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Testing high temperature lubricity of waterbased and synthetic hydraulic fluids
Lubricants and Lubrication / D. Dowson et al. (Editors) 0 1995 Elsevier Science B.V. All rights reserved.
327
Testing high temperature lubricity of waterbased and synthetic hydraulic fluids Eldar Onsoyen SINTEF Materials Technology, N-7034 Trondheim, Norway Two water-glycol- and two synthetic (polyalphaolefin) hydraulic fluids have been tested to determine the influence of 150°C temperature on the wear- and lubrication properties. The aim was to assure reliable operation of dynamic piston seals in "surface controlled subsurface safety valves" in petroleum production wells. To simulate the operating conditions of the piston seals, a new test equipment was developed, using a pinon-disc arrangement in a closed pressure- and temperature chamber. The lubrication properties were studied at 150°C and room temperature, using new fluids and fluids exposed to four months of degradation at 150°C. For the interacting parts, two material combinations relevant for the piston seals were used: two nickel alloys, and a nickel alloy against PTFE. For the synthetic fluids, the metal wear increased significantly when increasing the temperature from room temperature to 150°C. For the waterbased fluids, the temperature had no significant influence on metal wear. One waterbased fluid showed poorer metal wear performance than the three other fluids, and one synthetic fluid showed higher PTFE wear rate at 150°C than the other fluids. Four months of degradation at 150°C seemed to have little or no effect on the lubrication performance. The repeatability of the test system was not yet satisfactorily, and some modifications are desirable. 1. INTRODUCTION
In some future wells for petroleum production, high temperature (130 to 200°C) and high pressure are expected in the oil or gas reservoir. A critical component in the production system is the "surface controlled subsurface safety valve" (SCSSV), which is located typically 50 to 300 metres below seabottom level. The valve is operated by a hydraulic control system, and the valve reliability is to a great extent depending on the properties of the hydraulic tluid. The hydraulic control fluid is typically either a water-glycol fluid or a synthetic (water-free) polyalphaolefin (PAO) fluid. Important properties of such fluids include lubricity, chemical stability, corrosion properties, compatibility with elastomeric seal materials, toxicity and bacterial growth. All these properties have been studied in various projects in SINTEF. Regarding hydraulic fluid lubricity, i.e. boundnrv lubrication properties, the most critical parts of
a safety valve are the dynamic piston seals. The valve is held in open position by the control line hydraulic pressure, acting on one or more pistons. When the control pressure is released, the valve is closed by a spring. The pistons have dynamic seals, either metal-to-metal seals or non-metallic seals (e.g. PTFE-based), or a combination. Metal-to-metal dynamic seals are especially common in valves designed for high pressure (70 - 100 MPa) and high temperature (130 - 200OC). The hydraulic control line pressure is typically 15 - 20 MPa above the rated valve pressure. The metallic materials are selected to withstand the extreme corrosive environment, and may have poor friction- and wear properties. This paper presents the lubricity part of a project where lubricity, stability and corrosion properties of two water-glycol fluids and two synthetic fluids were studied. The fluids were tested at 150°C and room temperature, both in new condition and after four months of degradation at 150°C in N,atmosphere.
328 2. DEVELOPMENT OF TEST EQUIPMENT A main challenge in the project was to develop an equipment for testing lubrication properties of water-based fluids at 15OoC, which is far above the boiling point. The best would be to base the tests on piston / cylinder units from a safety valve manufacturer, but the costs would be too high. Therefore we decided to build a test equipment to simulate the interacting conditions between the dynamic piston seals and the cylinders.
2.1. Specifications and design basis Two material combinations relevant for safety valves were selected for the interacting parts: Metal-to-metal combination: Hastelloy C 276 to represent the seal, against Inconel 718, to represent the cylinder. Both are nickel alloys with good corrosion resistance. Elastomer to metal combination: PTFE (with 15% graphite) to represent the seal, against the Inconel 718 cylinder material. Other specifications to simulate the safety valve dynamic seals were: Temperature: It should be possible to test waterbased fluids at 150°C without boiling. Relative velocity: 0.1 m/s. Type of motion: Reciprocating. Contact pressure between the interacting parts: Adjustable from 0.1 MPa and up. The pin-on-disc principle was selected, in a closed chamber. A rotating disc (reciprocating rotation) with I 0 0 mm outer diameter represents the cylinder, and stationary pins represent the seal. The vapour pressure of water at 150°C is 0.48 MPa. To avoid local boiling due to frictional heat, 2 MPa pressure was decided inside the test chamber, using N, to limit the oxygen exposure. Special challenges during the development were: I.
The surface pressure between the pins and the disc should be uniform across the whole contact area. This calls for a rigid pin attachment to avoid tilting, and an accurate machining of the disc and pins, to assure parallel surfaces. In addition, no sharp pin edges should scrape
against the disc. It should be possible to test two or three pins simultaneously, with equal load. It should also be possible to run several sets of pins against fresh disc areas, without needing to open the test chamber. The pins should be inexpensive and easy to fabricate, to make it possible to use new pins in all the tests. In this way the "evidence" is not spoiled, and the pins can be re-examined if required. It should be possible to use the equipment for a variety of wear and lubrication research, with different material combinations, pin shapes, relative velocities, start/stop cycles, etc. The following test cycle was decided (10 seconds total): 60 mm sliding (0.6 s) along a circular track in one direction, 4.4 seconds idle, 60 mm sliding in the opposite direction and 4.4 seconds idle.
2.2. Test equipment design Photos of the test equipment are given in Figure 1 and 2, and Figure 3 shows a schematic sketch. The test disc is attached to a rotating shaft, mounted in a small lathe and operated by a low-speed hydraulic motor. Around the disc there is a stationary housing (test chamber), with ball bearings and seals against the shaft. The ball bearing inner rings may slide on the shaft in the axial direction, so that the test pins can be forced against the disc. The test load is achieved by a spring, which is adjusted by a nut. Only one test pin is seen in the left-hand part of Figure 3, screwed into the end cover. However, three pins are run at a time, and there are three sets of metal pins and one set of PTFE pins screwed into the end cover, as seen in the right-hand part of the figure. The pins A1 to C3 are all metal pins. At first, the A1 to A3 pins are tested against the disc at 68 mm diameter. Then the Bl-B3 pins are run at 80 mm diameter and the Cl-C3 at 90 mm. The metal pins are produced with a 5" cone. At first the Cl-C3 pins are mounted, with the three cone tips at equal elevation above the end cover inner surface. Then a flat surface with a very fine grinding paper is used to achieve a 2 mm diameter flat top surface (3.14 mm2) on each of the pins.
329
Figure 1 . Test equipment, assembled.
Figure 2. Test equipment, opened.
Figure 3. Test equipment, simplified (left); position of test pins (right). The same procedure is used for the Bl-B3 pins, at a 0.5 mm higher elevation than the C pins, and at last the Al-A3 pins are mounted and ground in the same manner. The D1-D3 pins are PTFE pins, mounted 0.5 mm below the Cl-C3 pins. They are run against the outer edge of the disc, and have such a position that they may be turned 180 degrees for another test on the same pins. The contact area for each pin is 17 m m 2. The test disc, which is ground and polished to a surface roughness of less than 0.1 p,has a flexible
attachment to the shaft. In this way the three pin top surfaces may slide against the disc with almost equal load, and with parallel interacting surfaces. Using a reciprocating disc rotation, each pin may run against a separate part of the disc, without disturbance from other pins. The housing is filled with 330 ml of test fluid, and is heated to the selected temperature by a heating cable around the housing. A thermocouple inside the chamber allows for accurate temperature control. Because a low-speed hydraulic motor is used directly on the shaft, the rotation starts and
330 stops immediately according to the chosen cycle. The motor is operated by a hydraulic power pack. A frequency controlled electric motor on the power pack allows for an accurate control of the sliding velocity. With the present test system, the torque on the test chamber (due to pin / disc friction) may be measured relatively accurate, after having measured and corrected for the torque caused by friction of the two shaft lip seals. A torque arm and a strain gage based load cell is used for this purpose. The normal (perpendicular) pin / disc force was however more difficult to measure accurately. Therefore, it was decided to control the test load based on friction force, i.e. shear stress at the pin / disc interaction, rather than on normal force. Implications of this decision are commented in Section 4.1.
2.3. Operating conditions An important aim when designing the test procedure was to be able to measure wear rate and friction coefficient in a mild wear situation, at least for the "best" fluids. Several preliminary tests were run, and the following loading conditions were concluded (in addition to those mentioned in Section 2. I ): Shear stress in the pin 1 disc interaction, and number of cycles - metal pins: * Running-in: 100 cycles (12 m sliding) at 0.5 f: 0.1 MPa shear stress * 200 cycles (24 m) at 1.0 f 0.1 MPa * Five to six tests (x 20 to 25 cycles = 100 to 150 cycles) at different shear stress levels in the 0 - 1.5 MPa range, to measure the friction coefficient Shear stress and number of test cycles - PTFE pins: * Running-in: 100 cycles at 0.15 f 0.03 MPa * 200 cycles at 0.3 f 0.05 MPa * Five to six tests (x 20 to 25 cycles = 100 to 150 cycles) at different shear stress levels in the 0 - 0.3 MPa range, to measure the friction coefficient Before each test: The pins and disk were ground and polished to a surface roughness of R, c 0. I pm, and the test equipment was rinsed
thoroughly and dried, An example of a torque measurement on the test chamber is shown in Figure 4. A PC based data logging system ("Labtech Notebook") is used to control the test cycle and to log the friction measurements at a 10 Hz rate (10 measurements per second). The torque peaks were excluded from the calculations.
I
-1,5
20
25
30
35 40 45 Time (sec)
50
55
Figure 4. Example of torque signal when running PTFE pins.
3. TEST RESULTS Four fluid types were tested in new condition and after four months degradation at 150°C (in N, atmosphere): Fluids A and B: Two makes of synthetic (polyalphaolefin) fluid Fluids C and D: Two makes of water / glycol fluid (about 40% glycol) The main results are given in Table 1 on the next page. The smaller numbers, the better. The main terms are explained and commented below:
Surface appearance: Surface appearance of the pin top surface: The values (0 - 5) are based on visual inspection, describing the fraction of the top surface having traces of severe wear. Zero means that the whole surface is polished (mild wear). Three examples of pin surfaces after tests are shown in Figure 5 .
Table 1 - Results from wear testing. The smaller numbers. the better. See text (Section 3) for further explanation.
ROOM TEI 'ERATURE Synthetic fluids A A B =w degr. new
Pin 1 Pin 2 Pin 3 Average Relative average DUghneSS pin (Ra, pm)
pin1
Pin 2 Pin 3 Average Relative average 'ear depth pin, pm Pin 1 Pin 2 Pin 3 Average Sal friction energy lhrough lest.J m m 2 ~rmaiiiedaverage pin wear deplh. pm elatie average pin wear depth O u g h w dlsC (Ra, urn) Against pin 1 Against pin 2 Against pin 3 Average Relative average Beforelest riction Coeffidenl. 96
.
.
0 0 0 0 0.0
0 1 0 0.3 02
B
deg
Synthetic fluids
C C D D ww degr. new deg
A A B B e w degr. new degr
0 1 4 5 2 1 0 1 3 4 1 1 4 1 1 1 0 0 0 0.7 3.7 3.3 1,3 1 0.0 0.4 22 2.0 0.8 0.6
0.4 0,l 0.2 0,4 0,3 0.1 0.1 0.6 02 0.1 0,l 0.1 D.3 0,l 0.1 0.3 0.3 0.1 0.1 0.4
0 4 2 2 12
2 2 3 3 4 3 0 1 0 1.7 2,3 2 1.0 1.4 12
1.2 0,l 0.5 1.1 0.8 02 12 1.0 0.6 1.3 1.1 0.6
3.5 1.4 3,O 1.4
2,3 1.5 12 0.1 02 1.0 1.1 1,4 1.3 1.2 1.6 1.5
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 41 87 41 49 0,O 1-0 0.0 0,o 0.0- 3.0 0.0 0.0
1.05 0,06 0.07 0.O 1.07 0,07 0.06 0.12 1.05 0.06 0.08 0,05 1,06 0.06 0.07 0,07 ),11 0,12 0.13 0.14 408 0.07 0.08 0,O 16 32 13 20 1,3 1,0 1.0 1,0
1.07 1.89 1.65 1.87
0.91 0,07 023 0.40
20 2 0 18 19
5 16 28 7 0 3 2 3 3 1 4 6 3 0 2 5 5 6 9 9 1 4 9 2 8 5 1.4 0.8 0.1 5,6 2.7 1.3 0.6 08 12 1.1 12 1,3 1.1 1.1 12 1.0
15 5 18 1 6 8 4 8 1 4 11 6 33 8 10 13 10 4 33 24 1 1 0.7 0.3 0.1 0.9 3.2 4,7 2.5 0.7 1.1 0.6 12 0.8 l a 0,8 1.3 0.9 1.0 0,9 12 0.8 0.9 0.8
letalwear TFE wear and friction
IS!
0.13 0,05 0.H
),85
0.12
w
Water/glycoi fluids C
094 0.45 0,60 0,05 0.17 0.05 0,57 0.18
169 1,11 0 3 0,78
0.3
12 1.1 1.0
14 14 11 13 9.1 14 0.4
D
2 3 1 1 4 4 1 3 2 2 0 0 2.7 3 0,7 1.3 1.6 1.8 0.4 08 1.6 1.3 1.4 1.5
1.1 02 0.6 0,l 1,7 02 1,l 02
1.3 0.2
0.1 1.4 0,l 0.5 0.6
0 15 0 10 0 20 0 5 0 0 7 10 0 35 31 39 11,4 0.0 6,8 2.7 32 0.0 1.9 10 10 1 10 41 9,8
0,82 0,05 02: 1,14 0.07 1.N 0,lO 0.05 0.E 0,69 0.06 0.4t 1,s 1p 0,ll 0,8L
1.95 1.69 1.32 1.32
20 1.4
17 70 22 21 18 45 1.4 0.9 0.9 0.8 0.8 1.0
25 33 30 29 11.0 27 0.8 5,6
106 41 90 97 11,o
134 110 110 118 9,8 120 3.4 3
1 5 3 7 9 5 3 7 9 5 1 5 5 5 9 2 0 1 2 7 9 5 5 15 33 76 50 12.0 11.0 11,o 12,l 88 1 2 3 0 6 9 4 2 0.3 0.9 2.0 1.2 2.5 7,5 6 3.5 5 5 13 12 0.7 1,8 1.4 0,8 12 1.5 1,4 1,l 1.0 1,4 0.9 0,8 0.9 1.4 1.4 12 1.0 1,l 0.8 0,8 0,s 1.3 1.3 1.1 0.9 1.0 0.8 0.7 0.9
sss
\vg. Max. Weigt fadoi
D
o m 052 l,42 032 1,lO 1,02 I* la? ap 0,s D$s 0,w 11,1: s 0937 031 4n a s 234
2#9
aec fa 0
12
0.0
1.07 0.86 0,s 0.93 ,40 1,19 0.99 0.91 1.55 0,07 0.06 0.05 1.67 0.71 0,57 0.63 31 137 1,11 1 2 2
1.12 0.11 0.05 0,08 85 70 19 21 40 1.6 1.4 1.1 1.0 1.4 2 9 5 5 1 5 0.1
0 0 0 0 46 0,o
C
DW
12 12 2,o 1,o 12 0.9 2.6 1.9 1.1
2.0 0.6 1,3 1.4
1 0 0 0 0 2 0 5 0 0 5 0 0 0 5 1 0 0 5 0 0 0 10 10 0 7 0 0 0 1 2 2 3 0 6 2 3 8 4 2 3 1 71 32 33 43 0.0 0.0 0.0 6.6 292 0.0 12 0.0 0.0 0.0 1.8 8.1 0.0
Pin 1 Pin 2 Pin 3 Average ~ t afriction i energy through test. J m m 2 ormalizedaverage pin wear depth, pm elalive average pin wear depth riction coefficient,% elative friction weffiaenl eal friction(toque, Nm) verage seal frid. (metal- and PTFE tests)
150"C, repeatd rn€
1%
Water/glycol fluids
0 0 0 0 1 . 0
d
r. new de r new d r. new
3 4 3 3 4 4 0 3 0 3 3 4 3 0 1 0 0 2 2 0 0 2.3 1.3 2 2.7 3.3 2.3 0 1.4 0.8 12 1.6 2.0 1.4 0.0
2.0 22
5 4 0.8 4 1.7 3,7 1
2
12 1.9 1.7 0.8 2,4 2,6 0.1 1,l 2.6 1.7 02 1.2 0,5 22 1,0 0,l 12 8.5 12 0,l 02 0.1 0.5 0.1 0.1 0.1 0,4 1.7 1.9 12
1.4 1.0 0.8 1,0 0.6 1.5 1.3 0.1 1.5 1.1 0.8 1.1 0.7 1.7 1.4 0.1 0 0 0 0
lo6
0 0 0 0 44
0 0 0 0 41
5 5 0 3 46 2.9
1.0 0,o 0.0 LO 0.0 0.0 0.8
0.9
1
3.0
1
3.6 292 1
1
1.61 1.67 123 1.52 1
1
134 1.69 1,32 1.32
1,08 0,12 31 85
35 120 1.00 4.3 7.5 1 1.1 1.5 1.1 1.4
1.00 2,9 1.01 2.5
4 1
w
w
L
332
Figure 5. Examples of pin top surfaces after tests.
Relative average: The average for that specific test (i.e. the number above), divided by the average for all fluids and tests (column no. 3 from the right) Roughness: The roughness (R,-value, pn) was measured along a 0.8 mm distance, perpendicular to the wear traces. For the pins, the roughness is measured at the "worst" spot. Wear depth: The wear depth (pn) of the metal pins is calculated from the measured increase of diameter of the top surface. Because of the 5" cone, a 0.1 mm diameter increase means about 5 pm wear depth. In practice, 5 pm is the sensitivity of these measurements. The wear depth (pm) of the PTFE pins is measured in an optical microscope. Because only a fraction of the pin top surface has rubbed against the disc, the step from the untouched area to the worn area can be measured relatively accurately. Total friction energy: The total friction energy in J/mm2, which the pins have suffered during the test. Normalized average pin wear depth: To compensate for differences in total friction energy, the average wear depths are multiplied by (40 / total friction energy) for metal pins and by (10 / total friction energy) for FTFE pins. Friction coejticient: The measured friction coefficients (%) for the metal pins do not have the desired accuracy. It is highly influenced by the degree of severe wear, and of metal debris ploughing into the disc surface. In addition, as described in Section 2.2, the accuracy of the normal force measurements is poor.
Seal friction: Torque (Nm) on the test chamber due to friction of the two shaft seals, at about 2 MPa internal pressure. The seals are lip seals, made of almost the same material as the PTFE pins (graphite filled FTFE). The measurements are included as a supplement to the PTFE friction coefficient measurements. SUMMARY (weighted relative average): The two bottom lines of the table give weighted average values of the "relative average" numbers, using the "weight factors" in the last column. Figure 6 and 7 give a graphic views of the two bottom lines in Table 1. The results are discussed in Section 4.2. The lubricant viscosity has often a major influence on friction and wear. The kinematic viscosity at 40°C (average of two measurements) for each fluid are given in Table 2. The main findings were: For new fluids, the viscosity is about 6 - 7 times higher for synthetic fluids than for water / glycol fluids The four months of degradation has not changed the viscosity of the synthetic fluids much, while the viscosity of the water / glycol fluids has increased by 25 - 40%. Table 2 Viscosity (cSt) at 40°C New
Degraded
Synthetic fluid A
14.2
14.5
Synthetic fluid B
16.8
16.2
Waterbased fluid C
2.3
2.9
Waterbased fluid D
2.5
3.5
333 Relative value
Relative value 0
1
2
3
0
4
Metal, 150°C - II PTFE, room temp
Fluid A - synthetic
Metal, 150°C - I Metal, 150°C - II PTFE, room temp PTFE, 150°C
Fluid B - synthetic
Metal, room temp Metal, 150°C - I Metal, 150°C - II PTFE, room temp PTFE, 150°C
-
M
Metal, 150°C - II PTFE, room temp
Metal, room temp Metal, 150°C - I Metal, 150°C II PTFE, room temp PTFE, 150°C
-
-
II
2
3
4
Metal, room temp
Metal, room temp
Metal I 1500
1
...................................
Metal, room temp Metal, 150°C - I Metal, 150°C II PTFE, room temp PTFE, 150°C
-
Fluid D water I glycol
-
-
Fluid A - synthetic
Fluid B - synthetic
Fluid C - water I glycol
Fluid D - water I glycol
I
10 New
IDegraded
I
Figure 6. Relative values for metal- and PTFE wear, based on the Summary in Table 1. The smaller values, the better.
Figure 7. Relative values for metal- and PTFE wear (average of the new- and degraded fluid values from Figure 6). The smaller values, the better.
4. DISCUSSION AND CONCLUSIONS
ating). The test equipment can also be used for a variety of other material combinations and operating conditions. Table 1 and Figure 6 show that further development of the test equipment is desired to improve the repeatability. One main reason for the variation of results between equal tests may be the poor friction- and wear properties of the selected material
4.1. Limitations of the test equipment Much effort was put into making the pin-ondisc test equipment relevant to piston seals in safety valves for petroleum production. This includes relevant temperature (1SOOC), material combinations, sliding velocity and type of motion (reciproc-
334 combination. A transition from a mild wear to severe adhesive wear occurs at a relatively low surface pressure, depending on the boundary lubrication properties of the fluids. The test conditions were close to these transition limits. A main disadvantage of the equipment was that the test load must be based on friction force rather than on normal load. This means that if the friction coefticient decreased, the load was increased manually. In other words, the test results show how the lubricant can reduce wear when the friction energy is kept constant, and this is not quite relevant for the piston seals. In the real life situation, the perpendicular surface pressure is a function of the hydraulic control pressure, and the friction force will vary, depending on the fluid type and other conditions. In spite of these limitations, the tests provided a lot of interesting and useful results. 4.2. The difference between fluid types, and the influence of degradation No statistical analyses of the results have been carried out. Yet, the conclusions below can be drawn from the test results, as presented in Figure 6 and 7: Influence of fluid degradation: For fluid B, it may seem that the performance has been reduced slightly due to degradation. For the other fluids, there is no reason to believe that four months of degradation at 150°C has had any major effect on the lubricity. However, we do not know the effect of several years of degradation, which will be the case in the real life situation.
Influence of temperature - synthetic fluids: For the synthetic fluids, the metal wear increased significantly when increasing the temperature from room temperature to 150°C. A reason for this may be the decreasing viscosity (Table 2). The F'TFE wear increased only for fluid B. Influence of temperature - waterbased fluids: For the waterbased fluids, the temperature seems to have no significant influence on metal wear. The PTFE wear has increased only for fluid D. Overall fluid Performance: Fluid C (waterbased) showed poorer metal wear performance than the three other fluids, and fluid B showed higher PTFE wear rate at 150°C than the other fluids. Apart from this, the performance of the four fluids at 150°C was fairly equal.
ACKNOWLEDGEMENTS I wish to thank: Statoil for economic support and for the permission to publish the results. Nils-Inge Nilsen (SINTEF Materials Technology) for major contributions on design and fabrication of the test equipment, and on running lubricity tests. Sven Morten Hesjevik (SINTEF Materials Technology) for supplying degraded fluids from a fluid stability and corrosivity sub-project.
327
Testing high temperature lubricity of waterbased and synthetic hydraulic fluids Eldar Onsoyen SINTEF Materials Technology, N-7034 Trondheim, Norway Two water-glycol- and two synthetic (polyalphaolefin) hydraulic fluids have been tested to determine the influence of 150°C temperature on the wear- and lubrication properties. The aim was to assure reliable operation of dynamic piston seals in "surface controlled subsurface safety valves" in petroleum production wells. To simulate the operating conditions of the piston seals, a new test equipment was developed, using a pinon-disc arrangement in a closed pressure- and temperature chamber. The lubrication properties were studied at 150°C and room temperature, using new fluids and fluids exposed to four months of degradation at 150°C. For the interacting parts, two material combinations relevant for the piston seals were used: two nickel alloys, and a nickel alloy against PTFE. For the synthetic fluids, the metal wear increased significantly when increasing the temperature from room temperature to 150°C. For the waterbased fluids, the temperature had no significant influence on metal wear. One waterbased fluid showed poorer metal wear performance than the three other fluids, and one synthetic fluid showed higher PTFE wear rate at 150°C than the other fluids. Four months of degradation at 150°C seemed to have little or no effect on the lubrication performance. The repeatability of the test system was not yet satisfactorily, and some modifications are desirable. 1. INTRODUCTION
In some future wells for petroleum production, high temperature (130 to 200°C) and high pressure are expected in the oil or gas reservoir. A critical component in the production system is the "surface controlled subsurface safety valve" (SCSSV), which is located typically 50 to 300 metres below seabottom level. The valve is operated by a hydraulic control system, and the valve reliability is to a great extent depending on the properties of the hydraulic tluid. The hydraulic control fluid is typically either a water-glycol fluid or a synthetic (water-free) polyalphaolefin (PAO) fluid. Important properties of such fluids include lubricity, chemical stability, corrosion properties, compatibility with elastomeric seal materials, toxicity and bacterial growth. All these properties have been studied in various projects in SINTEF. Regarding hydraulic fluid lubricity, i.e. boundnrv lubrication properties, the most critical parts of
a safety valve are the dynamic piston seals. The valve is held in open position by the control line hydraulic pressure, acting on one or more pistons. When the control pressure is released, the valve is closed by a spring. The pistons have dynamic seals, either metal-to-metal seals or non-metallic seals (e.g. PTFE-based), or a combination. Metal-to-metal dynamic seals are especially common in valves designed for high pressure (70 - 100 MPa) and high temperature (130 - 200OC). The hydraulic control line pressure is typically 15 - 20 MPa above the rated valve pressure. The metallic materials are selected to withstand the extreme corrosive environment, and may have poor friction- and wear properties. This paper presents the lubricity part of a project where lubricity, stability and corrosion properties of two water-glycol fluids and two synthetic fluids were studied. The fluids were tested at 150°C and room temperature, both in new condition and after four months of degradation at 150°C in N,atmosphere.
328 2. DEVELOPMENT OF TEST EQUIPMENT A main challenge in the project was to develop an equipment for testing lubrication properties of water-based fluids at 15OoC, which is far above the boiling point. The best would be to base the tests on piston / cylinder units from a safety valve manufacturer, but the costs would be too high. Therefore we decided to build a test equipment to simulate the interacting conditions between the dynamic piston seals and the cylinders.
2.1. Specifications and design basis Two material combinations relevant for safety valves were selected for the interacting parts: Metal-to-metal combination: Hastelloy C 276 to represent the seal, against Inconel 718, to represent the cylinder. Both are nickel alloys with good corrosion resistance. Elastomer to metal combination: PTFE (with 15% graphite) to represent the seal, against the Inconel 718 cylinder material. Other specifications to simulate the safety valve dynamic seals were: Temperature: It should be possible to test waterbased fluids at 150°C without boiling. Relative velocity: 0.1 m/s. Type of motion: Reciprocating. Contact pressure between the interacting parts: Adjustable from 0.1 MPa and up. The pin-on-disc principle was selected, in a closed chamber. A rotating disc (reciprocating rotation) with I 0 0 mm outer diameter represents the cylinder, and stationary pins represent the seal. The vapour pressure of water at 150°C is 0.48 MPa. To avoid local boiling due to frictional heat, 2 MPa pressure was decided inside the test chamber, using N, to limit the oxygen exposure. Special challenges during the development were: I.
The surface pressure between the pins and the disc should be uniform across the whole contact area. This calls for a rigid pin attachment to avoid tilting, and an accurate machining of the disc and pins, to assure parallel surfaces. In addition, no sharp pin edges should scrape
against the disc. It should be possible to test two or three pins simultaneously, with equal load. It should also be possible to run several sets of pins against fresh disc areas, without needing to open the test chamber. The pins should be inexpensive and easy to fabricate, to make it possible to use new pins in all the tests. In this way the "evidence" is not spoiled, and the pins can be re-examined if required. It should be possible to use the equipment for a variety of wear and lubrication research, with different material combinations, pin shapes, relative velocities, start/stop cycles, etc. The following test cycle was decided (10 seconds total): 60 mm sliding (0.6 s) along a circular track in one direction, 4.4 seconds idle, 60 mm sliding in the opposite direction and 4.4 seconds idle.
2.2. Test equipment design Photos of the test equipment are given in Figure 1 and 2, and Figure 3 shows a schematic sketch. The test disc is attached to a rotating shaft, mounted in a small lathe and operated by a low-speed hydraulic motor. Around the disc there is a stationary housing (test chamber), with ball bearings and seals against the shaft. The ball bearing inner rings may slide on the shaft in the axial direction, so that the test pins can be forced against the disc. The test load is achieved by a spring, which is adjusted by a nut. Only one test pin is seen in the left-hand part of Figure 3, screwed into the end cover. However, three pins are run at a time, and there are three sets of metal pins and one set of PTFE pins screwed into the end cover, as seen in the right-hand part of the figure. The pins A1 to C3 are all metal pins. At first, the A1 to A3 pins are tested against the disc at 68 mm diameter. Then the Bl-B3 pins are run at 80 mm diameter and the Cl-C3 at 90 mm. The metal pins are produced with a 5" cone. At first the Cl-C3 pins are mounted, with the three cone tips at equal elevation above the end cover inner surface. Then a flat surface with a very fine grinding paper is used to achieve a 2 mm diameter flat top surface (3.14 mm2) on each of the pins.
329
Figure 1 . Test equipment, assembled.
Figure 2. Test equipment, opened.
Figure 3. Test equipment, simplified (left); position of test pins (right). The same procedure is used for the Bl-B3 pins, at a 0.5 mm higher elevation than the C pins, and at last the Al-A3 pins are mounted and ground in the same manner. The D1-D3 pins are PTFE pins, mounted 0.5 mm below the Cl-C3 pins. They are run against the outer edge of the disc, and have such a position that they may be turned 180 degrees for another test on the same pins. The contact area for each pin is 17 m m 2. The test disc, which is ground and polished to a surface roughness of less than 0.1 p,has a flexible
attachment to the shaft. In this way the three pin top surfaces may slide against the disc with almost equal load, and with parallel interacting surfaces. Using a reciprocating disc rotation, each pin may run against a separate part of the disc, without disturbance from other pins. The housing is filled with 330 ml of test fluid, and is heated to the selected temperature by a heating cable around the housing. A thermocouple inside the chamber allows for accurate temperature control. Because a low-speed hydraulic motor is used directly on the shaft, the rotation starts and
330 stops immediately according to the chosen cycle. The motor is operated by a hydraulic power pack. A frequency controlled electric motor on the power pack allows for an accurate control of the sliding velocity. With the present test system, the torque on the test chamber (due to pin / disc friction) may be measured relatively accurate, after having measured and corrected for the torque caused by friction of the two shaft lip seals. A torque arm and a strain gage based load cell is used for this purpose. The normal (perpendicular) pin / disc force was however more difficult to measure accurately. Therefore, it was decided to control the test load based on friction force, i.e. shear stress at the pin / disc interaction, rather than on normal force. Implications of this decision are commented in Section 4.1.
2.3. Operating conditions An important aim when designing the test procedure was to be able to measure wear rate and friction coefficient in a mild wear situation, at least for the "best" fluids. Several preliminary tests were run, and the following loading conditions were concluded (in addition to those mentioned in Section 2. I ): Shear stress in the pin 1 disc interaction, and number of cycles - metal pins: * Running-in: 100 cycles (12 m sliding) at 0.5 f: 0.1 MPa shear stress * 200 cycles (24 m) at 1.0 f 0.1 MPa * Five to six tests (x 20 to 25 cycles = 100 to 150 cycles) at different shear stress levels in the 0 - 1.5 MPa range, to measure the friction coefficient Shear stress and number of test cycles - PTFE pins: * Running-in: 100 cycles at 0.15 f 0.03 MPa * 200 cycles at 0.3 f 0.05 MPa * Five to six tests (x 20 to 25 cycles = 100 to 150 cycles) at different shear stress levels in the 0 - 0.3 MPa range, to measure the friction coefficient Before each test: The pins and disk were ground and polished to a surface roughness of R, c 0. I pm, and the test equipment was rinsed
thoroughly and dried, An example of a torque measurement on the test chamber is shown in Figure 4. A PC based data logging system ("Labtech Notebook") is used to control the test cycle and to log the friction measurements at a 10 Hz rate (10 measurements per second). The torque peaks were excluded from the calculations.
I
-1,5
20
25
30
35 40 45 Time (sec)
50
55
Figure 4. Example of torque signal when running PTFE pins.
3. TEST RESULTS Four fluid types were tested in new condition and after four months degradation at 150°C (in N, atmosphere): Fluids A and B: Two makes of synthetic (polyalphaolefin) fluid Fluids C and D: Two makes of water / glycol fluid (about 40% glycol) The main results are given in Table 1 on the next page. The smaller numbers, the better. The main terms are explained and commented below:
Surface appearance: Surface appearance of the pin top surface: The values (0 - 5) are based on visual inspection, describing the fraction of the top surface having traces of severe wear. Zero means that the whole surface is polished (mild wear). Three examples of pin surfaces after tests are shown in Figure 5 .
Table 1 - Results from wear testing. The smaller numbers. the better. See text (Section 3) for further explanation.
ROOM TEI 'ERATURE Synthetic fluids A A B =w degr. new
Pin 1 Pin 2 Pin 3 Average Relative average DUghneSS pin (Ra, pm)
pin1
Pin 2 Pin 3 Average Relative average 'ear depth pin, pm Pin 1 Pin 2 Pin 3 Average Sal friction energy lhrough lest.J m m 2 ~rmaiiiedaverage pin wear deplh. pm elatie average pin wear depth O u g h w dlsC (Ra, urn) Against pin 1 Against pin 2 Against pin 3 Average Relative average Beforelest riction Coeffidenl. 96
.
.
0 0 0 0 0.0
0 1 0 0.3 02
B
deg
Synthetic fluids
C C D D ww degr. new deg
A A B B e w degr. new degr
0 1 4 5 2 1 0 1 3 4 1 1 4 1 1 1 0 0 0 0.7 3.7 3.3 1,3 1 0.0 0.4 22 2.0 0.8 0.6
0.4 0,l 0.2 0,4 0,3 0.1 0.1 0.6 02 0.1 0,l 0.1 D.3 0,l 0.1 0.3 0.3 0.1 0.1 0.4
0 4 2 2 12
2 2 3 3 4 3 0 1 0 1.7 2,3 2 1.0 1.4 12
1.2 0,l 0.5 1.1 0.8 02 12 1.0 0.6 1.3 1.1 0.6
3.5 1.4 3,O 1.4
2,3 1.5 12 0.1 02 1.0 1.1 1,4 1.3 1.2 1.6 1.5
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 41 87 41 49 0,O 1-0 0.0 0,o 0.0- 3.0 0.0 0.0
1.05 0,06 0.07 0.O 1.07 0,07 0.06 0.12 1.05 0.06 0.08 0,05 1,06 0.06 0.07 0,07 ),11 0,12 0.13 0.14 408 0.07 0.08 0,O 16 32 13 20 1,3 1,0 1.0 1,0
1.07 1.89 1.65 1.87
0.91 0,07 023 0.40
20 2 0 18 19
5 16 28 7 0 3 2 3 3 1 4 6 3 0 2 5 5 6 9 9 1 4 9 2 8 5 1.4 0.8 0.1 5,6 2.7 1.3 0.6 08 12 1.1 12 1,3 1.1 1.1 12 1.0
15 5 18 1 6 8 4 8 1 4 11 6 33 8 10 13 10 4 33 24 1 1 0.7 0.3 0.1 0.9 3.2 4,7 2.5 0.7 1.1 0.6 12 0.8 l a 0,8 1.3 0.9 1.0 0,9 12 0.8 0.9 0.8
letalwear TFE wear and friction
IS!
0.13 0,05 0.H
),85
0.12
w
Water/glycoi fluids C
094 0.45 0,60 0,05 0.17 0.05 0,57 0.18
169 1,11 0 3 0,78
0.3
12 1.1 1.0
14 14 11 13 9.1 14 0.4
D
2 3 1 1 4 4 1 3 2 2 0 0 2.7 3 0,7 1.3 1.6 1.8 0.4 08 1.6 1.3 1.4 1.5
1.1 02 0.6 0,l 1,7 02 1,l 02
1.3 0.2
0.1 1.4 0,l 0.5 0.6
0 15 0 10 0 20 0 5 0 0 7 10 0 35 31 39 11,4 0.0 6,8 2.7 32 0.0 1.9 10 10 1 10 41 9,8
0,82 0,05 02: 1,14 0.07 1.N 0,lO 0.05 0.E 0,69 0.06 0.4t 1,s 1p 0,ll 0,8L
1.95 1.69 1.32 1.32
20 1.4
17 70 22 21 18 45 1.4 0.9 0.9 0.8 0.8 1.0
25 33 30 29 11.0 27 0.8 5,6
106 41 90 97 11,o
134 110 110 118 9,8 120 3.4 3
1 5 3 7 9 5 3 7 9 5 1 5 5 5 9 2 0 1 2 7 9 5 5 15 33 76 50 12.0 11.0 11,o 12,l 88 1 2 3 0 6 9 4 2 0.3 0.9 2.0 1.2 2.5 7,5 6 3.5 5 5 13 12 0.7 1,8 1.4 0,8 12 1.5 1,4 1,l 1.0 1,4 0.9 0,8 0.9 1.4 1.4 12 1.0 1,l 0.8 0,8 0,s 1.3 1.3 1.1 0.9 1.0 0.8 0.7 0.9
sss
\vg. Max. Weigt fadoi
D
o m 052 l,42 032 1,lO 1,02 I* la? ap 0,s D$s 0,w 11,1: s 0937 031 4n a s 234
2#9
aec fa 0
12
0.0
1.07 0.86 0,s 0.93 ,40 1,19 0.99 0.91 1.55 0,07 0.06 0.05 1.67 0.71 0,57 0.63 31 137 1,11 1 2 2
1.12 0.11 0.05 0,08 85 70 19 21 40 1.6 1.4 1.1 1.0 1.4 2 9 5 5 1 5 0.1
0 0 0 0 46 0,o
C
DW
12 12 2,o 1,o 12 0.9 2.6 1.9 1.1
2.0 0.6 1,3 1.4
1 0 0 0 0 2 0 5 0 0 5 0 0 0 5 1 0 0 5 0 0 0 10 10 0 7 0 0 0 1 2 2 3 0 6 2 3 8 4 2 3 1 71 32 33 43 0.0 0.0 0.0 6.6 292 0.0 12 0.0 0.0 0.0 1.8 8.1 0.0
Pin 1 Pin 2 Pin 3 Average ~ t afriction i energy through test. J m m 2 ormalizedaverage pin wear depth, pm elalive average pin wear depth riction coefficient,% elative friction weffiaenl eal friction(toque, Nm) verage seal frid. (metal- and PTFE tests)
150"C, repeatd rn€
1%
Water/glycol fluids
0 0 0 0 1 . 0
d
r. new de r new d r. new
3 4 3 3 4 4 0 3 0 3 3 4 3 0 1 0 0 2 2 0 0 2.3 1.3 2 2.7 3.3 2.3 0 1.4 0.8 12 1.6 2.0 1.4 0.0
2.0 22
5 4 0.8 4 1.7 3,7 1
2
12 1.9 1.7 0.8 2,4 2,6 0.1 1,l 2.6 1.7 02 1.2 0,5 22 1,0 0,l 12 8.5 12 0,l 02 0.1 0.5 0.1 0.1 0.1 0,4 1.7 1.9 12
1.4 1.0 0.8 1,0 0.6 1.5 1.3 0.1 1.5 1.1 0.8 1.1 0.7 1.7 1.4 0.1 0 0 0 0
lo6
0 0 0 0 44
0 0 0 0 41
5 5 0 3 46 2.9
1.0 0,o 0.0 LO 0.0 0.0 0.8
0.9
1
3.0
1
3.6 292 1
1
1.61 1.67 123 1.52 1
1
134 1.69 1,32 1.32
1,08 0,12 31 85
35 120 1.00 4.3 7.5 1 1.1 1.5 1.1 1.4
1.00 2,9 1.01 2.5
4 1
w
w
L
332
Figure 5. Examples of pin top surfaces after tests.
Relative average: The average for that specific test (i.e. the number above), divided by the average for all fluids and tests (column no. 3 from the right) Roughness: The roughness (R,-value, pn) was measured along a 0.8 mm distance, perpendicular to the wear traces. For the pins, the roughness is measured at the "worst" spot. Wear depth: The wear depth (pn) of the metal pins is calculated from the measured increase of diameter of the top surface. Because of the 5" cone, a 0.1 mm diameter increase means about 5 pm wear depth. In practice, 5 pm is the sensitivity of these measurements. The wear depth (pm) of the PTFE pins is measured in an optical microscope. Because only a fraction of the pin top surface has rubbed against the disc, the step from the untouched area to the worn area can be measured relatively accurately. Total friction energy: The total friction energy in J/mm2, which the pins have suffered during the test. Normalized average pin wear depth: To compensate for differences in total friction energy, the average wear depths are multiplied by (40 / total friction energy) for metal pins and by (10 / total friction energy) for FTFE pins. Friction coejticient: The measured friction coefficients (%) for the metal pins do not have the desired accuracy. It is highly influenced by the degree of severe wear, and of metal debris ploughing into the disc surface. In addition, as described in Section 2.2, the accuracy of the normal force measurements is poor.
Seal friction: Torque (Nm) on the test chamber due to friction of the two shaft seals, at about 2 MPa internal pressure. The seals are lip seals, made of almost the same material as the PTFE pins (graphite filled FTFE). The measurements are included as a supplement to the PTFE friction coefficient measurements. SUMMARY (weighted relative average): The two bottom lines of the table give weighted average values of the "relative average" numbers, using the "weight factors" in the last column. Figure 6 and 7 give a graphic views of the two bottom lines in Table 1. The results are discussed in Section 4.2. The lubricant viscosity has often a major influence on friction and wear. The kinematic viscosity at 40°C (average of two measurements) for each fluid are given in Table 2. The main findings were: For new fluids, the viscosity is about 6 - 7 times higher for synthetic fluids than for water / glycol fluids The four months of degradation has not changed the viscosity of the synthetic fluids much, while the viscosity of the water / glycol fluids has increased by 25 - 40%. Table 2 Viscosity (cSt) at 40°C New
Degraded
Synthetic fluid A
14.2
14.5
Synthetic fluid B
16.8
16.2
Waterbased fluid C
2.3
2.9
Waterbased fluid D
2.5
3.5
333 Relative value
Relative value 0
1
2
3
0
4
Metal, 150°C - II PTFE, room temp
Fluid A - synthetic
Metal, 150°C - I Metal, 150°C - II PTFE, room temp PTFE, 150°C
Fluid B - synthetic
Metal, room temp Metal, 150°C - I Metal, 150°C - II PTFE, room temp PTFE, 150°C
-
M
Metal, 150°C - II PTFE, room temp
Metal, room temp Metal, 150°C - I Metal, 150°C II PTFE, room temp PTFE, 150°C
-
-
II
2
3
4
Metal, room temp
Metal, room temp
Metal I 1500
1
...................................
Metal, room temp Metal, 150°C - I Metal, 150°C II PTFE, room temp PTFE, 150°C
-
Fluid D water I glycol
-
-
Fluid A - synthetic
Fluid B - synthetic
Fluid C - water I glycol
Fluid D - water I glycol
I
10 New
IDegraded
I
Figure 6. Relative values for metal- and PTFE wear, based on the Summary in Table 1. The smaller values, the better.
Figure 7. Relative values for metal- and PTFE wear (average of the new- and degraded fluid values from Figure 6). The smaller values, the better.
4. DISCUSSION AND CONCLUSIONS
ating). The test equipment can also be used for a variety of other material combinations and operating conditions. Table 1 and Figure 6 show that further development of the test equipment is desired to improve the repeatability. One main reason for the variation of results between equal tests may be the poor friction- and wear properties of the selected material
4.1. Limitations of the test equipment Much effort was put into making the pin-ondisc test equipment relevant to piston seals in safety valves for petroleum production. This includes relevant temperature (1SOOC), material combinations, sliding velocity and type of motion (reciproc-
334 combination. A transition from a mild wear to severe adhesive wear occurs at a relatively low surface pressure, depending on the boundary lubrication properties of the fluids. The test conditions were close to these transition limits. A main disadvantage of the equipment was that the test load must be based on friction force rather than on normal load. This means that if the friction coefticient decreased, the load was increased manually. In other words, the test results show how the lubricant can reduce wear when the friction energy is kept constant, and this is not quite relevant for the piston seals. In the real life situation, the perpendicular surface pressure is a function of the hydraulic control pressure, and the friction force will vary, depending on the fluid type and other conditions. In spite of these limitations, the tests provided a lot of interesting and useful results. 4.2. The difference between fluid types, and the influence of degradation No statistical analyses of the results have been carried out. Yet, the conclusions below can be drawn from the test results, as presented in Figure 6 and 7: Influence of fluid degradation: For fluid B, it may seem that the performance has been reduced slightly due to degradation. For the other fluids, there is no reason to believe that four months of degradation at 150°C has had any major effect on the lubricity. However, we do not know the effect of several years of degradation, which will be the case in the real life situation.
Influence of temperature - synthetic fluids: For the synthetic fluids, the metal wear increased significantly when increasing the temperature from room temperature to 150°C. A reason for this may be the decreasing viscosity (Table 2). The F'TFE wear increased only for fluid B. Influence of temperature - waterbased fluids: For the waterbased fluids, the temperature seems to have no significant influence on metal wear. The PTFE wear has increased only for fluid D. Overall fluid Performance: Fluid C (waterbased) showed poorer metal wear performance than the three other fluids, and fluid B showed higher PTFE wear rate at 150°C than the other fluids. Apart from this, the performance of the four fluids at 150°C was fairly equal.
ACKNOWLEDGEMENTS I wish to thank: Statoil for economic support and for the permission to publish the results. Nils-Inge Nilsen (SINTEF Materials Technology) for major contributions on design and fabrication of the test equipment, and on running lubricity tests. Sven Morten Hesjevik (SINTEF Materials Technology) for supplying degraded fluids from a fluid stability and corrosivity sub-project.