Improving performance and development of two stage reciprocating compressors

International Journal of Advanced Research in Engineering and Technology (IJARET), ISSN 0976 –

6480(Print), ISSN 0976 – 6499(Online) Volume 3, Number 2, July-December (2012), © IAEME

119

IMPROVING PERFORMANCE AND DEVELOPMENT OF TWO-STAGE

RECIPROCATING COMPRESSORS

Ashraf Elfasakhany

*

Department of Mechanical Engineering, Faculty of Engineering, Taif University, Box 888, Al-Haweiah,

Taif, Saudi Arabia

*Corresponding author Tel.: +966 (02) 7272020; Fax: +966 (02)7274299

E-mail address: [email protected]

ABSTRACT

The most troublesome part in the development of a compressor technology depends strongly on

improvement of its performance. For this purpose, a performance characteristic evaluation of a two-

stage reciprocating compressor is carried out in this paper. The aims were to improve compressor

performance by illustrating the effects of various parameters: primary air tank, compressor running

time, background working condition, and air leakage. The effect of each parameter was compared

with the normal performance condition and, in turn, it was demonstrated the most/least important

parameters on the performance. The parameters were measured using three techniques: the digital

display unit, instruments fixed on system layout, and a PC-data acquisition system. The experiment

addressed some factors that led to the inefficient performance of the compressed air system and cause

energy losses. The results advocate the optimal time for starting each stage of the two-stage

compressors. This work, in addition, may give the insight for the development of the design of multi-

stage compressors and presents some key design parameters.

Keywords: Reciprocating compressor, Two-stage, Performance, Development, Experimental.

I. BACKGROUND

A compressor is a mechanical device that takes an ambient air and increases its pressure [1]. In the

early time, the compressor was bellow that used by blacksmiths to intensify the heat in their furnaces.

The first industrial compressor was simple, a reciprocating piston-driven machine powered by a water

wheel. In the early 1960s, modern engineering was first applied to air compressor, and hereafter its

design was enhanced significantly.

A current industrial compressor is a system composed of several sub-systems and many

components. Subsystems include prime mover, controllers and accessories, treatment equipment, and

distribution unit. Controllers serve to regulate the amount of compressed air being produced, and

accessories keep properly operated system. The prime mover powers the compressor and the treatment

INTERNATIONAL JOURNAL OF ADVANCED RESEARCH IN

ENGINEERING AND TECHNOLOGY (IJARET)

ISSN 0976 - 6480 (Print)

ISSN 0976 - 6499 (Online)

Volume 3, Issue 2, July-December (2012), pp. 119-136

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equipment removes contaminants from the compressed air. Distribution unit transports compressed air

to where it is needed.

Generally, the sub-systems and equipment of compressed air systems are organized as air

compressor followed by coolers, separator, air dryers, and air storage tank. Such organization is

common for both single stage and multistage compressors.

In a multiple stage compressor, the final discharge pressure is generated over several individual

steps/stages. The individual stages are mounted in tandem with the second stage driven directly from

the rear of the first stage. Alternatively, the stages may be mounted side by side either in separate

stators or within a common, multi bore housing stator. Multistage compressors save energy by cooling

the air between stages, reducing the volume and work required to compress the air. After final stage,

same as in a single stage compressor, compressed air passes though coolers, separator, air dryers, and

air storage tank.

Storage tank has many functions; it represents the available air that can be released or replenished

at any time as required. It can be used to control demand events (peak demand periods) in a

compressed air system by reducing both the amounts of pressure drop and the rate of decay. Storage

can be used to protect critical pressure applications from other events in the system. Storage can also

be used to control the rate of pressure drop to end uses.

Due air or gas compression generates heat. Heat must be removed to maintain the compressor

equipment tolerances and clearances, and the compressed air is cooled to make it suitable for the

intended use. Compressor equipment units are cooled with air, water, and/or lubricant. Liquid-cooled

compressors have jacketed cylinders, heads and heat exchangers through which liquid coolant is

circulated to dissipate the heat of compression. Lubricating oil is used for cooling as well as

lubricating of mechanical parts at contact moving. Air-cooled versions have external fins fixed on the

compressed cylinder and cylinder head. Air is blown by a fan across the fins for heat dissipation.

Separators are devices that remove suspended water droplets from streams of air or gas. A

separator is generally installed following each cooler to remove the condensed moisture by the cooler.

When air leaves the cooler and moisture separator, it is typically saturated. Any further radiant

cooling, as air passes through the distribution piping and exposed to colder temperatures, will cause

further condensation of moisture. Excessive water in compressed air, in either the liquid or vapor

phase, can cause a variety of operational problems when such compressed air is used. These problems

include freezing of outdoor airlines, corrosion in piping and equipment, malfunctioning of pneumatic

process control instruments, and fouling of processes and products. These problems can be avoided by

a proper using of compressed air dryers. Air dryer is a device used for removing water vapor, formed

after the separators, from compressed air by increasing air temperature. The higher the air

temperature, the more moisture the air is capable of holding. However, drying the compressed air

beyond the required pressure dew point will result in unnecessary energy and costs. Air dryers vary in

types, and their performance characteristics are typically defined by the dew point. It is worth to clear

that the separators and dryers could be classified as parts of the treatment equipment system or as

individual components and, in this case, treatment equipment is considered only as air filters used to

clean the coming air.

Subsystems and components of industrial compressor are common for all compressor types

although these types vary significantly. The compressor types could be classified as two main types:

positive-displacement and dynamic. Each type has a completely different working principle than the

other. In dynamic compressor type, impellers rotate at very high speeds and impart velocity energy to

flow air or gas. The velocity energy is changed into pressure energy both by the impellers and the

discharge volutes or diffusers. This process uses the speed or velocity of the air to increase the air

pressure. Positive-displacement type, on the other hand, trap a given quantity of air or gas in a

compression chamber and the volume which it occupies is mechanically reduced, causing a

corresponding rise in pressure before discharge. At constant speed, the air flow remains essentially

constant with variations in discharge pressure.

Dynamic compressors are available in two main types: centrifugal and axial. Centrifugal

compressors use a rotating disk or impeller in a shaped housing to force the gas to the rim of the



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impeller, and, in turn, increasing the velocity of the gas. A diffuser section converts the velocity

energy to pressure energy. The centrifugal air compressors rotate much faster and generate much more

energy than the other types of dynamic air compressors due to their extremely fast moving blade.

Since it produces an immense amount of energy, these compressors are used for applications that

require a large amount of energy. They are used for continuous, stationary service in industries such as

oil refineries, chemical and petrochemical plants and natural gas processing plants. Many large

snowmaking operations (like ski resorts) use this type of compressor. They are also used in internal

combustion engines as superchargers and turbochargers. Out of some other types of dynamic air

compressors, the commonest is the centrifugal compressors. Their application can be from 100

horsepower (HP) to thousands of horsepower. With multiple staging, they can achieve extremely high

output pressures that reach to 10,000 psi (69 MPa).

Axial-flow compressors are dynamic rotating compressors types that use arrays of fanlike airfoils

to progressively compress the air or gas. The arrays of airfoils are set in rows, usually as pairs: rotating

and stationary. The rotating airfoils, also known as blades or rotors, accelerate the air/gas. The

stationary airfoils, also known as stators or vanes, decelerate and redirect the flow direction of the

air/gas, preparing it for the rotor blades of the next stage. Axial compressors are usually multistage,

with the cross-sectional area of the air/gas passage diminishing along the compressor to maintain

optimum axial Mach number. Axial-flow compressors are used where there is a need for a high flow

rate or compact design. Such compressor types are relatively expensive, requiring a large number of

components, tight tolerances and high quality materials. Mainly, they can be found in medium to large

gas turbine engines, in natural gas pumping stations, and within certain chemical plants.

Dynamic air compressors (in both their main types, centrifugal and axial) are very useful and

widely used in many applications, but not as common as the positive displacement compressors. The

positive displacement compressors are the most universally used compressors; not only are they

common in the industry but also they are popular at home and they are widely used by the mechanics

and woodworkers.

The positive displacement compressors are available in two types: reciprocating and rotary.

Reciprocating compressors work by pumping air into an air chamber then reducing this chamber's

volume. The manner in which they work is very similar to that of an internal combustion engine but

more or less in a reverse manner (they were classified, in many classifications, in a separate category

of piston type compressors). They have pistons, valves, cylinders, housing blocks and crankshafts. The

piston is used to compress energy by moving up and down, and the air stored inside the compressor

becomes compressed and converted into energy. Based on this simple mechanism, these compressors

are capable of producing a large amount of energy which can be used for many purposes;

consequently, reciprocating compressors are the most common types available in the market today.

They are generally found in wide ranges that vary from fractional to very high horsepower. Small

reciprocating compressors from 5 to 30 HP are commonly seen in automotive applications, and they

are typically for intermittent duty. Larger reciprocating compressors, over 1000 HP, are commonly

found in large industries and petroleum applications.

Reciprocating compressors can be either stationary or portable, can be single or multistage, and

can be single acting, double acting or diaphragm. The reciprocating air compressor is single acting

when the compressor is accomplished using only one side of the piston. A compressor using both

sides of the piston is considered as double acting. In diaphragm compressor type (also known as a

membrane compressor) the compression of gas occurs by the movement of a flexible membrane

instead of an intake element. The back and forth movement of the membrane is driven by a rod and a

crankshaft mechanism. Only the membrane and the compressor box come in contact with the gas/air

being compressed. Diaphragm compressors are a less common type and are used for compressing

hydrogen and natural gas as well as in a number of other applications. Generally, multistage double-

acting reciprocating compressors are said to be one of the very efficient and the largest compressors

available, and more costly than comparable rotary types.

Rotary compressors have gained popularity and are now the “workhorse” of American industry.

Generally, the efficiency of the rotary compressor is higher than that of the single stage reciprocating



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compressor [2]. Rotary compressors are most commonly used in sizes from about 30 to 200 HP. The

commonest type of rotary compressor is the helical-twin screw-type (also known as rotary screw or

helical-lobe). This type works on the principle of air filling in a void that is present between two

helical mated screws. As these screws turned, the volume is reduced volume and, in turn, air pressure

increased. This compressor type is mostly used to inject oil into the compression area and bearing for

a function of lubrication, cooling and even creating a kind of seal to reduce any leakage. The rotary

screw is considered to be the commonest rotary types because it has low initial cost, compact size, low

weight, and easy to maintain. However, sliding-vane, liquid-ring, lobe, and scroll-types are considered

among less common rotary types.

2. INTRODUCTION

Currently, compressors are multipurpose tools used widely in industry for a variety of purposes.

Compressors are used to operate various machines, tools and hydraulic devices; and, in many cases,

compressors are so vital that the facility cannot operate without them. Most industrial plants, from a

small workshop machine to an enormous power plant, pulp, and paper mill, have some type of

compressed air system wherein the energy generated from these compressors is essential to operate the

mechanical equipment and power tools. In view of that, plant air compressor can vary in size from a

small unit of 5 horsepower (HP) to huge systems with more than 50,000 HP.

Unfortunately, running air compressors often uses more energy than any other equipment in

industrial facilities [3]. Energy savings by means of system improvements of air compressors can

range from 30 to 50 percent or more of the electricity consumption [4]. For many facilities this is

equivalent to thousands, or even hundreds of thousands of dollars of potential annual savings,

depending on use. Since compressing air is one of the most expensive sources of mechanical energy in

the industrial setting [3], it is often financially beneficial and more energy efficient to use all possible

methods to reduce the energy consumption. The energy consumption of any compressed air system

depends on several factors: the compressor type, model and size, the motor power rating, control

mechanisms, system design, and performance.

Minimizing the energy consumption of air compressors and thereby improving compressors

efficiency and performance has always been the researchers' goal. Hamilton et al. [5] Summarized

different ideas that were presented by other researchers, as a reviewing study, to improve compressors

performance. The study presents improvements in electric motor efficiency, internal losses, system

effects, speed variation, valve stresses, accelerated life tests, and interaction of valve stress and compressor performance. Hayano et al. [6] deserves attention on friction losses in scroll compressors

and compared with other designs such as the rotary compressors; different frictional losses of different

parts of the compressors were predicted mathematically with identifying the location of the maximum occurred frictional losses. Duggan et al. [7] evaluated the performance of compressors using two

different measuring methods: calorimeter and flow measuring techniques. Keribar and Morel [8]

improved the heat transfer in reciprocating compressors using the finite element analysis. Futakawa

[9] reported improvements in compressors with special emphasis on events in Japan. Etemad and

Neuter [10] discussed the optimum design of scroll compressors using a parametric study analysis. Hirano et al. [11] reported a study of the leakage problems on performance of scroll compressors.

While there are many papers, as presented early, discussed the improvement of scroll compressors,

there are very few papers discussed the improvement of reciprocating compressors although

reciprocating compressors are one of the most popular machines in use in industry [12,13]. For this

purpose, a performance characteristic evaluation of a two-stage reciprocating compressor is carried

out in this paper. The aims were to improve compressor performance by illustrating the effects of

various parameters. Specific attention is devoted to valve leakage, primary air tank, compressor

running time, and background working conditions. Even though some parametric studies of

reciprocating compressors have been presented by other researchers, see e.g. [14–20], such studies are

usually based on the global thermodynamic models other than experimental. The technique applied in

the current study is experimental. Besides, until today there are no theoretical methods currently



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available that guarantee predicting the efficiency and performance optimization [21]; thus, the

experimental technique currently is more believed.

Not only are the parametric studies of reciprocating compressor in demand to improve the

compressor’s performance, but also the energy losses are fundamental in the efficiency and

performance. The improvement of energy losses has been highly in demand since energy loss from

compressor was up to 80% [4]. For this reason, minimizing the energy consumption of compressor

stages for a specific pressure output and thereby improving compressor efficiency is covered in this

study. Development on system design of a two-stage reciprocating compressor is discussed as well.

3. EXPERIMENTAL

3.1. Experimental Setup

The experiment was carried out in a two-stage reciprocating air/gas compressor mounted in a V-

shape with two separated cylinders, as shown in Figs. 1-2. The compressor, manufactured by Kaeser

(Model type K 2502 H35), is capable of producing about 35 bar maximum output pressure and 13 bar

output working pressure. The compressor was mounted in tandem that the second stage was driven

directly from the rear of the first stage. The air is firstly drawn into the intake tank via a measuring

nozzle that used to determine the intake volume. The intake tank acts as a calming zone and housing

for the measuring sensors of the intake state, i.e., pressure transducer and manometer. Between the

first stage and the second stage, there is a small pressure vessel for intermediate cooling. After the

second stage, the compressed air is forced into a storage tank via a cooling tube. To achieve a steady

operating state, the compressed air is blown off via a bleeder valve with sound absorbing. Safety and

pressure regulator valves, which are compulsory components in any compressor, are installed for

safety and control.

Fig.1. Layout of the two-stage reciprocating compressor connected with PC



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Fig. 2. Schematic diagram of a two-stage reciprocating compressor showing different components

and locations of various measuring instruments

The entire test stand is mounted on a trolley, which is a welded square steel tube supported with

four castors for a simple movement and maneuvering of the unit. It also contains two rollers with

breaks for secure positioning. The overall dimensions of the unit are 1520 mm length, 800 mm width,

and 1500 mm height. Such dimensions are designed to be not so heavy (260 kg weight approx.) for

effortlessness moving and accessing through normal doorways. Cushioned unit suspension was added

to calm down the experimental operation; a large intake damper was installed to smooth the volume

flow and as a support for the measuring sensors. The compressor was fully instrumented with different

sensors (pressure transducers, manometers, and thermocouples) for the experimental intention, as

shown in Figs. 1-2. Table 1 summarizes the specifications and characteristics of different components

of the two-stage reciprocating compressor.

All electrical controls and displays are fitted into a switch cabinet, which contains, as shown in

Fig. 3, the master switch, emergency stop switch, and digital displays for whole measuring variables.

It also displays the electrical output data, and the electric motor switch for compressor. Switch cabinet

is connected to a PC (computer) via USB cable for displaying and recording the measured data.

3.2. Experimental Procedure

The test stand was placed on a ground level and secured against rolling away by locking the

brakes. By switching on the system by pulling the emergency button OFF and turning the master

switch ON, the drive motor with a rating of 2,2 kW and speed of 3000 RPM begins to drive the two-

stage compressor. The first stage strokes the piston (bore 78 mm and 150 mm of driving rod length)

with a supplied intake capacity of 15 m3/h (250 l/min) from the intake tank 1 (20 liters volume and 16

bar maximum pressure). Air is compressed in the first-stage cylinder and forced through the

intermediate cooler tank (5 liters volume and 16 bar maximum pressure). Once air cooled down in the

intermediate cooler tank, air moved into the smaller second-stage compressor (bore 45 mm, length of

driving rod 150 mm, and stroke 72 mm). The second stage compresses the air further, and then air is

directed to the after cooler, as shown in Figs. 1, 2 and 4. Finally compressed air is stored in the storage

tank (20 liter volume and 16 bar maximum pressure).

It is important to clarify that, sometimes the compressor was not able to work after switching on

the compressor. This may attribute to the over-current protection switch, which may have cut out

(1) Trolley

(2) Drive motor

(3) Acoustic

attenuator

(4) Pressure vessel

2nd

stage

(5) Pressure

transducer

(6) Manometer

(7) Safety valve

(8) Regulating

valve

(9) Inlet pressure

vessel

(10) Nozzle for

volume flow

measurement

(11) Differential

pressure transducer

(12) Switch cabinet

with digital

displays

(13) Resistance

thermometer

(14) Pressure vessel

and intermediate

cooler for 1st stage

(15) 1st Compressor

Stage

(16) Stage piston

compressor

(17) 2nd

Compressor Stage



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electricity on the motor. Hence, we need to restart the compressor by turning the button in the switch

cabinet ON and turning the button on the pressure switch ON (between the two upper pressure tanks).

Normally, the compressor cuts-out with closing a bleeder valve at approximately 13 bar and cuts-in at

approximately 10 bar with opening a bleeder valve. The cut-in and cut-out pressure can be adjusted by

the pressure switch.

Table 1 Characteristics of the two-stage reciprocating compressor

Characteristics Values

Overall dimensions

Power supply

Compressor layout:

Max. pressure

Working pressure

Intake capacity

Speed

Stage 1 details

Stage 2 details

Drive motor:

Inlet tank:

Intermediate cooler tank:

Outlet Pressure Vessel:

Differential pressure sensor:

Pressure sensor:

Resistance thermometer with

transducer:

Power transducer:

Length 1520 mm,

Width 800 mm,

Height 1500 mm,

Weight approx. 260 kg

400 V / 50 Hz / 3 phase

2 cylinders in V-shape

35 bar

13 bar

15 m3/h = 250 l/min

710 rpm

Bore 78 mm, Length of driving rod 150 mm

Bore 45 mm, Length of driving rod 150 mm,

Stroke 72 mm

Rating 2,2 kW , Speed 3000 rpm

Volume 20 l, Max. pressure 16 bar



Measuring range 0 - 10 mbar, Output signal 0 -

10 V DC, Supply 24 V DC

Measuring range 1x 0-1.6 bar abs. and 2x 0 - 16

bar abs., Output signal 0 - 10 V DC, Supply 24 V

DC

Type PT 100 , Measuring range 0 - 200 °C,

Output signal 0 - 10 V DC

Measuring range 0 - 2500 W, Output signal 0 -

10 V DC



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Fig. 3. Switch cabinet with digital displays and controls (presented in Fig. 2 No 12)

Fig. 4. Schematic diagram of a two-stage reciprocating compressor showing process diagram with

measuring the locations of different instruments

During experiment, the system is allowed to run until constant pressures have built up and

stabilized of the measured values. The parameters were measured and recorded using three techniques:

the digital display unit (Fig. 3), instruments on the system layout (Fig. 4), and a PC-data acquisition

system (Fig. 1).

The digital displays indicate the absolute pressure at three locations in bar (P1 at inlet condition,

P2 after first stage and P4 after second stage). The manometers on the system layout indicate the

pressure at the same three measured locations that presented in the digital displays. Four resistance

thermometers (Pt100) with transducers were used to measure the temperature at four locations (T1 at

inlet condition, T2 after first stage, T3 after intercourse but before second stage, and T4 after the

second stage); temperature values were indicated in the system layout using fine-wire thermocouples

(1) P1-Inlet pressure

(2) T1-Inlet

temperature

(3) P2-Pressure after

1st compressor stage

(4) T2-Temperature

after 1st compressor

stage

(5) P4-Pressure after

2nd compressor stage

(6) T3-Temperature

before 2nd

compressor stage

(7) dp-Differential

pressure across

Venturi nozzle

(8) T4-Temperature

after 2nd

compressor stage

(9) Emergency stop

switch

(10) Master switch

(11) Electric motor

switch

(12) Electrical

power

El electricity

Pl pressure

indicator

Tl temperature

indicator

PD pressure

differential M motor



127

(K-type). In addition, the same temperatures were indicated on the digital displays. Fig. 4 shows

different measuring locations of transient air temperatures and pressures, as well as hydraulic

pressures and flow rates. The measuring locations were chosen carefully to monitor the performance

of each stage and to evaluate the overall compressor performance. The effective power of the motor

for compressor is measured using a transducer and indicated on the digital displays.

Hydraulic pressure and flow rate were measured at the inlet condition using a pressure transducer

and a flow meter, respectively. The differential pressure in the system inlet is measured via the

differential pressure transducer and showed at the digital display in mbar. The differential pressure (DP) is the difference between the ambient pressure and the pressure at the smallest cross-section of

the Venturi nozzle, as shown in Fig. 5 and measured based on the following relations. The differential

pressure DP in the Venturi nozzle is related to flow rate as:

Fig. 5. Venturi nozzle shows pressure differential (∆∆∆∆P or PD)

ρ

DPAV p

××=

2& Eq. (1)

Where )/(:),(:),/(: 33 mkgdensityPapressurealdifferentiDPsmflowrateV ρ&

2410131.1nozzle Venturi theofsection -crosssmallest the: mAand p

−×=

The density ρ of the air depends on the temperature and pressure as:

)273(287

100

0

0

+×

×=

T

Pρ Eq. (2)

Where CinTandmbarinPmkgin0

00

3 ,,/ρ

The efficiency of the compressed air system is measured and indicated in the PC-data acquisition

system. The efficiency is measured based on the following relations. Firstly, the hydraulic power is

calculated as:



128

××=

1

411 ln

P

PVPPowerhydr& Eq. (3)

Where )(:,/, 3wattpowerhydraulicPowerandsminVPainP hydr

&

The overall efficiency is a relationship between the hydraulic output and the supplied electrical

power (measured using an electrical power transducer).

.Elect

hydr

Power

Power=η Eq. (4)

The electrical power is indicated in the digital displays and in the PC-data acquisition system.

However, the hydraulic power and overall system efficiency are shown only in the PC-data acquisition

system. To run the program for PC-data acquisition, the test stand must be connected to the PC's USB

port during the experiment, as shown in Fig. 1. The PC allows recording the transient gas temperatures

and pressures during the whole working time, and that lends a hand to trace the performance history.

In addition to transient gas temperatures and pressures, hydraulic pressures, flow rates, hydraulic and

electric powers, and overall compressor efficiency are also obtainable in the recorded PC data

acquisition.

The measured values (P1, P2, P4, T1, T2, T3, T4, and DP) were indicated, in addition to the PC-

data acquisition system and the digital display unit, in instruments placed at the system layout. The

response time of various instruments used in this study is significantly smaller than 10 seconds per

cycle time of the compressor and 22 seconds of thermal time constant. The measuring range of

pressure sensor is about 0-16 bar absolute with output signal of 0-10 VDC and power supply of 24

VDC. The measuring range of resistance thermometer with the transducer of type PT 100 is about 0-

200 °C and the output signal is about 0-10 VDC. The measuring range of power transducer is about 0-

2500 W and the output signal is about 0-10 VDC (see Table 1).

4. RESULTS AND DISCUSSIONS

The results from experimental measurements of a two-stage reciprocating compressor are

presented in Figs. 6-15. Firstly, the background working conditions before starting up the compressor

are investigated, as shown in Fig. 6. As seen, all measured parameters (pressures and temperatures)

have no changes with time since stabilizing of the background conditions. The inlet temperature (T1)

and inlet pressure (P1) are about 20 oC and 0.9 bar, respectively. The pressure after first and second

stages (P2 and P4, respectively) is identical to P1 (0.9 bar) and, in turn, all curves come over each

other in the figure. The temperature after the intercooler (T3) gets slightly higher (about 22 oC);

however, the temperatures after first and second stages (T2 and T4, respectively) get in a much higher

level (38 oC for both); that may be attributed to the accumulated heat in the compressor material,

which leads to the higher temperature of T2, T3 and T4 than background condition (T1).

After starting up the compressor, pressures and temperatures performance curves vary with time,

as shown in Fig. 7. As seen, at first 3 seconds, all performance curves had no changes with time since

the compressor did not start yet, i.e., the same conditions as in Fig. 6. Henceforward, all curves were

changed except the T1, P1 and T3. The T1 and P1 (not shown in the figure) have no changes since

they are backgrounding conditions. T3 has no changes with time due to that the intercooler cools

down the air after the first stage to reach a lower stable condition (28 oC). This stable temperature

depends on the intercooler efficiency and the air temperature introduced into the intercooler from the

first stage (T2). Accordingly, T3 in all figures shows constant values, which can be high (as shown in

Fig. 14) or low (as shown in Figs. 7, 11 and 15) related mainly to T2 values since its efficiency is

unvarying. On the other hand, T2 and T4, increase rapidly with time, as shown in Fig. 7. T2 increases



129

even faster than T4 (within the period 3-70 seconds) since the P2 increases faster than the P4 (in the

early 20 seconds). After long operating time (about 70 seconds) T4 increases to exceed T2 wherein

pressure P4 gets much higher than the P2 (P4 reaches 11 bar but P2 reaches 3 bar). From the period of

20 to 70 seconds, T2 is higher than T4 although P4 goes beyond P2. The reason may be due to that the

temperature level is a result of two main reasons, the amount of air pressed and its pressure level.

More amounts of air and/or higher air pressure cause a warmer air. In the first stage, more amount of

air was compressed with less pressure level since it is larger size; however, second stage presses fewer

amounts of air but with a higher pressure level since it is smaller size. From 20-70 seconds, the

amount of air dominating and in turn T2 is higher than T4 but after that the pressure level increases to

be dominating.

Fig. 6. Temperatures (T) and pressures (P) at background conditions (T1, P1 at inlet condition, T2.

P2 after 1st stage, T3 after the intercooler and before the 2

nd stage, and T4, P4 after the 2

nd stage)

Fig. 7. Performance in normal working conditions (captions are seen in Fig. 6.)

0123456789

101112131415

0 10 20 30 40 50 60 70 80 90 1000102030405060708090100110120130140150

T1T2T3T4P1P2P4

Time (seconds)

P (

ba

r)

Te

mp

. (C

)

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101112131415

0 10 20 30 40 50 60 70 80 90 1000102030405060708090100110120130140150

T2T3T4P2P4

Time (seconds)

P (

ba

r)

Te

mp

. (C

)



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It is interested in noting that during the early operating time of the compressor (0-20 seconds), the

P2 was higher than P4. This is mainly due to that the first stage delivers air to the second one. Since

second stage does not have enough air that was delivered in the early time, this stage works with very

little amount of air and, in turn, the pressure on second stage increases slowly. Accordingly, it is not

recommended to start both stages at the same time, but the second stage should have a shift time delay

by about 20 seconds from the first stage. In case of using a multi stage compressor, the same technique

should be applied by delaying each stage 20 seconds from the previous one. This may be fulfilled by

adding a controller at each stage, which sets the operation of each stage separately according to a

specific starting time. By that way, we may improve energy losses, performance, and efficiency since

this working period is ineffective; in addition, as more stages of compressor used as more saving on

energy gained in compressor performance. For further verification of this finding, more performance

curves are investigated, as shown in Figs. 8-10. As seen, the air pressure of the second stage is lower

than the air pressure from the first stage at the first 5 seconds of operation, as shown in Fig. 8.

However, at 25 seconds, as shown in Fig. 9, the second stage overcomes this lower pressure and the

pressure of second stage increases to exceed the first one. The second stage increases even faster after

35 seconds, as shown in Fig. 10.

Fig. 8. Pressure of outlet air from each compressor stage after 5 seconds of starting up the operation

Fig. 9. Pressure of outlet air from each compressor stage after 25 seconds of starting up the

operation



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Fig. 10. Pressure of outlet air from each compressor stage after 35 seconds of starting up the

operation

During experiments, results show that the compressor performance curves were not similar at all

running times with the same working condition. In the early start time, we have the best performance

but after frequent working time, we do not have that good performance. This is because that the

compressor systems and components get hotter. Figs. 11-13 investigate the influence of working

duration on compressor performance. As shown in Fig. 11, the performance temperatures of T2 and

T4 are high at the beginning of operation (42 oC for both temps.). In comparison with the same

characteristic curves on cold condition of compressor equipment, as shown in Fig. 7, we note that the

temperatures of T2 and T4 at the beginning of operation are 38 oC each for both temps. By the end of

running cycle (at 80 seconds of continuing operation) we can see almost the same shift in the final

values of T2 and T4 (T2=90 oC, T4=97 oC as presented in Fig. 7, and T2=96 oC, T4=102 oC as

presented in Fig. 11). However, we cannot see any differences in pressure characteristic curves in both

cases (P2 is similar in Figs. 7&11, and also P4). For further investigation of pressure performance in

both cases, Figs. 12-13 present the pressure of the both cases after 45 seconds of operation, and as

seen there are no any differences. However, comparing efficiencies, as shown at bottom of Figs. 12-

13, we found that in case of cold starting condition (Fig.12) we have a higher performance efficiency

(η=40.6) than in hot starting condition (η=40.1). This means that the running time influences

significantly compressor efficiency but not in pressure performance. Hence, in order to stabilize the

performance, we may suggest using larger storage tank where compressor works for shorter periods

and rest for a long one and, in turn, reducing system and equipment temperature.

Investigating of background condition on compressor performance shows that the intake air

temperature has a significant impact on compressor efficiency and performance, i.e., working in a

colder environment leads to higher compressor efficiency than in hotter case but compressor outlet

pressures will be the same. This may attribute to that the energy required to compress the cool air is

much less than that required compressing the warmer air. Reducing the intake temperature by moving

the compressor intake outside the building and into a shaded area may drastically lower the energy

required for compression and, in turn, improving the efficiency. Additionally, recovering the intake air

temperature using a heat exchanger can be used to preheat the process (boiler water or space heating)

and improving the whole system efficiency. Finally, we may conclude that the background working

condition and/or compressor running time can significantly affect on overall system performance and

efficiency.

Investigating of air leakage on compressor performance is examined as shown in Fig. 14. As seen,

when compressor works with air leakage in valves and/or other equipment, the compressor

performance is highly affected. The compressor works for a longer period of time to reach the set



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pressure and, in turn, the temperatures of T2 and T4 get to very high value (about 110 oC). Besides,

the temperature at the beginning of operation is also very high (about 90 oC), in comparison with no

leakage condition (38 oC), as shown in Fig. 7, and that influences significantly the compressor

efficiency and performance, as discussed earlier. On the other hand, the outlet pressure reaches about

7 bar after 80 seconds of continuing operation, comparing with 12.5 bar at no leakage condition (in

Fig. 7). In case of larger leakage, the compressor may work continuously to reach the set pressure and

that may cause failure to some controller and subsystems. The air leakage problem is one of the areas

where the most significant energy losses can occur. Fixing the leaks has often been relatively cheap

and that have immediate results compared with great impact on energy use.

Fig. 11. Performance after the definite running time of operation (captions are seen in Fig. 6.)

Fig. 12. Pressure of outlet air from each compressor stage and efficiency at hot working condition

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Te

mp

. (C

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Finally, investigating the influence of the primary tank on compressor performance is carried out,

as shown in Fig. 15. Comparing performance with and without installing the primary tank (Figs. 7 and

15, respectively) shows that there is no influence on the performance at all, i.e., there is no difference

between both cases (with and without primary tank). Although early researchers recommend using the

primary tank for stabilization of pressure input, the current study shows an insignificant effect on

pressure performance and, in turn, primary tank can be eliminated in the coming designs.

Consequently, the new design will be less cost and space.

Fig. 13. Pressure of outlet air from each compressor stage and efficiency at cold working condition

Fig. 14. Performance with air leakage from system (captions are seen in Fig. 6.)

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p. (C

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Fig. 15. Performance with no primary tank installed in the system (captions are seen in Fig. 6.)

5. CONCLUSIONS

A two-stage reciprocating compressor was investigated experimentally in this work in order to

develop and improve the compressor performance. Several process parameters were carried out to

investigate the performance and efficiency: effect of the primary air tank, compressor background

working condition, compressor running time, and air leakage. The parameters were measured using a

digital display unit, instruments fixed on system layout, and a PC-data acquisition system. The results

show that the primary tank could be eliminated from the compressed air system without influencing

on the performance. Although early researchers recommend using the primary tank for stabilization of

pressure input, the current study shows an insignificant effect and, in turn, primary tank can be

eliminated in the coming designs. Consequently, the new design will be less cost and space

Investigating of air leakage on performance shows dominate effect where it is one of the most

significant energy losses and, in addition, it causes to reduce the output pressure. Air leakage causes

the compressor to work for a longer period of time to overcome the leakage and, in turn, compressor

may work continuously in case of a large leakage. Therefore, failure may occur in some system

controllers and components. As a recommendation, fixing the leaks has often been relatively cheap

and that have immediate results compared with great impact on energy use and failure of the system.

Examining of background conditions on compressor performance shows that the temperature of

the inlet air has an effect on overall system performance and efficiency. In case of lower background

temperature, we have a higher efficiency. This may attribute to that the energy required to compress

the cool air is much less than that required compressing the warmer air. Accordingly, reducing the

intake temperature by moving the compressor intake to an outside of building and into a shaded area

may drastically improve efficiency and, in turn, lower the energy required for compression work. On

the other hand, background conditions have no influence on the outlet pressure from the first and

second stages.

The same findings of background condition are applicable for the compressor running time. More

running time means hotter compressor systems and equipments and that leads to less performance and

efficiency; however, the output pressure has no changes. Hence, in order to stabilize the performance,

we may suggest using a larger storage tank where compressor works for shorter periods and rest for a

long one and, in turn, reducing system and equipment temperature.

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P (

bar)

Tem

p. (C

)



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Finally, this study may conclude that it is not recommended to start both stages of the two-stage

compressor at the same time but the second stage should have a shifted delay time by about 20

seconds from the first stage. This is due to that the first stage delivers air to the second one. Since

second stage has no enough air received during the early working time, it works with very little

amount of air and, in turn, the air pressure of the second stage gets lower than the first one. Setting the

starting time of each stage is recommended and that may be occurred via controllers. By that way, we

may improve energy losses and efficiency wherever the energy saving may be enlarged in case of

using multistage compressors.

ACKNOWLEDGMENT

The author would like to appreciate the partial financial support from the Faculty of Engineering,

Taif University, Saudi Arabia

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Improving performance and development of two stage reciprocating compressors

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