Functional analysis and modelling of a dynamic seal ...

Functional analysis and modelling of a dynamic seal component for a reciprocating gas
compressor
Master of Science Thesis TRITA-ITM-EX 2018:685 KTH Industrial Engineering and Management
Machine Design SE-100 44 STOCKHOLM
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Master of Science Thesis TRITA-ITM-EX 2018:685
Functional analysis and modelling of a dynamic seal component for a reciprocating gas compressor
Chandramouli Suryanarayanan
Approved 2018-10-03
Contact person Andreas Söderberg
Abstract The rod packing seal in a reciprocating compressor plays a vital role to reduce the leakage of highly pressurized gas along the piston rod. Reciprocating compressor in the natural gas transmission and process industry is one of the crucial applications where the Nextseal technology can be put to greater use for the reduction of green-house gas emissions. Nextseal technology revolves around the notion of pressure balancing of seals or rings.
The purpose of this master thesis is to understand how the design of the novel rod-packing seals/rings behaves in correlation with the concept of pressure balancing. The work also includes developing a model, both CFD and analytical to obtain a relationship between fluid (viscous oil) pressure and displacement of the dynamic component. Functional analysis of the dynamic system is conducted by numerical simulation using MATLAB. Major parameters influencing the dynamic behaviour are identified at the beginning.
The scope for CFD model is defined and the developed method is used to obtain correlation between hydraulic fluid pressure and displacement of the dynamic component. A derived analytical model is solved, and the results are compared and validated. The validated correlation is employed to solve the dynamic system numerically and the results are analyzed. From this numerical method the effect of friction force and geometry of the dynamic component on pressure difference and displacement of the dynamic component is well analyzed and discussed in this thesis. From the influence of friction force on pressure difference study, a linear relation is observed. Also, by changing the geometry (chamfer length and angle of the dynamic component) of the dynamic component, it can be observed that design configuration with 60° chamfer angle gives smaller pressure difference value compared to the original design. Thus, the model developed can be used to obtain results for pressure difference, displacement, and to study the effects of friction, geometry and mass.
Keywords: Rod packing dynamic seals, Computational Fluid Dynamics, Pressure balance, Numerical analysis
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Chandramouli Suryanarayanan
Godkänt 2018-10-03
Sammanfattning Stångförpackningstätningen i en fram- och återgående kompressor spelar en viktig roll för att minska läckaget av gas under högt tryck längs kolvstången. Kolvkompressorer är viktiga komponeter för naturgasöverföring- och inom processindustrin. Teknik utvecklad av Nextseal kan potentiellt användas för att tryckbalansera tätningar eller ringar och därmed minska utsläpen växthusgas.
Syftet med detta examensarbete är att skapa kunskap kring hur utformningen av de nya stavförpackningstätningarna/ringarna påverkar tryckbalanseringen. Arbetet innefattar också modellutveckling, både av CFD-modeller och analytiska modeller för att analysera relationen mellan vätsketryck (viskös olja) och förskjutning av den dynamiska komponenten. Funktionsanalys av det dynamiska systemet utförs med numeriska simuleringar och experiment. Viktiga parametrar som påverkar det dynamiska beteendet identifieras i början.
Den utvecklade metoden och CFD-modellen används för att studera korrelationen mellan hydraulvätsketryck och förskjutning av den dynamiska komponenten. En härledd analytisk modell löses numeriskt och resultaten jämförs och valideras. Den validerade modellen används för att simulera det dynamiska systemet numeriskt och resultaten analyseras. Effekten av friktionskraft och geometri hos den dynamiska komponenten på tryckskillnad och förskjutning av den dynamiska komponenten analyseras och diskuteras i avhandlingen. Det visas att friktionskraften har en linjär effekt på tryckdifferensen.
Genom att ändra geometrin (fasens längd och vinkeln på den dynamiska komponenten) observeras att konstruktionskonfigurationen med 60° fasvinkel ger en lägre tryckskillnad jämfört med den ursprungliga konstruktionen. Således kan den utvecklade modellen användas för att studera relationerna mellan tryckskillnad och förskjutning och för att studera effekterna av friktion, geometri och tröghetsmassa.
Nyckelord: dynamiska tätningar, CFD, tryckbalans, numerisk analys
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Foreword
This thesis work was carried out over a period of eight months at Nextseal AB, Stockholm and KTH Royal Institute of Technology at the department of Machine Design. It has been a great experience and immense pleasure working with a novel technology and I would like to thank Nextseal AB and KTH, both the organization for providing various supports throughout my thesis work.
I would first like to thank my thesis supervisor Andreas Söderberg, CEO of Nextseal AB for providing me this opportunity to work with an interesting and challenging technology and also provided all the support to keep going forward with my thesis work. I would also like to thank Bengt Adolfsson, board member and senior consulting member of the firm, for providing valuable inputs for many problems that I faced with my experimental test setup. I would like to thank Tomas Östberg. He was very helpful with manufacturing of different parts required for the test-rig and also gave valuable suggestions related to design for manufacturing. I cannot thank them enough. I wish to express my sincere gratitude to my academic supervisor Stefan Björklund, who has always been available whenever I faced technical problems and has provided with valuable perspectives and solutions. He has continuously guided me to the right track. I would also like to thank Ulf Sellgren, my examiner at KTH, for providing timely support with the thesis and giving his insights whenever needed.
Finally, I would like to convey my sincere regards to my parents, sister and friends for continuous motivation and support throughout my years of education and through the process of writing this thesis.
Chandramouli Suryanarayanan Stockholm, Sweden
2-D Two Dimensional
FE Finite Element
PTFE Polytetrafluoroethylene
Abstract i
Sammanfattning iii
Foreword v
Nomenclature vii
1 Introduction 1 1.1 Background . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 1.2 Purpose . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3 1.3 Delimitations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3 1.4 Methodology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4
2 Frame of Reference 7 2.1 Rod Packing Technology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7 2.1.1 Packing Cups . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8 2.1.2 Packing rings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8 2.1.3 Flange . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9 2.2 Packing ring Designs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9 2.2.1 Single Acting Rings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10 2.2.2 Double Acting Rings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11 2.2.3 Oil Wiper Rings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12 2.3 Viscous Fluid Flows . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13 2.3.1 Flow Regimes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14 2.3.2 Flow through channels . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15 2.4 Governing Equations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16 2.4.1 Navier-Stokes Equation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16 2.4.2 Hagen-Poiseuille flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17 2.5 Analytical Background . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18
3 Methodology 19 3.1 Parameter Identification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19 3.2 Experimental Testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21 3.2.1 Test-Rig . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21 3.2.2 Design and Construction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23 3.2.3 Test Objective . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24 3.2.4 Instrumentation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25 3.2.5 Data Acquisition . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26 3.2.6 Drives and Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27
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3.3 Computational Analysis of the Dynamic System . . . . . . . . . . . . . . . . . . . . . . . . 28 3.3.1 Scope and Method Development for CFD . . . . . . . . . . . . . . . . . . . . . . . . 28 3.3.2 Fluid Model . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30 3.3.3 Meshing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30 3.3.4 Setting up Physics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31 3.3.5 Solving and Post-Processing. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32 3.3.6 Analytical model . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33 3.3.7 Geometry Design configurations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34 4 Results and Discussions 37 4.1 CFD Simulation of liquid domain . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37 4.1.1 Validation with analytical model . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38 4.2 Dynamic Response analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 39 4.2.1 Time-Domain Analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 39 4.2.2 Performance analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 41 4.3 Experimental Result . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 45
5 Conclusions 47
Introduction
The second largest source of energy in the United States is natural gas. The gas from both onshore and offshore sources, after refining process, is transported to supply households and industries all over the country where it is consumed for heat and electricity generation. The United States natural gas industry has undergone change of high magnitude and pace for the past decade. Natural gas production in US increased 33 percent between 2005 and 2013 [1]. Gas demand for power generation has grown from 15.8 billion cubic feet per day (Bcf/d) in 2005 to 22.2 Bcf/d in 2013 [1]. Natural gas pipeline network has been developed to transport natural gas, both interstate and intrastate, from production and storage areas to distribution systems and end users. As of 2007 there are 210 pipeline systems with 305,000 miles of high- pressure transmission [2]. Additionally, 4000 miles of interstate transmission pipelines are constructed from 2008-2013. The transmission network embraces more than 1,400 compressor stations to perpetuate high pressure in the pipelines. Methane is a potent green- house gas. Even with a short atmospheric lifetime of 10-12 years, methane is considered to be 50 times more effective than C2 at trapping heat in the atmosphere. This chapter covers essential background of this project, project scope, formulation of research question, methodology pursued and delimitations.
1.1 Background
The mitigation of methane gas emission is of major concern, as the Environmental Protection Agency (EPA) has stringent policies [3]. There are many ongoing researches on various compressor components for better performance by various compressor OEM’s and several other contributing players in oil and gas industry. Nextseal AB, a Swedish based company has an interesting solution for rod packing seals that could revolutionize this issue.
Reciprocating compressors are commonly and widely used compressors to build pressure to transport gas in a pipe. For transporting natural gas, there is a compressor station for every 100 kilometers. There are leakages from pipelines, joints and various compressor components, like valve, rod seals, and other fittings. From Transportation, storage and distribution, EPA has stated that most methane emission is from reciprocating compressors that account for over 1 to 6 standard cubic meters per hour (scm/h) [4] for large and high- pressure compressors. The largest contribution is from rod packing rings/seals in the compressor.
Current technology as shown in figure 1.1, uses sets of specifically-cut, dry-ring seals held in place with springs and cups. There is a trade-off between leakage reduction and friction in today’s technology. In the reciprocating compressors as piston moves back and forth, the pressure differential across the packing rings/seals creates a twisting effect that causes natural gas to leak into the casing. Ring/seal twisting also causes increased friction and wear to the sealing rings and piston rod. The gas leaking from rod packing case is vented to atmosphere.
Figure 1.1 Reciprocating compressor with existing rod packing seal technology Source: Image from U.S EPA 2006a
Nextseal technology revolves around the notion of pressure balancing of seals. A schematic representation of pressure balancing concept is shown in figure 1.2 below. It takes the concept of liquid sealing and combines it with a novel, patented arrangement for pressure balancing across a seal arrangement (Patent No: US 7,757,599 B2 [5]). The pressure balancing has two appealing characteristics. First, it stops the gas from leaking by letting the seal act as a separator of gas and liquid with the same pressure. Second, it reduces the friction caused by the seal running against the opposing surface. This results in both less leakage and less friction which will be a huge advantage over current solution for dynamic seals. This has several applications like waste heat recovery system- steam expanders, compressors in food industry, reciprocating compressors for gas transmission and so on.
Figure 1.2 Schematic representation of pressure balancing concept Source: Nextseal
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1.2 Purpose The main purpose of the thesis is to understand how the design of the rod-packing seals/rings behaves in correlation with the proof of concept (pressure balancing) through experimental and numerical analysis. Hereafter, rod-packing seals/rings will be referred as dynamic component in this report. Behaviour of the dynamic component means displacement and velocity based on the pressure acting on both sides of the dynamic component and how both the pressure curves behave. The test-rig almost reproduces the operating conditions of the real time application (reciprocating compressors of natural gas transmission). The objective also includes studying the effects of friction force and geometry changes on the dynamic response of the system.
Tools used to achieve the results are ANSYS Fluent and MATLAB. CFD tool is used to develop the relationship between fluid pressure and displacement of the dynamic component, which then can be used to analyze the dynamics and functioning of the system. Defining the scope of the thesis led to the formulation of following three research questions.
1. Can the behaviour of the dynamic component be understood through numerical analysis?
2. What is the relation between friction force and pressure difference over the dynamic component?
3. Does the geometry of the dynamic component influence the functioning of the system?
1.3 Delimitations
• Analysis of the sealing ring design will not be of focus as standard packing rings from leading market will be used.
• Product design influencing production cost is not included in this project work.
• Tests are undertaken for one configuration of fluid flow channel and design of dynamic component
• Parameters like dynamic component geometry, friction force that have an effect on the system functioning will be studied using the analytical model but does not include the effects of component mass, pipe geometry and flow rate.
1.4 Methodology The methodology acquired to achieve the above mentioned objectives are provided in the form of flow-chart as shown in figure 1.3 below. A short description on each context is published in this sub-topic.
Figure 1.3: Methodology flowchart
Literature study -
A comprehensive literature study is conducted to understand the basics of fluid dynamics & mechanics, mathematical equations, fluid-structure interactions, planning tests and how to implement CFD tool. It also includes learning about the existing rod packing seal technology in the market and the area of research being conducted in this field. Manuals of control drives (Donfoss frequency converter, Bosch Rexroth drive) are also learned as an outcome to set-up and operate them effectively.
Defining the Scope -
The scopes as defined in section 1.2 are vital, as they should be developed based on time constraint and knowledge limitations. Accordingly, research questions are formulated.
Identify system parameters -
The behaviour of the dynamic component will be studied and is affected by various parameters.
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• Flow rate of hydraulic fluid • Mass of the dynamic component • Friction acting on the dynamic component • Speed of the piston rod • Tube length of hydraulic fluid flow • Geometry of the dynamic component
Modify test-rig –
The test rig contains a motor, a modified Stirling single cylinder marine engine, test frame, a lubrication oil pump, and a Bosch-Rexroth hydraulic pump. A double-acting piston cylinder is designed and manufactured for testing. Additional components like control-volume adjuster, adapters, and cooling jackets are designed and manufactured. The testing reproduces the operating conditions as that of in a real time application.
Set-up control drives and instrumentation –
To set up control drives is an integral part of testing for remote access to operate the drives (hydraulic pump, motor). Setting appropriate parameter values for example, motor speed, pump flow-rate to the drives using the designated software is essential. The test-rig is instrumented at the desired location and data are collected using the data acquisition system.
Dynamic testing –
• The dynamic component is tested with the modified test rig • Assembling the double acting piston cylinder with appropriate instrumentation and
attachments • Setting up instrumentation and drives for remote access • Conducting tests and recording the outcomes
Simplified FE model and CFD analysis –
The dynamic component to be tested experimentally can be studied and analysed using any of the CFD tool commercially available. In this project ANSYS Fluent, is utilized for computational fluid dynamics. The fluid model of the dynamic system is developed using SolidEdge and is imported to ANSYS workbench. Meshing of an axisymmetric 2-D fluid model is performed with ANSYS mesh modeler. Pre and post processing are conducted with Fluent.
Analytical model –
An analytical model that represents the CFD model is formulated and its results are validated. The equation of motion which represents the dynamic system of the model (dynamic component) is derived. MATLAB numerical solver for ordinary differential equations is used to solve the differential equation. Result from validated analytical/CFD model will be used to solve the mathematical model.
Analyse the results –
• Validating the results obtained from CFD model with the analytical model. • Analyzing the dynamics and functioning of the component with computational
methods. • Varying the system parameters and studying its effect on the dynamic behaviour of the
model.
Conclusion and future work –
A brief discussion on various findings from this project will be reported. It will also describe the validity of the CFD and analytical model. Proposing possible improvements as future works based on the results obtained.
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Chapter 2
Frame of reference This chapter delineates the necessary details about the existing technology for basic understanding about rod packing rings and the background of technical knowledge obtained. The chapter provides information about the rod packing technology and ring design and continues with technical description on fluid dynamics. The governing equation behind numerical fluid dynamic analysis, namely Navier-Stokes equation for fluid motion is discussed.
2.1 Rod packing technology In reciprocating process gas compressors, one of the most critical technologies is rod pack seals. Based on the application, and the operating condition of the reciprocating compressors, the design and size of rod packing seals varies accordingly. When compressing gas at varying pressure (suction pressure to discharge pressure) using a reciprocating type piston compressor, there will be leakage along the piston rod due to the clearance between the cylinder and the rod [6]. The main function of the rod seals is to reduce the leakage rate. Rod seals are also called mechanical packings and their nomenclature is shown in figure 2.1. Mechanical packings comprise of three major components.
1. Packing case / cups 2. Packing rings 3. flange
Figure 2.1 Nomenclature of mechanical packing, Source: Image taken from CPI- Mechanical packing ‘design and theory of operation’ [6].
The design of mechanical packings depends on the following parameters [7]
• Lubricated or non-lubricated system • Operating pressure range • Environmental standards • Nature of gas being compressed • Rod size, stroke length and speed • Bolt load requirements • Project cost limits and timing restraints
The design should meet the requirements described in the API 618 standards for reciprocating compressors.
2.1.1 Packing cups
The packing cups which retain the packings rings are assembled back to back closing the rings in a confined space. Henceforth they are also referred as retainers [6]. The cups are machined to provide certain amount of radial clearance around the piston rod. This is to corroborate that the piston rod will have no contact with the packing retainers; if any lateral movement of the rod occurs when running. The radial clearance value is influenced by the rod diameter, operating pressure and the type of compressor. The packing ring set is held between the cups and some side clearance in the cup allocates for radial movement inside the cup. Since, the sealing rings can float around within the cup, this type of mechanical packing is termed as ‘floating’.
Packing cup faces are ground and lapped depending on the pressure and gas being used. Manufacturing of the component needs to be precise as per the drawings and quality check insures that faces are both flat and parallel such that it is assembled properly, and surfaces of the cup are perpendicular to the rod. They are precision machined with pertinent grade of material so that it meets the performance requirements [7]. Commonly used materials are carbon steel, alloy steel, stainless steel, cast iron, or bronze.
2.1.2 Packing rings
Pressure packing rings serves as crucial component in reciprocating compressor, which offers a dynamic, mechanical seal around piston rod and against the sealing surface of the packing cups to prevent leakage from the cylinder. The rings which are widely used till now are the design patented in late nineteenth century by A.W.France [8].
Different types of cut rings are assembled and held together within a cup which constitute a ring set. The cut rings are held together with a garter spring and is referred as segmental packing ring. Similarly, different ring sets are used in a rod packing technology based on the application. In section 2.2, a detailed discussion about different ring designs and existing ring pairs is presented.
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Figure 2.2 Pressure packing rings Source: Image taken from Compressor Products International – “CPI Packing Rings How to Install”
The common features of pressure packing rings are as follows [7]:
• Pressure loads the rings around the rod and against the sealing surface of the packing cup
• The compressed gas fills the gap between the cup and the rings and leaves as the rod reciprocates. In this gap (radial clearance) rings can float.
• Gas leakage is blocked by ring overlap. • End gaps allow the rings to self-adjust for wear which extends its life-time.
2.1.3 Flange
The flange is the end part of the mechanical rod packing. The packing cups, ring sets and flange are usually connected via threaded rods commonly called as tie rods as shown in figure 2.1. The packing rings assembly is bolted with the compressor cylinder through the flange. A wide range of flange sizes and number of bolts are available, based on the application, and operating conditions. Lube and vent connections are machined on the flange. Lube connection is furnished based on lubricated / non-lubricated type of packing rings.
2.2 Packing ring designs This section introduces common types of rod packing ring and ring sets. Different compressor duties require unique packing ring designs and suitable material of the rings. Other factors like gas properties, operating pressure, and compressor speed play a vital role in determining the appropriate combination of ring style [7].
Single acting rings, double acting rings and oil wiper rings are variety of segmented ring types and ring pairs which will be elucidated in the following sub-sections.
2.2.1 Single acting rings
As the name implies, single acting ring seals gas on one-side only. These rings prevent pressurized gas from being enclosed between the ring sets, during the suction stroke of the reciprocating compressor. Most common types in the market are enumerated as follows and refer figure 2.3 shown below.
Pressure breakers
Pressure breakers are usually installed in the first cup of the packing assembly. On one side of the pressure breaker is etched with a letter and that side faces high pressure. It throttles the gas pressure pulsations and do not seal them off. Hence it should be placed in the high-pressure side. It is typically required for applications where the pressure difference between the suction and discharge is over 20 bar [7].
Radial Tangent pair
Radial Tangent pair is a pair of 2 differently cut ring types. One is radially cut and is called radial ring and the other one is tangentially cut and hence tangential ring. These two are paired together and it provides sealing along the piston rod and against the sealing face of the next packing cup. It is a fundamental sealing element in mechanical rod packing. These two rings are doweled together, where the radial cut ring faces pressure which is followed by the tangential ring. A dowel pin in the tangential ring prevents rotation of one ring with respect to the other. It is manufactured for a wide variety of materials based on application.
Balanced-cap rings
Balanced cap rings are a four segmented radial cut ring held together around the rod with a spring. It comprises two caps (upper & lower) which bridges 2 side segments. This design produces a balanced pressure breakdown that does not diminish sealing capacity. The compact design provides advantages such as less frictional heat, easy to install and low leakage. It can be used for applications with high temperature, high load and can also be used for a wide range of piston speeds.
Backup rings
The main task of backup rings is to prevent the extrusion of non-metallic sealing rings into the clearance between the piston rod and packing cup. The bore diameter of the backup ring is slightly bigger than the sealing rings. Hence it does not seal on the rod. It is combined usually with a radial tangent pair. Then the backup ring will seal against the packing cup face. They are quintessentially manufactured using cast iron or bronze material.
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Figure 2.3 Single acting rod packing rings Source: Image taken from Hoerbiger, “Ring and packing”, Sealing systems for reciprocating compressors [7]
Side loaded rings
Side loaded rings comprises of three ring types, which are placed one behind the other, as mentioned below following the same order (from the pressure side)
1. Radial cut ring with chamfered recess 2. Radial cut ring with chamfered boss 3. Tangent cut ring
It provides effective sealing when pressure is below 4 bars. Hence, it is usually located in the low-pressure end of packing or behind the vent line.
2.2.2 Double acting rings
Double acting rings are commonly used at low pressures and as vent seals. They seal gas in both directions. Two types of double acting rings as shown in figure 2.4 are described below:
• Tangent pair • Double side loaded pressure rings
Figure 2.4 Double acting rod pressure packing rings Source: Image taken from Hoerbiger, “Ring and packing”,
Sealing systems for reciprocating compressors [7]
Tangent pair
Tangent pair is a common double acting ring configuration. The ring set incorporates 2 tangent rings pinned together so that the gaps do not align. Both the rings seal against the metallic cup faces and with the piston rod.
Double side loaded pressure rings
Double side loaded pressure ring sets uses two pairs of side loaded pressure rings with the tangent cut rings faces the low pressure. It is usually used in compressors with purge system. Purge gas usually used is nitrogen due to its inert nature.
2.2.3 Oil wiper rings
As the name suggest, oil wiper rings scrape off the oil from the piston rod, so that oil does not travel along the piston rod further away from the mechanical packing case. These wiper ring(s) as shown in figure 2.5 are held within oil scraper cup, which is similar to other packing cups. Oil scrapper cup with rings are usually placed after the flange. The oil wiper cup is either separated from the mechanical packing case or attached with the flange at the end. It is used when a lubricated type mechanical packing seals are installed in the compressor.
Tangent cut wiper
Tangent cut wipers have two scraping edges on the inner diameter of the ring, drilled internally with drain holes in the radially outward direction. Tangent cut wiper ring comprises of three segments, where the cut on each segment is tangential to the piston rod and these segments are held together with a spring. The front face of the ring is machined with drainage slots. The scrapped off oil along the piston rod is channeled radially outwards through these
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drainage slots. For double-acting rings, drainage slots are machined on both faces, but holes only on the front face
Figure 2.5 Oil Wiper rings Source: Image taken from Hoerbiger, “Ring and packing”, Sealing systems for
reciprocating compressors [7]
Radial High-volume wiper (HVOL)
Radial High-volume wiper constitutes two radial cut rings with a drainage hole milled on one face of the ring. The two rings are joined with each other through a dowel pin. It helps to suppress the rotation of rings and eradicate the issue of gaps from aligning. Rings are manufactured with variety of metallic and non-metallic materials, the most common is bronze.
Radial wiper rings set
The set of radial wiper rings provides effective oil wiping action. Usually it includes 3 radial wiper rings that are attached behind on another with dowel pin. The construction of the rings is almost similar to that of tangential wiper rings, except the fact that the cut is radial to the rod. In these rings, channelization of the scrapped oil is through the drainage slot. However, they do not contain oil drainage holes on the face of the ring. It is usually made of cast iron or bronze.
2.3 Viscous fluid flows It is indispensable to have background knowledge about fluid mechanics and dynamics before solving the problem description as stated in section 1.2, of chapter 1. Viscous flow of fluid [9] is an important topic. First, it is pre-requisite to understand fluid flow regimes, flow through channels and losses indulged, to design piping line connections for the test-rig to administer the experiments. The thesis work comprises the use of both compressible and incompressible fluids, but the methodology acquires the comprehensive understanding of incompressible viscous fluid flow to solve the dynamic system using numerical and experimental methods.
2.3.1 Flow regimes
The flow of fluid which can be an internal or external flow through a circular or non-circular channel, it is important to note that flow can be either laminar or turbulent in nature. The intermediate regime between these two flows is called transition. Now let’s review about various flow regimes in brief.
When fluid flows in a medium and the layers of fluid film are parallel to each other; then it is known as streamline / laminar flow. In other words, there is no disruption or lateral mixing (mixing at right angles to the flow direction) between the layers during the flow. In laminar flow, when examined microscopically, the fluid particles move in orderly path in straight lines parallel to the pipe walls. Lateral mixing is attributable to the action of diffusion of fluid layers. In general for laminar flow, diffusive mixing is slow. However, if diameter/ perimeter of the channel is small, then the diffusive mixing can be significant.
The turbulent flow regime is signalized by swift property changes. It means there will be rapid variation of velocity and pressure in space and time for turbulent flow. In contradiction to laminar flow, the flow is highly susceptible to diffusive mixing. It is hard to measure the mean velocity or pressure, since it requires highly sensitive instruments which complicates the process. Hot-wire anemometer or a piezoelectric pressure transducer can be used to measure the turbulence.
When the smooth and steady flow terminates, and becomes fluctuating and agitated, then the changeover phase is called transition. Transition depends on many effects [9], for example, wall roughness, fluctuations in the inlet stream or Reynolds number.The Reynolds number is the primary factor influencing the flow regimes. Hence, it is important to know the Reynolds number.
Reynolds number
In 1883, Irish scientist Osborne Reynolds discovered the number that predicts fluid flow based on static and dynamic properties like flow velocity, dynamic viscosity and density. It can be defined as the ratio of inertial force and viscous force. As the name indicates, it is a dimensionless number, which basically indicates whether the flow past an object or in a duct is steady or turbulent. If the flow is laminar, then the viscous force dominates over inertial force and vice-versa for turbulent condition. Thus in certain application where turbulence pertains, the fluid is inviscid. For internal fluid flow, the Reynolds number is computed by the following formula
Re = ρ∗v∗d µ
(2.1)
Where is the density (fluid property), v is the fluid velocity (flow property), is the dynamic viscosity (fluid property) and d is the diameter of pipe (geometrical property). For external flow L replaces d which is also a geometrical property. The critical Reynolds number is a value which specifies the transition of fluid flow from laminar to turbulent, diversifies
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regarding type of flow and geometry. Thus the applicability of the Reynolds number varies depending on the enumeration of fluid flow like alteration of density, variation of viscosity, flow being internal or external and so on. As an example, for internal flow of fluids in a pipe, transition occurs as indicated by critical Reynolds number of 2300.
2.3.2 Flow through channels
A hydraulic fluid (SAE 15W-40) is used as pressure balancing fluid for the dynamic component in the novel rod packing seal technology. The fluid flows between the pump and the component through steel tubes. Hence it is vital to perceive internal flow of viscous fluid through pipes.
Head loss or pressure drop prevails in a piping system and there are several reasons for this [10]. Forcing the fluid through a channel or pipe fittings consumes energy which results in a pressure drop along the channel and fittings. It is caused by elevation of the pipeline, shaft work, friction along the walls and turbulence provoked by abrupt changes in direction or cross-sectional area. For a fully developed pipe flow, the following relation for head loss is obtained from steady flow energy equation.
= + ∗
(2.2)
The pressure drop across a horizontal pipe can be computed by the equation given below:
∇ = ∗ ∗ (2.3)
Where is the head loss, is the height, is the pressure drop, is the density and is the acceleration due to gravity. The amount of kinetic energy contained in a stream of fluid is the velocity head [11]. Velocity head can be alternatively stated as the amount of potential energy necessary to drive a fluid to its flowing velocity. Thus the potential energy is converted to kinetic energy. The velocity of the stream is used to compute the velocity head, , as given below.
= 2
2∗ (2.4)
For pipes with fittings like pipe couplings, reducers, the excess loss in a fitting is indicated by a dimensionless ‘K-factor’.
= ∗ 2
2∗ (2.5)
Head loss due to friction in pipes and fittings should also be considered when computing the pressure drop across the flow. Thus to calculate the total head loss accounting in the piping system due to turbulence and friction, add to each sum of K factors the friction loss and multiply the sum by the velocity head.
= ( Σ + ∗
)( 2
2∗) (2.6)
By combining equations (2.6) with (2.3), the pressure drop in the piping system is given as follows
= Σ + ∗ ∗ ∗ 2
2 (2.7)
Friction factor
Friction factor f is a dimensionless parameter, which represents the effect of friction between the fluid and walls in a pipe flow. It can be used to compute the pressure drop and head loss in a flow. It primarily depends on the velocity v, diameter D, density , viscosity , and wall roughness ε. Since friction factor is dimensionless, the parameters it depends on should be in dimensionless form. Thus friction factor is a function of Reynolds number and relative roughness as given below.
= (,
) (2.8)
Friction factor as a function of Reynolds number and relative roughness is formulated by various scientists and experts, some are implicit and explicit relations. The most commonly used for flow through commercial pipes is based on Darcy’s friction factor or moody chart. It is based on the Colebrook-White equation. There are many variants of Colebrook-White equation which can be solved explicitly. For laminar flow through smooth pipes, the wall roughness tends to be zero and hence the friction factor depends on Reynolds number only and is inversely proportional to it as given in equation 2.9. Friction factor for turbulent flow of fluid in both smooth and rough pipes can be obtained either from the Moody diagram or by computing Colebrook equation.
= 64
2.4 Governing Equations Pivotal topic like Navier-Stokes equation, Hagen-Poiseuille flow will be briefed in the following sub-sections. The topics discussed in this chapter will also support background knowledge for numerical simulations carried out with ANSYS Fluent, a computational fluid dynamic tool.
2.4.1 Navier-Stokes equation
The motion of a fluid is basically described by the Navier-Stokes system of equations [9]. Application of three laws of conservation namely, (1) conservation of mass, (2) conservation of linear momentum and (3) energy conservation are applied to derive the system of
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equations. The law of conservation of mass and linear momentum are expounded in this section, as it is of major significance for this project to evaluate the pressure and velocity of flow.
Conservation of mass is often known as continuity relation, denotes that the fluid mass cannot change. It is basically a differential equation derived by deeming either an elemental control volume or an elemental system. It results in a partial differential equation involving the derivatives of density and velocity and is known as continuity equation. General forms of the equation for incompressible flow, in both Cartesian and cylindrical coordinates are given below.

+
+
The differential linear momentum for an infinitesimal element, constitutes the following terms
1. Gravity force per unit volume 2. Pressure force per unit volume 3. Viscous force per unit volume
The sum of these terms corresponds to density multiplied by acceleration. The component form of the law of conservation of momentum is represented as
− ∇ + ∇. =
(2.13)
For a Newtonian fluid, the viscous stresses () are proportional to element strain rates and the coefficient of viscosity. The differential momentum equations of an incompressible flow for one direction only in both rectangular and cylindrical coordinates are given below.
−
(2.15)
It can also be written for other dimensions for both the co-ordinates. This system of equations is essentially known as Navier-Stokes equations.
Navier-Stokes equations have only a limited number of analytical solutions, but computer numerical modeling is highly built upon these fundamental equations. It is possible to attain approximate and realistic results for various complex two- and three-dimensional viscous flows using any commercially available CFD tools.
2.4.2 Hagen-Poiseuille flow
G. Hagen in 1839 and J.L. Poiseuille in 1840 observed experimentally the incompressible flow in a straight circular pipe. For fully developed laminar pipe flow, the above mentioned Navier-Stokes equation is solved by stating initial assumptions to simplify the complex mathematical equation and then implementing the right set of boundary conditions according to the application. This leads to an empirical relation to evaluate the velocity of flow as given in equation 2.16 below.
= (−
= 2
8 (2.17)
This is known as Hagen-Poiseuille flow equation. From equation 2.16, we can interpret that velocity of flow in a pipe varies only in radial direction and its profile is parabolic. The corresponding relationship for velocity of flow in z-direction is valid only for incompressible laminar flow.
2.5 Analytical Background This chapter covers the analytical background to develop the analytical model for numerical analysis. The model provides a correlation between pressure difference and axial displacement that varies the orifice gap. The mathematical formula is derived from hydrostatic bearing design, which gives relationship between flow rate and pressure drop for a given gap height. In other words, the formula is obtained from pressure distribution in stepped film and radial flow through long thin film of hydrostatic bearings. Hence pressure distribution from these models will be discussed in brief in this section.
Stepped film –
The pressure distribution in the stepped film is derived based on Poiseuille flow with a film thickness of h [12]. In the part with film thickness of h, pressure drops from to . is the pressure in the supply side and is the pressure in the return side.

(2.18)
The above equation gives the flow rate of the fluid per unit length for a known pressure drop across the cross-section. It has a linear pressure drop across the fluid film of thickness h which has a length w.
Radial Flow through long thin gap –
For radial flow along long thin parallel film, the pressure distribution and flow resistance are derived from Reynolds Equation. Similar to the equation as obtained for stepped film, flow
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rate in relation with pressure difference and film thickness can be derived from RE in polar coordinates. For linear pressure distribution, the following equation is used.

= 3
12 1+(1 ) 1+(1 )
(2.19)
Chapter 3
Methodology This chapter expounds the implementation of methods and tools to accomplish desired outcomes; as well as helps to comprehend the behaviour of the dynamic component. In the beginning, in order to plan the testing and to solve numerically / analytically, identifying the influential parameters that produces an effect on the system is crucial. This is followed by exploring the experimental test-setup and design modification. This sub-section also furnishes a clear picture on control drives configuration / parameterization for pump and motor. It then extends to numerical/computational analysis of the dynamic system. This encompass various steps, namely modelling fluid domain, setting-up physics for the domain in the ANSYS fluent solver and solving differentials using appropriate solver in MATLAB.
3.1 Parameter identification In order to ascertain the effects (positive / negative) on the behaviour of the dynamic component, the essential part is to identify the influential parameters. The novel rod packing technology with the dynamic component is shown in figure 3.1 below. The result on the effects of the parameters will be covered in later chapter. It is significant to identify and analyse the parameters, so that it can facilitate product development / design improvements.
Figure 3.1 Dynamic component of the novel rod packing technology
• Flow rate of hydraulic fluid • Mass of the dynamic component • Friction acting on the dynamic component • Tube length of hydraulic fluid flow • Geometry of the dynamic component
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Flow rate of hydraulic fluid –
The fluid pressure (hydraulic oil), corresponds to flow rate and the axial distance between the orifice and the dynamic component. The displacement also tends to vary with the flow rate. This term is represented either as mass or volumetric flow rate. In CFD, for applying boundary condition, it’s usually given as mass flow rate.
Mass of the dynamic component –
The geometry and material selection for the dynamic component constitutes the mass. Mass of the dynamic component have an effect on the motion such as displacement and velocity of the component.
Friction on the dynamic component –
There will be two types of contacts established on the assembly of the dynamic component.
• Polymer-metal contact • Metal-rubber contact
The polymer-metal contact is between the piston rod and sealing rings. These rings are made of different materials. For example, a set of rings consists of cast-iron ring and polymer rings. Polymer rings are usually PTFE-filled, but the constituent of this ring is proprietary. The O- ring in association with high pressure oil cup (encompasses the dynamic component) results in metal-rubber contact. Material of the O-ring used is FKM-75. Friction is of paramount importance as it affects the system dynamics to greater extent. The problem is convoluted and so it is hard to model it analytically or to analyse numerically. In this thesis work, arbitrary friction force value is assigned to solve the equation of motion to simplify the complexity in formulation of friction force. In section 3.3.6, detail explanation on equation of motion is provided.
Tube length of the hydraulic fluid flow –
Viscous flow of fluid through pipes as reviewed in section 2.3.2 has an effect on the fluid dynamics. Thus the geometry (diameter and length) of the tube and wall roughness need to be taken into consideration.
Geometry of the dynamic component –
The design of the dynamic component (packing ring cups which holds the sealing rings) needs to be considered for analysis. The geometry of the design might have an effect on the dynamics of the component and on to the functioning (pressure balance). Different geometry configurations can be studied to assimilate its effects accordingly.
3.2 Experimental Testing
Experimentation is considered as an effective method for scientific study of a concept or to discern the principle behaviour of a system. Precise result is one of the main advantages of exerting experimental methods. The prototype of the novel rod packing technology for the Ariel reciprocating compressor JGH/4, which is manufactured for the original size, is used for testing. Manufacturing of the prototype requires special care, so that exact tolerances are maintained as specified in the drawings. Other components which forms up the test rig is also designed and manufactured. The experiments are conducted at KTH Royal Institute of Technology, in the Machine Design department, where the Test-rig is set up. The following sub-section will briefly describe the test-rig setup, instrumentation of the rig, data acquisition, and control drives.
3.2.1 Test-rig
The test-rig comprises of several parts / components which are assembled and connected mechanically with each other. The test frame provides support to the various parts and components as shown in figure 3.2 below. Steel beams of square and rectangular cross- section are welded to each other, which constitutes the frame. It is constructed in such a way that it is rigid, and the legs of the frame is provided with rubber dampers to reduce the vibration. Vibration is induced by the running compressor. A double acting piston compressor is used which simulates the principle function of the reciprocating compressor used in the methane gas transmission application.
Figure 3.2 Test-rig setup
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The experimental setup incorporates various parts which lead to successful working of the compressor and the rod packing technology. As shown in the schematic representation of the arrangement in figure 3.3, it contains an electric motor, a single cylinder modified marine engine, a Bosch Rexroth pump, a novel rod packing, a lube pump, and a double acting piston compressor. In this arrangement, the motor transmits torque to the single cylinder Stirling engine. Through a coupling, the piston rod of the compressor is connected to the modified piston-cylinder of the engine. The axial coupling selected should withstand a force of 4900 - 8000 N. This force (pull and push) is from compressing the gas across the piston cross- sectional area. Thus the rotational motion of the motor is converted to reciprocating motion which imparts movement to the double acting piston cylinder compressor. The Bosch Rexroth gear pump supplies hydraulic fluid to the dynamic component of the rod-packing seal, which recirculates back to the pump. Steel tubes of various dimensions are used as channels for fluid flow.
Figure 3.3 Schematic arrangement
Nitrogen is the gas, which is used for experimental testing, to prove the concept of pressure balancing of the dynamic component. Since, nitrogen is inert in nature; it neither reacts with oil nor ignites at high temperature and pressure. It is stored at 300 bar (g) pressure in a plastic cylinder which can contain a total volume of 6000 normal liters. At a regulated pressure, the gas is supplied through a steel reinforced flexible hose to the inlet of the double acting piston cylinder.
The electrical motor is controlled with a frequency controller (Danfoss drive) and speed is varied from 0 to 1500 rpm. The testing is conducted at 1000 rpm with a piston velocity of approximately 2 m/s. When the compressor runs at around 400 and 700 rpm, the whole set-up vibrates which is undesirable.
3.2.2 Design and construction
The following parts constitute the compressor, which is a reciprocating type, double acting cylinder, piston compressor. The cross-section of the design with the novel rod packing seal assembly is shown in 3.4 below.
• Piston and piston rod • Main cylinder
Figure 3.4 Cross-section of double acting piston compressor with novel rod packing assembly
The cylinder is designed based on the compression ratio required, stroke length of the engine and the rod packing assembly so that both can be fitted together conveniently. The cylinder bore is a known value based on the piston dimension which is fixed (39.5 X 40 mm). The required compression ratio on the bottom side of the piston (for a double acting cylinder) is set as 3, and the length of the cylinder bore is calculated based on the simple formula given below.
C.R = ∗2−2∗ ∗(2−2)∗
=
(3.1)
R is the cylinder bore radius, r is piston rod radius which is also fixed value. The main cylinder is designed such that it can be easily attached and detached with the existing cylinder head and rod packing assembly. They are assembled together with bolts. The cylinder needs
25
to equip a pressure transducer, a quick coupling for gas inlet, and a pressure relief / drain valve. The exact end connection needs to be machined. A volume adjuster is designed to improvise the control volume and adjust the compression ratio. Refer appendix A for detail drawings.
A double acting piston seal and piston rings are selected based on cylinder bore value and type of material required. Accordingly, piston is designed to fit the selected seal and guide rings. Proper tolerance is specified in the drawing as required between piston seal and cylinder bore. Tolerance on the piston seal is given in the manufacturer’s catalogue. Material selection is also a part in the design phase. The material used are steel, aluminum and stainless steel for different parts accordingly.
The rod packing sealing rings used are the existing market products as discussed in section 2.2 (packing ring designs) which are purchased from Ariel corporation. These rings are almost similar to the set of ring pairs used in a traditional packing design. Different ring pairs (double-acting BTR rings, single-acting BD ring pair and scrapper rings set) are placed within the rod packing assembly and the dynamic component. Design of rod packing assembly and dynamic component is shown in figure 3.5 below. Based on this design a fluid model is developed for CFD analysis which is discussed in detail in the following section 3.3. Manufacturing of the dynamic component and the pressure cups (low- and high-pressure cup) needs to be precise with regards to the specified tolerances
Figure 3.5 3D CAD model of a) Rod packing assembly b) Dynamic component
3.2.3 Test objective
The main goal to be achieved from experimental analysis is to study the pressure curves of hydraulic fluid and compressed gas and how it behaves regarding changes with parameters like flow rate, tube connection, and geometry. Another objective is to estimate the leakage rate of the product and compare with the benchmark result. The measured pressure data is analysed with LabVIEW software.
3.2.4 Instrumentation
The gas is compressed inside the main cylinder and cylinder head; hence it needs to be rigged with the required instruments. When the piston reciprocates, the bottom side of the piston is supplied with nitrogen at a constant pressure and flow rate from the storage cylinder. The reciprocating motion compresses the gas to high pressure up-to 150 bar. In the rod packing seal, the dynamic component is immersed in hydraulic oil. When the gas compresses and expands, the dynamic component also reciprocates which pressurizes the fluid. The pulsating pressure needs to be measured. The frequency of the pulsation depends on the speed of the motor. Hence a high frequency pressure measurement device is selected. The pressure transducer is installed at 3 locations, in the main cylinder, high pressure and low-pressure oil cups of the rod packing seal. The pressure transducer adopted can measure both liquid and gas pressure with a range of 0-250 bar. Pressure transducers need to be excited with an external DC supply. The input voltage range for the transducers is 0-30VDC. An AC to DC power supply unit is connected to these transducers to supply 24VDC. The output of the transducer is connected with National Instrument’s pressure module (data acquisition module) to analyze the measured signals.
Figure 3.6 Instrumentation of the test-rig parts
At the cylinder head the temperature of the compressed gas is measured over time. A thermocouple is used to record the temperature. J-type, in-built cold junction thermocouple is connected to the cylinder head. The other end of the thermocouple has two terminals, positive and negative terminal. The two terminals are connected to the temperature module to visualize the data using NI MAX tool. The modules will be discussed in brief in the following sub-section.
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Since the compressor will be working at high pressure, it is required to provide safety. Henceforth, a pressure relief valve is connected to the main cylinder as shown in figure 3.6 above. The pressure relief valve is set to relieve when the pressure reaches 200 bars.
The temperature on the oil is measured from the Bosch pump on the return line. Both the oil contamination sensor and temperature sensor are integrated with the pump. Thus when temperature reaches the limit it indicates a warning in the IndraWorks software and shut down the pump.
3.2.5 Data Acquisition
The pressure module and the temperature module as mentioned in the above section, are connected to the NI compact DAQ which is also known as chassis for the modules. This chassis is connected to the system to observe and analyze the measured signals. This compact DAQ can accommodate 4 modules. Four of the pressure sensors and thermocouple can be connected to the pressure module (NI-9239) and temperature module (NI-9211) respectively.
National Instruments- Measurement and automation explorer (NI-MAX) software interfaces with the hardware to analyze the signal. The working platform of the NI-MAX is shown in figure 3.7 below. Using c-DAQ assist feature, the device linked to the corresponding module can be connected to record the measured signals. The temperature data of the compressed gas is reviewed in NI-MAX.
Figure 3.7 NI-MAX platform
The gas and liquid pressures are both analyzed using LabVIEW software. DAQ-assist function from the express VI function palette is used to interface with the pressure module. LabVIEW has two windows, namely front panel and block diagram. A graphical program is developed with the help of function palettes in the block diagram window and the waveform charts / graphs to plot the pressure data is created and viewed in the front panel. To record,
analyze and store the pressure data of both fluids (gas and oil), the graphical program is developed as shown in figure 3.8. The data measured is stored as excel file using write to measurement file express VI. Write to measurement file express VI icon is seen in LabVIEW front panel of figure 3.8. It is used to edit and plot the required range of datas.
Figure 3.8 LabVIEW front panel and block diagram window
3.2.6 Drives and Control
The remote operation and control of the electrical motor and the Bosch Rexroth CytroPac pump, necessitates the use of control drives. The electrical motor is connected with Danfoss frequency converter. Using an Ethernet cable a connection is established with the computer. VLT motion control tool MCT 10 is the supporting software to interface with the frequency converter. The related parameters that allow access to remotely control the motor is set in the MCT 10 parameterization dialog box as shown in figure 3.9 below. The parameter file edited in the project is used to access the drive remotely. The motor speed is set and controlled using this tool, by varying the frequency value in terms of percentage. Desired working speed of the motor is 1000 to 1200 rpm. Update output frequency button of the remote controller is used to update the ongoing request to change the motor speed. Similarly, start and stop control option is used to switch on and off the motor intermittently.
29
Figure 3.9 VLT Motion Control Tool for remote control of the motor
Figure 3.10 Parameter editor dialog box for IndraWorks DS 14V20
Sytronix FcP 5020, the frequency converter is inbuilt with the Bosch CytroPac pump, controls the motor. USB cable connects the pump system with the computer. IndraWorks DS 14V20 is the software tool to communicate with the frequency converter. It is also used to view and analyze the working of the system and diagnose warnings and errors. There are many parameters related to the FcP 5020 converter, but the essential parameters are edited according to the application. Parameter values can be viewed in parameter viewer file as shown in figure 3.10 and is edited in the parameter editor dialog box. The inlet flow rate for the rod packing seal is set to 0.8 l/min using the flow command parameter (F1.12). The pump motor automatically shuts when the pressure in the return line reaches the set pressure limit of about 150 bar.
3.3 Computational analysis of the dynamic system
The sealing cups with the packing rings in the novel rod packing technology, together constitutes as dynamic system as discussed in section 3.2.2. It is expensive to analyze the dynamic behaviour and functionality of the system experimentally for different parameters. Hence it is essential to make computational analysis to vary certain parameters and analyze the outcome. For FE modelling and analysis, the model is developed with SolidEdge surface
modelling and imported to ANSYS. Fluent pre- and post-processor is used for setting-up physics and analyzing the results respectively. A mathematical model that represents the physics of the system is solved using MATLAB numerical solver. The analysis of both the fluid model and the mathematical model are discussed further in the following sub-section.
3.3.1 Scope and method development for CFD
Computational fluid dynamic analysis is required to understand the pressure development of the hydraulic fluid, which acts on the face of the dynamic component as shown in figure 3.11, and for various displacement of the dynamic component. Different methods are identified to build the FE model.
1. A 2D- axisymmetric model that constitutes a solid domain, a liquid domain, and air domain.
2. A 2D-axisymmetric model of the fluid domain 3. A 3D model of the fluid domain
Figure 3.11 Hydraulic fluid acting on the face of the dynamic component
The cross-sectional view of the novel rod packing technology as shown in figure 3.4 above is analyzed to construct a 2D- axisymmetrical model as described in the first method. In this method, the contour of the dynamic component which acts as a solid domain is modelled. The two fluid domains contour are modelled. All the domains need to be connected using the mesh interface. The simulation from this method will represent the function/behavior of the dynamic component as analyzed experimentally. One fluid domain is defined for the hydraulic fluid (viscous oil) properties and other domain represents the air. Aluminum is defined for the solid domain. This model is complex as it requires complex dynamic mesh motion and meshing methods for the successful simulation of the model. Also, the objective to obtain the correlation between the fluid pressure and the displacement can be obtained with the second method as stated above. In this method the fluid model is developed by analyzing the cross-sectional contour of the fluid region inside the pressure cups (high pressure and low- pressure cup). The contour comprises boundaries of different elements of the novel rod packing seal. Geometry of this model is explained further in the following sub-section. This model is simple, takes less computational time to achieve the above stated objective. Hence, it is the chosen method for CFD. The third method is just obtained by revolving the fluid model developed for second method into a 3D model. The result for fluid pressure obtained will be like that of the previous method, and it also has the disadvantage of consuming higher
31
computational time and meshing is limited. So, this method is also discarded. The model developed for the three methods is shown in figure 3.12 below.
Figure 3.12 Fluid model for CFD analysis a) 2D-axisymmetric model for fluid domain b) 3D model of the fluid domain c) 2D-axisymmetric model with both solid and fluid domains
3.3.2 Fluid model
For the second method as discussed in section 3.3.1 the fluid model is developed. The hydraulic fluid from pump as discussed in section 3.2.1 enters and leaves the novel rod packing seal in the radial direction i.e., radially outwards from the piston rod. Around the dynamic component the flow of this viscous fluid is both circumferential and tangential. As the flow in the direction of the cylindrical coordinates of the system is not of importance, a 2D-axisymmetric model (x-y plane of rectangular coordinates) is developed. SolidEdge surfacing feature is used to create the 2D model which is imported to ANSYS for further analysis. As shown in figure 3.12 a) above, the contour of the fluid model comprises of boundaries of various elements which faces the fluid. In simple terms, the face of the components that is in contact with the fluid forms the boundary of the defined model. To obtain the correlation of the high-pressure development with the displacement of the dynamic component, the boundary lines of the contour of dynamic component is modified for different position (termed as x). So, it is a repetitive method, where the pressure of the hydraulic fluid is analyzed for different displacements.
3.3.3 Meshing
The essential stage in a computational fluid dynamic analysis is to mesh the model. ANSYS mesh modeler is used to generate the 2D quadrilateral linear order mesh elements for the fluid domain. In the fluid model, the gap in the orifice is very small as shown in figure 3.13 below. The gap in the orifice means, the distance between two boundary lines (one is the slanting line of the dynamic component face and other is the slanting line of the low-pressure cup). Meshing in this gap needs to be refined since, it is where the fluid pressure drops, and the velocity profile is parabolic.
Figure 3.13 Linear quadrilateral mesh elements
Proximity and curvature mesh sizing option and fine meshing is chosen for the models. Mesh convergence study is done for a liquid model with 0.3mm axial gap in the orifice. It is observed that the meshing did not have major effect on the pressure drop along the orifice. So, for different models (with varying axial distance in the orifice) an average of about 100,000 mesh elements is generated. The computational time is also comparatively low compared to meshing the 3D fluid model for about 50,000 elements. Skewness in the mesh elements is maintained to a low value around 0.005. Further the meshing quality is also checked under fluent processor.
3.3.4 Setting up physics
A steady-state, pressure based, axisymmetric fluent solver is selected. Since the flow from the pump through the channel to the rod packing seal inlet is smooth and constant, the laminar flow model is considered to solve the 2D axisymmetric model. For the fluid model, the hydraulic oil used in the experiments is defined to the model by inputting the properties of the fluid as given in table 3.1 below.
Table 3.1 Hydraulic fluid properties
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Property Value Unit
Density 879 kg/m3
Viscosity 0.132 Kg/m-s
The boundary condition is a major step in setting-up physics to solve flow problems. In the experiments conducted, a known flow rate from the pump is supplied to the inlet line. At the inlet boundary (marked blue) of the fluid model as shown in figure 3.14, a mass flow inlet is defined as the boundary type. Mass-flow rate is given as 0.0103 kg/s. At the outlet the boundary type is set as pressure outlet, and a gauge pressure value of 300000 Pa is set.
Figure 3.14 Boundary conditions defined to the fluid model in setting-up physics.
3.3.5 Solving and Post processing
There are different numerical solvers for Fluent to solve the problem it encounters. The details about the solvers can be found in [13] and here the type of solution methods used will only be discussed. PISO type pressure velocity coupling scheme and for pressure spatial discretization PRESTO are selected for solution methods. For other values and methods, it is set to default options. This solution methods solves the flow equation. The numerical solver that the fluent implements is an iterative type to obtain converged solution. The problem needs to be initialized and a hybrid setting is provided for initialization.
The main objective for computational analysis is to obtain correlation of the fluid pressure build up for different displacements of the dynamic component. From the fluent post processor, the following results is obtained for different orifice gap.
• Pressure drop across the orifice and the maximum pressure developed surrounding the dynamic component.
• Average flow velocity and streamline function
In fluent post-processing, a contour plot is used to obtain the graphics for absolute pressure and total velocity component of the fluid model. As stated in the objective, it is essential to note down the maximum absolute pressure value for every model with different orifice gap. The values of pressure against the displacement is plotted in a line chart.
3.3.6 Analytical model and Equation of motion
Let’s consider the dynamic component (with sealing rings and packing cup) as a single body with mass m. Air pressure acts on one side of the dynamic component which is sinusoidal in nature due to compression and expansion of air over time . Due to this pressure the body moves towards the other side and the pressure builds up on the hydraulic fluid side, since an incompressible viscous oil is used. When the pressure is greater than air pressure, the body moves back. Henceforth, back and forth motion is obtained. The free-body diagram that describes the physics of the system is shown in figure 3.15
Figure 3.15 Free body diagram of the dynamic component.
From the above free body diagram, an equation of motion can be formulated that entails all essential parameters.
x = − + 0 − x − x () (3.2)
The above ordinary differential equation is solved using a numerical solver like ode 45 using MATLAB. Refer Appendix B for the code. The air pressure is generated using the equation 3.3 given below, which is sinusoidal in nature and is given as an input parameter. Similarly the damping coefficient, c of the hydraulic fluid is defined as constant. A frictional force value is guessed and is given as arbitrary input to the friction component of the above given equation of motion.
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= + ( ∗ sin (2 ∗ ∗ )) (3.3)
With M as mean pressure, A is the amplitude of sine wave, and T is the periodicity. Thus, air pressure is generated which is a function of time. Oil pressure term is calculated within the numerical solver loop, since it is a function of x displacement. So everytime when x is solved it will be given as input to compute the oil pressure. So a correlation between oil pressure and displacement x is obtained and the formula is presented below.
= ∗ ∗∗∗
+ (3.4)
The formula obtained as given in section 2.5 is the analytical model that represents the CFD model developed to obtain the pressure accretion in the hydraulic fluid domain based on the displacement of the dynamic component. So the CFD model is validated by this analytical model.
Hydraulic fluid pressure force projected on to the cross-sectional area of the component is not the same area as the air pressure force projected area as shown in figure 3.16 below. Hence for the given geometrical design of the component, the pressure developed in the hydraulic fluid side is slightly greater than the generated air pressure.
Figure 3.16 Fluid pressure acting on the dynamic component
The 0 component of the equation of motion, where the projected area on the oil side () is divided into three sections. 1 is the projected area where maximum oil pressure acts; 2 is the projected area where minimum oil pressure acts. The force acting on the orifice gap is computed by integration of varying oil pressure force (pressure multiplied with the circumference). This constitutes the third section. Thus the equation 3.2 is rewritten as
× = − + 10 + 20 + ∫ () × 2 × − × − × ()
The integration of pressure force is performed over the limits of and , which represent the inner and outer radius of the orifice respectively.
3.3.7 Geometry design configurations
Different design configurations are developed by varying the geometry outline of the dynamic component. This is used to study the geometry’s effect on the dynamics and as well for functional analysis of the system. For basic understanding of different configurations, the outline of the fluid domain is developed using surfacing feature with SolidEdge as shown in figure 3.17 below. Instead of representing the geometrical change for different configurations in the 3D model of the dynamic component it is represented through the 2D outline of the hydraulic fluid domain. The 2D fluid model is a result of cross-section of the rod-packing assembly as shown in figure 3.4.
Configuration 1 represents the chamfer edge of the orifice with an angle of 45°; configuration 2 represents the chamfer edge of the orifice with an angle of 30°; configuration 3 represents the chamfer edge of the orifice with an angle of 60°. More geometry configurations can be studied, but for this thesis work analysis of only three configurations are conducted to check if the developed analytical model can be used to study their effects.
Figure 3.17 2D-fluid model for different configurations a) configuration 1 b) configuration 2 c) configuration 3
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Results and Discussions
This chapter provides the results from numerical simulation carried out by CFD and solving the analytical model. The first section describes the oil pressure development in the hydraulic fluid side of the dynamic component using ANSYS Fluent. The CFD model developed for the fluid domain is validated with an analytical model. Then the results from transient simulation for the dynamic system, by solving the equation of motion with MATLAB, are presented. After validating the model with the previous work of experimental results, analysis of pressure difference and displacement for different configurations are also presented in this chapter.
4.1 CFD simulation of Liquid Domain The result from ANSYS Fluent gives the pressure distribution of oil in the developed liquid domain over a range of orifice gap. Varying the orifice gap represents the displacement of the dynamic component. Thus from this static CFD analysis, a corresponding relationship between the maximum pressure and displacement is obtained. The pressure distribution in the liquid (hydraulic oil) domain for different gaps is shown in figure 4.1 below focus the orifice gap.
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Figure 4.1: Pressure distribution for different orifice gap a) x= 0.07mm b) x= 0.12mm c) x=0.16 d) x= 0.22mm
The pressure in the red region as shown in figure 4.1 above represents the maximum pressure formed that acts on 3/4th of the projected area of the dynamic component. The maximum pressure acquired for different displacement is plotted and an exponential curve is obtained as shown in figure 4.2 below.
Figure 4.2 Displacement versus oil pressure plot
4.1.1 Validation with Analytical Model
As discussed in section 3.3.6, analytical models that represent the pressure drop over the orifice, as obtained with the CFD model, are solved analytically with MATLAB. The results of various analytical models are compared with the CFD result as shown in figure 4.3. The three analytical models closely match the CFD result. The models are solved using the same boundary conditions. Hence the CFD model is validated.
3,00E+05
8,00E+05
1,30E+06
1,80E+06
2,30E+06
2,80E+06
3,30E+06
Figure 4.3 Comparison of analytical model with CFD model
4.2 Dynamic Response Analysis In this chapter the dynamics of the novel sealing solution is analyzed through numerical simulation of the equation of motion of the system which is solved using MATLAB. At first the pressure curves for both generated air pressure and computed oil pressure is plotted together for analyzing the pressure difference. Then the displacement of the dynamic component is analyzed. The performance of the component dynamics for varying friction force and geometry (chamfer angle) is also discussed and analyzed.
4.2.1 Time-domain analysis
The major analysis presented in this thesis work involves the pressure balancing of the fluid for the rod packing seal and the displacement of the dynamic component. Pressure balancing plot as shown in figure 4.4 below is obtained for initial design configuration of the dynamic component as discussed in section 3.4 with inlet flow rate of 0.8l/min of hydraulic fluid, supplied air pressure ranges from 5.67 bar to 6.72 bar. The force acting on the component from the hydraulic side is greater than the force acting on the other side of the component. This is reflected in the pressure plot, as the oil pressure is slightly greater than the compressed air pressure. The plot shown below is generated for an arbitrary friction force value of 200N acting on the component. This plot is comparable with the experimental results from previous work and hence the model is validated. The average pressure difference is 9.5e4 Pascals as given in figure 4.4 below.
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Figure 4.4 comparison of air pressure with hydraulic fluid pressure
The corresponding displacement of the dynamic component which is influenced by the forces acting on the component is analyzed in the figure 4.5 given below. For the base configuration of the design with a friction force of 200N the travel of the component is found to be 9 µm.
Figure 4.5 Displacement plot of the dynamic component
4.2.2 Performance Analysis
The dynamic behavior and the functionality of the sealing component is analyzed for various frictional force and for three different configurations. As shown in figure 4.5, with increase in friction force the travel of the component drops which is indicated by flattening out of the oil pressure when friction force reaches about 500N. Thus, the amplitude of the oil pressure decreases over increasing friction. The plots shown in figure 4.6 are obtained for the base configuration of the dynamic component.
Figure 4.6 Oil pressure and air pressure plots for different friction force starting from top left corner with friction force of a) 50N b) 100N c) 300N d) 500N
The corresponding displacement plot for 50, 100, 300 and 500 Newton of friction force is shown in figure 4.7 below. Travel of the dynamic component for these friction forces are 0.013mm, 0.011mm, 0.005mm and 0.001mmrespectively.
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Figure 4.7 Displacement of the dynamic component for different friction forces starting from top left corner with friction force of a) 50N b)100N c) 300N d) 500N
By comparing the results with previous experimental results for the base configuration, based on the amplitude of oil pressure and the pressure difference (Δp), the friction force on the dynamic component is estimated to be around 200 N. For this frictional force, the displacement of the dynamic component and pressure balance for three different configurations is plotted and analyzed. As shown in figure 4.8 below, the computed oil pressure for three configurations and generated air pressure are plotted over time. In the plot, configuration 1 represents a chamfer angle of 45°; configuration 2 represents a chamfer angle of 30°; configuration 3 represents a chamfer angle of 60°.
Figure 4.8 Comparison of oil pressure for three different configurations
The corresponding travel of the dynamic component for the three configurations are plotted in the figure 4.9 shown below. The data acquired by computing the equation of motion, is solved
with a friction force of 200N acting on the component. The axial orifice gap between the chamfer edge and the component varies for different configurations which can be noticed from the plot given below. The values for displacement range and pressure difference for the three configurations as obtained from the plots as shown in figure 4.8 & 4.9, are tabulated below.
Figure 4.9 Comparison of travel of the component for different configurations
Table 4.1 Values of average pressure difference and displacement range for different configurations
Types Average Pressure difference (Pa)
Average displacement range (mm)
Configuration 1 (45° chamfer) 9.4527E+04 0.009
Configuration 2 (30° chamfer) 6.5229E+04 0.011
Configuration 3 (60° chamfer) 5.9466E+04 0.007
To analyze the effect of friction force on the pressure balancing, an assumption is made to the model. The plot as shown in figure 4.6, cannot be used to find the effect of friction force onto the pressure difference. Reason is computing average pressure difference for different friction forces will give similar values which doesn’t not make sense. So force for maximum oil pressure 0 is assumed to act on the entire cross-sectional area of the component. But in reality, oil pressure force varies over the face of the component as given in section 3.3.6.
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Normalized pressure plot for the base configuration with different friction force is obtained as shown in figure 4.10 below.
Figure 4.10 Normalizing pressure plot for different friction forces a) 50N b) 150N c) 250N d) 500N
From the plots as given in figure 4.10 above, average pressure difference is computed which is plotted against friction forces as shown in figure 4.11 below. From the influence of friction force on pressure difference plot, a linear relation between friction force and pressure difference is obtained. Also from this plot it is noted that the pressure difference value will not change once a limit is reached. In this case, 450 N is the limiting friction force.
4,00E+03
9,00E+03
1,40E+04
1,90E+04
2,40E+04
2,90E+04
3,40E+04
0 1 0 0 2 0 0 3 0 0 4 0 0 5 0 0
Δ P_
A V
E R
A G
E I
N P
A SC
A L
ΔP for 45degree(Pa)
Figure 4.11 Influence of friction force on pressure difference for base configuration
4.3 Experimental result Test-rig modification was the beginning part of the work which involved designing and fabricating parts for double acting piston cylinder with additional attachments as per the requirements. Refer to appendix A, for detailed drawings of the components. Benchmarking test for the competitor’s product was conducted, in which long time test run was unsuccessful due to heat buildup and the temperature did not stabilize as expected. Tried couple of water- cooling methods to bring down the temperature, which did not work as well.
Before starting the tests for the novel rod-packing seal technology, lot of time and work was spent on troubleshooting the test-rig. Moreover, at the beginning of testing the technology, problems were faced with oil leakage along the rod out from the packing flange and scrapper cup. A reason could be a manufacturing defect of the packing flange, where proper tolerance needs to be maintained at the BD ring seat.
These problems resulted in that the experimental part of thesis work could not be completed as planned
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Chapter 5
Conclusions This chapter delineates the conclusions which are inferred from the computational analysis of the analytical model developed for the design of the novel rod packing seal technology. In summary from the previous chapters, an analytical model was framed based on the CFD analysis and was used to solve the physics of the system for transient analysis. The behavior of the dynamic component is analyzed with an analytical model through numerical simulations but could not obtain results through physical testing. Also parameters like friction force and geometry are studied with the validated analytical model. From the influence of friction force on pressure difference study, a linear relation between friction and pressure difference is observed. Also, by changing the geometry (chamfer length and angle of the dynamic component) of the dynamic component, it can be observed that design configuration with 60° chamfer angle gives smaller pressure difference value compared to the original design.
• Results from CFD analysis closely match with the analytical model which provides correlation between hydraulic fluid pressure and axial displacement of the dynamic component.
• The model developed and analysed in this thesis is suitable for functional analysis of the component and influence of certain parameters like geometry, friction force, mass and flow-rate can be studied.
• The friction force value is estimated to be around 100-200 N, from the model which represents the dynamics of the component and on comparison with the previous testing results.
• From the performance analysis for different friction forces and its influence on pressure difference, it can be noted that, a linear relation is obtained between friction force and pressure difference. Also friction affects the dynamics and for the case analysed in this thesis it can be seen that when friction force is above 450 N oil pressures become constant. This means that the component does not perform as it should.
• The developed model provides a good estimate of the pressure balance and trav