Any mobile crane lift that involves two cranes working in tandem must be treated as a complex lift. In the United Kingdom, the lift must be performed in accordance with British Standard 7121. To complete such a lift all personnel must be experienced, trained, qualified and competent to complete their role during the complex lift. An Appointed Person (AP) is in charge of the overall lifting operation and must produce the lift plan. Note that the AP does not have to be on site during the time of the lifting operation, they may delegate authority of the lift to a lift supervisor. However, even though they have delegated onsite supervision of the lift, they still assume full responsibility for the lift plan and this is the document that would be scrutinised if the lift had to fail.

The first role of the AP is to select the most appropriate cranes to perform the task. To do this, the AP must consider many variables:

• What is the ground condition like, i.e. is the ground suitable to be able to withstand the load of the crane whilst the crane is lifting the heaviest load at the maximum working radius.
• What about access to the site. The heavier the load that the crane is designed to lift will mean that the overall dimensions of the crane will be larger. Will the crane actually be able to get to the site? Think about any weak bridges the crane needs to drive over or height of the bridge if the crane has to drive under the bridge, is the terrain suitable for your crane selection, are there any items onsite that have to be moved to allow crane access, must the crane travel on any narrow roads?
• What is the weight of the load and what are the maximum and minimum radii that the crane has to operate to lift the load and position the load at its final location?  What is the length of boom required to lift the load and place at its final destination? Does the crane have to travel on wheels to position the load or will the crane be working on block duties (i.e. with outriggers down)?
• Will the selected crane still be able to lift the load when the actual load that the crane can lift has been de-rated due to the weight of the extended boom, the weight of the hook block and the weight of any lifting accessories?

Fortunately, to complete this current task, both cranes could be situated on hard standing.  This hard standing provided adequate ground bearing properties for the loads to be lifting with the cranes working on block duties.

The loads where transported from one part of the client’s site to their final destination. This route did include a pipe track that didn’t give much clearance for the loads as they were transported by low loader. The loads just cleared the underside of the pipe track by 150 mm. In addition, several pallets of whiskey drums had to be cleared by the client’s fork lift truck driver in order to get the loads as close as possible to their final destination so the cranes to perform their tasks on block duties.

As discussed the loads were 11 tonnes and the max-working radius was 12 metres and the min working radius was 6 metres.

Crane Capacity Guide Chart

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The chart above (provided by Ainscough) details possible cranes that may be used to lift specific loads at specific radii. Note that the individual crane specs must be examined to ensure that the crane can still lift the load at max and min radii and once the de-ratings have been calculated. Using, the chart and crane specs it was decided that a 90 tonne Liebherr would be suffice to lift the 11 tonne loads at 12 meters. A point to note is that this crane would fail with the load at 14 meters, therefore illustrating the criticality of selecting the right crane. As a 90 tonne crane could not be provided, a 100 tonne crane was sent to site. Normally, a larger crane will be able to complete the task, though we must still examine the configuration of the crane and remember that the ground bearing loads will increase and the dimensions of the crane will probably be greater.

Before siting the cranes the AP must also consider all the proximity hazards around about the working envelopes of both cranes. These include:

• Overhead power lines including wooden telegraph poles and steel electricity pylons. The crane must be sited at a minimum of 9 meters from a wooden telegraph pole and 15 metres from a metal electric pylon. Permission would have to be sought from the relevant power companies if the task had to be performed inside these designated distances.
• Any trenches dug beside the crane outriggers must be considered. As a rule of thumb, and subject to ground conditions, the crane must be kept away from the trench at a distance equal to the depth of the trench plus one metre.
• Underground services must be considered along with overhead pipelines.
• Consider if the load has to be transferred over the top of any buildings or is the lifting operation to take place adjacent to dangerous chemicals, live gas pipelines, low/high voltage power cables or any other plant/personnel.

 
Now that we’re happy with the siting of the crane and we have also taken a look at possible proximity hazards we have to consider some more risks associated with the job. For example how are we going to attach and remove our lifting accessories? For this particular job we hired a man cage and harnesses to allow us to safely remove the lifting accessories using the 50 tonne crane after the 100 tonne crane had positioned the load on top of the supporting steelwork bases.

Another major concern is keeping other personnel away from the work whilst the lifting is progressing. The working area should be cordoned off and a road marshal appointed to keep both personnel not associated with the works and other site traffic away whilst the lifting operation is taking place.

Mark Rae removing lifting accessories.

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As with all lifting operations the lifting accessories should be checked before and after use. Also, the load should be able to take it’s own weight, if not, the load may have to be supported, for example, placed into a container designed for lifting the load. Tag lines should be also used to control loads that may start to sway in the wind to make sure that no damage can be done to surrounding plant, personnel or buildings during the lift. The signaller should be easily identifiable, usually by wearing a different coloured hi-viz vest to the rest of the personnel. Only one signaller at anytime should be giving signals to the crane drivers. These signals should be as designated by BS 7121.

BS 7121 Hand Signals.

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Note, before the lift begins, the load should be securely fastened to the lifting accessories and the crane should take the strain before the load holding down straps are removed from the low loader. This is of extreme importance as people have been killed in the past by heavy unsecured loads rolling off trailers and low loaders before the lifting operation commences. Once the load is secured to the crane, the strain is taken and the holding down straps removed, a test lift should be performed. The test lift is to establish that the load is going to lift evenly with the weight of the load evenly distributed. This process can be tricky if the load’s centre of gravity is not easily identifiable. Very often chains with shortening clutches have to be used and adjusted until a perfect lift is achieved.

We have discussed a lot of the standard risks involved with lifting, however during a tandem lift with two or more cranes, greater risks have to be identified and dealt with. When conducting a tandem lift it is best practice to:

• Conduct the lift during stable low wind conditions.
• Move cranes using only one motion at a time.
• Try and avoid any slewing motions.
• Jib up rather than jib down. Don’t forget that jibbing down increases the load radius and therefore increase the load placed on the cranes.
• Carry out all crane movements at a slow pace.
• Make sure that the cranes’ axis remain fully aligned with each other during pick and carry operations.
• If the signaller cannot observe all necessary locations, additional trained personnel should be positioned to feedback to the signaller any potential problems.

 
Due to the tight setup of the cranes for this particular operation, both cranes were set up perpendicular to each other. The 100 tonne crane did perform a minor slewing action during the lift. This slewing action was towards the 50 tonne crane and did not exert any twisting action onto the other crane’s boom and also decreased the load radius of the 50 tonne crane therefore reducing the load experience by the smaller crane.

First of all, may I say, the root cause can be excruciatingly difficult to pinpont. Let me begin by telling you of one such problem that I had with high interstage pressure that was causing the machine safety valves to pop (as designed to do!) and also causing the machine to trip. I had just overhauled the machine with a new set of reconditioned suction and discharge valves on both first and second stages. I had also replaced the piston rider and piston rings. I had replaced all of the packing in the stuffing box and intermediate packing. The compressor used to be lubricated, however it had been converted to run as a non-lubricated machine (and I must say that the mean time between failures had improved substantially). Although, the cylinder heads were now both non-lubricated the crank was still lubricated – so I had also changed out the oil in the frame and fitted new duplex filters. Finally, one of the actuating piston rods on the second stage inlet valves was bent, so I fitted a new oem (original equipment manufacturer) part that I got straight from the on site stores stock.

All was looking fine, I checked my piston standouts, crosshead shoe clearance, piston rod run out and clearance between the double acting piston head and the frame and outer ends (on both stages). So why was may pristine machine crashing out on high interstage pressure? When such a thing happens the first thing I do is head to the archive cabin and pull out the compressors manufacturer’s manual. A good manual is normally brimming with information regarding maintenance, operation and troubleshooting tips to help get one’s head into the operating characteristics of the machine – pinpointing exactly what makes the machine tick. Therefore, I browsed the common causes of high interstage pressure.

A higher than normal interstage pressure indicates problems in the cylinder that follows the interstage cooler. As this is a two stage compressor, the problem had to be with the second stage cylinder. If the compressor is loaded, then the problem normally lies with the suction valves. If the interstage pressure is high whilst the cylinder is unloaded – then the problem normally lies with the discharge valves. From my discussions with the Duty Shift Engineer, I was informed that the problem was occuring once the compressor was loaded. Therefore, I looked back at all the maintenance that I had performed on the machine and noted that I had changed out a second stage piston unloader. I quickly examined the piston rod with the piston rod that had been removed from the machine and there lied the problem. The new piston rod was 6mm (1/4″) longer than the piston rod that I had removed from the machine. Therefore, the piston was holding the suction valve open, allowing the pressure to flow back to the interstage and blow the safety valve. I cut the piston rod down to size, installed and started the compressor. This time all pressures were fine and the compressor ran as expected.

Basically, the liquid flows through the suction piping and arrives at the suction nozzle. Note that a pump cannot suck the liquid into the pump, the liquid must have sufficient energy to allow the pump to take the energy and work the liquid’s energy. This concept is called Net Positive Suction Head (NPSH).

Positive displacement pumps have some form of movable enclosure. They’re many types of PD pumps; some consist of rotating gears, others may be as simple as a pulsing diaphragm. The fluid enters the suction nozzle at a low pressure zone, the fluid is then mechanically transported through the pump to the discharge nozzle. This mechanical movement basically contracts and as a fluid can’t be compressed the pressure inside the pump’s housing increases and the fluid is expelled into the discharge piping. The head or pressure that a PD pump can generate is predominately a function of the tolerances and strength of the pump components coupled with the thickness of the pump case.

A centrifugal pump contains an impeller and volute. The impeller is attached to a shaft. The shaft is connected to a driving unit, normally but not exclusively, an electric motor. The fluid enters into the eye of the impeller and is captured between the impeller blades. The blades impart velocity to the fluid as the fluid is transferred from the eye to the outside diameter of the impeller. As the fluid accelerates, a low pressure zone is created in the eye of the impeller. This principle is known as the Bernoulli principle – as velocity increases, pressure decreases. Therefore, the fluid leaves the outside diameter of the impeller with increased velocity and as the fluid hits the internal casing of the volute the fluid stops abruptly and the velocity is converted into pressure – the Bernoulli Principle in reverse. As the impeller is spinning, rotary velocity is also occuring and the fluid is transfered around an ever increasing escape channel within the volute. The pathway is increasing and consequently the rotary velocity is decreasing adding additional pressure to the fluid – Bernoulli’s Principle yet again!

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The pressure and flow that a centrifugal pump can deliver is mostly governed by the diameter of the impeller and the speed of the motor. As we have seen from the David Brown DB37 pump above, several impellers can be used in sequence to build up to the required pressure in several stages.

Force (F) is equal to Pressure (P) multiplied by Area (A). So F= P x A.

Therefore, to work out pressure we divide the Force (F) by the Area (A). Pressure = F/A.

When we apply pressure to the surface of a liquid the pressure is transmitted uniformly across the surface and also through the liquid to the walls of the vessel. The pressure can be expressed by a number of imperial and metric units, the most popular bar (metric) and pounds per square inch (psi) – imperial.

Atmospheric Pressure (ATM)

Atmospheric Pressure (ATM) is the force exerted by the weight of the atmosphere on a unit of area. ATM is equal to 14.7 psia at sea level. As elevation rises above sea level, the atmospheric pressure reduces.

Absolute Pressure (psia)

Absolute pressure is the pressure measured from a zero pressure reference. Compound pressure gauges record absolute pressure and absolute pressure is 14.7 psia at sea level.

Gauge Pressure (psig)

Gauge pressure is the pressure indicated on a simple pressure gauge. psig equals psia – ATM.

Vacuum

Understanding vacuum can lead to confusion very often as most people think that vacuum is expressed as a negative psi. This is true with regards to a simple pressure gauge which will record vacuum as a negative psig. However, vacuum is any pressure less than atmospheric pressure. Therefore, anything less than 14.7 psia is a vacuum. Compound gauges record vacuum as a positive psia. When we listen to the daily forecast we often see areas of low pressure described in millibars (1000 millibars is atmospheric pressure). Vacuum can be expressed in many ways. It is probably easiest to think of vacuum in pumping terms as a positive number less than 14.7 psi.

Pump manufacturers often talk in terms of pump head. The basic formula for pump head is Head (H) equals Pressure (psi) divided by Density (D). To convert pressure into head we can use the following equation:

Head ft = (2.31 x Pressure psi ) divided by (specific gravity). The 2.31 is a conversion factor.

Specific Gravity

When we compare the density of a liquid with the density of water, we are detailing the liquid’s specific gravity. The formula for specific gravity is:

Specific gravity = density liquid / density water (i.e. water 60 degrees Fahrenheit at sea level).

Water has a specific gravity of 1.0. Other liquids are either denser or lighter than water. The specific gravity affects the pressure in relation to the head as illustrated in the formula above.

The above is the reason why pump manufacturers sell pumps that will produce a certain feet of head. The pump manufacturer will not know the final liquid that the pump will be pumping therefore, as the above formula details, the working pressure of the pump will vary depending on the specific gravity of the liquid to be pumped.

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Of course, a good condition monitoring system should hopefully pinpoint bearing wear through oil and vibration analysis, however, as we all know even with the best detection and preventative maintenance programmes – failure can still occur. One of the things we always try and do is to attempt to pinpoint the root cause of our bearing failure.

• Foreign Matter: When abrasive particles get into the oil and start to circulate around the bearing housing.
• Bearing Misalignment: An all too common problem that can often be spotted by examining the bearing raceways, the ball bearing track will display signs of the balls oscillating from side to side. The shaft should be checked by lasers or clock gauges to make sure it isn’t bent. The shaft should be true, as should any shaft shoulders, spacers and locking nuts.
• Machine Train Misalignment: Machine train misalignment is indeed a whole subject in its own right. Simply put, if your pump and motor train are misaligned, then you will experience increased temperatures and vibrations and your bearing life will plummet dramatically.
• Bearing Lubrication: Getting the bearing lubrication wrong can lead to overheating, subsequently leading to excessive wear. So selecting the wrong lubrication can be detrimental, also when the correct lubrication is selected the oil level must be correct. Oil level should be addressed by the implementation of constant oil levellers, that when correctly set up will keep the oil and the right level, especially at times when oil is lost through leakage. The oil should be visually checked by competent plant operators to ensure that lubricants have not been contaminated or broken down due to oxidisation or exposure to atmospheric conditions.
• Bearing Fatigue: Fatigue develops due to the magnitude of force exerted on the bearings and the frequency that the forces are repeated. The bearing’s rolling elements spread the metal in front of them as they roll, putting the metal components in compression and tension with this repeated action resulting in the flaking of the metal.
• High Temperatures: Bearing lubrication can be greatly affected by high temperatures. The formation of carbon can occur and lubricating elements are lost. The carbon formed can cause increased friction and has a tendency to jam the bearing. Also, the bearing metals are affected by the high temperatures. A change in the molecular structure of the metal can alter the metal’s strength and hardness characteristics substantially. A final point to note is that high temperatures have a tendency to expand the bearing metals leading to a reduction in bearing clearances resulting in increased preloading within the bearing. Note that this can also occur at low temperatures when dealing with cryogenic pumps.
• Bearing Corrosion:Ball and roller bearing surfaces are finished to a high standard in order to meet the tolerances required to perform with high efficiency. The bearing high quality surface finishes are subjected to corrosion by a whole range of different mediums; water, acid, broken down oils and greases, condensation from excessive temperature reversals. Corrosion causes abrasion and can account for excessive wear within bearings.
  

The following video details the 3D Modelling of a Scottish Distillery.