Heavy equipment machinery

What is a grapple?

A grapple is simply a hook or claw that is used to catch and hold something. This generic definition applies to many different devices, the simplest being a ship’s anchor. Other simple examples include a throwing grapple or a multi-spiked hook used for climbing. In engineering or construction it almost exclusively applies to the powered claw that is often seen on large, earth moving machinery.

Grapplers and heavy machines

A grapple is a powered claw that makes use of hydraulic technology by using two or more opposing levers to pinch and then pull or drag materials. Grapples are commonly mounted to tractors and excavators. They are attached to the movable arm where they can extend or retract in addition to moving side to side, which is referred to as a pivot or rotate. Higher end engineering machines may also contain a separate control for rotating the grapple by itself.

Simple grapple machines are made-up of a hydraulic powered fork, also called a "grapple rake," or bucket in addition to an opposing "thumb" or multiple hooks and levels that enclose and grip different materials for lifting and dragging purposes.

The bucket on the loader or other heavy machine, otherwise known as a multi-purpose or demolition bucket, operates a grapple that forces the hinges on the rear side of the bucket to be forced apart or forced together using cylinders operating on hydraulic technology.

Are their different types of grapples?

There are several other grapples that have specific abilities, like lifting, pushing or digging. The major one is called a lifting grapple, and it functions as an attachment to large, heavy objects so it can support heavy material landing equipment.

Lifting grapples can also stand in as a tie down that hold large objects in place. This is done via attaching ropes or chains to a point on the grapple, securing the bulky item.

Thus grapplers serve many functions and are highly versatile, serving as aids to large, earthmoving machinery or secure supports for large, bulky machinery. Since their inception they have provided a powerful, valuable tool that functions as an essential accessory in any large, engineering or construction project that uses heavy, powerful machinery.

http://www.dunkelbros.com/grapple-tools-heavy-machines.html

Rebuilding a hydraulic component involves reworking or replacing all parts necessary to return the component to an "as new" condition in terms of performance and expected service life. In other words, a rebuilt component should perform as well and last as long as a new one. In this column, the terms "rebuild" and "repair" will be used interchangeably with the same meaning.

In many cases, rebuilding a component can result in significant savings when compared to the cost of purchasing a new one. The economics of proceeding with repairs ultimately depends on the cost of the repair relative to the cost of a new component. As a rule, the more expensive a new component is, the more likely that a repair will be cost-effective.

Factors That Influence Component Repair Costs
A component’s rebuild cost is determined by many factors including:

• extent of wear or damage to the component

• facilities and expertise of the repairer

• repair techniques employed

If you are responsible for keeping one or more hydraulic machines running and are serious about minimizing operating costs, then at some point, a hydraulic component will need to be rebuilt by a specialist.

There are generally three options for rebuilding: the machine dealer, the component manufacturer or an independently owned hydraulic shop.

Machine Dealers
With hydraulic component repairs, the capabilities of machine dealers vary. Some have the necessary expertise and equipment to perform most hydraulic rebuilds in-house, while others are completely reliant on outside suppliers to provide this service. If a dealer relies on a third party for repairs to hydraulic components, it could end up costing more. Exchange Programs
If machine downtime is an issue and the dealer offers an exchange or "reman" program, this can be an attractive option. The benefit of an exchange program is the dealer carries inventory of rebuilt components. Therefore, instead of waiting for a component to be repaired, it can be exchanged with a reman unit from the dealer’s stock.

It’s important to understand the exchange price is not necessarily based on the cost of repairing a particular component. It is usually based on the average rebuild cost for that component plus an inventory charge, and may be conditional on certain parts of the component being reusable.

Because the dealer’s exchange price is based on an average rebuild price, if you were to have a component repaired it could cost more than the exchange price, or it may cost less. Of course, machine downtime costs money and the main advantage of an exchange program is that downtime is minimized.

If downtime is not a concern, then it is wise to obtain a repair quote for the component before opting for an exchange unit. This is one reason for scheduling component change-outs upon completion of their expected service life. If component rebuilds can be scheduled when machine downtime is not an issue, it will not be necessary to pay a premium for an exchange unit.

Hydraulic Component Manufacturers
When a hydraulic component needs to be repaired, a logical option is to send it to the company that manufactured it. Note here that the shop being referred to is a repair shop owned and operated by the hydraulic component manufacturer, rather than an independent hydraulic shop that is an authorized repairer or distributor of the manufacturer’s products.

By using the component manufacturer, a repair to OEM standards will be achieved, but in some cases this can mean foregoing the opportunity to save money. Consider the following example.

Let’s say a radial piston motor has been removed from a hydraulic machine and shipped to the motor manufacturer for repair. Upon dismantling the motor, it is discovered that several of the pistons and bores are badly scored as a result of contamination. The manufacturer suggests that because of this damage, the motor housing and all of its pistons need to be replaced. As a result, the manufacturer’s price to repair the motor is almost as much as a new one.

You don’t want to buy a new motor if you can avoid it, so you decide to get a second repair quote. This time you take the unit to an independent hydraulic shop. After inspecting the motor, the repair shop advises that it can rebuild it for around 60 percent of the cost of new one - a significant saving on a $10,000 motor!

The lower cost is possible by employing a money-saving repair technique that involves machining the piston bores to remove scoring, and fitting a set of oversize pistons, thereby salvaging the housing. The oversize pistons are nongenuine or aftermarket parts, however their quality has been proven in a large number of rebuilt units. The repair shop’s confidence in this method of repair is supported by a 12-month warranty.

Independent Hydraulic Shops
This example highlights the type of savings possible by using a reputable, independently-owned hydraulic shop to carry out your component rebuilds. The motor manufacturer didn’t offer this lower-cost solution because the oversize pistons are nongenuine parts, therefore the repair is not to OEM standard. However, just because a repair is not carried out to OEM standard doesn’t mean the rebuilt component won’t perform and last like a new one.

It is important to distinguish between a properly engineered and proven repair technique that will save you money and a dubious repair that will likely cost twice as much in the long-run. Always ask the repair shop two questions:

• Has the repair technique been proven successful in terms of performance and achieved service life?

• Is the repair covered by warranty?

If the repair technique is unproven, or the shop offering the repair is not willing to back it up with a warranty, think twice before proceeding. You need to make an informed decision based on the amount of money the repair technique will save you if it works, versus the cost if it doesn’t. The repair shop may be willing to share some of the risk involved in finding out, given that if the technique proves to be successful, it can offer the same solution to other customers. These issues should be discussed with the repair shop before making a decision.

http://www.machinerylubrication.com/

Throughout the world, injection molding machine (IMM) operators work with high-temperature plastics, high injection pressures, powerful clamping and ejection forces and rapidly moving steel masses. Though these would seem to present a potential for injury, today’s workers enjoy many safeguards against the potential hazards associated with machine operation. These safe-guards are the result of IMMs that are manufactured and used in accordance with today’s safety standards, such as ANSI/SPI B151.1-1997.

These IMM safety standards are the result of the collaborative efforts of manufacturers and molders to:

1) identify potential hazards relating to IMM operation and use, and

2) develop performance criteria for safeguards to protect workers from those potential hazards.

In North America, the Society of the Plastics Industry, Inc. (SPI) plays a key role as a coordinator of such activities.

SPI believes this same sort of collaboration should be applied to the development of a mold safety standard.

Background
IMMs that comply with the ANSI/SPI B151.1-1997 safety standard have multiple safety systems that protect operators from potential mold area hazards, such as platen closing motion and machine ejector action. Though the IMM clamp is prevented from closing by multiple interlocks, the mold-ejection system, which could include springs or motions powered independently from the IMM, may create pinch points with the IMM gate open. Also, as the mold is sometimes operated at high temperatures, access to hot mold surfaces could cause serious burns. In response to these concerns, the moldmakers of North America, under the coordination of SPI, have established a mold safety committee to develop an ANSI safety standard for the integration, care and use of molds used with IMMs.
In developing this standard, the moldmakers face a challenge that is unique to their industry. This challenge is that no mold safety standard now exists. Thus, this mold standard must be developed without the benefit of experience from previous standards. By contrast, the IMM manufacturers and molders have a long tradition of safety standards for IMMs, with the first one being published in 1976.

Unlike the IMM industry, the moldmaking industry is made up of many small companies, which makes SPI’s task of achieving communication, organization and consensus for a safety standard much more difficult. However, SPI has accepted these challenges and has taken the lead in this effort, as it did in the creation of the IMM standards.

Safety Issues
The following are some of the mold safety issues that the SPI mold safety committee will be addressing in developing this standard:

• Mold Mounted Guards: In the event a mold has components that could be potentially hazardous when the machine’s operator gate is open or has components that extend beyond the guarding of the machine, guarding attached to the mold may be required to prevent injury.
• Interlocking Mold Motion: Many molds use hydraulic or pneumatic cylinders to move mold elements. These may be independent of the machine control. If these mold elements are energized when the operator gate is open, potential pinch points may result.
• Mold Electrical Safety: Many molds, particularly those with hot runners or hot sprues, use electrical power supplied either by the IMM or by a separate controller. This power can be substantial and very often is located in a high-temperature area of the mold, which could result in a high-temperature and/or high-voltage hazard.
• Mold Mounting and Handling: Practices and devices for safe lifting, handling and storing of the assembled mold and mold sub-assemblies will be addressed.
• Warning Signs and Instructions: Where potential mold hazards cannot be eliminated by design, effective warning signs and instructions are needed.

SPI Takes Action
To address safety issues surrounding proper usage of molds, SPI has established a committee to develop an "American National Standard for Molds Used with Horizontal Injection Molding Machines - Safety Requirements for the Integration, Care and Use."

The standard’s objective will be to minimize hazards to personnel associated with mold activity by establishing recommendations for the manufacture, care and use of molds. To accomplish this goal, the SPI Machinery and Moldmakers Divisions’ Standards Development Committee decided to approach mold safety from two directions:

1. Eliminating recognized hazards and establishing standard approaches to design so that molds available from competitive manufacturers will have similar safety features, and
2. Safeguarding personnel from recognized mold hazards.

Acknowledging the impossibility of updating equipment and changing operational methods associated with existing molds immediately after the approval date of this standard, a grace period will be provided to employers for updating existing molds. Likewise, recognizing the impossibility of immediate updating of design and manufacturing methods, the Manufacture, Remanufacture and Modification and the Safety Signs clauses would become effective one year after the approval date of this standard.

The standard is based upon the following facts:

• Molds are an essential element in the production of plastic parts and goods.
• They are installed in horizontal molding machines.
• They are complicated instruments with moving parts and mechanisms.
• They require trained, skilled operators/technicians.
• They can be large and extremely heavy.
• There are numerous safety considerations for molds, including electrical, thermal, mechanical and pneumatic/hydraulic.
• The molds may come in human contact on every cycle of the machine.

Among its requirements, the proposed standard would:

• Establish responsibility for instructions, maintenance and inspections.
• Seek to ensure that modifications and repairs do not diminish safety levels.
• Require guarding or signage where hot surfaces create hazards.
• Require an indication of mold weight on the mold.
• Set guidelines for lifting assembled molds, sub-assemblies and components.
• Require that molds be designed and packaged to ensure safe storage.
• Set parameters for indication of "top" and "operator’s side" on the mold and in instructions.
• Require that mold vents be designed to protect operators from hot plastic spray.
• Set safety guidelines for ejector housing and mechanisms and certain other moving parts.
• Set requirements for stack molds, including the design and function of drop-limiting devices and guarding of hot spruebars to prevent burns.
• Address safety issues involved with unscrewing mold mechanisms.
• Require that hot-runner protection, including plugs, be designed for the pressures and conditions of the process.
• Address mold-cavity safety issues.
• Establish guidelines for special mold function mechanisms, motion/no motion, air systems and air, water and hydraulic service identification.
• Address modification of HIMM and/or robot guards to fit a mold.
• Set parameters for electrical wire channel moisture drains for hot-runner molds.
• Develop electrical standards unique to the high temperature requirements of hot runners.
• Address electrical service routing and sharp edges.
• Address the use of springs and window materials.
• Establish guidelines for training in the use of ancillary equipment and protective gear.
• Mandate guarding to restrict access to pinch and shear points in the mold.
• Address issues involved in interrupted cycles, automatic mold changes, motion interlocks and valve gates.
• Set requirements for appropriate signage.

At the time of this writing, the SPI Committee on Mold Safety is developing a working draft of this standard. While schedules are difficult to predict, it is anticipated that a final draft of a Mold Safety Standard could be balloted in the year 2000. A Mold Electrical Safety Standard will follow.

SPI members involved in the standard’s development include representatives from Husky Injection Molding Systems, Blue Water Plastics, Milacron Inc., DME, HPM, Landis Plastics, National Tool & Manufacturing, Plastool, Pleasant Precision, Progressive Components, RPK Tool & Die, Fast Heat, Stott Tool & Machine Co., Superior Die Set, Dynisco Hot Runners, Incoe and Van Dorn Demag.

As stated, creation of a mold standard is not an easy task. It is, however, a necessary one. That is why these representatives met this challenge and took the initiative to develop a safety standard for the industry.

http://www.moldmakingtechnology.com/articles/010006.html

Many industrialized countries have regulations restricting noise levels in the workplace. The high-power density and corresponding high noise emission of hydraulic components cause industrial hydraulic systems to be the target of efforts to reduce mean noise levels.

The pump is the dominant source of noise in hydraulic systems. It transmits structure-borne and fluid-borne noise into the system and radiates air-borne noise.

All positive-displacement hydraulic pumps have a specific number of pumping chambers, which operate in a continuous cycle of opening to be filled (inlet), closing to prevent back flow, opening to expel contents (outlet) and closing to prevent back flow. These separate but superimposed flows result in a pulsating delivery, resulting in a corresponding sequence of pressure pulsations. These pulsations create fluid-borne noise, which cause downstream components to vibrate. The pump also creates structure-borne noise by producing vibration in any component it is mechanically linked to, for example, the tank lid. The transfer of fluid- and structure-induced vibration to the adjacent air mass results in air-borne noise.

Reducing Fluid-borne Noise
While fluid-borne noise caused by pressure pulsation can be minimized through hydraulic pump design, it cannot be completely eliminated. In large hydraulic systems or noise-sensitive applications, the propagation of fluid-borne noise can be reduced by the installation of a silencer. The simplest type of silencer is the reflection silencer, which eliminates sound waves by superimposing a second sound wave of the same amplitude and frequency at a 180-degree phase angle to the first.

Reducing Structure-borne Noise
Structure-borne noise created by the vibrating mass of the power unit (the hydraulic pump and its prime mover) can be minimized through the elimination of sound bridges between the power unit and tank, and the power unit and valves. This is normally achieved with the use of flexible connections, such as rubber mounting blocks and hoses. However, it is necessary to introduce additional mass in certain situations, where the inertia reduces the transmission of vibration at bridging points.

Reducing Air-borne Noise
The magnitude of noise radiating from an object is proportional to its area and inversely proportional to its mass. Reducing an object’s surface area or increasing its mass can therefore reduce its noise radiation. For example, constructing the hydraulic reservoir from thicker plates, which increases its mass, will reduce its noise radiation.

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Reducing%20Hydraulic%20System%20Noise

As a product group, cylinders are as common as pumps and motors combined. Therefore, if a plant operates a lot of hydraulic equipment, cylinder repair expense is likely a significant portion of total maintenance costs.

It is often stated that up to 25 percent of mechanical equipment failures are design related. With regards to hydraulic cylinders, this suggests as many as one in four are not adequately designed for the application they are operating in. This doesn’t mean the cylinder won’t perform the required job asked of it, it will - but not with an acceptable service life. If a particular cylinder requires frequent repair, one or more of the following design-related problems need to be addressed.

Bent Rods
The bending of cylinder rods can be caused by insufficient rod diameter, material strength, improper cylinder mounting arrangement or a combination of all three. Once the rod bends, excessive load is placed on the rod seal resulting in premature failure of the seal.

Rod straightness should always be checked when a hydraulic cylinder is being repaired. To test for straightness, place the rod on rollers and measure the run-out with a dial gauge. Position the rod so the distance between the rollers (L) is as large as possible, then measure the run-out at the midpoint between the rollers (L/2).

The rod should be as straight as possible, however 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) = 0.5 x L ÷ 1,000
Where L equals the distance between rollers in millimeters

For example, if the distance between the rollers is 1.2 meters, then the maximum, allowable run-out measured at L/2 would be given by 0.5 x 1,200 ÷ 1,000 = 0.6mm.

If a rod is bent, actual rod loading should be examined against permissible rod loading based on the cylinder’s mounting arrangement and the tensile strength of the rod material. The formulas and procedure for this are available from www.IndustrialHydraulicControl.com. If actual rod 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.

Figure 1. Testing Rod Straightness

Ballooned Tubes
Ballooning of the cylinder tube is typically caused by insufficient wall thickness and/or material strength for the cylinder’s operating pressure. Once the tube balloons, the correct tolerance between the piston seal and tube wall is lost and high-pressure fluid bypasses the seal. This high velocity fluid can erode the seal, and localized heating caused by the pressure drop across the piston reduces seal life.

Testing Tube Integrity
The conventional way of testing the integrity of the piston seal in a double-acting cylinder is by pressurizing the cylinder at the end of stroke and measuring any leakage past the seal. This is often referred to as an end- of-stroke bypass test.

A major limitation of the end-of-stroke bypass test is that it generally does not reveal ballooning of the cylinder tube caused by hoop stress as a result of under-designed cylinder wall thickness or reduction of wall thickness through excessive honing. The ideal way to test for ballooning of the cylinder tube is to conduct a piston-seal bypass test midstroke. A disadvantage of this procedure is that the force developed by the cylinder has to be mechanically resisted, which in the case of large-diameter high-pressure cylinders, is impractical.

However, a midstroke bypass test can be conducted hydrostatically using the intensification effect. The necessary circuit is shown in Figure 2.

Pressure Intensification
Force produced by a hydraulic cylinder is a product of pressure and area (F = p x A). In a conventional double-acting cylinder, the effective area and therefore force produced by the piston and rod sides of the cylinder are unequal. It follows that if the rod side of the cylinder has half the effective area of the piston side, it will produce half the force of the piston side for the same amount of pressure.

The equation F = p x A can be transposed as p = F/A: that is, pressure equals force divided by area. For the rod side of the cylinder to resist the force developed by the piston side, with only half the area, then the pressure needs to be doubled. This means that if the piston side is pressurized to 3,000 PSI, a pressure of 6,000 PSI will be required on the rod side to produce an equal force, which explains why pressure intensification can occur in a double-acting cylinder.

If for any reason the piston side of a double-acting cylinder is pressurized and at the same time fluid is prevented from escaping from the rod side, pressure will increase in the rod side of the cylinder until the forces become balanced or the cylinder fails catastrophically. Intensification of pressure in a double-acting cylinder is a dangerous phenomenon and the concept must be thoroughly understood when testing hydraulic cylinders using the following procedure:

1. Secure the cylinder with its service ports up.

2. Fill both sides of the cylinder with clean hydraulic fluid through its service ports.

3. Connect ball valves (1) and (2), gauges (3) and (4), relief valve (5) and directional control valve (6) as shown in Figure 2.

   
   Figure 2. Cylinder Test Circuit
4. With ball valves (1) and (2) open, stroke the cylinder using the directional control valve (6) multiple times to remove all remaining air from both sides of the cylinder – take care not to “diesel” the cylinder.

5. Position the piston rod midstroke and close ball valve (2).

6. With the adjustment on the relief valve (5) backed out, direct flow to the rod side of the cylinder.

7. Increase the setting of relief valve (5) until the cylinder’s rated pressure is seen on gauge (3).

8. Close ball valve (1) and center directional control valve (6). Note: it is assumed that the hydraulic power unit used to conduct the test has its own overpressure protection.

9. Record the respective pressure readings on gauges (3) and (4) and monitor any change over time.

If the ratio of effective area between the piston and rod side of the cylinder is 2:1, and if the rod side of the cylinder has been pressurized to 3,000 PSI, gauge (4) on the piston side should read 1,500 PSI. If the differential pressure across the piston seal is not maintained, this indicates a problem with the piston seal and/or tube.

Under no circumstances should flow be directed to the piston side of the cylinder with ball valve (1) closed because failure of the cylinder and personal injury could occur. When conducting this or any other hydrostatic (pressure) test, always wear appropriate personal-protection equipment.

Insufficient Bearing Area
If the internal bearing areas in the gland and at the piston are insufficient to carry the torsional load transferred to the cylinder, excessive load is placed on the rod and piston seals. This results in deformation and ultimately premature failure of the seals.

Rod Finish
The surface finish of the cylinder rod can have a dramatic effect on the life of the rod seal. If the surface roughness is too low, seal life can be reduced through inadequate lubrication. If the surface roughness is too high, contaminant ingression is increased and an unacceptable level of leakage can occur.

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How%20to%20Maximize%20Hydraulic%20Cylinder%20Service%20Life

When things go awry with a piece of hydraulic equipment, the maintenance technician is usually the first on the scene. For the technician’s troubleshooting efforts to be effective, he or she must understand how the equipment operates. One type of hydraulic control system in widespread use, but not well understood, is load-sensing control.

Load-sensing describes a type of variable pump control used in open circuits. It is also termed this 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 typically a directional control valve with proportional flow characteristics, but a needle valve or even a fixed orifice can be employed, depending on the application.

Power-saving Control
In hydraulic systems subject to wide fluctuations in flow and pressure, load-sensing circuits can save substantial amounts of input power (Figure 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 required only intermittently, the remaining 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).

A similar situation occurs in systems fitted with pressure-controlled (pressure-compensated) variable pumps, where 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 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 Figure 1).

Figure 1. Flow-pressure-power Diagrams for Fixed,
Variable and Load-sensing Controlled Pumps (Peter Rohner)

Circuit Configuration
A load-sensing circuit typically has 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 (Figure 2).

Figure 2. Typical Load-sensing Circuit

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 the control valve sections through a series of shuttle valves. This ensures the actuator with the highest load pressure is sensed and fed back to the pump control.

To understand how the load-sensing pump and directional control valve operate together, 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 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.

Constant Pressure Drop Equals Constant Flow
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 slightly increasing pump displacement (flow) 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.

High-end load-sensing directional control valves feature a pressure compensator at the inlet to each valve section. The section pressure compensator works with the spool-selected orifice opening to maintain a constant flow rate, independent of the pressure variations caused by the operation of multiple functions at the same time. This is sometimes referred to as “sensitive load-sensing”.

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859&pagetitle=Understanding%20Load-sensing%20Control

Hydraulic systems are often considered perennial consumers of oil and in turn, makeup fluid is an inherent cost of operating hydraulic equipment. But what is the real cost of one or more minor leaks on your hydraulic equipment? To answer this question, the costs associated with the following factors need to be considered:

• Makeup fluid
• Cleanup
• Disposal
• Contaminant ingress
• Safety

Makeup Fluid
Makeup fluid should be the most obvious cost of hydraulic system leaks. I say “should be” because many hydraulic equipment users fail to consider the accumulative effect of the cost of one or more slow leaks over time.

Consider a piece of hydraulic equipment losing six cubic centimeters of oil per minute. Over 24 hours, the loss is 0.9 liters, which perhaps is not a significant amount. But over a month this equates to 27 liters, and 330 liters over the course of a year. Assuming a fluid cost of three dollars per liter, the annual cost is around $1,000.

Cleanup
Where oil leaks occur, there are almost always cleanup costs to consider, which include:

• Labor
• Equipment required to empty sumps and drip trays and degrease machine surfaces
• Consumables such as detergents and absorbent materials

Assuming it costs $20 per week in labor, equipment and consumables to clean the piece of equipment discussed above, the annual cleanup bill totals more than $1,000.

Disposal
I remember when waste oil companies paid for the privilege of emptying waste hydraulic oil tanks. These days, companies must pay for their waste to be discarded. Environmentally acceptable disposal of waste oil and absorbent material containing waste oil costs money.

Assuming a transport and disposal cost of one dollar per liter, the annual disposal costs attributable to the leakage discussed above amounts to $330.

Contaminant Ingress
When oil leaks out, contaminants such as air, particles and water can get in. The costs to consider here include:

• component damage and fluid degradation as a result of contaminant ingress
• equipment reliability problems
• removal of ingested contaminants

Safety
Oil leaks regularly pose a safety hazard. Like the costs associated with contaminant ingress, the costs associated with the hazards of oil leaks are difficult to quantify. However, active management of the safety risk posed - for example, more frequent cleanup than may otherwise be necessary - skews this cost to a quantifiable area.

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The%20Real%20Cost%20of%20Fluid%20Power%20Leaks

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 fullfilm lubrication is improved and boundary lubrication conditions are less likely to occur.

Hydrostatic force is the product of pressure and area. Expressed mathematically, the equation is 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 percent of Area A. This means that the force exerted on the lubricated surfaces at the end of the cylinder rod is 20 percent 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 (Area C) and fluid viscosity are adequate, full-film lubrication of the sliding surfaces is achieved.

The same principle applied to a typical axial design piston is illustrated in Figure 2.

Figure 2. Cross-section of an axial design piston.

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 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 percent of Area A and Area C is 140 percent of Area A. This means that the force transmitted to Area C is half of the force 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, for example, 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 percent of the force developed by the pressure acting on Area A. If full-film lubrication is dependent on the hydrostatic balance of the cylinder, boundary lubrication conditions will eventuate and two-body abrasion is likely.

Figure 4. Cross-section of axial piston
showing blocked balance drilling.

Applied to an axial piston, this is equivalent to blocking the balance drilling by particle contamination (Figure 4). As a consequence, 100 percent of the force developed by the pressure acting on Area A is transferred to Area C, most likely resulting in boundary lubrication and twobody abrasion between slipper and swash plate.

Because hydrostatic force is a product of pressure and area (F = P x a), hydrostatic balance is affected by changes in either pressure or area. As illustrated in Figures 3 and 4, blockage of balance drillings by contamination results in loss of hydrostatic balance. Wear caused by two-body and three-body abrasion can affect hydrostatic balance by altering the areas where balancing pressure acts or by reducing the effective balancing pressure.

To illustrate this, consider the sliding surface of an axial piston slipper. The inner edge of the slipper’s sliding surface is a load concentration point. Deformation of this area can result in localized contact (two-body abrasion) between slipper and swash plate.

Figure 5. Mushrooming of
slipper surface reduces
effective balancing area.

This causes the slipper’s surface to mushroom (Figure 5), creating an area that acts to increase the hydrostatic force, and therefore load, on the lubricated surfaces of the slipper and swash plate. Once the slipper begins to mushroom, a cycle of increased load and wear follows (Figure 6), leading to loss of hydrostatic balance and slipper failure.

Figure 6. Advanced mushrooming
and wear of piston slipper.

The slipper’s sliding surface acts as a seal for hydrostatic balancing pressure.

Figure 7. Heavy scoring
of piston slipper.

If this surface becomes severely scored as a result of threebody abrasion (Figure 7), leakage can increase to the point where a pressure drop develops between the piston area and the slipper’s balancing area. This reduces the hydrostatic balancing force and increases the load on the lubricated surfaces of the slipper and swash plate. The resulting cycle of increased load and wear leads to loss of hydrostatic balance and slipper failure (Figure 8).

Figure 8. Failure of piston
slipper caused by loss of
hydrostatic balance.

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Hydrostatic%20Balance%20in%20Hydraulic%20Component%20Design