# Avoiding Problems with Hydraulic Cylinders

As hydraulic system working pressures increase, individual component design details become more important.

The term “high pressure” in hydraulics is quite subjective. Among the strictest definitions is any system with operating pressure exceeding 10,000 psi. Perhaps a more general interpretation is 6,000 psi, which will be assumed for this discussion. For this article, high system pressure will be defined as 6,000 psi (410 bar). This is the pressure setting on some newer excavator functions. A system that operates at 6,000 psi will see brief pressure spikes that are two to three times the 6,000-psi working pressure. This means the components need to accommodate 12,000 to 18,000 psi, 827 to 1,241 bar pressure spikes, for a reasonable service life of 8,000 hours (about four years), and should not fail in less than 2,000 hours, about one year of service.

Because most of my extensive learning experience is from working with logging and construction equipment attachments, the examples I have set forth are from those industries. I chose not to blame the customer for using the cylinders differently than I specified, and I accepted full warranty responsibility for any poor design choices. Several cylinders were repaired under warranty several times, and a few cylinders were scrapped and completely new designs implemented to meet the self-imposed one-year or 2,000-hour warranty.

The concepts and areas of focus being presented here can be applied to most hydraulic cylinder applications.

Cylinder Design Considerations

What is actually being done with the cylinder and how is the cylinder being loaded? When a machine or attachment is being designed, the designer has an idea of what the unit should do and how it should be done. As soon as you hand the machine to a different operator, the application, duty cycle, and operating parameters have changed. In some cases the change is enough that things start to fail. Frequently, the cylinder will be subjected to a work-induced load that is much greater than what is possible by direct control.

• Will the cylinder need load-holding valves? Both counterbalance valves and pilot-operated check valves are used to hold a cylinder in position, and both valves can produce additional loads on the cylinder components when in operation. Some load-holding valve settings are increased by back pressure on the pilot port.
• What is the cylinder stroke? A short-stroke cylinder does not typically need to be concerned with buckling load design. A long-stroke cylinder will require running a buckling load calculation. Use standard NFPA standard T3.6.37 R1-2010 for buckling calculations.
• How fast is the cylinder going to move? The speed the cylinder will move, or the flow going into the cylinder, must be considered to determine port and plumbing size. This speed is also important when designing cushions if they are required.
• What are the appropriate bore and rod diameters? Choosing the correct bore and rod size can be the single most difficult aspect of the cylinder to determine because of the forces required when extending and retracting, the stroke required, and the pin centers’ distance needed to be met. If the cylinder needs cushioning, this could change the bore or rod size requirements. Sometimes a larger bore and rod will be required to meet all the design needs. One side of the piston may need to run at a lower pressure to meet the force requirements as well as structural requirements, such as the piston-to-rod interface, the piston-fastening needs, and buckling of the rod.
• What are the side loads that the cylinder will see over the service life of the device? It may be difficult to determine the load, or load vectors, for a highly accurate model, but it is important to determine this data if you expect to create a robust design. Some designs will require enough mounting flexibility that either adequate side clearance, self-alignment capabilities, or both need to be incorporated into the anchor and pin-joint design. If possible do a Finite Element Analysis (FEA) of the structure to determine its deflection.
• Are the pin joints adequate? Pin-joint design needs to be applied to both the cylinder ends and the pin anchors being actuated by the cylinder. For boom pivot joints you will need to look at not only the loads created by the cylinder, but also by other loads of the system such as torsion loading. For cylinder pins, you need to design for the maximum load the cylinder will see, which may be the work-induced loads. We use 6,000 psi projected area loading on the bushing, or bushings. We recommend using a bushing length that is about twice as long as the pin diameter if possible. The fixed pin holes’ width should be 0.75 to 1.25 times the pin diameter. Longer, smaller-diameter pins and bushings tend to bend more than shorter, larger-diameter pins and will make the manufacturing and maintenance more expensive over the life of the cylinder.
• Does the cylinder need a deceleration cushion on one, or both, ends of travel? For many applications, slowing the cylinder at the end of stroke is not required. If the application has high inertia loads and no speed-control feedback, building deceleration cushions into the cylinder can be an inexpensive solution. Designing an adequate cushion will require a good understanding of the application and loads to be controlled. To cushion a cylinder, there has to be a way to meter the oil going out of the cylinder. There are several methods of metering the oil flow and controlling cushion pressure. We typically use the piston to close the port to increase the area decelerating the load. The pressure on the side of the piston opposite the side being cushioned is additive to the mass being decelerated and must be considered during design.

1. This sectional drawing shows main components of a high-pressure hydraulic cylinder.

The Rod-to-Piston Interface

The rod-to-piston interface is one area of high-pressure cylinder design that seems to be the most difficult to manage well. A cylinder’s bore and rod diameter often are chosen from a prior lower-pressure design. This can the make interface prone to premature failure if the fastener is too small or the interface area is too small.

Figure 2 shows pistons from a high-pressure hydraulic cylinder with a 5-in. bore and 3-in.-long rod from a log grapple. This cylinder has a short stroke and does not have extreme side load, but it does have a pilot-operated check valve hard-plumbed to the barrel. This prevents the load from dropping if a hose to the cylinder fails. Operators sometimes operate the grapple so that the cylinder sees high extension forces acting against the threads. They also may use the grapple to pull stumps, so the cap end of the cylinder sees pressures up to about 20,000 psi before a catastrophic failure occurs.

2. These four images show three different rod-to-piston designs of a high-pressure hydraulic cylinder (5-in. bore, 3-in. long rod) used in a log grapple. Using a larger nut and tapered rod-to-piston interface prevented failure from high pressure spikes.

Figure 2A shows the 5,000-psi design I developed in 1985. It uses a 1½-NF nut to hold the piston in place. As shown in Fig. 2B, the rod-to-piston interface area was cold-formed. When the rod and piston bearing area were loaded above the yield point of the steel, material movement occurred, which allowed the piston to come loose and cause further damage. The 5,000-psi piston design did not have a large enough thread, so the rod broke at the end of the threads. Also, the rod bearing area was too small, allowing the piston to cold-form over the rod.

Figure 2C shows a competitive design developed in the early 2000s with a threaded-on piston. The internal thread area of the piston in 2C has an even smaller rod-bearing area. There must be clearance in the thread for assembly. This clearance, and the clamping force of tightening the piston, will allow the piston to cold-form over the rod even sooner than other designs. The deflection in the threads causes the piston to loosen, leading to breakage in the rod as shown.

Figure 2D shows my 6,000-psi design developed in the late 1990s with a 1¾-NF nut. In addition to a larger fastener, it uses a tapered interface for a larger bearing area. We have upgraded the piston to 4150 HT on the 6,000-psi design for additional protection from pressure spikes. Some users specify all three designs and have never had a failure with the 6,000-psi rod-to-piston interface. The limitations of the current 6,000-psi design are at about 18,000 psi, which is when the barrel will yield or the pilot-operated check valve will fail anyway.

Considerations for Cylinder Components

With high stress on the rod-to-piston interface, the materials used for the rod and piston should have a similar hardness, a high impact capability, and may require a high Charpy impact value at low temperatures. This is required if the cylinder is being used at low temperatures. The barrel should be made from a material with a high yield strength, good weldability, and adequate impact capabilities. Do not use materials that contain free-machining elements in parts that will be welded. These free-machining elements include, but are not limited to sulfur, lead, manganese, calcium, selenium, tellurium, and bismuth. The wall thickness choice should be thick enough to keep barrel swell low to reduce seal failure. This graph is barrel swell in inches and psi for a 4-in. ID barrel with a 5-in. OD. Barrel swell can reduce seal life by increasing the gap between the piston and barrel. This increases the diameter of the element, or elements, contacting the barrel causing an increased seal groove volume.

Head gland attachment—There are many different methods to attach the head gland to the barrel. The most reliable design will handle the maximum load the cylinder sees when extended by pressure in the blind end, plus the stopping load. If the cylinder has an extension hard stop, then the stopping load is 0. If the cylinder has to stop the movement of the structure it is driving, this must be calculated and added to the extend force. In the case of excavator boom, and bucket, cylinders this can be a large force.

Welding procedures—Welding failures are common in high-pressure hydraulic cylinders for many different reasons but I will only address the more common ones and provide some repair information.

• The weld chamfer should allow for correct wire stickout and shield gas control, but it should be no larger than necessary to prevent increased welding time and the resultant weld-generated distortion.
• Early in the development of high-pressure cylinders, one of the weld failures was from using the incorrect yield-strength filler materials. Most tubing for high-pressure cylinders will need a yield strength greater than 70,000 psi. This will require a filler material that is 80,000 psi yield or higher. E80 wire is typically a good choice.
• To ensure the weld does not have too much stress from welding, I recommend using a multi-pass pulsed-spray weld procedure. This will also reduce distortion of the parts. There should be a small contact area where the barrel tube and the blind end connect, so, if the barrel needs to be replaced, the barrel pilot and locating shoulder machined on the blind end will still exist after weld removal. This does require the weld to be larger in diameter than the barrel, which may not work in some applications.
• Weld repair in hydraulic cylinders can be a challenge because of oil saturation. Low hydrogen welding materials will not seal oil-contaminated metals. The first pass should be done with a 6010 or 6011 stick rod using a whipping motion. Cover passes using higher-strength filler materials are required for structural integrity.

Challenges of High-Pressure Sealing

The materials chosen for sealing and guiding the rod and piston can be the single most difficult portion of high-pressure cylinder design. Cylinders for off-highway equipment can undergo some very high-pressure spikes during operation that will reduce seal life. Proper use of step-cut or lap-cut wear rings, zero-split clearance bushings, or wear rings can be a better solution than buffer seals. Buffer seals can trap high-pressure oil against the rod seal, causing premature rod-seal failure. A backup ring behind the rod seal to reduce extrusion is a good option for all cylinders.

Too much space for movement in the seal groove can reduce seal life. Filled seal materials tend to break down under high pressure, scratch components, and contaminate the hydraulic oil. Glass-filled material tends to be brittle where the glass bonds to the parent material and will deposit glass into the hydraulic system as they fail.

The PV value (combined pressure and velocity) of the sealing material is an important consideration for high-pressure sealing. When the piston and rod assembly are moving, the seal surface becomes a bearing that has no running clearance and must run almost dry to stop fluid bypass. Every sealing material will have a different PV value relative to different contacting materials.

Compression set—Most elastomers have increased compression set when cooled below freezing and raised to operating temperature daily. Many O-ring grooves do not have dimensions to allow for the cyclic temperature-compression set. If you see O-rings damp enough to collect dust in static applications, this leak is probably from cyclic temperature-compression set.

Size change—Elastomers have a higher coefficient of thermal expansion than steel. This means the seal will shrink more when cold than steel does, and expand more when hot. When looking at seal groove dimensions, it is important to consider the size of the elastomer will be at during the lowest operating temperature as well as at the highest temperature. Too little volume or too much volume can reduce seal life.

Seal failures—Extrusion, nibbling, and breakage are common failures. Figure 4A is an example of extrusion—in less than 300 hours of operation—from excessive pressure applied to a urethane seal. Figure 4B shows an example of seal extrusion from honing a cylinder to clean it up during repair without increasing the diameter of the piston. Figures 4C and 4D are examples of nibbling. Nibbling can be caused by barrel swell, axial movement of the seal in the groove, or incorrect groove dimensions. Figure 4E is an example of a filled material breaking and the sides of the expander show evidence of axial movement in the groove.

4. Examples of premature seal failure. View A shows extrusion damage to a urethane seal after only 300 hr of operation from excessive pressure. View B shows extrusion damage from inadequate honing of the cylinder bore during repair. Views C and D show nibbling, caused by barrel swell, axial movement of the seal in the groove, or incorrect groove dimensions. View E shows an example of a broken filled material breaking with sides of the expander showing evidence of axial movement in the groove.

Assembly Considerations

Lubrication—When assembling hydraulic components, use oils or greases that are compatible with the hydraulic fluid used in the machine. Do not use materials that contain mineral-based additives or thickeners. We use a Polyurea-based grease to meet these requirements.

When faced with the task of redesigning a cylinder to handle higher working pressures, use FEA or other simulation tools to reverse-engineer failed components as well as qualify new designs. Remember that just because component design, material choice, seal choice, or assembly method worked at 5,000 psi does not guarantee it will perform well at 6,000 psi.