The use of experimental results in solving tribological problems

Wear, 56 (1979) 167 - 175 0 Elsevier Sequoia S.A., Lausanne

167 -

Printed

THE USE OF EXPERIMENTAL TRIBOLOGICAL PROBLEMS*

A. ZALAI

and GY.

Hungarian

Oil and Gas Research

in the Netherlands

RESULTS

IN SOLVING

BARMOS Institute,

Budapest-Veszprbm

(Hungary)

J. CZkGI Technical

University,

(Received

March

Budapest

(Hungary)

27, 1979)

Summary The use of extreme pressure (EP) additives resulted in increased gear life for the driving unit of a hot roll stand in a steel work. For large diameter sliding bearings operational problems caused by the surface roughness of the journal and bearing linings can be eliminated by using oil of the correct viscosity together with careful filtering. Seizure of the jaw coupling of a steam turbine can be eliminated by an appropriate material combination which is assessed by seizure investigations. Structural changes based on the results of an experimental investigation of sticking have eliminated unreliable operation of a starting device for a steam turbine.

1. Introduction Tribological problems arise in industry. We investigated such a problem for two different driving units from a steel works: an expansion jaw coupling and a starting device for a power plant turbine. The reduction of wear and friction and the correct choice of lubrication have a significant effect on energy conservation. Fewer parts need to be replaced and operation is more efficient if wear is reduced. The energy loss due to stopping and restarting an industrial unit because of wear or damage must also be considered. Reduction of wear and friction together with proper lubrication will produce economies in industry and transportation and thus allow energy saving.

*Paper 10

- 15.1979.

presented

at the 4th International

Tribology

Conference,

Paisley,

Sept.

168

2. Increasing mill

the life of the gear set of a horizontal

roll stand of a hot rolling

The horizontal roll stand is run at the upper limit of its rated capacity and therefore the gears which are 1000 mm in diameter have to be replaced after 1 - 1.5 a operation because of the damage caused by pitting and scoring. Calculations suggested that gear life can be increased by increasing the surface hardness, by using a higher viscosity transmission oil or by using an oil containing a wear-reducing or extreme pressure (EP) additive. Experiments with the EP oil were carried out. Details of the original oil and the new oil are given in Table 1. The original oil was a high viscosity bright stock containing no additives and the new oil was a high viscosity industrial transmission oil containing 3.5% EP additive. Because of the additives t,here is a marked difference in the four-ball test results for the two oils. TABLE

1

Properties

of the oils

Specification

Bright stock

Industrial EP transmission oil

Specific gravity at 20 “C Flash point (“C) Viscosity at 50 “C (mm2 s-l) Viscosity at 100 “C (mm2 s-l) Viscosity index Conradson number (%) Ash content (%) Acid value (mgKOH g--.l) Four-ball test

0.904

0.904

280 209.12 25.34 85 0.88 0.06 0.05

227 118.91 17.48 87 0.85 0.14 0.91

2200

3600

2000

3400

Welding load (N) Welding

load 200 N (N) __~

The gear wear was measured in June 1977 after 360 days operation. The tooth wear was 0.13 mm for the upper gear and 2.14 mm for the lower gear. Figure 1 shows the tooth profiles and their contrast marks. Intensive damage, pitting and material removal can be clearly seen. Figure 2 shows the tooth profile of a newly installed gear. During the investigation carried out with a new gear and the new lubricating oil samples of oil were taken and the more important oil test results are indicated in Table 2. The properties of the oil did not change with time. Both the welding load and the wear diameter were practically unaltered. The wear of the gear wheels was indicated by the incre ased iron content of the oil. The volume of oil make-up (1500 2000 1 r’(Brmonth) aff?ci,s the

(a) Fig. 1. (a) Profile

and (b) contrast

marks of the gear tooth.

Fig. 2. Profileof a tooth from a newly installedgear. oil changes. Laboratory test results show that the oil does not undergo any essential alteration. The changes of tooth profile of the lower gear were examined visually after six months operation. Sharp scratches were observed on the loaded side of the tooth, at the tooth tip and at the dedendum, and at the middle of the tooth both shallow and deep pits were observed. On comparing these observations with the results obtained from using additive-free lubricating oil, we found lower tooth wear and an increase in gear life by using an EP oil. After 12 months operation it was found that gear tooth wear was less than with an additive-free oil. 3. Lubrication of large diameter sliding bearings from supporting rolls In the first half of 1977 some of the large diameter sliding bearings from the supporting rolIs of the cold rolling mills failed. Geometrical and surface roughness data were obtained and the lubrication conditions of the bearings under the operating conditions were checked. Failure of the bearings can be classified into three groups. (1) Surface scratches, spots and discolourations. These defects can be attributed to the presence of solid material in the oil.

170

171

(2) Plastic flow of the tin-based bearing metal which reduced its thickness and caused displacement towards the oil pockets. Deformation of the bearing could have been caused partly by the high bearing temperature and partly by loading in excess of the load specifications. (3) Melting and wiping of the bearing metal lining. Failure was due to deterioration of lubrication and frictional conditions which produced bearing surface temperatures in excess of the melting point of the bearing material. The surface roughnesses R, of ten new journal bushings were measured. The R, values varied between 0.050 and 0.068 pm with an average value of 0.059 pm. The values of R, were 0.7 - 2.0 pm and 1.08 pm respectively. The R, values of one used bushing were 0.3 - 1.5 pm with an average of 0.86 pm, while the values of R, for the same bushing were 5 30 pm and 13.7 pm respectively. The surface roughness of the used bearing bushing increased by more than a factor of 10 (R, increased by a factor of 14.6 and R, increased by a factor of 12.7) compared with a new bushing owing to the deep scratches caused by the solid contaminants in the oil. The measured surface roughness values of the bearing linings were as follows: for the 26 new bearings R, was in the range 0.48 - 0.65 pm with an average of 0.56 pm and R, was in the range 3 - 18 pm with an average of 7.7 pm; for the five used bearings R, was in the range 0.8 - 7.6 pm with an average of 2.53 pm and R, was in the range 7 - 100 with an average of 30.5 pm. The finer particles contained in the oil are embedded in the white metal and only the larger particles cause scratching and seizure. For this reason the roughness deterioration of bearing metal surfaces is less than that of the hardened steel journal bushings. Increased roughness causes increased friction and an increase in the temperature of the rubbing surfaces so that the bearing metal softens and eventually melts. From the calculated allowable minimum oil-film thickness and the relative bearing clearance for a reversing roll stand loading of 2000 ton at a circumferential speed of 10 m s-’ it is found that an oil of viscosity 160 - 250 mm2 s-l at 50 “C can be used. With a finishing roll stand loading of 500 ton and a speed of 20 m s-l a viscosity of 30 - 160 mm2 s-l at 50 “C!is required. Regular inspection and filtering of oil as well as maintaining the specified oil pressure and temperature is advisable. 4. Investigation of seizure of the expansion jaw coupling of a steam turbine Seizure of the jaw surfaces occurred at the expansion jaw coupling of a 44 MW 3000 rev min-’ steam turbine (Fig. 3). With a calculated average surface loading of 500 N cmP2 and a turbine oil of viscosity 30 mm2 s-l at 50 “C no seizure should occur with a properly set coupling with jaws of hardness 420 - 508 HV. Owing to improper manufacture and assembly or deformation caused by loading the surface loading can exceed 2000 N cmh2. Rapid displacements with small amplitude due to spinning can cause fretting corrosion of the steel surfaces.

172

Fig. 3. Schematic

diagram

of the jaw coupling

The damage can be reduced by increasing surface hardness or by improving the frictional conditions. As it was not possible to increase hardness attempts were made to reduce friction by surface coating or by using a suitable combination of materials. Frictional model tests were carried out using a plane wear tester with a ground steel sliding plate combined with steel specimens with different coatings and with reinforced phenolic resin and polyamide specimens. The wear tester is shown schematically in Fig. 4 and the results are summarized in Table 3.

Fig. 4. Schematic

arrangement

of the wear tester:

1, specimens;

2, sliding plate.

The friction measured on Teflon-coated, polyamide and phenolic resin specimens was satisfactory. MoSz lacquer coatings were effective only when the surfaces of both the specimen and the sliding plate were coated. On the basis of these laboratory measurements material combinations giving the four lowest friction coefficients were recommended for industrial testing. 5. Prevention of the periodic sticking of the double-action switchgear of a steam turbine in a thermal power station

start and release

The function of the equipment is to ensure the operation of steam valves by mechanical and hydraulic controls when starting the turbine and to ensure immediate valve closing when stopping the equipment or in the case of breakdown. The equipment is shown schematically in Fig. 5. If the equipment is out of use for a period of about one to three days the rotary slide valve sticks and operation of the equipment becomes unreliable. This problem does not occur if the slide valves are moved regularly each day. The manufacturer is now modifying the design.

Tinned

10

--

Tinned

-

9

7 8

6

Coating with Coating with MO& lacquer MoS2 lacquer Teflon fabric coating Phenolic resin Polyamide

5

Rubbed with MoS2 paste

Rubbed in with MoS2 paste

Coating with MO& lacquer

Coating with MoS2 lacquer

-

_._

Test specimens

Sliding plate

Preparation

4

3

2

Sample no.

Test results

TABLE 3

2120

2120

2120 2120

2120

2120

2120

2120

2120

Loading force (N)

265

265

265 265

265

265

265

265

265

279.5

202.0

125 163.4

94.6

73

163

149

387

294

162 215

172

94.6

650

187

102

Surface Force moving loading sliding plate (N) (N cm-2)-~-Min. Max.

0.066

0.048

0.029 0.039

0.022

0.017

0.038

0.035

Min.

0.091

0.069

0.038 0.051

0.041

0.022

0.153

0.044

0.024

Max.

Friction coefficient

The tin coating separated; seized at 3450 N load

Also undamaged by load of 10040 N although the tin coating wore off

Also undamaged by load of 10040 N Also undamaged by load of 10040 N

Also undamaged by load of 10040 N

Also undamaged by load of 3450 N

Seized

Seized at 3450 N load

Seized at 10040 N load

Remarks

Fig. 5. Schematic

arrangement

of the steam turbine

start and release switchgear.

Movement of the slide valves can be hindered by the combined effect of two phenomena: the frictional moment acting on the slide valve as a result of the spring force and variation of the coefficient of friction. Sticking is caused partly by squeezing of the oil film as a result of the spring forces and partly by a sticking effect between the contacting surfaces due to aging or oxidation of the oil. Compensation of the spring force acting on the sliding valve was proposed to improve valve operation. Compensation of the forces acting on the slide valve eliminated the frictional force and prevented squeezing of the oil film. The slide valves can be kept in motion by the operational vibrations of the turbine and therefore instead of static friction (sticking) only the friction of the valve movement is applicable. The turbine vibration may prevent the formation of a resin film on the surface of the oil. The forces were compensated by installing an additional spring on both slide valves. A test rig has been designed and installed to simulate double-action start and release switchgear. The rig is shown schematically in Fig. 6. The aim of the investigation was to determine the magnitude and variation in terms of the period of the moment rotating the slide valves. Therefore the slide valves were not fixed by latches but were supported by a dynamometer through the balance beam. Forces proportional to the rotational moment (torque) were measured on the dynamometer. Slewing of the sliding valve was measured by indicators on the beams. The results are shown in Fig. 7. Curves 1 show the moment (torque) measured in motion (during wind-up) which is independent of the operating time. Two separate curves were required because the pre-tension of the

175

Fig. 6. Schematic diagram of the rig: 1, test stand; 2, release device; 3, starter wheel; 4, axle for torque measurement; 5, dynamometer; 6, heating-oil jet; 7, thermostat; 8, pump; 9, pressure control valve; 10, manometer. M fNt?lj 3

_=--Q

20

:;_o

Fig, 7. Test results: curve 1, average value of the torques measured in motion during wind-up; curve 2, average value of the torques measured in motion during cut-off; curve 3, cut-off torque measured at rest; curve 4, cut-off torque measured for the original design at rest. was different for both couplings. Curves 2 show the moment measured in motion which acts in the direction of the cut-out. This value is less than the torque of the spring by an amount equal to the frictional torque caused by movement. Curves 3 show the cut-out moment measured at rest. This value determines the smooth operation of the equipment. The static friction depends on the operating time. The cut-out torque decreases significantly with time but after about 48 h operation it did not change appreciably in the two weeks during which the investigation was carried out. Curves 4 show the original cut-out torque which was measured to evaluate the effect of the design change. During the measurements carried out on the test rigs both slide valves stuck after 24 h operation and in one instance one slide valve did not disengage even after a standstill of several minutes. The measurements carried out indicate that a design change in the double-action start and release switchgear would improve. This should ensure reliable and proper operation of the equipment. springs