*Corresponding author. Copyright © 2016 by Northumbria University at Newcastle upon Tyne and ISEC International Stirling Engine Committee. All right reserved.

Upgrading of Stirling Engine Dynamic Seals – Swedish Development since 40 years

Andreas BAUMUELLER a *, F. Xavier BORRASb

Per ESKILSON c,Martin NILSSON d, Andreas VERNERe Cleanergy AB, Forsbrogatan 4, 662 34 Åmål, Sweden aandreas.baumueller@cleanergy.com bxavier.borras@cleanergy.com cper.eskilson@cleanergy.com dmartin.nilsson@cleanergy.com eandreas.verner@cleanergy.com

Topic/s: Area 4 “Components and systems” Keywords: Stirling history, piston rod seal, dry piston ring

ABSTRACT

The first Stirling cycle hot air engines worked with gas pressures close to atmospheric conditions. The design approach at Phillips raised working pressures significantly - up to 20 MPa; these Stirlings performed comparable to Otto and Diesel engines in power per weight. In modern Stirling concepts, suitable working gas pressures range between 10 and 20 MPa in order to obtain high work output. It is desirable to keep the pressurized volume low to allow power control by pressure variation and limit risk. As a consequence of using unpressurized crankcases, a rod seal separating the working space from the crankcase is needed which has been a hard nut to crack for decades.

Stirling piston rings require self-lubricating materials which wear out and define the ser-vice interval for replacement; they do not allow any lubrication as the wear rate would in-crease with oil.

Kockums/USAB in Sweden continued with the Philips concept, but went from the her-

metic roll-sock to a sliding seal between working space and crankcase. After a couple of steps, the “Pumping Leningrader” (PL) seal allowed good levels of tightness. Cleanergy is still using this type of piston rod seals.

Cleanergy started with the basic design and materials selected in the 80-ies and it was time to upgrade the seals technology of the 160 cc V-twin alpha engine. Even though the gas leak through the rod seal is low, the oil- migration into the cylinder is determining the limits of pis-ton ring life.

This paper is going to describe in the following the project which aims to increase service

interval of the seals. An analysis of forces and movements, alignments and clearances was made, temperatures were measured. The sealing gap was studied using an elasto- hydro-dynamic model. This resulted in a first implementation of improvements to the piston rod seal. Further investigations could quantify friction forces and compare the oil migration for

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different settings of mean working pressure and shaft speed. Simulations advanced from the linear elastic material model to a visco- plastic material model. The new model allowed simu-lating short-time elastic response plus the long-time visco- plastic deformation of seals mate-rials.

There has been an impressive advance in materials science for dry-lubricated piston rings in the last years. 16 alternative compounds for dry operation were compared in an oscillating rig in He and H2 environment, and the best choices were run in real engines for validation.

The intermediate result is a 50% increase of the service interval of the seal and another step in the same order can be expected within a short time. Cleanergy expects service inter-vals far over 10.000 operating hours due to careful modifications of a proven concept. 1. INTRODUCTION

It is said that the intention of Robert Stirling was to create a power source with lower risk of explosion than steam engines and their vessels allowed in those days. Hence the first Stir-ling- cycle hot- air engines worked with gas pressures close to atmosphere, and therefore no advanced sealing technology was needed. But, when the internal combustion engines were developed towards the end of the 19th century, hot air machines could not compete in power/ weight- ratio and disappeared for a while.

The new design approach at Philips raised pressures significantly up to 22 MPa, the light

gases helium and hydrogen were used; these Stirlings performed at least comparable to Otto and Diesel engines in power per weight and per volume. Even higher pressures were dis-cussed but dry running piston rings and rod seals at the time set limits on the upper end.

In modern Stirling engines, suitable working gas pressures range between 5 and 20 MPa in order to balance high work output and length of service intervals. At this level it is desirable to keep the pressurized volume low and reduce safety risks; therefore a crankcase with atmos-pheric pressure is preferred. Also the power control by pressure variation sets limits for the working gas volumes which can be supplied and discharged. The separation also allows the use of proven oil lubricated journal bearings. Last but not least, crankcases without pressure load save significant material and costs.

Figure 1. Early versions of “Leningrader” and “Pumping Leningrader” piston rod seal

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As a consequence of using atmospheric pressure in the crankcases, a rod seal is needed to

maintain the high pressure in the working space which has been a hard nut to crack for dec-ades.

Kockums/USAB in Sweden continued working with the Phillips concept but went from the hermetic roll-sock to a sliding seal between working space and crankcase, called “Lenin-grader seal” with a reference to the first publication. A packing, made from a PTFE com-pound seals against the oscillating rod. The seal was lubricated and cooled with oil spray on the piston rod which required a secondary scraper seal to remove the oil on the working gas side and a third seal on the dry section of the rod; still not a completely satisfying concept.

As a next step, the “Pumping Leningrader” (PL) seal got a slightly tapered inside which enabled it to pump the lubrication oil back to the crankcase. This concept provided good lev-els of tightness and lifetime [1, 2].

The Swedish development lines were split into high performance double acting four cyl-

inder engines (V4X, 4-95, 4-275, V4 275) and a cost efficient V twin alfa- engine with one cycle, which was much easier to produce and to maintain. This “SPS V160” with 160 cc swept volume used PL - seals and demonstrated its high service time potential in the 1980s. Solo in Germany redesigned the engine in the 1990s to the Stirling 161 with 9 kW@ 1500 rpm but kept all proven details without significant modifications, including the rod seal and the piston rings [3]. Cleanergy continued the development in 2010 with an improved version of the Stirling 161. A gas heated and a solar powered version is being produced. With over 2.5 million accumulated operating hours, this is still the best proven engine in the 10 kW segment.

The company is currently using the rod seal in quite a compact design of piston rod seal assembly. It is one of the most important components which also represents a key technology.

Figure 2. Cleanergy’s Stirling engine, rod- and piston seals

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Another specific sealing technology in Stirling engines are piston rings, requiring self-

lubricating materials which wear over time and thus define the service interval; they do not allow any lubrication as wear rate would increase; oil in the cycle would also carbonize. Thus, the service- time for piston rings depends from oil migration through the piston rod seal too.

Even though gas leaks through the rod seal around 0,1 lN/h (normalised conditions) are not

of major concern, any oil migration into the cylinder is lowering piston ring life. This aspect is today in the focus of improvement of piston rod seals. The original design and materials of the dynamic seals were selected around 1980 and the ad-vances in material science and design experience in the compressor industry prompted a de-velopment effort of the piston rings. 2. OPERATING CONDITIONS FOR SEALS IN THE CLEANERGY STIRLING

ENGINE AND IMPROVEMENTS To create a database about working conditions for piston rings and rod seals, the pressures

in the cycle and in the buffer space were measured with piezo- sensors. This Vinnova project at Chalmers University in Gothenburg was also base for Cleanergy’s Stirling cycle simulation SQUID.

From the pressures, indicated and effective power could be deducted and the friction loss-

es determined. An analysis of forces and was performed, alignments and clearances were used to study lat-eral movements.

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0 15 30 45 60 75 90 105 120 135 150 165 180 195 210 225 240 255 270 285 300 315 330 345 360

Gas pre

ssure (

bar)

Crankshaft rotation (degrees)

Pressures on the expansion and compression pistons

P_Exp (bar)P_BufExp (bar)P_Com (bar)P_BufCom (bar)

Figure 3. Pressure variation in cycle (pExp and pCom) and below pistons (pBuf) for average helium pressure 15 MPa, top temperature 625 °C, coolant out 50°C, speed 1500 rpm

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As the piston rod seal is installed between the buffer space (which is a volume below the

pistons) and the crankcase, the pressure to be sealed is between 14 and 16 MPa at full power. We can also see that pressures in the cycle peak 21 MPa max and 10,5 MPa min. Result is that the pistons and rings are loaded with 7 MPa in the down stroke and 5,5 MPa in the up-stroke.

Temperatures were measured with irreversible strips and thermocouples on different posi-tions in a gasheated engine. Piston rings work over 100 °C and rod seals were close to this value. Certainly there are impacts from working gas pressure, speed, oil- and water tempera-ture.

The measured data was used to build a model for the conductive heat flow. Radiation is difficult to calculate or measure but simplified estimations could be done. Cleanergy con-cludes that convection is the most important heat transfer through the piston dome.

Figure 4. Temperatures measured and conductive heat flow in expansion piston

As the openly available theoretical knowledge about rod seals with high pressure on the dry side was limited, [4,5], Lulea’s Technical University (LUT) was engaged to study the sealing gap using an elasto- hydrodynamic model, i.e. a combination of Reynolds equations for the hydrodynamic contact between seal and rod and a linear elastic FEM analysis for the seal.

Figure 5. Simulation model and result for hydrodynamic pressure in the outstroke

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The hydrodynamic pressure in the rod seal peaking 25 MPa is quite high for a PTFE based

material but such materials deform plastically to release stress. These studies resulted in a first implementation of improvements of the piston rod seal. The material was changed to a more pressure resistant PTFE compound and temperatures could be lowered some 20 K. 3. INVESTIGATIONS ABOUT OIL MIGRATION To compare the oil migration for different settings of mean working pressure, tempera-tures and shaft speeds, white oil- indicator paint was applied on washers attached on top of the piston rod seal housing and on the pistons bottom side to quantify small amounts of oil by a colour change. Relatively short test runs in a driven rig provided a variety of results which allowed optimisation of the operation conditions and some design changes.

Figure 6. Oil detection washers on rod seal housing and piston, results over temperature Together with an improved control of production tolerances, service times could be expanded some 50% and the expected service life is in excess of 6000 h in gas fired units. Solar units in the field surpassed service intervals longer than 3200h, tests are ongoing and the upper limit is yet to be seen.

Piston Seal housing

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4. FRICTION FORCES ON THE PISTON ROD SEAL

Further investigations could quantify friction forces by measuring them with strain gauges applied on the taper ring, the part supporting the rod seal. The friction losses for both seals range around 250 W, approximately 10% of the total friction losses.

Figure 7. Strain gauges applied to measure friction forces and results

5. VISCO - PLASTIC – MODELLING FEM simulations use normally linear elastic material models which is adequate for metals,

but plastic materials behave very differently.

Figure 8. Time- and temperature- dependent strain response of a PTFE with fillers

and FEM analysis for displacement of the rod seal The strain response on stress variations depends typically from strain rate (stress change

per time), temperature and direction (tension and compression, homogeneity). FEM simula-tion for a PTFE material needs this information to produce useful results. Thus, such data

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were measured for relevant materials and the new model allowed simulating short time elastic response plus the long time viscoplastic deformation of seal materials.

Based on such tools, the classic Swedish “Pumping Leningrader” seal can still be opti-mised in several details to provide improvements.

Simultaneously, Cleanergy works in development of new concepts of rod seals.

6. DRY RUNNING MATERIAL IMPROVEMENTS The development of materials for piston rings has advanced materially over the last 30

years. This is valid for polymer materials in general and specifically for materials used in dry running compressors for hydrogen and helium. After careful analysis of the materials and suppliers available on the market, 16 alternative compounds for dry operation were tested in an oscillating rig in hydrogen and helium and Hydrogen atmosphere. The material samples were pressed against a plate and moved up and down with application realistic speed and pressure. Friction forces and wear can be measured. It turned out that friction sliding force can differ between materials by a factor 10 and wear even more. The reference material, which is the polymer- in- use can most certainly be beaten in wear resistance.

Figure 9. Oscillation test rig and results for piston ring materials, friction and wear

7. RESULTS

When Cleanergy started to build engines, the expected time between services was 4000 operating hours for gas engines and one year (approx. 2000h) for solar units. Based on the analysis described above, service intervals now increased by over 50%.

The next step is close to implementation: A lowered oil migration due to some modifica-tions on the piston rod seal and improved materials for piston rings and guides lead to expec-tations of yet again twice as long service intervals and a cost reduction per service.

The development will continue: Meanwhile, even better materials have been identified and advanced designs can be tested.

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8. CONCLUSIONS

The Swedish concept of Stirling engine design, introduced by United Stirling / Kockums and using crossheads and piston rod seals seems to be out of fashion at first glance. But in re-ality, it provides the highest performance, good power control options and has proven during the longest experience to secure outstanding service- and lifetime.

Even more, relatively small updates in materials and design allow nowadays significant additional steps in service time and operational costs without risking the achieved reliability. This is in focus of Cleanergy’s development.

Similar to internal combustion engines, it takes decades to fully optimise an engine gener-ation but in contrast to these, the Stirling is still a relatively unexplored technology.

The Stirling technology is still young and the potential is large. It is mostly a question of

optimisation to become successful 200 years after invention.

REFERENCES [1] Lundholm, G., Lund University: The Experimental V4X Stirling Engine – A Pioneeing

Development, ISEC 2003 in Rome [2] Lundholm, G.,United Stirling AB: Statistical Life-Testing of Stirling Engine Main Seals

(B4ASESS9TR29); SAE Proceedings of the 21st ATD CCM, Nov 14-17, Dearborn l983 [3] Lundholm, G.; Lundström, L.; Schiel, W.; Baumueller, A.: Development History of the

V160 and Solo Stirling 161 Engines; ISEC 1999 in South Africa [4] Nikas, G.K.: Eighty years of research on hydraulic reciprocating seals: review of tribolog-

ical studies and related topics since the 1930s; Mechanical Engineering Department, Im-perial College London, Exhibition Road, London SW7 2AZ, UK, 2008

[5] M.W. Euseip, J.A. Walowit, O. Pinkus, and P. Holmes: Performance of Oil Pumping Rings, An Analytical and Experimental Study; Mechanical Technology Incorporated, 1986