Guidelines to Avoid Those Hydraulic-Cylinder Headaches

Guidelines to Avoid Those Hydraulic-Cylinder Headaches

Increases in hydraulic-system working pressures make individual component design details even more important.

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The term “high pressure” is quite subjective when it comes to hydraulics. 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. In this article, high system pressure will be defined as 6,000 psi (410 bar), which 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 2000 hours—about one year of service.

Because most of my extensive learning experiences are 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 were implemented to meet the self-imposed one-year or 2,000-hr warranty.

The concepts and areas of focus being presented here can be applied to most hydraulic cylinder applications. Components of a typical cylinder are illustrated in Figure 1.

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

Cylinder Design Considerations 

What is actually being done with the cylinder and how is the cylinder being loaded? When designing a machine or attachment, the designer has an idea of what the unit should do and how it should be done. As soon as you hand the machine over 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, a cylinder will be subjected to a work-induced load that is much greater than what is possible via 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 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 diameter? Choosing the correct bore and rod size can be the single most difficult aspect of the cylinder to determine. That’s because of the forces that are needed when extending and retracting, the stroke required, and the pin center’s distance having 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 of the design needs. One side of the piston may have to run at a lower pressure to meet the force 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 at the cylinder pin joints.
  • 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 other system loads 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 whenever possible. The fixed pin-hole width should be 0.75 to 1.25 times the pin diameter. Long, small-diameter pins and bushings tend to bend more than short, large-diameter ones, and thus 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 to the side being cushioned is additive to the mass being decelerated and must be considered during design.

The Rod-to-Piston Interface

Perhaps the most difficult-to-manage area of high-pressure cylinder design is the rod-to-piston interface. 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 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 include 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 that I developed in 1985. It uses a 1½-NF nut to hold the piston in place. As shown in Figure 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 caused the piston to come loose and create 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 Fig. 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 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 when using the cylinder 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.

3. Barrel swell is measured in a 4-in. bore steel tube with ½-in. thick wall. Barrel swell can degrade seal life.

The wall-thickness choice should be thick enough to keep barrel swell low enough to prevent seal failure. Figure 3 plots barrel swell 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—Many different methods can be used 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 zero. 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, these forces can be substantial.

I do not recommend screwing the head into the barrel. As pressure rises, it swells the barrel and increases thread clearance, which allows movement that will cause thread wear and reduce the service life of the cylinder. Barrel, or gland, nuts work if properly sized for wall thickness. It also helps to use Acme or square threads. Buttress threads have a sharp thin cross-section, so they may be weaker than other thread forms. Using a ring of cap screws is a common method that works well if properly designed.

The thickness of the flange clamped by the fasteners must be thick enough to not deflect during loading. The 5K cylinder design had flanges that were marginally thick enough. As a result, in an application where high forces were at the end of the cylinder travel, the deflection caused the head and cap screws to fail sooner than in other applications. Using a flange that is twice as thick as the bolt diameter will usually be adequate for most applications. 

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 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 cap end connect. Therefore, if the barrel needs to be replaced, the barrel pilot and locating shoulder machined on the cap 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 may 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 is often brittle where the glass bonds to the parent material, and will deposit glass into the hydraulic system as the bonds fail.

4. Examples of premature seal failure: extrusion damage to a urethane seal after only 300 hours of operation from excessive pressure (a); extrusion damage from inadequate honing of the cylinder bore during repair (b); and nibbling caused by barrel swell, axial movement of the seal in the groove, or incorrect groove dimensions (c and d). Also shown is an example of filled material breaking, with sides of the expander showing evidence of axial movement in the groove (e).

The PV value (pressure × 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 than steel when it’s cold, and expand more when hot. Regarding seal groove dimensions, it’s important to consider what the size of the elastomer will be 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 shows an example of extrusion—in less than 300 hours of operation—from excessive pressure applied to a urethane seal. Figure 4b illustrates an example of seal extrusion from honing a cylinder to clean it up during repair without increasing the piston’s diameter. 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; the sides of the expander show evidence of axial movement in the groove.


When assembling hydraulic components, use oils or greases that are compatible with the hydraulic fluid used in the machine. Do not opt for 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.

Ed Danzer is president of 6K Products, Tenino, Wash. For more information, call (360) 264-2141, or visit the 6K Products site.

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