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March 2014

GE Buying Cameron’s recip line

Frigid Wintertests gas industry

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A lot of flack seems to be appearing in the mainstream media of late having to do with redistribution of wealth.

In my favor, you have to be wealthy in order to be subjected to this seem-ingly appealing exercise of justice — to lift up the poor among us. In the process, someone or some bu-reaucracy determines if you have too much or too little. Always, someone who likes to dispose of other peo-ple’s money makes these decisions. Fortunately, at my level of wealth, I won’t have to deal with this ideology. My belief is that it is our right to be wealthy if we so desire to be — to keep the fruits of our labor.

“The poor aren’t poor because Bill Gates and Warren Buffett are rich.” I’m not sure of the source of this quote, but, I’ll give credit to Kathleen Parker, an opinion writer for The Washington Post.

We adore the rich and famous, pure and simple, and in the end, despite our envy, we grant them their due. History does not reveal any success with the attempt at wealth distribution. On the contrary, the poor remain poor and the wealthy either are lowered to poverty or relocate.

But, how should we view being rich from a moral perspective? We have been warned that riches are danger-ous, dangerous to the soul and dan-gerous to society. So what should be our attitude towards having wealth, both as it pertains to the very rich and as it pertains to us?

First, we must never idealize pov-erty and see wealth as a bad thing in itself. Then, we must avoid politi-cizing both poverty and wealth. Our lens must always be moral rather than

political, though wealth and poverty have huge political implications. The position of wealth is not a bad thing of itself; it is how we use it and what it can do to us as authentic human per-sons that can be bad. There is a huge cavern that separates generosity from miserliness.

Generally speaking, it is the rich that provide the engine for economic growth through their investments and risk taking. It is through their efforts that society in general progresses and is the best hope for lifting up the bot-tom rung of society. The poor will al-ways be with us and they should be the target of our generosity regardless of our level of wealth.

Bill Gates Sr. puts it, “Society has an enormous claim upon the fortunes of the wealthy. This is rooted not only in most religious traditions, but also in an honest accounting of society’s substan-tial investment in creating fertile ground for wealth-creation.”

Government is limited, in my opin-ion, in solving the problem of poverty. It is very inefficient, it doesn’t manage its vast economic power very well and try and try as it does from time to time, it always fails in its effort to eliminate poverty.

On the contrary, depending upon the power of the decision-makers, the wealth-distribution ideology often morphs into despotism if the perpetra-tors aren’t thrown out of office. History reveals some prime examples of where to gain compliance, a major portion of the population had to be eliminated.

May the Lord hold you in the hol-low of His hand throughout the rest of 2014. CT2

What’s Yours Is Yours,What’s Mine Is Yours

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evolving solutions around your world

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www.c-p-i.com

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Featured Articles 16 GE Acquiring Cameron’s Recip Division

18 Murphy, EControls Assimilation Pays Dividends

22 The Importance Of Motor Dynamics In Reciprocating Compressor Drives

26 Motortech Adds High-End Ignition Controllers

28 Dresser-Rand Makes Major Move Into Gas Liquefaction

32 Custom Packager Strikes Gold In The Marcellus Shale

42 Flowmeter Can Be Sized For Any Engine

62 Recip Compressor Performance, Safety Predictions For Control Panels

TECHcorner 36 Motor Dynamic Influence On Torsional Vibration Analysis

46 Friction-Surface Coatings In Dry-Running Recips

58 Solving Compressor Impeller Rub Problems During Mechanical Run Tests

76 Case Study: Intake/Exhaust Silencer Redesign Mitigates Noise

Departments 4 Page 4 — What’s Yours Is Yours, What’s Mine Is Yours

8 Global Perspective — EOR Could Revitalize Middle Eastern Oil Industry

10 Meetings & Events

12 About The Business — Frigid Winter Challenges U.S. Gas Industry

14 Monitoring Government — U.S. LNG Exports Remain On The Back Burner

56 Featured Products

73 Recent Orders

73 Web Headlines

74 Prime Movers

84 Scheduled Downtime

85 Marketplace

86 Advertisers’ Index

88 Cornerstones Of Compression — The Heart Of The High-Speed Recip

COMPRESSORDedicated To Gas Compression Products & Applications

March 2014

Follow Us @ Compressortech2

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Compressortech 2 ( ISSN 1085-2468) Volume 19, No. 2 — Published 10 issues/year (January-February, March, April, May, June, July, August-September, October, November, December) by Diesel & Gas Turbine Publications, 20855 Watertown Road, Waukesha, WI 53186-1873, U.S.A. Subscription rates are $85.00 per year/$10.00 per copy worldwide. Periodicals post-age paid at Waukesha, WI 53186 and at addi-tional mailing offices. Copyright © 2014 Diesel & Gas Turbine Publications. All Rights Reserved. Materials protected by U.S. and international copy-right laws and treaties. Unauthorized duplication and publication is expressly prohibited. Canadian Publication Mail Agreement # 40035419. Return Undeliverable Canadian Addresses to: P.O. Box 456, Niagara Falls, ON L2E 6V2, Canada. E-mail: slizdas@dieselpub.com. POSTMASTER: Send address changes to: Circulation Man ager, COMPRESSORtech2, 20855 Watertown Road, Suite 220, Waukesha, WI 53186-1873 U.S.A.

www.compressortech2.com

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MARCH 2014 8 CoMpRessoRtech2

Advanced enhanced oil recovery (EOR) has not been widely applied in the Middle East due to con-tinuing strong levels of production and the addition-

al costs of installing of high-pressure compressors, pumps and other equipment.

EOR injects associated gas, water, carbon dioxide, steam, etc., underground in order to maintain pressure in the formation and/or fluidize the heavy hydrocarbons to en-able their flow toward the production wells. Not only can this increase the rate of production, but it also can extend the commercial life of the oilfield.

In some oil-rich nations, the high price of crude and steep depletion rates at giant, mature oilfields have pushed na-tional oil companies to reconsider their investment plans. Oman has somewhat taken a lead in EOR in the region and the United Arab Emirates has implemented a project.

Qatar is moving forward aggressively. In mid-January, Maersk Oil opened a digital core laboratory in Doha. The facility will put particular focus on researching EOR in car-bonate formations and the marine environment. This should develop better technologies for reservoirs like Maersk’s Al Shaheen oilfield on offshore Block 5, which is one of the most complex carbonate structures in the world.

The Middle East is the world’s largest carbonate oil-pro-ducing region. Historically, EOR technologies have been applied to sandstone formations, where they have largely been successful due to the relatively standard porosity throughout the geology. However, it has been challenging to apply the same techniques to carbonate reservoirs as they are generally more naturally fractured than sandstone.

Fractures in a formation act as conduits to injected water, polymers or steam, often bypassing the target oil zone and reducing the efficiency of the EOR sweep. This research into subsurface chemistry and fluid flow relationships in carbonate structures will be essential to sustaining Qatar’s longer term oil output.

Qatar’s hydrocarbon industry is heavily centered on the re-sources in the largest single deposit of gas in the world, the North Field. However, since 2005, the government has main-tained a moratorium on new developments pending a study on field optimization and the geological impact of gas extrac-tion. The final sanctioned project, the Barzan Gas Project, is due to come on stream in two phases this year and in 2015.

As a result of reduced oil and gas activity, Business Moni-tor International (BMI) expects Qatar’s hydrocarbon sector to stagnate and its contribution to the nation’s economy to continue to decline.

Despite the slowing investment in Qatar’s upstream, the country still has plenty to offer. French major Total remains upbeat about prospects offshore Qatar and this year will conduct exploratory drilling with partner China National Off-shore Oil Corp. on Block BC.

An exploration and production sharing agreement signed with the Qatari government requires at least three explora-tion wells to be drilled this year. The first is due to spud soon, with results expected to be available before the end of the year. Total’s efforts could therefore bring welcome upside po-tential to the country’s reserves and production outlook.

Perhaps of greater importance is Qatar’s drive to move its oil and gas sector to the forefront of technology. In particu-lar, as Qatar’s oilfields begin maturing, the nation is placing substantial focus on maximizing production from existing fields and exploring EOR opportunities. Both tactics could make a major contribution to increasing reserves and sus-taining production over a longer period.

Redevelopment efforts are already underway. Qatar Pe-troleum is budgeting US$13 billion to maximize production from Bul Hanine oilfield. While the project is only expected to add 50,000 bbl/d of output, it will extend production life by 25 years. Increased output is due in 2018.

Meanwhile, Occidental Petroleum is working on a six-year project to redevelop the Idd El Shargi field after signing an agreement with Qatar Petroleum in mid-2013. The US$3 billion project will add 200 production and injection wells. While the development is not expected to increase produc-tion, it should sustain oil production at around 100,000 bbl/d and extend field life.

Total is also involved in Qatar’s production optimization drive. The company secured a 25-year extension to its li-cense at Al Khalij field effective in January 2014. No rede-velopment plans have been announced, but the terms of the new agreement require Total to take additional steps to maximize oil output.

Because of these efforts, BMI expects Qatari oil produc-tion (including NGLs and other liquids) will be sustained above 1.7 million bbl/d over the next 10 years. CT2

EOR Could Revitalize Middle Eastern Oil Industry >By ROBERTO CHELLINI

ASSOCIATE PUBLISHER

Global Perspective

Qatar Pushes Ambitious Program

CT338.indd 1 2/21/14 2:27 PM

rolls-royce.com

Trusted to deliver excellence

It’s all in the name.

A proud past leads to a new future. There may be a proud old name on the outside, but the driving force within the world’s best-engineered, most efficient, pipeline gas compressors is Rolls-Royce. The heritage name, Cooper-Bessemer, still carried by older machines, echoes the engineering excellence that has earned Rolls-Royce an unparalleled reputation for quality. Today, in a business where productivity and dependability mean so much, the unsurpassed engineering experience of the past makes Rolls-Royce the compressor name of the future.

Cooper-Bessemer is a registered trade name of Cameron Corporation, used under license by Rolls-Royce plc

RollsRoyce.indd 1 10/11/13 1:55 PM

Meetings & Events*Indicates shows and conferences in which Compressortech2 is participating

For a complete listing of upcoming events, please visit our website at www.compressortech2.com

MARCHMarch 19-21*China International Offshore Oil &

Gas Exhibition — Beijing

Tel: +86 10 5823 6555

Web: www.ciooe.com.cn/2014/en

March 23-27 Sour Oil & Gas Advanced Technology 2014 — Abu Dhabi, United Arab Emirates

Tel: +971 2 674 4040

Web: www.sogat.org

March 24-25

*Gas Transport & Storage 2014 —

Berlin

Tel: +44 20 7202 7690

Web: www.gtsevent.com

March 24-27

*Gastech — Seoul, Korea

Tel: +44 203 615 2872

Web: www.gastechkorea.com

March 26-27

Georgian International Oil, Gas, Energy

and Infrastructure Conference —

Tbilisi, Georgia

Tel: +44 207 596 5135

Web: www.giogie.com

ApRIlApril 7-9

*Gas Compressor Association Expo

& Conference — Galveston, Texas

Tel: +1 (972) 518-0019

Web: www.gascompressor.org

April 9-10Turkish International Oil and Gas

Conference 2014 — Ankara, Turkey

Tel: +44 207 596 5147

Web: www.turoge.com

April 13-16

*Gas processors Association

Annual Convention — Dallas

Tel: +1 (918) 493-3872

Web: www.gpaglobal.org

April 15-16

*Gas Compressor Institute —

Liberal, Kansas

Tel: +1 (620) 417-1171

Web: www.gascompressor.info

April 28-30

Black Sea Oil and Gas Forum

— Bucharest, Romania

Tel: +44 203 615 2988

Web: www.blackseaoilgas.com

April 28-May 2

*Gulf South Rotating

Machinery Symposium —

Baton Rouge, Louisiana

Meetings & Events

March 2014 10 coMpressortech2

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For a complete listing of upcoming events, please visit our website at www.compressortech2.com

Tel: +1 (225) 578-4853Web: www.gsrms.org

MAYMay 5-8*Offshore Technology Conference — HoustonTel: +1 (972) 952-9494Web: www.otcnet.org

May 13-15*Eastern Gas Compression Roundtable — Moon Township, PennsylvaniaTel: +1 (412) 372-4301Web: www.egcr.org

May 13-15Uzbekistan International Oil & Gas (OGU) Exhibition — Tashkent, UzbekistanTel: +44 207 596 5144Web: www.oguzbekistan.com

JUNEJune 3-5

*Sensor+Test 2014 —

Nuremberg, Germany

Tel: +49 5033 9639 0

Web: www.sensor-test.de

June 3-5

*Power-Gen Europe 2014 —

Cologne, Germany

Tel: +44 1992 656 617

Web: www.powergeneurope.com

June 3-6

Caspian Oil & Gas 2014 —

Baku, Azerbaijan

Tel: +44 207 596 5000

Web: www.caspianoil-gas.com

June 10-12

*Global Petroleum Show —

Calgary, Alberta, Canada

*Indicates shows and conferences in which Compressortech2 is participating

Tel: +1 (403) 209-3555

Web: www.globalpetroleumshow.com

June 16-20

*ASME Turbo Expo 2014 —

Dusseldorf, Germany

Tel: +1 (404) 847-0072

Web: www.asmeconferences.org/

TE2014

June 24-26

Sensors Expo & Conference —

Rosemont, Illinois

Tel: +1 (617) 219-8375

Web: www.sensorsmag.com/

sensors-expo

June 25-26

Energy Exposition — Billings, Montana

Tel: +1 (307) 234-1868

Web: www.energyexposition.com

March 2014 11 coMpressortech2

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March 2014 12 coMpressortech2

An exceptionally cold winter in much of the U.S. has driven increased demand for natural gas.

When a polar vortex struck the upper Midwest and Northeast in January, some areas experienced low gas pres-sures and near brownout conditions as pipelines strained to deliver enough gas to heat homes and produce electricity.

In December, Williams Partners had reported record nat-ural gas flows on its Transco interstate gas pipeline system, driven by demand for home heating and power generation in the eastern U.S.

Energy Information Administration (EIA) data shows that gas storage levels have been below the recent five-year aver-age since the beginning of January. Since October, net with-drawals from storage in the eastern U.S. slightly exceeded the previous record noted during the 2002-03 heating sea-son. Henry Hub spot prices reached the highest level in four years, hovering in the US$4.50 to US$5.50/MMBtu range since January and spiking to US$7.90 on Feb. 5.

Despite the surging winter demand, IHS Cambridge En-ergy Research Associates predicted in January that shale production would restrain natural gas prices to the US$4 to US$5/MMBtu range for at least 20 years. Even at the low end of the price range, the report said that about 900 Tcf (2.55 x 1013 m3) of unconventional gas could be economically produced, inferring that significant demand increases can be accommodated without requiring a substantially higher price.

Ironically, much of the recent high demand and the result-ing strain on the pipeline system have been in the eastern U.S., the region that has developed into the most prolific producer of gas in the nation.

The EIA reported production from the Marcellus region reached 12 Bcfd (3.4 x 108 m3/d) earlier this year, more than six times the 2009 production rate. If the Marcellus Shale region were a nation, its natural gas output would rank third in the world, after Russia and the rest of the U.S.

With supplies growing in new shale gas regions, inadequate oil and gas infrastructure continues to cause problems. Those

infrastructure challenges were discussed during the Gas/ Electric Partnership conference in Cypress, Texas, on Feb. 5-6.

A presentation by McKinsey & Co. showed 65 gas pipe-line projects under construction, approved, filed or pro-posed in the U.S. Those included a mix of expansions and laterals, principally in the Appalachian region or routes to New England and the Midwest from the Marcellus area (See COMPRESSORtech2, January-February 2014, p. 40).

Meanwhile, the long-established, long-haul pipelines from the Gulf Coast to the Northeast and Canada are underuti-lized. Some of those may be reversed or converted to liquid or ethane lines.

McKinsey projects that Appalachian shale gas and associ-ated gas elsewhere in the nation will drive U.S. gas produc-tion through this decade. Large quantities of Marcellus gas will flow out of the Northeast while associated gas from the Bakken Shale in the Williston Basin will increasingly displace Canadian gas that historically has supplied the Midwest market. Volumes from both will potentially flood Chicago and the Gulf South.

Appalachian production from the Marcellus and Utica will continue to displace gas coming from the southwest.

A presentation by the Wood Mackenzie consultancy showed that U.S. upstream capital investment, led by oil development, is almost 25% of the global total. It said that by 2020, output of “tight oil” will grow significantly, with the largest supplies coming from the Eagle Ford Shale, Bakken Shale and Permian Basin.

This oil production also will boost the gas flow since those shales produce large volumes of associated gas. The pipe-line infrastructure to take away all the Eagle Ford and Bak-ken associated gas currently is insufficient and flaring is prevalent. McKinsey predicts that the Permian gas flow will also be bottlenecked by next year.

Speakers said conversions and repurposing of gas pipe-lines are proceeding very quickly; new laterals are taking longer. More underground gas storage or local liquefied nat-ural gas storage facilities also may have to be considered to meet power generation demand peaks.

This burst of production and pipeline growth is driving the need for new compressors as well as the refurbishment and resizing of existing pipeline stations in these regions, all of which is good news for the gas compression industry. CT2

Frigid Winter Challenges U.S. Gas Industry >By NORM SHADE

About The Business

Production growth, new infrastructure driving compression demand

By NORM SHADE

Norm Shade is senior consultant and president emeritus of ACI Services Inc. of Cambridge, Ohio. A 44-year veteran of the gas compression industry, he has written numerous papers and is active in the major industry associations.

CT342.indd 1 2/21/14 2:30 PM

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Monitoring governMent

Energy Department’s permitting continues at leisurely pace

By patrick crow

U.S. LNG Exports Remain On The Back Burner >

MARCH 2014 14 CoMpRessoRtech2

twelve months ago, this column discussed the gla-cial pace in approvals for liquefied natural gas (LNG) export permits at the U.S. Department of Energy

(DoE). Little has changed since then, and that of course continues to be the problem.

the shale gas boom has bestowed ample natural gas and gas liquids supplies upon the nation at bargain pric-es, much to the delight of the chemical industry and other hydrocarbon-intensive manufacturers (who predictably op-pose LNG exports).

with even greater shale gas production in prospect, en-trepreneurs have proposed a number of projects to export domestic LNG into a thirsty world market. Given the long lead times required to build liquefaction plants, sponsors are clamoring for DoE approvals now.

So far, the department has cleared six projects in the key category, exports to non-Free trade agreement nations, and about 25 are pending. in the past 12 months, DoE has approved only five permits.

in reality, all 30 export projects won’t go forward. perhaps the market will be able to accommodate a half dozen; may-be a few more. proponents argue that’s not the point: DoE should approve the permits and let the market determine which projects bloom in the right place at the right time.

although associations representing gas producers have been pressing for LNG exports, an ad hoc coalition was launched in February with the same goal objective.

the “our Energy Moment” alliance is a convergence of producer associations, businesses and state economic de-velopment organizations. they are mostly based in texas and Louisiana, where much of the surplus gas productivity is pooling.

the group said LNG exports could generate up to US$47 billion in net benefits to the U.S. economy and create up to 450,000 jobs.

it said exports would help reduce the U.S. trade deficit by as much as US$27 billion. it argued exports would have little impact on domestic prices, due to the mushrooming supplies now recoverable at low production costs.

in the House of representatives, the Energy and com-merce committee recently issued a report urging the obama administration to approve all of the pending LNG export applications by the end of the year.

the report said, “our friends and allies around the globe desperately need a more stable, reliable, and affordable supply of natural gas, and american consumers and manu-facturers need continued robust demand to bring additional resources into competitive production.

“this window of opportunity will not remain open indefi-nitely, therefore the committee is urging the Department of Energy to approve all remaining export licenses by the end of the year and is considering legislative action to modern-ize the process and remove barriers.”

committee chairman Fred Upton (r-Mich.) said, “in es-sence we’re sending a friendly shot across the bow for DoE to take it up.”

two key senators seem to be firing shots aimed at the bow. they are Mary Landrieu (D-La.), the new chairman of the Senate Energy and Natural resources committee, and Lisa Murkowski (r-alaska), the senior republican on the panel.

Landrieu recently ascended to the chair when ron wyden (D-ore.), whose cautious resistance toward gas ex-ports constituted passive opposition, switched to the chair of the Finance committee.

the Landrieu-Murkowski duo is a washington energy lobbyist’s fantasy team. Both represent major producing states and hold key positions. over the years, the Senate panel has been the most influential entity in the national energy policy realm.

Landrieu supports the “our Energy Moment” gas-export campaign. Murkowski recently issued a white paper urging the administration to lift the ban on crude oil exports and expedite LNG permitting.

So from outward appearances, it might seem that mo-mentum is building in congress to force faster action at DoE. well, not really.

the recovery of the U.S. economy is partially due to surge in cheap shale gas production. the obama adminis-tration recognizes that fact and has every incentive to keep natural gas supplies plentiful and cheap, especially with the congressional midterm elections approaching in November.

No matter what House republicans bluster, nor what Landrieu and Murkowski can push through their energy committee, Senate Democratic leaders can be expected to keep the lid on legislation to expedite LNG exports. Ct2

ct340.indd 1 2/21/14 2:29 pM

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Motortech.indd 1 2/11/14 11:13 AM

Cameron has agreed to sell its reciprocating compression division to GE Oil & Gas for

US$550 million and, at press time, was considering the sale of its cen-trifugal compression business.

The company’s recip business pro- vides reciprocating compression equip- ment and aftermarket parts and serv-ices for oil and gas production, gas processing, gas distribution and inde-pendent power industries.

The division, which generated sales of about US$355 million in 2012, has 900 employees and operates from 20

global locations. If U.S. regulatory agen-cies approve, the acquisition is expect-ed to close during the third quarter.

GE said the acquisition comple-ments GE Oil & Gas’ existing high-speed reciprocating (HSR) business, which focuses on low-horsepower units that are used predominately in gas lift applications. Cameron’s recip portfolio will enable the company to offer higher-horsepower models used in gas gath-ering, processing and transmission.

GE noted that high-speed recips are used in applications from gas gather-ing, gas lift and injection, as well as

transmission and storage. The devel-opment of shale oil and gas fields, particularly in North America, has increased demand for high-speed re-ciprocating compressors. It said that as shale continues to develop in other regions of the world, such as Asia and South America, the acquisition will po-sition GE to serve the industry globally.

Lorenzo Simonelli, president and CEO of GE Oil & Gas, said, “The new (Cameron recip) business positions GE to more effectively focus on key downstream industry trends and to better anticipate customer needs.”

GE Acquiring Cameron’s Recip Division > Cameron may sell centrifugal business, too

BY PATRICK CROW

march 2014 16 compressortech2

n Cameron employees assemble Superior compressors at the company’s Houston plant.

CT337.indd 1 2/21/14 3:49 PM

march 2014 17 compressortech2

per year in 1994 to nearly US$20 bil-lion today and profits growing at an av-erage 16% over the past three years.

Hasan Dandashly, DTS vice presi-dent, said, “Cameron’s reciprocating compression division will become an important part of our downstream of-fering to customers. Its services and geographic presence will expand our distributed gas portfolio and enhance our shale capability and services ex-pertise for our customers.”

In addition to Cameron’s reciprocat-ing compression division, DTS also will include the newly acquired Salof prod-uct lines of small-scale, modular lique-fied natural gas plants for the rail, ma-rine, trucking and industrial industries.

In late January, Cameron reported 2013 net income of US$724.2 million, compared to US$750.5 million in 2012. Revenues were a record US$9.8 bil-lion, up 16% from US$8.5 billion in the prior year. CT2

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Cameron said its after-tax proceeds from the sale would be US$400 mil-lion. The recip business had sales of US$355 million for the year ended Dec. 31, 2012.

Citi is helping Cameron evaluate options for its centrifugal compression business, which had sales of US$365 million in 2012.

Jack Moore, Cameron chairman and CEO, said, “These actions will streamline the company’s operations and are consistent with our strategy of building on our strong sales and order momentum in our core markets while selectively expanding product and service offerings in strategic growth areas. The proceeds from the trans-action will provide us with greater fi-nancial flexibility and afford us the op-portunity to drive additional value for our shareholders.

“Exploring strategic alternatives for the centrifugal compression business is part of our ongoing effort to opti-mize our asset base with a focus on our core markets. We are committed to directing resources to businesses where we have the best opportunities to achieve sales growth, higher mar-gins and market leadership.”

Richard Stegall, president of Cam-eron’s reciprocating compression divi-sion, said, “We have achieved good growth over recent years, and we an-ticipate that this trajectory will contin-ue with the support of GE Oil & Gas.”

Cameron’s recip operations will be-come part of GE Oil & Gas’ recently formed Downstream Technology So-lutions (DTS) business, which will supply equipment and services to the US$10 to US$11 billion downstream and distributed gas industry segments — including unconventional oil and gas activities.

Headquartered in Houston, DTS offers technologies and services in-cluding steam turbines, blowers, dis-tributed gas solutions, reciprocating compressors, pumps, valves and dis-tribution systems. The business also provides maintenance services and remote monitoring and diagnostics.

GE said oil and gas is one of its fastest-growing businesses, with or-ders rising from less than US$1 billion

CT337.indd 2 2/24/14 4:16 PM

The business formed by the 2009 merger of FW Murphy, Tulsa, Oklahoma, and EControls, San

Antonio, Texas, has taken the name of Enovation Controls to reflect the now fully integrated company.

Enovation Controls is positioning it-self as a global provider of engine and engine-driven equipment management and control products and services. The company serves markets that include natural gas compression and liquids, natural gas commercial ve-hicles, off-highway, material handling, recreational and commercial marine, power generation and agriculture.

The privately held company employs 1000 people worldwide and was on track to exceed US$250 million in rev-enue in 2013. Board member Patrick W. Cavanagh is president and CEO.

“Enovation Controls combines Mur-phy’s line of controls and instrumen-tation with EControls’ engine control expertise to go beyond components to a fully integrated engine control and in-strumentation provider,” said Frank Mur-phy III, Enovation Controls executive chairman of the board. Murphy added that Enovation Controls would continue to promote and utilize the Murphy and

EControls brands while strengthening the company’s positioning in the market.

“Although we officially merged four years ago, we wanted to be very careful and deliberate with our integration pro-cess in order to make sure we truly pre-served the strengths both companies brought to the table,” Murphy said. “The Enovation Controls name now repre-sents our fully integrated company as a single, focused team.

“By capitalizing on our synergies, Enovation Controls delivers the same products and services our customers have always known, along with the more fully integrated solutions our markets demand.”

In order to facilitate the company’s approach to engine control, protection and monitoring, Enovation Controls is realigning its technical resources into Solution Groups to work more closely with the company’s Natural Gas Pro-duction Controls, Display and Power Controls, and Engine Controls and Fuel Systems business segments.

“The goal in creating integrated teams of the company’s product de-sign and development engineers is for Enovation Controls to develop and deliver solutions to customers faster

than its competitors,” said Kennon Guglielmo, chief technology officer.

“The Solution Groups are focused on the development of more integrat-ed and tailored solutions for custom-ers, to deliver improvements in engine management, fuel efficiency, reliabil-ity, drivability and emissions.”

A recent example of Enovation Con-trols’ approach to engine governance is its Engine Integrated Control System (EICS) to optimize emissions compli-ance and performance for natural gas compressors. EICS is a full-authority engine control system that incorpo-rates ignition, air/fuel ratio and speed control along with diagnostics, sen-sors and catalyst into one package de-signed for specific engine models.

The system includes the company’s engine control modules, PowerView displays and related proprietary soft-ware that displays critical engine in-formation. The system is precalibrated to meet emission and performance requirements for the application and typically requires no field calibration or adjustment.

With the introduction of EICS-equipped engines, customers have

Murphy, EControls Assimilation Pays Dividends > One example: meshed emissions, performance

systems for gas compressors

march 2014 18 compressortech2

n Enovation Controls’ Engine Integrated Control System helps companies meet emissions standards as well as optimize performance for their natural gas en-gines driving compressors. The system incorporates everything needed in one high-tech package designed specifically for selected engine models.

continued on page 20

CT333.indd 1 2/21/14 3:26 PM

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experienced short installation times with improved engine performance and fuel economy in the presence of load swings and changes in natural gas well flow, said Terry Baldwin, vice president of the Natural Gas Produc-tion Systems segment. Baldwin said EICS’ ability to monitor all integrated system components could simplify troubleshooting and reduce engine downtime while enabling customers to meet emissions regulations.

“Our customers are reporting a 20 to 30% savings in fuel and operating costs by utilizing the EICS system,” Baldwin said, adding that over 1500 natural gas compressors have been retrofitted with the EICS.

In addition to EICS, the merger has resulted in other synergies. EControls’ electronics manufacturing capabilities, which include in-house, surface-mount assembly lines, electronic quality-inspection equipment and direct re-lationships with electronic parts sup-pliers, are now used to improve the quality, cost and lead time of Murphy- branded products.

Enovation Controls has used this combination of capabilities to launch a generation of PowerView color and monochrome displays with new capa-bilities and price points.

Another benefit of the merger, Mur-phy said, is that Enovation Controls has taken engine control module de-signs from EControls and transformed them into general-purpose controllers and I/O modules for off-highway and marine applications. These new mod-ules have been combined with the next

generation PowerView displays to form a package that is designed to work as a single system. The customer also can program these components as a single integrated system by using the proprietary PowerVision Configuration Studio development software.

The combination of intellectual prop-erty from the two companies is also yielding value for Enovation Controls customers, Guglielmo said. “The inte-gration of the engine control, vehicle systems and dashboard displays can yield performance advantages that a collection of components from a vari-ety of vendors cannot match,” he said. “The ability to cross over the boundary between vehicle systems and engine control is a unique Enovation Controls’ capability; a capability made possible by the merger.”

Some of Enovation Controls’ most promising growth opportunities are

in providing control and fuel system solutions for natural gas powered ve-hicles, Guglielmo said.

Of the more than 1 million engine control systems Enovation Controls has shipped to date, more than 170,000 are found on heavy-duty natural gas fu-eled buses and trucks. Guglielmo said Enovation Controls expects the num-ber of natural gas-powered vehicles to grow significantly in the U.S., China and worldwide.

To that end, Enovation Controls has invested in the relocation and significant upgrades of its engine development fa-cilities in San Antonio. It is spending an-other US$4 million (in addition to more than US$15 million already invested in China) in an engine development cen-ter in Hangzhou, China, that is sched-uled to open at midyear.

With these investments, Enovation Controls will be positioned to support the expected growth in the natural gas vehicle sector. The company has sales and manufacturing facilities in North America, China, Europe and In-dia, and said it will continue to expand and invest its international operations.

“At Enovation Controls, we’re lever-aging our synergies in order to expand our offering so that more customers can experience the benefits of a total engine control and instrumentation solution,” Cavanagh said. “However, we remain committed to providing our customers with the quality and service that earned us the business in the first place.” CT2

march 2014 20 compressortech2

n The EControls manufacturing and assembly operations in San Antonio, Texas, have been expanded and upgraded.

n China is a major target for Enovation Controls’ global marketing effort. It has an application engineering and sales operation in Shanghai and will expand its plant at Hangzhou, shown above, at midyear.

CT333.indd 2 2/21/14 3:26 PM

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Investments in further developing the North American natural gas supply continued at a strong pace

in 2013. Major investments are being made to process, transmit and store the new gas supplies.

The need for highly accurate torsion-al analyses to produce the most reli-able drivetrain is paramount to achiev-ing best-in-class uptime.

The use of electric motor-driven

compression continued to increase 2013, a trend which is primarily being driven by availability of electricity, cost of electricity, emissions regulations and the strategies of end users.

When utilizing motor-driven drive-trains, proper torsional dynamics and system analysis should lead to:

• improved uptime due to much lower maintenance time,

• improved uptime due to decreas­

ed torsional issues — coupling alignment, resonance and vibra-tion alarms,

• improved uptime due to increased equipment reliability.

The following technical summary is based on a presentation that Gerhard Knop, head of development projects for Neuman & Esser (NEA), made at the 2012 International Rotating Equipment Conference in Düsseldorf, Germany. The paper reviewed the analysis and decisions that NEA takes to ensure that torsional designs for motor driven appli-cations minimize unforeseen downtime.

The impact of using a motorBecause of their enormous torque

variations, reciprocating compressors introduce torsional vibrations into the drivetrain that must be thoroughly

The Importance Of Motor Dynamics In Reciprocating Compressor Drives > Exacting torsional analyses can ensure reliable drivetrains

By SWAMy SuBRAMANyAM AND GERhARD KNop

Swamy Subramanyam is vice president of technical management and procurement at Neu-man & Esser USA Inc. He holds a master’s degree in mechanical engineering and has more than 20 years of experience in compressor application engineering and design of reciprocating gas compressor packages and associated equipment. He is a mem-ber of the GMRC Torsional Analysis subcommittee. Contact him at: ssubramanyam@ neuman-esser.com. Gerhard Knop is head of development projects at the Neuman & Esser Machine Factory in Übach-Palenberg, Germany. He has more than 18 years of experi-ence in numerical and analytical calculations, with a focus on drivetrain vibrations. He is member of the GMRC Torsional Analysis sub-committee and has a Dipl.-Ing.(FH) degree in mechanical engineering. Contact him at: gerhard.knop@neuman-esser.de.

march 2014 22 compressortech2

n NEA’s products, including this 10,500 hp (7.8 MW) compressor train for natural gas storage, all undergo detailed torsional analy-ses, including motor dynamics.

continued on page 24

022.indd 1 2/21/14 4:04 pM

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Prognost.indd 1 6/7/13 11:42 AM

investigated by simulation before these machines are put into opera-tion. In recent years, but interestingly, sometimes even today, the reacting torque of the driving induction or syn-chronous motor was considered to be constant at steady-state operation.

This assumption most often leads to completely incorrect shaft torque dynamics and gives no information on electric current fluctuations. Both quantities, however, are the key de-sign features of a drivetrain (API 618, paragraphs 7.1.1.7 and 7.1.2.6).

Torque dynamics must be limited in order to:

• avoid coupling overloading or early rubber wear,

• avoid shaft fractures,• avoid rotor and stator mechanical

overloading.Electric current fluctuations must be

limited to:• avoid excessive loading and flicker,• avoid thermal overloading of the

frequency converter,• semiconductor malfunction.The motor reactions on torsional vi-

brations can create highly amplified or attenuated shaft torque variations. This diverse behavior was modeled with a focus on induction motors, which are much more frequently used than syn-chronous motors.

Why to consider motor dynamics in drivetrain simulations

The introduction of motor dynamics into the drivetrain simulation is increas-ingly important and requires precise calculations and sophisticated simula-tion models for the following reasons:

• Higher compressor rotational speeds are being used, which in-creases the probability of torsional resonances.

• The use of speed controls contin-ues to increase, which results in resonant conditions that need to be properly investigated.

• Improved drivetrain simulations allow reduction of flywheel size and weight, which reduces com-pressor main bearing load, crank shaft bending load and lateral vi-bration tendency.

• More complete and sophisticated drivetrain simulations lead to in-creased reliability. It should be noted that the number of coupling or shafting failures due to com-plete simulation models (with mo-tor dynamics) has gone down to practically zero.

Motors react on torsional vibra-tions. If the torque of the compres-sor were constant, the motor air gap torque would be constant as well. However, the fluctuating compressor torque produces angular deflections of the rotor within its magnetic field and causes the motor air gap torque to change its height.

The resulting motor air gap torque variations represent an external exci-tation, similar to compressor torque, and influences the torsional vibrations of the drivetrain.

The coupling type has an immense effect on the motor air gap torque which reduces or increases torsional vibra-tions and shafting torque load depend-ing on the given drivetrain configuration.

Soft coupling and motor — two friends

The most amazing influence on the torsional vibration behavior of the motor dynamics can be observed at drivetrains being equipped with high flexible (rubber in shear type) cou-plings. When properly chosen, they would even attenuate the shaft torque loading to an almost constant value.

Negative dampingWhen using all-steel disc couplings,

which are quite rigid, advanced ana-lytical design approaches reveal that coupling and motor shaft torque varia-tions go up significantly without much external excitation.

Whether during start-up or steady-state, these self-induced vibrations are driven by feedback of the motor.

Where positive damping results in attenuation of vibrations in resonance near condition, negative damping would amplify the vibrations due to unfortunate phase shift between exci-tation and vibration.

Variable-frequency drivesIf a frequency converter is used, the

motor air gap torque is affected by the respective frequency converter con-trol methods. The options are:

• Control of motor current fluctua-tions. In this case, the motor air gap torque is smoothened in the same way as the motor current. If the control is fast enough, the be-havior approaches that of a con-stant motor air gap torque.

• Control of the rotational speed fluc-tuations. In order to reduce speed variations, the motor air gap torque must vary accordingly. Therefore, smooth speed is achieved by vary-ing current and shaft torque.

• Setting the control very slow so that no fluctuations are controlled down. In this case the motor be-haves in the same way as if con-nected directly to the main power network, only with different fre-quency and voltage supply.

• As with connecting directly to the main power network, for variable- frequency drive (VFD) drives it is

march 2014 24 compressortech2

n This cutaway drawing shows a high flexible (rubber in shear) coupling.

CT335.indd 2 2/21/14 4:08 PM

march 2014 25 compressortech2

not only the dynamic shafting torque but also the motor electric current fluctuations that have to be considered dur-ing the drivetrain design.

ConclusionsIt could be demonstrated that the motor dynamics rep-

resent the major influence on the torsional vibrations of drivetrains — especially on the shaft torque loading and on electric current pulsations.

The inclusion of motor dynamics in drivetrain simula-tions has increased equipment reliability. This has be-come even more important today, as modern compressor installations provide a stronger demand on calculation accuracy, due to better strength utilization. Therefore, the motor electromagnetics should always be included into the calculations.

When using torsionally soft couplings, the torque loading value can become quite low due to the motor reaction. Motor and soft coupling represent a very good combination and pro-duce robust systems that show only small torque pulsations at coupling and shafts.

When torsionally rigid couplings like all-steel, disc-type couplings are used, self-induced vibrations may occur with sometimes harmful amplitudes. The resulting shaft and coupling torque variations can be multiples of the calculated values due to the false assumption of constant motor torque.

At VFD, the frequency converter control method must be thoroughly chosen in order to promote or reduce the mo-tor magnetic field influence, depending on the individual requirements.

As a compressor and motor can only be simulated in a single combined model, compressor manufacturers must un-derstand the dynamic motor behavior, allowing them to opti-mize the drivetrain on every relevant aspect such as proper motor inertia, coupling, flywheel and crankshaft designs.

Neuman & Esser has developed expertise in evaluat-ing motor dynamics for drivetrain simulations (especially motor electromagnetic characteristics), increasing the reli-ability of high-pressure reciprocating compressors used in gas applications. CT2

n This is an all-steel, disc-type coupling.

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CT335.indd 3 2/21/14 4:06 PM

Motortech has extended its range of high-end ignition con-trollers with two new series.

They are the MIC3, which has 12 ig-nition outputs and 200 mJ of primary energy for small and mid-sized gas engines, and the MIC5 with 20 ignition outputs and 500 mJ primary energy for engines with up to 670 hp (500 kW).

The Celle, Germany-based com-pany said that high ignition energy, accurate spark timing and diversified online diagnostics can combine to improve engine efficiency, spark plug life and availability of the equipment under the strict emissions regulations. The controller is fully customer con-figurable via a laptop.

Motortech has also extended its VariFuel2 air/fuel ratio mixer series with two new sizes to cover, or re-spectively overlap, a greater range of application areas.

The main task of the gas mixer is to mix the fuel (gas) and air so that the gas engine achieves optimal combustion.

The VariFuel2 is a high-tech, vari-able Venturi-type mixer that can con-stantly adjust to any fuel changes. It allows the engine to operate at its most efficient point.

Series 100,140, 200 and 250 Vari-Fuels are available for engines with an air requirement up to 176 Mcfh (5000 m3/h).

Coupled to an air/ fuel ratio control-ler, lean-burn or stoichiometric, it pre-cisely regulates the mixture and, ac-cording to the company, is suitable for applications with constant changes in the calorific value of fuel.

It can be used with nearly all gas types, including natural, biogas, land-fill, sewage, wood, wellhead or mine and it employs a high precision step-per motor drive with a reprogramma-ble controller board called VariStep.

Various flow bodies and flexible in-let and outlet configurations allow fully flexible cross section adjustment.

Motortech said the decisive optimi-zation parameters are a high degree

of efficiency and low emissions that comply with relevant regulations.

In the VariFuel2, gas and air are mixed based on the Venturi effect. Based on the suction vacuum of the engine, the air is sucked through the air inlet into the Venturi nozzle.

The Venturi effect produces an un-derpressure at the narrowest point, sucking the gas in through the gas inlet. This way, gas and air are mixed and re-leased at the Venturi outlet. Based on different design sizes and different flow bodies in the Venturi nozzle, it is pos-sible to achieve various volume flows.

Another recent Motortech prod-uct launch was a handheld device to support gas engine mechanics who are trouble shooting engine problems or going through preventive mainte-nance procedures. SparkView moni-tors the high voltage required by the spark plug when firing.

To see a video of the SparkView application, go to: www.youtube.com/watch?v=YcC94wnhzQc. CT2

Motortech Adds High-End Ignition Controllers > It also expands offerings of

fuel mixersBY IAn CAMeron

n This is a grouping of VariFuel2s available from Motortech.

MARCH 2014 26 CoMpRessoRtech2

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Dresser-Rand has announced the successful operation of its newly developed small-scale

liquefied natural gas (LNG) production system, named LNGo.

“We are very excited about this technology for small-scale LNG pro-duction, which allows for very small stand-alone plants that are portable and can be moved to support chang-ing requirements and needs,” Presi-dent and CEO Vincent Volpe Jr. said. “The standard LNGo plants are sized to produce approximately 6000 gal-lons (22,712 L) of LNG per day.”

Dresser-Rand designed, construct-ed and commissioned its demonstra-tion plant, which in December reached a milestone with the initial production of LNG. Now, extended performance and endurance tests are being per-formed as the final step prior to full market release.

According to Charles Ely, general manager of midstream SBU, the LNGo

system operates on the open-loop methane concept. The process is iden-tified as a methane letdown cycle.

The process module contains a me-chanical chiller, turboexpander, Joule-Thomson valve, stainless-steel, brazed welded plate heat exchangers and liq-

uefaction process piping to drop the pressure and reduce the temperature of the natural gas.

The first phase of cooling is ac-complished by the mechanical chiller, which uses an ammonia refrigerant loop and evaporative cooler. The in-tegrated ammonia chilling system consists of an evaporative condenser and screw compressor that removes the heat of compression from the methane prior to expansion. The ad-vantage of this first stage is lower cost and smaller footprint.

The second phase of cooling is ac-complished with the turboexpander. The expander wheel uses variable in-let guide vanes to control inlet gas flow and maximize efficiency, and is fitted with a gas seal and anti-surge control systems to assure reliable operation. Final subcooling is accomplished with a Joule-Thomson valve that rapidly expands natural gas and reduces tem-perature to below -250°F (-157°C). It is used to supply the final cryogenic heat exchanger with cooling medium.

The conditioning module, also known as the molecular sieve, cleans and

Dresser-Rand Makes Major Move Into Gas Liquefaction > Tests completed for small-scale LNG

production systemBy JOE KANE

n The compressor module contains a four-throw Dresser-Rand 7MOS4TM reciprocating com-pressor to provide all stages of compression and the interconnecting pipe to the process module.

MARCH 2014 28 CoMpRessoRtech2

continued on page 30

n The LNGo system is shown with three-dimensional CAD software created by the development team at Painted Post, New York.

CT345.indd 1 2/25/14 2:15 PM

Kilowatts.indd 1 10/23/13 11:01 AM

n The liquefaction process module does all the cooling and subcooling of the natural gas in the system.

sep arates the incoming gas into two streams. The product stream, now free of H2O or CO2, is fed to a Dress-er-Rand MOS compressor to begin the liquefaction process. The purge stream is essentially a waste stream that contains H2O, CO2 and heavier hydrocarbons that are not permis-sible for making LNG. It is blended with makeup natural gas and used to fuel the Guascor engine, avoiding ad-ditional gas treatment or cleanup.

The power module contains a Dresser-Rand Guascor, 16-cylinder, four-cycle, turbocharged and after-cooled, rich-burn engine (rated 870 kW at 1800 rpm and bmep of 151 psi or 10.4 bar). Bore and stroke are 6.3 x 6.9 in. (160 x 175 mm) giving a displacement of 3436 cu. in. (56.35 L) and a compression ratio of 9.3:1. Rat-ed emissions are 0.1/0.2/0.2 g/bhph (0.13/0.27/0.27 g/kWh) for NOx/CO/HMHC. A fan-driven ambient air cooler is used to cool engine jacket and lube oil circuits. Dresser-Rand’s Enginuity PLC-based control system serves both the power module and the LNGo pro-cess. The power module also contains the motor control center (MCC).

With an engine configured to han-dle natural gas within a lower heating value (LHV) range of 800 to 1150 Btu/cf (29,800 to 42,870 kJ/m3), the power module provides power for the entire LNGo system. Fuel gas is a blended mixture of mole sieve purge (waste) gas and feed natural gas.

The compressor module satisfies all four stages of gas compression re-

quired by the LNGo liquefaction system.Dresser-Rand’s four-stroke model

7MOS4 reciprocating compressor has a 7 in. (172 mm) stroke and bore sizes (C1/C2/C3/C4) of 10.5/10.5/17.5/11.5 in. (267/267/444.5/281.6 mm). Com-pression ratios of the four stages are (C1/C2/C3/C4): 3.25/3.5/2.74/2.85. The compressor features Dresser-Rand’s Magnum XF series valves for proven combination of efficiency and durability in high-pressure ratio applications.

It uses nonlubricated cylinders to avoid process contamination and elim-inate lube oil consumption. A water- glycol coolant is used to cool compres-sor cylinders, rod packings and the crankshaft and rod bearings lube oil is fed by a mechanically driven pump during normal operation and electric motor-pump during start-up.

The compressor is driven by a 12-pole, three-phase WEG induction mo-tor rated for 750 hp (559 kW) at 480V. It includes VFD with harmonic filter.

“Our development process began in earnest less than one year ago,” Volpe said. “In this time frame, our or-

ganization has taken the process and designed, built and commissioned an entire plant, with a target capacity of 6000 gpd (2.2 kpd). This development cycle time to market is amazingly short, and is a tribute to our internal processes, as well as to the men and women of our development team, and our Painted Post, New York, opera-tions. We are proud of our people and their outstanding efforts.”

On the commercial side, the com-pany has been talking to potential clients for several months. Depend-ing upon the nature of the applica-tion, these potential opportunities are broadly classified as modules that would either be for sale or for lease.

For direct sale, Dresser-Rand pro-vides the liquefaction process and associated ancillary gas processing equipment, a power module and full turnkey installation and commissioning.

Dresser-Rand currently offers the LNGo system for direct sale but is evaluating other market strategies such as rental or lease options with market channel partners.

For all users, Dresser-Rand can pro-vide full turnkey installation and com-missioning services, as well as routine operations, monitoring and mainte-nance contracts to ensure ongoing re-liable and available operations.

Upstream applications include, among others, the monetization of flared gas to increase revenues for oil companies and reduce their en-vironmental impact, the production of stranded natural gas fields that are not close to existing pipeline in-frastructures, on-site fuel supply for

MARCH 2014 30 CoMpRessoRtech2

n Dresser-Rand’s Guascor generator set takes natural gas and converts it into electrical energy for the rest of the system to operate.

CT345.indd 2 2/25/14 2:20 PM

www.ACIServicesInc.com 740-435-0240

Don’t Just Go With The FlowExceed Your Expectations

ACI provides a robust line of performance control devices to help youoptimize the operation of your reciprocating compressors.

ACI.indd 1 2/4/14 8:27 AM

drilling and hydraulic fracturing equip-ment converted to run on LNG and applications for coal bed methane for fueling mining vehicles. Downstream applications include the production of vehicle-grade LNG, allowing LNG to compete effectively with diesel fuel on a cost-per-energy-content (Btu) basis.

As LNGo plants enable the “distrib-uted” production of LNG on a small scale, the technology eliminates the need for the costly trucking of LNG

long distances from large, central-ized plants to LNG fueling depots, as is the practice today. Further, Dresser-Rand believes that its ap-proach to short cycle time will enable LNGo plants to be installed and op-erating in months rather than years. The short cycle times will allow own-ers to see quick returns on their in-vestment, as well as matching the supply and demand of LNG as local markets develop.

Driven by rapidly expanding global shale gas development and continued price differentials between natural gas and oil, Dresser-Rand predicts that the market for distributed, small-scale LNG production plants will grow from early adopters in North America to a broad, robust market for users around the world. North America is the most rapidly growing market. The substantial price disparity between diesel fuel and low-priced natural gas has oil-field service operators, oil and gas companies, shipping and deliv-ery companies, and downstream fuel distributors/marketers across the region converting drilling rigs, industri-al mining equipment, transportation fleets and retail fueling stations along the United States interstate highway system to LNG fuel.

Dresser-Rand said orders booked for LNGo in the next several months will convert to shipments in 2014. As those units are placed into service and gain operating experience and runtime, the company expects the incoming stream of orders to grow over time. CT2

n A simple schematic representation of the process shows the progress from input natural gas to the output of LNG.

CT345.indd 3 2/21/14 5:41 PM

Richard (Rick) H. Dearing Jr., the current president and grandson of the founder, joined the company in 1985 af-ter receiving petroleum engineering and MBA degrees. Rebecca Dearing Wall is the executive vice president, chief fi-nancial officer, sister and co-owner with Rick. She joined the company in 1981.

“The company has grown steadily since we purchased it from our father in 1996,” Dearing Wall said. “Dearing is established as a leader in the engi-neering, packaging and installation of quality compressor equipment.”

Rick Dearing said until about 2005,

packages in the 100 to 300 hp (75 to 224 kW) range were common, and 500 hp (373 kW) was a big unit for the com-pany. “But customers demanded that we be able to do it all or we wouldn’t get any of their other business, so we kept expanding to larger and larger units,” Rick Dearing said.

Until August 2006, Dearing operated in a 26,000 sq.ft. (2416 m2) facility with limited crane capacity. A 10,000 sq.ft. (929 m2) expansion helped the Appala-chian Basin packager continue several years of steady growth. And then the Marcellus Shale emerged. “Business

n Dearing’s plant has eight fabrication bays with 180 tons (163 tonnes) of crane capacity and is being expanded to 12 large bays in 2014. The large package in the foreground includes a Waukesha 9390 gas engine, Ariel JGD/4 inlet gas compressor, and AXH air cooler.

MARCH 2014 32 CoMpRessoRtech2

Dearing Compressor and Pump Co. of Youngstown, Ohio,

has found itself in the right place and the right time to supply the Marcellus and Utica shale plays.

Albin P. Dearing III, who sold and serviced Gardner Denver industrial air compressors, formed Dearing in 1945. The family owned company’s introduc-tion to the oil and gas business came in 1960 by working on Gardner Denver oil rigs and compressors. This later led to an expansion into the natural gas compressor market within Ohio.

Dearing Compressor has served the Appalachian Basin since 1945

BY NORM SHADe

Custom PackagerSTRIkeS GOlD

In The Marcellus Shale

CT341.indd 1 2/21/14 4:20 PM

COMPRESSORDedicated To Gas Compression Products & Applications

PACKAGER GUIDE2014www.compressortech2.com

CT2_PackagerGuide_2014.indd 1 2/24/14 2:20 PM

Establishing a comprehensive listing of compressor packagers, their locations and package capacity ranges is an important service to the end users of this equipment. This addition of our services to the industry is based on several inquiries we received regarding compressor packagers. A listing of packagers follows, along with contact information, types of compressors offered and the capacity range of the packages they produce. An important note — if your company is missing from this listing, please let us know, as it will be updated on a regular basis.

COMPRESSORDedicated To Gas Compression Products & Applications

• AG Equipment Co., Broken Arrow, Oklahoma. Principal contact: Kent Bright. E-mail: kbright@agequipmentcompany.com. Types of compressors: reciprocating and rotary screw. Capacity range: 20 to 10,000 hp (15 to 7456 kW).

• ANGI Energy Systems LLC, Janesville, Wisconsin. Principal contact: Jared Hightower, vice president, domestic CNG Sales. E-mail: jhightower@angienergy.com. Types of compressors: reciprocating and rotary screw. Capacity range: 40 to 800 hp (30 to 597 kW).

• Abby Services Inc., Canonsburg, Pennsylvania. Principal contact: Don Fulmer, president. E-mail: sales@abbyservicesinc.com. Types of compressors: rotary screw, reciprocating, liquid rings and blowers. Capacity range: 3 to 500 hp (2 to 373 kW).

• ABC Compressors, Eibar, Spain. Principal contact: Javier Cuevas. E-mail: jcuevas@abc-compressors.com. Types of compressors: reciprocating. Capacity range: 70 to 1600 hp (50 to 1200 kW).

• Alegacy Equipment, Waller, Texas. Principal contact: Bo Pierce. E-mail: b.pierce@alegacy.biz. Types of compressors: reciprocating. Capacity range: 30 to 400 hp (22 to 298 kW).

• Arrow Engine Co., Tulsa, Oklahoma. Principal contact: Terry Kerbo, general manager, compression. E-mail: tkerbo@arrowengine.com. Types of compressors: reciprocating. Capacity range: 25 to 300 hp (19 to 224 kW).

• Bidell Gas Compression, Calgary, Alberta, Canada. Principal contact: Mat Clark, vice president, sales and applications. E-mail: mclark@bidell.com. Types of compres-sors: reciprocating and rotary screw. Capacity range: up to 10,000 hp (7456 kW).

• Brahma Compression, Calgary, Alberta, Canada. Principal contact: Phil Me-loche, sales. E-mail: philm@brahmacompression.com. Types of compressors: blow-ers, vane and rotary screw. Capacity range: 5 to 400 hp (4 to 298 kW).

• Cameron, Houston. Principal contact: Mike Gerzina. E-mail: Mike.Gerzina@c-a-m.com. Type of compressor: reciprocating. Capacity range: 148 to 9000 hp (110 to 6711 kW).

• Cobey Inc., Buffalo, New York. Principal contact: Eric McKendry, director of mar-keting and sales. E-mail: emckendry@cobey.com. Types of compressors: recipro-cating, screw, centrifugal, axial, with expanders and turbine-generator sets, lube oil systems. Capacity range: up to 30,000 hp (22,065 kW).

• Comoti – Romanian Research and Development Institute for Gas Tur-bines, Bucharest, Romania. Principal contact: Marius Teodorescu, marketing and sales manager. E-mail: marius.teodorescu@comoti.ro. Types of compressors: ro-tary screw, centrifugal and blowers. Capacity range: 30 to 2448 hp (22 to 3280 kW).

• Com-Pac Systems Inc., Odessa, Texas. Principal contact: Jack Motley, president. E-mail: jackmotley@compressorpackaging.com. Types of compressors: rotary vane, rotary screw, and reciprocating. Capacity range: 25 to 4000 hp (19 to 2983 kW).

• Compass Compression Services Ltd., Calgary, Alberta, Canada. Principal contact: Scott Douglas, vice president, sales. E-mail: sdouglas@compasscompres-sion.com. Types of compressors: reciprocating, rotary screw and vane. Capacity range: 5 to 8000 hp (4 to 5965 kW).

• Compressor Systems Inc., Midland, Texas. Principal contact: Hank Sheeran, vice president, sales. E-mail: hank.sheeran@compressor-systems.com. Types of com-pressors: reciprocating and screw. Capacity range: 26 to 8500 hp (19 to 6338 kW).

• ConPackSys, Dordrecht, Netherlands. Principal contact: Michel Bezemer. E-mail:

michel.bezemer@conpacksys.nl. Type of compressors: reciprocating. Capacity range: 44 to 9383 hp (33 to 7000 kW).

• Custom Compression Systems, New Iberia, Louisiana. Principal contact: Bob Carter. E-mail: bcarter@customcompressionsystems.com. Type of compressors: reciprocating. Capacity range: 95 to 5000 hp (71 to 3728 kW).

• Dearing Compressor & Pump Co., Youngstown, Ohio. Principal contact: Rich-ard H. Dearing Jr., president. E-mail: rick@dearingcomp.com. Types of compres-sors: rotary screw, reciprocating, blowers and liquid ring. Capacity range: 30 to 8000 hp (22 to 5965 kW).

• Dresser-Rand, Houston. Principal contact: Colman DeJong, VP of sales – The Americas. E-mail: cdejong@dresser-rand.com. Type of compressors: centrifugal, reciprocating. Capacity range: 20,100 to 181,000 hp (15,000 to 135,000 kW).

• Elliott Co., Jeannette, Pennsylvania. Principal contact: Tom Brown, marketing manager. E-mail: tbrown@elliott-turbo.com. Type of compressors: centrifugal. Ca-pacity range: 10,000 to 120,000 hp (7456 to 89,500 kW).

• Enerflex Ltd., Calgary, Alberta, Canada. Principal contact: Trevor Hunt, manager, Sales, Compression and Process. E-mail: thunt@enerflex.com. Houston, Texas. Principal contact: Peter Kourkoubes, manager, Sales, USA and Latin America. E-mail: pkourkoubes@enerflex.com. Types of compressors: reciprocating and rotary screw. Capacity range: up to 10,000 hp (7456 kW).

• Enerproject SA, Mezzovico, Switzerland. Principal contact: Vito Notari, sales manager. E-mail: vito.notari@enerproject.com. Types of compressors: centrifugal, rotary screw and rotary vane. Capacity range: up to 4024 hp (3000 kW).

• Euro Gas Systems SRL, Targu Mures, Romania. Principal contact: Roger Wachter, general manager. E-mail: roger.wachter@eurogassystems.com. Types of compressors: reciprocating and rotary screw. Capacity range: 100 to 5000 hp (75 to 3728 kW).

• Exterran Compression, Houston. Principal contact: Susan Nelson, marketing/communications. E-mail: susan.nelson@exterran.com. Types of compressors: re-ciprocating and rotary screw. Capacity range: 250 to 9000 hp (186 to 6710 kW).

• FIMA Maschinenbau GmbH, Obersontheim, Germany. Principal contact: Mi-chael Loercher, sales engineer. E-mail: m.loercher@fima.de. Types of compressors: centrifugal and seal-less. Capacity range: 10 to 6800 hp (8 to 5000 kW).

• Flatrock Compression Ltd., Houston. Principal contact: Brian McDonald, presi-dent. E-mail: brian.mcdonald@flatrockcompression.com. Type of compressors: re-ciprocating. Capacity range: 26 to 500 hp (19 to 373 kW).

• Flogistix, Oklahoma City. Principal contact: Drake Andarakes, vice president of sales and marketing. E-mail: sales@flogistix.com. Type of compressors: rotary screw. Capacity range: 20 to 800 hp (15 to 597 kW).

• GEA Refrigeration Italy, Castel Maggiore, Italy. Principal contact: Ivano Cama-ggi, president, Power Technology Center. E-mail: ivano.camaggi@geagroup.com. Type of compressors: screw, reciprocating and centrifugal. Capacity range: up to 13,412 hp (10,000 kW).

• GE Oil & Gas, Florence, Italy. Principal contact: Sara Hassett, Communications Leader. E-mail: sara.hassett@ge.com. Type of compressors: centrifugal and axial compressors; integrated electric motor driven compressors; reciprocating and high- speed reciprocating compressors; gas and steam turbines, turboexpanders and hot gas expanders. Capacity range: 20,400 to 95,200 hp (15,000 to 70,000 kW).

Packager Guide 2014• GE Oil & Gas do Brasil Ltd., Rio de Janeiro. Principal contact: Rogerio Freitas da Fonseca, regional senior sales leader for Latin America. E-mail: rogerio.fonse-ca@ge.com. Type of compressors: high-speed reciprocating. Capacity range: 60 to 7200 hp (44 to 5371 kW).

• G.I. & E. S.p.A., Porto Recanati, Italy. Principal contact: Donatello Vocca, sales and marketing director. E-mail: donatello.vocca@gie.it. Type of compressors: recip-rocating. Capacity range: 70 to 7000 hp (50 to 5000 kW).

• Great Plains Gas Compression Inc., Hugoton, Kansas. Principal contact: Ter-ry R. McBride, vice president of sales and marketing. E-mail: tmcbride@greatplain-sgas.com. Types of compressors: reciprocating, rotary screw, rotary vane, blowers, vapor recovery, CNG fueling stations, high spec, and electric. Capacity range: 5 to 5000 hp (3.7 to 3729 kW).

• HBR Equipamentos ltda, Sao Paulo, Brazil. Principal contact: Valdir Zuffo. E-mail: valdir@hbrarcomprimido.com.br. Type of compressors: rotary screw, recipro-cating, centrifugal. Capacity range: up to 5000 hp (3728.5 kW).

• Henry Production Inc., Farmington, New Mexico. Principal contact: Sam Henry. E-mail: samh@hpi1.com. Types of compressors: rotary screw, reciprocating and scroll. Capacity range: up to 250 hp (186 kW).

• Howden Process Compressors, Renfrew, U.K. E-mail: (new systems) hpc@howden.com; (service and parts) hpc.aftersales@howden.com. Type of compres-sors: screw. Capacity range: up to 6705 hp (5000 kW).

• Industrias Juan F. Secco S.A., Rosario, Argentina. Principal contact: Augusto F. Beni. E-mail: abeni@jfsecco.com.ar. Type of compressors: reciprocating. Capacity range: up to 6500 hp (4846 kW).

• J-W Energy Co., Dallas. Principal contact: James R. Barr. E-mail: jbarr@jwen-ergy.com. Type of compressors: reciprocating and rotary screw. Capacity range: 25 to 4500 hp (19 to 3355 kW).

• Kingsly Compression Inc., Cambridge, Ohio; Saxonburg, Pennsylvania. Principal contact: Jeffrey B. Sable. E-mail: jeffrey.b.sable@kingslycompression.com. Types of compressors: reciprocating, rotary screw. Capacity range: 5 to 1000 hp (3.7 to 745.7 kW).

• McClung Energy Services, Longview, Texas. Principal contact: Tim McDonald, production manager. E-mail: tim@mcclungenergy.com. Types of compressors: recip-rocating. Capacity range: 50 to 400 hp (37.3 to 298.3 kW).

• Natural Gas Compression Systems Inc., Traverse City, Michigan. Prin-cipal contact: Bill Jenkins, vice president, sales. E-mail: jenkins@ngcsi.com. Types of compressors: rotary screw and reciprocating. Capacity range: up to 2500 hp (1864 kW).

• Neuman & Esser U.S.A Inc., Katy, Texas. Principal contact: Scott DeBaldo. E-mail: sdebaldo@neuman-esser.com. Type of compressors: horizontal, vertical and V-type reciprocating. Capacity range: up to 20,000 hp (15,000 kW).

• OTA Compression LLC, Irving, Texas. Principal contact: Vickie L. Gage-Tims, vice president, sales. E-mail: sales@OTACompression.com. Types of compressors: reciprocating. Capacity range: up to 100 hp (75 kW).

• Palmero San Luis S.A., Buenos Aires, Argentina. Principal contact: Matias Maggi. E-mail: mmaggi@palmero.com. Types of compressors: reciprocating and rotary screw. Capacity range: up to 6500 hp (4846 kW).

• Propak Systems Ltd., Airdrie, Alberta, Canada. Principal contact: Ron Delisle. E-mail: rdelisle@propaksystems.com. Types of compressors: rotary screw and re-ciprocating. Capacity range: up to 10,000 hp (7456 kW).

• PSE Engineering GmbH, Hannover, Germany. Principal contact: Dirk Heyer,

CT2_PackagerGuide_2014.indd 2 2/26/14 3:49 PM

Establishing a comprehensive listing of compressor packagers, their locations and package capacity ranges is an important service to the end users of this equipment. This addition of our services to the industry is based on several inquiries we received regarding compressor packagers. A listing of packagers follows, along with contact information, types of compressors offered and the capacity range of the packages they produce. An important note — if your company is missing from this listing, please let us know, as it will be updated on a regular basis.

michel.bezemer@conpacksys.nl. Type of compressors: reciprocating. Capacity range: 44 to 9383 hp (33 to 7000 kW).

• Custom Compression Systems, New Iberia, Louisiana. Principal contact: Bob Carter. E-mail: bcarter@customcompressionsystems.com. Type of compressors: reciprocating. Capacity range: 95 to 5000 hp (71 to 3728 kW).

• Dearing Compressor & Pump Co., Youngstown, Ohio. Principal contact: Rich-ard H. Dearing Jr., president. E-mail: rick@dearingcomp.com. Types of compres-sors: rotary screw, reciprocating, blowers and liquid ring. Capacity range: 30 to 8000 hp (22 to 5965 kW).

• Dresser-Rand, Houston. Principal contact: Colman DeJong, VP of sales – The Americas. E-mail: cdejong@dresser-rand.com. Type of compressors: centrifugal, reciprocating. Capacity range: 20,100 to 181,000 hp (15,000 to 135,000 kW).

• Elliott Co., Jeannette, Pennsylvania. Principal contact: Tom Brown, marketing manager. E-mail: tbrown@elliott-turbo.com. Type of compressors: centrifugal. Ca-pacity range: 10,000 to 120,000 hp (7456 to 89,500 kW).

• Enerflex Ltd., Calgary, Alberta, Canada. Principal contact: Trevor Hunt, manager, Sales, Compression and Process. E-mail: thunt@enerflex.com. Houston, Texas. Principal contact: Peter Kourkoubes, manager, Sales, USA and Latin America. E-mail: pkourkoubes@enerflex.com. Types of compressors: reciprocating and rotary screw. Capacity range: up to 10,000 hp (7456 kW).

• Enerproject SA, Mezzovico, Switzerland. Principal contact: Vito Notari, sales manager. E-mail: vito.notari@enerproject.com. Types of compressors: centrifugal, rotary screw and rotary vane. Capacity range: up to 4024 hp (3000 kW).

• Euro Gas Systems SRL, Targu Mures, Romania. Principal contact: Roger Wachter, general manager. E-mail: roger.wachter@eurogassystems.com. Types of compressors: reciprocating and rotary screw. Capacity range: 100 to 5000 hp (75 to 3728 kW).

• Exterran Compression, Houston. Principal contact: Susan Nelson, marketing/communications. E-mail: susan.nelson@exterran.com. Types of compressors: re-ciprocating and rotary screw. Capacity range: 250 to 9000 hp (186 to 6710 kW).

• FIMA Maschinenbau GmbH, Obersontheim, Germany. Principal contact: Mi-chael Loercher, sales engineer. E-mail: m.loercher@fima.de. Types of compressors: centrifugal and seal-less. Capacity range: 10 to 6800 hp (8 to 5000 kW).

• Flatrock Compression Ltd., Houston. Principal contact: Brian McDonald, presi-dent. E-mail: brian.mcdonald@flatrockcompression.com. Type of compressors: re-ciprocating. Capacity range: 26 to 500 hp (19 to 373 kW).

• Flogistix, Oklahoma City. Principal contact: Drake Andarakes, vice president of sales and marketing. E-mail: sales@flogistix.com. Type of compressors: rotary screw. Capacity range: 20 to 800 hp (15 to 597 kW).

• GEA Refrigeration Italy, Castel Maggiore, Italy. Principal contact: Ivano Cama-ggi, president, Power Technology Center. E-mail: ivano.camaggi@geagroup.com. Type of compressors: screw, reciprocating and centrifugal. Capacity range: up to 13,412 hp (10,000 kW).

• GE Oil & Gas, Florence, Italy. Principal contact: Sara Hassett, Communications Leader. E-mail: sara.hassett@ge.com. Type of compressors: centrifugal and axial compressors; integrated electric motor driven compressors; reciprocating and high- speed reciprocating compressors; gas and steam turbines, turboexpanders and hot gas expanders. Capacity range: 20,400 to 95,200 hp (15,000 to 70,000 kW).

• GE Oil & Gas do Brasil Ltd., Rio de Janeiro. Principal contact: Rogerio Freitas da Fonseca, regional senior sales leader for Latin America. E-mail: rogerio.fonse-ca@ge.com. Type of compressors: high-speed reciprocating. Capacity range: 60 to 7200 hp (44 to 5371 kW).

• G.I. & E. S.p.A., Porto Recanati, Italy. Principal contact: Donatello Vocca, sales and marketing director. E-mail: donatello.vocca@gie.it. Type of compressors: recip-rocating. Capacity range: 70 to 7000 hp (50 to 5000 kW).

• Great Plains Gas Compression Inc., Hugoton, Kansas. Principal contact: Ter-ry R. McBride, vice president of sales and marketing. E-mail: tmcbride@greatplain-sgas.com. Types of compressors: reciprocating, rotary screw, rotary vane, blowers, vapor recovery, CNG fueling stations, high spec, and electric. Capacity range: 5 to 5000 hp (4 to 3728 kW).

• HBR Equipamentos ltda, Sao Paulo, Brazil. Principal contact: Valdir Zuffo. E-mail: valdir@hbrarcomprimido.com.br. Type of compressors: rotary screw, recipro-cating, centrifugal. Capacity range: up to 5000 hp (3728 kW).

• Henry Production Inc., Farmington, New Mexico. Principal contact: Sam Henry. E-mail: samh@hpi1.com. Types of compressors: rotary screw, reciprocating and scroll. Capacity range: up to 250 hp (186 kW).

• Howden Process Compressors, Renfrew, U.K. E-mail: (new systems) hpc@howden.com; (service and parts) hpc.aftersales@howden.com. Type of compres-sors: screw. Capacity range: up to 6705 hp (5000 kW).

• Industrias Juan F. Secco S.A., Rosario, Argentina. Principal contact: Augusto F. Beni. E-mail: abeni@jfsecco.com.ar. Type of compressors: reciprocating. Capacity range: up to 6500 hp (4846 kW).

• J-W Energy Co., Dallas. Principal contact: James R. Barr. E-mail: jbarr@jwen-ergy.com. Type of compressors: reciprocating and rotary screw. Capacity range: 25 to 4500 hp (19 to 3355 kW).

• Kingsly Compression Inc., Cambridge, Ohio; Saxonburg, Pennsylvania. Principal contact: Jeffrey B. Sable. E-mail: jeffrey.b.sable@kingslycompression.com. Types of compressors: reciprocating, rotary screw. Capacity range: 5 to 1000 hp (4 to 746 kW).

• McClung Energy Services, Longview, Texas. Principal contact: Tim McDonald, production manager. E-mail: tim@mcclungenergy.com. Types of compressors: recip-rocating. Capacity range: 50 to 400 hp (37 to 298 kW).

• Natural Gas Compression Systems Inc., Traverse City, Michigan. Prin-cipal contact: Bill Jenkins, vice president, sales. E-mail: jenkins@ngcsi.com. Types of compressors: rotary screw and reciprocating. Capacity range: up to 2500 hp (1864 kW).

• Neuman & Esser U.S.A Inc., Katy, Texas. Principal contact: Scott DeBaldo. E-mail: sdebaldo@neuman-esser.com. Type of compressors: horizontal, vertical and V-type reciprocating. Capacity range: up to 20,000 hp (15,000 kW).

• OTA Compression LLC, Irving, Texas. Principal contact: Vickie L. Gage-Tims, vice president, sales. E-mail: sales@OTACompression.com. Types of compressors: reciprocating. Capacity range: up to 100 hp (75 kW).

• Palmero San Luis S.A., Buenos Aires, Argentina. Principal contact: Matias Maggi. E-mail: mmaggi@palmero.com. Types of compressors: reciprocating and rotary screw. Capacity range: up to 6500 hp (4846 kW).

• Propak Systems Ltd., Airdrie, Alberta, Canada. Principal contact: Ron Delisle. E-mail: rdelisle@propaksystems.com. Types of compressors: rotary screw and re-ciprocating. Capacity range: up to 10,000 hp (7456 kW).

• PSE Engineering GmbH, Hannover, Germany. Principal contact: Dirk Heyer,

division manager, compression systems. E-mail: info@pse-eng.de. Type of com-pressors: reciprocating. Capacity range: 100 to 10,000 hp (75 to 7456 kW).

• Reagan Power and Compression Inc. Broussard, Louisiana. Principal contact: Francis A. Orso, director of sales. E-mail: forso@reaganp ower.com. Type of com-pressors: reciprocating, screw (gas and electric motor driven); Capacity range: up to 10,000 hp (7456 kW).

• Safe, San Giovanni in Persiceto (BO), Italy. Principal contact: Michele Petraccone. E-mail: michele.petraccone@safe-ita.com. Type of compressors: reciprocating. Ca-pacity range: 67 to 4700 hp (50 to 3500 kW).

• SAGE Energy Corp., Calgary, Alberta, Canada. Principal contact: Trent Bruce. E-mail: trent.bruce@sageenergy.ca. Types of compressors: rotary screw and recip-rocating. Capacity range: 30 to 5000 hp (22 to 3728 kW).

• S&R Compression LLC, Tulsa, Oklahoma. Principal contact: G. Durland. E-mail: gdurland@sandrequipment.com. Types of compressors: vapor recovery, screw and reciprocating. Capacity range: up to 400 hp (298 kW).

• SEC Energy Products & Services, Houston. Principal contact: Frank Northup, director, sales and marketing. E-mail: fnorthup@sec-ep.com. Type of compressors: reciprocating. Capacity range: 50 to 10,000 hp (37 to 7456 kW).

• Sertco, Okemah, Oklahoma. Principal contact: Matt Smith. E-mail: matt@sertco.com. Types of compressors: natural gas and reciprocating. Capacity range: 20 to 200 hp (15 to 149 kW).

• Siemens AG, Energy, Oil & Gas, Turbo Equipment; Duisburg, Germany (rotating equipment); Hengelo, Netherlands (packaging). Principal contact: Michaela Niss. E-mail address: michaela.niss@siemens.com. Type of compressors: turbocompres-sors. Capacity range: 6635 to 40,348 hp (4950 to 30,100 kW).

• Siemens Sp. z o.o., Elblag, Poland. Principal contact: Tomasz Grzegolkowski. E-mail: tomasz.grzegolkowski@siemens.com. Type of compressors: reciprocating. Capacity range: 670 to 10,700 hp (500 to 8000 kW).

• Siad Macchine Impianti S.p.A., Bergamo, Italy. Principal contact: Mauro Ac-quati, Compressor Division sales manager. E-mail: mauro_acquati@siad.eu. Type of compressors: reciprocating. Capacity range: 27 to 3220 hp (20 to 2400 kW).

• Solar Turbines Incorporated, San Diego. Principal contact: sales department, applications engineering. E-mail: powergen@solarturbines.com. Type of compres-sors: centrifugal. Capacity range: 13,097 to 77,707 hp (9800 to 57,900 kW).

• Speir Energy Solutions, Okemah, Oklahoma. Principal contact: Thompson Speir, owner. E-mail: thompson@speircompression.com. Types of compressors: reciprocating, screw, rotary. Capacity range: up to 75 hp (56 kW).

• Startec Refrigeration Services Ltd., Calgary, Alberta, Canada. Principal con-tact: Mike Tearoe, vice president, corporate development, compression and process division. E-mail: mtearoe@startec.ca. Types of compressors: reciprocating, rotary screw and vane. Capacity range: 5 to 8000 hp (4 to 5965 kW).

• UE Compression LLC, Henderson, Colorado. Principal contact: Steven Tyler, sales. E-mail: styler@uecompression.com. Types of compressors: rotary vane, ro-tary screw and reciprocating. Capacity range: 50 to 7500 hp (37 to 5592 kW).

• Valerus, Houston. Principal contact: Maggie Seeliger, director of marketing. E-mail: mseeliger@valerus.com. Type of compressors: reciprocating. Capacity range: 75 to 5000 hp (56 to 3728 kW).

• VPT Kompresssoren GmbH, Remscheid, Germany. Principal contact: Carsten Kollenbach, sales engineer, export. E-mail: c.kollenbach@vpt-kompressoren.de. Type of compressors: screw. Capacity range: 7 to 5365 hp (5 to 4000 kW).

COMPRESSION TECHNOLOGY SOURCING SUPPLEMENTCTSSnet.net

CT2_PackagerGuide_2014.indd 3 2/26/14 3:55 PM

Untitled-2 1 2/19/14 4:12 PMCT2_PackagerGuide_2014.indd 4 2/24/14 2:21 PM

It provided 50,000 sq.ft. (4645 m2) of shop space and 10,000 sq.ft. (929 m2) of office and break area.

The expansion quickly enabled more business as typical order sizes of four to six units were replaced with orders of eight to 15 units. In addition, there were more engine driven units, 69 and 79%, respectively, in 2010 and 2011, with average size growing to 1383 hp (1031 kW) in 2010 and 1530 hp (1141 kW) in 2011.

In 2012, the company built 92 units averaging 1881 hp (1403 kW). Even though a number of large Caterpil-lar G3616 engine driven packages were built, electric motor-driven fuel gas boosters and gas plant residue compressors for the western Mar-cellus and Utica Shale resulted in a mix of 58% engine and 42% electric drive. “We continued to see more

MARCH 2014 33 CoMpRessoRtech2

has skyrocketed since companies start-ed drilling in earnest in the Marcellus Shale,” Dearing Wall said.

The company received its first Mar-cellus Shale compressor order for 15 units in early 2009. That year, it built 64

units averaging 868 hp (647 kW), with 45% of them electric motor-driven.

Seeing demand growing rapidly, Dearing nearly tripled its plant space with a 60,000 sq.ft. (5574 m2) addi-tion that was completed in July 2010.

n This Ariel KBZ/6 reciprocating compressor package was installed at a Marcellus wet gas-processing facility near Washington, Pennsylvania, in 2010. The cryogenic gas process-ing residue compressor is driven by a 5000 hp (3729 kW), 900 rpm, 4160V GE VFD capable electric motor. continued on page 34

n This Waukesha P9390 and Ariel JGD/4 package was shipped in late 2013 to a Marcellus site in West Virgin-ia. The engine was removed to reduce the shipment weight and height, which is a common requirement for large packages going to remote Appalachian Basin destinations.

CT341.indd 2 2/25/14 2:37 PM

electric-driven units in 2013 for Utica Shale gas plants,” Rick Dearing said.

Dearing’s primary market covers western Pennsylvania, West Virginia, Ohio, western New York, Virginia, Kentucky and Tennessee.

“In 2012, we packaged units rang-ing from 40 to 5000 hp (30 to 3729 kW),” Dearing Wall said. “We don’t want to box ourselves out of any market. We want small units too. We build air skids, and we see grow-ing markets in vapor recovery units (VRUs) and compressed natural gas (CNG) compressors.”

With the acquisition of adjacent properties, the company now has 125,000 sq.ft. (11,613 m2) under roof on 16 acres (6.5 ha). Current capacity is 100 large units per year.

Seeing more demand, the company is adding 20,000 sq.ft. (1858 m2) in 2014 to expand the parts warehouse and package pipe welding areas, moving them from the current fabri-cation plant to free more packaging space there.

The facility has eight fabrication bays, although multiple small units can be built in a single bay. When the new

addition is complete, fabrication capac-ity will grow to 12 large bays with 180 tons (163 tonnes) of crane capacity. The company has 190 employees.

Other parts of the complex include an ASME coded pressure vessel shop, a hydrostatic test shop, a qual-ity department and inspection area, a paint line for prepainting many items prior to assembly on the packages and a skid fabrication area. Some skids are subcontracted.

Dearing offers packaged com-pressors for field gas gathering, gas processing plants, pipeline transmission, gas storage, VRU and CNG, including complete CNG sta-tions with the dispensers and fuel management systems provided by a partner company.

It also assembles air starting sys-tems, including complete buildings, air drilling primary and booster com-pressor packages, and blower and vacuum packages for environmental services. In addition, the company provides pump packages for well ser-vicing, salt water disposal and high pressure hydraulic applications. It de-livered its first sour gas units last year.

Dearing is a distributor for Ariel, LeROI and Gardner Denver compres-sors. The company has packaged other compressors when specified by the customer.

An engineering team of 14 to 16 provides custom designs for many applications. Packages are complete-ly designed on Autodesk Inventor 3-D CAD. The company’s factory-trained technicians provide assistance with installation, startup and training.

“We have based our reputation on service, reliability, integrity and in-novation, and we have responsibly served our industrial and energy cus-tomers with dependable equipment and systems for compressed air, gas, process gas and hydraulic applica-tions,” Rick Dearing said. “We match quality compressors, drivers and other equipment with engineering, de-sign expertise, installation experience and quality service.” CT2

MARCH 2014 34 CoMpRessoRtech2

n This modular building, shown nearly complete at the Dearing plant in late 2013, contains a complete air system for a com-pressor installation. Included is a Gardner Denver rotary screw compressor, drier, con-trols, and all HVAC equipment and environ-mental controls.

CT341.indd 3 2/25/14 2:42 PM

NEA UNDERSTANDS MOTOR DYNAMICS TO REDUCE DRIVE TRAIN VIBRATIONSMotor dynamics have a major impact on torsional vibrations especially on shaft torque and electric current pulsations. That’s why for NEA it is mandatory to include motor electro-magnetic characteristics in drive drain simulations. NEA compressors are made in Germany, and we have local engineering, packaging, service and parts to deliver reliable, vibration free solutions.

THINK GERMAN, ACT LOCAL.VIBRATIONS TROUBLING THE NATURAL GAS MARKET?

NEUMAN & ESSER USA, Inc.Located in Katy, Texaswww.neuman-esser.comContact me for Latin and North America:Swamy SubramanyamVice President of Technical Management and Procurementssubramanyam@neuman-esser.comDirect Phone: +1 713-554-9636

NEA GROUP Headquarters in Germany

BLUESTROKECOMPRESSOR SYSTEMS

NuemanEsser.indd 1 2/11/14 11:23 AM

Ed Hauptmann is director of engineering development for Lo-Rez Vibration Control Ltd., Vancouver, British Columbia, Canada. He has a Bachelor of Science in mechanical engineering from the University of Alberta, Edmonton, Canada and a PhD in ap-plied mechanics from the California institute of Technology. Hauptmann has 35 years of experience as a torsional vibration consultant. Contact him at: ed@lo-rez.com. Bill Eckert is princi-pal engineer for Beta Machinery Analysis Ltd., Calgary, Alberta, Canada. He has a Bachelor of Science and a PhD in mechanical engineering from the University of Alberta. Eckert has extensive experience in finite element modeling of flexible rotors. Con-tact him at: beckert@betamachinery.com. Brian Howes is chief engineer for Beta Machinery. He has Bachelor of Science and Master of Science degrees in mechanical engineering from the University of Calgary. Howes recognized the problem matching actual torsional natural frequencies with predicted TNFs. This difference could not be explained until Dr. Knop’s paper was examined. Contact him at: (bhowes@betamachinery.com)

Motor Dynamic Influence On Torsional Vibration Analysis >

Editor’s Note: This article is based on a paper given at GMRC’s Gas Machinery Conference in Albuquerque, New Mexico, Oct. 6-9, 2013. The authors wish to thank Andrew Mancini for his help in rerunning past analyses and Fabian Claussen for translating several references from German to English.

The steady-state torque and power output of a poly-phase induction motor are the result of electromag-netic fields that act across the air gap between stator

and rotor. If the rotor has a torsional vibration superimposed over the steady rotation, the same electromagnetic fields across the air gap can produce additional torques on the rotor, which act in the same way as a torsional spring or damper. These additional electromagnetic (em) effects are not usually included in standard torsional vibration analy-ses, as no analytical methods for estimating their magni-tudes have been available to date.

Such unsteady em effects have nevertheless been ex-tensively studied in the past by direct numerical integration of the differential equations representing stator and rotor currents and their mutually induced stator and rotor mag-

netic fields. Past studies have typically included start-up of drive systems, estimation of the resulting transient motor torques, and dynamic effects such as limit cycles and drive instabilities (“negative” damping).

As important as these past works are, the direct em effects on torsional vibration in a drive train have not been extensive-ly explored. The importance of em effects can be shown by first referring to simpler mass-elastic models. Figure 1 shows an idealized two-mass torsional model of a compressor and motor rotor, with the addition of an em “spring” and “damper” acting between the rotor and stator.

n Figure 1. This is an idealized two-mass model of a motor-compressor drive, with an additional spring and damper between the rotor and stator.

The additional spring and damper add a second vi-brational mode, so that the system now has two natural frequencies, one above and one below the single mode frequency, as shown in Figure 2.

Since coupling stiffness is often chosen to place the cou-pling mode frequency below the operating speed range (heavy line a in Figure 2), the additional em spring could well raise the higher natural frequency into the operat-ing range (the right-hand dashed lines b in Figure 2). As an example, Figure 3 shows that for a coupling with stiff-ness lower than that of the em spring, kM ≥ kC, and de-pending on the ratio of motor to compressor inertias,

TECHcornerA method of estimating electromagnetic damping mag-nitudes and their effect on torsional vibration response

By Ed HAuPTmAnn, BIll EckErT And BrIAn HowES

MARCH 2014 36 CoMpRessoRtech2

cT326.indd 1 2/24/14 9:40 Am

TECHcorner

MARCH 2014 37 CoMpRessoRtech2

JM / JC, the upper natural frequency could be as much as 1.5 times higher than normally estimated.

These unsteady em motor effects were described by Knop [1] in a 2012 EFRC presentation. The numerical ex-amples given by Knop (ratio of rotor to compressor inertias, JM / JC, less than 1.0; ratio of em stiffness to coupling stiffness, kM / kC, greater than 1.0) are found in many European compressor-motor drivelines, but are not as com-mon in North American installations.

For example, for a 1200 hp (900 kW), 1200 rpm motor driving a four-throw compressor through a steel-spring coupling, the ratio of motor to compressor inertia is approximately 1.85 and the ratio of em spring to cou-pling stiffness is approximately 0.8. Figure 3 shows that the (upper) nat-ural frequency in this case would be raised by about 6% above that without including the em effect.

Keep in mind that the above re-marks are based on a simplified two-mass model, with a constant value of em spring stiffness and no accounting of damping effects. In the following we present a way to develop estimates of em spring and damping magnitudes, and consider their effect on torsion-al vibration response for a range of typical North American motor- compressor installations.

Estimating electromagnetic spring and damping values

In the earlier reference cited, Knop [1], suggested that by linearization of the differential equations for the stator and rotor currents, analytical expressions

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n Figure 2. Shown here are the torsional natural frequencies of a two-mass model: curve (a), single frequency, and no em effect; curves (b) including em effects, with the higher natural frequency possibly raised into the system operating range.

n Figure 3. The effect of including an em spring in the two-mass mod-el (neglecting damping, and taking kM as constant) is demonstrated.

continued on page 38

CT326.indd 2 2/24/14 9:40 AM

MARCH 2014 38 CoMpRessoRtech2

could be generated for em spring and damping values. This would not allow study of certain dynamic aspects (limit cy-cles, “negative” damping), however practical estimations of torsional response in a wide range of driveline installations might still be made.

Knop [1] proposed that the em spring stiffness kM, and damping dM \ could be expressed as:

kM = (MSt /Ωs){(v2TL) / [1 + (vTL)2] } … Equation 1,dM = (MSt /Ωs){1 / [1 + (vTL)2] } … Equation 2.

where: v = frequency of the superimposed torsional vibration,

Ωs = electrical supply frequency, MSt = a motor circuit constant with dimensions of

torque, TL = a motor circuit electrical constant with units

of time.

No further information was given in Equation 1 or re-garding how the expressions MSt and TL could be evalu-ated, nor was further information made available to us.

To uncover the meaning and get estimates for the vari-ables MSt and TL in Equations 1 and 2, we have referred to the fundamental analysis and equations in Jordan et al [2, 3], and have also carried out a further linearization analy-sis of expressions for spring and damping effects used in their studies. This involves considerable algebraic manipu-lation not worthy of inclusion here, and interested parties should contact the lead author for details.

As a result, we have derived the following approximate equations for MSt and TL:

with: TR = rated (full load) motor torque, TB = breakdown motor torque, sR = slip at rated load,

then: TL ≈ (1/Ωs)[1/(2sR)]( TR / TB ) … Equation 3, andMSt /(ΩsTL) ≈ (# stator poles)(TB ) … Equation 4.

Using Equations 3 and 4 in Equations 1 and 2 will allow ready estimations to be made of the em effect on torsional vibration in a range of drive installations.

With the further substitution of:

x = (vTL), Equations 1 and (2) become,kM = (# stator poles)( TB )[ x2 / (1 + x2) ] … Equation 5, and dM = kM / (v2TL) = kM (TL) / (x)2 … Equation 6.

The main motor torque characteristics TB, TR and sR are either tabulated or readily available, but the time constant TL requires a further evaluation of Equation 3. Figure 4 shows estimates made using Equation 3 for two types of motors, with time constants TL ranging from 0.07 to 0.14 seconds for motors in the range of 670 to 2000 hp (500 to 1500 kW).

n Figure 4. This shows the estimated time constant TL for typical induction motors; data taken from published information [4], [5]. The dashed line is for reference only.

For torsional frequencies, v of interest from 10 to 60 Hz, the variable x2 = (vTL)2 in Equation 5 can vary from approximately 20 to 2000, but the quantity [x2/(1+x2)] is always within a few percent of 1.0 and in such cases can be neglected. However, for very soft couplings between large motor-compressor setups, the frequency effect on kM, shown in Equation 5, can be significant and should not be neglected.

Case studiesEquations 5 and 6 have been used to make estimates

of kM and dM in order to find how important em effects may have been in past torsional vibration studies based on conventional mass-elastic models (Figure 5). A represen-tative sample of these studies is shown in Table 1, which summarizes the results for three types of installations:

1. Drive A; high power, 5000 hp (3728 kW), soft (rubber)

coupling, (450 to 900 rpm),2. Drive B; mid-range power, 1250 hp (932 kW), steel

spring coupling, (1200 rpm),3. Drive C; low power, 450 hp (340 kW), stiff coupling,

mid-speed (720 rpm).

n Figure 5. This is the conventional mass-elastic model used for the torsional vibration case studies in Table 1, with the addition of an em spring and damper acting on the rotor.

CT326.indd 3 2/25/14 2:50 PM

MARCH 2014 39 CoMpRessoRtech2

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n Table 1. This is a summary of em effects on some previous tor-sional vibration studies. The effect is pronounced for higher power installations with ultrasoft rubber couplings.

Results of em effects on natural frequency (coupling mode) — Table 1 indicates that for systems like C with stiff couplings, that is, kM << kC, em effects are relatively unimport-ant. On the other hand for systems like A with soft couplings, kM >> kC, the coupling mode frequency can be as much as 90% higher, much as predicted by Knop [1]. The variable stiff-ness characteristics of rubber couplings are a particularly im-portant influence on the wide range of frequency estimates for this type of system. Intermediate systems like B are modestly affected. These results are also shown in Figure 6, where kM in the abscissa is calculated from Equation (5).

Table 2 shows a summary of the torsional natural fre-quencies compared to the orders of run speed. As previ-ously noted the largest shift in torsional natural frequency was for the Ariel JGC/6 with a very soft rubber coupling; the natural frequency was up to 1.86 times that without the em effects in the model. For this case, the torsional natu-ral frequency was shifted enough to result in resonance at 1 x shaft speed at the lower end of the run speed range.

Drive A Drive B Drive C

Compressor Ariel JGC/6

Ariel JGK/4

Cameron MH64

Speed Range (rpm) 450-900 1185-1192 716

Coupling Soft; rubber-in shear Steel-Spring

Stiff: Disc pack

Cold Nominal Warm

kC – kNm/rad 136 106 72 124 15,117

1st TNF, v0 – Hz 5.58 4.93 4.04 14.75 77.98

Motor Toshiba Reliance WEG

Rated Power (kW) 3728 900 336

Line Frequency (Hz) 60 60 60

Number Of Poles 8 6 10

Rated Torque, TR – Nm 40,230 7,210 4474

Rated Slip, sR - % 1.67 0.67 0.56

Breakdown Torque, TB – Nm 88,507 16,593 8948

Time Const., TL – sec. 0.036 0.086 0.118

xE = (vETL)* 3.40 3.40 3.40 10.8 8.88

xE2 / (1 + xE

2) 0.92 0.92 0.92 0.991 0.987

kM – kNm / rad 651.8 651.8 651.8 98.7 88.4

dM – kNms / rad 2.03 2.03 2.03 0.07 0.133

kM / kC 4.79 6.15 9.05 0.80 0.0058

v / v0 1.55 1.66 1.86 1.089 1.002

* Main excitation harmonic assumed to be 1 x run speed; vE = 94.25 rad/s for drive A.

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continued on page 40

CT326.indd 4 2/24/14 9:41 AM

n Figure 6. The representative em effects for the three different types of drives are described in Table 1. The inertia ratio lines are taken from the two-mass model and are shown for reference only.

Predicted Natural Frequencies (Hz)

Orders Of Run Speed At

450 rpm 900 rpm

Without em Stiffness

4.04 (Coupling Mode) 0.54 0.27

With em Stiffness

2.61 (Rigid Body Mode) 0.35 0.17

7.51 (Coupling Mode) 1.00 0.50

n Table 2. A summary of em effects on the predicted torsional natural frequencies of drive A (soft rubber coupling) is given. The addition of the em stiffness results in the coupling mode being resonant at 1 x run speed.

Compressor Vibra-tory Torque; lbf-in.

Coupling Vibra-tory Torque; lbf-in.

Coupling Heat Load (HP)

Motor Shaft Stress Design(Bagci) Factor

Design Operating Conditions

No em Effect 516,900 28,720 0.77 51.5

With em Effect 516,730 27,090 0.71 54.2

Upset Operating Conditions (Single Acting Cylinders)

No em Effect 512,440 77,500 3.19 11.3

With em Effect 513,300 125,980 4.67 15.3

n Table 3. This is a summary of how em effects alter the predicted response of drive A (soft rubber coupling). In this case, the ad-dition of the em stiffness and damping was insignificant for the design operating conditions, but important for upset conditions (particularly on the coupling vibratory torques and heat load).

Results of em effects on system stress and torque levels — For each of the previous design cases A, B, C, the predicted torques, stress design factors and coupling heat loads did not change to any significant degree dur-ing design operating conditions. This is because, although the system is now resonant with the coupling mode at run speed, there is very little 1 x torque demand from the com-pressor due to the staging and double acting loading for the

MARCH 2014 40 CoMpRessoRtech2

given design operating conditions. However, if single acting conditions are considered (possibly as a result of a cylin-der valve failure), the coupling vibratory torques and heat loads are predicted to be much higher when including em effects. In this case, the coupling heat load is approaching the manufacturer’s limit during the upset conditions.

Table 3 shows a comparison of the predicted torques, design factors and heat loads for drive A (soft rubber coupling). The added em stiffness does have the potential to cause a reso-nant condition with resulting failures of driveline components during upset operating conditions. At this point, the authors are not aware of any actual cases where that has occurred.

Comparison with field data

n Figure 7. This shows rpm versus time for a 1000 hp (746 kW) mo-tor driving a four-throw compressor through a steel-spring coupling, upper left, with em effects present; lower right, with the power off.

n Figure 8. This graph shows measured TNF during run-up (with em effects present), versus predicted TNF, including em effects. Proximity to the 45° dotted line indicates good agreement. Open markers are for predicted TNF without em effects.

Field measurements of torsional resonance are made con-veniently during start-up or shutdown sweeps of a system. In many cases, such sweeps produce significantly different re-sults: during run-up, em effects are present, while during run-down (with power off), they are not. This difference is graphi-cally illustrated in Figure 7, which shows rpm measured on a 1000 hp (746 kW) motor driving a four-throw compressor through a steel spring coupling. During run-up a TNF is indi-cated at 615 rpm, while during rundown the TNF is shown at approximately 500 rpm.

CT326.indd 5 2/24/14 9:42 AM

Hilliard.indd 1 8/1/13 3:35 PM

3. Enter these values on Figure 6 to find how much the natural frequency might be affected. This is a very conservative estimate; very soft couplings might have much greater frequency shifts and in those cases, calculate kM from Equation 5.

Further field measurement of the coupling mode natural frequency for fixed speed induction motors is needed to con-firm the validity of the approach used to develop Equations 5 and 6. Such field work requires advanced measurement and data processing techniques. The authors plan to carry out such work, which will be reported in a future paper. CT2

References[1] Knop, G.; “The importance of motor dynamics in recip-

rocating compressor drives,” EFRC, Düsseldorf, 2012.[2] Jordan, H.; Müller, J; Seinsch, H.O.; “About Electro-

magnetic and Mechanical Transient Processes with Three-Phase Drives” (translated from German); Wiss. Ber. AEG-Telefunken 53, 1979, 5.

[3] Jordan, H; Müller, J.; Seinsch, H.O.; “The Behavior of Three Phase Asynchronous Motors in Torsionally Elastic Drives” (translated. from German); Wiss. Ber. AEG-Telefunken 53, 1980, 3.

[4] Siemens, “Three-phase induction motors,” Catalog D 84.1, 2009.

[5] ABB HV induction motors, technical catalog for IEC motors EN, September 2011-3.

We have compiled a few field measurements made dur-ing run-up (with em effects) and compared them with pre-dictions made, including em springs and dampers, in a full torsional analysis of the system. Figure 8 shows that the predicted and measured values fall close to the dotted 45° line, indicating good agreement. The solid symbols show measured versus predicted results when em effects are included in the analysis. The omission of em effects is par-ticularly significant with very soft couplings.

ConclusionsThe electromagnetic stiffness and damping effects on

torsional vibration can be estimated by Equations 5 and 6. The predictions agree reasonably well with field measure-ments in a few cases known to us.

Systems with “very soft” couplings can be extremely sen-sitive to the em effect, particularly during upset conditions. Steel spring couplings may have their coupling mode fre-quency shifted upward by 8 to 10%, while “stiff” couplings are not significantly affected.

Reasonable preliminary estimates of the importance of em effects in a particular system can be made by taking the following steps:

1. Combine the inertias on either side of the coupling to form a simple two-mass model and note the inertia ratio.

2. Evaluate the abscissa in Figure 6, that is; # poles x breakdown torque/coupling stiffness.

CT326.indd 6 2/26/14 2:26 PM

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sized to measure flow and fuel con-sumption for nearly any reciprocating gas engine or small- to medium-sized gas turbines.

The venturi-based FM 50 provides near instant flow momentary mea-surements, as well as measurements either averaged or totalized over a period of hours, days or months, the company said.

The device provides a true mass flow calculation that corrects for tem-perature or pressure inconsistencies, resulting in an accuracy of 3% of read-ing or 0.5% of full scale, whichever is greater.

The measurement tool, which was specifically designed for single en-gines, works with engines and gas turbines from 100 to 10,000 hp (75 to 7457 kW). The basic design of the FM 50 remains the same, no matter the size engine or turbine it is installed

on, and the unit is sized to match the needs of a particular engine.

Features of the flowmeter include a pressure drop of 2 psi (0.14 bar), a response time under 10 ms, and a turndown ratio up to 20-1. CCC also calibrates the meter, which automati-cally makes adjustments to re-zero the transducers each time the meter is powered down.

The 8 in. (20.3 cm) device can be dropped into a 2 in. (5 cm) inlet pipe rather than having to re-pipe the in-let and find a way to fit a larger me-ter in around existing equipment, the company said.

The FM 50 comes with a heated, backlit display built into the body, al-lowing the device to operate in ex-treme subzero temperatures. Data can be stored locally on a USB por-table drive or can be communicated via CAN J-1939, Modbus, Ethernet, satellite or cellular modem, or through the included 2.6 GHz wireless radio.

In addition to environmental com-

pliance, the FM 50 is useful for ap-plications where a compressor is us-ing a bi-fuel gas substitution system. In such a case, the FM 50 provides measurements of the fuel flow and cost-saving estimates.

For a gas turbine with a low NOx

combustor, the FM 50 can measure the flow to each part of the combustor and provide that data to the turbine control system for determining efficiency.

The release of the CCC product follows proposed U.S. Environmental Protection Agency standards issued last September to reduce carbon di-oxide emissions from new and exist-ing power plants.

According to the proposed rule, new large natural gas-fired turbines would need to meet a limit of 1000 lb. (454 kg) of CO2 per MWh, while new small natural gas-fired turbines would need to meet a limit of 1100 lb. (499 kg) of CO2 per MWh.

New coal-fired units would need to meet a limit of 1100 lb. (499 kg) of CO2 per MWh, and operators would have the option to meet a tighter limit if they choose to average emissions over multiple years.

The agency is also developing stan-dards to address carbon emissions from existing power plants. They are due by June 1. CT2

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n The FM 50 flowmeter from Continental Controls Corp. is shown installed on a gas engine application.

MARCH 2014 42 CoMpRessoRtech2

n The FM 50 flowmeter is designed to measure flow and fuel consump-tion for gas engine and turbines.

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Norbert Feistel has a degree in mechanical engineering from the University of Karlsruhe and a doctorate from the University of Erlangen-Nuremberg. In 1988, he joined the R&D group of Burckhardt Compression in Winterthur. This paper was present-ed at a conference of the European Forum for Reciprocating Compressors on Sept. 27-28, 2012, in Düsseldorf, Germany.

Friction-Surface Coatings In Dry-Running Recips >

Due to increasing demands on the reliability of pro-cess gas compressors, measures to protect the counter-body surfaces of the sealing and rider rings

are becoming increasingly important.In this regard, the influences of different coatings on the

processes involving dry-running friction contact are not clear. Bench tests with dry-running sealing systems have revealed notable changes in surface texture, typically ac-companied by severe wear of the sealing elements after just 500 hours of testing in the case of some coatings, despite high hardness values significantly in excess of 1000 HV.

In addition to surface topography, the chemical resis-tances of the various coatings appear to be an important factor influencing the formation of transfer film. Especially with high-pressure loads, all coatings recommended for use in corrosive media achieved poorer wear rates compared to nitrided-steel piston rods.

IntroductionThe application of coatings to prevent damage during

compression of corrosive gases has proven successful over many years. Due to ever-increasing demands on the reli-ability of process gas compressors, coatings are now also being used increasingly in the absence of corrosive media to protect surfaces subjected to tribological loads.

API 618 [1] recommends wear-reducing coatings for pis-ton rods, regardless of the base material. A wear-resistant coating inside the cylinder might also become necessary in case of a high load exerted on the sealing system, a use of piston and rider rings with abrasive fillers, or a presence of abrasive contaminations in the gas.

Today’s complex set of requirements concerning the use of coatings in process gas compressors and involving, in particular, a diversity of gases, gas mixtures and impurities, are met through numerous combinations of coating materi-als and processes.

Especially in the case of dry-running reciprocating com-pressors, the interactions taking place between the friction pair and ambient medium must be considered, too. The coat-ing’s surface texture, of importance to stable dry-running conditions, also raises questions about optimal parameters and feasibility. To be taken into account furthermore is the large bandwidth of parameters influencing production of the desired coating quality.

Coatings requirementsA sound knowledge of materials as well as detailed infor-

mation regarding the load parameters involved are needed to determine the most appropriate combination of coating material and procedure for a given application.

High corrosion resistance and/or high wear resistance are required depending on the specific application. The maximum permissible layer thickness which already pre-cludes individual coating processes or at least necessitates use of a special intermediate layer is of importance depend-ing on whether a new coating or repair of a worn compo-nent is involved.

In addition to the piston rod’s outer coating, the diameter of the cylinder to be coated on the inside can range from less than 1.96 in. (50 mm) to more than 39.3 in. (1 m).

Finally, also to be considered in the case of oil-free com-pression is the suitability of the friction surface’s coating for dry-running operation with a broad spectrum of filled plastic materials.

Corrosion resistanceCoatings of the highest possible chemical resistance are

used to avoid damage to the metallic friction surfaces inside the cylinder and the piston rod during compression of cor-rosive gases such as chlorine, hydrogen chloride, hydrogen sulfide, etc. Especially ceramic coatings such as chromium (III) oxide (Cr2O3) have proven suitable for such applica-tions for many years.

In addition to ceramics, hard-metal coatings are being used increasingly for compression of corrosive media. In the most common variant of friction-surface coating comprising tungsten carbide in a cobalt matrix, however, the base cobalt constitutes a weakness in terms of corrosion resistance.

To mitigate this drawback, about one-third of the me-

TECHcorner

An assessment of the benefits and risksBy NOrBErT FEIsTEl

MARCH 2014 46 CoMpRessoRtech2

CT317.indd 1 2/21/14 5:45 PM

TECHcorner

MARCH 2014 47 CoMpRessoRtech2

tallic cobalt matrix is replaced with chromium. In reciprocating compres-sors, these WC/CoCr coatings quickly proved successful in many applica-tions involving corrosive media.

Another very corrosion-resistant variant of a hard-metal coating is chro-mium carbide in a matrix of nickel and chromium. Suitable especially for high-temperature applications, this wear-re-sistant layer has performed well in the chemical and aerospace industries, and is also used now as a friction-sur-face coating in compressors.

Ceramics and hard metals suitable for corrosion protection are usually applied through thermal coating. Due to the nature of this process, how-ever, thermal spray coatings always exhibit a porosity whose degree is in-fluenced, in particular, by the type of process employed.

The porosity ranges from 0.5% in the case of very dense layers to more than 2%. Through these pores and microcracks in the spray coating, cor-rosive media can penetrate as far as the substrate.

To prevent this kind of underlying corrosion, it is necessary to imple-ment additional measures such as surface sealing, thermal aftertreat-ment by means of self-fluxing alloy powders, or use of a dense, corro-sion-resistant intermediate layer be-tween the base material and coating. Intermediate layers comprising, for ex- ample, nickel-chromium-molybdenum alloys or chemically deposited nickel are employed as diffusion barriers.

Available finally as an alternative coating process is high-temperature CVD, in which repeated spraying and annealing of a ceramic suspension fol-lowed by sealing also makes it possible to create a layer almost free of pores.

Wear resistanceDry-running materials with abra-

sive fillers such as ceramics, glass or carbon fiber can already cause severe wear-related damage on counter- body surfaces made of cast iron. If abrasive gas contaminations such as aluminum oxide are involved ad-ditionally, the wear resistance limits

of even high quality nitrided steel are rapidly exceeded [7].

Even hard chrome coatings offer no protection here. The only remedy in this case is to use a coating with a significantly higher wear resistance. The highest possible degree of hard-ness, in particular, is usually favored during selection of this kind of wear-protection coating.

Thin layers deposited using the PVD or CVD process, and possess-ing a maximum thickness of only a few micrometers, theoretically offer a very good performance. The well-known titanium nitride layers are specified to have micro hardness values of over 2000 HV. The extremely hard diamond-like carbon (DLC) coatings have significantly higher micro hard-ness values of 4000 to 6000 HV.

However, high hardness values alone are no guarantee of adequate protection for the base material, for example, if requirements for bonding strength are not met so that the layer fails by flaking during operation.

In addition to ceramic coatings with a high hardness of about 2000 HV, some hard-metal coatings have prov-en themselves superbly during opera-tion with the aforementioned alumina particles present in the gas flow.

Though a hard metal comprising 88% WC and 12% Co has a hardness of just 1300 to 1400 HV, the hetero-geneous structure of this sintered composite of hard carbides in a soft metal matrix has proven to be wear-resistant even in the presence of very hard gas impurities.

Coating qualityRequirements for the quality of a

friction-surface coating are described by coating specifications, which stipu-late agreed limiting values for the lay-er’s composition, thickness, hardness, porosity, tensile adhesive strength, etc.

In fact, however, a coating’s prop-erties depend on a variety of further parameters. In the case of thermal spraying, for example, the coating’s quality is also influenced to a large ex-tent by the process, spray gun, fuel,

continued on page 48

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MARCH 2014 48 CoMpRessoRtech2

application rate, etc. In dependence on these boundary conditions, the coater selects a powder of the most appro-priate particle size.

Despite a constant powder composition, any change to these parameters can significantly affect coating quality, as Figure 1 already shows in terms of the differences between the structures of two WC/Co coatings applied using differ-ent processes.

n Figure 1. WC/Co coatings applied using a detonation process (bottom), and using high-velocity oxygen fuel spraying (HVOF) [9].

The coating applied using a detonation process has a much coarser structure, with carbide sizes ranging from 10 to 25 µm (Figure 1, bottom), whereas a WC/Co coating of a nearly identical composition applied by high-velocity oxygen fuel spraying (HVOF), exhibits a very fine structure with carbide sizes ranging from just 4 to 5 µm [9].

Process-specific variations in the coating quality are not always avoidable in practice. From a critical value on-ward, the internal coating of a cylinder of a small diameter can no longer be applied using the process, spray gun or powder particle size optimal for a piston rod’s outer coat-ing, for example.

To achieve reproducible coating qualities, it is therefore essential that all process parameters, the preparation of the substrate surface and the aftertreatment of the coating are

defined by means of a spraying instruction, in addition to defining a coating specification [2, 3].

Suitability for dry-runningPractice has shown that cast iron or nitrided steel as the

counter-body material in dry-running applications achieves a very good service life for the sealing and rider rings. At least the same expectations are placed on friction-surface coatings.

The requirement to produce a functional friction pair from various available dry-running materials is therefore a key criterion that significantly restricts the choice of avail-able coatings.

Key requirements for friction-surface coatings intended for dry-running can generally be described as follows:

• Resistance to local, high temperatures above 572°F (300°C)

• Highest possible thermal conductivity• Favorable properties for formation of a stable transfer film• Favorable influence on tribochemical processesThe thermal conductivity values of eligible friction-surface

coatings might differ notably from those of conventional me-tallic counter-body materials. For instance, carbon-based coatings exhibit significantly better values. WC/Co coatings also have a very good thermal conductivity. Ceramics range from materials with excellent thermal conductivity, such as silicon carbide, to insulators (e.g., zirconia).

An optimal roughness range instead of an extremely smooth surface has proven successful during dry-running [7]. Appro-priate precise surface finishing by means of grinding, lapping, honing etc., is required after the application of the coating.

At present, it is unclear whether proven roughness values for metallic counter-body materials can be implemented for specific coatings, and whether the resultant combinations have optimal effects.

Attention must also be drawn to the fact that tribochemi-cal processes take place during dry frictional contact, sig-nificantly influencing friction and wear, and can thus have a positive effect, for example, through formation of a protec-tive layer on the sealing elements’ running surface [6].

However, adequate empirical data regarding tribochemi-cal interaction with a variety of combinations of filled plastic materials and gases are not available for all coatings. Es-pecially in the case of ceramics with a low chemical activity, as well as coatings with pronounced anti-adhesive proper-ties, such as hard chrome and composite materials made of nickel-phosphorus with embedded PTFE particles, sig-nificant differences to conventional metallic counter-body materials can be expected.

Particularly with regard to coatings containing PTFE, it is necessary to note that they are not identical to the tribo-chemically modified transfer film formed during operation, and therefore do not advance the process of transfer film formation, but instead can affect it negatively.

Friction-surface coatings in practiceThe aforementioned requirements for friction-surface coat-

CT317.indd 3 2/21/14 5:48 PM

MARCH 2014 49 CoMpRessoRtech2

ings in dry-running reciprocating com-pressors are now met mainly by hard-metal and ceramic layers applied by a thermal spray process. This method offers the greatest variation in terms of coating materials as well as layer prop-erties and thicknesses, thereby hold-ing a lot of potential, especially for the repair of worn friction surfaces.

Hard metals are handled differently than ceramics, however. Thus, hard metals are now favored increasingly as piston-rod coatings. For instance, their lower impact sensitivity compared to ceramics has a positive effect highly valued during everyday handling.

In addition, hard-metal powders ap-plied using the HVOF technique can be used to produce very dense, low-porosity layers. Spraying of ceramic layers requires higher temperatures, achievable by plasma processes. How- ever, smaller spray-gun dimensions here also make it possible to realize internal coatings of cylinders with a diameter of slightly less than 1.96 in. (50 mm).

Optimized multicomponent sys-tems are also continually broadening the scope of applications for ceramic coatings. An alternative process to thermal coating of cylinders with small diameters is the high temperature CVD method.

Despite the additional costs, the application of friction-surface coat-ings in reciprocating compressors has

increased significantly. For example, coated piston rods are increasingly replacing nitrided piston rods, which were still widespread until a few years ago, especially in dry-running applica-tions involving low loads. When invest-ing in such surface finishing, however, customers expect added value in the form of notably longer lifespans of the coated components.

Operational experience with coat-ings has shown that demands placed on them, especially in terms of in-creased wear protection compared to uncoated friction surfaces, are usually fulfilled in a most excellent manner. However, practical experiences with the performance of dry-running pack-ings in practice have been mixed. In applications involving pressure dif-ferences in excess of about 290 psi (20 bar), packings occasionally had a very short service life, the piston rod exhibiting conspicuous, very smooth surface areas.

Figure 2 shows the average wear rate of all sealing elements forming part of dry-running packings of a hy-drogen compressor with a pressure difference of 783 psi (54 bar).

Evidently, the wear rate deterio-rates continuously despite constant conditions for assembly and operation of the packings, until premature fail-ure of the fourth sealing system. Each packing was operated on the same

n Figure 2. Average wear rates of four sets of packing rings for sealing a piston rod coated with hard metal.

continued on page 50

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MARCH 2014 50 CoMpRessoRtech2

rod with a hard metal coating, which was in excellent con-dition except for a low roughness along the packing rings’ friction path.

InteractionsInsights into the phenomena observed on coated piston

rods in practice were to be provided by bench tests on a dry-running compressor.

For this purpose, the compressor’s 216 in. (550 mm) long tail rod with a diameter of 1.96 in. (50 mm) was coated with the hard metal WC/Co using both the detonation as well as the HVOF process. For comparison with the two chromium-free coatings, the hard metal WC/CoCr containing chromi-um was applied to another piston rod, also using the HVOF process. Finally, a quaternary ceramic consisting of alumi-num oxide, chromium oxide, zirconium oxide and titanium oxide was included in the test series.

A piston rod made of nitrided steel served as a reference for a metallic counter-body surface. An arithmetical mean deviation of the profile Ra in the range from 0.20 to 0.30 µm was specified for all surfaces (Table 1).

A suction pressure of 232 psig (16 barg), final pressure of 580 psig (40 barg) and average piston velocity of 10.4 fps (3.18 m/s) were specified as the load parameters for the tests conducted with dry nitrogen. Used in each case was only one sealing element comprising PTFE/PPS polymer blend in a cooled packing, to ensure clearly defined condi-tions during subjection of the piston rod to friction power.

The PTFE/PPS polymer blend selected for the packing rings and filled with graphite and carbon fiber [7] has proven to be nonabrasive in many applications, even in combina-tion with cast iron of a low hardness.

Counter-Body Material

Nitrided Steel

WC/Co WC/CoWC/CoCr

Ceramic

Coating Process - HVOF Detonation HVOF Plasma

Hardness HV0.3 731 1128 1156 1279 926

Ra [µm] 0.237 0.263 0.223 0.277 0.280

Rz (DIN) [µm] 1.797 2.093 1.793 1.957 2.487

Rk [µm] 0.773 0.840 0.667 0.933 0.303

Ratio Rk / Rz 0.430 0.401 0.372 0.477 0.122

Rmr [%] 31.29 46.86 53.47 32.15 82.12

n Table 1. Hardness and roughness values of the tested counter-body materials.

Each test was to have a duration of 500 hours, though this was not possible with the WC/CoCr-coated piston rod. In this case, test had to be interrupted due to a sudden rise in the temperature of the sealing element’s chamber (Figure 3).

The removed sealing element had been thermally dam-aged by the excess temperature, and therefore had to be replaced with a new one.

On the piston rod, the roughness in the region of the seal-ing element’s friction surface had declined significantly; this

is clearly visible in comparison with the surface in its new state (Figure 4).

There was another temperature rise later on, and the test eventually had to be cancelled after 424 hours due to high leakage. The piston rod’s friction surface was polished smooth and glossy (Figure 4, bottom).

Such steep rises in temperature to values above 212°F (100°C) were not observed in the other tests. Here, the chamber temperatures remained stable in the range be-tween 140° to 158°F (60° to 70°C) after running in. Figure 3 therefore only shows the temperature profile of the piston

n Figure 3. Temperatures measured in the sealing element’s cham-ber, for piston rods coated with WC/CoCr and WC/Co using the HVOF process.

n Figure 4. Friction surface of a piston rod coated with WC/CoCr in the new condition (top), after an operating period of 270 hours (center) and after 424 hours (edge length 300 x 225 µm).

CT317.indd 5 2/21/14 5:50 PM

MARCH 2014 51

rod coated with WC/Co (HVOF pro-cess) for the purpose of comparison.

Especially during the running-in peri-od of a dry-running friction pair, surface topography is altered through breakage of roughness peaks, deposition of wear particles and reorganization of surfac-es, even on extremely hard counter- body materials. Because these ef-fects can vary highly along the friction surface, roughness measurements in this region usually provide results that fluctuate greatly. Consequently, zones exhibiting minimal roughness were sought and evaluated on the surfaces of all piston rods.

Comparing the thus obtained ar-ithmetical mean deviation of the pro-file Ra to the values in the original state reveals a reduction in the case of all surfaces, except the ceramic coating (Figure 5).

On this coating, formation of the transfer film even led to a slight in-crease in the Ra value. Especially evi-dent is the reduction in roughness on the WC/CoCr-coated piston rod, al-though it turned out to have the high-est hardness values (Table 1). The two chromium-free, hard-metal coat-ings and even the nitrided steel exhibit a much smaller drop in Ra values.

A further result shown in Figure 6 comprises the wear rates of the seal-ing elements operated on the various surfaces under otherwise identical

conditions. The lowest value was ob-tained in the test of the ceramic-coated piston rod. Very good results were also obtained with the two chromium- free, hard-metal variants.

The wear rates of the sealing ele-ments are also significantly lower than that of the nitrided steel surface. The highest wear rate was established in the test of the WC/CoCr-coated pis-ton rod involving high chamber tem-peratures at times, and revealing the greatest drop in the Ra value.

Influencing parametersAlthough a functional design of

counter-body surfaces is of par-ticular importance for the dry-run-ning friction pair comprising soft and hard, the surface of the hard counter-body material is usually de-scribed only by specifying a permis-sible range for the Ra value.

Ignored in this case is the known fact that the arithmetical mean de-viation of the profile Ra is totally un-suitable for characterizing surfaces, as it does not permit any distinction between different profiles. Peaks and valleys are treated equally during de-termination of the Ra value.

The arithmetical mean deviation of the profile is therefore only suitable as a quality control parameter if the character of the roughness is known,

n Figure 5. Ra values measured on the fric-tion surfaces of dif-ferent counter-body materials, compared with the original finish in each case.

continued on page 52

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i.e., whether a fissured or plateau-like surface is involved, for example [8].

Several surface texture parameters permit a character-ization of surfaces and therefore serve to supplement the results of an Ra measurement. This is achievable, for ex-ample, by means of the following parameters:

• Ratio of core roughness depth Rk to average roughness depth Rz (DIN)

• Ratio of average peak-to-valley height Rp to average roughness depth Rz (DIN)

• Relative material ratio of the profile RmrEspecially in the case of the porous ceramic coatings ap-

plied using the plasma process, the surface after comple-

tion of processing might be characterized by a plateau-like roughness profile comprising a high proportion of valleys attributable to the pores (Figure 7).

However, the upper limit of the permissible Ra value is of-ten exceeded by this kind of surface. While an inadmissibly large Ra value of a profile with a pronounced proportion of peaks leads to high abrasive wear on the soft dry-running sealing element, it is not necessarily a drawback in the case of a plateau-like roughness profile.

A quick initial conclusion about the nature of the rough-ness profile can be drawn from the ratio between the core roughness depth Rk and the average roughness depth Rz (DIN) [8]. In practice, surfaces achieving a value of less than or equal to 0.25 in terms of this ratio have also proven suc-cessful in dry-running if their Ra value exceeds the permis-sible upper limit (Table 1).

Conclusions about the characteristics of the various pis-ton rod surfaces can also be drawn from the material ratio of the profile Rmr indicated in Table 1. In each case, the ma-terial ratio was determined with a reference line c0 shifted by about 2% with respect to the highest peak at an profile section level of c1 = Rz/4.

A comparison between the Rmr values of the three hard-metal layers reveals that the material ratio of the WC/CoCr-coated piston rod is significantly lower than that of the two chromium-free variants, and only slightly higher than the

MARCH 2014 52 CoMpRessoRtech2

n Figure 6. Wear rates of a single packing ring in dry nitrogen on various counter-body materials.

n Figure 7. Porous surface of a ceramic applied using a plas-ma process (bottom) compared to a nitrid-ed steel surface (edge length 300 x 225 µm).

n Figure 8. Different surfaces after processing of a WC/Co coating and a WC/CoCr coating (bottom) [9].

CT317.indd 7 2/21/14 5:51 PM

BG Services Island.indd 1 1/31/14 8:25 AM

minimum value of 31.3% ascertained for the nitrided-steel piston rod. The highest value of 82.1% was obtained with the ceramic coating.

Hardness as average from 10 mea-surements, roughness parameters as average from 6 measurements by pro-file method [4, 5].

A comparison of the surface to-pographies of both layers applied by means of the HVOF process reveals that the piston rod coated with WC/CoCr has a coarser structure (Fig-ure 8, bottom). Powder with a parti-cle size of -45 +15 µm was employed for both hard-metal variants; as a re-sult, the carbides present in the WC/CoCr coating and sized about 3 µm were slightly smaller compared to the sizes of 4 to 5 µm in the chromium-free coating.

The differences in surface topogra-phies observed after processing can therefore hardly be explained by their structure, but rather by the different composition of the matrix material and the resultant material properties.

Whereas the cobalt content of the chromium-free variant is 12% by weight, it drops to 10% by weight in the WC/CoCr-powder, and is supple-mented there by a chromium con-tent of 4% by weight. The chromium content has a negative effect on the composite’s toughness, apparently making it difficult to produce a sur-face profile with a high material ratio, and resulting in the observed coarse surface topography [9].

An analysis of the test results ob-tained with the ceramic coating and the two chromium-free WC/Co coat-ings shows that plateau-like surfaces with the greatest possible material ratio offer favorable conditions for dry-running. Plateau-like structures can be realized, for example, by levelling the roughened profile’s peaks using a suitable superfinishing process, there-by creating a surface with a dominant proportion of valleys.

Influence of coatings on the performance of dry-running sealing systems

The single-ring test results con-

cerning the influence of various coatings’ surface texture on the per-formance of dry-running sealing sys-

tems were studied in more detail in a hydrogen compressor.

MARCH 2014 53 CoMpRessoRtech2

Counter-Body MaterialNitrided Steel

WC/Co WC/CoCr CrC/NiCr

Coating Process - HVOF HVOF HVOF

Rmr, original finish [%] 33.89 64.36 40.60 43.58

Ra, original finish [µm] 0.220 0.227 0.170 0.237

Ra, friction surface [µm] 0.157 0.140 0.102 0.077

n Table 2. Roughness values of the counter-body materials tested in dry hydrogen.

continued on page 54

CT317.indd 8 2/21/14 5:51 PM

For this, the piston rods possessing a diameter of 1.96 in. (50 mm) were coated with both hard-metal variants with and without a chromium content using the HVOF process, and their surfaces furnished with highest possible material ratio Rmr in an additional superfinishing process.

Table 2 shows that this was achieved more successfully with the chromium-free coating. To counter the conspicu-ously high roughness loss on the WC/CoCr-coated piston rod, slightly lower Ra values in the range from 0.15 to 0.25 µm were selected for all surfaces. A conventionally finished piston rod made of nitrided steel again served as a refer-ence for a metallic surface.

A suction pressure of 203 psig (14 barg), final pressure of 580 psig (40 barg) and average piston speed of 1.5 fps (3.41 m/s) were specified for the tests, conducted in the first stage of a horizontal compressor that was operated with dry hydrogen with dew point of about -85°F (-65°C).

The piston rod sealing system in each case comprised a cooled packing of six packing rings made of a PTFE/PPS polymer blend. An operating period of 500 hours was planned for each test.

In this series of tests, too, the WC/CoCr-coated piston rod’s behavior was conspicuous. Though the test in this case did not pose any problems, the average wear rate of the six packing rings was more than twice that of the chromium-free variant, and higher by about one-third even compared to the result obtained for the nitrided-steel piston rod (Figure 9), despite the higher material ratio Rmr.

Accordingly, laboratory analyses were conducted to reveal whether friction and wear are influenced not only by different surface topographies, but also the different compositions of the two hard-metal variants via tribo-chemical interactions between the friction pair and the ambient medium [9].

n Figure 9. Average wear rates of packings in dry hydrogen on various counter-body materials.

However, investigations carried out using energy disper-sive X-ray analysis (EDX) and micro-probe analysis (WDX) revealed no differences, such as a depletion of oxygen or chromium, between the coatings’ used and original areas. Nor was there any change in the layer structure near the surface. The test series was nonetheless supplemented by a further hard-metal coating that, like WC/CoCr, is highly resistant to corrosion.

Though the average wear rate of the sealing elements operated on this CrC/NiCr-coated piston rod is somewhat lower than the result obtained for the WC/CoCr coating, it is still significantly higher compared with the values obtained for the chromium-free hard-metal variant and the nitrided-steel piston rod.

In addition, the friction surface of the CrC/NiCr-coated piston rod exhibited a significant decrease in Ra value from 0.237 to just 0.077 µm (Table 2). This loss of surface roughness was accompanied by formation of a striking, silver-colored layer on the sealing elements’ friction sur-faces (Figure 10).

n Figure 10. Silver-colored layer on the friction surface of a dry- running packing ring after operation on a CrC/NiCr-coated piston rod.

The quaternary ceramic that successfully passed the single-ring tests also has an excellent corrosion resis-tance. Due to the very good results obtained at a final pressure of 40 barg, this coating was to subsequently prove its dry-running suitability under a significantly higher load comprising a suction pressure of 580 psig (40 barg) and a final pressure of 1450 psig (100 barg) in the hy-drogen compressor’s second stage. The sealing system employed here comprised a cooled packing with a total of 10 packing rings made of a PTFE/PPS polymer blend and optimized for this application.

Already during the test, unfavorable operating character-istics were indicated by conspicuously high temperatures on the piston rod’s surface and at the outlet for leakage gas and cooling water.

The average wear rate of the sealing elements oper-ated on the ceramic coating is more than twice the value obtained for a conventional piston rod made of nitrided steel (Figure 11). This poor result was confirmed in a re-peated test.

MARCH 2014 54 CoMpRessoRtech2

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At variance to the compact structure of the transfer film observed in all other tests, both tests here revealed a large amount of loose, powdery particles on the ceramic coat-ing’s surface.

Summary and conclusionsDespite high hardness values significantly in excess of

1000 HV, the arithmetical mean deviation of the profile Ra of some hard-metal coatings dropped by more than 0.1 µm in the region of the sealing element’s friction surface after only 500 hours. This change in surface texture was typi-cally accompanied by high wear rates on the packing rings, thereby confirming the observations made in practice.

Differences in roughness loss indicate an influencing role played by the composition and structure of the hard-metal composite. For example, hard-metal coatings with a pure cobalt matrix exhibited a considerably smaller de-cline in roughness than those with a matrix material con-taining chromium.

Another important factor influencing roughness loss is the counter-body surface’s topography. The tests showed that plateau-like surfaces with the highest possible material ratio Rmr offer favorable conditions for dry-running, besides reducing roughness losses.

n Figure 11. Average wear rates of two packings operated on a ceramic coating in comparison to a nitrided-steel piston rod.

Coating composition turns out to play an influential role here too, a generation of such surfaces with the highest possible material ratio Rmr being more successful in the case of WC/Co coatings than in the case of the two hard-metal variants WC/CoCr and CrC/NiCr.

These WC/Co coatings possessing a high material ratio

MARCH 2014 55 CoMpRessoRtech2

continued on page 56

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Rmr also exhibited lower wear rates compared with the tests of a conventional piston rod made of nitrided steel. A quater-nary ceramic tested as an alternative to hard-metal coatings, and possessing a significantly higher material ratio Rmr yet due to its porous surface topography, achieved the best test results with a final pressure of 580 psig (40 barg). After the load pressure was increased to 1450 psig (100 barg), howev-er, the ascertained wear rates were reproducibly worse com-pared to the result obtained for the nitrided-steel piston rod.

Large amounts of loose particles on the counter-body surface suggest that the corrosion-resistant ceramic coat-ing adversely affects a formation of transfer film. In ad-dition to surface topography, variations in chemical resis-tance among the coatings therefore also appear to be an important factor influencing wear characteristics. Espe-cially with high-pressure loads, all coatings recommended for use in corrosive media achieved poorer wear rates than a nitrided-steel piston rod, despite higher values for the material ratio Rmr. CT2

NotationPTFE PolytetrafluorethylenePPS PolyphenylensulfideCVD Chemical vapor depositionPVD Physical vapor depositionHVOF High velocity oxygen fuelDLC Diamond-like carbonRa Arithmetical mean deviation of the profile (AA, CLA)Rz (DIN) Average roughness depthRk Core roughness depthRp Average peak-to-valley height (Rpm)Rmr Relative material ratio of the profilec0 Reference levelc1 Profile section level

References[1] American Petroleum Institute, “Reciprocating Com-

pressors for Petroleum, Chemical, and Gas Industry Services,” API Standard 618, Fifth Edition; Washington, D.C. 20005, December 2007.

[2] DIN EN 15311 Thermisches Spritzen –

Bauteile mit thermisch gespritzten Schichten – Technische Lieferbedingungen Beuth-Verlag GmbH, Berlin, Juni 2007[3] DIN EN 15648 Thermisches Spritzen –

Bauteilbezogene Verfahrensprüfung Beuth-Verlag GmbH, Berlin, April 2009[4] DIN EN ISO 3274 Geometrische Produktspezifikation –

Oberflächenbeschaffenheit: Tastschnitt-verfahren –

Nenneigenschaften von Tastschnittgeräten Beuth-Verlag GmbH, Berlin, April 1998[5] DIN EN ISO 4287 Geometrische Produktspezifikation –

Oberflächenbeschaffenheit: Tastschnitt-verfahren – Benennungen, Definitionen und Kenngrößen der Oberflächen-beschaffenheit Beuth-Verlag GmbH, Berlin, Juli 2010

[6] Tomschi, U.: Verschleißverhalten von Trockenlaufwerkstoffen für

Abdichtelemente in Kolben-kompressoren Dissertation Universität Erlangen-Nürnberg, 1995[7] “Dry-running sealing systems in practice —

new challenges by new materials,” Feistel, N.; fith EFRC-Conference, Prague, Czech Republic 2007, p. 51-62

[8] Volk, R.: Rauheitsmessung Beuth Verlag GmbH, Berlin Wien Zürich, 2005[9] Kühnert, B.: Schichtbeurteilung von Kolbenstangen

MARCH 2014 56 CoMpRessoRtech2

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Mantosh Bhattacharya is a rotating equipment technical special-ist for Petrofac in the United Arab Emirates. He has 22 years of experience with rotating equipment.

Solving Compressor Impeller Rub Problems During Mechanical Run Tests >

To validate the rotor dynamic design of centrifugal compressors, mechanical run tests generally are conducted in OEM test facilities. In a nutshell, the

test validates the critical speed calculations and amplitudes in steady state (four hours at maximum continuous speed) and transient conditions (while passing through critical speed). The agenda generally follows clause 4.3.6 of chap-ter 1 and clause 4.3.1 of chapter 2 in API 617, 7th Edition.

The intent of this test is to ascertain rotor stability with-out it being influenced by process gas. Some OEMs follow this procedure using a partial vacuum as a medium. Some end users insist that a partial vacuum be used to ensure that fluid medium does not influence the results.

The common understanding was that mechanical run tests are to be conducted in partial vacuum by a leading OEM and some of the more capable end users. A recent issue changed this philosophy regarding running of a mechanical run test. This article investigates the issue in simple form and provides a means to mitigate the problem with technical backup.

The machine in question is a single-shaft multistage centrif-ugal compressor handling hydrocarbon gas in a back-to-back design. It has three impellers at each section with a suction impeller of high-quality, super-duplex alloy material. The gas handled is a hydrocarbon gas with molecular weight of 18, compressor speed is 10,800 rpm, impeller diameters are 21.2 in. (540 mm) for stages 1, 2 and 3 and 24.8 in. (630 mm) for stages 4, 5 and 6. Interstage labyrinths are stainless steel be-cause of the presence of mercury and CO2 in the process gas.

First incidentThe subject was undergoing a four-hour mechanical run

test as per end user approved and API 617-based proce-dure. During this test, when approaching maximum con-tinuous speed, vibration amplitude shot up from 7 to 40 µm at NDE side. During rundown, it even reached 80 µm pk-pk (higher than the trip limit set).

It was decided to dismantle the bundle as telltale evi-dence of rub was shown in the form of a phase-amplitude

plot taken during that period. The horn shape of the peak at critical speed suggests a possible rub as shown in Figure 1.

n Figure 1. This is the phase-amplitude plot during rub.

n Figure 2. Damage occurred in the impeller eye labyrinth area.

Using a fishbone analysis tool after this failure, it was shown that rotor-stator clearance was correct as per API 617 requirements. Alarm and trip settings were correct to prevent incipient rub. These included setup, the chemi-cal and mechanical properties of the impellers, post-weld heat treatment procedures of impellers and measured rotor and gas path radial and axial clearances. All were found to be correct.

The OEM undertook the replacement of the subject im-peller and in parallel increased the clearance between im-peller eye and labyrinth by 100 µm to avoid rub. The calcu-lated efficiency was now just at the guaranteed point.

Second IncidentAfter the replacement and subsequent modification, an-

other mechanical run test was conducted on rotor No. 2.

TECHcornerHeat dissipation improved, impeller expansion limited by shifting molecular weight of compressed medium

By MANTOSH BHATTACHAryA

March 2014 58 coMpressortech2

continued on page 60

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Similar rub was observed at maximum continuous operat-ing speed. The test was aborted. Upon dismantling the ro-tor, similar rub with a lower severity was observed.

n Figure 3. Graph shows rotor mode shape X — NDE side, Y — DE side of compressor rotor.

An analysis of failures with observations that were prob-ably overlooked by the OEM:

1. It was observed in both cases of damage; most se-vere was the first stage, 19.7 in. (500 mm) diameter impeller. This part is lightly loaded with respect to the drive end part. (It has three 630 mm diameter impel-lers, balance drum and coupling half.) During the wit-nessed mechanical test, the DE vibrations were higher than NDE side due to high centrifugal forces of com-paratively high diameter impellers (540 mm/630 mm). In spite of comparatively high vibration at the DE side, rotor rub occurred in the NDE side only.

2. The ratio of clearance shown in the lateral critical speed analysis did not suggest any probability of ro-tor rub at the first critical speed, which was confirmed in test run-up mode and orbit plots.

3. If the second critical speed was getting excited due to some unknown reason, the second stage suction impeller should have rubbed before the first stage suction impeller based on mode shape and their com-parative amplifications, as shown in Figure 3.

In reality, operating mode shape is a linear combination/ interpolation of the first and second critical speed mode shapes.

4. The geometrical differences between the first stage impeller and second stage suction impeller were:

First stage impeller — 540 mm diameter, 0.065 flow coefficient.

Second stage impeller — 630 mm diameter, 0.025 flow coefficient.

The high flow coefficient clearly indicated that eye open-ing diameter was bigger in the first suction impeller than the second stage suction impeller. Diametrical clearance between impeller neck and labyrinth were 0.027 in. (700 µm) for all impellers. RCV analysis and vibration amplifica-tion factor showed that even at trip level, the rotor cannot rub on stationary close clearance parts. Only an unex-plained local heating and thermal expansion could have lessened the clearance and initiated rubbing.

The bearing pad temperatures record did not show any local heating (Morton Effect). The attention shifted to close clearance labyrinth seals as the interstage HC seal was found new and intact.

The OEM procedure was to use a partial vacuum me-dia and lubricant media used for dry-gas seals was air. It is known that 70% of seal air goes to the process stage based on differential pressure. Hence, air being at higher pressure than partial vacuum, it could easily ingress into the clear-ance between labyrinth and impeller eye and cause a wind-age effect. Local windage can cause high local heating. This leads to thermal expansion of the impeller hub and laby-rinth. Consequently, they tend to come closer. In a vacuum, heat dissipation is much less by convection. Air, too, is a poor conductor of heat.

n Figure 4. This shows the test medium and seal-gas path with windage location.

This observation led to further calculation. Super duplex

alloy material has a mean coefficient of thermal expansion a of 13.1 x 10-6 mm/°C. Both stages of super duplex alloy material impellers have the same diametrical clearances, 0.027 in. (700 µm).

For example, the initial outside diameter of the impeller eye is X mm = 0.00X m if measured in ambient tempera-ture of A in °C. During the test, if the local temperature (at impeller area) goes up to B, in °C due to windage and improper heat dissipation.

The final diameter after circumferential expansion may be calculated as:

d1 = d0 (dt a + 1) = 0.X [(B - A))x 13.1) + 1] = Z mmWhere d1 is the final diameter after thermal expansion,

d0 is the original diameter and dt is the temperature differ-ential. Expansion was substantially higher for the first-stage impeller (higher flow coefficient) and was higher than that of the second-stage suction impeller (lower flow coefficient).

So, if the shaft displacement reading at the probes is M microns, which is far less than the API 617 criteria of 25 µm, still the impeller eye can rub with the labyrinth if the

CT315.indd 2 2/24/14 9:45 AM

March 2014 61 coMpressortech2

shaft vibration amplitude plus thermal expansion is greater than “as built” clearance.

To confirm that local heating and subsequent thermal ex-pansion has an undesirable effect, the OEM conducted an FEA of the first stage suction impeller using the mechanical properties measured in an experimental test. This was done to evaluate the possibility of radial deformation at the impel-ler eye during full speed and over speed in a bunker. Mea-sured values were a very close match with FEA calculation results, which suggested that bore/eye radial deformation does not occur that might have led to impeller/labyrinth rub.

Solution/corrective actionBased on the previous, it was evident that localized heat gen-

erated must be removed. This was to be done by using a me-dium which does not influence the rotor behavior by damping and the intent of the mechanical run test is not compromised.

Hence a gas mixture of molecular weight in MRT — 8.805 (80% He + 20% N2) with suction pressure of 14.5 psia (1 bara) and discharge pressure of 23.2 psia (1.6 bara) was se-lected. There was almost no change in compressor loading from the previous test. The gas mixture entrapped between labyrinth and impeller eye had a higher Nusselt number than air so heat generated due to windage dissipates faster.

The Nusselt number, Nu, is the dimensionless param-eter for fluids characterizing convective heat transfer. It is defined as:

Nu = a. L / l.Where a is convective heat transfer coefficient, L is rep-

resentative dimension of a conduit/pipe, and l is the ther-mal conductivity of the fluid. Nusselt number is a measure of the ratio between heat transfer by convection (a) and heat transfer by conduction alone (l/L).

Regarding the rotor dynamic changes due to use of this gas mixture it was concluded that:

1. There will be no shift in rotor critical speeds (minimum to maximum bearing stiffness).

2. Force coefficients are a function of shaft diameter at labyrinth, width of labyrinth, clearance of labyrinth and kinematic viscosity and density of gas. The gas mixture, MW, suction pressure and flow is so low that it will not affect the rotor behavior. The rest of the parameters are based on rotor construction and fixtures.

3. Alford Forces are a function of power, density ratio, impeller diameter, minimum width of gas passage in bundle and rpm. As mentioned in point 2, the power and density ratio is in the very lower side. Hence it will not affect the rotor by aero cross coupling forces.

Based on the above understanding, an internal test with 80% He and 20% N2 with molecular weight of 8.8 was con-ducted. It was successful in all aspects. Later, a complete set of four compressors were tested with the same modi-fied procedure and tests were successful in all aspects to the satisfaction of the OEM and end user. CT2

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Real-time modeling of perfor-mance and safety allows the PLC to predict healthy unit per-

formance and to identify where it is safe to run that unit ahead of time. Real-time monitoring reviews unit performance based on actual unit operations (healthy or damaged) but generally provides no insight of what actions as safe to take — other than announcing alarms and shutdowns. The first is required action; the second is desired.

There are three general methods of modeling reciprocating compressor performance and safety prediction com-monly used in unit control panels: Fast and Easy, Covers 90% of Concerns, and Covers Everything according to the Compressor Manufacturer (OEM).

Compressor station supervisors need to know which method is imple-mented in each control panel (UCP), as anything inferior to the OEM meth-od could lead to compressor damage.

Many compressor stations have UCPs based on the Fast and Easy method. This method is often imple-mented by System Integrators (SI) whose programmers curvefit a handful of calculated points, and then apply a few generic rules (such as maximum pressure differentials and maximum compression ratios) to keep units running and “safe.” This simple method typically does not pre- estimate interstage pressures be-fore load steps are changed, or look ahead to determine if changing

unit speed creates pin reversal issues.For simple, single-stage, double-

acting units that operate over narrow ranges without having too much cylin-der clearance applied, this approach is not completely without some merit. It is commonly implemented in UCPs for many slow-speed, transmission and gas-gathering applications.

The real problems begin when this Fast and Easy method is applied to more generic units and/or wider op-erating ranges with the belief that this simple method is actually safe to use for all applications. It’s not.

The Covers 90% of Concerns meth-od involves more calculations regard-ing unit performance and safety, and is the most frequently used method by experienced end users and their selected experienced SIs. The calcula-tions and methods for determining safe operating limits are more robust, better understood and are based on actual modeling equations that calculate per-formance in real time versus a static curvefit approach.

Recip Compressor Performance, Safety Predictions for Control Panels > Real-time monitoring is often not enough

By DwaynE a. HICkMan

Dwayne A. Hickman is director of soft-ware development with ACI Services Inc. He taught university math and com-puter science courses for 13 years and has been in the reciprocating compressor industry for the past 20 years.

March 2014 62 coMpressortech2

Here, thermodynamics are consid-ered, interstage pressures are rea-sonably predicted rod loads are bet-ter predicted, and the ability to model both double-acting and single-acting compressor modes is standard. This approach works well for gas gather-ing, gas boosting, transmission, in-jection, withdrawal and even some process services.

For slow-speed units, legacy com-pressor manufacturers typically re-viewed the operating ranges (often indicated on unit’s nameplating) for complex issues like internal rod loads, pin loading forces, degrees of pin re-versal and crank pin forces. as such, control panels could effectively ignore those issues and just concentrate on overload, high pressures, static (gas force) rod loads, low volumetric effi-ciencies and high discharge tempera-tures. The problem with this method begins when it is applied to high-speed units under the belief that a slow-speed method is generically valid for high-speed units. Unfortunately, it’s not.

The OEM-based method covers unit performance and safety accord-ing to all OEM limits and methods. while this seems like the obvious choice to implement, a higher level of modeling complexity may come with a premium price.

Furthermore, how does one model unit performance and safety for each OEM’s line of compressors if that knowledge is not readily available? and, how can one ex-pect a low-level (and computa-tionally slower) PLC to handle the

complex mathematical methods that take even a high-end desktop

PC a few seconds to compute?Fortunately, the industry does offer a

couple approaches to satisfy these con-cerns. One approach is via unit-specific

Editor’s Note: This was derived from a paper published on the COMPRESSOR- tech2 website (www.compressortech2.com) under Featured Articles.

continued on page 64

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n Figure 1A. This chart shows safety based on rod loads determined by dynamic internal gas pressures and inertia, and crank pin forces (as specified by OEM).

PLC algorithms based on OEM perfor-mance and safety limits programmed directly into the UCP. Another approach is via use of an add-on micro-controller that computes complex performance in real time using the actual OEM- specified equations and methods.

Using the correct OEM methods for performance and safety is important for safety, warranty and optimization. OEM-X does not warranty its units based on OEM-Y’s method of model-ing mechanical issues like rod loads, crank pin forces, and gas thermody-namics. Its warranty is based on its modeling methods.

Consequentially, if the wrong OEM method is applied, the potential out-comes are:

• The results closely match the OEM’s results. Good luck, head to Las Vegas and buy a lottery ticket — but even a broken clock has the correct time twice each day.

• The results create a conservative operating map. That is, the calcu-lations prevent the unit from run-ning in areas that it could safely operate. This means that the unit is likely not always optimized and

thus not delivering maximum flows and revenues.

• The results allow the unit to op-erate in unsafe areas. The UCP calculations indicate that certain areas in the operating map are safe, when in fact, according to the OEM, operating the unit in those areas results in specific safety lim-its being exceeded — the worst-case scenario for safety.

• The results are conservative in some areas of the operating map while allowing unsafe operations in other areas. This is likely the com-mon case for many current PLC implementations of compressor performance and safety in UCPs — limiting the available operating map on one hand and unknowing-ly exposing the unit to cumulative and potentially catastrophic dam-age on the other hand.

Today’s modern, high-power recip-rocating engines and motors usually include an OEM designed, developed and approved control panel with the purchase of the driver. This OEM con-trol panel maximizes the driver’s ability while also maintaining proper safety

March 2014 64 coMpressortech2

based on current operating conditions — and it can also act independent of the UCP, if necessary, to shut down the driver if serious issues are identified.

The equivalent of a driver OEM control panel does not readily exist for reciprocating compressors. This leaves control, performance predic-tion, safety prediction and dynamic compressor reconfiguration left to the end user, the packager or a third- party automation company.

Since two very expensive pieces of equipment must be coupled together (a driver with a compressor) along with other numerous auxiliary items, control concepts necessary for a wide variety of reciprocating compressor applica-tions makes it difficult for a single set of control logic to address all situations.

However, it is quite possible for two complimentary and interactive control panels (one for the driver and one for the compressor) to coexist with the UCP. With today’s open distributed control architectures, this is a very reli-able approach.

For some situations (e.g., gas gather-ing) only a few system elements, such as speed governors, suction pressure regulators and recycle valves, are auto-mated and thus controllable by a con-trol panel. Some standardized UCPs are encoded with generic control logic for these three basic devices, allowing the UCP to adjust one, two or all three types of devices to keep the unit run-ning within the end user’s requirements.

This type of solution works well when the only issues of concern are overload, low suction pressures, high discharge pressures and maintaining desired flow rates. Nevertheless, since operators can still manually adjust manual vari-able volume clearance pockets (VVCP) to set head end clearances, add/ remove valve spacers or pull valves to deactivate ends, all combinations of pressures vs. speed vs. suction gas temperatures vs. end clearance set-tings vs. ends deactivated must be re-viewed beforehand.

This is required to make sure that running the unit in one of the potential hardware configurations cannot lead to mechanical failure. Areas that lead

continued on page 66

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GCI.indd 1 2/25/14 4:14 PM

March 2014 66 coMpressortech2

n Figure 1B. Specifications are given for three-stage compressor tested.

to violations of the OEM’s safety limits must then be “locked out” from allow-able operations.

Consider a typical two-stage, two-throw, high-speed unit running from 50 to 150 psig (345 to 1035 kPag) suction pressure, 800 to 1100 psig (5516 to 7584 kPag) discharge pressure, 1100 to 1500 rpm, suction gas tempera-tures ranging from 40° to 90°F (4.4° to 32.2°C), VVCPs on both stages and a need to pull suction valves from the first stage’s head end at times.

A review should be made that cov-ers approximately 2 to 16 million dis-tinct operating points before any UCP is allowed to arbitrarily adjust speed or suction pressure. Unless the UCP has been reviewed for all possible combi-nations before it was allowed to control the unit, how can one be confident that there are no areas where the panel may turn down the driver speed in an effort to control low suction pressure only to put the unit in pin nonreversal conditions that burn up bushings and possibly destroy crossheads, rods, pis-tons or even crankshafts? And if those fail, then there exists the possibility of a fire, an explosion, loss of station opera-tions and even injury or loss of life.

And if such failures were to occur, who would accept responsibility: the end user for not requiring the pre-checks across their entire operating range, the packager for not requiring the checks from the UCP provider, the control panel provider, or the com-pressor OEM?

It is not in the best interest for any

party associated with the design, in-stallation and operation of the com-pressor to have that unit fail. OEMs build robust hardware to meet the challenges of end user applications. However, all mechanical systems have limits. Reciprocating compressors of-ten run under conditions of high stress on the components.

OEMs design their units to handle those high stresses. But if the unit is operated in conditions that exceed al-lowed limits, then damage can occur. Continued running of the unit under excessive forces weakens the compo-nents. Thus, the unit may fail later even when it is running within OEM specifi-cations — effectively making the unit’s safety unpredictable. That is not a good situation for anyone, especially the operators.

End users and packagers certainly expect that any installed UCP would act to protect the compressor unit, the station and personnel, and not arbi-trarily adjust control valves, actuate unloading hardware or change speed in such a manner that the mechani-cal forces acting on rods and pins ex-ceeded allowed OEM limits — dam-aging or potentially destroying the compressor. Therefore, end users and packagers must make sure that their expectations are clearly specified in UCP requests (via in-house or third party): provided UCPs must protect the unit according to all of the OEM’s specifications for safe operations.

SIs want their installed UCP to be able to properly start up the unit,

safely operate it and properly shut down the compressor according to the OEM’s specifications. While OEM manuals provide a lot of useful infor-mation in regards to lubrication rates, warmup time periods, etc., acquiring actual OEM equations and methods is not so straightforward.

Most compressor OEMs provide high-end, Windows-based performance modeling software for determining if a specific condition is safe or not. How-ever, they do not generally provide all of the actual equations and modeling methods for others to duplicate the OEM’s calculations and methods.

Many OEMs have openly released sanitized versions of their modeling equations. This is a good start, but SI pro- grammers still do not have all of the data and methods to be 100% compliant with that OEM’s safety requirements. Typical critical methods not released by OEMs are equations or methods relat-ed to: thermodynamics, degrees of pin reversal, pin loading forces, expected interstage pressures, dynamic cylinder end internal pressures and internal gas force plus inertia rod loads.

Nevertheless, SI programmers are being asked to fully incorporate OEM safety limits for each OEM based on that OEM’s developed methods to determine unit safety. Consequently, an SI must: go to the OEM for guid-ance; ask the end user what meth-ods they want to implement; or work with an OEM-approved, compressor performance expert.

continued on page 68

CT339.indd 3 2/24/14 9:52 AM

Gulf South Rotating Gulf South Rotating Machinery SymposiumMachinery Symposium

Baton Rouge, LA.—Crowne Plaza Executive Center

Hands-on technical course training and tutorials covering topics, including:

SEMS (Safety Environmental Management Systems)

Turbo Machinery Reciprocating Engines & Compressors Cranes Pumps Centrifugal Compressors Diesel, Gas, & Bi-fuel Engines Generation Control Systems Diagnostics & Reliability Emissions Alignment Compressor Auxiliary Equipment &

Components View the full agenda at gsrms.org

Visit www.gsrms.org to register today!

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Who should attend: Rotating and static equipment

engineers Maintenance managers, coordinators

and planners Engineering managers Equipment operators Technical specialists Reciprocating machinery analysts

The GSRMS Symposium provides education and training to improve the compression and rotating machinery industry though workshops, hands-on technical courses and tutorials.

Experience. Learn. Grow.

Sponsorships available! Contact gsrms@outreach.lsu.edu for information.

GSRMS/LSU Continuing Education

Attn: Lisa Graves

1225 Pleasant Hall | Baton Rouge, LA 70803

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GSRMS.indd 1 2/19/13 3:33 PM

April 28 – May 1, 2014

Course presenters include representatives from the industry’s leading companies, including Southwest Research Institute, Ariel Corp., Sinor Engine, CPI, Solar Turbine, M&J Valve, Conoco Phillips, FW Murphy, Cameron Compression Systems, Ludeca, Hytorc, Simmons Plating & Grinding, Governor Control Systems, Inc., RuhRPumpen, John H. Carter, and Total Specialties, Inc.

GSRMS.indd 1 2/11/14 8:25 AM

March 2014 68 coMpressortech2

n Figure 2A. This table shows safety based on rod loads based on static gas pres-sures (using old, inappropriate rod load methods). Green area represents all load steps predicted to be safe at all speeds, and all suction gas temperatures.

n Figure 2B. This table shows safety based on rod loads based on static gas pressures (using old, inappropriate rod load methods). When overlaid on actual OEM’s safety map, the majority of the “predicted safe” area is not always safe. In this case, the unit can fail due to crank pin forces being exceeded at various speeds for various load steps.

The first option makes the most sense, as the OEM knows more about its equipment and safety limits than anyone else. However, asking for OEM guidance can be awkward since while OEMs want end users to control their compressors safely, OEMs also need to protect their proprietary meth-ods (developed via costly R&D) that may give them certain advantages over their competitors in select areas (e.g., process markets).

Furthermore, some items, such as gas thermodynamic routines used in OEM software, cannot run efficiently on PLCs, and the OEM may not even have access to the complex source code for those routines to provide the automation programmers.

The second option, asking the end user for guidance, works better with slow-speed units (less than 500 rpm) as end users have had many years of experience with slow-speed units to develop and perfect how they want those units modeled. But even reap-plication of slow-speed units to differ-ent operating ranges from which they were originally applied causes con-cern. Also, many of the slow-speed, end-user performance gurus have retired, leaving a void in many end users’ knowledgebase of compressor performance and safety prediction.

The third option, consulting with compressor experts, is what most end users are opting for in today’s high-speed compressor era. As such, end users often direct packagers and/or automation specialists to consult with specific experts to create appropriate models for the UCP. A compression expert will know the calculation differ-ences between the various OEMs’ rod load and pin load calculation methods. They can review thousands to millions of operating points so as to prereview where the compressor may experi-ence problems. They can calculate valve losses similar to each OEM’s method, and they know how to pre-dict interstage pressures and safety issues for all hardware unloading configurations (load steps) before any physical hardware is changed.

The end-all expert for a compres-sor’s performance and safety is the

CT339.indd 4 2/24/14 9:52 AM

March 2014 69 coMpressortech2

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OEM of that compressor, as the OEM determines the limits, the methods for predicting forces and stresses, and the modeling equations used. Thus, any selected compressor expert should have an excellent working re-lationship with the compressor OEMs.

In regards to performance and safety modeling, where does the com-pressor OEM’s responsibility begin and end? In short, they cannot rea-sonably expect anyone to be able to model and safely control their com-pressors without full disclosure of all limits involved in protecting the unit, and all methods for calculating those limits. However, these disclosures do not necessarily have to be made pub-lic thus revealing proprietary informa-tion. That is, the OEMs can work with a few trusted companies with whom they maintain secrecy agreements.

These select companies can then employ the correct methods and limits and develop simplified equations and methods for the PLC programmers on a per-job basis, providing the desired safety for that specific unit, yet without disclosing proprietary OEM methods. This approach works well for many PLC programmers as it reduces the number of calculations needed to model the unit.

The main drawback is that this ap-proach can potentially reduce the size of the allowed safe operating map. That is, in favor of simpler model-ing equations and limiting constraint equations, the prospective operating map where the compressor can safe-ly run may be reduced by 5 to 50%. Why? To keep the modeling equations as simple as possible, often the worst cases are considered when creat-ing the array of constraint equations needed to keep the unit safe.

Overly conservative constraint equa- tions can effectively reduce the al-lowed operating map, and hence the operational flexibility often needed by the end user to fully achieve its goals.

Well, at least that’s the way it has been up until about 2012. Now, a grow-ing number of current OEMs (Cameron/Ajax, GE Oil & Gas, Knox Western, LeROI Gas Compressors and Arrow Engine) have incorporated their full set

of compressor performance and safety limits into an actual micro-controller add-on for UCPs. Thus, the UCP can now achieve the exact same perfor-mance and safety checks that those OEMs’ Windows-based performance modeling software packages provide.

Most other major high-speed OEMs have provided many of their modeling equations and safety check methods so that the micro-controller add-on provides similar results when calculat-

ing performance and safety limits for those OEMs. As expected, the UCP add-on also provides methods for a host of legacy OEMs.

Having an approved third-party SI handle the controls of the compressor makes sense. Customizing the actual UCP to best meet the needs of the station can lead to optimal use of the equipment. UCP logic should not be confused with compressor performance

continued on page 70

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March 2014 70 coMpressortech2

n Figure 3A. This table shows safety based on rod loads based on adjusted static gas pressures (using another OEM’s method, which is inappropriate since the calculation method is not appropriate for the limits used). The green area represents all load steps predicted to be safe at all speeds, and all suction gas temperatures.

n Figure 3B. This table shows safety based on rod loads based on adjusted static gas pressures (using another OEM’s method, which is inappropriate since the calculation method is not appropriate for the limits used). When overlaid on actual OEM’s safety map, the majority of the “predicted safe” area is not always safe. In this case, the unit can fail due to crank pin forces being exceeded at various speeds for various load steps.

and safety. The latter is a basic neces-sity for the former. UCP logic deals with “what do I want to achieve?” and “what sequence of events, actions and time delays are required to try to achieve those goals?”

Calculating compressor performance and safety predictions by a real-time, micro-controller add-on provides an-swers to the UCP questions: If the com-pressor were to run under the proposed operating conditions, what hardware configurations would be safe? And for the ones that are safe, how much load is required? How much flow is predicted? How efficient is the unit at those condi-tions? And, does the compressor need to be shut down for any safety reasons?

SIs of reciprocating compressors are challenged to design, code and test the control logic for properly start-ing compressors, safely operating them across a defined operating map, and properly shutting them down. Part of most control logic involves adjust-ments to operating speed, adjust-ments to suction pressure regulators, actuation of unloading hardware and control of recycle valves.

Often, SIs and end users assume that because there is a plethora of pressure, temperature and vibration sensors installed throughout the sys-tem, that issues would be caught by those sensors — a bad and poten-tially dangerous assumption.

Very few reciprocating compres-sors actually have real-time internal pressure sensors. Standard pres-sure sensors only measure the suc-tion and discharge pressures nearby the cylinder flanges. These are useful for checking against low- and high- pressure limits, but not directly useful for OEM models that base rod loads on dynamic internal pressures.

Discharge temperature sensors can catch high discharge temperatures, but only if the discharge valves open and release the hot gas. Deactivated ends and ends operating in very low suc-tion volumetric efficiencies do not lead to hot gases exiting out the discharge valves. As such, internal tempera-tures can climb to 350°, 400°, 500°F (175°/204°/260°C) and hotter. Eventu-ally, valves, packing and rings will fail,

CT339.indd 6 2/24/14 9:53 AM

binations remain unsafe. Green areas reflect when all load steps are safe at all speeds. The key issues noted on the graphic are: C: Crank Pin Forces Ex-ceeded, R: Rod Load Forces Exceed-ed, U: Gas is Throttling through one of the stages, and O: Unit is Overloaded. Excessive crank pin forces can damage the crankshaft, and excessive rod loads can damage the piston, rod, crosshead and even crankshaft. Either one can lead to catastrophic damage to the unit.

Figure 2A shows the same compres-sor, but this time it is being modeled by standard compressor performance and safety calculations one would find by searching the Internet, or in older compression theory textbooks. Notice that Figure 2A shows no areas of ex-cessive crank pin forces. And, areas marked as rod load concern (R) are quite different than those indicated in Figure 1A. Figure 2B overlays Figure 2A over Figure 1A with some translu-cency. The blue boundary line helps to highlight where the non-OEM, generic method would indicate to the end user and the UCP that the unit is safe even

though the unit may very well be under excessive stresses.

Figure 3A shows the same compres-sor, but this time it is being modeled by a different OEM’s equations and safety limits. Again, notice that Figure 3A shows no areas of excessive crank pin forces. And, areas marked as rod load concern (R) are quite different than those indicated in Figure 1A. Figure 3B overlays Figure 3A over Figure 1A with some translucency.

The blue boundary line helps to highlight where the alternate-OEM method would indicate to the end user and the UCP that the unit is safe even though the unit may very well be under excessive stresses. That is, for some combinations of load step and speed, the unit is safe, while for other combinations of load step and speed the unit is unsafe.

Since the earliest control panels, end users held high hopes that UCPs could keep their units safe, but knew, due to limitations in computing power, that some safety items were not being

leading to consequential and potentially catastrophic damage. Thus, a few sen-sors cannot make up for not implement-ing the correct performance and safety models in the PLC — any expectation otherwise can be a dangerous choice.

Figure 1A shows the Safe Oper-ating Map of an OEM’s three-stage compressor (see specifications in Figure 1B) across a suction pres-sure range of 0 to 100 psig (0 to 690 kPag), a discharge pressure range of 100 to 1000 psig (690 to 6895 kPag), a speed range of 700 to 1000 rpm, six load steps and suction temperatures for the three stages varying from 100 to 140°F (38° to 60°C).

The Safe Operating Map is gener-ated after reviewing over 8.5 million dis-tinct operating points using the OEM’s exact methods for calculating safety issues. Red areas are where the unit cannot safely operate regardless of what load step is used and regardless of what speed the unit is running. Yellow areas reflect where some combinations of load step and speed can be used to safely operate the unit while other com- continued on page 72

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checked. However, since at least 2004, any new UCP is expected to correctly know when it is safe and unsafe to op-erate the unit being controlled.

Knowingly operating compressors in unsafe areas is unacceptable. Un-knowingly operating them in unsafe ar-eas also is unacceptable. Operating a compressor by using overly conserva-tive methods and thus restricting its op-erating map (lost revenues, potentially higher emissions, more shutdowns, more wasted energies from pressure throttling and recycling, etc.) is better, but not ideal. Operating a compressor by implementing that compressor’s OEM methods is ideal.

End users expect maximum use of their compressors, which includes safely covering the widest operating map, running the unit most efficiently and accurately predicting load, flow and unit safety. If the unit is going to be controlled by a UCP, then automa-tion specialists should be required to integrate into their control logic the proper OEM-approved methods for determining unit safety. Anything short of full implementation of the OEM- approved performance and safety methods is not an acceptable best practice for today’s control panel.

Finally, many optimization and/or monitoring service companies provide valuable information and insight into the fleet of compressors being mod-eled. These service-based companies take operating data from the field, run the data through a series of calcula-tions, trend the data, compare to his-toric values and generate reports.

However, just like automation com-panies, too often inappropriately ap-plied equations are used, thus poten-tially rendering the information from those reports anywhere from mean-ingless to misleading, and even pos-sibly dangerous.

The concluding rule to use for end users is clear and simple: When auto-mating a unit, or adding it to a moni-toring service, make sure that your selected vendor applies the correct OEM safety and performance meth-ods when modeling the unit. Ask for it. If the bidding vendor cannot provide it, then select a vendor that can. CT2

March 2014 72 coMpressortech2

PRODUCTSFEATUREDTemperature Sensor

Motortech has developed a new temperature sensor compatible with Caterpillar GS3500 series gas en-gines. The sensor is an alternative to Caterpillar’s P/N 3832989/2419591 and serves as a replacement with-out modifications, according to Mo-tortech. The temperature sensor P/N 56.01.092-28 is compatible with Cat-erpillar gas engines G3508, G3508B, G3512, G3512B, G3516, G3516B and G3520B, the company said.

www.motortech.de

Alignment Application

Ludeca has introduced tab@lign, a tablet-based application for pump-motor alignment, which combines the Prüftechnik laser measurement technology with a tablet and smart-phone devices. Three steps — enter dimensions, rotate shafts and display measurement result — are all that’s needed for the app to align horizontal machines, the company said.

Other features include active clock measurement mode, Bluetooth capa-bility, the ability to measure, correct and save soft foot results, live mode for real-time corrections and preassembled brackets for quick mounting.

The tab@lign app runs on Apple and Android mobile devices. It can be downloaded free from the Apple App Store or Google Play.

www.ludeca.com

Emissions AnalyzerE Instruments has released the

E5500 portable emissions analyzer for industrial flue gas.

The user can choose several elec-trochemical gas sensors — O2, CO2, CO, NO, NOx and SO2. Temperature measurements for flue gas and air, as well as the differential temperature, are standard features. An internal pressure sensor is designed to allow the analyzer to measure pressure and stack draft. Gas velocity can be mea-sured using the differential pressure and an optional pitot tube.

The standard EGAS software pack-age offers the ability to save and graph data in real time in the field with a laptop or in a laboratory with a PC. The device can be configured with low-range NOx sensors with 0.1 ppm resolution, the company said.

www.e-inst.com

Temperature Measurement System

Photon Control Inc. has released the PalmSense2, a portable, hand-held fiber optic temperature mea-surement system.

The PalmSense2 is designed for high temperature RF environments or EMI applications in labs and field serv-ice. The portable nature of the device, as well as its 24-hour battery life (in continuous use mode) allows the user to move while taking measurements.

The PalmSense2 features a tem-perature range up to 842°F (450°C), accuracy of ± 32.09°F (0.05°C) (probe dependent) and resolution of ± 32.02°F (0.01°C). The PalmSense2 can be paired with both contact and immersion temperature measurement probes.

www.photon-control.com

CT339.indd 8 2/24/14 9:53 AM

MARCH 2014 73 CoMpRessoRtech2

Recent OrdersMAN Diesel & Turbo

MAN Diesel & Turbo’s first her-metically sealed compressor to be installed in the North Sea will go to Det Norske Oljeselskap’s Ivar Aasen production platform.

The high-speed, oil-free, inte-grated motor (HOFIM) compres-sor unit is similar to MAN’s subsea compressors currently under sys-tem integration tests in Norway.

The order for Ivar Aasen consists of a multistage radial compressor (1x100%) arranged in tandem con-figuration around a centrally posi-tioned 12,740 hp (9.5 MW) high-speed electrical motor. MECOS, which MAN acquired in early 2012, provided the integrated active mag-netic bearings.

The HOFIM unit will be delivered to the module packager SMOE Pte Ltd. in Singapore, which is building the Ivar Aasen platform. In 2016, the module will be towed to the Nor-wegian North Sea.

Valerus Field SolutionsValerus Field Solutions has re-

ceived a US$62 million contract to provide engineering, procurement, construction and commissioning of two compressor stations in Dod-dridge County, West Virginia, for Crestwood Midstream Partners.

The Marcellus Shale facilities, each rated at 120 MMcfd (3.4 x 10.6 m3/d), are scheduled for com-pletion in 2014.

The turnkey compressor stations will together include more than 35,000 hp (26 MW) of compres-sion, inlet separation and filtration, gas dehydration, power genera-tion and distribution and station in-strumentation and controls.

Kentz Corp. Ltd., the holding com-pany of the Kentz engineering and construction group, completed its US$435 million purchase of Valerus Field Solutions in January. The re-maining Valerus contract services and aftermarket service businesses have been renamed Axip. CT2

(918) 283-9200 Fax (918) 283-9229 www.axh.com

Experience, Reliability, Integrity...

Now with three plants totaling over 400,000 SF on 53 acres

AXH.indd 1 2/17/14 9:06 AM

The stories listed below recently ap-peared on the COMPRESSORtech2 internet page. Visit www.compres-sortech2.com to read the complete ar-ticles. Subscribers to the digital maga-zine can simply click on the headline.

GE To Acquire Cameron’s Reciprocating Compression DivisionCat Issues Upgrade For G3516 LE EnginesEmerson Introduces Vapor Recovery UnitsValerus Gets Work On Two StationsBolivia To Get GE’s ICL CompressorsGE Starts Downstream Tech Solutions Valerus Compression Services Renamed ‘Axip’Compressor Stations Lambasted At MeetingsCameron May Sell Centrifugal BusinessMAN Compressor Set For Norway FieldU.S. Gas Storage Withdrawals Set Record

HEADLINESWEB

PRODUCTSFEATURED

continued on page 82

Fluid Ends

Caterpillar Global Petroleum has added stainless-steel fluid ends to the Cat XD line of well stimulation pumps and fluid ends. These fluid ends are compatible with all Cat XD Series components.

Benefits of the stainless-steel fluid ends include improved corrosion re-sistance, reduced incidents of crack-ing due to corrosion and improved durability, the company said.

www.catoilandgasinfo.com

CT343.indd 1 2/24/14 9:55 AM

Laborde ProductsLaborde Products Inc., headquartered

in Covington, Louisiana, has opened a division — Laborde Equipment Serv-ices (LES) in San Antonio, Texas.

The facility provides service to the San Antonio metro area and the Eagle Ford Shale market. The 18,000 sq.ft. (1672 m2), 3 acre (1.2 ha) facility has a 15 ton (13,608 kg) overhead crane capacity.

LES provides service for all makes of diesel and gas engines and engine- driven equipment, as well as hy-draulic and electrical service. It also stocks and provides parts for a broad range of support.

Roger Markwardt heads the division. He has more than 30 years of experience in the diesel engine and rental industries.

Laborde Products has also operated a branch location in Channelview, Tex-as, for the past seven years.

FS-ElliottFS-Elliott Co., a manufacturer of oil-

free centrifugal air and gas compres-sors, has expanded into Southeast Asia by establishing an office in Selan-gor, Malaysia.

Led by Greg Baldwin, director of Business Development for Southeast Asia, the office will seek to expand partnerships.

FS-Elliott also has offices at Bas-ingstoke, U.K.; Houston; Jubail, Saudi Arabia; Los Angeles; Mangalore, India; Shanghai; and Taipei, China.

Detechtion TechnologiesDetechtion Technologies has re-

ceived an undisclosed investment by Element Partners, a growth equity fund focused on energy and industrial tech-nology companies.

Detechtion provides software-based optimization and fleet monitoring serv-ices for about 20% of the active com-pression horsepower in North America and manages assets for 15 major gas producers.

Concurrent with the investment, Chris Smith joined the company as president and CEO and Gerry Conroy in the newly created role of senior vice president of products and portfolio.

Smith was formerly CEO of Cygnet Software, a provider of natural gas SCADA products and services. Conroy was vice president of global products at P2 Energy Solutions.

Detechtion founder Brian Taylor, previously president and CEO, will be-come chairman of the board. Andrew Miles will continue as senior vice presi-dent of operations.

Dresser-RandDresser-Rand has joined Gaelectric,

a renewable energy firm, in the develop-ment of its compressed air energy stor-age (CAES) site near Larne, Northern Ireland. They also formed an alliance to develop other European CAES projects.

When completed, the £300 million Larne facility will comprise a 268 MW

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twin powertrain storage and electricity generation facility. Target commissioning date is 2017.

Dresser-Rand will provide project management, sched-uling and technical support services through the planning phase of the Larne project, including front end engineering and design in 2014.

Wood Group GTSWood Group GTS has appointed Mike

Fisher as president of its U.S. oil, gas and industrial services (OGIS) business. This is a new position for the Houston-based business that provides maintenance, re-pair and overhaul for Solar Mars, Saturn 10, Centaur 40, Centaur 50 and Taurus 60 gas turbines, which are used primarily by the oil and gas industry.

Fisher joined Wood Group in 2005, working initially in Venezuela where he managed a two-year project to increase the reliability, performance and uptime of the Solar and Siemens turbine fleets on Lake Maracaibo for PDVSA, the Venezuelan national oil company.

He subsequently served within the operations group for power plant services and most recently has been vice presi-dent of Solar operations within OGIS.

Axip Energy ServicesValerus Compression Services LP selected Axip Energy

Services LP as the new name for its Contract Services and Aftermarket Services business, which began operating as a stand-alone company on Jan. 3, following the sale of Valerus Field Solutions LP.

Axip is focused on delivering service-based offerings for customers, including contract compression, contract production and processing solutions, fee-based gathering, processing and pipeline solutions and aftermarket field ser-vices globally. Axip also retains the Command performance optimization offering, which was released to the market in late 2013.

Axip’s 700 employees include Valerus contract services and aftermarket services personnel plus relevant support services personnel. The current Valerus Compression Serv-ices management team will continue with Axip and oversee the transition and growth of the company.

Caterpillar Oil & GasCaterpillar Oil & Gas has released a low-emission upgrade

kit for select G3516 LE petroleum engines used in gas com-pression applications.

The upgrade kit allows operators to modify existing en-gines to a lower emission configuration, enabling operation at 0.5 or 1.0 g/bhp-hr NTE NOx levels.

Caterpillar said the upgrade is an economic, minimally invasive solution that allows the engine to offer the same

Einstruments.indd 1 3/7/13 9:49 AM

MoversPRIME

MARCH 2014 75 CoMpRessoRtech2

power, 1340 bhp (1 MW) at 1400 rpm, and fuel tolerance while increasing alti-tude and turndown capability.

New components in the kit include a two-stage aftercool-er, venture fuel mixer, turbochargers, oil cooler, air cleaners, gas shutoff valve and software. Existing engines must have the ADEM A3 engine control module, air fuel ratio control (AFRC) and a NOx sensor. Upgrade kits also are available to bring engines up to this configuration.

MonicoMonico Inc. has promoted Bill Dicken to technical sup-

port manager.He will oversee phone support, field service and the com-

pletion of special configuration projects.Dicken has 17 years of experience working with Rolls-

Royce engines, with 12 spent in customer service for aviation gas turbine engines.

M. Fisher

PrimeMovers.indd 2 2/24/14 3:25 PM

Eugene Broerman is a senior research engineer with Southwest Research Institute (SwRI). He holds a bachelor’s degree in me-chanical engineering from Texas A&M University-Kingsville. He has been involved with reciprocating compressor pulsa-tion control research work since 2004. Contact him at: eugene. broerman@swri.org. Ray Durke is a senior research engineer with SwRI. He has more than 30 years of experience with vibration-related problems. He has a bachelor’s degree in mechanical engineering from Texas A&M University and a master’s from the University of Texas at San Antonio. Con-tact him at: ray.durke@swri.org. Richard Baldwin retired from SwRI in 2012 and is an engineering consultant. He has ex-perience with pulsation, vibration, corrosion, and reliability problems in plant systems. He has a bachelor’s degree in en-gineering from Trinity University and a master’s from Georgia Institute of Technology. Contact him at: engineerbaldwin@ satx.rr.com

Case Study: Intake/Exhaust Silencer Redesign Mitigates Noise >

Southwest Research Institute (SwRI) was contracted to investigate complaints of rattling and shaking of windows, doors, cabinets, etc., from residents near

a gas transmission station.Compression at the station was provided by two models

of integral compressors: Ingersoll Rand KVG (10-cylinder) and KVS (12-cylinder) units and by two gas turbine-driven centrifugal compressors. Field tests were conducted to identify the sources of the neighbors’ complaints and to aid in development of modifications to reduce complaints.

This article presents field testing activities to verify the excitation sources, a pulsation analysis to develop modifications for reducing offending pulsations, and follow-up field tests for verification. The subject problem is somewhat unique in that the complaints of rattling are vibration-related, but the primary driving source is inaudi-ble pulsations produced in the engine intake and exhaust manifold piping.

Field testing to identify problem sourcesIn an effort to characterize the local complaints, sound

pressures and vibration measurements were recorded at nearby home sites and outside the compressor building as the operating conditions of each type of compressor was varied.

Vibration measurements were recorded on the founda-tion of the residences as well as in the ground near each home to assess the potential of vibration transmitted from the compressor station. In summary, it was found that ground-borne vibration was not present.

Airborne sound pressure measurements identified a peak occurring near 12.5 to 13 Hz at the residences and near the compressor building as seen in Figure 1. This frequency is below the range of human hearing, but was considered a likely source of rattling. The noise measurements were made in engineering units of Pascals due to the low pressure levels. One Pascal (Pa) is equal to 0.000145 psi (0.000009 bar).

n Figure 1. Graphs show sound pressure data recorded at resi-dence with 4 KVG, 2 KVS and turbine operating.

In order to identify the sources of the airborne pulsations, testing was conducted with various units operating over a range of conditions to track the pulsations near 12.5 Hz. Field measurements showed that the elevated pulsations near 12.5 Hz corresponded to 2.5 times (2.5x) the running speed of the KVG units.

TECHcornerSouthwest Research Institute solves problem of airborne pulsations from compressor station

By EuGEnE ‘Buddy’ BROERmAn,

RAy duRKE And RICHARd BAldwIn

Editor’s Note: This article is based on a paper given by the authors at the Oct. 6-9, 2013, Gas Machinery Research Council meeting at Albuquerque, New Mexico.

MARCH 2014 76 CoMpRessoRtech2

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TECHcorner

MARCH 2014 77 CoMpRessoRtech2

When the engine running speed was varied over the op-erating range of 300 to 330 rpm, the 2.5x component in the sound and pulsation data tracked speed and varied from about 12.5 to 13.75 Hz. A summary of those pulsation mea-surements near 2.5x is presented in Table 1. The testing effort identified the KVG units as the primary source of the elevated 2.5x energy and eliminated the KVS units and the turbine-driven units as potential sources.

n Table 1. Noise Pressures (Pascals, pk-pk) Detected near 2.5x at various test locations.

Airborne sound pressure measurements near the in-let duct to the engines could not distinguish whether the source was from the intake or the exhaust. A pressure measurement taken inside the inlet duct shows high pul-sation at 2.5x the running speed and much lower ampli-tudes at several harmonics of running speed as displayed in Figure 2. The indication is that at least a portion of the noise at the residences originates from the inlet duct.

n Figure 2. Pulsation data recorded in Unit 4 intake duct showing 2.5 multiple pulsations.

Field testing to map air intake and exhaust pulsationsA separate testing effort was conducted utilizing high

temperature transducers to measure pulsations in the en-gine inlet and exhaust manifolds in order to define pulsation characteristics contributing to the high 2.5x pulsations and to provide data for correlation with an acoustic model of the intake/exhaust manifolds. A schematic of the exhaust sys-tem is provided in Figure 3.

The pulsations at 2.5x the running speed dominated the spectrum and reached a maximum at the capped end of the manifold, location P9, as seen from the data in Figure 4.

n Figure 3. Schematic shows engine exhaust manifold.

n Figure 4. This graph shows the pulsation spectra recorded in the exhaust manifold.

From the summary of pulsation amplitudes in Table 2, note that the 2.5x pulsation amplitudes measured in the exhaust piping are thousands of times more than that mea-sured at the residences.

n Table 2. This is a summary of maximum 2.5x pulsations mea-sured in the exhaust manifold.

n Figure 5. The engine intake manifold schematic is shown.continued on page 78

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MARCH 2014 78 CoMpRessoRtech2

The inlet manifold is represented in Figure 5 and is a somewhat more complex piping system. The field measure-ments identified a strong acoustic response at 2.5x running speed that contains a pulsation maximum at the capped end of each cylinder bank (test locations P5 and P8), and a minimum near the header between the cylinder banks (be-tween test points P3 and P6).

Spectra pulsation plots of data recorded inside the intake manifold show the 2.5x pulsations present over the speed range, which is similar to the pulsation data measured in the exhaust system. A summary of the maximum pulsation am-plitudes near 2.5x is provided in Table 3. The pulsation am-plitudes in the intake manifold are lower than those in the ex-haust by about a factor of three. Data from the field studies was used for calibration of an acoustic model of the manifold systems. Comparisons of the field measurements with the model predictions are included in subsequent sections.

n Table 3. This is a summary of maximum 2.5x pulsations mea-sured in the inlet air manifold.

n Figure 6. The pressure transducer installed at the end of the manifold (test point P5).

Figure 6 shows an example installation of a pressure transducer in the intake manifold at the capped end of the cylinder bank near cylinder one, test point P5.

Pulsation modelAn acoustic model of both the inlet and exhaust mani-

folds was developed to explore the sources of the energy

near 2.5x running speed and to investigate potential modi-fications for pulsation mitigation. The extent of the piping included in the modeling is summarized in Figures 3 and 5. A summary of model results are presented in the follow-ing sections.

Exhaust manifold model

The pulsation model of the exhaust piping indicated that an acoustic response was present in the system near 12.5 Hz. A pulsation maximum was predicted at the capped end of the header and lower amplitudes near the stack. The red plot in Figure 7 shows model predictions of the pulsa-tion in the stack for the existing piping system. The spec-trum is dominated by the response peak near 12.5 Hz

n Figure 7. The graph shows predicted pulsations in the ex-haust stack.

Modifications to reduce pulsations by changing the mani-fold length, adding orifice plates, and providing an acous-tic filter installed as part of the silencer were investigated. The operating company desired the most reliable solution, which was an acoustic filter designed into the silencer to attenuate pulsations associated with 2.5x engine running speed and higher harmonics.

n Figure 8. This exhaust piping pulsation summary gives predic-tions and field data. (Reference Figure 3 for test point locations.)

Figure 8 shows pulsation amplitudes near 12.5 Hz predict-ed by the model in red for the existing manifold and in blue with the new silencer installed. Field measurements acquired at four locations in the manifold piping, indicated by stars, show good correlation with the model and field data. The im-portant aspect of the design is the predicted amplitudes in the exhaust stack are reduced by a factor of about eight with the new silencer installed. It was assumed that the primary coupling of the excitation energy near 2.5x running speed to

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MARCH 2014 79 CoMpRessoRtech2

the atmosphere occurs at the exhaust stack outlet. A spectral plot of pulsation predicted in the stack is shown in Figure 7, where the red line is the model prediction of pulsations with the original silencer and the blue line is the predicted ampli-tude with the new silencer installed.

Replacing the existing silencer with the silencer present-ed in Figure 9, results in predictions of significant reduction of 2.5x running speed pulsations in the piping downstream of the filter; therefore, installation of the new silencer should reduce the pulsations that couple into the atmosphere at the end of the stack.

Inlet manifold pipingThe pulsation analysis of the engine inlet air piping indi-

cated that an acoustic response should be expected in the system near the top range of 2.5x running speed. The acous-tic resonance is associated with the length of piping from the capped end of the cylinder six through 10 bank to the capped end of the cylinder one through five bank. The response pre-dicted at the inlet to the original intake system is given in red in Figure 10. As shown, the response amplitude is significant from approximately 11 to 15.5 Hz. Field data matches rela-tively well with the model predictions, showing higher 2.5x pulsations at higher running speeds.

n Figure 9. Drawing shows the new exhaust silencer to attenuate 2.5x pulsations.

n Figure 10. This gives the predicted pulsations at the intake manifold.

The acoustic mode shape of the pulsation response in the inlet manifold model is provided in Figure 11 for the cylin-der bank of cylinders six through 10. The model predictions (shown as a red line) match the field data (plotted as red stars) fairly closely and show a pulsation maximum at the closed end of the cylinder bank and a minimum near the inlet.

n Figure 11. This intake piping pulsation summary shows predic-tions and field data. (Reference Figure 5 for test point locations.)

Acoustic modifications were investigated to mitigate the 2.5x pulsations, which included orifices to damp the reso-nance, a piping length change to shift the frequency of the response, and the addition of side branch resonators to ab-sorb the pulsations. A more reliable modification to provide pulsation attenuation for airborne pulsations from the inlet piping was developed.

The design utilized an acoustic filter with an inlet silenc-er similar to that developed for the exhaust system. The acoustic filter was designed to attenuate the 2.5x engine pulsations upstream of the silencer that would pass into the atmosphere. A drawing of the new inlet silencer design is provided in Figure 12.

Pulsation predictions within the intake piping with the si-lencer installed, the blue line in Figure 11, do not vary sig-nificantly from pulsations in the original manifold model (the red line in Figure 11). However, the objective of the new inlet silencer is to attenuate the pulsations near 2.5x that pass into the atmosphere. Therefore, the key pulsation data is at the inlet to the intake system.

The predicted pulsation spectra at the 10 in. inlet piping can best be seen in Figure 10 for both the original piping (in red) and with the new silencer installed (in blue). Adding the new silencer to the acoustic model results in a reduction of

continued on page 80

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MARCH 2014 80 CoMpRessoRtech2

the 2.5x pulsations by approximately 2:1 at the inlet to the intake system. These predictions are an indication that the pulsations transmitted into the atmosphere at the engine intake will be significantly reduced.

n Figure 12. This drawing is the new inlet silencer.

Field test to verify model predictionField measurements were recorded with the recommend-

ed inlet and exhaust silencers installed on one unit to verify performance prior to installation on all six units. Field mea-sured pulsations in the intake and exhaust piping systems are summarized by the blue stars in Figures 8 and 11.

Pulsation amplitudes at the inlet and outlet of the intake and exhaust systems, respectively, were significantly re-duced after the new silencers were installed, as predicted in the pulsation modeling. Field sound measurements were recorded at locations similar to those recorded in the previ-ous field tests with microphones located outside the com-pressor building and at Homes 2 and 3 to assess the ef-fects of the new silencers. Table 4 provides a summary of the measurements and shows a significant decrease in the 2.5x pulsations with the silencers installed.

n Table 4. These sound pressures were at 2.5x at noted test locations.

The test results indicate that the sound pressures at 2.5x are lower with the new silencers by factors of about 3:1 near the compressor building, 2:1 at Home 2, and about 6:1 at Home 3. With the lower levels present, the sound pres-sure measurements are more erratic and can be affected by local noises. Figure 13 shows spectral plots recorded at Home 2, while Figure 14 shows sound pressure measured in the compressor station.

The data distinctly indicates that the new silencers atten-uate the 2.5x sound component, and also show the pres-ence of more random low-level sound pressures present over the lower frequency band.

n Figure 13. This graph is of the sound pressure spectra at Home 2.

n Figure 14. This shows sound pressure spectra at the compres-sor station.

Based on the results of the tests, the decision was made to install the silencers on the remaining units. A follow-up field study was performed to evaluate the sys-tem when all six units were fitted with the new silenc-ers. Sound data was measured at the station and at two home sites. A reduction of at least 2:1 was measured at the compressor station, a reduction of more than 6:1 at one home site, and a reduction of approximately 5 to 30% at the second residence.

It is not clear why the noise measured at one of the nearby

CT308.indd 5 2/24/14 10:06 AM

able option for reducing the pulsations transmitted into the atmosphere was replacement of the existing silencer with a low-pass acoustic filter type silencer designed to attenuate the 2.5x running speed pulsations (and higher frequency pulsations). This design was predicted to pro-vide a reduction of about 8:1 in the 2.5x pulsations as measured in the exhaust stack.

• Acoustic modeling of the engine intake manifold indi-cated that a half-wave acoustic resonance existed in the system near the 2.5x frequency. A low-pass silencer was designed to attenuate the 2.5x pulsations similar to that of the exhaust system. A maximum reduction of about 2:1 in the 2.5x pulsations was predicted at the inlet.

• Field measurements conducted with the new silencers on one unit revealed that the sound pressure spike near 2.5x was eliminated. The measured reduction of the 2.5x pulsation amplitudes varied from a factor of 2:1 to 6:1 as measured near the compressor building and at two nearby homes.

• Field tests conducted with the new silencers installed on all six units revealed that the sound pressure spike near 2.5x was eliminated at measurement locations in the com-pressor station and at one of the home sites. At a second home site, a 2.5x sound pressure spike remained. Since the sound pressure spike was eliminated in the compres-sor station and at one home site, questions remain regard-ing another potential sound source. CT2

MARCH 2014 81 CoMpRessoRtech2

home sites was not significantly reduced by the installation of the new silencers; however, it is clear that the 2.5x pulsa-tions transmitted into the atmosphere were reduced signifi-cantly by the installation of the newly-designed silencers.

Summary The key points of the field investigation, pulsation analysis,

and follow-up testing were:• Initial field tests to address community annoyance

complaints identified airborne sound pressure peaks occurring at homes near the compressor station that corresponded to 2.5x running speed of one type of com-pressor unit. Although the 12.5 to 14 Hz frequency was well below the range of human hearing, this lone peak in the sound spectrum was considered the most likely cause of the rattling and shaking complaints.

• Pulsation measurements inside of the engine intake and exhaust manifold piping revealed strong pulsa-tions at 2.5x running speed and was shown to be the source of the elevated sound pressure energy detected at the local houses. The field data provided correlation for acoustic models of the air ducts that identified acoustic responses in both the intake and exhaust manifold piping.

• An acoustic model of the exhaust gas manifold identified a quarter wave pulsation response in the piping system. Of several potential design modifications, the most reli-

EGCR.indd 1 2/4/14 3:34 PM

CT308.indd 6 2/24/14 10:06 AM

PRODUCTSFEATURED

MARCH 2014 82 CoMpRessoRtech2

Bearing Fault DetectorIMI Sensors has launched the

bearing fault detector PLUS, which is a sensor designed to monitor roll-ing element bearings and provide a 4 to 20 mA signal to plant monitor-ing and control systems for early fault alarms.

The model 649A03 contains an accelerometer and 4 to 20 mA transmitter in a single housing and has multiple specialized outputs. When used with a control system, these outputs provide an early warning of bearing defects such as cracked races, brinelling, looseness and spalling.

The bearing fault detector PLUS is designed for applications with critical rotating machinery requiring continu-ous monitoring, the company said. It is also a simple drop-in replacement for traditional vibration transmitter units as it uses the same loop power and 4 to 20 mA output.

www.imi-sensors.com

Oil Drain Valve

Global Sales Group has introduced its engine oil drain valve for industrial engines and construction equipment. The EZ Oil Drain Valve replaces the traditional oil drain plugs. The nickel-plated valve features a ball valve mechanism for leak-proof op-eration, an O-ring seal and rubber lever cover.

Once installed, users only have to turn the lever to drain oil, and turn it again to close it, the company said. A hose may be attached to the nip-ple end to drain the oil away from the engine.

www.ezoildrain.com

Bearings And Wear Rings

Graphite Metallizing Corp. has re-leased its Graphalloy bearings and wear rings, which are designed to ad-dress pumping problems caused by low viscosity, light hydrocarbon liquids such as natural gas liquids (NGL), LNG and CO2.

Graphalloy bearings are used in horizontal and vertical pumps. The nongalling, self-lubricating features of Graphalloy allow pumps to con-tinue working even when experienc-ing run-dry, flashing or cavitation, the company said.

According to Graphite Metallizing, light hydrocarbons can be difficult for metallic bearings because the hydro-dynamic film provided by these low-lubricity fluids is unable to provide enough lubrication, which can lead to metal-on-metal contact, galling or seizing of the pump.

Using Graphalloy bearings and wear rings allows tighter clearances, improv-ing reliability, lowering vibration and in-creasing efficiency in vertical and hori-zontal pumps, the company said.

www.graphalloy.com

LED Fixture

Dialight revealed its latest LED lin-ear fixture, designed for Class I Div. 2 certified hazardous applications such as on- and offshore drilling rigs, plat-forms and other oil, gas, chemical, petrochemical and hazardous loca-

tion facilities. The fixture is intended to replace traditional fluorescent and HID lighting fixtures, and is also avail-able for nonclassified general purpose industrial applications.

Featuring 106 lumens per watt, the SafeSite LED linear fixture is available in 2 and 4 ft. (0.6 and 1.2 m) models with a weight of 11 lb. (5 kg). The 100 to 277 Vac linear fixtures also feature an integrated wiring box to provide ac-cess for making electrical connections in a separate compartment.

The SafeSite is L70 rated for more than 100,000 hours and features a T4a temperature rating of 40° to 65°C (-40° to 149°F).

www.dialight.com

Protective Shield

Clark-Reliance Corp. has released a new protective shield for use on armored glass, liquid level gages. Safe View shields protect nearby op-erators from high-pressure leaks, and can be retrofitted to Jerguson brand level gages, as well as many others, the company said. Made of Lexan polycarbonate, Safe View shields are available in lengths to fit flat glass gage styles.

www.clark-reliance.com

Analytics SoftwareVisage Information Solutions re-

leased Visage 2014, a new version of its visual analytics software, which is designed for analyzing oil and gas data.

Building on the company’s experi-

Products.indd 1 2/24/14 10:08 AM

ence in the oil and gas industry, the software provides thousands of ways to examine data and generate analy-ses to inform better decision-making and increase competitive edge, the company said.

www.visageinfo.com

Turbine FiltersChampion Laboratories (Europe)

Ltd. has launched its FRAM Indus-trial E10 filter line, designed for gas turbines.

Featuring a 100% synthetic me-dia constructed from polyester and polyethylene, the filter has been in-

dependently tested and certified to EN1822:2009.

Key features include interchange-able components; FRAM-Clench, an inner and outer clenched liner that provides a mechanical lock with-out damaging the liner surface; and FRAM-Bead, a technology that maxi-mizes dust-release space to prolong the life of the filter.

www.framindustrial.com

Digital Process MeterH. G. Schaevitz LLC Alliance Sen-

sors Group has expanded its line of sensor support electronics with the release of the M-100 series single-channel digital process meter.

The M-100 digital meter is de-signed for measurement applica-tions that display an analog sensor’s output or the output of most types of thermocouples.

It offers a 0.3 in. (7 mm), four-digit

LED display with 16-bit resolution in a 1/32 DIN size that can be mounted. It also can be calibrated to display an analog sensor’s dc voltage or current output in engineering units, or tem-perature in Fahrenheit or Celsius. The M-100 meter comes with peak-hold functions standard, along with internal relay contacts for set-point control of an alarm or annunciator. It also fea-tures the ability to retransmit a 4 to 20 mA analog signal to a control system or data logger.

www.alliancesensors.com

PRODUCTSFEATURED

MARCH 2014 83 CoMpRessoRtech2

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march 2014 87 compressortech2

Compressor suction and discharge valves are high-speed check valves that seal gas inside the cylinder dur-ing gas compression and expansion. The valve plates open when the gas pressure on one side overcomes the spring forces and the pressure on the other side.

Conversely, the plates close when that pressure is no lon-ger high enough to resist the spring force and the pressure on the back side of the plates. This action happens many times each second (e.g., 5 at 300 rpm and 30 at 1800 rpm).

Early compressors operated at low speeds using valves with concentric rings, reeds, channel plates or other ap-proaches with multiple coil springs. Most U.S. compressor manufacturers at the time, e.g., Ajax, Cooper-Bessemer, Superior, Worthington, Clark and Ingersoll Rand (IR), made their own valves.

As higher-speed gas engines were developed, compres-sor manufacturers responded with separable compressors to match these newer, more cost-competitive drivers. HCA supplied high-speed plate valves 800 rpm IR RDH and RDS frames, as well as the Knight KOC and the EI and Gemini compressors. These were mostly called CGD valves, which employed metal plates with a unique double damping system.

But the incumbent valve technology threatened to limit the size and speed of compressors. The valve lift (plate opening distance) for steel plate valves was limited by a number of factors including impact forces during opening and closing events where two hard and rigid surfaces collide. The high cost of steel valve plates also made them increasingly unat-tractive to compressor operators. Another problem was that when the steel plates failed, broken pieces could damage the valve seat and render it unrepairable.

While steel valve sealing elements (plates, rings, chan-nels, strips) had served as workhorses for valve plate tech-nology since the early 1900s, in the late 1970s and early 1980s (then) relatively new nonmetallic valve plates, were proving to be much better and more reliable.

The nonmetallic valve plates not only cost less, they were more resistant to fatigue and could accommodate higher lifts that increased the valve flow area and reduced valve losses. And the likelihood of severe valve seat damage was much less in the event of failure of the nonmetallic plates.

U.S. reciprocating compressor manufacturers were be-ginning to embrace the separable compressor concept and were looking to push from the “high speed” of 800 rpm to unheard of speeds as high as 1800 rpm to match the emerging gas engines. This new requirement pushed the

Hoerbiger R&D group in Vienna to investigate alternative valve designs, taking advantage of the new nonmetallic valve plate materials.

The result was a compressor valve concept never before seen by the reciprocating compressor industry. The valve, designated CT, was designed from the ground up. It was never intended to be used with steel sealing elements and because of this the valve designers were able to offer sig-nificant increases in “valve lift area.” The new design used a thin 0.079 in. (2 mm) polyether ether ketone (PEEK) plastic plate with a steel cushion plate and a wafer spring that also helped eliminate the phenomenon then recently rec-ognized as stiction.

Stiction is a phenomenon that can occur in lubricated compressors when the valve plate comes into contact with either the seat lands or the guard and where a film of lubri-cation can cause a delay of the opening or closing event of the valve plate. This delay can result in the premature failure of the valve plate due to higher impact forces. The wafer spring, fitted between the guard and the steel cush-ion plate in the CT valve, worked beautifully in breaking the stiction effect.

The CT valve had another very unique feature. Rather than having the closing springs dispersed throughout the valve guard, all the springs were arranged very close to the outside diameter of the valve guard. This allowed the valve plate to deflect from the inside out, keeping the valve plate in a flatter opening position and stabilizing the opening impact as the outside diameter of the valve plate tended to see the highest opening impact. The design lessened these typical high opening impacts, al-lowing the valve plate to last longer under extreme oper-ating conditions, especially in the new short-stroke, high-speed compressors.

Few people understood the impact that this breakthrough would have on the compressor industry. Hoerbiger put some out on test and sold a few. Cooper was developing a new 1200 to 1500 rpm Superior RAM compressor in 1985 and first applied the new CT valve. At about the same time Ariel, which was developing ever-larger compressors, started applying them in its JGR and then the newer JGK models. Soon the valves were being used in Gemini and EI compressors as well.

A number of developments occurred as the CT valve evolved over the years. Plate thickness was increased to 0.157 to 0.236 in. (4 to 6 mm), and the steel cushion plate was eliminated. PEEK was brittle in cold weather start-ups, so nylon plates were offered. Soon, taking advantage of new material technologies, a special MT material was developed that was closer to the strength and high tem-perature capability of PEEK and the low temperature flexibility of nylon.

Truly a game changer that enabled the successful appli-cation of modern high-speed separable compressors, the CT valve remains the “heart” of most natural gas compres-sors operating today. It is produced in a range of sizes and used in most gases, including sour gas. CT2

CT Valve Application Limits

Size Range (Relates to diameter of the seat land area)

2.65 to 8.70 in.

65 to 221 mm

Maximum operating pressure 5075 psig 350 bar

Maximum differential pressure 2175 psi 150 bar

Maximum lift 0.11 in. 2.8 mm

Maximum speed 2000 rpm

Cornerstones Of Compression story continued from page 88

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C ornerstones Of Compression

march 2014 88 compressortech2

In 1895, Hanns Hörbiger developed an innovative valve for reciprocating compressors. His steel plate valve over-came all the disadvantages of other common valve de-

signs at the time.In 1900, Hörbiger, together with engineer Friedrich Wil-

helm Rogler, founded an engineering office in Budapest, which they relocated to Vienna in 1903.

Hanns Hörbiger devoted himself to continually enhanc-ing the steel plate valve he had invented and patented. His office issued licenses for use of the technology to business partners domestically and abroad. In 1925, the engineering office became Hoerbiger & Co. trading company. In 1931, Alfred Hörbiger, oldest son of Hanns, started the in-house production of valves.

Between 1925 and 1945, 171 patents were granted for Hoerbiger & Co. for inventions and developments in the field of compressor valves and controls. After the destruction of the Vienna production plant during the last year of World War II and the sudden passing of Alfred Hörbiger, his wife, Martina, managed to rebuild the plant.

Martina Hörbiger was a take-charge entrepreneur who oversaw the development of a worldwide company over many decades. Hoerbiger Co. expanded sales throughout much of the world and by the end of the 1950s had developed its own distribution system in North America. Hoerbiger Corp. of America (HCA) was founded in 1963 and the development of a production operation followed, led by Hubert Wagner.

The Heart Of The High-Speed Recip >

Subhead line subhead lineor to a second or third lineor even to a third line

Hoerbiger CT valve was a game changerBy NoRM SHAde

n The CT valve is a critical component for tens of thousands of high-speed separable compressors operating today.

continued on page 87

CT344.indd 1 2/24/14 10:35 AM

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