Pulsation and vibration study of reciprocating compressor according to api 618

International
OPEN ACCESS Journal
Of Modern Engineering Research (IJMER)
| IJMER | ISSN: 2249–6645 | www.ijmer.com | Vol. 4 | Iss.7| July. 2014 | 1|
Pulsation and Vibration Study Of Reciprocating Compressor
According To API 618 5th
Edition
Pradip P. Shejal1
, Dr. A. D. Desai2
1
(Mechanical Department, P. G. M. College of Engineering / Pune University, India)
2
(Mechanical Department, S. R. College of Engineering / Pune University, India)
I. INTRODUCTION
The high vibrations were reported at thee piping of compressor which was build according to API 618
Standard. This compressor is using to compress Hydrogen gas from 19 bar to 70 bar in two stages for feeding a
hydrocracker in a refinery. The flow rate of this compressor is varying from 226 kg/hr to 1136 kg/hr. The
compressor rotation speed is 742 RPM. The motor power for this 2 crank compressor is 355 kW.
The objective of this project is to reduce the vibration of piping system of reciprocating compressor.
To accomplish this, the following specific objectives are defined and completed.
 Fundamentals of Pulsation and Mechanical Vibration Theory
 Pulsation Analysis as per API 618 5th
Edition approach 2 : To reduce the pulsation across the piping
system by using orifices at different locations
 Vibration Analysis as per API 618 5th
Edition approach 3 : To reduce the Vibration across the piping
system by using supports at different locations
 Compare the results with before and after pulsation and vibration analysis plotting graphs in Velocity-
Time Domain.
Abstract: The compressor package contains the reciprocating compressor, pulsation dampers, gas
coolers and the connected pipe system which are often the heart of an installation and should be
operate smoothly and reliably. The compressor piping vibrations can contribute to fatigue failure of the
system or entire package which can lead to unsafe situations for human being as well as environment,
loss of capacity and increase in maintenance as well as repair cost and to avoid this situation
compressor piping vibration analysis to be carried out at a very early stage of the design of an
installation. The pulsation analysis should be carried out before the piping vibration analysis. The
guidelines for the pulsation analysis are given in API 618 Approach 2 and the guidelines for the
vibration analysis are given in API 618 Approach 3. The original system layout is checked with respect
to pulsations with all operating cases that are characterized by steady-state operating conditions. The
different gas properties and operation cases are considered with valve unloading cases as operation
cases. The measures are proposed to reduce pressure pulsation by installation of orifices. The shaking
forces are used in the subsequent vibration study. After the analysis of pulsation results, find out the
worst case for the vibration study means we used the shaking forces induced by pressure pulsation
excite the mechanical piping system of worst case. The Vibration Study determines the effect on the
mechanical piping system and proposes measures to avoid stresses possibly leading to deformation or
rupture by fatigue. The finite element program ANSYS is used for modeling of the mechanical system.
The model is built of several types of basic piping elements (e.g. pipes, beams, elbows, T-pieces)
connected at node points. The modifications are proposed to meet agreed criteria of vibration. This
paper demonstrated that by properly analysising the compressor piping vibration in an accurate and
economic way using Pulsim and Ansys software. The accuracy of the analytical solution is validated by
means of experimental results by using B & K Analyser for the measurement of compressor piping
vibration.
Keywords: ANSYS, API 618, APL LANGAGUGE, FFT ANALYSER, PULSE, PULSIM

Pulsation And Vibration Study Of Reciprocating Compressor According To API 618 5th
Edition
II. LITERATURE OVERVIEW
This literature review summarises, interprets and critically evaluates existing ―literature‖ in order to
establish correct knowledge of subject. The many experts present these views about the pulsation and vibration
in different literatures. Following are literature review is to go through the main topics of interest.
Shelley Greenfield and Kelly Eberle (2008) summarized an interesting literature overview of the
new API standard 618 (5th
edition) and its impact on reciprocating compressor package design. The new API
618 5th
Edition includes many improvements over the 4th Edition in the specification for engineering studies to
minimize pulsation and vibration. This provides the changes of The Pulsation, Vibration, Torsional, Skid and
Engineering studies. Also this gives the Summary of API 618 standard and key changes in the 5th
edition.
James D. Tison and Kenneth E. Atkins (2008) described the control philosophy for the pulsation
and vibration as per new fifth edition of API 618 for reciprocating compressors. This paper provide the user
with a working knowledge of good engineering practices for pulsation and vibration control of reciprocating
machinery in relatively high mole weight gases as well as an in depth understanding of the postposed changes
in API 618 and differing design philosophies. The conclusions for this paper are A) The intent of API 618 at
the Third Edition was that Design approach 3 meant effective pulsation control. This usually required that
reactive filtering be used in relatively high mole weight systems. B) The Fourth Edition attempted to define the
steps required to quality a piping system, in the event that the allowable pulsation levels were exceeded. This
created confusion and led to systems being designed with less emphasis on pulsation control, and justified with
mechanical response calculations of questionable validity. C) The fifth Edition will clarify the confusion that
resulted from the addition of the language concerning mechanical forced response calculations. The User will
now be able to determine if the systems mattes Design Approach 3 by the use of the technically sound pulsation
and shaking force control philosophy or through the use of the higher risk philosophies based on mechanical
forced response calculations.
Paul Alves (2006) studied the acoustical and mechanical Analysis of reciprocating compressor
installation as per API 618 4th Edition. This paper gives basic about the Pulsation, Dynamic Forces, Resistive
elements, Reactive elements, analysis cycle. Also he explained different clause which are mentioned in API
618 for Pulsation and Vibration analysis. We can easily understand API 618 M3 is applicable for pulsation
analysis and API 618 M4 to M8 are applicable for vibration analysis.
Enzo Giacomelli et al. (2006) studied the forced response of cylinder manifold for reciprocating
compressor applications by considering Crosshead guides, distance pieces, cylinder flanges, joint, supports etc
for the analysis. The study performed frequencies and amplitudes of pulsation induced shaking forces defined
by acoustical simulation, internal gas forces in the cylinder, and unbalanced mechanical forces and moments
allows a proper forced response analysis of the cylinder manifold system. The study applied these forces to the
finite element model to calculate the relevant vibrations and stress amplitudes by performing a harmonic
analysis. The paper covered the existing procedures, application experience and recommendations for properly
considering the applied loads.
J. C. Wachel and J. D. Tison (1999) investigated a wide variety of vibration and failure problems
occur in reciprocating machinery and piping systems. Excessive piping vibration problems usually occur when
a mechanical natural frequency of the piping system or compressor manifold system is excited by a pulsation or
mechanical excitation source since reciprocating compressors and pumps generate high pulsation forces,
vibration and failure problems in these systems are common. As per this paper, whenever high vibrations are
encountered in reciprocating compressors, pumps or piping, it is necessary to determine if vibrations and
dynamic stresses are acceptable.
W. W. von Nimitz (1982) described pulsation and vibration control requirements in the design of
reciprocating compressor and pump installations. The paper provided the basis for evolving improved methods
for assuring the reliability of reciprocating compressor and pump installations at the design stage. This
contains new methods for sizing surge volumes, maximum allowable pulsation levels at compressor valves and
in the piping systems, and improved pressure drop criteria based on performance. The introduction of the
paper explained the nature of the problem, previous work, purpose, and the contribution of the paper. The
contents of each section may be provided to understand easily about the paper.

Edition
III. FUNDAMENTALS OF PULSATION AND MECHANICAL VIBRATION THEORY
In order to understand how to control pulsation and vibration in positive displacement machinery systems, it is
imperative that one understands the differences between acoustical and mechanical concepts. In 2.1, the
acoustic issues, along with acoustic control techniques, will be addressed. In 3.2, the elements of the
mechanical system will be explained along with the concept of acoustic-mechanical coupling and the
mechanical techniques for controlling vibration.
Overview of Pulsation Concepts
Pressure variations that result from oscillatory flow of positive displacement machinery are the subject
of this section. These variations in pressure are referred to as pulsation. The pulsation occur in systems
handling both gases and liquids. High vibration, support degradation and fatigue failures caused by dynamic
forces induced by the pulsation are the most common problems resulting from pulsation. In order to reduce the
possibility of detrimental pulsation and vibration at the design stage, it is necessary to understand several
technical concepts Excitation mechanisms are first addressed in 2.1.1. Acoustic response is then explained in
2.1.2. In 2.1.3, the most common results of excessive pulsation are reviewed and the concept of acoustic-
mechanical coupling is explained.
General Brief about pulsation
Pulsations are the pressure and flow variations in gases and liquids that propagate in the pipe systems
and fluid machinery. Every pulsation consists of a pressure pulsation wave and a flow pulsation wave. It found
that where the pulsation is high, flow pulsation is low and vice versa. Always pulsation propagate with speed
of sound id the gas. Pulsation should be controlled in order to avoid dangerous vibration & fatigue in the pipe
system, ensure the integrity of the pipe system, get optimum performance of machinery like compressor and
pump, achieve high flow meter accuracy, Control Noise.
Effect of Pulsation
The pulsation that results in high shaking forces can cause excessive vibration in a piping system.
However, excessive vibration can occur even in cases where the dynamic forces are low if an excitation
frequency is close to, or coincides, with a mechanical natural frequency. In this case, vibration will be
amplified, typically a factor of 5 through 10 compared to the off-resonance condition. The amplitude at
resonance is limited by the damping of the system. The Pulsation cause pipe vibrations and subsequently
failures due to fatigue of the material, reduce compressor efficiency, produce noise, cause errors or inaccuracy
in the flow metering, reduce the lifecycle of the compressor valves.
Pulsation Control Methods
The pulsation control in compressor piping systems can be accomplished by application of the basic
acoustic elements of acoustical compliance (volume), acoustical iterance (choke tube), and resistance (pressure
drop). These elements can be used individually or combined in various manners to achieve pulsation control.
The pulsation suppression devices range from single surge volumes (empty bottles) to acoustic filters (bottles
with internals or utilizing secondary volumes), often used in conjunction with orifice plates. The user should
understand that this discussion is not intended to enable one to design these elements themselves.
Spread of Pressure Wave
Excitation Sources
In systems utilizing positive displacement machinery, the flow of gas or liquid is not steady. Instead,
the fluid moves through the piping in a series of flow pulses (dynamic or time varying), which are
superimposed upon the steady (average) flow. As an example, the magnitude and shape of the flow pulses
through the compressor valves in a reciprocating compressor cylinder are determined by physical, geometrical
and mechanical characteristics of the compressor (rotational speed, bore, stroke, loading, compression ratio,
etc.). These flow pulses act as excitations which create pressure and flow modulations (acoustic waves) that
move through the process fluid as it moves through the piping system. Generally, the predominant pressure
and flow modulations generated by a reciprocating compressor are at frequencies which can be modeled as
one-dimensional waves. An important part of the acoustic analysis is the development of a compressor model
that accurately predicts the dynamic flow excitation (flow versus time) delivered by the compressor. Some
simplified examples are shown in Fig.: 3.1, Fig.: 3.2, Fig.: 3.3, and Fig.: 3.4.

Edition
Fig.: 3.1 — Piston Motion and Velocity for a
Slider Crank Mechanism
Fig.: 3.2 — Single Acting Compressor Cylinder
with Rod Length/Stroke = ∞ and No Valve Losses
Fig.: 3.3 — Symmetrical, Double Acting
Compressor Cylinder with Rod Length/Stroke = ∞
and No Valve Losses
Fig.: 3.4 — Unsymmetrical, Double Acting
Compressor Cylinder with Rod Length/Stroke = 5
and No Valve Losses
Acoustic Response and Resonance
The flow pulses caused by the reciprocating action of the compressor or pump create pressure pulses
or waves that move through the piping system as shown in Fig.: 5.
Fig.: 3.5 — Traveling Wave in Infinite Length Pipe
While the flow pulse frequencies generated by the compressor are a function of the mechanical
properties of the compressor, the acoustical response in the piping is a function of the mechanical properties of
the compressor, the thermo physical properties of the gas, and the acoustical network defined by the attached

Edition
piping. When a particular harmonic of running speed is near or coincident with an acoustical natural
frequency, the acoustic response (dynamic pressure amplitude) is amplified. These resonances that occur when
an excitation frequency coincides with a natural frequency can be simple organ-pipe type resonances or
complex modes involving all of the piping. For simple constant diameter lines with open and/or closed
boundary conditions, specific pipe lengths determine acoustical natural frequencies. If a line length coincides
with integer multiples of one half or one quarter of the wavelength, depending on the combination of open or
closed end conditions, an acoustical resonance can be excited. End conditions are defined as either open or
closed. For half wave resonances, both end conditions must be the same, i.e. open-open or closed-closed. For
quarter wave resonances, the end conditions must be opposite, i.e. one open end and one closed end. Examples
of these configurations are shown in Fig.: 6 and Fig.: 7, and are defined by following Equations.
Fig.: 3.6 — Mode Shapes of Half Wave Responses
Fig.: 3.7 — Mode Shapes of Quarter Wave
Responses
Formula for half wave (closed-closed and open-open acoustic response frequency):
Formula for quarter wave (open-closed acoustic response frequency):
Fig.: 3.8 — Pulsation Wave Form Fig.: 3.9 — Propagation of Pulsation Wave

Edition
Fig.: 3.10 — Pulsations and Vibrations – the mainline
Pulsation and Vibration Study – The Mainline
Pulsation study
The Pulsations Study defined the height of the pulsations and describes measures to reach the
pulsation limits.
The pulsation limits are most according API 618, Approach 2. The calculation program used for this analysis
is Pulsim 3.1 which is developed by M/S TNO, Netherland. The more details about this will discuss in chapter
III.
Vibration study
Vibrations are mechanical oscillation in a plant, which can occur in pipes, vessels and beam
structures. The vibration provides cyclic stress and can lead to fatigue break of the piping system.
In the vibrations study the dynamic behavior of the piping system is investigated and the necessary measures
are described to reach the vibration limits. Base of the vibration study is the previous performed Pulsation
Study. The Limits are most according API 618, Approach 3. The calculation program used for this analysis is
ANSYS. The more details about this will discuss in chapter IV.
Fig.: 3.11 — Pulsation and Vibration Study – the mainline

Edition
IV. STUDY OF PULSATION ACCORDING TO API 618 (ANALYTICAL METHOD)
According Design Approach 2 of API Standard 618 5th
edition 2007, an acoustic simulation is
required to evaluate and control pulsation in piping systems at reciprocating compressors. The pulsation report
contains a summary of the pressure pulsations in the piping system as well as modification proposals to reduce
pulsation to the agreed level. The original system layout will be checked with respect to pulsation. All
operating cases that are characterized by steady-state operating conditions will be checked, thus transient cases
(e.g. start-up-cases) are excluded. Different gas properties and operation cases are considered. Valve unloading
cases are considered as operation cases. Measures are proposed to reduce pressure pulsation, e.g. installation of
orifices, changes in piping length or diameter. Shaking forces are used in the subsequent vibration study.
Used Documents for Pulsation Study
The exact gas data means gas composition and gas properties like velocity of sound, density, pressure,
temperature flow for which compressor is going to use is considered for the pulsation simulation. Also the
Isometric drawings of piping and compressors are used for the pulsation simulation.
Table : 4.1 — Gas Properties / Input Data for Pulsation Simulation
Load case: Unit
1 2 3 4 5 6
EOR
100%
SOR
100%
Low Purity
alt design
case
EOR
50%
SOR
50%
Low Purity
alt design
case 50%
Index Name
1
1st stage
suction
VOS [m/s] 1358.8 1358.8 876.9 1358.8 1358.8 876.9
density [kg/m3] 1.521 1.521 3.531 1.521 1.521 3.531
FFF [-] 0.005 0.005 0.005 0.005 0.005 0.005
pressure [bar] 19.8 19.8 19.8 19.8 19.8 19.8
temperature [°C] 40 40 40 40 40 40
flow [kg/h] 486 487 1136 226 227 530
wave
length
[m] 109.9 109.9 70.9 109.9 109.9 70.9
2
1st stage
discharge
VOS [m/s] 1506.5 1504.4 960.6 1510.8 1508.7 963.4
density [kg/m3] 2.466 2.457 5.831 2.499 2.491 5.923
FFF [-] 0.005 0.005 0.005 0.005 0.005 0.005
pressure [bar] 39.21 38.98 39.24 39.97 39.73 40.09
temperature [°C] 107 106 100 109 108 102
flow [kg/h] 486 487 1136 226 227 530
wave
length
[m] 121.8 121.6 77.7 122.2 122.0 77.9
3 Cooler 1-2
VOS [m/s] 1440.9 1439.8 924 1444.5 1443.3 926.2
density [kg/m3] 2.669 2.658 6.260 2.708 2.696 6.365
FFF [-] 0.005 0.005 0.005 0.005 0.005 0.005
pressure [bar] 38.74 38.515 38.74 39.49 39.255 39.58
temperature [°C] 73.5 73 70 75 74.5 71.5
flow [kg/h] 486 487 1136 226 227 530
wave
length
[m] 116.5 116.4 74.7 116.8 116.7 74.9
4
2nd stage
suction
VOS [m/s] 1372.2 1372 885.5 1374.8 1374.7 887.2
density [kg/m3] 2.917 2.900 6.772 2.962 2.945 6.895
FFF [-] 0.005 0.005 0.005 0.005 0.005 0.005
pressure [bar] 38.27 38.05 38.24 39.01 38.78 39.07
temperature [°C] 40 40 40 41 41 41
flow [kg/h] 486 487 1136 226 227 530
wave
length
[m] 111.0 110.9 71.6 111.2 111.2 71.7
5
2nd stage
discharge
VOS [m/s] 1516.5 1506.9 969.3 1514.6 1505.8 968.1
density [kg/m3] 4.436 4.327 10.465 4.448 4.401 10.494
FFF [-] 0.005 0.005 0.005 0.005 0.005 0.005
pressure [bar] 70.2 67.68 70.2 70.2 68.68 70.2
temperature [°C] 100 96 94 99 95 93
flow [kg/h] 486 487 1136 226 227 530
wave
length
[m] 122.6 121.9 78.4 122.5 121.8 78.3
Calculation and Interpretation of Analytical Results
The study of the gas pulsation is carried out with the digital simulation program PULSIM which has
been developed by the TNO/TPD Institute in Delft (NL). A system of pipes is built of several complicated parts
and elements which must be simplified for calculations. For example volumes of a cylinder casing between
valves and connecting flange, heat exchangers, separators, dampers, etc. The data of the gas might also change
during operation. In order to take these uncertainties into account, we make several calculations at different
velocities of sound. The range of variation is  12 % and it is varied in steps of 2%. The results are presented
in tables.
Admissible limits
The main requirements of API618 are as follows:
[API618, section 7.9.4.2.5.2.2.2] ―For systems operating at absolute line pressures between 3.5 bar and 350 bar
(350-35’000 kPa), the peak-to-peak pulsation level of each individual pulsation component shall be limited to
that calculated by (in SI units):‖

Edition










fDP
aP
IL
400
3501
With the following equation for the frequency f
60
Nrpm
f


For absolute pressures less than 3.5 bar, the peak-to-peak levels of individual pulsation components need only
meet the levels calculated for an absolute pressure of 3.5 bar.
Outcome of Pulsation Study for 1st Stage Suction
With the original system layout the pulsations are too high. The dominant order of pulsation is
primarily the 1st
& 2nd
order. The 1st
stage suction side showed slightly increased pulsation, the worst case was
the EOR 50% regulation of case 4 which reached 4.3% ptp. At the state of realization a change in piping was
not practicable; orifices were used to dampen the pulsation to the possible minimum. The residual pulsation is
still above the API limit. With orifices the worst case reached values 1.4 times higher than the API Limit. For
the calculation of the vibration analysis we use the case with the highest pulsation with its shakingforces. With
this procedure we ensure that the increased pulsation does not affect the piping stability. The vibration study
however showed throughout acceptable values for the suction system. A reason for this is the relatively small
shaking forces due to the relatively low pressure pulsation. No modifications of the piping were made on the
suction side. The original system and orifices with only a small pressure drop gives satisfying pulsation results.
With the recommended modification no. 5, we achieve satisfied results. The following table shows the highest
value of the pulsation at different deviation of velocity of sound for design case at some representative nodes..
Table : 4.2 — A Few Suction Calculation Results
for Different Load Cases
Following are the recommended orifices with
locations :
Table : 4.3 — Recommended Orifices for Suction
Side
Outcome of Pulsation Study for Interstage
primarily the 1st & 2nd order. The interstage showed high pulsation, the worst case was the EOR 50%
regulation of case 14 which reached 6.0% ptp. At the state of realization a change in piping was not
practicable; orifices were used to dampen the pulsation to the possible minimum. For the calculation of the
vibration analysis we use the case with the highest pulsation with its shakingforces. With this procedure we
ensure that the increased pulsation does not affect the piping stability. The vibration study however showed
throughout acceptable values for the interstage system. A reason for this is the relatively small shaking forces
due to the relatively low molecular weight. No modifications of the piping were made on the interstage side.
The original system and orifices with only a small pressure drop gives satisfying pulsation results. With the
recommended modification no. 5, we achieve satisfied results. The following table shows the highest value of
the pulsation at different deviation of velocity of sound for design case at some representative nodes.

Edition
Table : 4.4 — A Few Interstage Calculation
Results for Different Load Cases
Following are the recommended orifices with
locations :
Table : 4.5 — Recommended Orifices for
Interstage
Outcome of Pulsation Study for 2nd
Stage Discharge
primarily the 1st
& 2nd
order. The discharge stage showed high pulsation, the worst case was the Low Purity alt
design case 50% regulation of case 26 which reached 4.3% ptp. At the state of realization a change in piping
was not practicable; orifices were used to dampen the pulsation to the possible minimum. For the calculation of
the vibration analysis we use the case with the highest pulsation with its shakingforces. With this procedure we
ensure that the increased pulsation does not affect the piping stability. The vibration study however showed
throughout acceptable values for the interstage system. A reason for this is the relatively small shaking forces
due to the relatively low molecular weight. No modifications of the piping were made on the discharge stage.
The original system and orifices with only a small pressure drop gives satisfying pulsation results. With the
recommended modification no. 5, we achieve satisfied results.
Table : 4.6 — A Few Discharge Calculation
Results for Different Load Cases
We recommend to install the following orifices:
Table : 4.7 — Recommended Orifices for
Discharge Side
Sketches of Piping for the Nodes of Pulsation Study
Fig.: 4.1 — Sketch of Suction Piping for Pulsation Study

Edition
Fig.: 4.2 — Sketch of Interstage Piping for Pulsation Study
Fig.: 4.3 — Sketch of Discharge Piping for Pulsation Study

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V. STUDY OF PIPING VIBRATION ACCORDING TO API 618
In this chapter the results for the vibration study for the compressor type can be found. The vibration
study gives an overview of the dynamic behaviour of the piping systems. The natural frequencies and the
corresponding modes are determined. In a second step the response of the system towards shaking forces which
are caused by the pressure pulsation is calculated. The resulting displacement, velocity and bending stress
values should not exceed certain limits.
The most critical case of the pressure pulsation is used for the dynamic analysis. Uncritical parts of
the piping are not further investigated.
General Information
Modelling the vibration system
The system is modelled within the Finite Element Program ANSYS. It consists of a piping modelling
tool, with several different elements and modelling possibilities. Elements are the connection between different
nodes. The vibration study consists in general of the following ANSYS elements:
 BEAM4 : supporting structures
 COMBIN14 : spring supports
 PIPE16 : straight pipes & flanges / valves with specific mass and flexibility factors
 PIPE18 : pipe elbows
Fig.: 5.1 — Typical Finite Element Model
On the compressor skid our supports are modelled with beam elements, like in fig.: 1. On the
customer side of the piping, supports are modelled by restricting at least one degree of freedom of a node. On
request we model even the customer piping with supporting structures. Without restriction a node has six
degrees of freedom, translations and rotations in each direction. These are the following four typical types of
supports:
 Rest Support restricting only translation in vertical direction
 Directional Guide restricting translation in vertical and transversal directions
 Fixed Support restricting translations in all directions
 Anchored Support restricting translations and rotations in all directions
If friction forces are high enough, a support can possibly restrict an additional direction. For cases with a slight
excess in vibration we check if this additional restriction can be furnished by the expected friction. For static
thermal analyses, with in general higher force values than for vibration calculations, the friction is neglected.
Calculation procedure
The calculation for the vibration analyses is carried out in two steps.
The first step is the calculation of natural frequencies and mode shapes. This indicates which excitation
frequencies will cause resonance effects, and which part of the system will be the most affected. This gives a
general overview of the vibration behavior of the system.

Edition
Fig.: 5.2 — Typical Finite Element Model (Wireframe)
In step two we perform a dynamic response analysis. The excitation comes from the shaking forces,
which are a result of the pressure pulsation in the piping system. The forces are calculated with the PULSIM
program (see pulsation study) and are given as harmonics of the compressor speed for the first twelve orders.
Shaking forces occur in case of a change of the flow area or direction (inlet/outlet of a volume, elbows,
orifices,...). We use the ―worst case‖ shaking forces that means the forces from the calculated deviation of
velocity of sound that delivered overall the highest values.
The results of this calculation are the dynamic displacements, vibration velocities and bending stresses
of the pipes as well as the reaction forces on the supports. The peak-to-peak values are given in tables, and for
nodes of special interest the time functions are plotted over one crank revolution. The peak-to-peak values are
compared to the maximum allowable limit.
Admissible limits
According API 618 5th
ed. the stress in the pipes should not exceed the endurance limit of the material
used. For steel pipes the peak to peak cyclic stress should be less than 179 N/mm2
including all stress
concentration factors. Normally use a much lower limit of 45 N/mm2
, because we do not know all locations of
stress concentration, especially at welded pipes, and we allow for inaccuracies to a certain degree. Stress
intensification factors at pipe elbows and tee pieces are taken into account, according to ANSI B31.1. An
additional limit we use, which is not limited by API, is the maximum displacement value of 1 mm. A pipe
vibrating at such a high level, may cause problems and will generally leave a bad impression of plant
operation, even if stress limits are not exceeded. Moreover, we survey the vibration velocities at all nodes,
which should not exceed 30 mm/s, according to our experience. The support loads are given as information
only, and have to be checked individually by the customer, if supports could carry the calculated reaction force.
Guide to the Sketches
The node numbers for the following sketches are sorted by the following criterions:
 Node 1 - Node 84 and Node 118 - Node 374 for straight pipes
 Node 85 - Node 117 for the compressor model
 Node 375 - Node 440 for support constructions
 Node 441 - Node 647 for bends
 Node 648 - Node 713 for additional nodes near T connections
The numbers arise in the direction of the flow in the pipes.
Further details are given on the sketches.

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Sketches of Piping for Vibration Study
Fig.: 5.3 — Sketches of Suction Piping for Vibration Study
Fig.: 5.4 — Sketches of Interstage piping for Vibration Study

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Fig.: 5.5 — Sketches of Discharge Piping for Vibration Study
Fig.: 5.6 — Sketches of Support and Compressor

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Detailed Results
Summary of Results
The investigation of the piping showed no severe problems with vibrations. One reason for this is the
reduced pulsation level which is achieved from the modifications recommended in the pulsation study.
The model for this vibration study includes the suction, interstage and discharge side. Therefore natural
frequencies could be calculated once for whole system. For the dynamic response analysis we choose deviation
of the velocity of sound with the highest pulsations of the worst case. These are the following load cases:
 for suction : Case 4, EOR 50% - Sys 4 Mod 5, variation of velocity of sound +12%
 for Interststage : Case 4, EOR 50% - Sys 14 Mod 5, variation of velocity of sound +12%
 for discharge : Case 6, Low Purity alt design case 50% - Sys 26 Mod 5, variation of velocity of sound
+12%
The model contains the suction side starting at 6‖-P-551093-B2A1-IT existing header, suction piping,
interstage piping with cooler, KOD and PSV line, the discharge side with PSV-line and with the exiting
header 4‖-P-551107-B4A1-IH40. Relevant sidelines are also included in the calculation.
No change of piping layout is required except orifices which are already discussed for the pulsation study and
also following support should be added into the piping layout.
 Support : X & Y at Node 345 : Refer Sketch
 Support : X & Z at Node 367 : Refer Sketch
After several calculations with the support modification all vibration results are within the limits.
The supports with its restricted direction of the last version are indicated in the attached isometrics. At the
piping parts it was necessary to implement additional supports to achieve satisfied results. The details of all
supports (blocked translations and rotations) in the calculation model can be found in the sketches of the
system. These were discussed with the customer, including blocked translations and rotations for all supports.
The highest displacement occurs at the node 145 for the safety valve 554 for interstage piping before the
interstage cooler EA-45555. The detailed picture of the displacement for this node (Appendix A4-1) shows
influences of the 1st
order. The model analysis of the part at node 145 shows resonance for a frequency around
12.942 Hz (16th
eigenmode on fig.: 14). This value is close to the compressor frequency of the 1st
order (12.37
Hz). All values of displacement at this nodes are within the limit.
The highest velocity occurs at the node 297 in the side line of the main discharge piping around Y-
direction. The detailed picture of the displacement for this node (Appendix A4-1) shows influences of the 1st
order. The model analysis of the part at node 297 shows resonance for a frequency around 12.882 Hz (15th
order (12.37 Hz). All values
of velocity at this nodes are within the limit.
The highest stress occurs at the node 140 in the safety valve 554 piping for interstage piping before
the interstage cooler EA-45555. The detailed picture of the displacement for this node (Appendix A4-1) shows
influences of the 1st
order. The model analysis of the part at node 145 shows resonance for a frequency around
12.942 Hz (17th
order (12.37
Hz). All values of stress at this nodes are within the limit. The highest reaction force occurs at the node 92 for
the 1st
stage cylinder support near to the 2nd
stage discharge damper Z- direction. The detailed picture of the
displacement for this node (Appendix A4-1) shows influences of the 1st
order. The model analysis of the part at
node 145 shows resonance for a frequency around 12.942 Hz (16th
eigenmode on fig.: 14). This value is close
to the compressor frequency of the 1st
order (12.37 Hz).
An overview of the results is shown in following table 5.1.
Table : 5.1 — Maximum values for the dynamic response analysis

Edition
Many supports are described to be loose. Therefore the horizontal stiffness is extremely low. In reality
the friction between pipe and supports will give an additional stiffness. That means, in reality lower
displacements can be expected.
Plot of the model of calculation
Following are the few plots of the compressor package which is modelled in Ansys.
Fig.: 5.7 — Display of Elements for compressor
package
Fig.: 5.8 — Display of Elements for Compressor
System
Plot of the model of calculation – Boundary Conditions
Following are the few plots of the compressor package with boundary conditions.
Fig.: 5.9 — Boundary Conditions for compressor
package
Fig.: 5.10 — Boundary Conditions with Master Nodes for compressor package
Fig.: 5.11 — Boundary Conditions with Master Nodes
for Compressor System

Edition
Natural frequencies and natural modes of the system
Table : 5.2 — The first 179 eigenmodes (out of 219) of the piping system model
SET Frequency [Hz]
1 1.6074
2 5.1959
3 5.8717
4 5.9192
5 6.2108
6 6.9121
7 6.9535
8 8.6063
9 8.8348
10 9.5087
The first 10 eigenmodes (out of 219) of the piping system model. The shapes of some Eigenmodes are
plotted on the following pages. Some Eigenmodes have to be considered as unrealistic since some boundary
conditions are assumed. Therefore only selected Eigenmodes are plotted.
Plot of the mode shapes for the most important frequencies
Fig.: 5.12 — Plot of the 1st Eigenmode (1.607
Hz)
Fig.: 5.13 — Plot of the 2nd Eigenmode (5.196
Hz)
Plot of dynamic responses at selected nodes
Table : 5.3 - Dynamic displacements Peak-to-Peak in mm
Nodes UX UY UZ ABS
145 0.8888 0.0315 0.3887 0.9706
144 0.788 0.0001 0.3887 0.8786
143 0.6874 0.0295 0.3887 0.7902
142 0.5468 0.062 0.3886 0.6737
297 0.0908 0.61 0.2325 0.6591
298 0.091 0.5822 0.217 0.6279
598 0.3515 0.4971 0.0296 0.6095
577 0.0911 0.5115 0.3093 0.6047
578 0.0931 0.5061 0.3167 0.6043
599 0.3601 0.4822 0.0352 0.6028

Edition
Plot of dynamic displacements at selected nodes Peak-to-Peak in mm
Fig.: 5.14 — Plot of Dynamic Displacements At Selected Nodes Peak-To-Peak in mm
Table : 5.5 - Dynamic velocities RMS in mm/s
Nodes UX UY UZ
297 3.37 28.32 7.47
298 3.38 26.38 6.49
595 13.45 25.64 0.98
318 0 25.64 0
597 14.72 25.64 1.19
598 16.98 25.43 1.63
594 15.85 24.7 1.1
599 17.42 24.67 1.91
145 24.56 0.87 10.66
622 23.96 9.28 1.64

Edition
Plot of dynamic velocities at selected nodes Peak-to-Peak in mm/s
Fig.: 5.15 — Plot of Dynamic Velocities At Selected Nodes Peak-To-Peak in mm/s
Table : 5.6 - Dynamic bending stresses Peak-to-Peak in N/mm2
Node Stress
Elemen
t
140 16.29 129
582 13.26 490
141 13.12 129
144 13.01 134
501 12.96 449
583 12.59 490
581 12.5 489
138 11.77 548
593 11.5 495
361 11.38 585

Pulsation and Vibration Study of Reciprocating Compressor According to API 618 5th
Edition
Plot of dynamic bending stresses at selected nodes Peak-to-Peak in N/mm2
Fig.: 5.16 — Plot of Dynamic Bending Stresses At Selected Nodes Peak-To-Peak in N/mm2
VI. EXPERIMENTATION
Experimental set-up and Instruments
The arrangement of the experimental setup used
for the measurements is shown below
Fig.: 6.1 — Arrangement of Experimental Setup
for Vibration Measurements
Table : 6.1 - List of Instruments used
Sr. No. Instrument / Software Remarks
1
B & K (Denmark) Make 5 channel
PULSE Data Acquisition System
(Analyzer) with BNC Connectors – Type
3560B with LAN Interface
Data Acquisition System
2
B & K (Denmark) Make IEPE
Accelerometer – Type 4514B
Transducer for Vibration
measurement
3
B & K (Denmark) Make Pulse FFT &
CPB Analysis Labshop – Type 7700
Vibration & Noise Analysis
Software
Experimental Procedure
The vibration measurements were carried out at rated load (EOR case) since EOR case is the design
case for this compressor. The Experimental set-up and Instruments details are shown in above chapter VI.
The measurements locations at different Nodes for vibration are shown in chapter V Sketches of Piping for
Vibration Study.
Vibration Measurements:
Only few nodes 297, 595, 318, 145 and 599 are considered for vibration measurements, which are
having very high vibrations. The vibration measured in all three directions (vertical, horizontal & axial) in
accelerations for 1/3rd
Octave center frequency over a range of 0 Hz to 8 KHz.

Edition
Result Analysis and Discussion
Results before Vibration Study (Experimental)
0
10
20
30
40
50
0 5000
Velocity(mm/s)
Freq (Hz)
297 UX
297 UY
297 UZ
0.00
10.00
20.00
30.00
40.00
50.00
0 5000
Velocity(mm/s)
Freq (Hz)
595 UX
595 UY
595 UZ
0.00
10.00
20.00
30.00
40.00
50.00
0 5000
Velocity(mm/s)
Freq (Hz)
318 UX
318 UY
318 UZ
0.00
10.00
20.00
30.00
40.00
50.00
0 5000
Velocity(mm/s)
Freq (Hz)
145 UX
145 UY
145 UZ
0.00
10.00
20.00
30.00
40.00
50.00
0 5000
Velocity(mm/s)
Freq (Hz)
599 UX
599 UY
599 UZ
Nodes UX UY UZ
297 10.09 45.33 13.21
595 18.93 36.47 0.84
318 0.38 42.85 0
145 34.64 2.89 36.15
599 23.84 37.15 0.34
Table : 6.3 – Mean Values Before Vibration Study
The above graphs are the experimental results before Vibration Study means before the modifications
into the support. The graphs shown for the five nodes and found that the velocities are the much higher than
the specified limit i.e. 30 mm/s in API 618. The above table 6.3 is shown the mean values of the graphs.
Results after Vibration Study (Experimental)
0
10
20
30
40
0 5000
Velocity(mm/s)
Freq (Hz)
297 UX
297 UY
297 UZ
0.00
10.00
20.00
30.00
40.00
0 5000
Velocity(mm/s)
Freq (Hz)
595 UX
595 UY
595 UZ
0.00
10.00
20.00
30.00
40.00
0 5000
Velocity(mm/s)
Freq (Hz)
318 UX
318 UY
318 UZ
0.00
10.00
20.00
30.00
40.00
0 5000
Velocity(mm/s)
Freq (Hz)
145 UX
145 UY
145 UZ
0.00
10.00
20.00
30.00
40.00
0 5000
Veklocity(mm/s)
Freq (Hz)
599 UX
599 UY
599 UZ
Nodes UX UY UZ
297 3.21 27.52 6.88
595 12.90 24.08 0.77
318 0 24.08 00
145 24.08 0.77 9.46
599 16.34 16.34 0
Table : 6.3 – Mean Values after Vibration Study

Edition
The above graphs are the experimental results after Vibration Study means after the modifications
into the support. The graphs shown for the five nodes and found that the velocities are the much lower than the
specified limit i.e. 30 mm/s in API 618. The above table 6.4 is shown the mean values of the graphs.
Results after Vibration Study (Analytical)
Table : 6.4 - List of Nodes with Maximum Velocities RMS in mm/s after Pulsation and Vibration Study
(Analytical)
Nodes UX UY UZ
297* 3.26 28.14 7.55
298 3.27 26.24 6.37
595* 13.29 25.29 0.97
318* 0 25.29 0
597 14.54 25.29 1.18
598 16.78 25.08 1.61
145* 24.55 0.87 10.66
594 15.66 24.36 1.09
599* 17.21 24.33 1.89
577 3.28 22.97 10.39
*- nodes are used for comparison with experimental results
The model for this vibration study includes the suction, interstage and discharge side. The model
contains the suction side starting at 6‖-P-551093-B2A1-IT existing header, suction piping, interstage piping
with cooler, KOD and PSV line, the discharge side with PSV-line and with the exiting header 4‖-P-551107-
B4A1-IH40. Relevant sidelines are also included in the calculation.
The natural frequencies could be calculated once for the whole system. For the dynamic response
analysis we choose the deviation of the velocity of sound with the highest pulsations of the worst case. The
details about the worst case are mentioned in chapter V. No change of piping layout is required except orifices
which are already discussed for the pulsation study and also following support should be added into the piping
layout.
 Support : X & Y at Node 345 : Refer Sketch
 Support : X & Z at Node 367 : Refer Sketch
After several calculations with the support modification all vibration results are within the limits.
The supports with its restricted direction of the last version are indicated in the attached isometrics. At piping
parts it was necessary to implement additional supports to achieve satisfied results. The details of all supports
(blocked translations and rotations) in calculation model can be found in the sketches of the system (chapter
V).
VII. CONCLUSION
The investigation of the piping showed no severe problems with vibrations. One reason for this is the
reduced pulsation level which is achieved from the modifications recommended in the pulsation study. The
combination of the acoustic simulation with a mechanical analysis as defined in Design Approach 3 of API
618 is the content of the vibration study. Shaking forces induced by pressure pulsation excite the mechanical
piping system. The vibration study determines the effect on the mechanical piping system and proposes
measures to avoid stresses possibly leading to deformation or rupture by fatigue.
The finite element program ANSYS is used for modeling of the mechanical system. The model is
built of several types of basic elements (e.g. pipes, beams, elbows, T-pieces) connected at node points.
Modifications are proposed to meet agreed criteria of vibration as per API 618 5th
edition 2007).
Since shaking forces are a result of the previous pulsation calculation, vibration study can only be
done in combination with a preceding pulsation study.
Also the accuracy of the analytical solution had been validated by means of experimental results by
using B & K Analyser for the measurement of compressor piping vibration.
REFERENCES
[1.] Shelley Greenfield and Kelly Eberle, New API Standard 618 (5TH
ED.) And Its Impact On Reciprocating
Compressor Package Design, Compressor TechTWO
, 2008, pp. 55-67.
[2.] Paul Alves, Acoustical And Mechanical Analysis Of Reciprocating Compressor Installation - API 618,
Compressor TechTWO
, 2006, pp. 64-69.
[3.] James D. Tison and Kenneth E. Atkins, The New Fifth Edition of API 618 for Reciprocating Compressors –
Which Pulsation and Vibration Control philosophy should you use?, Engineering Dynamics Incorporated Seminar
Manual, San Antonio, Texas, 2008.

Edition
[4.] J. C. Wachel and J. D. Tison, Vibrations In Reciprocating Machinery And Piping Systems, Engineering
Dynamics Incorporated Seminar Manual, San Antonio, Texas, 1999.
[5.] Enzo Giacomelli et al., Forced Response of Cylinder Manifold For Reciprocating Compressor Applications, 8TH
Biennial ASME Conference on Engineering Systems Design and Analysis, Torino, Italy, 2006.
[6.] W. W. von Nimitz , Pulsation And Vibration Control Requirements In The Design Of Reciprocating Compressor
And Pump Installations, International Compressor Engineering Conference, Purdue University, Indiana, USA,
1982.
[7.] Reciprocating Compressors for Petroleum, Chemical, and Gas Industry Services. API Standard 618, Fourth
Edition , June 1995.
[8.] Reciprocating Compressors for Petroleum, Chemical, and Gas Industry Services. API Standard 618, Fifth edition,
December 2007
[9.] http://www.betamachinery.com/
[10.] http://www.tno.nl/
[11.] Ansys Software

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