A popular misconception about hydraulic cylinders is that if the piston seal is leaking, the cylinder can creep down.
Fact is, if the piston seal is completely removed from a double-acting cylinder, the cylinder is completely filled with oil and the ports are plugged, the cylinder will hold its load indefinitely – unless the rod-seal leaks.
What happens under these conditions – due to the unequal volume either side of the piston, is fluid pressure equalizes and the cylinder becomes hydraulically locked. Once this occurs, the only way the cylinder can move is if fluid escapes from the cylinder via the rod seal or its ports.
If you grasp the theory at work here, you’ll probably realize there are a couple of exceptions. The first is a double-rod cylinder – where volume is equal on both sides of the piston.
And the second is when a load is hanging on a double-acting cylinder. In this arrangement, the volume of pressurized fluid on the rod side can be easily accommodated on the piston side. But as the cylinder creeps a vacuum will develop on the piston side – once again due to the unequal volumes – and depending on the weight of the load, this vacuum may eventually result in equilibrium that arrests further creep.
This is not quite the end of the story though, but it’s important to at least grasp this theory before we move on.
Not withstanding the two exceptions mentioned above, if a double-acting cylinder’s service ports are blocked – by a closed-to-actuator or cylinder spool and the piston seal does bypass, pressure will eventually equalize on both sides of the cylinder. As already stated, at this point a hydraulic lock is effected and no further creep can occur – unless fluid is allowed to escape from the cylinder or the cylinder circuit.
But because of the loss in effective area – due to pressure now acting on the rod-side annulus area, the static pressure in the cylinder must increase to support the same load.
For example, if the load-induced pressure on the piston side of the cylinder was 2,000 PSI and zero on the rod side when the DCV closed, assuming no leakage past the spool, the equalized pressure may be 3,000 PSI – depending on the ratio of the areas.
But what if this circuit has a service port relief valve set at 2,500 PSI? As pressure starts to equalize across the piston seal and the increasing static pressure on the piston side of the cylinder reaches the cracking pressure of the port relief, the cylinder WILL creep down.
While the root cause of the problem IS the leaking piston seal, the physics that applies is fundamentally different to what many people believe. And if you understand the theory, you can see how the humble pressure gauge can be extremely useful when troubleshooting cylinder creep.
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A few months ago, I was involved – in a supervisory capacity – in the planned change out of components on a hydraulic machine.
The design of the machine’s hydraulic power unit was the all-too-common, cheap and nasty, everything mounted on the tank lid variety. You know the ones – electric motors mounted vertically, with pumps submerged in the tank. They’re cheap and easy to build, but an eternal pain in the butt for anyone who has to work on them. I could rant about this … but I’ll save it for another day.
The scope of work included changing out a tandem gear pump submerged in the tank. On this machine, a 15 minute job – after you’ve spent two hours disconnecting everything to enable the tank lid to be lifted.
I was offsite when this pump was changed, but the technician doing the work was experienced and knew what had to be done. After the shut down was completed, the same technician was scheduled to re-commission the machine. But due to conflicting commitments he became unavailable.
It was early in the afternoon when I got the call requesting I go to site to supervise start-up. Thinking this would be a case of push the button and watch everything behave as it should, I left the office in my ‘civilian’ clothes, took my good car and no tools. BAD mistake.
Commissioning didn’t begin well. The rear section of the tandem gear pump that had been replaced, charged an accumulator. But we weren’t getting any charge pressure on start-up. The direction of rotation of the electric motor was correct. But what about the rotation of the new pump? Bit hard to tell when it’s submerged in the tank.
After speaking with the technician on my cell, I was reasonably confident the new pump’s rotation was correct. But just to be sure, I decided to try reversing the rotation of the electric motor. No change.
Now what? I don’t have a flow-tester with me and it’s a major job to lift the tank lid. It appears that new pump is either faulty or not priming. Unlike a vane pump which needs a head of oil on its outlet for the vanes to ‘throw’ and start pumping, a gear pump is normally considered self-priming – especially when its submerged in hydraulic oil.
To give the new pump the best possible chance of priming, I topped off the oil level at maximum. Still no charge pressure after running for a minute. In the absence of a flow-meter, I decided to lift the electric motor off its bell housing, disconnect the pipes from the pump’s outlet penetrations in the top of the tank and turn the pump by hand.
This was revealing. The front section was pumping, but the rear section wasn’t. Joy. The rear section was faulty. I continued turning the pump by hand, while I wondered what the chances were of this happening, and contemplated the work involved in lifting the tank lid.
Just as I was about to give up and call for reinforcements, I noticed oil was being displaced from the rear pump’s outlet penetration in the top of the tank. It WAS pumping. I reconnected the plumbing, dropped the electric motor back into position and bingo. We had accumulator charge.
Had I been told this story about a gear pump that wouldn’t prime itself – I would have struggled to believe it. Yes it IS strange … but true.
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In a previous article, I wrote about the advantages of defining your maintenance and reliability objectives for a piece of hydraulic equipment before you even order it. In response to this article, I received the following from one of our members:
“As an engineer for a heavy equipment manufacturer, I have to take exception with your advice in the last newsletter. I believe selecting a hydraulic fluid, and then requesting the equipment manufacturer to design around that is putting the cart before the horse.
It makes much more sense for the equipment manufacturer to design the system for temperature, life, component availability, cost, and the million other things that go into machine design.
Would you tell Ford that you have a set of brake pads, and would like a car designed that could use them? No, you rely on the manufacturer to provide a system, and take their recommendation.”
Hmmm… looks like I’m in trouble – again. I feel a bit like David, being pounded by Goliath. This member didn’t reveal which OEM he works for… but the mind boggles with possibilities.
Needless to say, I don’t agree with my colleague’s assertion that the reliability-based strategy I advocated in last month’s newsletter is “putting the cart before the horse”. Let me explain with an example:
Say I am about to acquire a 25 ton hydraulic excavator. And let’s say for example, this machine is fitted with Rexroth pumps and motors.
According to the pump manufacturer, optimum performance and service life will be achieved by maintaining oil viscosity in the range of 25 to 36 centistokes. I also know that in my location I expect to use a VG68 weight hydraulic oil and the brand of oil I use has a viscosity index of 95.
This being the case, Rexroth are telling me – indirectly of course, that if my new machine runs any hotter than 70 Celsius the performance and reliability of their pumps and motors will be less than ideal. Not only that, with 70 Celsius as the maximum operating temperature, the oil will last longer, the seals will last longer, the hoses will last longer and almost every lubricated component in the hydraulic system will last longer.
So being the sophisticated buyer that I am, I say to the OEM – before I order the machine: “I expect ambient temperatures at my location as high as 45 Celsius and under normal conditions (no abnormal heat load in the system) I want this machine to run no hotter than 70 Celsius. If you deliver it to site and it runs at 85 Celsius (or whatever) on a 45 Celsius day, then you’ll have a problem on your hands.”
I’m not suggesting this is in the interests of the OEM – clearly it’s not. It’s going to make their life more complicated and cut into their after sale revenue. No, it’s totally in the interests of the guy signing the checks to keep the machine running. Luckily for all the OEM’s out there, very few machine buyers will approach a new equipment acquisition with this level of sophistication.
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During a recent Hydraulic Breakdown Prevention Blueprint seminar, one of the attendees, a maintenance manager for a large open-cut mining operation, mentioned that he was considering upgrading the filtration on their fleet of hydraulic mining shovels – to achieve a higher level of fluid cleanliness.
There’s lots of documented evidence to suggest that increasing hydraulic fluid cleanliness increases the service life of hydraulic components – all other things equal. Whether such an initiative would yield an acceptable return on the investment required, for the machines in question depends on a number of variables, which I don’t have room to go into here. But this got me thinking about a bigger issue.
The maintenance routines I teach in my books and Workshops are about equipping people with the knowledge the need today to optimize the reliability and service life of the hydraulic equipment they have right now. And that’s fair enough – it’s rarely helpful and not very instructive to tell someone what they should have done yesterday. With this is mind, one of the exercises we do during my one-day Workshop is carry out a maintenance and reliability audit on an existing hydraulic machine.
Even though equipment design and equipment maintenance are often viewed in isolation, the reality is, certain aspects of hydraulic machine design have a significant impact on the machine’s operating cost and reliability, and ultimately, its life-of-machine cost.
Over lunch, the same maintenance manager mentioned that his mine is starting to think about the replacement of their aging fleet of hydraulic shovels. And it occurred to me, the best time to carry out a maintenance and reliability audit on a piece of hydraulic equipment is BEFORE you buy it.
By starting with the end in mind, you get the maintenance and reliability outcomes you desire – before the machine even gets delivered. For example:
You specify the contamination control targets you want to achieve based on your reliability objectives for the piece of equipment. And instruct the manufacturer to deliver the machine appropriately equipped to achieve these targets.
Based on the weight and viscosity index of the hydraulic oil you plan to use, you determine the minimum viscosity and therefore the maximum temperature you want the machine to run at. And instruct the manufacturer to deliver the machine equipped with the necessary cooling capacity, based on ambient temperatures at your location. Rather than accepting hydraulic system operating temperatures dictated by the machine’s ‘design’ cooling capacity – as is the norm.
If you don’t think the viscosity/temperature issue is this important – you’re mistaken. Lubrication failure resulting from low fluid viscosity is one of the biggest causes of premature failure in hydraulic components. If you’re not on top of this issue it could be costing you big.
And we could continue by specifying things like flooded inlet for all pumps and so on. But you get the idea.
So the next time you or the company you work for are looking to acquire hydraulic equipment, begin with the end in mind. Define your maintenance and reliability objectives in advance and make them an integral part of your equipment selection process.
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When a hydraulic system sees a spike in pressure it won’t necessarily blow up with a bang. But damage can occur in a number of ways. In fact, a single pressure spike of sufficient magnitude can render a piston pump or motor unserviceable. Here’s how:
In axial and bent axis piston pump and motor designs, the cylinder barrel is hydrostatically loaded against the valve plate. To maintain full-film lubrication between the rotating cylinder barrel and the stationary valve plate, the hydrostatic force holding them in contact is offset by a hydrostatic force acting to separate the parts. This is achieved by making the effective area of half the total number of piston bores slightly larger than the effective area of the pressure kidney in the valve plate.
The higher the operating pressure, the higher the hydrostatic force holding the cylinder barrel in contact with the valve plate. However, if operating pressure exceeds design limits, the cylinder barrel will separate from the valve plate.
Design geometry prevents a perfect alignment of the opposing hydrostatic forces. This misalignment creates a twisting force (torque) on the cylinder barrel. During normal operation, this torque is supported by the drive shaft (axial designs) or center pin (bent axis designs). If operating pressure exceeds design limits, the magnitude of the torque created causes elastic deformation of the drive shaft or center pin. This allows the cylinder barrel to tilt, bearing hard against the outlet side of the valve plate and separating from the inlet side (exhibit 1).
Exhibit 1. Separation of cylinder barrel and valve plate
due to overpressurization (Bosch Rexroth)
Once separation occurs, the lubricating film is lost and the resulting two-body abrasion damages (scores) the sliding surfaces of the cylinder barrel and valve plate. Erosion of the kidney area of the valve plate can also occur as high-pressure fluid escapes into the case at high velocity. This surge of flow into the case can cause excessive case pressure, resulting in shaft seal failure.
Note also that separation can also occur at operating pressures within design limits due to distortion (loss of flatness) of the valve plate, over-speeding or excessive wear of the cylinder barrel drive-spline in axial designs.
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Rupert Murdoch, the boss of global media giant News Corporation was a neighbor of ours where I grew-up. Not that my family was particularly well off. It’s just that my father’s farm happened to be situated close to a group of “sheep stations” the media mogul owned. But compared to the 300,000 acres Mr Murdoch controlled, Dad’s land holding was modest indeed.
In 1981, just in time for the wheat harvest, Dad took delivery of a new combine harvester. It was one of many he owned over the years, but this one was different. It was the first I’d seen equipped with a hydrostatic transmission for the ground drive. The infinitely variable and step-less control afforded by a hydrostatic transmission was quite an advance over the mechanical gearbox with a variable speed input used in earlier models.
Anyway, in its second season the hydrostatic transmission gave trouble. Downtime during harvest was always guaranteed to elevate Dad’s stress level to 11 out of 10. And that wasn’t a pretty sight. I didn’t know much about hydraulics then.
Of course in the 25 years since, I’ve accumulated a bit of knowledge on hydrostatic transmissions. And an issue that is often overlooked and one that came up in a job I was involved in recently, is the combined effect of fluid compressibility and the ‘accumulator effect’ of conductors (the increase in volume of a hose or pipe as pressure increases).
When a hydrostatic transmission is subject to a sudden increase in load, the motor stalls instantaneously and system pressure increases until the increased load is overcome or the high pressure relief valve opens – whichever occurs first.
While the motor is stalled, there is no return flow from the outlet of the motor to the inlet of the pump. This means that the transmission pump will cavitate for as long as it takes to make-up the volume of fluid required to develop the pressure needed to overcome either the increased load or the high-pressure relief valve. How long the pump cavitates depends on the output of the charge pump, the magnitude of the pressure increase, and its influence on the increase in volume of the conductor and the decrease in volume of the fluid. This is illustrated in the following example.
A hydrostatic transmission operating the drill head on a drill rig is delivering a flow of 35 GPM at a pressure of 1000 PSI. A sudden increase in load on the drill bit instantaneously stalls the motor until sufficient pressure is developed to overcome the increase in load, which for the purposes of this example is 3000 PSI.
In order to increase system pressure from 1000 PSI to 3000 PSI, the transmission pump must make-up additional volume, due to the compression of the hydraulic fluid and the volumetric expansion of the high-pressure hose between the pump and the motor. But because the motor is momentarily stalled, there is no return flow from the outlet of the motor to the inlet of the pump. The only fluid available at the inlet of the transmission pump is 7 GPM from the charge pump, which is around 80% less than required!
In this example, the high-pressure hose between the pump and motor is SAE 100R9AT-16, 36 feet long. The volumetric expansion of this hose, due to the increase in pressure, is 9.7 in³ and the additional volume required due to compression of the fluid within this hose is 2.8 in³. Therefore the total, additional fluid volume required to increase the operating pressure from 1000 to 3000 PSI is 12.5 in³ (9.7 + 2.8 = 12.5).
To calculate the time taken for the operating pressure to increase from 1000 to 3000 PSI, which is equivalent to the length of time the transmission pump will cavitate, we divide the required make-up volume (12.5 in³) by the volume available from the charge pump per second (27 in³). In this example, the transmission pump cavitates for 0.46 seconds every time a sudden increase in load demands an increase in system pressure from 1000 to 3000 PSI (12.5 ÷ 27 = 0.46).
This problem occurs in applications where there are sudden fluctuations in load on the transmission. Typical examples include drill rigs, boring machines, and cutter wheels on dredgers. The solution involves increasing available charge volume – usually through the installation of an adequately sized accumulator in the charge circuit.
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Would you part with $50 to save $70,000? This is a ‘no-brainer’ for most of us. Well, here’s a story for you:
I recently conducted failure analysis and a reliability audit on a 300 kilowatt hydrostatic transmission. The hydraulic system was running a synthetic ester, biodegradable hydraulic fluid. This $45/gallon hydraulic fluid had been destroyed in under 12 months and a set of pumps shortly after. So with $20,000 of hydraulic fluid and $50,000 of pumps ruined in short order, my client was understandably wondering what went wrong.
The system was built and installed by a reputable distributor. From a hydraulic engineering perspective the circuit was adequately designed and the system well built. But from a maintenance and reliability perspective it left a lot to be desired. My client, the end user, didn’t have a lot of experience with hydraulic equipment and was reliant on the company that built the system to guide them on its maintenance.
An oil analysis program had been set up, but it seems the only thing anyone was taking any notice of was particle contamination. If you’ve been reading this newsletter for a while, you’ll know there’s a lot more to hydraulic equipment reliability than just monitoring and controlling hard particle counts.
All the warning signs pointing to oxidative failure of the oil went unnoticed. The oil started polymerizing, coating internal components with sludge. These gum-like deposits block lubrication passages, reduce heat transfer and cause valve stiction. The oxidation process diminishes foaming resistance and air release properties of the oil, which in turn causes damage through aeration and gaseous cavitation.
By the time I got involved, the original set of pumps had already failed and the hydraulic fluid had a TAN of 10 and water content of 6,500 ppm. So I set about establishing the root cause of failure and instigating measures to ensure it didn’t happen again.
It’s not rocket science. But the problem for my client was the information they needed to prevent this maintenance disaster is not widely available. You won’t find it in the machine manual and it’s not taught in everyday, how-it-works, hydraulics classes.
As a consequence, most owners, operators, mechanics, technicians and engineers are clueless when it comes to proper maintenance of hydraulic equipment. It’s not their fault – they just haven’t been given the opportunity to learn.
The principles I applied and the procedures I put in place to prevent a reoccurrence of the failures described above. Had my client had this information and applied it diligently at the outset, they could have saved more than $70,000 – and a lot of downtime and aggravation.
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A recent client had a set of pumps worth fifty grand fail after achieving only half of their expected service life. And they wanted some answers. At the initial meeting the client opened proceedings with a brief history on the machine, an account of the events leading up to the failures and then pushed a stack of oil analysis reports across the table.
After I finished taking notes on what I’d just been told, I fired off my first question:
“What is the system’s normal operating temperature?”
Stunned silence. Client shrugs his shoulders.
“O-K … what’s the system’s usual operating pressure range?”
Blank look from client. “Err dunno … we don’t monitor either of those things.”
At the end of the meeting we took a walk through to the control room. Turns out, both operating pressure and temperature were displayed on the default PLC screen – albeit along with a lot of obviously more important production information. Say no more.
But could YOU answer these two simple questions about the “vital statistics” of your hydraulic equipment? If not, I strongly recommend you make the effort to get to know your equipment better.
This information is easy to collect, can give valuable insight to the health of your equipment and is essential data if failure analysis is required. Here’s how I recommend you do it:
First you need an infrared thermometer, also called a heat gun. If you don’t have, you’ll need to invest around 100 bucks to get one.
Next, using a permanent marker, draw a small target on the hydraulic tank below minimum oil level and away from the cooler return. Label it 1. This marks the spot where you’ll take your tank temperature readings. The idea behind these targets is that regardless of who takes the readings they’ll be taken from the same place each time.
If the system is a closed circuit hydrostatic transmission, mark a convenient location on each leg of the transmission loop and number them 2 and 3. Skip this step for open circuit systems.
Next, mark a target on the cooler inlet and outlet and number them 4 and 5. This records the temperature drop across the cooler.
With that done, now draw up a table like the one below to record these temperatures and a few other essential parameters. Note that there is little point in recording the temperature across the cooler (4 & 5) if the fan isn’t running. And charge pressure is only relevant to closed circuit hydrostatic transmissions.
| Date |
02-06-07 |
08-06-07 |
| Time |
1615 |
0640 |
| Ambient Temperature |
11 |
39 |
| Operating Pressure |
3000 |
2900 |
| Charge Pressure |
270 |
250 |
| (1) Tank Temperature |
30 |
68 |
| (2) Transmission A |
41 |
60 |
| (3) Transmission B |
44 |
65 |
| (4) Cooler In |
|
64 |
| (5) Cooler Out |
|
53 |
| Fan On Yes/No |
No |
Yes |
I recommend you take readings on the hottest and coldest days of the year and on a couple of average temperature days in between. This provides a baseline of data. Beyond that, taking readings at regular intervals – daily, weekly or monthly, can provide early warning of system problems. And if the system starts to give trouble, taking a set of readings will reveal if it’s operating outside of its normal parameters.
In a previous blog post, I discussed the procedure I use when preparing hydraulic cylinders for storage.
In response to this article, one of our members sent in this question:
“One issue I feel you left out of the cylinder storage issue is the orientation question. How should a cylinder be orientated for short term or long term storage?
Our company repairs and evaluates cylinders and their associated failures. We try to provide solutions for breakdowns as well as repairing those that have failed. We evaluate and repair approximately 1000 cylinders annually. One common issue has been seal failure particularly in large pneumatic and hydraulic cylinders. We have found that allowing cylinders to lay flat has a direct effect on piston and rod seal failures. We have instituted a cylinder storage standard that adheres to your recommendations as far as ports plugged, rods wrapped but in addition mandates that all cylinders are stored vertically – in such a position as not to distort or place the weight of the cylinder on the rod seals. This verticality also helps, we feel, the piston seals.
I wouldn’t mind hearing your argument on this issue.”
Hmmm. I intentionally didn’t mention it because I didn’t want to do anything to perpetuate the myth. Because based on my experience, that’s exactly what it is. Two cases I was indirectly involved in come to mind. In both cases the cylinders in question were off 400 ton mining-size hydraulic excavators. We are talking here about cylinders that weigh between two and three tons. The piston rod typically weighs well over a ton by itself.
So you have a situation where big, expensive, high-pressure cylinders are suffering premature seal failures. In both cases, the machine operators sought the advice of “seal experts”. The recommendation of these supposed experts was to store the cylinders vertically.
Let’s consider the reality of this nonsense:
Someone drops a three ton cylinder with a closed length of four meters at your feet and tells you to store it vertically – so it doesn’t fall over and destroy itself – or worse still, kill someone. Not a five-minute job, but possible I suppose.
A truck arrives to transport the cylinder to a remote mine-site. The route consists of 1,000 miles of rough, unsealed road. Given you have gone to the trouble of storing the cylinder vertically in the warehouse, surely you must insist that it is transported in the same orientation? I mean, if it can’t be stored horizontally in a shed, then surely the pounding it is going to get if it’s laid down on the back of a truck will turn the seals into mush, right? The truck driver thinks you’re crazy but he doubles his rate and obliges anyway.
The cylinder arrives at the mine-site in the mandated vertical position. Trouble is it’s a stick cylinder so it’s orientation on the machine is horizontal. If the bearing bands on the piston and in the gland can’t adequately support the piston rod and prevent it from distorting the seals when the cylinder is sitting in a shed or bouncing around on the back of a truck, how on earth will it cope with the thrust developed when it goes into service on a 400 ton excavator?
Common sense would tend to suggest that if the bearing bands have sufficient area and are correctly tolerenced to adequately support load-induced thrust without distorting the seals, then surely they will cope with the static weight of the piston-rod in storage and any dynamic loading that may occur during transport?
Whether you agree with this assessment or not, you know troubleshooting is a process of elimination. So when seal failures continue to occur even after the cylinders have been stored vertically … well it’s safe to say that’s not the root cause of the problem. And that was the outcome in the two situations I mentioned above. No surprise to me.
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A question I’ve been asked several times in recent months by equipment owners, is the procedure for storing spare hydraulic cylinders for an extended period. So here’s what I recommend:
- Always store fully retracted.
- Store indoors in a clean, dry area.
- Smear the internal surfaces of eye/clevis bushes or bearings with grease – particularly if they’re steel.
- Protect any exposed chrome on the rod. Oil-impregnated tape such as Denso tape can be used for this purpose. Before applying, make sure the rod is fully retracted. If a product like Denso tape is applied to the rod when the rod is not fully retracted, subsequent retraction of the rod can result in damage to the rod seal.
- Plug the service ports with steel – not plastic, plugs or blanking plates.
- Consider filling the cylinder with clean hydraulic oil through rod-end service port. Particularly if it is an expensive, large diameter or high pressure cylinder. I say “consider” because there are a few issues to understand before you do this.
If the cylinder is not filled with oil it will obviously be filled with air. If this air is not perfectly dry, then as ambient temperature decreases the air can reach due point. This results in moisture forming on the inside of the cylinder tube. This can cause spot rusting and pitting of the tube surface, which will reduce the volumetric efficiency of the cylinder, the service life of the piston seal, and ultimately, the life of the tube itself.
Completely filling the cylinder with clean hydraulic oil prevents this from occurring, however there’s a major caution with this. It’s best illustrated by an example:
Say a cylinder is prepared for storage during the winter months. When the cylinder is filled with oil, the ambient temperature is 10 degrees Celsius. A year and a half later, during the middle of summer, the same cylinder is set down beside the machine to which it is to be installed. In the heat of the midday sun the temperature of the cylinder rises to 40 degrees Celsius. Assuming an infinitely stiff cylinder, the pressure of the oil in the cylinder resulting from the rise in temperature can be approximated by the formula:
p (bar) = 11.8 x (T2 – T1)
So the theoretical pressure of the oil in the cylinder is now: 11.8 x (40 – 10) = 354 bar or 5134 PSI! When it comes time for the unsuspecting mechanic to crack loose the blanks on the service ports … well let’s just say that’s more excitement that he signed up for.
That said, cylinders CAN be safely filled with oil for storage provided you:
- Check that the worst-case temperature rise in storage won’t result in a static pressure that exceeds the cylinder’s working pressure.
- Only fill the cylinder when fully retracted and ONLY through the rod-end port. This avoids potentially dangerous pressure intensification.
- Use service port plugs or blanks that are rated for the cylinder’s working pressure.
- Attach appropriate warning tags to BOTH service ports.
- Provide a means to check and vent any pressure before each of the service port blanks is removed. A simple way to do this is to fit each port blank with a pressure test-point. This enables the quick attachment of a pressure gauge to check the pressure in the cylinder. And if necessary, the pressure can be safely vented into a drum using a test-gauge hose.
As you can see, this procedure is somewhat involved and so the decision to fill a cylinder with oil is something you have to weigh up based on the value of the cylinder and how long you expect it to be in storage.
Oh, and the moral to the above story is: if you get involved in installing hydraulic components, when it comes time to remove blanking plates or plugs – always assume there’s a possibility the component contains oil under pressure. And take the necessary precautions.
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