GUIDELINE

Screw Compressors Testing

Contents

1. Principle of operation and classification of the compressors .............................................. 2

2. Basic parameters of the compression process. Nature of the compression ......................... 4

3. Screw compressors .............................................................................................................. 6

3.1. Construction ................................................................................................................. 6

3.2. Fundamentals of operation........................................................................................... 7

3.3. Volume Ratio .............................................................................................................. 11

3.4. Capacity Control ......................................................................................................... 13

4. Methods for testing of screw compressors ......................................................................... 17

4.1. Testing of pressure distribution in twin screw compressors for multiphase

duties ................................................................................................................................. 17

4.2. Testing of screw refrigeration compressors under superfeed conditions ................... 19

4.3 Testing of gas screw compressors ( for natural gas or other types of gas ) ................. 23

5. Conclusion .......................................................................................................................... 26

6. Bibliography ....................................................................................................................... 27

1. Principle of operation and classification of the compressors

Compressors are operating pneumatic machines that increasing the potential energy,

the pressure of gas or steam by converting them held to the shaft mechanical energy drive

motor.

All compressors according to the principle of action, which means according to

physical phenomena, which are used to transfer energy to gas can be divided into three

groups:

1) volume compressors ( positive displacement compressors);

2) dynamic compressors;

3) thermal compressors.

In some types of compressors combinations of several ways to increase the pressure

have been used. The volume compressors increase the gas pressure by reducing the

volume of closed chambers containing a certain amount of gas, which means definite

number of molecules of the gas. Reducing the chamber increases the concentration of

molecules in unit volume. The pressure of the ambient gas on the walls according to the

laws of kinetic energy of the gas is proportional to the total (aggregate) impact energy of

the gas molecules on the wall. Therefore, when reducing the volume and increasing the

gas molecules per unit volume increases the number of strokes of molecules per unit area,

resulting in increased gas pressure. The process of increasing the pressure in the volume

compressors is periodic. It is the first volume of the working chamber to increase and be

filled with gas, then reduce, increase gas pressure and tight (compressed) kind to be

forced from the compressor.

The pneumatic volume machines are rotational and non rotational. Typical

representative of non rotational compressors is the piston compressor. It features a large

economy, simple construction, maintenance, repair and reliability. Operating cycle of the

piston compressor consists of three phases: suction, compression of gas from the cylinder

and ejection from the chamber in the compression line.

In rotary compressors (lammelate, screw, eight figurative, at water ring, vortex)

working cycle occurs continuously rolling the working cameras with rotation of the rotor.

Compressors can be classified more of the following signs:

a) depending from the way of use - general and special (refrigeration, etc..);

b) a type - landline or mobile;

c) the type of compressible fluid - air, gas and steam;

d) according to the drive - electrically adjustable, with internal combustion engines,

steam and gas turbine;

e) according to the cooling - air and water cooled;

f) according to the stages of pressure - single stage and multistage;

g) according to the size of the flow - low flow (to 10 m3/min), with an average flow

rate (from 10 to 30 m3/min) and with high flow (from 30 to 25000 m3/min);

h) according to the developed pressure - vacuum, shrinking diluted gas to a pressure

slightly above atmospheric; compressors, shrinking gas to 0.3 MPa (without cooling); low

pressure compressors - from 0.3 to 1 MPa; an average pressure compressors - from 1.0 to

10 MPa; high pressure compressors - from 10 to 100 MPa; ultra high pressure compressors

- up to 250 MPa [3].

2. Basic parameters of the compression process. Nature of the compression

Thermodynamic compression is the process whereby the action of external forces

forcibly reducing the volume of gas to increase its pressure.

According to the kinetic theory of gases, their pressure is proportional to the total

kinetic energy of the molecules, which is converted into potential energy of pressure as a

result of their shots from the walls limiting gas volume. Upon compression of the gas

increases the number of hits of their molecules falling on unit area of restrictive walls, ie

increasing the gas pressure. As a result, the molecules hit the walls, and as a result of

blows between them, heat is released. This heat is converted back into kinetic energy,

which causes a further increase in pressure. The further increase of pressure is to increase

the impact energy on the walls as a result of the conversion of heat into kinetic energy of

the gas molecules.

Therefore, the increase in gas pressure in the ratio is due, and the reduction of its

volume, and by increasing its temperature.

If the pressure in the compression process is p1 and at the end of the process – p2,

then the ratio p2/p1 in theory compressors is we are referring to degree of pressure

increase and is written as pk :

1

2

pkp

p . (1)

Similar concepts are introduced and the relationship of the volume and temperature

of the gas at the beginning and end of the process compression. If the volume of gas in the

trial ratio is V1 and at the end of the process is V2, then the relation V1/ V2 is called the

degree of compression and marked withVk

:

1

2

VkV

V . (2)

If the gas temperature at the beginning of the process of compression is T1 and at

the end of the process - T2, then the ratio T1/T2 is called degree of the temperature

increasing and was marked with Tk

:

1

2

TkT

T . (3)

If in equation (1) the pressures p1 and p2 will be expressed from the equation of state

of ideal gas:

p.V = m.R.T, (4)

and taking into account equations (2) and (2) for the rate of pressure rise is obtained [3]:

TkVkpk

. . (5)

3. Screw compressors

Rotary screw compressors are widely used today in industrial refrigeration for

compression of ammonia and other refrigerating gases, petrol and gas industry or for

compressed air for industrial applications (and,function of final parameters and

compositions of gas screw compressors are two types : oil free and oil injected-flooded-

screw compressors)

3.1. Construction

A typical oil flooded twin screw compressor consists of male and female rotors

mounted on bearings to fix their position in a rotor housing which holds the rotors in

closely toleranced intersecting cylindrical bores (figure 1).

Fig.1. Construction and principle of work of rotary screw compressor

The rotors basic shape is a screw thread, with varying numbers of lobes on the male

and female rotors. The driving device is generally connected to the male rotor with the

male driving the female through an oil film. In refrigeration, four or five lobed male rotors

generally drive six or seven lobe female rotors to give a female rotor speed that is

somewhat less than the male speed. Some designs connect the drive to the female rotor in

order to produce higher rotor speeds thus increasing displacement. However, this

increases loading on the rotors in the area of torque transfer and can reduce rotor life.

3.2. Fundamentals of operation

A screw compressor is best described as a positive displacement volume reduction

device. Its action is analogous to a reciprocating compressor more than any of the other

common compressor types. It is helpful to refer to the equivalent recip. process to

visualize how compression progresses in a screw. Gas is compressed by pure rotary motion

of the two intermeshing helical rotors. Gas travels around the outside of the rotors,

starting at the top and traveling to the bottom while it is transferred axially from the

suction end to the discharge end of the rotor area.

3.2.1. Suction process

Suction gas is drawn into the compressor to fill the void where the male rotor

rotates out of the female flute on the suction end of the compressor. Suction charge fills

the entire volume of each screw thread as the unmeshing thread proceeds down the

length of the rotor. This is analogous to the suction stroke in a reciprocating compressor

as the piston is drawn down the cylinder (figure 2).

Fig.2. Beginning of suction

The suction charge becomes trapped in two helically shaped cylinders formed by the

screw threads and the housing as the threads rotate out of the open suction port. The

volume trapped in both screw threads over their entire length is defined as the volume at

suction, (Vs). In the recip. analogy the piston reaches the bottom of the stroke and the

suction valve closes, trapping the suction volume, (Vs) (figure 3).

Fig.3. Maximum suction

The displacement per revolution of the recip. is defined in terms of suction volume,

by the bore times the stroke times the number of cylinders. The total displacement of the

screw compressor is the volume at suction per thread times the number of lobes on the

driving rotor.

3.2.2. Compression

The male rotor lobe will begin to enter the trapped female flute on the bottom of

the compressor at the suction end, forming the back edge of the trapped gas pocket. The

two separate gas cylinders in each rotor are joined to form a "V" shaped wedge of gas with

the point of the "V" at the intersection of the threads on the suction end (figure 4).

Fig.4. Beginning of compression

Further rotation begins to reduce the trapped volume in the "V" and compress the

trapped gas. The intersection point of the male lobe in the female flute is like the piston

in the recip. that is starting up the cylinder and compressing the gas ahead of it (figure 5).

Fig.5. Compression

3.2.3. Discharge Process

In the recip. compressor, the discharge process starts when the discharge valve first

opens. As the pressure in the cylinder exceeds the pressure above the valve, the valve

lifts, allowing the compressed gas to be pushed into the discharge manifold. The screw

compressor has no valves to determine when compression is over. The location of the

discharge ports determine when compression is over (figure 6).The volume of gas

remaining in the "V" shaped trapped pocket at discharge port opening is defined as the

volume at discharge, (Vd).

A radial discharge port is used on the outlet end of the slide valve and an axial port

is used on the discharge end wall. These two ports provide relief of the internal

compressed gas and allow it to be pushed into the discharge housing. Positioning of the

discharge ports is very important as this controls the amount of internal compression.

Fig.6. Beginning of discharge

In the recip., the discharge process is complete when the piston reaches the top of

the compression stroke and the discharge valve closes. The end of the discharge process in

the screw occurs as the trapped pocket is filled by the male lobe at the outlet end wall of

the compressor (figure 7).

Fig.7. End of discharge

The recip. always has a small amount of gas, (clearance volume), that is left at the

top of the stroke to expand on the next suction stroke, taking up space that could have

been used to draw in more suction charge. At the end of the discharge process in the

screw, no clearance volume remains. All compressed gas is pushed out the discharge

ports. This is a significant factor that helps the screw compressor to be able to run at

much higher compression ratios than a recip.

3.3. Volume Ratio

In a reciprocating compressor, the discharge valves open when the pressure in the

cylinder exceeds the pressure in the discharge manifold. Because a screw compressor does

not have valves, the location of the discharge ports determine the maximum discharge

pressure level that will be achieved in the screw threads before the compressed gas is

pushed into the discharge pipe.

Volume ratio is a fundamental design characteristic of all screw compressors. The

compressor is a volume reduction device. The comparison of the volume of trapped gas at

suction, (Vs) to the volume of trapped gas remaining in the compression chamber when it

opens to discharge, (Vd) defines the internal volume reduction ratio of the compressor.

This volume index or "Vi" determines the internal pressure ratio of the compressor and the

relationship between them can be approximated as follows:

Vi = Vs/Vd, (6)

where: Vi is volume ratio or index;

Vs is volume at suction;

Vd is volume at discharge.

pi = Vi

k, (7)

where: pi is internal pressure ratio;

k is specific heat ratio of the gas being compressed.

Only the suction pressure and the internal volume ratio determine the internal

pressure level in the trapped pocket before opening to the discharge port. However, in all

refrigeration systems the condensing temperature determines the discharge pressure in

the system, and the evaporating temperature determines the suction pressure.

If the internal volume ratio of the compressor is too high for a given set of operating

conditions the discharge gas will be kept trapped too long and be raised above the

discharge pressure in the piping. This is called overcompression and is represented in the

pressure-volume curve in figure 8.

Fig.8. Over compression “p-V” diagram

In this case the gas is compressed above discharge pressure and when the port

opening occurs, the higher pressure gas in the screw thread expands out of the compressor

into the discharge line. This takes more energy than if the compression had been stopped

sooner, when the internal pressure was equal to the system discharge pressure.

When the compressor volume ratio is too low for the system operating pressures this

is called undercompression and is represented in figure 9.

Fig.9. Under compression “p-V” diagram

In this case the discharge port opening occurs before the internal pressure in the

compressor trapped pocket has reached the system discharge pressure level. The higher

pressure gas outside the compressor flows back into the lower pressure pocket, raising the

thread pressure immediately to the discharge pressure level. The compressor then has to

pump against this higher pressure level, rather than pump against a gradual build up to

discharge pressure level if the volume ratio had been higher, keeping the trapped pocket

closed longer.

In both cases the compressor will still function, and the same volume of gas will be

moved, but more power will be required than if the discharge ports are correctly located

to match the compressor volume ratio to what the system needs. Variable volume ratio

compressor designs are used in order to optimize discharge port location and minimize

compressor power.

3.4. Capacity Control

Capacity control is used in screw compressors to vary the amount of gas drawn into

the compressor. This is necessary in order to provide accurate suction temperature control

as evaporator load varies. Common capacity control methods are:

- slide valve controlling discharge port;

- slide valve controlling discharge port and volume ratio;

- slide valve not controlling discharge port;

- plug valves;

- variable speed.

Slide valves controlling the discharge port are a very common type of capacity

control device used in screw compressors. They are popular because they can give

infinitely adjustable control of capacity, often from 10 to 100 %. This type of slide valve

works by opening a recirculation passage in the high pressure cusp which allows a portion

of the trapped gas in the "V" shaped compression chamber to be recirculated back to the

suction cavity before it begins compression (figure 10).

Fig.10. Slide valve controlling discharge port

This method offers good efficiency at part load for two reasons. First, the

recirculated gas only has to overcome a slight pressure drop in order to bypass back to

suction since the recirculation slot opens to the trapped pocket before compression has

started, avoiding a precompression loss. Second, as the slide valve moves, the radial

discharge port is also being moved. As the trapped volume at suction is decreased, the

discharge port opening is also delayed, thus maintaining approximately the same volume

ratio at part load as at full load for optimum part load efficiency.

A compressor designed to control capacity and volume ratio is shown in figure 11. In

this design a movable slide stop is adjustable in the same bore as the slide valve. In this

design the discharge port position and the recirculation slot position can both be adjusted.

This allows an infinite number of adjustable positions for both valves, which provides

volume ratio and capacity adjustment from full load to approximately 40 % load, with

continuing capacity adjustment down to 10 % load. This arrangement offers improved

energy efficiency at full and part load.

Fig.11. Slide valve controlling discharge port and volume ratio

Slide valves that do not control the discharge port come in several varieties, the

most common is a round slide valve intersecting with slots in the rotor bore (figure

12).This type of unloader still gives good reduction of capacity but not as good a reduction

in part load power because it does not maintain the volume ratio during unloading. There

can also be some leakage across the slots in the rotor bore which can hurt performance at

all loads. These devices are lower in cost than conventional slide valves and used in some

smaller compressors.

Fig.12. Slide valve not controlling discharge port

Plug valves are radial or axial devices which lift to open a recirculation passage from

the trapped pocket back to suction. They will typically give unloading in steps of 75, 50,

and 25 % of full load as each progressive plug is opened. These devices also do not give

part load volume ratio correction like the first slide valves, thus part load efficiency is

comparable to slide valves that do not regulate the discharge port. Plug valves also tend

to be lower cost and simple in control method (figure 13).

Fig.13. Controlling with plug valves

Variable speed is occasionally used as a method of capacity control with screws. This

can be provided with speed controlled engines, steam turbines, or variable frequency

electric drives. Compressor power does not decrease linearly with speed reduction but

rather decreases as a function of rotor tip speed, and operating compression ratio. In

general, the compressor part load efficiency will be slightly better at low compression

ratio and significantly better at high compression ratio with reduced speed compared to

slide valve control, but this is before taking into account the losses in the driver at

reduced speed.

Variable speed control with screws should not be implemented without consulting

the compressor manufacturer. There are lower speed limits for compressors below which

bearings may fail due to inadequate bearing lubrication. Large compressors will have

lower minimum speeds than small compressors. Many small compressors may be able to

accommodate drive speeds above the input line frequency, but separator limits, oil cooler

size, and other package limitations must be investigated. It is also possible to fill a

compressor up with oil and cause failure if the speed is reduced below an acceptable

range with the compressor unloaded. Many of these limits are not published but should be

investigated early in a variable speed proposal or study [4].

4. Methods for testing of screw compressors

4.1. Testing of pressure distribution in twin screw compressors for multiphase duties

Multiphase production system (MPS) is a relatively new technology which is seeing

increased industrial application in recent years. Multiphase transportation technology

enables the transport the mixture of oil, water and gas, occasionally sand, natural gas

hydrates and waxes from wells via a single flow line to the processing facilities. MPS can

reduce the cost of exploring and conveying about 70 % that of a conventional facility.

Especially designed for application in multiphase transportation are a rotor dynamic pump

of the helicon-axial type and a positive displacement pump of twin screw type. The twin-

screw pump works on the principles of enclosing a defined volume on the suction side and

moving it to the discharge end by adding energy. It is used for liquids ranging in viscosity

from water-like consistency to polymers with a viscosity of millions of centipoise, and it

does a creditable job over a fairly wide range of gas void fractions (GVFs) including 100 %

gas for short periods.

Pressure distribution is showing large influences on the multiphase pumping behavior

especially during the boosting of two phase mixture. The pressure build-up has a very

important contribution to the fluid backflow recirculated through the internal clearance

back to the inlet of screws, which leads to the varieties of pump’s delivered volume flow

rate, volumetric efficiencies and the power consumption. On the other side, the

inlet/outlet pressure, rotational speed and gas volume fraction (GVF) have a direct

influence on the pressure distribution inside the working chamber. Nevertheless, the

prediction of pressure distribution of twin-screw pump on multiphase mixture presents a

challenge. Since the twin-screw multiphase pump with isometric process do face problem

with changes in temperature as pressure increase when the pump is operating with high

GVFs, the twin screw compressor with built-in volume ratio are used for multiphase

transport in this work. With the help of unloading valve, the twin screw compressor for

multiphase duties can ran at a variety of flow conditions covering GVFs ranging from 30 %

to 100 %.

A schematic diagram of the multiphase compressor test rig is shown in figure 14:

Fig.14. Multiphase compressor test rig layout

The gas is taken from the pressurized air network, mixed with liquid phase in mixer

just in front of the compressor suction. There is a separator set up downstream of the

compressor to separate the multiphase flow.

The measurement of global operation parameters, like the suction and discharge

pressure are recorded by data acquisition system (DAS). The shaft power is measured using

a torque meter. The volumetric flow rates of both phases are measured by turbine flow

meters. The temperature of the oil phase can be controlled by means of a water heat

exchanger.

Figure 15 presents the measuring system for recording the pressure distribution in

the working chamber. With aid of a slip ring on the rotational female rotor, the pressure

signal from the working chamber is led out of the multiphase compressor and then

conveniently processed in the standard data acquisition system.

Fig.15. Transducer installation and data acquisition

By this method of testing the following results of investigation can be achieved:

1) Pressure distribution under different GVFs.

2) Effects of discharge pressure on pressure distribution.

3) Effects of inlet pressure on pressure distribution.

4) Effects of rotational speed on pressure distribution.

5) Multiphase compressor performances with different GVFs [1].

4.2. Testing of screw refrigeration compressors under superfeed conditions

The refrigeration equipments are widely employed in both commerce and industry.

To give rise to an increase in the cooling capacity and COP of the refrigeration system, an

economizer is usually arranged in the refrigeration plants. Due to the advantages of twin

screw refrigeration compressor, such as high efficiency, stable operation, high reliability,

and so on, especially the superfeed process is achieved easily in twin screw compressor,

the twin screw refrigeration compressors are broadly used in various refrigeration systems,

which have gradually substituted for the reciprocating refrigeration system employed in

small cooling capacity and part of centrifugal refrigeration system employed in large

cooling capacity.

In refrigeration system, the twin screw refrigeration compressor with an

economizer arrangement is the key equipment. The performance of twin screw

refrigeration compressor is very sensitive to a number of design parameters governing the

thermodynamic and flow process, especially when the compressor is working under

superfeed condition. With the help of the simulation model and the experimental

recording of “p-V” diagrams, the effects of the process of superfeed with an economizer

on the compressor’s performance can be described clearly.

The working process of twin screw refrigeration compressors can be described by

mathematical models.

The geometrical parameters of the compressors can be calculated by computer

programs and the results can be input to the simulation program as functions of the male

rotor angle of rotation, such as volume curve, inlet port area, outlet port area, superfeed

port area, oil port area, slide valve by-pass port area, and all at-rest clearance areas

(contact line, rotor tips, cusp blow hole, and rotor end faces).

In the refrigeration systems, the capacity and COP can be improved by use of

economizer. On the other hand, the performance of twin screw compressors is affected by

the superfeed gas injection into compression chamber from the economizer. The one-

stage compression from suction to discharge in the compressor becomes the quasi-two-

stage compression. A superfeed process is added in the compression process originally.

The superfeed process in the control process of the compressor is shown in figure

16.

Fig.16. Schematic plan for superfeed process

When the control volume is connected with the superfeed port, the superfeed

process is beginning. If this control volume is disconnected with the superfeed port, the

superfeed process is finished. As the refrigerant gas is added in this control volume, the

pressure of the refrigerant gas is increased too.

Due to the superfeed mid-pressure refrigerant gas injection into the compressor, the

refrigeration circle is changed, so the “p-h” diagram is changed as shown in figure 17.

Fig.17. “p-h” diagram of refrigeration circle with economizer

In this figure, the point 9 stands for the beginning of the superfeed process and the

point 3 stands for the end of the superfeed process. The superfeed gas mixes with the gas

in the control volume, which results in the increase in pressure of gas in the control

volume from point 2 to point 3. And then, the mixed gas in the control volume is

compressed to point 4 in the compressor.

As shown in figure 17, it can be seen that the liquid refrigerant from the condenser is

throttled two times. After the first throttling action, the intermediate pressure saturated

gas is drawn into the compressor and the superfeed process occurs, the rest of the liquid

is throttled again and then it is evaporated in the evaporator. In this way, the evaporator

is provided with a larger percentage of refrigerant liquid, which gives an increased cooling

capacity.

Using this method of testing the following results of investigation can be achieved:

1) Comparison between the calculated and measured graphical dependences “mass

of superfeed gas – superfeed pressure”.

2) Drawing of “p-V” diagrams without and under superfeed process using the system,

shown in figure 18.

Fig.18. System of “p-V” diagram recorded

3) Comparison between the calculated and measured graphical dependences “shaft

or indicated power – superfeed pressure”.

4) Comparison between the calculated and measured graphical dependences

“isentropic efficiency – superfeed pressure” [2].

4.3 Testing of gas screw compressors ( for natural gas or other types of gas )

Fig.19 Diagram of experimental stand

Under the rules governing the positive displacement compressors-API619 and ISO

10440-1 - compression equipment before delivery must be subjected to tests, mechanical

running test, and- able to function in conditions certainly,at the nominal operating

parameters- the performance test. Stand configuration is according “Diagram of

experimental stand”, which takes into account the configuration mode of compressor: oil

required for injection compression chamber for lubrication of bearings, sealing is ensured

that during operation the compressor is pumped to the lubrication points the gas pressure

in the separator cushion of oil discharged from the vessel. Injected oil helps to dissipate

the heat from the compressor, the lubrication of bearings, providing space compression

seal, drive shaft seal. During operation must maintain a minimum pressure gradient,

specific to each lubrication point. Also it emphasizes that in order to avoid unwanted

seizure because mechanical consequences, it is recommended that, coupling between the

drive motor and compressor to make electromagnetic.

4.3.1.Mechanical test

To conduct mechanical compressor test on bench mounted, whose configuration is

according “Diagram of experimental stand”

Mechanical test takes at least 4 hours at rated speed. If measurements are made at other

speeds, steps will be 10% of nnom and measurements will be made only after approx.15

min constant speed operation in order to stabilize the system (temperatures, vibration,

etc.)

During the test, the following requirements must be met:

-mechanical operation must be correct

-vibrations must be within their normal limits. The values recorded in sample report

-buyer and seller may require variation should be able to pressure / temperature / flow

within the limits specified in the operating manual. The equipment shall be run for a

minimum 4 hours.

Parameters are recorded during the test, measurements are presented in sheets

If replacement or modification of bearings or seals or dismantling of the case or

modify other parts is required to correct mechanical or performance deficiencies, the

initial test will not be acceptable, and the final shop test shall be run after these

deficiencies are corrected.

4.3.2. Performance Test

This test is conducted conf.ISO -1217:1996 E. We believe that performance testing should

be performed for the first compressor in a range of parameters. Among the most important

parameters when we refer to a compressor, parameters are guaranteed by the supplier of

the equipment are:

a) gas flow delivered

b) power consumption for gas compression

c) volumetric efficiency

a) for metering the flow, the stand must have required instrumentation. It is

recommended that flow information can be used to calculate volumetric efficiency as to

measure the discharge flow (corrected to standard conditions of aspiration). Discharged

flow rate is due to the operator. Using that information flow intake, can introduce errors

because the compressor internal losses occur, due to report pressure / pressure difference

across the rotor, which reduces the flow discharged (compared to the measured flow at

inlet)

b) power consumption for gas compression

There are several ways to measure power absorbed. Stand measurement function

facilities can be made by:

- Electronic torqmeter, between engine and compressor

- Double-watt meter method or equivalent method

With measured power consumption, can be calculated isentropic efficiency as the

ratio between theoretical isentropic power and measured power. Theoretical

Isentropic power is given by:

Pt, is = Gxk/k-1 xRxTaspx (π k-1/k-1) [kW] where

G = mass flow [kg / s}

k = isentropic exponent of the gas

R = gas constant [kj/kg 0K]

Tasp = temperature at the inlet [0K]

π= compression ratio

c) Volumetric efficiency is defined as the ratio of the volume of gas delivered to

the discharge (corrected to standard conditions suction) and theoretical volume, specific

to each type of compressor (swept volume or theoretical volumes)

Because, in practical terms of measurement of volume is much more difficult on

the discharge side compared to the suction side-and this leakage is usually small- it is

often acceptable to simply use the suction volume in the calculation of volumetric

efficiency instead of the delivered volume ,converted to suction conditions

4. Conclusion

Rotary screw compressors are widely used today in industrial refrigeration because

of its inherent efficiency, safety, and flexibility.

The conditions in the refrigeration industry are changing and screw compressors are

also changing to meet customer's demands.

Modern machine tools and automated inspection equipment are making it possible to

hold tighter tolerances in day to day manufacturing environments. This improves

compressor performance and consistency from one compressor to the next.

The testing of the screw compressors is very important activity for checking of their

condition and work parameters. Therefore the improving of the compressor testing

methods is one of the ways for improving of the compressors.

6. Bibliography

[1] Feng Cao, Tieyu Gao, Songshan Li, Ziwen Xing, Pengcheng Shu. Experimental

analysis of pressure distribution in a twin screw compressor for multiphase duties.

Sciencedirect.com, 2010.

[2] Huagen Wu, Jianfeng Li, Ziwen Xing. Theoretical and experimental research on

the working process of screw refrigeration compressor under superfeed condition.

Sciencedirect.com, 2007.

[3] Grozev G., S. Stoyanov, G. Gujgulov. Hydro and pneumo machines and driving

installations. Sofia, 1990.

[4] Pillis J. W. Basics of Operation, Application & Troubleshooting of Screw

Compressors. 1998.

[5] API 619

[6] ISO 10440

[7] ISO 1217