬂow,

• shell and coil, draw-through, vertical airﬂow, and

• tube in tube, static, or forced airﬂow.

The type of condenser selected depends largely on the following considerations:

• size of the cooling load,

• refrigeration used,

• quality and temperature of available cooling water (if any), and

• amount of water that can be circulated, if water use is acceptable.

130 Refrigeration Systems and Applications

ELT

tube in tube

SST

shell and tube

VSE

shell and coil

AMC

ammonia

application

SCH/SCS

coaxial

Figure 3.18 Various water-cooled condensers (Courtesy of Standard Refrigeration Company ).

3.6.1 Water-Cooled Condensers

Water-cooled condensers are of many different types as shown in Figure 3.18. The most common

condensers are generally shell and tube type heat exchangers with refrigerant ﬂow through the

shell and water (as coolant) ﬂow through the tubes (SST type in Figure 3.18). The lower portion

of the shell acts as a liquid receiver. These condensers are widely used in large heat capacity

refrigerating and chilling applications. If a water-cooled condenser is used, the following criteria

must be examined:

• requirement of cooling water for heat rejection,

• utilization of a cooling tower if inexpensive cooling water is available,

• requirement of auxiliary pumps and piping for recirculating cooling water,

• requirement of water treatment in water recirculation systems,

• space requirements,

• maintenance and service situations, and

• provision of freeze protection substances and tools for winter operation.

In general, water-cooled condensers are used with cooling towers or groundwater (well, lake,

river, etc.).

3.6.2 Air-Cooled Condensers

The air-cooled condensers ﬁnd applications in domestic, commercial, and industrial refrigerat-

ing, chilling, freezing, and air-conditioning systems with a common capacity of 20−120 tons

(Figure 3.19). The centrifugal fan air-cooled condensers (with a capacity of 3−100 tons) are par-

ticularly used for heat recovery and auxiliary ventilation applications. In fact, they employ outside

air as the cooling medium. Fans draw air past the refrigerant coil and the latent heat of the refrig-

erant is removed as sensible heat by the air stream. The advantages of air-cooled condensers are

the following:

• no water requirement,

• standard outdoor installation,

• elimination of freezing, scaling, and corrosion problems,

• elimination of water piping, circulation pumps, and water treatment,

• low installation cost, and

• low maintenance and service requirement.