Improving performance and development of two stage reciprocating compressors

International Journal of Advanced JOURNAL OF ADVANCED RESEARCH ISSN 0976 –
INTERNATIONAL Research in Engineering and Technology (IJARET), IN
6480(Print), ISSN 0976 – 6499(Online) Volume 3, Number 2, July-December (2012), © IAEME
ENGINEERING AND TECHNOLOGY (IJARET)
ISSN 0976 - 6480 (Print)
ISSN 0976 - 6499 (Online)
Volume 3, Issue 2, July-December (2012), pp. 119-136
IJARET
© IAEME: www.iaeme.com/ijaret.html
Journal Impact Factor (2012): 2.7078 (Calculated by GISI)
©IAEME
www.jifactor.com

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: ashr12000@yahoo.com

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

119

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

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

120


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

121


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

122


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

123


(1) Trolley
(2) Drive motor (11) Differential
(3) Acoustic pressure transducer
attenuator (12) Switch cabinet
(4) Pressure vessel with digital
2nd stage displays
(5) Pressure (13) Resistance
transducer thermometer
(6) Manometer (14) Pressure vessel
(7) Safety valve and intermediate
(8) Regulating cooler for 1st stage
valve (15) 1st Compressor
(9) Inlet pressure Stage
vessel (16) Stage piston
(10) Nozzle for compressor
volume flow (17) 2nd
measurement Compressor Stage

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

124


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 Length 1520 mm,
Width 800 mm,
Height 1500 mm,
Weight approx. 260 kg

Power supply 400 V / 50 Hz / 3 phase

Compressor layout: 2 cylinders in V-shape

Max. pressure 35 bar

Working pressure 13 bar

Intake capacity 15 m3/h = 250 l/min

Speed 710 rpm

Stage 1 details Bore 78 mm, Length of driving rod 150 mm

Stage 2 details Bore 45 mm, Length of driving rod 150 mm,
Stroke 72 mm

Drive motor: Rating 2,2 kW , Speed 3000 rpm

Inlet tank: Volume 20 l, Max. pressure 16 bar

Intermediate cooler tank: Volume 5 l, Max. pressure 16 bar

Outlet Pressure Vessel: Volume 20 l, Max. pressure 16 bar

Differential pressure sensor: Measuring range 0 - 10 mbar, Output signal 0 -
10 V DC, Supply 24 V DC

Pressure sensor: Measuring range 1x 0-1.6 bar abs. and 2x 0 - 16
bar abs., Output signal 0 - 10 V DC, Supply 24 V
DC

Resistance thermometer with Type PT 100 , Measuring range 0 - 200 °C,
transducer: Output signal 0 - 10 V DC

Power transducer: Measuring range 0 - 2500 W, Output signal 0 -
10 V DC

125


(1) P1-Inlet pressure
(2) T1-Inlet (7) dp-Differential
temperature pressure across
(3) P2-Pressure after Venturi nozzle
1st compressor stage (8) T4-Temperature
(4) T2-Temperature after 2nd
after 1st compressor compressor stage
stage (9) Emergency stop
(5) P4-Pressure after switch
2nd compressor stage (10) Master switch
(6) T3-Temperature (11) Electric motor
before 2nd switch
compressor stage (12) Electrical
power

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

El electricity
Pl pressure
indicator
Tl temperature
indicator
PD pressure
differential
M motor

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

126


(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)

& 2 × DP Eq. (1)
V = Ap ×
ρ

Where &
V : flowrate (m3 / s ), DP : differential pressure ( Pa), ρ : density (kg / m3 )
and Ap : the smallest cross - section of the Venturi nozzle = 1.131×10−4 m 2

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

100 × P0
ρ= Eq. (2)
287 × (T0 + 273)

Where ρ in kg / m3 , P0 in mbar , and T0 in 0C

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:

127


P 
Powerhydr = P × V1 × ln 4 
1
&
P Eq. (3)
 1

Where &
P in Pa, V in m3 / s, and Powerhydr : hydraulic power ( watt )

The overall efficiency is a relationship between the hydraulic output and the supplied electrical
power (measured using an electrical power transducer).

Powerhydr
η= Eq. (4)
PowerElect .

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

128


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.

15 150
14 140
13 T1 130
12 T2 120
11 T3 110
T4
10 100

Temp. (C)
P1
9 90
P (bar)

P2
8 P4 80
7 70
6 60
5 50
4 40
3 30
2 20
1 10
0 0
0 10 20 30 40 50 60 70 80 90 100

Time (seconds)

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 2nd stage, and T4, P4 after the 2nd stage)

15 150
14 140
13 130
T2
12 120
T3
11 110
T4
10 100
P2
Temp. (C)

9 90
P (bar)

P4
8 80
7 70
6 60
5 50
4 40
3 30
2 20
1 10
0 0
0 10 20 30 40 50 60 70 80 90 100

Time (seconds)

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

129


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

130


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

131


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.

15 150
14 140
13 T2 130
12 T3 120
11 T4 110
P2
10 100

Temp. (C)
P4
9 90
P (bar)

8 80
7 70
6 60
5 50
4 40
3 30
2 20
1 10
0 0
0 10 20 30 40 50 60 70 80 90 100

Time (seconds)

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

132


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

15 150
14 T2 140
13 T3 130
12 T4 120
11 P2 110
P4
10 100
Temp. (C)

9 90
P (bar)

8 80
7 70
6 60
5 50
4 40
3 30
2 20
1 10
0 0
0 10 20 30 40 50 60 70 80 90 100

Time (seconds)

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

133


15 150
14 140
13 T2 130
12 T3 120
11 T4 110
P2
10 100

Temp. (C)
P4
9 90
P (bar)

8 80
7 70
6 60
5 50
4 40
3 30
2 20
1 10
0 0
0 10 20 30 40 50 60 70 80 90 100

Time (seconds)

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.

134


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

REFERENCES

[1] B.N. Royce, Compressors selection and sizing, Library of Congress Cataloging-in-Publication
Data Compressors, selection & sizing / Royce N. Brown, 2nd ed. (1997) ISBN 0-88415-164-6.
[2] M. Ozu, T. Itami, Efficiency analysis of power consumption in small hermetic refrigerant rotary
compressors, International Journal of Refrigeration 4 (1981) 5- 14.
[3] Department of Energy: Office of Industrial Technologies, The N.C. Department of Environment
and Natural Resources, Division of Pollution Prevention and Environmental Assistance. The State
of North Carolina. Energy Efficiency in Air Compressors, Report - 3 January, 2004.
[4] M. Yang, Air compressor efficiency in a Vietnamese enterprise, Energy Policy 37 (2009) 2327-
2337.
[5] J.F. Hamilton, W. Soedel, R. Cohen, Recent research and development in positive displacement
compressors in the USA, Proceedings of the Conference on Design and Operation of Industrial
Compressors, The Institution of Mechanical Engineers, London (1978) Paper CI 0/78.
[6] M. Hayano, H. Sakata, S. Nagatomo, H. Muraski, An analysis of losses in scroll compressor
Proceedings of the International Compressor Conference at Purdue, Purdue University, West
Lafayette (1988) 189-197.
[7] M.G. Duggan, G.F. Hundy, S. Lawson, Refrigeration compressor performance using calorimeter
and flow rater techniques, Proceedings of the International Compressor Conference at Purdue,
Purdue University, West Lafayette (1988) 206-215.
[8] R. Keribar, T. Morel, Heat transfer and component temperature prediction in reciprocating
compressors Proceedings of the International Compressor Conference-at Purdue, Purdue
University, West Lafayette (1988) 454-463.
[9] A. Futakawa, Improvements in compressors with special emphasis on interesting developments in
Japan Proceedings of the International Compressor Conference-at Purdue, Purdue University,
West Lafayette (1984) 339-351.
[10] S. Etemad, J. Nieter, Computational parametric study of scroll compressor efficiency, design and
manufacturing issues Proceedings of the International Compressor Conference at Purdue, Purdue
[11] T. Hirano, N. Matsumura, K. Takeda, Development of high efficiency scroll compressors for air
conditioners Proceedings of the International Compressor Conference at Purdue, Purdue
[12] M. Elhaj, F. Gu, A.D. Ball, A. Albarbar, M. Al-Qattan, A. Naid, Numerical simulation and
experimental study of a two-stage reciprocating compressor for condition monitoring, Mechanical
Systems and Signal Processing 22 (2008) 374-389.
[13] R. Damle, J. Rigola, C.D. Pe´rezSegarra, J. Castro, A. Oliva, Object-oriented simulation of
reciprocating compressors: Numerical verification and experimental comparison, International
journal of refrigeration 34 (2011) 1989-1998.

135


[14] J. Rigola, C.D. Pe´rezSegarra, A. Oliva, Parametric studies on hermetic reciprocating
compressors, International Journal of Refrigeration 28 (2005) 253–266.
[15] F. Fagotti, Performance evaluation of reciprocating compressors operating with hydrocarbon
refrigerants, Proceedings of the International Compressor Engineering Conference, Purdue
University, USA (1994) 611–616.
[16] K.M. Igmatiev, M. Chrustalev, M.M. Pererozchikov, V.B. Zdalinsky, Simulation and parametric
analysis of the suction valve and muffler of a small reciprocating compressor, Proceedings of the
International Compressor Engineering Conference, Purdue University, USA (1996) 205–210.
[17] S. Manepatil, G.S. Yadara, B.C. Nakva, Theoretical study of design and operating parameters on
the reciprocating compressor performance, Proceedings of the International Compressor
Engineering Conference, Purdue University, USA (1998) 821–826.
[18] P. Grolier, A method to estimate the performance of reciprocating compressors, Proceedings of
the International Compressor Engineering Conference, Purdue University, USA (2002) C4–C5.
[19] K. Hyup, J.H. Lee, I.W. Lee, I.S. Lee, S.C. Park, Performance prediction of reciprocating
compressor, Proceedings of the International Compressor Engineering Conference, Purdue
University, IN, USA (2002) C4–C7.
[20] S. Porkhial, B. Khastoo, M.R.M. Razavi, Transient characteristics of reciprocating compressors
in household refrigerators, Appl Thermal Eng. 22 (2002) 1391–1402.
[21] L. Chen, J. Luo, F. Sun, C. Wu, Optimized efficiency axial-flow compressor, Applied Energy 81
(2005) 409–419.

136

Improving performance and development of two stage reciprocating compressors

Recomendados

Recomendados

Más contenido relacionado

La actualidad más candente

La actualidad más candente (20)

Similar a Improving performance and development of two stage reciprocating compressors

Similar a Improving performance and development of two stage reciprocating compressors (20)

Más de iaemedu

Más de iaemedu (20)

Improving performance and development of two stage reciprocating compressors