Cryogenic Tribology of High-Speed Bearings and Shaft Seals ...

Tribology Online, 6, 2 (2011) 133-141. ISSN 1881-2198

DOI 10.2474/trol.6.133

Review

Cryogenic Tribology of High-Speed Bearings and

Shaft Seals in Liquid Hydrogen

Masataka Nosaka*

Department of Mechanical Engineering, The University of Tokyo 7-3-1 Hongo, Bunkyo-ku, Tokyo 113-8656, Japan

*Corresponding author: [email protected]

( Manuscript received 30 November 2009; accepted 21 September 2010; published 31 January 2011 ) ( Presented at Minisymposium on Hydrogen Tribology for Future Energy at the World Tribology Congress 2009, Kyoto, 6-11 September, 2009 )

Liquid hydrogen (LH2) is excellent in extensive storage and transportation systems, such as being used as fuel in highly efficient liquid-rocket engines. To consider the use of LH2 required for future hydrogen-energy systems, this paper presents a topical review of previous cryogenic tribology studies on the research and development of the bearings and shaft seals for LH2 turbopumps. Cryogenic tribology studies were conducted for the LE-5/LE-7 rocket engines of the Japanese H-2 rocket as well as for an advanced ultra-high-speed LH2 turbopump. The tribo-chemical formation of CaF2/FeF2 film due to the reduction power of LH2 showed excellent self-lubrication, resulting in a sound wear condition within the turbopump bearing. A new hybrid ceramic bearing with Si3N4 balls, which bearing had a single-guided retainer, tested excellently at an ultra-high speed of 120,000 rpm (3 million DN) in LH2. Seal performance for the floating-ring seals using Ag-plated metal seal-rings at ultra-high speeds was also tested.

Keywords: cryogenic tribology, solid lubrication, PTFE, ball bearing, hybrid ceramic bearing, shaft seal, mechanical seal, floating-ring seal, liquid hydrogen, liquid oxygen

1. Introduction

As earth scale issues of energy conservation and environment preservation, such as the exhaustion of petroleum resources and the CO2 reduction to prevent global warming, prevail nowadays; a breakaway from the conventional fossil-fuel society becomes a big problem. Clean hydrogen energy is attractive due to its energy efficiency and its smaller impact on the environment, and it is expected to be a key technology in the 21 st century.

Liquid hydrogen (LH2) shows excellent properties in extensive storage and transportation systems, and is currently being used as fuel in highly efficient liquid-rocket engines. For LH2 turbopumps used in liquid oxygen (LO2)/LH2 rocket engines, many cryogenic tribo-components, such as self-lubricated bearings and shaft seals which must sustain high-speed rotor activity in extreme conditions, have been researched and developed. These turbopump bearings and shaft seals, which were used in the LE-5/LE-7 rocket engines of the Japanese H-2 rocket as well as in an advanced ultra- high-speed LH2 turbopump required for future space transportations, have been under

development since the mid-1970’s. To consider the use of LH2 technology required for

future hydrogen-energy systems, previous studies on cryogenic tribology in the research and development of the bearings and shaft seals for LH2 turbopumps are reviewed and discussed. It is proposed that, to build hydrogen infrastructure for LH2 storage and distribution, development of an industrial tribo-system with long durability and high reliability is essential and advances by using cryogenic tribology studied for LH2 rocket system.

2. Turbopumps and cryogenic tribo-components

An LO2/LH2 pump-fed engine (LE-7) with 980 kN-thrust for the first stage of the H-2 rocket had been developed. Because this LE-7 engine had a staged- combustion cycle with high chamber pressure, the engine required high-pressure LO2/LH2 turbopumps, which delivered high-pressure propellants from the propellant tanks to the combustion chamber.

Figure 1 shows the high pressure LH2 turbppump of the LE-7 engine1). The pump discharge pressure was 27 MPa, the discharge LH2 flow rate was 510 liters/s, and the shaft power was 19,700 kW. The 35 mm bore paired

Copyright © 2011 Japanese Society of Tribologists 133

Masataka Nosaka

bearings were at the inducer side, and 40 mm bore paired bearings were at the turbine side. These bearings were all-steel bearings (SUS 440C) which were self- lubricated by the transfer film from the reinforced PTFE (polytetrafluoroethylene) composite retainer. Both pairs, cooled by circulating LH2 in the turbopump, had to operate at a speed of 42,000 rpm which was beyond the third critical speed of 32,000 rpm and had to support a high shaft-power in LH2.

In actual LE-7 LH2 turbopump operation, conventional bearings using a retainer with circular pockets showed a significant temperature rise. As the radial load increased, ball excursion occurring in the retainer pocket due to ball-speed-variation (BSV) was significantly large. For the 40-mm-bore bearing, a retainer having elliptical pockets with a large pocket clearance was developed to allow maximum ball excursion due to BSV and to stabilize wobbling of the retainer. Consequently, the turbopump bearings with the elliptical-pocket retainer exhibited excellent performance by reducing severe frictional heating and high wear of bearing components at a high-speed level of 50,000 rpm (2 million DN)1).

Fig. 1 High pressure LH2 turbpump of LE-71)

Figure 2 shows the high pressure LO2 turbopump of the LE-7 engine2). The discharge pressure was 18 MPa for the main pump and 26 MPa for the preburner pump, respectively. The total discharge LO2 flow rate was 240 liters/s, and the shaft power was 4,700 kW. The 32 mm bore paired bearings were located at the inducer side and the 45 mm bore paired bearings were at the turbine side. The paired ball bearings sustaining the rotor operated at a speed of 18,000 rpm in LO2.

It has been noted that violent frictional heating in LO2 can lead to the ignition of tribo-elements due to burn-out phenomenon (overheat occurring in a transition from nucleate boiling to film boiling at critical heat flux), which is defined by engineering heat transfer.

For the turbopump bearings, the durability and fatigue life were evaluated by applying heavy radial loads at a speed of 20,000 rpm for about 2.2 hours in LO2

2). In the high pressure LO2 turbopump, the shaft seal

system was set up between the cryogenic pumps and the hot turbine to prevent the mixing of the leakage of LO2 and the high-pressure turbine hot gas. The shaft seal system of the high pressure LO2 turbopump of the LE-7 engine is shown in Fig. 33). The LO2 seal was composed of a floating-ring seal with a carbon seal ring and a wear-ring seal made of hard plastic material (Kel-F). The turbine gas seal used two floating-ring seals to seal the low temperature hydrogen gas (GH2), which made a barrier to the turbine hot gas, so that the turbine gas seal was kept at a lower temperature against the hot turbine section and the reliability of the shaft-seal system was further increased.

Between the LO2 seal and the turbine gas seal, the segmented circumferential seal, which had shrouded Rayleigh step hydrodynamic lift-pads to increase opening force, was paired and purged with helium gas to prevent mixing of the leakage of LO2 and GH2. The segmented carbon seal-rings were coated with a sprayed MoS2 film to reduce the carbon wear at low temperatures3).

Fig. 2 High pressure LO2 turbpump of LE-72)

Fig. 3 Shaft seal system of LO2 turbopump of LE-73)

Japanese Society of Tribologists (http://www.tribology.jp/) Tribology Online, Vol. 6, No. 2 (2011) / 134

Cryogenic Tribology of High-Speed Bearings and Shaft Seals in Liquid Hydrogen


Fig. 4 Tribo-components and solid lubricants used in turbopumps4)

Tribo-components and solid lubricants used in the turbopumps are topically indicated in Fig. 44). High-speed components, such as bearings, shaft seals, Labyrinth seals, wear rings and balance pistons, used the proper solid lubricants to protect them from severe adhesive wear in the oxidation or reduction environment of the cryogenic propellants.

3. Cryogenic tribology problems

LH2 is a particularly poor lubricant due to its extremely low viscosity and chemical reducing effect, resulting in a severe lubricating condition. It has been noted that lower hydrodynamic force and fresh frictional surface result in tribological problems at the frictional interfaces. Furthermore, at extremely low temperatures in LH2, the specific heats and thermal conductivities of tribo-materials drop off rapidly, and it is easy to produce local hot spots at frictional interfaces. Consequently, frictional conditions would be more severe in addition to the vaporization of LH2

5). Under burn-out phenomenon, an extreme rise in

surface temperature was experienced because a marked reduction occurred in heat transfer during in a transition from nucleate boiling to film boiling. For example, in liquid nitrogen (LN2), the sliding surface of Ag - 10%Cu alloy pin (melting point was most low, 1,155 K) against the Ti alloy disk melted due to burn-out wear, as shown in Fig. 55). The pin-on-disk test was conducted with various pins (3 mm diameter) at a sliding speed of 1 m/s under a load of 98 N. The surface coating of TiN or TiO2 to the Ti alloy disk, which had a high resistance to adhesive welding, was able to protect the Ag - 10%Cu pin from burn-out wear. The results were applied to the balance-piston system in the LE-7 LH2 turbopump.

It has also been noted that the tribo-characteristics at

cryogenic temperatures tend to complexly change. For example, Fig. 6 shows the change of friction and wear of PTFE as a function of pin temperature in gaseous nitrogen (GN2)

2). This figure denotes the glass transition temperature of PTFE (about 170 K, 230 K and 260 K), which are defined by its amorphous structure. Thus, it is very interesting that the friction and wear properties of PTFE change characteristically at its glass transition temperature.

Sliding direction

(a) Ag pin (melting point, 1,234K) (b) Ag-7.5%Cu pin (1,173K)

At cryogenic temperatures, sufficient cooling and restriction of frictional heat generation are essential. Therefore, to solve these cryogenic tribology problems, (1) understanding the complex characteristics of tribology at low temperatures, (2) selection of the proper solid-lubricants against the reduction or oxidation power, and (3) active cooling to remove

(c) Ag-10%Cu pin (1,155K) (d) Cu pin (1,356K)

Fig. 5 Wear of various pins in LN25)

Masataka Nosaka

severe frictional heat at local hot spots are very important.

Fig. 6 Friction and wear of PTFE in cryogenic GN22)

4. High-speed bearings

The turbopump bearings are self-lubricated by the transfer film from the retainer which is made of solid- lubricants such as PTFE. For the rocket-turbopump bearings, a laminated glass cloth with PTFE binder was used because of its mechanical strength to protect against dangerous retainer rupture. Figure 7 shows the glass cloth-reinforced PTFE retainer (glass cloth, 45 mass% and PTFE, 55 mass%)6). The bearing operating in LH2 showed unstable high-temperature rise and poor lubrication was observed due to the abrasive glass cloth, resulting in severe wear of the balls. Average ball wear (of 25 mm bore bearing tested for 62 minutes at 50,000 rpm) was about 7.8 μm6).

Fig. 7 Glass cloth-reinforced PTFE retainer

To improve the self-lubricating performance of the retainer, the abrasive retainer surfaces with the exposed glass cloth, which had a depth of 0.10 - 0.15 mm, were

chemically etched with hydrofluoric acid (HF), as shown in Fig. 86). Following this treatment, smooth surfaces for the retainer were obtained. The sliding friction and wear between the ball and ball-pocket surfaces was reduced, resulting in a sufficient supply of PTFE transfer film from the retainer to the rolling balls.

Fig. 8 Chemical etching of retainer with HF

Fig. 9 XPS depth analysis for ball of LH2 bearing (Etching depth was 30 nm by SiO2 rate)

Figure 9 shows the XPS (X-ray photoelectron

spectroscopic) depth analysis for the ball taken from the LH2 turbopump bearing used in LE-78). This bearing was operated in LH2 for the engine test time of 34.4 minutes with 20 start-stop times7). F(1s) and Fe(2p) spectra show the significant formation of thick CaF2 and FeF2 transfer film. It is interesting to note that, due to the reduction power of LH2, the reacted oxide material (CaO) of etched glass fiber was tribo-chemically changed to CaF2 with the F of PTFE during bearing

Glass cloth

HF treatment (Etched depth: 0.10-0.15 mm)

PTFE Glass Cloth

Reacted material of etched glass fiber



operation. The reacted material remained on the etched retainer surface as indicated in Fig. 86). A large amount of this material was formed (49 mass% of glass fiber) and consisted mainly of an oxide of Ca2). Furthermore, a FeF2 transfer film was formed by a chemical reaction between the F of PTFE and the Fe of bearing steel (SUS 440C)8). On the contrary, for the LO2 turbopump bearings, the oxidization power of LO2 prohibited the tribo-chemical formation of CaF2 /FeF2 transfer film8).

Thus, the LH2 turbopump bearings used in the LE-7 rocket engine demonstrated excellent performance due to the formation of thick CaF2 and FeF2 transfer film. The tribo-chemical formation of CaF2/FeF2 in transfer film possibly reduced wear at frictional interfaces within the bearings (both all-steel and hybrid ceramic bearings) used in LH2

8). The basic tribo-chemical reaction seemed to be determined as follows: (-CF2-)n + CaO + Fe → (-CF2-CO-)n + CaF2 + FeF2 (1)

Fig. 10 Wear of PTFE composite with various fillers2)

To determine the effect of tribo-chemical formation of CaF2 in transfer film, as shown in Fig. 9, additional friction tests were conducted. Figure 10 shows the wear of PTFE composite with various fillers (15 mass%) in cryogenic oxygen gas (GO2, 123 K) under a high sliding speed of 10 m/s2). The PTFE composites with CaO and MgO fillers showed excellent wear resistance (progression of the pin-wear was stopped) due to the formation of good transfer film even in both cryogenic GO2 and GN2 (123 K), because alkali-earth-metals such as Ca and Mg were able to react easily with F, resulting in the formation of CaF2 and MgF2 within the transfer film. The tribo-chemical formation of CaF2 and MgF2 might enhance adhesion of transfer film. When CaF2 and MgF2 added as fillers to PTFE, there was no tribo-chemical reaction, resulting in poor wear resistance. Furthermore, oxidation of the Mo filler in GO2 seemed to be extremely effective except in GN2

2). In another test, the lubrication of fluorine-passivated

bearings (single-guided, 25 mm bore) with FeF2 layer was estimated in LO2, LN2 and LH2

9). The FeF2 layer (thickness, 50 nm) of SUS 440C bearings was formed in hot pure F2 gas (327 K) at a pressure of 0.1 MPa for two hours. The fluorine-passivated bearings showed excellent self-lubrication in both LH2 and LN2; however, poor lubrication was seen in LO2. These results indicated that excellent lubrication in LH2 depended on the formation of the FeF2 layer due to tribo-chemical reaction.

Under sufficient cooling conditions using jet cooling in LO2, untreated bearings exhibited excellent self- lubrication and durability. It was observed that sound surface conditions with hardly any wear were capable for a total test time up to 11.7 hours at a speed of 50,000 rpm, as a result of the intense formation of a Cr2O3 layer rather than the bearing before testing. The Cr2O3 layer provided an increased resistance to metal-to-metal adhesion beneath an extremely thin CaF2 film

9).

Limit A: bearing torque and temperature increased. Limit B: local film rupture occurred.

Fig. 11 Load capacity of transfer film in LH2

1)

The load capacity of the transfer film under inner race ball-spinning in LH2 was determined as shown in Fig. 111). This figure shows the critical load capacity, that is, maximum Hertz stress (Smax) vs. maximum spinning speed (Vmax). Under high thrust loads, an increasing of the bearing torque and bearing temperature (limit A) was determined by the bearing tester. The film local rupture (limit B) was also defined by the electrical resistance monitoring between the inner race and outer race.

Since the bearing was under the outer-race ball control at high speed, the transfer film of the inner raceway was damaged due to the spinning of the ball.


Masataka Nosaka

Up to a Vmax of 5 m/s at 50,000 rpm, the transfer film was able to sustain a Smax up to 2 GPa. It was determined that the load capacity of the transfer film depended more on Smax than on Vmax. So, in order to increase durability of the bearing, it is important to limit the stress level to a Smax of 2 GPa to prevent transfer-film rupture and sufficiently to cool the frictional heat due to high Vmax

1).

5. Ultra-high-speed bearings

Advanced rocket engines required for future high- performance space transportation would need higher- speed (ultra-high-speed) turbopumps operating at a level of 100,000 rpm. Modified hybrid ceramic bearings with Si3N4 balls (inner-outer races, SUS 440C) were studied and showed excellent performance even at 3 million DN (120,000 rpm) in LH2

10). Research results of hybrid ceramic bearings were applied to the LH2 turbopump of the RL-60 demonstrator engine（thrust power, 270 kN). The RL-60 demo-engine was developed in USA with international collaboration (USA, Japan, Russia and Sweden) and the LH2 turbopump (rotational speed, 90,000 rpm) was developed in Japan11）.

The modified bearing had a single-guided retainer with elliptical ball pockets, as shown in Fig. 1212). The single-guided retainer was guided only on one side of the outer-ring bore (land) to reduce land friction and to increase the cooling ability within the bearing. For the elliptical ball pocket of the single-guided retainer, the circumferential clearance (1.2 - 1.3 mm) was twice as large as that of the conventional circular pocket to reduce ball-to-pocket interaction under high BSV. Furthermore, the axial clearance (0.1 - 0.2 mm) was narrow to stabilize wobbling of the single-guided retainer at high speeds12).

Fig. 12 Modified bearing with single-guided retainer12) Performance of the retainer using elliptical ball

pocket was previously studied to develop the LE-7 LH2 turbopump bearings (40 mm bore) with a large clearance (1.8 mm) at a speed of 50,000 rpm1). The retainer using elliptical ball pocket with a more large

clearance (2.2 mm) was able to prevent the retainer fracture due to high BSV under high radial load as well as to high tilted misalignment (in a level of 1.9 - 3.5 × 10-3 mm/mm) caused by inclination of the outer race13).

Figure 13 shows the performance of single-guided bearings (hybrid ceramic bearing and all-steel bearing) at up to a speed of 50,000 rpm under a thrust load of 784 N in LH2

12). It was observed that the bearing torque of the single-guided bearing decreased to about one-half of that of the conventional double-guided bearing. The hybrid ceramic bearing showed lower bearing torque than the all-steel bearing at high speeds.

For the single-guided hybrid ceramic bearing, the critical thrust load (1,960 N) with a Smax of 2.7 GPa, without a significant rise of the bearing torque and bearing temperature at a speed of 50,000 rpm, was two times higher than that of the conventional double-guided all-steel bearing. Furthermore, even under high bearing torque with high-temperature rise, the hybrid ceramic bearing was able to sustain a thrust load of 2,840 N (Smax, 3.2 GPa) at a speed of 50,000 rpm without seizure in LH2

12). It was found that a considerably thick transfer film consisting of CaF2/FeF2 was tribo-chemically formed for the hybrid ceramic bearing8).

Fig. 13 Bearing torque of single-guided bearings12)

Based on previous bearing tests at high speeds up to

50,000 rpm8,12), the hybrid ceramic bearing was tested at ultra-high-speeds up to 120,000 rpm, and results were compared with the all-steel bearing10). Figure 14 shows the change of the bearing temperature at a steady speed of 120,000 rpm with a thrust load of 2,160 N10). The hybrid ceramic bearing showed excellent performance at high thrust load, compared to the seized all-steel bearing. At a higher thrust load of 3,140 N, the hybrid bearing showed slight damage with a spiky rise of the bearing temperature. It was found that the critical load capacity (Smax) without seizure at a speed of 120,000 rpm was 3.0 GPa (thrust load, 2,160 N) for the hybrid ceramic



bearing and 2.0 GPa (980 N) for the all-steel bearing, respectively.

The components of the hybrid ceramic bearing were in excellent condition with regard to wear. On the contrary, the seized all-steel bearing exhibited severe adhesive wear. For the ceramic balls tested under thrust loads up to 3,140 N, superficial micro-cracks which extended visually like meshes of a net, as shown in Fig. 15, were observed at the contact track10). A marked feature of these superficial micro-cracks was that they were very shallow (about 3 μm at maximum) and did not extend deeply into the ball.

Fig. 14 Change of bearing temperature at 120,000 rpm10)

Fig. 15 Superficial micro-cracks on Si3N4 ball10)

An advanced study was conducted to select a tough Si3N4 ball capable of restraining crack propagation as well as to evaluate the efficient bearing cooling with nozzles14). A Si3N4 ball having higher thermal-shock resistance, as well as higher fracture toughness, was found to reduce the propagation of superficial micro- cracks, resulting in a decrease of ball wear. Furthermore, it was observed that the cooling ability of the LH2 jet- flow aimed at the retainer was superior to that aimed at the inner raceway, further reducing the propagation of thermal micro-cracks on the Si3N4 balls. Under the same cooling rate, the four nozzles achieved a higher cooling ability than the two nozzles with increasing jet speed above 208 m/s, which jet-speed reached to twice the

sliding speed (108 m/s) at the retainer outer-land. It was indicated that micro-cracks on the balls were

possibly generated at the trace contacting the outer raceway under insufficient cooling conditions. Decreasing the maximum outer-race stress to 2.0 GPa (thrust load, 1,960 N and Smax, 2.9 GPa at inner race) at a speed of 120,000 rpm by the limited outer-race curvature of 0.51 could prevent the formation of superficial micro-cracks even under insufficient cooling conditions14).

Fig. 16 Face-contact mechanical seal15)

6. Ultra-high-speed shaft seals

For the LH2 shaft seals, the contact-type mechanical seals and the noncontact-type floating-ring seals (annular seals) were developed and used in the LE-5 and LE-7 engines15). Figure 16 shows the face-contact mechanical seal (seal diameter, 43.2 mm) developed for the LE-5 LH2 turbopump, which has a modified seal nose that reduced the seal face distortion at low temperature and high pressure15). Furthermore, a modified vibration damper made of PTFE sheets is attached around the seal nose to prevent the fluttering of the seal nose during the rapid start or stop of the turbopump. For this mechanical seal, experimental and analytical studies on the friction power loss and dynamic sealing performance were conducted and showed an excellent durability (carbon-ring wear of maximum 8 μm) with an extremely small LH2 leakage rate (7.7 - 19.2 milliliters/ min) for 83 minutes at a speed of 50,000 rpm with a seal pressure of 1.37 MPaG16).

On the other hand, the floating-ring seal leaking in a large amount due to noncontact-type is suitable for high-pressure turbopumps; however, conventional seals using carbon seal-rings were weak under high speed and high pressure conditions. Since metal seal-rings have higher mechanical strength and durability, advanced floating-ring seals (one-seal and two-seal rings) which used Ag-plated metal seal-rings were studied at ultra- high-speeds up to 120,000 rpm in LH2

10). Figure 17 shows the configuration of a floating-ring seal consisting of two-seal rings with a seal diameter of 30 mm10). The test seal had an Ag-plated seal ring made of Ni alloy (Inconel 718), which was the same material


Masataka Nosaka

Fig. 17 Floating-ring seal with two metal-seal rings 10)

used for the runner. The runner was coated with a Cr2O3 plasma spray, and this coating exhibited excellent friction and wear without adhesion to Ag in LN2. To obtain smooth radial movement of the seal ring, the static seal surface of the housing was coated with a sprayed MoS2 film.

Figure 18 shows the seal performance of the one- ring seal vs. the two-ring seal10). These seals had a straight bore with a seal clearance of 110 - 120 μm. Figure 19 also shows the phase change models of leakage flow within the seal clearance. Seal performance depended on the two-phase flow (gas/liquid phase) of leakage, because the vaporization of leakage was generated by the viscous friction heat and by the seal pressure drop. At lower speeds, the leakage of the one- ring seal was relatively greater than that of the two-ring seal; however, with increasing speed, the leakage of the one-ring seal was drastically decreased and approached the same level of the two-ring seal due to enlargement of the two-phase flow.

For the two-ring seal, the two-phase flow was fully enlarged within the secondary seal ring which was at the downstream of the primary seal ring. Seal leakage was reduced within limits; however, the hydrodynamic force of the liquid phase flow which sustained the seal ring was lost and resulted in seal-ring seizure at a relatively lower speed of 98,700 rpm10).

Also, shaft vibration for the two-ring seal, which was likely produced by wobbling of the seal ring under severe rubbing conditions, abruptly increased at speeds of more than 92,000 rpm before resulting in seal-ring seizure at a speed of 98,700 rpm. Furthermore, in the two-ring seal with a seal clearance of 70 - 80 μm, the primary seal-ring seized a speed of 108,600 rpm, because the hydrostatic force decreased due to a low differential pressure10).

In contrast, the one-ring seal successfully functioned with no abnormal signs of seizure during tests, because the liquid-phase flow remained within a seal clearance even though the two-phase flow increased. As a result,

the hydrodynamic force in the liquid-phase flow as well as the hydrostatic force due to high differential pressure possibly helped to prevent seal-ring seizure.

At a steady speed of 120,000 rpm, the one-ring seal exhibited a stable leakage in a range of 0.21 - 0.24 liters/s, which is similar to leakage in the two-ring seal as shown in Fig. 18. Thus, the one-ring seal was superior to the two-ring seal, preventing seal-ring seizure due to an increase of two-phase flow within the sealing clearance10). These results indicated that more spring load was necessary for the static seal to prevent wobbling of the seal ring under the two-phase condition at ultra-high speeds.

Fig. 18 Seal performance of floating-ring seals10)

Fig. 19 Phase-change models of leakage flow

7. Concluding remarks

In order to consider the use of LH2 technology required for future hydrogen-energy systems, previous cryogenic tribology studies (since the mid-1970’s) on the research and development of the bearings and shaft seals for LH2 rocket-turbopumps are reviewed and discussed. Concluding remarks are as follows:




(1) Due to the LH2 reduction power, the reacted CaO of etched glass fiber was tribo-chemically changed to CaF2 in the transfer film. Furthermore, a FeF2 layer was formed. Such tribo-chemical formation of CaF2/FeF2 showed excellent solid-lubrication and reduced the bearing wear in LH2.

(2) The bearing torque of the single guided bearing decreased to one-half of that of the double-guided bearing. The single-guided hybrid ceramic bearing with Si3N4 balls showed excellent performance at an ultra-high speed of 120,000 rpm (3 million DN) in LH2.

(3) For the two-ring seal, at ultra-high speeds in LH2, the two-phase flow was increased in the second seal ring, and seal leakage was decreased; however, the hydrodynamic force as well as the hydrostatic force tended to reduce and resulted in seal-ring seizure. The one-ring seals showed superiority to the two-ring seals as they did not result in seizure.

8. Acknowledgments

This paper is based on previous cryogenic tribology studies carried out by Japan Aerospace Exploration Agency (JAXA) at Kakuda Space Center. These studies were also supported by IHI Corporation for turbopumps, by NTN Corporation for bearings and by Eagle Industry Co., LTD. for shaft seals, respectively. The author is indebted to workers engaged for their valuable support, to organizations for their enthusiastic cooperation as well as to late Prof. Miyakawa, Y. of Hhosei University for his guidance to cryogenic tribology. 9. References [1] Nosaka, M., Oike, M., Kikuchi, M., Kamijo, K.

and Tajiri, M., “Tribo-Characteristics of Self-Lubricating Ball Bearings for the LE-7 Liquid Hydrogen Rocket-Turbopump,” Trib. Trans., 36, 3, 1993, 432-442.

[2] Nosaka, M., Oike, M., Kikuchi, M., Kamijo, K. and Tajiri, M., “Self-Lubricating Performance and Durability of Ball Bearings for the LE-7 Liquid Oxygen Rocket-Turbopump,” Lubr. Eng., 49, 9, 1993, 677-688.

[3] Oike, M., Nosaka, M., Kikuchi, M. and Watanabe, Y., “Performance of a Shaft Seal System for the LE-7 Rocket Engine Oxidizer Turbopump,” Proc. of 18th ISTS, Kagoshima, 1992, 143-154.

[4] Nosaka, M., “Tribology in Low Temperature Environment,” J. of Japanese Society of Tribologists, 52, 11, 2007, 759-764. (in Japanese)

[5] Nosaka, M., “Tribological Burn-Out of Wear,” J. of Japanese Society of Tribologists, 36, 9, 1991, 689-691. (in Japanese)

[6] Nosaka, M., Oike, M., Kamijo, K., Kikuchi, M. and Katsuta, H., “Experimental Study on Lubricating Performance of Self-Lubricating Ball Bearings for Liquid Hydrogen Turbopump,” Lubr. Eng., 44, 1, 1988, 30-44.

[7] Nosaka, M., Oike, M., Kikuchi, M., Nagao, R. and Mayumi, T., “Evaluation of Durability for Cryo- genic High-Speed Ball Bearings for LE-7 Rocket Turbopumps,” Lubr. Eng., 52, 3, 1996, 221-233.

[8] Nosaka, M., Kikuchi, M., Oike, M. and Kawai, N., “Tribo-Characteristics of Cryogenic Hybrid Ceramic Ball Bearings for Rocket Turbopumps: Bearing Wear and Transfer Film,” Trib. Trans., 42, 1, 1999, 106-115.

[9] Nosaka, M., Kikuchi, M., Kawai, N. and Kikuyama, H., “Effect of Iron Fluoride Layer on Durability of Cryogenic High-Speed Ball Bearings for Rocket Turbopumps,” Trib. Trans., 43, 2, 2000, 163-174.

[10] Nosaka, M., Takada, S., Kikuchi, M., Sudo, T. and Yoshida, M., “Ultra-High-Speed Performance of Ball Bearings and Annular Seals in Liquid Hydrogen at up to 3 Million DN (120,000 rpm),” Trib. Trans., 47, 2004, 43-53.

[11] Ohta, T., Kimoto, K., Kawai, T., Motomura, T., Russ, M. and Paulus, T., “Design, Fabrication and Test of the RL60 Fuel Turbopump,” AIAA- 2003-5073, 2003.

[12] Nosaka, M., Kikuchi, M., Oike, M. and Kawai, N., “Tribo-Characteristics of Cryogenic Hybrid Ceramic Ball Bearings for Rocket Turbopumps: Self-Lubricating Performance,” Trib. Trans., 40, 1, 1997, 21-30.

[13] Nosaka, M., Takada, S., Yoshida, M., Kikuchi, M., Sudo, T. and Nakamura, S., “Effect of Tilted Misalignment of Tribo-Characteristics of High- Speed Ball Bearings in Liquid Hydrogen,” Tribology Online, 5, 2, 2010, 71-79.

[14] Nosaka, M., Takada, S., Yoshida, M., Kikuchi, M., Sudo, T. and Nakamura, S., “Improvement of Durability of Hybrid Ceramic Ball Bearings in Liquid Hydrogen at 3 Million DN (120,000 rpm),” Tribology Online, 5, 1, 2010, 60-70.

[15] Suzuki, M., Nosaka, M., Kamijo, K. and Kikuchi, M., “Research and Development of a Rotating- Shaft Seals for a Liquid Hydrogen Turbopump,” Lubr. Eng., 42, 3, 1986, 162-169.

[16] Nosaka, M., Miyakawa, Y., Kamijo, K., Suzuki, M. and Kikuchi, M., “Study on Sealing Characteristics of High Speed, Contacting Mechanical Seals for Liquid Hydrogen (Part 1),” J. of Japan Society of Lubrication Engineers, 29, 1, 1984, 35-42. (in Japanese)

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