12TH CONFERENCE on

DYNAMICAL SYSTEMS THEORY AND APPLICATIONS

December 2 5, 2013. ód , Poland

Dynamic analysis of the chemical plant piping with hot hydrocarbons

Andrzej B aszczyk, Adam Papierski, Maciej Rydlewicz, Mariusz Susik

Abstract: This paper presents an analysis of the dynamic state of suction and discharge piping of the process pump in a chemical plant of hydrocarbon mixture at 270 ° C. Analysis was made on the basis of a measured force, exciting on facility in a shape of fast changing pressure of suction and discharge process pumps. In order to calculate the static and dynamic structures there was used a commercial software, Bentley AutoPIPE. Theoretical modal analysis was performed on the 1.5 high-frequency component, with the load of the measured pressure pulsation. Results of theoretical modal analysis have shown that the frequency components which are similar to extortion, their modal shapes are in favor of occurrence of the mechanical resonance, caused by pressure pulsations. In order to verify the numerical vibration, there had been made the spectral analysis on the pipe supports, which has later been compared with numerical computations.

1. Introduction

Pipelines in chemical process plants are being designed on the basis of guidelines from the standard

EN 13480-3 entitled "Metallic industrial piping – Part 3: Design and calculation” [1]. At the design

stage, there normally are considered computations of strength of a structure (pipings, supports,

displacements and creep). Due to the lack of demand by the Office of Technical Inspection, the

dynamic analysis of piping is usually not carried out in the case of installation, in which process

pumps are the impeller type. Dynamic loads result in greater reactions of the structure than within

static loads of the same size. Value of reaction is dependent not only on the value of the imposed

load, but also on its frequency. Vibrations of compressors and pumps have a pattern of sinusoidal

vibration with a constant frequency of excitation. In a single system, there might occur various loads

and usually they are not in phase. Frequency of excitation is usually high. In order to determine the

total dynamic reaction of the structure there should be analyzed all the free vibration frequencies of

the system. Modal analysis indicates characteristics of structures subjected to dynamic loads and its

tendency to oscillate.

2. Description of hydrocarbons plant

Plant pipelines which were the subject of analysis are filled with a mixture of hydrocarbons with a

high temperature of about 270 ° C. Agent is forced through one or two of the three process pumps

installed in parallel. During normal operation, the system is based on two pumps, and the third is the

so-called pump hot reserve. It is also possible during the system start-up, that a single pump is

operating. In the Figure 1 there is visible a layout of suction pipelines of process pumps, Figure 2

shows discharge pipeline and in the Figure 3 the assembly of both systems.

Figure 1. The view of suction pipes modeled in AutoPipe software

Figure 2. The view of discharge pipelines modeled in AutoPipe software

Figure 3. The assembly of discharge and suction pipelines

In the system there have been installed three single-stage pumps with six-bladed, double-suction

impellers. Synchronous frequency of rotations of these pumps engines is 1500 rpm. Flow system of

pumps consists of the centrifugal impeller, the inlet volute casing and the double outlet volute casing.

The rotating assembly of the pump is beared:

• on the opposite drive side, two angular, single-row ball bearings,

• on the drive side, barrel bearing.

In the Figure 1 there has been shown the meridional cross-section of the pump with marked basic

elements.

Figure 4. Meridional cross-section of the process pump

a – Discharge nozzle, b – impeller, c – shaft, d – inlet volute casing, e – double outlet volute

casing

Designers of pipelines usually do not consider the stiffness of pump nozzles, operating in these

systems due to not available data from pump manufacturers. In computations there is anticipated that

pump nozzles are perfectly rigid elements. This anticipation is causing that results of stress

computations in pipeline elements are significantly overestimated. On the other hand, results of

computations of free vibration frequencies, on the basis of modal analysis, are also far from real

values.

On the basis of documentation and measurements of the outline and wall thickness by the use of

"REVERSE ENGINEERING" method there has been made a three-dimensional solid model of the

pump case shown in the Figure 5. Computations by a use of the finite element method with a mesh of

number of nodes equal to 282640.

Discharge Suction

nozzle nozzle

Figure 5. Finite Element Model of process pump

In order to determine the stiffness of pump casing nozzles, there have been carried out the

numerical computation of displacements and angles for both nozzles. The computations were

performed for admissible loads given by the pump manufacturer. Computed values of stiffness are

shown in the table T-1. These values are introduced in the coordinate system consistent with the

standard ISO 13709 (API 610) [2].

Table T-1 Calculated process pump stiffness by FEM method

Pump Nozzle stiffness

Thermal displacement

(22°C - 272°C)

Fx/ x Fz/ z Fy/ y Mx/ x Mz/ z My/ y y z x

N/mm N/mm N/mm Nm/1° Nm/1° Nm/1° mm mm mm

Suction

nozzle

98 887 299 279 1 892 149 1 541 669 289 442 1 544 873 3,86 -1,52 0

Discharge

nozzle

74 811 189 241 1 742 274 579 934 155 973 414 507 3,86 1.48 0

3. Modal analysis

Modal analysis is commonly applied in practice by the use of technique of dynamic characteristics

examination technique for mechanical objects. As a result of the modal analysis there is obtained a

model as a set of natural frequencies and their shapes.

Knowledge of these parameters allows for prediction of the behavior of the object as a result of

any imbalance. It is used for the purpose of modifying the structure, structural health diagnosis, active

vibration reduction and for the purposes of verification and validation of numerical models such as

finite element or boundary elements method.

In order to identify the excitation frequency there have been carried out measurements of satic

pressure at the outlet and the inlet of the pump.

Measurements were done by the use of fast changing pressure probes, resistant to high

temperatures. These probes were assembled directly to the pipeline at the suction and discharge of the

pump. Collecting the data was carried out continuously, and the figures show only a few seconds

sample form the measurement time. Sampling frequency was 2 kHz. Measurement card had the

installed anti-aliasing filter, which value was set to 500 Hz.

Exemplary plot of pressure measured on site, at the time of start-up of the pump, is shown in the

Figure 6 and in a stable state is shown in the Figure 7.

Figure 6. Suction pressure (blue line) and discharge pressure (red line) of process pump measured at

pump start-up

0

500

1000

1500

2000

2500

3000

0 1 2 3 4 5 6 7 8 9 10 11

pt,ps[kPa]

t[s]

Because of the complexity of the model and resulting from that fact, high number of frequencies

and shapes of vibrations, there are introduced only these shapes and frequencies of natural vibrations

close to 150 Hz, which is the main system excitation frequency, caused by a sixth harmonic excited

by a flow of the fluid agent through the pump.

Suction nozzle Discharge nozzle

Figure 7. Suction pressure (left) and discharge pressure (right) measured at pump A during normal

operating time

In following figures there have been shown the chosen frequencies and their shapes for basic

operation variants.

The introduced natural vibration frequencies of the system were determined with the accuracy

resulting from an adopted computational model (linear model). Real facility might be excited with

every of the shown below frequencies of modal vibrations. For verification of numerical

computations there have been used the results of vibration measurements at supports, which are

shown in the next chapter.

3.1. Experiment verification

For every of the operation variants, there has been identified from 6 to 7 natural vibration

frequencies, which values held in the scope from 145 to 155 Hz. Values for every of the variants are

set together in the table T-2.

Figure 8 shows the exemplary impeller with a natural vibration frequency of 150.65 Hz, obtained for

the operation variant in which pumps A and B are running and pump C is standby.

Table T-2

Variant of pumps operation free vibration frequency

Pump A Pump B Pump C

running running standby 144.98, 146.95, 148.02, 150.65, 153.25, 155.00

running standby running 145.14, 146.49, 150.73, 152.56, 154.95, 155.00

standby running running 145.24, 146.67, 147.95, 148.67, 150.80, 152.56, 154.96

For computations there have been adopted the following pipelines temperatures:

a) 272OC - pipelines temperature, which hold a flowing agent, b) 150OC - pipelines temperature with a pump standby, c) 5OC - pipelines temperature which do not hold a flowing agent

Figure 8. Free vibrations shape of discharge pipelines. Pumps A and B running and pump C standby

Measurements of vibrations in pipelines were made by a purpose of verification. The question

was, which from the shapes of natural vibrations of pipelines, is being excited in a real facility. Full

verification of shape of vibrations could not have been carried out due to the fact that pipelines were

thermally insulated and the measurements could not have been performed in positions where the

deflections or deflection nodes occurred. Measurements could have been carried out in positions

where the access to a pipeline material was available, that is the supports. Despite the above, these

measurements helped to identify the excited shapes of pipelines vibrations. In the future, these results

will serve with modifications of pipelines supports such that they could be offset from the forced

vibration of the pressure pulsation. Exemplary results of vibration, together with the position of

measurement are shown in Figures 9 and 10.

Figure 9. Measuring position of vibrations at discharge pipelines

Figure 10. Measuring position of vibrations at suction pipelines

4. Conclusions

Consideration of stiffness of process pumps nozzles, helped to precisely identify shapes and

frequencies of natural vibrations of pipeline systems. Measured pressure pulsation values (the

exciting force of vibrations in the system) allowed for their consideration in strength computations

and carrying out the so-called acoustic analysis which has been described in a different article.

In order to reduce vibration in the analyzed systems, it is essential to strive to reduce the pressure

pulsations as the source of exciting force. This can be achieved by the change of the process itself or

building in special pressure pulsation dampers. A further possibility is to change the length of the

pipelines and supports so that they could offset the vibration frequency which enforces. The latter

method is difficult to implement due to the lack of space for installation.

References

1. The European Standard EN 13480-3:2002 "Metallic industrial piping – Part 3: Design and calculation”

2. API Standard 610 Tenth Edition “Centrifugal pumps for petroleum, petrochemical and natural gas industries”

3. aszczyk A.., Najdecki S., Papierski A., Kunicki R., Susik M.: Ekspertyza techniczna orurowania ssawnego i t ocznego pomp 1 P3 A/B/C zlokalizowanych na instalacji – zadanie inwestycyjne 11327 pn. Budowa instalacji do produkcji Paraksylenu – Instalacja podstawowa.

ód , Instytut Maszyn Przep ywowych 2011.

Andrzej B aszczyk, Professor: Lodz University of Technology, Institute of Turbomachinery, Wólcza ska 219/223 Street, 90-924 ód , Poland (andrzej.blaszczyk@p.lodz.pl).

Adam Papierski, D.Sc, Ph.D.: Lodz University of Technology, Institute of Turbomachinery, Wólcza ska 219/223 Street, 90-924 ód , Poland, the author gave a presentation of this paper during one of the conference sessions (adam.papierski@p.lodz.pl).

Mariusz Susik, Ph.D.: Lodz University of Technology, Institute of Turbomachinery, Wólcza ska 219/223 Street, 90-924 ód , Poland, (mariusz.susik@p.lodz.pl).

Maciej Rydlewicz, Ph.D.: System Center Softdesk, Lodowa 101 Street, 93-232 ód , Poland, (maciej.rydlewicz@softdesk.pl).