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Novel Concepts & Research
Chapter 16
Novel Concepts & Research Sarah Simons*, J. Jeffrey Moore*, Karl Wygant†, Harry Miller‡, Eugene "Buddy" Broerman* and Aoron Rimpel§ *
Southwest Research Institute, San Antonio, TX, United States, †Hanwha Power Systems Americas Inc., Houston, TX, United States, ‡Dresser-Rand, A Siemens Business, Olean, NY, United States, § Southwest Research Institute, San Antonio, TX, United States
The following novel concepts for oil and gas machinery are under development but are not (as of the publication date of this book) offered in production machines.
Cooled Diaphragms for Centrifugal Compressors Isothermal compression is one of the recent areas of research and development in centrifugal compressors. To reduce the power penalty associated with a large number of compression stages, internally cooled diaphragms were developed under a Department of Energy (DOE) research project to increase the efficiency of CO2 sequestration from power plants. Through three-dimensional (3D) computational fluid dynamics analysis, a design can be achieved to provide good heat transfer while adding no additional pressure drop [1]. This technology allows a barrel compressor to achieve performance close to an integrally geared compressor. Fig. 16.1 shows the conceptual cooled diaphragm design where cooling flow (blue) is routed through the diaphragms adjacent to the gas flow path (red). The total temperature increases, due to the work input of the impeller, are reduced through the diaphragm flow path thereby reducing the temperature into the downstream stage. Concepts of printed circuit heat exchangers were used for the basis of developing an efficient heat transfer mechanism to ensure that the compression process is nearly isothermal. A new compressor was fabricated based on a Dresser-Rand DATUM D12 frame size. The compressor consisted of a six-stage, back-to-back centrifugal compressor (D12R6B) that incorporated the cooled diaphragms. The compressor and associated piping flow loop are shown in Fig. 16.2 [2]. Aerodynamic testing of the compressor in several operating configurations was completed with and without the cooling diaphragms activated. Results showed that the Compression Machinery for Oil and Gas. https://doi.org/10.1016/B978-0-12-814683-5.00016-X © 2019 Elsevier Inc. All rights reserved.
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FIG. 16.1 Cooled diaphragm concept [1].
FIG. 16.2 Left—installed Dresser-Rand DATUM compressor. Right—DATUM compressor test loop [2].
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cooled diaphragm technology reduces compressor power consumption by 3.0% near surge to 8.0% near choke when compared with the adiabatic case with intercooling between the two sections. Additional savings would result if the machine was configured as an eight-stage compressor, which would provide two additional internal heat exchangers. A 9.0% power savings was measured when the compressor was operated as a straight-through compressor with no intercooling at the design point and would be higher (>10%) at larger flows. The heat exchanger effectiveness for the cooled diaphragm was measured between 12% and 30%, depending on the stage, operating point, and backto-back vs straight-through intercooling configurations. The cooled diaphragms removed 297–311 K (28%–35%) of the temperature rise within each section when compared to the adiabatic case. The measured temperature drop, heat exchanger effectiveness, and power savings were all slightly higher than predicted values. Operation of the cooled diaphragms changed the characteristics of the multistage machine, increasing flow capacity and pressure ratio compared to adiabatic performance at the same speed. Additional performance gains may be realized by designing the compressor aerodynamic flow path for the cooled case rather than the adiabatic case. No reliability issues associated with the cooled diaphragm design were encountered during testing including diaphragm leakage. While these results provided were for a CO2 application, benefits can be realized for any high-pressure ratio application or where reducing discharge temperatures is desired (e.g., preventing ethylene polymerization).
Subsurface Compression To increase the hydrocarbon recovery in oil fields, a subsurface process and reinjection compressor (SPARC) tool was developed to process, compress, and reinject up to 30% of the gas in a high gas-to-oil ratio well (defined by greater than 15,000 standard cubic feet of gas per stock tank barrel of oil). This would allow additional black oil to be produced to the increased gas rate production and increase the gas condensate production since the recycled gas will carry the condensates to the production stream [3]. The primary goal of the tool design is to provide additional gas handling capacity while maintaining reservoir pressure to existing production fields at a fraction of the cost of a surfacebased operation. Fig. 16.3 represents the SPARC concept presented by Brady et al. [3] and will be used to describe the tool’s basic functions. The SPARC device includes a three phase separator at the bottom of the tool which will prevent most liquids and solids from passing through the turbo-expander or turbine (separator 2 is labeled as pre-swirl auger in Fig. 16.3). The remaining solids and liquids bypass around the turbomachinery section. The energy in this production stream is usually wasted by passing through a series of chokes within the production system. A turbine bypass valve (not shown) allows the gas flow to enter the turbine only after the well has been cleaned up and is free from debris usually seen while
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FIG. 16.3 Schematic representation of SPARC concept for functional description [3].
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bringing a shut-in well back online. Until these conditions are achieved, the valve remains closed and flow bypasses the turbomachinery section entirely. Reservoir gas is filtered (not shown) and used to cool the oil-free bearings. The production stream is expanded through the turbine, rejoins with the bypass flow, and is separated once again by an auger separator—the gas is directed to the compressor suction while the other solids/liquids continue to the surface. A recycle valve is utilized to prevent the compressor from operating in surge conditions. The gas which exits the compressor is finally reinjected back into the formation. The SPARC test rig components and rotor assembly are shown in Fig. 16.4. The design speed of the turbine and compressor is 105,000 rpm (105 krpm), and the design compressor power consumption at down hole conditions is approximately 343 kW. Given that the turbine weighs approximately 0.907 N and the combined compressor impeller weight is approximately 0.703 N, the power-toweight ratio of the SPARC is an order-of-magnitude larger than a typical commercial centrifugal impeller. Gas-lubricated foil bearings are utilized using the gas following clean-up process. Aerodynamic open-loop testing was performed on the SPARC tool. Several performance curves were generated for each component ranging from 73.5 to 115.5 krpm (70%–110% of design speed) [4]. In general, measured compressor head was approximately 10% lower than predictions, and measured efficiencies for the 84 krpm speed line were very close to the predicted values. The turbine had higher flow than predictions, due in part to a larger flow area in the manufactured turbine wheel. Rotordynamic open-loop testing was also performed which showed an instability of the conical rigid rotor mode at the design operating speed of 105,000 rpm [5]. It was found that the amount of stiffness the seals contribute varies with operating pressures which changes the frequencies of the conical and cylindrical modes. This was resolved by replacing the holepattern seals with higher clearance labyrinth seals and stiffer bearings; subsequent testing showed no stability issues over the operating speed range.
FIG. 16.4 Left—SPARC test rig components: (A) test rig housing assembly, (B) large labyrinth seal, (C) compressor balance piston hole-pattern seal, (D) stage 2 compressor impeller, (E) stage 1 compressor impeller, (F) interstage diaphragm section, (G) stage 2 diffuser and de-swirl cascade, (H) turbine hub hole-pattern seal, and (I) complete rotor assembly. Right—Rotor assembly.
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Advanced Seals The current state-of-the-art method for sealing flow in a centrifugal compressor between the rotating and stationary components consists of several types or styles of seals, such as plain stationary labyrinth seals, rotating labyrinth seals, stepped labyrinth seals, pressure activated leaf seals, finger seals, and brush seals. An alternative dynamic pressure-balanced (DPB) seal uses the flow and pressure drop associated with the fluid it is sealing to generate a force-balanced seal system with an operational clearance as tight as a few thousandths of an inch (hundredths of a millimeter) over the rotor surface. A cross-section of the DPB seal is shown in Fig. 16.5. The DPB seal is made up of a circular array of shoe segments (see Figs. 16.5 and 16.6). Each shoe segment acts independently and is connected to the static hardware via a spring support. The seal is typically installed with a nominal gap that is larger than the desired running clearance. The acceleration of fluid between the seal and the rotor creates a lowpressure region that draws the seal shoes toward the rotor. As the seal-surface approaches the rotor surface, features within the seal surface reduce the velocity of the fluid, and when combined with static upstream pressure results in a pressure rise that increases the radial outward force on the seal shoes. The operational clearance is achieved when this outward force is balanced with the inward force from the upstream and downstream pressures acting on the backside of the seal. Seal dimensions are tuned to achieve the desired operational clearance. This method of creating a force balance over the rotor surface facilitates a seal that maintains a set clearance during dynamic changes in pressure,
Retaining ring
Secondary seals
Scalloped front plate Shoe (15 pl.)
Rotor
Flow direction Beams (2 pl. per shoe)
Radial stop
FIG. 16.5 Cross section of the DPB seal.
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FIG. 16.6 DPB seal photo.
temperature, centrifugal growth, and rotor eccentricity. The nature of operation is not negatively impacted by swirl ratio or surface velocity associated with high-speed operation. DPB seals are currently being developed and tested for the following reasons: DPB seals have been shown in laboratory settings to maintain design flow leakages (minimizing wear) for longer run times thus extending service intervals, increase aerodynamic efficiency by up to 2.0%, and allow a more forgiving operating envelope for the rotor which increases rotor dynamic stability [6–8]. Testing of these seals in an ASME PTC-10 Type 1 test of a 31-MPa absolute discharge pressure gas lift centrifugal compressor found that 2.8% lower power was required for the same head level across the entire range of inlet flows and pressure ratios, compared to the same test with conventional labyrinth seals [9]. In addition, the system rotordynamic stability, obtained via operational modal analysis (OMA), showed the DPB) seals exhibited log decs similar to standard labyrinth seals across the entire range of flows and pressure ratios.
Gas Bearings Typical fluid film bearings utilize oil as a lubricant to generate the necessary hydrodynamic pressure to support a load. However, bearings utilizing gaseous lubricant is also possible. This is advantageous when the process gas of a machine is able to serve in this role. For example, air cycle machines on aircraft have utilized air bearings since the 1980s [1]. Gas bearings for turbomachinery have notable advantages over traditional oil film bearings, namely, the elimination of sealing systems to separate oil lubricant from the process and reduced parasitic power loss due to lower fluid viscosity. Also, compared to rolling element bearings, gas bearings do not have speed limits due to high contact stresses from centrifugal loading. The only other “oil-free” competitor to gas bearings
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are active magnetic bearings. Both gas bearings and magnetic bearings can be utilized in hermetically sealed machines, but the magnetic bearing requires control systems and back-up bearings, which add complexity and cost. The advantages of gas bearings over other bearing types allow gas-bearing machines to be more compact, less complex, operate cleaner and more efficiently, and operate at higher relative speeds. However, due to the lower viscosity of a gas compared to oil lubricants, disadvantages of gas bearings are low load capacity and low damping. This has limited gas bearing use to socalled microturbomachines, that is, machinery with outputs on the order of hundreds of kilowatts or less. Advancements in gas-bearing technology have tried to address the load capacity and damping issues in order to permit their use in larger machines. In compliant gas bearings, such as foil bearings, progress has been achieved by tailoring the compliance structure stiffness profile to the gas film stiffness [10]. This ensures a more uniform hydrodynamic film thickness and enhances load capacity. However, hydrodynamic-only load capacity is usually insufficient for gas-bearing machines with larger, heavier shafts (shaft weight scales with length to the third power). For example, shaft sizes larger than 100 mm generally require additional lift capability via hydrostatic pressure [11–13]. Fluid film damping in a gas bearing is often supplemented by external damping mechanisms. For example, foil bearings have friction damping at the interfaces with the underspring, which may be corrugated bump foils, layers of foil leaves, or metal meshes. However, the most-effective supplemental damping scheme for gas-bearing applications targeting larger-size machines appears to be with the incorporation of squeeze film dampers. In this configuration, the squeeze film damper utilizes oil like a conventional squeeze film damper, but the damper has to be hermetically sealed for the bearing to remain “oil-free.” Delgado and Ertas [12] presented damping results for a 110-mm hydrostatic gas bearing with hermetically sealed squeeze film dampers, and they showed 200%–300% increase in damping compared to a similar bearing with metal mesh dampers [11]. Gas-bearing technology offers numerous advantages over conventional oillubricated fluid film bearings, rolling element bearings, and magnetic bearings. The drawbacks of lower load capacity and limited damping have prevented gasbearing use in larger-sized turbomachinery. However, technical advancements have improved the performance of gas bearings over the years, where continually larger applications seem realistic. Specifically, utilizing hydrostatic lift features and hermetically sealed squeeze film dampers represents the current state of the art for large-scale gas-bearing machinery.
Additive Manufacturing Additive manufacturing has the potential to be the single largest driver of change to the turbomachinery industry in the past 100 years. The reason for this
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FIG. 16.7 Closed compressor impeller manufactured by direct metal laser sintering of metal particles. Left—raw manufactured, right—cleaned.
bold statement is that additive manufacturing has a direct impact on potentially every component of the turbomachine design. That impact extends to allowing substantial reductions in stress and the ability to implement design concepts that could not previously be manufactured. Various additive manufacturing techniques are emerging. Fig. 16.7 shows a compressor impeller that was manufactured by direct metal laser sintering (DMLS). DMLS uses a high power-density laser to melt and fuse metallic powders into a solid component. The process can leave micro-voids in the part that are closed by a hipping process. There is variations between different metals, but in general the DMLS process produces a part with comparable yield and ultimate strength as forged material properties while some decrement is present in fatigue strength. The ability to apply localized fillet and strengthening can more than offset the loss of fatigue strength margin. Manufacturers are investigating the use of additive manufacturing for a wide range of turbomachinery components. These components are: guide vanes, impellers, integral impellers and shafts, etc. For gas turbines there is active work on applying the technology for combustor nozzles and turbine blades. The impact to the industry is that greater levels of customization is present, novel configurations can be explored, welds and brazed connections can be eliminated, and the process produces very little waste material as unused metal powders is salvaged after the build is complete. Currently the obstacles in the technology are related to relatively small components 300 300 300 mm and quality assurance challenges. Since the parts created can be complex shapes with internal geometries the ability to check for voids and/or deformities is challenging. As the size of the additive manufacturing machines increases and inspection processes mature the application of the technology will only increase in the turbomachinery industry eventually necessitating that design engineers rethink the entire design manufacturing process.
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Casing Treatment (Range Extension) Casing treatment has been applied in automotive turbocharger industry for decades. The technology allows the ability to substantially extend compressor range at minimal impact to efficiency and head rise. The technology has not previously been applied in industrial turbomachinery due to the need for extensive optimization of the configurations. In the automotive industry extensive design optimization occurs and then turbochargers are mass produced and the relatively high-pressure ratio of those turbochargers. In the industrial machinery industry the mass production volumes are substantially smaller and the end users more reticent on accepting innovation and typically pressure ratios smaller (thereby naturally giving better range than higher-pressure ratio stages.) Casing treatment is a compressor range extension technology that can be easily integrated in to most compressor systems without significant additional expense. Casing treatments have gained acceptance in some applications since the technology can be implemented with few additional stationary parts to maintain a robust and reliable configuration. The casing treatment can extend the range of a stage by delaying the onset of stall in the inducer of the compressor. Inducer stall occurs due to the high levels of incidence and loading near the leading edge of the blade. Flow recirculation through the casing treatment increases the inducer flow rate and helps reduce the leading edge loading. This delays the onset of inducer stall and results in enhancement of compressor map width. Fig. 16.8 (left) shows the flow recirculating from the inducer back to the inlet through the casing treatment cavity near stall for a typical design. Fig. 16.8 (right) shows the amount that the compressor range was extended for two different impellers, Type A with a high-specific speed design, Ns ¼ 1.00 and Type B with a medium-specific speed design, Ns ¼ 0.85 [15].
Gas Property Testing Equations of State (EOS) are used to determine the performance of compressors, size and select various types of equipment, and evaluate pipeline hydraulics. There are several different commonly available models developed over time with the most commonly used ones being GERG, NIST, AGA8, PengRobinson (P-R), Benedict-Webb-Rubin-Starling (BWRSE), and RedlichKwong (R-K). EOS were first developed based on empirical data for pure substances as a method of relating the pressure, density, and temperature of a gas. Modern EOS extend these original equations to mixtures by combining the properties of each mixture component using mixing laws to approximate the behavior of the entire mixture. The resulting semi-empirical models calculate thermodynamic and physical properties, such as density, enthalpy, and entropy of gas mixtures for known pressures and temperatures.
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FIG. 16.8 Left—recirculation path used for casing treatment, right—predicted and measured range extension for two different impellers, type A with Ns 1.00 and type B with Ns 0.85 [14].
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For typical dry natural gas used in midstream transmission, the most recently developed EOS—GERG, NIST, and AGA8—produce highly accurate thermophysical property calculations, in part due to the significant amounts of recently available empirical data. However, there is still some error in predicting the critical point and behavior of natural gas mixtures near the phase envelope; the (P-R) EOS typically provides more accurate results for this calculation. In addition, various software that use the same equation of state, implement the mixing laws and associated calculations in different ways thus giving varying results. For example, the thermophysical properties of a typical natural gas mixture calculated by the implementation of the SRK EOS using three different commercially available software, Multiflash, PVT Sim, and Refprop, was shown by Ridens et al. to result in a deviation in centrifugal compressor performance calculations of over 5% [16]. For mixtures containing heavier hydrocarbons, sour or acid gas components, or high CO2 content as well as for all gas mixtures with operating points near the critical phase, high pressures, or dense phase (supercritical) operation the EOS predictions still have significant error margins since mixing rules and binary mixture functions do not provide sufficient estimations of the behavior of the multicomponent mixture and there is little data available for model calibrations [2]. Highly accurate gas property testing can be performed for specific gas mixtures rather than using EOS calculations to obtain density, speed of sound and enthalpy rise at particular operating points. Directly measuring the properties of a specific gas mixture at various temperatures and pressures is particularly valuable when performed prior to purchasing and sizing large equipment or when operating custody transfer metering sites. The following methods are used to test physical properties of gasses. Density (or specific volume) is measured using separate measurements of gas mass and internal volume of a pressurized test fixture. The masses of the gas mixture and test fixture are measured using a precision industrial laboratory mass comparator scale. Speed of sound (or sound velocity) measurements employ the fundamental acoustic theory by exciting resonant acoustic modes and recording their respective amplitudes and frequencies. A standing wave is created in a high-pressure fixture that is amplified at frequencies associated with the speed of sound of the gas within the fixture. Enthalpy rise measurements are made by using a near instantaneous compression process to compress a test gas in an adiabatic autoclave.
Linear Motor Compressor A linear motor reciprocating compressor (LMRC) was developed meet the DOEs goal of increasing the efficiency and reducing the cost of forecourt hydrogen compression. This advanced compression system utilizes a novel and patented concept of driving a permanent magnet piston inside a hermetically sealed compressor cylinder through electromagnetic windings. The LMRC is an
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improvement over conventional reciprocating compressors as it minimizes the mechanical part count, reduces leakage paths, and is easily modularized for simple field installation (US Patent 8,534,058) [17]. The LMRC has the following advantages over conventional reciprocating compressors: l
l
l
System Complexity—Conventional state-of-the-art reciprocating compressors utilize double-acting pistons, each connected to a rod, crankshaft, coupling, and separate motor/driver. This arrangement is mechanically complex and relatively inefficient, as it consists of multiple moving parts that require bearings, seals, and lubrication, which can require frequent maintenance intervals. The fundamental piston actuation of the LMRC is achieved through magnetic forces. Since the LMRC replaces a crankshaft-type piston design with a sealed cylinder, deterioration of the wear bands is minimized and piston rod packing is avoided. Therefore, component wear and associated maintenance are significantly reduced. Mechanical Performance—Due to the many moving parts of a conventional reciprocating compressor, speed variation and flow capacity control make it difficult to avoid vibration problems. Each moving part has its own natural vibration frequency and high-cycle fatigue limitation. The LMRC uses fewer moving parts, and thus, it will be easier to design the system to avoid coincidence with component mechanical natural frequencies. Pulsation Control—Pulsations generated by a conventional reciprocating compressor have to be damped/attenuated using pulsation filter bottles, orifice plates, choke tubes, and/or Helmholtz resonators, all of which add cost and complexity and reduce compressor efficiency. The LMRC will operate at very low speeds. As a result, pulsations are expected to be much less problematic, and most of the above-mentioned pulsation attenuation devices should not be necessary.
In addition to the advantages over conventional reciprocating machinery noted above, the LMRC is able to improve the efficiency of the hydrogen compression process by using low-friction bearings, low loss valves, an optimized piston motion profile, and a slow piston stroke, which minimizes aerodynamic losses.
Project Goal The goal of the LMRC is to meet the following metrics: l
l
Improve isentropic efficiency above 95% by minimizing aerodynamic losses and using low-friction bearings Reduce capital costs to half those of conventional reciprocating compressors by minimizing part count
l
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Reduce required maintenance by simplifying the compressor design to eliminate common wear items Design a system using the LMRC to compress hydrogen from 2 MPa absolute to 87.5 MPa absolute with flow rates greater than 10 kg/h
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Technical Overview The LMRC is a novel concept compared to conventional reciprocating compression technology. The basic principles of reciprocating compressor design have shown that lower piston speeds and gas flow velocities are necessary to maintain isentropic efficiencies within 5% points of the isentropic limit. In a low-speed reciprocating compressor, the piston imparts energy on a stationary gas resulting in minimal aerodynamic losses, especially when recirculation and friction losses are well controlled. Utilizing interstage cooling reduces the initial enthalpy of the gas per stage, which keeps the gas at a lower energy state and requires less compression power. The LMRC system uses these principles to keep parasitic losses minimized, using reduced piston speeds, low-pressuredrop contoured valves, and interstage cooling manifolds. Working at low reciprocating speeds of approximately 300 cycles per minute (CPM) (5 Hz), the LMRC is expected to meet an isentropic efficiency target of greater than 95% [14]. That efficiency can be compared with current state-of-the-art technology that typically has an efficiency of closer to 73%. Fig. 16.9 shows a screen shot of a 3D model of the currently designed LMRC. The compression system replaces the functions of an electric motor drive and reciprocating compressor with an integrated, linear, and electrically actuated piston. It has a magnetic piston within a cylinder and a gas compression chamber at each end of the piston. The compressor cylinder is comprised of an electromagnetic coil that is operable with the piston to convert an input of electrical power to a reciprocating movement of the piston. This uses the same technology seen in magnetic bearings in turbomachinery and does not require oil for lubrication. Since the driver and compressor are integrated into the same hermetically sealed component, there is a significant reduction in the number of parts and materials needed to construct this device. In addition, the simplicity of the design reduces required maintenance, minimizes seal leakages and wear, and allows for oil-free operation. As mentioned previously, the isentropic efficiency is improved with the LMRC design and mechanical losses are reduced by reducing secondary systems. This results in an increase in overall efficiency for the system. The technology is based primarily on conventional parts used in many electromagnetic devices combined with reciprocating compressor cylinders and differential-pressure-activated valves. Traditional reciprocating compressor components, such as a separate driver, crankshaft, and connecting rods, are not required in the LMRC design. A photo of the prototype LMRC mounted in the test stand is depicted in Fig. 16.10. Upon final development of the LMRC, it should be feasible to scale the LMRC to a full-scale version. It is anticipated that the compact nature of the design would create a smaller footprint than any existing reciprocating compressor design. In addition, the low operating speeds, direct power input to the piston, and aerodynamic valves would make the overall compression
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FIG. 16.9 Schematic representation of LMRC concept.
process more efficient. Reducing the parts count by integrating the driver and compression chamber is expected to not only reduce the overall package size, but also significantly reduce the system maintenance costs. Once the LMRC is in full production mode, it is anticipated that the compressor package cost and maintenance costs will be approximately half those of a standard, conventional compressor.
Advanced Pulsation Control for Reciprocating Compressors Several advances in pulsation control for reciprocating compressors have been developed over the last 10 years to reduce the pressure losses associated with the orifice plates and choke tubes in a traditional pulsation filter bottle system.
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FIG. 16.10 Test stand construction.
The first concept is the tunable choke tube. The tunable choke tube varies the diameter of an internal or external pulsation bottle choke tube by using multiple choke tubes that can be opened or closed. This concept was developed in 2012 as a method of tuning the choke tube diameter as a function of the compressor speed [18] to allow for effective pulsation control over the entire speed range, while maintaining minimal losses throughout the speed range. The virtual orifice (VO) and tunable side branch absorber (TSBA) are pulsation control devices developed under Gas Machinery Research Council (GMRC) funding [19]. The acoustic response of the cylinder nozzle is typically damped using orifice plates, which can create significant pressure losses. The VO was developed to virtually eliminate pressure loss while attenuating the cylinder nozzle response. A prototype and subsequent versions of the VO were developed and installed in an operating compressor resulting in significant pulsation attenuation. An example set of data from a field installation is shown in Fig. 16.11. The side branch absorber is a commonly used pulsation device to attenuate a single frequency (small frequency range) problematic resonance in the piping system. This concept was advanced through incorporating a variable volume
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Pulsation amplitude (psi pk-pk)
Max. amplitude = 52 psi 50
No orifice installed Orifice installed VO installed
40
Max. amplitude = 30 psi 30
20
Max. amplitude = 15 psi 10
0 0
20
40
60
80
100
120
140
160
180
200
Frequency (Hz)
FIG. 16.11 Virtual orifice prototype and test results.
FIG. 16.12 Tunable side branch absorber conceptual design. (Image from ACI Services Inc.)
(Fig. 16.12) to expand the frequency range of attenuation for the side branch absorber thus making it more effective over a range of operating speeds.
Advanced Reciprocating Compressor Valve Technology Valve performance, which includes both life and efficiency, must be considered when operating a reciprocating compressor. In a reciprocating compressor valve, desirable functional attributes include good sealing, rapid opening and closing, sustained high flow area (when open), minimum bouncing upon impact, toleration of impact forces and maximum temperatures, and low flow resistance. In most traditional valve designs, such as the ring, plate or poppet designs, a trade-off is made between life and efficiency.
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Gas flow in
Valve housing
Valve plate
Valve shaft
Gas flow out Control box Coil Magnets FIG. 16.13 Semi-active valve concept.
Advances in existing valve designs have been made through coatings and improved materials that reduce the impact wear, allow for corrosive and erosive gas compositions, and higher temperatures and pressures; additionally, various valve profiles have been developed to reduce performance losses [20]. Novel designs include a semi-active valve and the reed valve [21, 22]. Valve life is reduced when the valve passively slams shut or open due to the pressure differential across the valve; the semi-active valve employs magnets to slow down the closing of the valve, thus extending its life. The semi-active valve conceptual design is shown in Fig. 16.13 with the Fig. 16.14 showing the semi-active valve opening motion profile in red as compared to that of a typical plate valve shown in blue. The reed valve is based on the design of reeds commonly used in wind instruments. This valve consists of a reed contained in an easily replaceable module that is contained in the valve seat; since the design is a direct flowpath, there are low pressure losses associated with this type of valve (Fig. 16.15).
Surge Force Predictions Centrifugal compressor impellers and shafts are subject to fluctuating axial and radial forces when operating in surge. These forces can cause severe damage to the close clearance components of a centrifugal compressor such as the thrust and radial bearings, interstage and dry gas seals, and balance piston. Being able
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1400
Position [Volts]
1200 1000 800 Lower impact reduces plate stresses
600 400 200 0 120.00
140.00
160.00
180.00 Time [ms]
200.00
220.00
FIG. 16.14 Typical plate valve opening motion (blue) vs. semi-active valve opening motion (Red) [23].
FIG. 16.15 Reed valve concept [24].
to accurately quantify the frequency and amplitude of the cyclic surge forces on the close clearance components of the compressor allows the user to determine whether an accidental surge event, or emergency shut-down (ESD) transient, has caused damage requiring inspection, repair, or part replacement. In addition, the results can be used when designing a surge control system to determine if a hot gas bypass line is necessary or for choosing the anti-surge valve and control system parameters. In a compressor with no unbalance, misalignments, or rubs, surge- and axialsymmetric phenomenon-causes minor radial vibration and high-axial
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TABLE 16.1 Potential Damage Γ <0.001
Mild surge (low potential damage)
0.01<Γ < 0.001
Classic surge (discrete potential damage (requires further analysis)
Γ >0.01
Deep surge strong (high potential damage)
displacements due to cyclic thrust reversal. Thus, the strongest surge forces are those in the axial direction and the highest risk of failure component is the compressor’s axial thrust bearing. A method was developed, based on experimental test data performed for the GMRC [3], to quantify the cyclic axial surge forces on the close clearance components of the compressor allowing the user to determine whether an accidental (or ESD transient) surge event causes sufficient damage to require inspection or repair or as justification for omitting it. The methodology can be captured accurately in a transient one-dimensional (1D) lump sum parameter bond graph method dynamic analysis as demonstrated by Pinelli et al. [25]. In this work by Brun et al. [3], a surge severity coefficient was derived as shown in Eq. (16.1), based on energy conservation nondimensional analysis. This coefficient, Γ, takes into account the main parameters which affect the potential damage due to a surge event. Calculations above 0.001 suggest that an in-depth analysis be performed to determine the actual force amplitude in relation to the characteristics of the compressor to evaluate the risk of damage (see Table 16.1). Γ¼
r 2 VΔP2 smkΔh
(16.1)
where k—axial thrust-bearing stiffness m—mass of the rotor r—impeller tip radius s—axial thrust-bearing normal operational clearance V—compressor discharge flange to check valve volume Δ h—compressor enthalpy rise (head) Δ P—pressure difference across compressor (Pd Ps).
Liquid Packing Seals for Reciprocating Compressors The highest source of methane emissions in reciprocating compressors is leakage of process gas through the sealing components in the packing systems around the piston rods as shown in Fig. 16.16 [25]. Current technology uses a series of specifically-cut, dry-ring seals held in place with springs and cups (see Fig. 16.17). However, there is a trade-off made between leakage reduction with minimal gaps between the seals and the rod versus allowing sufficient gaps
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FIG. 16.16 Gas leakage path of reciprocating compressors. (Image courtesy of Dresser Rand.)
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FIG. 16.17 Cut away view of currently used reciprocating compressor rod packing sealing system [16]. (Image courtesy of EPA.)
such that the friction between the parts is sufficiently reduced to allow movement. Once the piston moves, the pressure differential across the packing seals creates a twisting effect on the seal, allowing substantial amounts of natural gas to leak into the casing. Ring twisting also causes increased friction and wear to the sealing rings and compressor rod. Using liquid, rather than gas, for essentially eliminating process gas leakage through the piston rods is not a novel concept, but does come with several challenges that have prevented its widespread implementation in production machines. The primary challenge is creating a system in which the pressure between the process gas and the sealing liquid is continuously balanced such that gas does not leak into the liquid and liquid does not leak into the gas. Reciprocating compressor pistons and valves create continuous high-frequency pressure changes that make instantaneous pressure balancing difficult. Other challenges include incorporating the support and monitoring systems for the liquid (pump, tank, filter, valves, flow-rate monitoring). The DOE has sponsored the development of a novel liquid packing seal prototype (2017–2019) and as of 2017, and one commercial company has manufactured and installed a small number of a different prototype of a liquid seal in currently operating compressors.
References [1] G.L. Agrawal, Foil air/gas bearing technology—an overview, in: Paper No. 97-GT-347, International Gas Turbine & Aeroengine Congress & Exhibition, June 2–5, Orlando, FL, 1997. [2] N. Baltadjiev, C. Lettieri, Z. Spakovszky, An investigation of real gas effects in supercritical CO2 centrifugal compressors, in: Proceedings of the ASME Turbo Expo 2014: Turbine Technical Conference and Exposition, Paper No. GT2014-26180, June 16–20, Dusseldorf, Germany, 2014. [3] J.L. Brady, J.M. Klein, M.D. Stevenson, S.P. Petullo, N.C. D’Orsi, R.C. Skinner, K.D. Eager, R.A. Perry, Downhole gas separation and injection powered by a downhole turbo expander, in: Proceedings of SPE Annual Technical Conference and Exhibition, SPE 49051, New Orleans, Louisiana, USA, September 27–30, 1998. [4] A. Rimpel, T. Allison, J. Moore, J. Grieco, P. Shy, J. Klein, J. Brady, Open-loop aerodynamic performance testing of a 105,000 RPM oil-free compressor expander for subsurface natural gas
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[5]
[6]
[7] [8]
[9]
[10]
[11] [12]
[13] [14] [15]
[16]
[17] [18] [19] [20] [21] [22]
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compression and reinjection, in: Proceedings of ASME Turbo Expo 2012, GT2012-69017, Copenhagen, Denmark, June 11–15, 2012. A. Rimpel, J. Moore, J. Grieco, P. Shy, J. Klein, J. Brady, Rotordynamics of a 105,000 RPM oil-free compressor-expander for subsurface natural gas compression and reinjection, in: Proceedings of ASME Turbo Expo 2012, GT2012-69119, Copenhagen, Denmark, June 11–15, 2012. L. San Andres, A. Anderson, Comparison of leakage between a labyrinth seal and an all-metal compliant gas seal at high temperature, in: Report to TRC-S-02-2011, Turbomachinery Research Consortium, Texas A&M University, May, 2011, 2011. L. San Andres, A. Anderson, Leakage for an all-metal compliant gas seal operating at high temperature, in: STLE Annual Meeting & Exhibition, Detroit, MI, May 5–9, 2013. L. San Andres, A. Anderson, An all-metal compliant-seal versus a labyrinth seal: a comparison of gas leakage at high temperatures. ASME J. Eng. Gas Turb. Power 137 (2015) 052504, https://doi.org/10.1115/1.4028665. ASME Paper GT2014-25572. D. Stiles, J. Justak, M. Kuzdzal, H. Miller, M. Sandberg, E. Wilcox, C. Rohrs, Application of dynamic pressure balanced seals in a multi-stage centrifugal compressor, in: 45th Turbomachinery & 32nd Pump Symposia, Houston, TX, September 12–15, 2016. C. DellaCorte, M. Valco, Load capacity estimation of foil air journal bearings for oil-free turbomachinery applications, in: STLE International Joint Tribology Conference, October 1–4, Seattle, WA, 2000. A. Delgado, Experimental identification of dynamic force coefficients for a 110 mm compliantly damped hybrid gas bearing, ASME J. Eng. Gas Turb. Power 137 (7) (2015) 072502. Delgado, A., and Ertas, B., 2018, “Dynamic characterization of a novel externally pressurized compliantly damped gas-lubricated bearing with hermetically sealed squeeze film damper modules,” Paper No. GT2018-7212, ASME Turbo Expo 2018, June 11-15, Oslo, Norway. Y. Wang, D. Kim, Experimental identification of force coefficients of large hybrid air foil bearings, ASME J. Eng. Gas Turb. Power 136 (3) (2014) 032503. Deffenbaugh, D., et al., Advanced Reciprocating Compression Technology, DOE Award No. DE-FC26-04NT42269, SwRI Contract No. 18.11052, 2005. S. Jung, R. Pelton, Numerically Derived Design Guidelines of Self Recirculation Casing Treatment for Industrial Centrifugal Compressors, in: Proceedings of the ASME Turbo Expo 2016: Turbomachinery Technical Conference and Exposition, Paper No. GT2016-56672, June 13–17, Seoul, South Korea, 2016. B. Ridens, S. Simons, S. Coogan, K. Brun, R. Kurz, GMRC/PRCI project: equation of state comparisons and evaluations for applications through gas property testing and derivations, in: Gas Machinery Conference, 2016. K. Brun; M.A. Wilcox; E.L. Broerman. U.S. Patent 8,534,058. Energy Storage and Production Systems, Apparatus and Methods of Use Thereof, Patented in United States of America, 2013. Deffenbaugh, D., McKee, R., Nored, M., Tunable Choke Tube for Pulsation Control Device Used with Gas Compressor. Publication number: 20090191076, 2009. E. Broerman, M. Nored, R. McKee, Advancements in pulsation control technology, in: Gas Machinery Conference, 2007. K. Brun, M. Nored, Valve performance and life of reciprocating compressors, in: Proceedings of the 41st Turbomachinery Symposium, September 24–27, Houston, Texas, 2012. K. Brun, J. Platt, R. Gernentz, M. Smolik, Semi-active compressor valve development and testing, in: Gas Machinery Conference, Oklahoma City, Oklahoma, October 2–4, 2006. World Oil Magazine, Modular Reed Valves Improve Efficiency for Gas Gathering and Distribution, 2013.
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[23] J. Moore, A. Lerche, H. Delgado, T. Allison, J. Pacheco, Development of advanced centrifugal compressors and pumps for carbon capture and sequestration applications, in: Proceedings of the Fortieth Turbomachinery Symposium, September 12–15, Houston, Texas, 2011. [24] J. Moore, N. Evans, J. Pacheco, T. Allison, J. Kerth, Development and testing of multi-stage internally cooled centrifugal compressor, in: 44th TurboMachinery and 31st Pump Symposia, Houston, Texas, 2015. [25] M. Pinelli, M. Morini, E. Munari, K. Brun, S. Simons, R. Kurz, Measurement and prediction of centrifugal compressor axial forces during surge: part 2-surge force measurements, in: Proceedings of ASME Turbo Expo 2017: Turbomachinery Technical Conference and Expositions GT2017-63061, June 26–30, Charlotte, North Carolina, 2017.
Further Reading [26] K. Brun, S. Simons, R. Kurz, M. Pinelli, M. Morini, E. Munari, Measurement and prediction of centrifugal compressor axial forces during surge: part 1-surge force measurements, in: Proceedings of ASME Turbo Expo 2017: Turbomachinery Technical Conference and Expositions GT2017-63061, June 26–30, Charlotte, North Carolina, 2017. [27] R. Subramanian, L.L. Williams, T.L. Vaughn, D. Zimmerle, J.R. Roscioli, S.C. Herndon, et al., Methane emissions from natural gas compressor stations in the transmission and storage sector: measurements and comparisons with the EPA greenhouse gas reporting program protocol, Environ. Sci. Technol. 49 (2015) 3252–3261. [28] US EPA, Reducing Methane Emissions from Compressor Rod Packing Systems, US Environmental Protection Agency, 2006. http://www3.epa.gov/gasstar/documents/ll_rodpack.pdf. [29] Zahroofvalves.com/pioneering-design.html. [30] D.J. Zimmerle, L.L. Williams, T.L. Vaughn, C. Quinn, R. Subramanian, G.P. Duggan, et al., Methane emissions from the natural gas transmission and storage system in the United States, Environ. Sci. Technol. 49 (2015) 9374–9383.
Novel Concepts & Research Sarah Simons*, J. Jeffrey Moore*, Karl Wygant†, Harry Miller‡, Eugene "Buddy" Broerman* and Aoron Rimpel§ *
Southwest Research Institute, San Antonio, TX, United States, †Hanwha Power Systems Americas Inc., Houston, TX, United States, ‡Dresser-Rand, A Siemens Business, Olean, NY, United States, § Southwest Research Institute, San Antonio, TX, United States
The following novel concepts for oil and gas machinery are under development but are not (as of the publication date of this book) offered in production machines.
Cooled Diaphragms for Centrifugal Compressors Isothermal compression is one of the recent areas of research and development in centrifugal compressors. To reduce the power penalty associated with a large number of compression stages, internally cooled diaphragms were developed under a Department of Energy (DOE) research project to increase the efficiency of CO2 sequestration from power plants. Through three-dimensional (3D) computational fluid dynamics analysis, a design can be achieved to provide good heat transfer while adding no additional pressure drop [1]. This technology allows a barrel compressor to achieve performance close to an integrally geared compressor. Fig. 16.1 shows the conceptual cooled diaphragm design where cooling flow (blue) is routed through the diaphragms adjacent to the gas flow path (red). The total temperature increases, due to the work input of the impeller, are reduced through the diaphragm flow path thereby reducing the temperature into the downstream stage. Concepts of printed circuit heat exchangers were used for the basis of developing an efficient heat transfer mechanism to ensure that the compression process is nearly isothermal. A new compressor was fabricated based on a Dresser-Rand DATUM D12 frame size. The compressor consisted of a six-stage, back-to-back centrifugal compressor (D12R6B) that incorporated the cooled diaphragms. The compressor and associated piping flow loop are shown in Fig. 16.2 [2]. Aerodynamic testing of the compressor in several operating configurations was completed with and without the cooling diaphragms activated. Results showed that the Compression Machinery for Oil and Gas. https://doi.org/10.1016/B978-0-12-814683-5.00016-X © 2019 Elsevier Inc. All rights reserved.
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FIG. 16.1 Cooled diaphragm concept [1].
FIG. 16.2 Left—installed Dresser-Rand DATUM compressor. Right—DATUM compressor test loop [2].
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cooled diaphragm technology reduces compressor power consumption by 3.0% near surge to 8.0% near choke when compared with the adiabatic case with intercooling between the two sections. Additional savings would result if the machine was configured as an eight-stage compressor, which would provide two additional internal heat exchangers. A 9.0% power savings was measured when the compressor was operated as a straight-through compressor with no intercooling at the design point and would be higher (>10%) at larger flows. The heat exchanger effectiveness for the cooled diaphragm was measured between 12% and 30%, depending on the stage, operating point, and backto-back vs straight-through intercooling configurations. The cooled diaphragms removed 297–311 K (28%–35%) of the temperature rise within each section when compared to the adiabatic case. The measured temperature drop, heat exchanger effectiveness, and power savings were all slightly higher than predicted values. Operation of the cooled diaphragms changed the characteristics of the multistage machine, increasing flow capacity and pressure ratio compared to adiabatic performance at the same speed. Additional performance gains may be realized by designing the compressor aerodynamic flow path for the cooled case rather than the adiabatic case. No reliability issues associated with the cooled diaphragm design were encountered during testing including diaphragm leakage. While these results provided were for a CO2 application, benefits can be realized for any high-pressure ratio application or where reducing discharge temperatures is desired (e.g., preventing ethylene polymerization).
Subsurface Compression To increase the hydrocarbon recovery in oil fields, a subsurface process and reinjection compressor (SPARC) tool was developed to process, compress, and reinject up to 30% of the gas in a high gas-to-oil ratio well (defined by greater than 15,000 standard cubic feet of gas per stock tank barrel of oil). This would allow additional black oil to be produced to the increased gas rate production and increase the gas condensate production since the recycled gas will carry the condensates to the production stream [3]. The primary goal of the tool design is to provide additional gas handling capacity while maintaining reservoir pressure to existing production fields at a fraction of the cost of a surfacebased operation. Fig. 16.3 represents the SPARC concept presented by Brady et al. [3] and will be used to describe the tool’s basic functions. The SPARC device includes a three phase separator at the bottom of the tool which will prevent most liquids and solids from passing through the turbo-expander or turbine (separator 2 is labeled as pre-swirl auger in Fig. 16.3). The remaining solids and liquids bypass around the turbomachinery section. The energy in this production stream is usually wasted by passing through a series of chokes within the production system. A turbine bypass valve (not shown) allows the gas flow to enter the turbine only after the well has been cleaned up and is free from debris usually seen while
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FIG. 16.3 Schematic representation of SPARC concept for functional description [3].
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bringing a shut-in well back online. Until these conditions are achieved, the valve remains closed and flow bypasses the turbomachinery section entirely. Reservoir gas is filtered (not shown) and used to cool the oil-free bearings. The production stream is expanded through the turbine, rejoins with the bypass flow, and is separated once again by an auger separator—the gas is directed to the compressor suction while the other solids/liquids continue to the surface. A recycle valve is utilized to prevent the compressor from operating in surge conditions. The gas which exits the compressor is finally reinjected back into the formation. The SPARC test rig components and rotor assembly are shown in Fig. 16.4. The design speed of the turbine and compressor is 105,000 rpm (105 krpm), and the design compressor power consumption at down hole conditions is approximately 343 kW. Given that the turbine weighs approximately 0.907 N and the combined compressor impeller weight is approximately 0.703 N, the power-toweight ratio of the SPARC is an order-of-magnitude larger than a typical commercial centrifugal impeller. Gas-lubricated foil bearings are utilized using the gas following clean-up process. Aerodynamic open-loop testing was performed on the SPARC tool. Several performance curves were generated for each component ranging from 73.5 to 115.5 krpm (70%–110% of design speed) [4]. In general, measured compressor head was approximately 10% lower than predictions, and measured efficiencies for the 84 krpm speed line were very close to the predicted values. The turbine had higher flow than predictions, due in part to a larger flow area in the manufactured turbine wheel. Rotordynamic open-loop testing was also performed which showed an instability of the conical rigid rotor mode at the design operating speed of 105,000 rpm [5]. It was found that the amount of stiffness the seals contribute varies with operating pressures which changes the frequencies of the conical and cylindrical modes. This was resolved by replacing the holepattern seals with higher clearance labyrinth seals and stiffer bearings; subsequent testing showed no stability issues over the operating speed range.
FIG. 16.4 Left—SPARC test rig components: (A) test rig housing assembly, (B) large labyrinth seal, (C) compressor balance piston hole-pattern seal, (D) stage 2 compressor impeller, (E) stage 1 compressor impeller, (F) interstage diaphragm section, (G) stage 2 diffuser and de-swirl cascade, (H) turbine hub hole-pattern seal, and (I) complete rotor assembly. Right—Rotor assembly.
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Advanced Seals The current state-of-the-art method for sealing flow in a centrifugal compressor between the rotating and stationary components consists of several types or styles of seals, such as plain stationary labyrinth seals, rotating labyrinth seals, stepped labyrinth seals, pressure activated leaf seals, finger seals, and brush seals. An alternative dynamic pressure-balanced (DPB) seal uses the flow and pressure drop associated with the fluid it is sealing to generate a force-balanced seal system with an operational clearance as tight as a few thousandths of an inch (hundredths of a millimeter) over the rotor surface. A cross-section of the DPB seal is shown in Fig. 16.5. The DPB seal is made up of a circular array of shoe segments (see Figs. 16.5 and 16.6). Each shoe segment acts independently and is connected to the static hardware via a spring support. The seal is typically installed with a nominal gap that is larger than the desired running clearance. The acceleration of fluid between the seal and the rotor creates a lowpressure region that draws the seal shoes toward the rotor. As the seal-surface approaches the rotor surface, features within the seal surface reduce the velocity of the fluid, and when combined with static upstream pressure results in a pressure rise that increases the radial outward force on the seal shoes. The operational clearance is achieved when this outward force is balanced with the inward force from the upstream and downstream pressures acting on the backside of the seal. Seal dimensions are tuned to achieve the desired operational clearance. This method of creating a force balance over the rotor surface facilitates a seal that maintains a set clearance during dynamic changes in pressure,
Retaining ring
Secondary seals
Scalloped front plate Shoe (15 pl.)
Rotor
Flow direction Beams (2 pl. per shoe)
Radial stop
FIG. 16.5 Cross section of the DPB seal.
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FIG. 16.6 DPB seal photo.
temperature, centrifugal growth, and rotor eccentricity. The nature of operation is not negatively impacted by swirl ratio or surface velocity associated with high-speed operation. DPB seals are currently being developed and tested for the following reasons: DPB seals have been shown in laboratory settings to maintain design flow leakages (minimizing wear) for longer run times thus extending service intervals, increase aerodynamic efficiency by up to 2.0%, and allow a more forgiving operating envelope for the rotor which increases rotor dynamic stability [6–8]. Testing of these seals in an ASME PTC-10 Type 1 test of a 31-MPa absolute discharge pressure gas lift centrifugal compressor found that 2.8% lower power was required for the same head level across the entire range of inlet flows and pressure ratios, compared to the same test with conventional labyrinth seals [9]. In addition, the system rotordynamic stability, obtained via operational modal analysis (OMA), showed the DPB) seals exhibited log decs similar to standard labyrinth seals across the entire range of flows and pressure ratios.
Gas Bearings Typical fluid film bearings utilize oil as a lubricant to generate the necessary hydrodynamic pressure to support a load. However, bearings utilizing gaseous lubricant is also possible. This is advantageous when the process gas of a machine is able to serve in this role. For example, air cycle machines on aircraft have utilized air bearings since the 1980s [1]. Gas bearings for turbomachinery have notable advantages over traditional oil film bearings, namely, the elimination of sealing systems to separate oil lubricant from the process and reduced parasitic power loss due to lower fluid viscosity. Also, compared to rolling element bearings, gas bearings do not have speed limits due to high contact stresses from centrifugal loading. The only other “oil-free” competitor to gas bearings
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are active magnetic bearings. Both gas bearings and magnetic bearings can be utilized in hermetically sealed machines, but the magnetic bearing requires control systems and back-up bearings, which add complexity and cost. The advantages of gas bearings over other bearing types allow gas-bearing machines to be more compact, less complex, operate cleaner and more efficiently, and operate at higher relative speeds. However, due to the lower viscosity of a gas compared to oil lubricants, disadvantages of gas bearings are low load capacity and low damping. This has limited gas bearing use to socalled microturbomachines, that is, machinery with outputs on the order of hundreds of kilowatts or less. Advancements in gas-bearing technology have tried to address the load capacity and damping issues in order to permit their use in larger machines. In compliant gas bearings, such as foil bearings, progress has been achieved by tailoring the compliance structure stiffness profile to the gas film stiffness [10]. This ensures a more uniform hydrodynamic film thickness and enhances load capacity. However, hydrodynamic-only load capacity is usually insufficient for gas-bearing machines with larger, heavier shafts (shaft weight scales with length to the third power). For example, shaft sizes larger than 100 mm generally require additional lift capability via hydrostatic pressure [11–13]. Fluid film damping in a gas bearing is often supplemented by external damping mechanisms. For example, foil bearings have friction damping at the interfaces with the underspring, which may be corrugated bump foils, layers of foil leaves, or metal meshes. However, the most-effective supplemental damping scheme for gas-bearing applications targeting larger-size machines appears to be with the incorporation of squeeze film dampers. In this configuration, the squeeze film damper utilizes oil like a conventional squeeze film damper, but the damper has to be hermetically sealed for the bearing to remain “oil-free.” Delgado and Ertas [12] presented damping results for a 110-mm hydrostatic gas bearing with hermetically sealed squeeze film dampers, and they showed 200%–300% increase in damping compared to a similar bearing with metal mesh dampers [11]. Gas-bearing technology offers numerous advantages over conventional oillubricated fluid film bearings, rolling element bearings, and magnetic bearings. The drawbacks of lower load capacity and limited damping have prevented gasbearing use in larger-sized turbomachinery. However, technical advancements have improved the performance of gas bearings over the years, where continually larger applications seem realistic. Specifically, utilizing hydrostatic lift features and hermetically sealed squeeze film dampers represents the current state of the art for large-scale gas-bearing machinery.
Additive Manufacturing Additive manufacturing has the potential to be the single largest driver of change to the turbomachinery industry in the past 100 years. The reason for this
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FIG. 16.7 Closed compressor impeller manufactured by direct metal laser sintering of metal particles. Left—raw manufactured, right—cleaned.
bold statement is that additive manufacturing has a direct impact on potentially every component of the turbomachine design. That impact extends to allowing substantial reductions in stress and the ability to implement design concepts that could not previously be manufactured. Various additive manufacturing techniques are emerging. Fig. 16.7 shows a compressor impeller that was manufactured by direct metal laser sintering (DMLS). DMLS uses a high power-density laser to melt and fuse metallic powders into a solid component. The process can leave micro-voids in the part that are closed by a hipping process. There is variations between different metals, but in general the DMLS process produces a part with comparable yield and ultimate strength as forged material properties while some decrement is present in fatigue strength. The ability to apply localized fillet and strengthening can more than offset the loss of fatigue strength margin. Manufacturers are investigating the use of additive manufacturing for a wide range of turbomachinery components. These components are: guide vanes, impellers, integral impellers and shafts, etc. For gas turbines there is active work on applying the technology for combustor nozzles and turbine blades. The impact to the industry is that greater levels of customization is present, novel configurations can be explored, welds and brazed connections can be eliminated, and the process produces very little waste material as unused metal powders is salvaged after the build is complete. Currently the obstacles in the technology are related to relatively small components 300 300 300 mm and quality assurance challenges. Since the parts created can be complex shapes with internal geometries the ability to check for voids and/or deformities is challenging. As the size of the additive manufacturing machines increases and inspection processes mature the application of the technology will only increase in the turbomachinery industry eventually necessitating that design engineers rethink the entire design manufacturing process.
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Casing Treatment (Range Extension) Casing treatment has been applied in automotive turbocharger industry for decades. The technology allows the ability to substantially extend compressor range at minimal impact to efficiency and head rise. The technology has not previously been applied in industrial turbomachinery due to the need for extensive optimization of the configurations. In the automotive industry extensive design optimization occurs and then turbochargers are mass produced and the relatively high-pressure ratio of those turbochargers. In the industrial machinery industry the mass production volumes are substantially smaller and the end users more reticent on accepting innovation and typically pressure ratios smaller (thereby naturally giving better range than higher-pressure ratio stages.) Casing treatment is a compressor range extension technology that can be easily integrated in to most compressor systems without significant additional expense. Casing treatments have gained acceptance in some applications since the technology can be implemented with few additional stationary parts to maintain a robust and reliable configuration. The casing treatment can extend the range of a stage by delaying the onset of stall in the inducer of the compressor. Inducer stall occurs due to the high levels of incidence and loading near the leading edge of the blade. Flow recirculation through the casing treatment increases the inducer flow rate and helps reduce the leading edge loading. This delays the onset of inducer stall and results in enhancement of compressor map width. Fig. 16.8 (left) shows the flow recirculating from the inducer back to the inlet through the casing treatment cavity near stall for a typical design. Fig. 16.8 (right) shows the amount that the compressor range was extended for two different impellers, Type A with a high-specific speed design, Ns ¼ 1.00 and Type B with a medium-specific speed design, Ns ¼ 0.85 [15].
Gas Property Testing Equations of State (EOS) are used to determine the performance of compressors, size and select various types of equipment, and evaluate pipeline hydraulics. There are several different commonly available models developed over time with the most commonly used ones being GERG, NIST, AGA8, PengRobinson (P-R), Benedict-Webb-Rubin-Starling (BWRSE), and RedlichKwong (R-K). EOS were first developed based on empirical data for pure substances as a method of relating the pressure, density, and temperature of a gas. Modern EOS extend these original equations to mixtures by combining the properties of each mixture component using mixing laws to approximate the behavior of the entire mixture. The resulting semi-empirical models calculate thermodynamic and physical properties, such as density, enthalpy, and entropy of gas mixtures for known pressures and temperatures.
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FIG. 16.8 Left—recirculation path used for casing treatment, right—predicted and measured range extension for two different impellers, type A with Ns 1.00 and type B with Ns 0.85 [14].
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For typical dry natural gas used in midstream transmission, the most recently developed EOS—GERG, NIST, and AGA8—produce highly accurate thermophysical property calculations, in part due to the significant amounts of recently available empirical data. However, there is still some error in predicting the critical point and behavior of natural gas mixtures near the phase envelope; the (P-R) EOS typically provides more accurate results for this calculation. In addition, various software that use the same equation of state, implement the mixing laws and associated calculations in different ways thus giving varying results. For example, the thermophysical properties of a typical natural gas mixture calculated by the implementation of the SRK EOS using three different commercially available software, Multiflash, PVT Sim, and Refprop, was shown by Ridens et al. to result in a deviation in centrifugal compressor performance calculations of over 5% [16]. For mixtures containing heavier hydrocarbons, sour or acid gas components, or high CO2 content as well as for all gas mixtures with operating points near the critical phase, high pressures, or dense phase (supercritical) operation the EOS predictions still have significant error margins since mixing rules and binary mixture functions do not provide sufficient estimations of the behavior of the multicomponent mixture and there is little data available for model calibrations [2]. Highly accurate gas property testing can be performed for specific gas mixtures rather than using EOS calculations to obtain density, speed of sound and enthalpy rise at particular operating points. Directly measuring the properties of a specific gas mixture at various temperatures and pressures is particularly valuable when performed prior to purchasing and sizing large equipment or when operating custody transfer metering sites. The following methods are used to test physical properties of gasses. Density (or specific volume) is measured using separate measurements of gas mass and internal volume of a pressurized test fixture. The masses of the gas mixture and test fixture are measured using a precision industrial laboratory mass comparator scale. Speed of sound (or sound velocity) measurements employ the fundamental acoustic theory by exciting resonant acoustic modes and recording their respective amplitudes and frequencies. A standing wave is created in a high-pressure fixture that is amplified at frequencies associated with the speed of sound of the gas within the fixture. Enthalpy rise measurements are made by using a near instantaneous compression process to compress a test gas in an adiabatic autoclave.
Linear Motor Compressor A linear motor reciprocating compressor (LMRC) was developed meet the DOEs goal of increasing the efficiency and reducing the cost of forecourt hydrogen compression. This advanced compression system utilizes a novel and patented concept of driving a permanent magnet piston inside a hermetically sealed compressor cylinder through electromagnetic windings. The LMRC is an
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improvement over conventional reciprocating compressors as it minimizes the mechanical part count, reduces leakage paths, and is easily modularized for simple field installation (US Patent 8,534,058) [17]. The LMRC has the following advantages over conventional reciprocating compressors: l
l
l
System Complexity—Conventional state-of-the-art reciprocating compressors utilize double-acting pistons, each connected to a rod, crankshaft, coupling, and separate motor/driver. This arrangement is mechanically complex and relatively inefficient, as it consists of multiple moving parts that require bearings, seals, and lubrication, which can require frequent maintenance intervals. The fundamental piston actuation of the LMRC is achieved through magnetic forces. Since the LMRC replaces a crankshaft-type piston design with a sealed cylinder, deterioration of the wear bands is minimized and piston rod packing is avoided. Therefore, component wear and associated maintenance are significantly reduced. Mechanical Performance—Due to the many moving parts of a conventional reciprocating compressor, speed variation and flow capacity control make it difficult to avoid vibration problems. Each moving part has its own natural vibration frequency and high-cycle fatigue limitation. The LMRC uses fewer moving parts, and thus, it will be easier to design the system to avoid coincidence with component mechanical natural frequencies. Pulsation Control—Pulsations generated by a conventional reciprocating compressor have to be damped/attenuated using pulsation filter bottles, orifice plates, choke tubes, and/or Helmholtz resonators, all of which add cost and complexity and reduce compressor efficiency. The LMRC will operate at very low speeds. As a result, pulsations are expected to be much less problematic, and most of the above-mentioned pulsation attenuation devices should not be necessary.
In addition to the advantages over conventional reciprocating machinery noted above, the LMRC is able to improve the efficiency of the hydrogen compression process by using low-friction bearings, low loss valves, an optimized piston motion profile, and a slow piston stroke, which minimizes aerodynamic losses.
Project Goal The goal of the LMRC is to meet the following metrics: l
l
Improve isentropic efficiency above 95% by minimizing aerodynamic losses and using low-friction bearings Reduce capital costs to half those of conventional reciprocating compressors by minimizing part count
l
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Reduce required maintenance by simplifying the compressor design to eliminate common wear items Design a system using the LMRC to compress hydrogen from 2 MPa absolute to 87.5 MPa absolute with flow rates greater than 10 kg/h
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Technical Overview The LMRC is a novel concept compared to conventional reciprocating compression technology. The basic principles of reciprocating compressor design have shown that lower piston speeds and gas flow velocities are necessary to maintain isentropic efficiencies within 5% points of the isentropic limit. In a low-speed reciprocating compressor, the piston imparts energy on a stationary gas resulting in minimal aerodynamic losses, especially when recirculation and friction losses are well controlled. Utilizing interstage cooling reduces the initial enthalpy of the gas per stage, which keeps the gas at a lower energy state and requires less compression power. The LMRC system uses these principles to keep parasitic losses minimized, using reduced piston speeds, low-pressuredrop contoured valves, and interstage cooling manifolds. Working at low reciprocating speeds of approximately 300 cycles per minute (CPM) (5 Hz), the LMRC is expected to meet an isentropic efficiency target of greater than 95% [14]. That efficiency can be compared with current state-of-the-art technology that typically has an efficiency of closer to 73%. Fig. 16.9 shows a screen shot of a 3D model of the currently designed LMRC. The compression system replaces the functions of an electric motor drive and reciprocating compressor with an integrated, linear, and electrically actuated piston. It has a magnetic piston within a cylinder and a gas compression chamber at each end of the piston. The compressor cylinder is comprised of an electromagnetic coil that is operable with the piston to convert an input of electrical power to a reciprocating movement of the piston. This uses the same technology seen in magnetic bearings in turbomachinery and does not require oil for lubrication. Since the driver and compressor are integrated into the same hermetically sealed component, there is a significant reduction in the number of parts and materials needed to construct this device. In addition, the simplicity of the design reduces required maintenance, minimizes seal leakages and wear, and allows for oil-free operation. As mentioned previously, the isentropic efficiency is improved with the LMRC design and mechanical losses are reduced by reducing secondary systems. This results in an increase in overall efficiency for the system. The technology is based primarily on conventional parts used in many electromagnetic devices combined with reciprocating compressor cylinders and differential-pressure-activated valves. Traditional reciprocating compressor components, such as a separate driver, crankshaft, and connecting rods, are not required in the LMRC design. A photo of the prototype LMRC mounted in the test stand is depicted in Fig. 16.10. Upon final development of the LMRC, it should be feasible to scale the LMRC to a full-scale version. It is anticipated that the compact nature of the design would create a smaller footprint than any existing reciprocating compressor design. In addition, the low operating speeds, direct power input to the piston, and aerodynamic valves would make the overall compression
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FIG. 16.9 Schematic representation of LMRC concept.
process more efficient. Reducing the parts count by integrating the driver and compression chamber is expected to not only reduce the overall package size, but also significantly reduce the system maintenance costs. Once the LMRC is in full production mode, it is anticipated that the compressor package cost and maintenance costs will be approximately half those of a standard, conventional compressor.
Advanced Pulsation Control for Reciprocating Compressors Several advances in pulsation control for reciprocating compressors have been developed over the last 10 years to reduce the pressure losses associated with the orifice plates and choke tubes in a traditional pulsation filter bottle system.
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FIG. 16.10 Test stand construction.
The first concept is the tunable choke tube. The tunable choke tube varies the diameter of an internal or external pulsation bottle choke tube by using multiple choke tubes that can be opened or closed. This concept was developed in 2012 as a method of tuning the choke tube diameter as a function of the compressor speed [18] to allow for effective pulsation control over the entire speed range, while maintaining minimal losses throughout the speed range. The virtual orifice (VO) and tunable side branch absorber (TSBA) are pulsation control devices developed under Gas Machinery Research Council (GMRC) funding [19]. The acoustic response of the cylinder nozzle is typically damped using orifice plates, which can create significant pressure losses. The VO was developed to virtually eliminate pressure loss while attenuating the cylinder nozzle response. A prototype and subsequent versions of the VO were developed and installed in an operating compressor resulting in significant pulsation attenuation. An example set of data from a field installation is shown in Fig. 16.11. The side branch absorber is a commonly used pulsation device to attenuate a single frequency (small frequency range) problematic resonance in the piping system. This concept was advanced through incorporating a variable volume
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60
Pulsation amplitude (psi pk-pk)
Max. amplitude = 52 psi 50
No orifice installed Orifice installed VO installed
40
Max. amplitude = 30 psi 30
20
Max. amplitude = 15 psi 10
0 0
20
40
60
80
100
120
140
160
180
200
Frequency (Hz)
FIG. 16.11 Virtual orifice prototype and test results.
FIG. 16.12 Tunable side branch absorber conceptual design. (Image from ACI Services Inc.)
(Fig. 16.12) to expand the frequency range of attenuation for the side branch absorber thus making it more effective over a range of operating speeds.
Advanced Reciprocating Compressor Valve Technology Valve performance, which includes both life and efficiency, must be considered when operating a reciprocating compressor. In a reciprocating compressor valve, desirable functional attributes include good sealing, rapid opening and closing, sustained high flow area (when open), minimum bouncing upon impact, toleration of impact forces and maximum temperatures, and low flow resistance. In most traditional valve designs, such as the ring, plate or poppet designs, a trade-off is made between life and efficiency.
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Gas flow in
Valve housing
Valve plate
Valve shaft
Gas flow out Control box Coil Magnets FIG. 16.13 Semi-active valve concept.
Advances in existing valve designs have been made through coatings and improved materials that reduce the impact wear, allow for corrosive and erosive gas compositions, and higher temperatures and pressures; additionally, various valve profiles have been developed to reduce performance losses [20]. Novel designs include a semi-active valve and the reed valve [21, 22]. Valve life is reduced when the valve passively slams shut or open due to the pressure differential across the valve; the semi-active valve employs magnets to slow down the closing of the valve, thus extending its life. The semi-active valve conceptual design is shown in Fig. 16.13 with the Fig. 16.14 showing the semi-active valve opening motion profile in red as compared to that of a typical plate valve shown in blue. The reed valve is based on the design of reeds commonly used in wind instruments. This valve consists of a reed contained in an easily replaceable module that is contained in the valve seat; since the design is a direct flowpath, there are low pressure losses associated with this type of valve (Fig. 16.15).
Surge Force Predictions Centrifugal compressor impellers and shafts are subject to fluctuating axial and radial forces when operating in surge. These forces can cause severe damage to the close clearance components of a centrifugal compressor such as the thrust and radial bearings, interstage and dry gas seals, and balance piston. Being able
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1400
Position [Volts]
1200 1000 800 Lower impact reduces plate stresses
600 400 200 0 120.00
140.00
160.00
180.00 Time [ms]
200.00
220.00
FIG. 16.14 Typical plate valve opening motion (blue) vs. semi-active valve opening motion (Red) [23].
FIG. 16.15 Reed valve concept [24].
to accurately quantify the frequency and amplitude of the cyclic surge forces on the close clearance components of the compressor allows the user to determine whether an accidental surge event, or emergency shut-down (ESD) transient, has caused damage requiring inspection, repair, or part replacement. In addition, the results can be used when designing a surge control system to determine if a hot gas bypass line is necessary or for choosing the anti-surge valve and control system parameters. In a compressor with no unbalance, misalignments, or rubs, surge- and axialsymmetric phenomenon-causes minor radial vibration and high-axial
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TABLE 16.1 Potential Damage Γ <0.001
Mild surge (low potential damage)
0.01<Γ < 0.001
Classic surge (discrete potential damage (requires further analysis)
Γ >0.01
Deep surge strong (high potential damage)
displacements due to cyclic thrust reversal. Thus, the strongest surge forces are those in the axial direction and the highest risk of failure component is the compressor’s axial thrust bearing. A method was developed, based on experimental test data performed for the GMRC [3], to quantify the cyclic axial surge forces on the close clearance components of the compressor allowing the user to determine whether an accidental (or ESD transient) surge event causes sufficient damage to require inspection or repair or as justification for omitting it. The methodology can be captured accurately in a transient one-dimensional (1D) lump sum parameter bond graph method dynamic analysis as demonstrated by Pinelli et al. [25]. In this work by Brun et al. [3], a surge severity coefficient was derived as shown in Eq. (16.1), based on energy conservation nondimensional analysis. This coefficient, Γ, takes into account the main parameters which affect the potential damage due to a surge event. Calculations above 0.001 suggest that an in-depth analysis be performed to determine the actual force amplitude in relation to the characteristics of the compressor to evaluate the risk of damage (see Table 16.1). Γ¼
r 2 VΔP2 smkΔh
(16.1)
where k—axial thrust-bearing stiffness m—mass of the rotor r—impeller tip radius s—axial thrust-bearing normal operational clearance V—compressor discharge flange to check valve volume Δ h—compressor enthalpy rise (head) Δ P—pressure difference across compressor (Pd Ps).
Liquid Packing Seals for Reciprocating Compressors The highest source of methane emissions in reciprocating compressors is leakage of process gas through the sealing components in the packing systems around the piston rods as shown in Fig. 16.16 [25]. Current technology uses a series of specifically-cut, dry-ring seals held in place with springs and cups (see Fig. 16.17). However, there is a trade-off made between leakage reduction with minimal gaps between the seals and the rod versus allowing sufficient gaps
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FIG. 16.16 Gas leakage path of reciprocating compressors. (Image courtesy of Dresser Rand.)
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FIG. 16.17 Cut away view of currently used reciprocating compressor rod packing sealing system [16]. (Image courtesy of EPA.)
such that the friction between the parts is sufficiently reduced to allow movement. Once the piston moves, the pressure differential across the packing seals creates a twisting effect on the seal, allowing substantial amounts of natural gas to leak into the casing. Ring twisting also causes increased friction and wear to the sealing rings and compressor rod. Using liquid, rather than gas, for essentially eliminating process gas leakage through the piston rods is not a novel concept, but does come with several challenges that have prevented its widespread implementation in production machines. The primary challenge is creating a system in which the pressure between the process gas and the sealing liquid is continuously balanced such that gas does not leak into the liquid and liquid does not leak into the gas. Reciprocating compressor pistons and valves create continuous high-frequency pressure changes that make instantaneous pressure balancing difficult. Other challenges include incorporating the support and monitoring systems for the liquid (pump, tank, filter, valves, flow-rate monitoring). The DOE has sponsored the development of a novel liquid packing seal prototype (2017–2019) and as of 2017, and one commercial company has manufactured and installed a small number of a different prototype of a liquid seal in currently operating compressors.
References [1] G.L. Agrawal, Foil air/gas bearing technology—an overview, in: Paper No. 97-GT-347, International Gas Turbine & Aeroengine Congress & Exhibition, June 2–5, Orlando, FL, 1997. [2] N. Baltadjiev, C. Lettieri, Z. Spakovszky, An investigation of real gas effects in supercritical CO2 centrifugal compressors, in: Proceedings of the ASME Turbo Expo 2014: Turbine Technical Conference and Exposition, Paper No. GT2014-26180, June 16–20, Dusseldorf, Germany, 2014. [3] J.L. Brady, J.M. Klein, M.D. Stevenson, S.P. Petullo, N.C. D’Orsi, R.C. Skinner, K.D. Eager, R.A. Perry, Downhole gas separation and injection powered by a downhole turbo expander, in: Proceedings of SPE Annual Technical Conference and Exhibition, SPE 49051, New Orleans, Louisiana, USA, September 27–30, 1998. [4] A. Rimpel, T. Allison, J. Moore, J. Grieco, P. Shy, J. Klein, J. Brady, Open-loop aerodynamic performance testing of a 105,000 RPM oil-free compressor expander for subsurface natural gas
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[5]
[6]
[7] [8]
[9]
[10]
[11] [12]
[13] [14] [15]
[16]
[17] [18] [19] [20] [21] [22]
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compression and reinjection, in: Proceedings of ASME Turbo Expo 2012, GT2012-69017, Copenhagen, Denmark, June 11–15, 2012. A. Rimpel, J. Moore, J. Grieco, P. Shy, J. Klein, J. Brady, Rotordynamics of a 105,000 RPM oil-free compressor-expander for subsurface natural gas compression and reinjection, in: Proceedings of ASME Turbo Expo 2012, GT2012-69119, Copenhagen, Denmark, June 11–15, 2012. L. San Andres, A. Anderson, Comparison of leakage between a labyrinth seal and an all-metal compliant gas seal at high temperature, in: Report to TRC-S-02-2011, Turbomachinery Research Consortium, Texas A&M University, May, 2011, 2011. L. San Andres, A. Anderson, Leakage for an all-metal compliant gas seal operating at high temperature, in: STLE Annual Meeting & Exhibition, Detroit, MI, May 5–9, 2013. L. San Andres, A. Anderson, An all-metal compliant-seal versus a labyrinth seal: a comparison of gas leakage at high temperatures. ASME J. Eng. Gas Turb. Power 137 (2015) 052504, https://doi.org/10.1115/1.4028665. ASME Paper GT2014-25572. D. Stiles, J. Justak, M. Kuzdzal, H. Miller, M. Sandberg, E. Wilcox, C. Rohrs, Application of dynamic pressure balanced seals in a multi-stage centrifugal compressor, in: 45th Turbomachinery & 32nd Pump Symposia, Houston, TX, September 12–15, 2016. C. DellaCorte, M. Valco, Load capacity estimation of foil air journal bearings for oil-free turbomachinery applications, in: STLE International Joint Tribology Conference, October 1–4, Seattle, WA, 2000. A. Delgado, Experimental identification of dynamic force coefficients for a 110 mm compliantly damped hybrid gas bearing, ASME J. Eng. Gas Turb. Power 137 (7) (2015) 072502. Delgado, A., and Ertas, B., 2018, “Dynamic characterization of a novel externally pressurized compliantly damped gas-lubricated bearing with hermetically sealed squeeze film damper modules,” Paper No. GT2018-7212, ASME Turbo Expo 2018, June 11-15, Oslo, Norway. Y. Wang, D. Kim, Experimental identification of force coefficients of large hybrid air foil bearings, ASME J. Eng. Gas Turb. Power 136 (3) (2014) 032503. Deffenbaugh, D., et al., Advanced Reciprocating Compression Technology, DOE Award No. DE-FC26-04NT42269, SwRI Contract No. 18.11052, 2005. S. Jung, R. Pelton, Numerically Derived Design Guidelines of Self Recirculation Casing Treatment for Industrial Centrifugal Compressors, in: Proceedings of the ASME Turbo Expo 2016: Turbomachinery Technical Conference and Exposition, Paper No. GT2016-56672, June 13–17, Seoul, South Korea, 2016. B. Ridens, S. Simons, S. Coogan, K. Brun, R. Kurz, GMRC/PRCI project: equation of state comparisons and evaluations for applications through gas property testing and derivations, in: Gas Machinery Conference, 2016. K. Brun; M.A. Wilcox; E.L. Broerman. U.S. Patent 8,534,058. Energy Storage and Production Systems, Apparatus and Methods of Use Thereof, Patented in United States of America, 2013. Deffenbaugh, D., McKee, R., Nored, M., Tunable Choke Tube for Pulsation Control Device Used with Gas Compressor. Publication number: 20090191076, 2009. E. Broerman, M. Nored, R. McKee, Advancements in pulsation control technology, in: Gas Machinery Conference, 2007. K. Brun, M. Nored, Valve performance and life of reciprocating compressors, in: Proceedings of the 41st Turbomachinery Symposium, September 24–27, Houston, Texas, 2012. K. Brun, J. Platt, R. Gernentz, M. Smolik, Semi-active compressor valve development and testing, in: Gas Machinery Conference, Oklahoma City, Oklahoma, October 2–4, 2006. World Oil Magazine, Modular Reed Valves Improve Efficiency for Gas Gathering and Distribution, 2013.
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[23] J. Moore, A. Lerche, H. Delgado, T. Allison, J. Pacheco, Development of advanced centrifugal compressors and pumps for carbon capture and sequestration applications, in: Proceedings of the Fortieth Turbomachinery Symposium, September 12–15, Houston, Texas, 2011. [24] J. Moore, N. Evans, J. Pacheco, T. Allison, J. Kerth, Development and testing of multi-stage internally cooled centrifugal compressor, in: 44th TurboMachinery and 31st Pump Symposia, Houston, Texas, 2015. [25] M. Pinelli, M. Morini, E. Munari, K. Brun, S. Simons, R. Kurz, Measurement and prediction of centrifugal compressor axial forces during surge: part 2-surge force measurements, in: Proceedings of ASME Turbo Expo 2017: Turbomachinery Technical Conference and Expositions GT2017-63061, June 26–30, Charlotte, North Carolina, 2017.
Further Reading [26] K. Brun, S. Simons, R. Kurz, M. Pinelli, M. Morini, E. Munari, Measurement and prediction of centrifugal compressor axial forces during surge: part 1-surge force measurements, in: Proceedings of ASME Turbo Expo 2017: Turbomachinery Technical Conference and Expositions GT2017-63061, June 26–30, Charlotte, North Carolina, 2017. [27] R. Subramanian, L.L. Williams, T.L. Vaughn, D. Zimmerle, J.R. Roscioli, S.C. Herndon, et al., Methane emissions from natural gas compressor stations in the transmission and storage sector: measurements and comparisons with the EPA greenhouse gas reporting program protocol, Environ. Sci. Technol. 49 (2015) 3252–3261. [28] US EPA, Reducing Methane Emissions from Compressor Rod Packing Systems, US Environmental Protection Agency, 2006. http://www3.epa.gov/gasstar/documents/ll_rodpack.pdf. [29] Zahroofvalves.com/pioneering-design.html. [30] D.J. Zimmerle, L.L. Williams, T.L. Vaughn, C. Quinn, R. Subramanian, G.P. Duggan, et al., Methane emissions from the natural gas transmission and storage system in the United States, Environ. Sci. Technol. 49 (2015) 9374–9383.