I presented a workshop on minimizing hydraulic equipment operating costs at a local University recently. During that presentation, I shared with the attendees what I consider to be THE most important proactive maintenance routine for hydraulic equipment.
No, it’s not contamination control. These days, best-practice contamination control is an accepted precondition for reliability. And given contemporary advances in technology for excluding and removing contaminants, it could be said that failure to control contamination is a failure of machine design – rather than a failure of maintenance.
The maintenance routine that I believe ranks above contamination control in order of importance these days – largely due to its neglect, is: maintaining fluid temperature and viscosity within optimum limits. This involves:
- Defining an appropriate fluid operating temperature and viscosity range for the ambient temperature conditions in which the hydraulic machine operates;
- Selecting a hydraulic fluid with a suitable viscosity grade and additive package; and
- Ensuring that both fluid temperature and viscosity are maintained within the limits defined.
In order to determine the correct fluid viscosity grade for a particular application, it is necessary to consider:
- starting viscosity at minimum ambient temperature;
- maximum expected operating temperature, which is influenced by system efficiency, installed cooling capacity and maximum ambient temperature; and
- permissible and optimum viscosity range for individual components in a system.
For example, consider an application where the minimum ambient temperature is 15°C, maximum operating temperature is 75°C, the optimum viscosity range for the system’s components is between 36 and 16 centistokes and the permissible, intermittent viscosity range is between 1000 and 10 centistokes.
From the temperature/viscosity diagram , it can be seen that to maintain viscosity above the minimum, optimum value of 16 centistokes at 75°C, an ISO VG68 fluid is required. At a starting temperature of 15°C, the viscosity of VG68 fluid is 300 centistokes, which is within the maximum permissible limit of 1000 centistokes at start up.
Having established the correct fluid viscosity grade, the next step is to define the fluid temperature equivalents of the optimum and permissible viscosity values for the system’s components.
By referring back to the temperature/viscosity curve for VG68 fluid shown in it can be seen that the optimum viscosity range of between 36 and 16 centistokes will be achieved with a fluid temperature range of between 55°C and 78°C. The minimum viscosity for optimum bearing life of 25 centistokes will be achieved at a temperature of 65°C. The permissible, intermittent viscosity limits of 1000 and 10 centistokes equate to fluid temperatures of 2°C and 95°C, respectively (see exhibit 2).
| Viscosity Value |
cSt |
Temperature (VG68) |
| Min. Permissible |
10 |
95ºC |
| Min. Optimum |
16 |
78ºC |
| Opt. Bearing Life |
25 |
65ºC |
| Max. Optimum |
36 |
55ºC |
| Max. Permissible |
1000 |
2ºC |
Exhibit 2. Correlation of typical operating viscosity values for a piston pump with fluid temperature, based on fluid viscosity grade.
Going back to our example, this means that with an ISO VG68 fluid with a viscosity index similar to that shown in exhibit 1 in the system, the optimum operating temperature is 65°C. Maximum operating efficiency will be achieved by maintaining fluid temperature in the range of 55°C to 78°C. And if cold start conditions at or below 2°C are expected, it will be necessary to pre-heat the fluid to avoid damage to system components. Intermittent fluid temperature in the hottest part of the system, which is usually the pump case, must not exceed 95°C.
Having defined the parameters shown in exhibit 2 for a specific piece of hydraulic equipment, damage caused by high or low fluid temperature (low or high fluid viscosity) can be prevented, and recurring PM tasks in respect of this routine can be virtually eliminated, by installing fluid temperature monitoring instrumentation with alarms and shutdowns.
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One of our readers wrote to me recently about the following problem:
” We have a hydraulic power unit that runs a calendar roll on a paper machine. The calendar roll has multiple zones inside it to vary the pressure on the paper to maintain uniform thickness. Moog servo-valves control the adjustments to the different zones. The problem we have is high vibration in the pumps, lines and tank. Lines have broken and we had to disconnect one bank of filters from the tank to keep it from breaking off. This has been going on for some time. Do you have any ideas?”
As always, there are a number of possibilities and issues to consider. Some background reading on vibration and noise in hydraulic systems is available in this tutorial. One explanation that jumps to the top of the list in this application is water hammer.
Coining a phrase
Water hammer is the term used to describe the effect that occurs when the velocity of the fluid moving through a pipe suddenly changes. Sudden change in fluid velocity causes a pressure wave to propagate within the pipe. Under certain conditions, this pressure wave can create a banging noise, similar to that you would expect to hear when beating a pipe with a hammer. Hence the phrase. Not surprisingly, common symptoms of this problem are high noise levels, vibration and broken pipes.
Hitting the wall
When a moving column of fluid hits a solid boundary – when a directional control valve closes suddenly for example, its velocity drops to zero and the fluid column deforms, within the rigid cross-sectional area of the pipe, to absorb the (kinetic) energy associated with its motion – similar to a car hitting a concrete wall. However unlike a car, the fluid is almost incompressible so the deformation is small and a store of energy accumulates in the fluid – similar to compression of a spring. The magnitude of the pressure rise that results from the subsequent release of this stored energy can be expressed mathematically as follows:
Pr = P + u p c
where P is initial pressure, u and p are initial fluid velocity and density respectively and c is the speed of sound through the fluid.
In our reader’s application, uniform paper thickness is dependent on the constant adjustment of the calendar roll zones by the servo-valves. Under certain conditions, rapid switching of these valves could result in something that resembles peening a pipeline with a thousand hammers.
Speed kills
Accumulators and other damping devices are sometimes installed in an effort to deal with this problem. However, the significance of the pressure rise equation shown above is that fluid velocity is the only variable that can be altered to address the root cause. Put simply, reducing the velocity of the fluid column that hits the solid boundary, reduces the magnitude of the subsequent pressure rise. Returning to the traffic crash analogy – the slower the car is travelling when it hits the wall, the less damage is caused.
In hydraulics, the easiest way to do this – on paper at least, is to increase the diameter of the pipe. This reduces fluid velocity for a given flow rate. The other alternative is to control deceleration of the fluid column by choking valve switching time to the point where the pump’s pressure compensator and/or system relief valve reacts fast enough to reduce flow rate through the pipe and therefore velocity of the fluid.
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“In one of our applications we are using NG 40 cartridge valves (sleeve, poppet and logic cover). With the valve closed and the inlet port pressurized to 315 bar, we are seeing a leakage from the outlet port in the order of half a liter per minute. Is this level of leakage acceptable?”
The first thing to consider is whether the logic element has been configured for leakless operation. If the direction of flow is from A to B this is referred to as base flow. If flow is from B to A this is know as annulus flow (see figure 1). A logic element can be configured for flow in either or both directions.
Figure 1. Logic element base and annulus flow configurations (Industrial Hydraulic Control).
To establish whether a logic element is configured for zero leakage, it is necessary to consider the direction of pressure drop across the poppet when it is closed. Consider a logic element configured for check valve function in both base and annulus flow directions. When configured as a check valve for base flow (A to B) see figure 2, the direction of pressure drop across the poppet when it is closed is from B to A. In this configuration the logic element is leakless.
Figure 2. Logic element; check valve function; base flow (Industrial Hydraulic Control).
When configured as a check valve for annulus flow (B to A) see figure 3, the direction of pressure drop across the poppet when it is closed is from A to B. In this configuration the clearance between the poppet and its sleeve results in leakage from A to B. The magnitude of this leakage may increase over time as a result of wear between the poppet and sleeve.
Figure 3. Logic element; check valve function; annulus flow (Industrial Hydraulic Control).
Assuming our reader’s logic elements have been configured for leakless operation, other possible explanations for the leakage include:
- damage to the poppet and/or its seat
- degradation or damage to the elastomeric seal at the base of the sleeve
- incorrect machining tolerance in the logic housing
In this, and all other troubleshooting situations, the first place to look for guidance is the machine’s circuit diagram and your reference library. From there on, it is a logically process of elimination.
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“Are there any specific considerations I should be aware of prior to replacing pumps and motors on an excavator with aftermarket units? If I order a pump from a supplier other than the OEM, how can I be certain that the pump I’m getting is precisely the same as the one I’m replacing?”
Aftermarket units
A machine owner will usually get a better price on a replacement hydraulic component if they are able to buy it ‘aftermarket’, from a fluid power distributor. However, this is not always as easy as it might seem. Original equipment manufacturers (OEM’s) usually do all they can to control the distribution and sale of spare parts for the machines they build. This is particularly true in the case of mass-produced hydraulic machines. The OEM knows that if you can identify the make and model of the hydraulic pump fitted to your excavator, you will be able to shop around for the best price on a replacement pump and as a result, there is a good chance they will lose the business.
In an effort to prevent this from happening, OEM’s usually identify hydraulic pumps, motors, cylinders and other components fitted to their machines with their own part numbers. In most cases these numbers are meaningless to anyone else. Therefore the first thing you need to do is to identify the component’s manufacturer and model code.
If you know another operator or company that owns the same machine you need the replacement component for, you could ask them if they have replaced this component with an aftermarket unit. If so, they will probably be happy to give you the model code from the aftermarket component’s identification tag.
If you have a good idea of what you’re looking for, you can identify the component yourself, using information that is available on the Internet. This involves first measuring and identifying the component’s physical attributes such as shaft type, mounting flange, ports, displacement and control. You then need to match these variables to the dimensional and technical data contained in the manufacturer’s product catalog. The catalog will show you how to compile a model code (sometimes called an order code) that corresponds to the component you want to replace. Sounds easy, but if you don’t know who the manufacturer is or what the component does, this can be a difficult task. In this case, it is easier to let a fluid power distributor do the work and earn their profit margin in the process.
OEM specials
It is not always possible to source aftermarket hydraulic components for OEM equipment. OEM’s sometimes use components that are manufactured with a unique difference, known as ‘OEM specials’. This means that even if you do identify the make and model of the component, the only way you can buy an identical unit is through the machine dealer.
The difference may be something obvious, such as the shaft type or the orientation of the ports – or not so obvious, such as the control set-up in a variable displacement pump or motor. I can think of at least one example where the main pump for a particular excavator was, to the casual observer, a standard unit. However, if a pump from the component manufacturer’s standard product line was installed on the machine, it would cause the engine to stall. Reason being, the standard pump was fitted with a hydraulic displacement control with a control range (minimum to maximum displacement) of 10 bar, whereas the OEM pump was fitted with a special control range of 35 bar. With a standard pump installed, the excavator’s electronic power management system could not effectively control the pump’s displacement and therefore power draw.
Obvious or not, these differences are usually enough to make it either impossible or uneconomic to adapt a unit from the component manufacturer’s standard product line. Some large OEM’s also manufacture their own hydraulic components. As with OEM specials, these components are a captive market for the machine dealer.
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In the case of oil leaks, the cost areas that need to be considered include:
- Make-up fluid;
- Clean-up;
- Disposal;
- Contaminant ingress; and
- Safety.
But what about hydraulics’ fluid power cousin – pneumatics? One of the advantages that pneumatics has over hydraulics is its cleanness. Air leaks are much easier to ignore than oil leaks because they don’t draw attention to themselves in the same way. You don’t need to worry yourself with clean-up and disposal costs. Contaminant ingression is possible, but is generally not a major concern. And unless the leak is significant, safety is not usually a big issue either. So that leaves make-up fluid (air).
Make-up air
While air is free – clean, dry compressed air is not. In considering the cost of make-up air for a pneumatics system the following need to be considered:
- Depreciation (wear and tear) of the compressor;
- Conditioning costs – filtration, drying and lubrication; and
- Energy cost of compression.
The ideal leakage rate is of course zero, but when calculating the free air delivery (FAD) required by a pneumatic system a rule of thumb is to allow for leakage of 10% of total flow rate. Consider a 10 cubic meter/minute system leaking one cubic meter/minute. The power required to compress one cubic meter (35.3 cubic feet) of air per minute to a pressure of 6 bar (90 PSI) is approximately 5.2 kW. At an electricity cost of $0.10/kWh this leakage is costing over 50 cents per hour in electricity consumption alone. In a 24/7/365 operation that amounts to $4500 per year!
Quantifying losses
While a leakage rate of 10% of flow rate may sound high and would be unsustainable in a hydraulic system, air leakage rates as high as 25% are not unheard of – even in apparently well maintained pneumatic systems. The actual leakage rate of a system can be calculated using the following formula:
QL = QC* t / (T+t)
Where:
QL = System leakage rate (cubic meters/minute)
QC = Compressor FAD (cubic meters/minute)
T = Leakage time – time between compressor cut-out and cut-in (minutes)
t = Charging time – time between compressor cut-in and cut-out (minutes)
Note: this formula assumes that all system consumption is suspended while the leakage test is conducted.
Conclusion
As demonstrated by the above example, the annual cost of air leaks in pneumatic systems can be significant – in power consumption alone. Conduct regular leakage tests on your pneumatic systems and take necessary action to locate and rectify leaks as required.
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Load sensing is a term used to describe a type of variable pump control used in open circuits. It is so called because the load-induced pressure downstream of an orifice is sensed and pump flow is adjusted to maintain a constant pressure drop (and therefore flow) across the orifice. The ‘orifice’ is usually a directional control valve with proportional flow characteristics, but a needle valve or even a fixed orifice can be employed, depending on the application.
In hydraulic systems that are subject to wide fluctuations in flow and pressure, load-sensing circuits can save substantial amounts of input power. This is illustrated in Exhibit 1. In systems where all available flow (Q) is continuously converted to useful work, the amount of input power lost to heat is limited to inherent inefficiencies. In systems fitted with fixed displacement pumps where 100 percent of available flow is only required intermittently, the flow not required passes over the system relief valve and is converted to heat. This situation is compounded if the load-induced pressure (p) is less than the set relief pressure – resulting in additional power loss due to pressure drop across the metering orifice (control valve).
Exhibit 1. Flow-pressure-power diagrams for fixed, variable and load sensing controlled hydraulic pumps (Peter Rohner).
A similar situation occurs in systems fitted with pressure controlled (pressure compensated) variable pumps, when only a portion of available flow is required at less than maximum system pressure. Because this type of control regulates pump flow at the maximum pressure setting, power is lost to heat due to the potentially large pressure drop across the metering orifice.
A load sensing controlled variable pump largely eliminates these inefficiencies. The power lost to heat is limited to the relatively small pressure drop across the metering orifice, which is held constant across the system’s operating pressure range (see bottom of Exhibit 1).
A load sensing circuit typically comprises a variable displacement pump, usually axial-piston design, fitted with a load sensing controller, and a directional control valve with an integral load-signal gallery (Exhibit 2). The load-signal gallery (LS, shown in red) is connected to the load-signal port (X) on the pump controller. The load-signal gallery in the directional control valve connects the A and B ports of each of the control valve sections through a series of shuttle valves. This ensures that the actuator with the highest load pressure is sensed and fed back to the pump control.
Exhibit 2. Typical load sensing circuit. Enlarge
To understand how the load-sensing pump and directional control valve function together in operation, consider a winch being driven through a manually actuated valve. The operator summons the winch by moving the spool in the directional valve 20 percent of its stroke. The winch drum turns at five rpm. For clarity, imagine that the directional valve is now a fixed orifice. Flow across an orifice decreases as the pressure drop across it decreases. As load on the winch increases, the load-induced pressure downstream of the orifice (directional valve) increases. This decreases the pressure drop across the orifice, which means flow across the orifice decreases and the winch slows down.
In a load sensing circuit the load-induced pressure downstream of the orifice (directional valve) is fed back to the pump control via the load-signal gallery in the directional control valve. The load-sensing controller responds to the increase in load pressure by increasing pump displacement (flow) slightly so that pressure upstream of the orifice increases by a corresponding amount. This keeps the pressure drop across the orifice (directional valve) constant, which keeps flow constant and in this case, winch speed constant. The value of the pressure drop or delta p maintained across the orifice (directional valve) is typically 10 to 30 Bar (145 to 435 PSI). When all spools are in the center or neutral position the load-signal port is vented to tank and the pump maintains ‘standby’ pressure equal to or slightly higher than the load sensing control’s delta p setting.
Because the variable pump only produces the flow demanded by the actuators, load-sensing control is energy efficient (fewer losses to heat) and as demonstrated in the above example, improves actuator control. Load-sensing control also provides constant flow independent of pump shaft speed variations. If pump drive speed decreases, the load-sensing controller will increase displacement (flow) to maintain the set delta p across the directional control valve (orifice), until displacement is at maximum.
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Troubleshooting hydraulic systems can be a complex exercise. It involves a lot of science and sometimes, a bit of art. Incorrect diagnosis prolongs downtime and can result in the unnecessary repair or replacement of serviceable components. To avoid these costly mistakes, the correct equipment and a logical approach are required.
Assess the problem and eliminate the obvious
Before you incur the expense of hiring a technician, assess the problem and eliminate all of the obvious, possible causes. I have lost count of the number of times that I’ve been called to a problem and found that the cause was something quite simple. A wire broken off a solenoid valve, a pin fallen out of a mechanical linkage, an isolation valve that had vibrated closed, a blocked heat exchanger… and so the list goes on.
Your oversight won’t bother the technician, because his hourly rate is the same, regardless of how easy or difficult the problem is to find. But you may be annoyed with yourself for not checking something so obvious, knowing that you could have easily saved the cost of the call-out.
Quality is more important than quantity
Paying for a technician’s time when it is not required is certainly not desirable. But it is nowhere near as costly as paying for the unnecessary repair or replacement of serviceable components, as a result of incorrect diagnosis of a problem. Incorrect diagnosis in a troubleshooting situation is usually the result of the technician’s incompetence, insufficient investigation of the problem or a combination of both.
Unfortunately it is not possible to determine a technician’s competency from the badge on his shirt or his charge-out rate. While charge-out rates may be a factor in deciding whose technician you hire, from an overall cost perspective it is far more important to evaluate the technician and his diagnosis, so that you don’t end up paying for his mistakes.
Let me illustrate how this can happen with an example. Several years ago, I was asked for a second opinion on the condition of a set of pumps operating a processing plant. The customer had called in a technician to check the performance of these pumps and was alarmed when the technician advised that all four pumps were in need of repair.
The pumps in question were variable-displacement units fitted with constant power control. The power required to drive a hydraulic pump is a product of flow and pressure. A constant power or power limiting control operates by reducing the displacement, and therefore flow, from the pump as pressure increases, so that the power rating of the prime mover is not exceeded.
Pump performance is checked using a flow-tester to load the pump and measure its flow rate. As resistance to flow is increased, pressure increases and the flow available from the pump to do useful work decreases because of internal leakage. The difference in the measured flow rate between no load and full load determines the volume of internal leakage and therefore pump performance.
I tested all four pumps, recording flow against pressure from no load through to maximum working pressure. In my report I explained to the customer that the tests revealed that pump flow did decrease significantly as pressure increased, but that this is a normal characteristic of a pump fitted with constant power control. I further advised that apart from the constant power control requiring adjustment on two of the pumps, the performance of all four pumps was acceptable.
The first technician’s assessment can only be explained by fraud or incompetence. I suspect it was the latter, with the technician failing to either establish or understand that the pumps he was testing were fitted with constant power control. This ignorance led to an incorrect interpretation of the test results. Whatever the explanation, the customer could have paid thousands of dollars for unnecessary repairs, if they had not sought a second opinion.
Conclusion
When you have a problem with your hydraulic equipment, carry out an informed assessment of the problem and eliminate the obvious before you call for a technician. And if you do need to hire a technician, be sure to evaluate the technician and his diagnosis so you don’t end up paying for his on-the-job-training or worse, his mistakes!
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Continuous monitoring of the filter elements in a hydraulic system can provide valuable clues to the performance of the filter and the condition of the system. Before I discuss this, let’s consider some of the advantages and disadvantages of common filter locations.
Pressure filtration
Locating filtering media in the pressure line provides maximum protection for components located immediately downstream. Filtration rates of two microns or less are possible, due to the positive pressure (in comparison to an intake line filter) available to force fluid through the media (in comparison to an intake line filter). Filter efficiency may be reduced by the presence of high flow velocities, and pressure and flow transients, which disturb trapped particles. The major disadvantage of pressure filtration is economic. Because the housings and elements (high-collapse type) must be designed to withstand system operating pressure, pressure filtration has the highest initial and ongoing cost.
Return filtration
The rationale for locating filtering media in the return line is this: if the reservoir and the fluid it contains start out clean and all air entering the reservoir and returning fluid is adequately filtered, then fluid cleanliness will be maintained. The other advantage of the return line as a filter location is that sufficient pressure is available to force fluid through fine media – typically 10 microns, but pressure is not high enough to complicate filter or housing design. This combined with relatively low flow velocity, means that a high degree of filtering efficiency can be achieved at an economical cost. For these reasons, return filtration is a feature of most hydraulic systems.
Off-line filtration
Off-line filtration enables continuous, multi-pass filtration at a controlled flow velocity and pressure drop, which results in high filtering efficiency. Filtration rates of two microns or less are possible, and water absorbent filters and heat exchangers can be included in the circuit for total fluid conditioning. Off-line filtration has a high initial cost, although this can often be justified on a life-of-machine cost basis.
Filter condition monitoring
Warning of filter-bypass is typically afforded by visual or electric clogging-indicators. These devices indicate when pressure drop (delta P) across the element is approaching the opening pressure of the bypass valve (where fitted). In the case of a return filter for example, if the bypass valve opens at a delta P of 3 Bar, the clogging indicator will typically switch at 2 Bar.
Advanced filter condition monitoring
Replacing standard clogging-indicators with differential pressure gauges or transducers enables continuous, condition monitoring of the filter element. This permits trending of fluid cleanliness against filter element pressure-drop, which may be used to optimize oil sample and filter change intervals. For example, the optimal change for a return filter in a particular system could be higher or lower than the clogging indicator switching pressure of 2 Bar.
Continuous monitoring of filter pressure drop can also provide early warning of component failures and element rupture. For example, if the delta P across a pressure filter suddenly increased from 1 to 3 Bar (all other things equal), this could be an indication of an imminent failure of a component upstream. Similarly, a sudden decrease in delta P could indicate a rupture in the element – something that a standard clogging indicator will not warn of.
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Hydraulic components are unique in that it is often possible to offset or balance hydrostatic forces to reduce loads on lubricated surfaces. By reducing surface loading, the maintenance of full-film lubrication is improved and therefore boundary lubrication conditions are less likely to occur.
Hydrostatic force is the product of pressure and area. Expressed mathematically: F = P x a. The balancing or offsetting of hydrostatic force is achieved by exposing opposing areas to the same pressure. The double-acting cylinder in Figure 1 illustrates this concept.
Figure 1. Hydrostatically balanced cylinder loading two lubricated surfaces.
The rod-side area of the piston, area B, is 80% of area A. This means that the force exerted on the lubricated surfaces at the end of the cylinder rod is 20% of the force developed by the pressure acting on area A. This is due to the balancing or offsetting force developed by the same pressure acting on area B. Assuming the speed of the rotating surface (C) and fluid viscosity are adequate, full-film lubrication of the sliding surfaces is achieved.
Figure 2. Typical cross-section of an axial design piston.
The same principle applied to a typical axial design piston is illustrated in Figure 2. Area A is exposed to system pressure during outlet (pump) or inlet (motor) and the force developed is transmitted to the lubricated surfaces of the slipper and swash plate. System pressure also acts on area B, the balancing area of the slipper, via the drilling through the center of the piston. Area C is the sliding (lubricated) area of the slipper. While the ratio of these three areas varies, in this particular piston, area B is 50% of area A and area C is 140% of area A. This means that the force transmitted to area C is half that developed by area A and is spread over 1.4 times the area, further reducing the load on the lubricated surfaces.
Figure 3. Loss of hydrostatic balance increases load on the lubricated surfaces.
If the hydrostatic balancing force is lost, that is there is no pressuring acting on area B (Figure 3), the force exerted on the lubricated surfaces at the end of the cylinder rod will be 100% of the force developed by the pressure acting on area A. If full-film lubrication was dependent on the hydrostatic balance of the cylinder, boundary lubrication conditions will eventuate and two body abrasion is likely.
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One of the topics discussed was cylinder rod buckling loads. In response to this article, many or our readers wanted to know how to deal with bent rods in a repair situation.
Checking rod straightness
Rod straightness should always be checked when a hydraulic cylinder is being re-sealed or repaired. This is done by placing the rod on rollers and measuring the run-out with a dial gauge (Figure 1). Position the rod so that the distance between the rollers (L) is as large as possible and measure the run-out at the mid-point between the rollers (L/2).
Figure 1. Checking rod straightness.
Allowable run-out
The rod should be as straight as possible, but a run-out of 0.5 millimeters per linear meter of rod is generally considered acceptable. To calculate maximum, permissible run-out (measured at L/2) use the formula:
Run-out max. (mm) = 0.5xL/1000
Where: L equals distance between rollers in millimeters.
For example, if the distance between the rollers was 1.2 meters, then the maximum, allowable run-out measured at L/2 would be given by 0.5 x 1200 / 1000 = 0.6mm.
Dealing with bent rods
In most cases, bent rods can be straightened in a press. It is sometimes possible to straighten hydraulic cylinder rods without damaging the hard-chrome plating, however if the chrome is damaged, the rod must be either re-chromed or replaced.
If a rod is bent, then it is wise to check actual rod loading against permissible rod loading based on the cylinder’s mounting arrangement and the tensile strength of the rod material. The formulas and procedure for doing this are explained in detail in Industrial Hydraulic Control. If actual road load exceeds permissible load then a new rod should be manufactured from higher tensile material and/or the rod diameter increased to prevent the rod from bending in service.
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