3. THE SHUT-OFF HEAD METHOD FOR MONITORING OF PUMPS

 

Measuring the Head at zero flow is a simple test (2). It is only possible where it can be tolerated, which is not so for high energy pumps nor for pumps of high specific speed where the power at shutoff is greater than that at duty point.

With the discharge valve closed fully for no longer than 30 seconds os so, suction and discharge pressures are read once steady. The liquid temperature is also needed to find the density, which is used to convert the pressure readings into head values.

Wear of vane outer diameters will show readily, as the head-flow curve of a worn pump moves towards the zero flow axis. To show sealing ring wear, the pump head/flow curve needs to be relatively steep. (Note that if the pump has a rising curve, internal leakage will initially give an increase in shutoff head).

4. THE THERMODYNAMIC METHOD FOR MONITORING OF PUMPS

Another method of pump monitoring is to measure the temperature rise of the liquid through the pump, which reflects the inefficiency of the pump. As the differential temperature is very small, great care is required to measure it. Any effects of recirculation at pump inlet and outlet must be eliminated, and tests are not possible at very low flows or zero flow. The efficiency can be calculated from the measured data of inlet temperature, differential temperature and head. Comparisons if it changes with time can be made on plots of Efficiency vs Head. For high head pumps, an allowance must be made for the isentropic temperature rise which occurs as a result of pressure increase (2).

A commercially available device is widely used in the UK water industry (3). The tappings at suction and discharge, where the pressure/temperature probes are installed, are required to be two diameters away from pump flanges. Tong-type detectors are placed to measure motor power. Pump efficiency is then found from the precise measurement of the head and temperature rise through the pump. From assessment of motor losses, the power absorbed by the pump is computed. From all this data, the pump flow can be found.

For condition monitoring, data at around normal operating point is usually sufficient. The thermodynamic method would be more attractive economically if no special tapping points were required. Research at Monash University on high head pumps using special semi-conductor temperature probes on the outside surface of the piping, covered with insulation, gave usable results, provided the pump is allowed to run at steady operation conditions for 30 minutes in order for the piping temperature to stabilise.

5. MEASUREMENT OF BALANCE FLOW FOR MONITORING OF PUMPS

Multi-stage pumps with the impellers facing in the one direction usually have a balance disk or drum arranged such that final stage discharge pressure counteracts the axial thrust on the shaft line. Another method for pump condition monitoring is to measure the leakoff from the balance device (4). The basis is that if there is increased wear in the annular space to the balance device which is evident from increased leakoff flow, then the interstage clearances are also worn.

As the leakoff line is quite small compared to the pump main flow piping, a permanent flowmeter is relatively inexpensive. For some years, overhauls have been scheduled on this basis on some boiler feed pumps. Flows are read manually, and trends plotted using a database program (Figure I). Note that here the balance flow of 15 L/s corresponds to about 10% of the duty flow, and about 250kW of extra power. When added to the likely internal recirculation, this would mean that an even larger proportion of the power absorbed being wasted. These pumps are variable speed and tests show that the measured flows must be corrected in direct proportion to the speed.

On a set of pumps of another design elsewhere, both head-flow and balance flow were measured for some years, but no correlation was found between the two.

On yet another pump, of 11 stages, the head-flow performance was tested as well below the datum curve. As the pump was dismantled, measurements showed that the interstage clearances were not worn. Results were reviewed, but then the balance seat area was reached and found to be severely eroded. Balance flow had obviously been very high. For the best monitoring, it is therefore considered that both head-flow and balance flow should be measured, particularly if the balance area can be separately dismantled in the field.

Figure I. Condition monitoring of a high energy multi-stage pump by measurement of balance device leakoff flow. (Note: flows are corrected to a standard pump speed)

6. HOW TO CALCULATE THE OPTIMUM TIME FOR OVERHAUL

The most economic time to restore lost performance by overhaul will vary with the circumstances.
If the deterioration is constant over time, then a cash flow analysis can be done to ensure that the investment in overhaul will give the required rate of return. The same process is used in deciding on any investment in plant improvement.

If the deterioration rate is increasing with time, then the optimum time for overhaul will be when the accumulated cost of the increased electricity consumption equals the cost of the overhaul.

The method is now described for some of the situations which occur:

6.1 Pump deterioration results in a reduction in plant production. Where the cost of overhaul is insignificant in proportion to the cost of lost production, prompt overhaul is usually simply justified at a convenient "window" .

6.2 Pump which runs intermittently to meet a demand. In a pumping installation such as topping up a water supply tank or pumping out, deterioration will result in the pump taking more time to do its duty. The extra service time required therefore results in increased power consumption which can be related to the cost of overhaul.

6.3 Pump deterioration does not affect plant production, at least initially: constant speed, throttle valve controlled pump. The internal wear does not cause any loss in production from the plant, as the control valve opens more fully to ensure that pump output is maintained. Eventually, as wear progresses, pump output may be insufficient to avoid loss of production, or the power taken will exceed the motor rating.

Figure II shows the Head-Power-Flow site test characteristics of such a pump. Its output is controlled using a throttle control valve. The duty flow is 825 m³/h, and the duty point in the new condition is A. The power absorbed by the pump is read off the Power-Flow curve as 2150kW: B. The power-flow curve should ideally be found on site, but the works tests information may have to suffice.

After some service, the "Test points -worn pump" plotted indicate that internal wear has occurred. When worn to this extent, the operating point moves to C, as the system resistance curve lowers when the throttle valve is opened further.
The increased power required in the worn condition can be estimated by extending from the Head-Flow curve at constant head from the operating point to D, and then intersecting the Power-Flow curve for new condition at constant flow: E. Follow the arrowed line in Figure II. This assumes that the original curve still represents the flow through the impellers (of which less is leaving the pump to the system). The power could of course be measured on test at extra expense if the pump was motor-driven.

In our example, the power required for this duty in the worn condition is shown in Figure II by the projection from the duty flow of 825 m3/h to the test curve to find 640m head, then across to the "Site test - new pump" curve, then down to the power curve, to find 2300 kW.

The extra electricity consumption is therefore 2300 - 2150 = 150kW ÷ motor efficiency (say 90% ) = 167kW.

If the sealing clearances are known, by previous experience of correlation with measured performance, or if the pump is opened up already, the extra power consumed likely to be saved by overhaul can be estimated (5).

Using this method, a number of pumps of varying wear conditions could be prioritised for maintenance, based on their increased power consumption and their relative costs of overhaul, ie. the cost/benefits.

The optimum time for overhaul can be determined from this data. The following conditions and data apply for this example: the above test points were obtained following 24 months of service since the pump was known to be in new condition; an overhaul costs $50 000; electricity costs 10c/kWh; and the pump runs for 27% of the time on average.

Our test shows that the rate of increasing cost/month has reached 167 × 0.10 × 0.27 × 720 = $3240/month (taking an average month as 720h).

As the time now is 24 months, $3240 ÷ 24 gives the average cost rate of deterioration as $135/month/month.

The optimum time for overhaul can be calculated (6), but it is more practical to calculate the average total cost/month values for a range of times. If these are plotted, you can see clearly the cost impact of doing the repairs at some other time, such as at a scheduled plant shutdown. (The time value of money could also be taken into account if required). Usually the total cost curve is fairly flat for ± 20% or so.

Here's how to calculate the total average cost per month, month by month.

In this example, take the time at 22 months. The average cost of overhaul is: $50 000 ÷ 22 = $2273/month.

The average cost of extra energy used is : $135 × ½ × 22 = $1485/month.

The total average cost/month is therefore the sum of these two figures, or $3578.

Repeat this calculation for several months, perhaps using a spreadsheet, and look for the minimum total cost, which is at 27.2 months. If plotted as cost/month against time, the resulting curves will show the cost per month of overhaul dropping with time, with the cost of lost energy increasing with time. The sum of these two curves will be seen to be fairly flat around the minimum.

If the overhaul was delayed until, say, 30 months, then the accumulated cost of lost energy would have reached $135 × ½ × 30² = $60 750. At 27.2 months, the cost is $135 × ½ × 27.2² = $49939. The cost of delaying overhaul is thus the difference, $10811.

Note that this calculation is only correct if the wear progresses at a uniformly increasing rate with time. Information may not be available to make any other assumption, but decision makers have to start somewhere. Other formulae for rates of change which are not linear (6).

Note that some relatively small pumps may never justify overhaul on savings in energy use alone, but may be justified on reduced plant production rate.

6.4 Pump deterioration does not affect production, at least initially: variable speed controlled pump. For a pump where the speed is varied to meet its desired duty, the effect of wear on power required is much more dramatic than for the case of a constant speed throttle controlled pump. This is because the power increases in proportion to the speed ratio cubed.

Unless the pump output is limited by the pump reaching its maximum speed, or by its driver reaching its highest allowable power output, then no production will be lost. However, power consumed will increase more dramatically for a given wear state than for a constant speed pump.

To estimate the power required in the worn state, the Head-Flow curve must be drawn for the current higher speed in the new condition. Select a Head-Flow point on the original new condition curve, and correct it to the higher speed: multiply the Flow by the speed ratio, multiply the Head by the (speed ratio)². Repeat this for some other points at flows above duty flow to draw the new condition Head-Flow curve.

Follow the same method and calculations as before to find the time for overhaul for minimum total cost. The operating point is projected from the worn curve to the new curve at the same speed as the worn curve.

This optimisation approach can also be applied to any item of plant where deterioration results in loss of efficiency, and is a valuable tool for maintenance engineers and managers in their role of managing assets to provide capacity for production.

7. OPTIMISATION USING SHUT-OFF HEAD TEST RESULTS

The shut-off head test information can also be used to estimate power used in the worn state, and do the optimisation calculations as explained in 6.3 and 6.4 above.

Head-Power-Flow characteristics in the "new" state are needed as before, and the operating point must be known. Note the power required at operating point as before.

Make an overlay trace of the Head-Flow curve in the new condition. Place it over the "new" curve and move to the left horizontally until the curve cuts the Head axis at the value of shut-off head obtained on the test. The trace is now in the position of the "worn" Head-Flow curve which is being experienced. Exactly the same process can be followed as explained above.

8. THE CASE OF THE WORN VERTICAL MULTI-STAGE PUMP

Here is an example of the use of both vibration and performance analysis to help solve a case where a pump's vibration was excessive. This pump is a 9-stage mixed flow pump, rated at 70 L/s @ 155m, at a speed of 1480 r/min. Only the motor and the pump discharge piping are visible above the plant floor, with the base flange bolted to the floor.

For condition monitoring by performance analysis, a permanently installed orifice plate flowmeter is used. An ultrasonic flowmeter shows close agreement. Vibration measurements are made in the horizontal and axial directions on the motor flange.

The pump had not yet been added to the routine vibration program. High vibration was reported, so vibration measurements were made, showing 14 mm/s horizontal vibration, all at 7.5Hz. This level is not acceptable for reliable service.

Lower down the casing, the vibration was half this level, but mostly at 50Hz, with some at 7.5Hz. No vibration was evident at the running speed frequency of 25Hz.

More frequent monitoring commenced, and a week later, the vibration had increased to 22 mm/s, all at 6.5Hz. The performance at the usual operating point had also decreased by 18%.

As an alternative flow path was available without reducing production, albeit reducing plant efficiency, the pump was removed for overhaul. The impellers and sealing rings were all damaged beyond repair, and the shaft bent.

After repairs, the highest vibration level was below 1 mm/s, and at 25Hz and 50Hz frequencies, and tested performance had also returned to datum level. Routine measurements were commenced at 3 month intervals.

It is suspected that the base mounting bolts had worked loose, sufficiently to allow the pump to vibrate at its natural frequency of about 7.5Hz. Consequent internal damage resulted.

9. THE CASE OF THE VIBRATING BOOSTER FEED PUMP

10. REFERENCES

 

The paper was originally presented at the International Conference of Maintenance Societies 1998 - ICOMS 98 - in Adelaide, Australia.