Preventing brush seal damage starts with proven and stable production processes with suitable quality controls. Important characteristics of the quality of a brush seal are the evenness of the brush in the brush angle, packet density, and secure fastening of the bristles. The bristle tips must have a suitable geometry and surface structure opposite the glide surface. Secure fastening of the brushes is a prerequisite for use of the seal near roller bearings (Ref. 7.3.3-1) and in the flow of cooling air to the turbine rotor blades. With regard to this principle, brush seal designs in which the bristles are positive-fit into the seal can be seen as especially safe. Even seemingly minor changes to the production process should be undertaken only after sufficient testing of their effectiveness in operation.
Damage and mistakes during brush seal mounting are to be avoided through appropriate precautions. Special assemblies may be necessary for installing and removing brush seals. The brush seal must be sufficiently accessible so that the clearance gap can be measured and confirmed.
A large portion of brush damage can be avoided by proper configuration of the seal, so this should be the starting point for remedies (Fig. "Designing brush seals", Ref. 7.3.3-1). For example, overloading of the brush seal due to overly high pressure differences is to be avoided (Fig. "Brush seal limits"). At high pressure differences, the supporting backing plate is subjected to bending loads and must be designed so that it is sufficiently stiff. Otherwise, the backing plate plastically bends through backwards.
Rubbing is a critical phase, in which the goals must be low wear and low friction heat. This can be achieved by selection of a tribo-system that is matched to the operating conditions (Fig. "Brush seal determines operation"). Finding the right material combination for the brush/rotor may require extensive realistic trials in suitable brush testing rigs (Fig. "Designing brush seals"4). If operating experience is already available, it should be made use of.
The rub surface of the rotor is important for the operating behavior of the brush. Thickness, structure, thermal conductivity, and tribo-properties are a few of the criteria for material selection. Thermal spray
coatings with appropriately matched compositions can minimize the wear during rubbing and ensure acceptable long-term performance.
The wear-related life span of a brush is dependant on the operating temperatures. Oxidation weakens the thin bristles. Rub surfaces are worn out more rapidly and spalling oxide layers accelerate erosion in the brush (Fig. "Brush seal bristle tips damage symptoms"). For material selection, it is hard to avoid long-duration tests under realistic oxidizing conditions on original brushes, since the wear behavior must be seen in connection with fretting occurring inside the brush.
The stiffness of the brush acts against radial deflection by the rotor and affects rubbing in this way. High brush stiffness increases the tolerable pressure differences, but also leads to intensive rubbing. It is especially dangerous if the frame encloses the brush so much that bristles jam inside it and further increase the rubbing process. In order to avoid this, the bristle heights (difference in radial height between the front and backing plates; Fig. "Brush seal specific terms") must be sufficiently large.
Proper selection of the fence height is extremely important. If the fence height is too large, it decreases the load capacity of the brush due to blow down (Fig. "Brush seal limits"). Insufficient fence height, i.e. insufficient clearance between the rotor and backing plate, must definitely be avoided, since otherwise there is danger of self-increasing rubbing occurring ( ).
Excess brush length has many drawbacks. These brushes are at high risk of being damaged by accidental reverse rotation of the rotor. Wear will eliminate the improved seal effectiveness the excess brush length provides after very few start-up/shut-down cycles.
Whirled inflow into the brush causes uneven wear around the circumference (Fig. "Brush seal damaging turbulence"), as well as tangling of the bristles. Therefore, rotating screw heads, nuts, and discrete air jets must be avoided in the space in front of the brush (Fig. "Brush seal problems by turbulence"). Installing suitable deflectors in front of the brush can prevent both whirled inflow and powerful blow-down (Fig. "Preventing bending of brush seal bristles").
Figure "Designing brush seals" (Ref. 7.3.3-1): The depicted methodical procedure during design and construction of a brush seal is based on understanding of the physical processes that influence brush function, experience with brush seal behavior in gas turbines, and testing rig results. This systematic approach is used for brush seals that are designed to seal air. The depicted method makes designs for the use of brush seals in hot and cool engine areas possible.
If the necessary data for brush design are not available and/or were not provided by the brush seal manufacturer, then they must be determined by separate tests, which are often expensive and time-consuming. This is the case with special, application-specific tribo-systems with rotor coatings. The direction of the leakage flow must prevent loose bristles or bristle particles from damaging the roller bearings.
The depicted designed system makes it possible to take advantage of the excellent seal effectiveness and mechanical security (relative to labyrinths) for the engine. However, the seal effectiveness of brush seals does not compare with that of floating ring seals, which are used primarily for sealing oil/air in bearing chambers (Fig. "Brush seal limits").
Different analytic disciplines are used during design. Analysis of the stresses in the brush and analysis of the security against failure due to dynamic fatigue in the HCF region are both conducted within the framework of the geometric data, which take into account the available space for installation, the friction heat, and the seal requirements. The actual leakage air flow during operation must be determined with the aid of data that were empirically derived from testing rig experiments.
The same is true for wear data (Fig. "Brush seal determines operation"). If necessary, these must be determined for the specific tribo-system (e.g. special coatings on the rub surface of the rotor; also see Ref. 7.3.3-3).
Coefficients of friction and the calculated evaluation of the friction heat created during rubbing are covered in Refs. 7.3.3-2, 7.3.3-4, and 7.3.3-6).
Several iteration steps are necessary to fulfill the requirements for tolerable pressure differences, sufficiently small whirl in front of the brush, and good performance under extreme conditions and unsteady operating conditions.
Figure "Brush seal determines operation": During rubbing, the heat created and its dissipation control the heating-up of the brush and, therefore, its reliability (Fig. "Brush seals rules preventing damages"). The top diagram (Refs. 7.3.3-2 , 7.3.3-4 and 7.3.3-6) shows the increase of the friction heat along with the excess brush length. The shorter the “effective” bristle length (depending on the brush stiffness, which is also influenced by the friction inside the brush and against the backing plate), the stiffer the brush (higher friction forces). The higher the friction forces, the more friction heat is created. Large leakage air flows can dissipate more heat and prevent dangerous overheating of the brush. In order to minimize heat creation, the rotor coating must be as smooth as possible and suitably matched with the brush material. If the rotor coating dissipates heat well, it is also good for the brush. Uncoated rotors must also be viewed with regard to this positive aspect.
The middle diagram (Ref. 7.3.3-3) shows that the coefficient of friction, which is an important value with regard to heat creation during rubbing, is clearly, material-specifically, dependent on the rubbing speed. The given data were determined for “miniature brush seals” with 0.25 mm radial excess and a testing rig temperature of 426 °C. The coating-specific, very different progression of the coefficient of friction is surprising.
The low coefficient of friction of NiCrAlY brushes is due to the tough oxide film on the contact surface, as well as a lubricating film of Ba- and Ca-fluoride. This film forms on the rub coating of the rotor, which is made of a thermal-sprayed Cr-carbide coating with Ba and Ca additives.
Al2O3 and Cr2O3 rub-tolerant coatings exhibit micro-spalling (grain outbreaks), which create a high coefficient of friction and heavy brush wear (Fig. "Brush seal track damages"). Co-alloys such as Haynes 25 are known for their low coefficient of friction at temperatures under 400°C. The steep increase in the coefficient of friction above 400°C is due to a structural change.
The bottom diagram shows brush wear opposite the wear of various thermal-sprayed rotor coatings (Ref. 7.3.3-5). The data were taken from a testing rig with complete brush seals running for 100 hours at room temperature.
Figure "Brush seal bristle oxidation" (Ref. 7.3.3-3): The left diagram shows the typical oxidation behavior of a nickel alloy:
A relatively short incubation time “I”, during which the weight of the probe stays almost unchanged.
At first, accelerated weight increase in region “II”, where an oxide film forms. In region “III” the weight increase is continual, and the weight-time curve has a corresponding linear progression. Here, the oxidation is balanced with oxygen diffusion through the oxide layer.
The right diagram shows the oxidation behavior of three potential bristle alloys at an annealing temperature of 1038°C. The oxidation resistance depends on the type of oxide film that forms
The NiCrAlY alloy with the highest oxidation resistance forms strongly-adhering dense(Al or Cr)2O3 oxides, which are stabilized by rare earth metal supplements.
On the other hand, Inconel 718 and Haynes 25 (Co-alloy) form non-stabilized Cr2O3 films, which are considerably less oxidation resistant. These oxides do not form a diffusion barrier for oxygen ions. Therefore, the oxidation process continues, which is the deciding difference between these materials and those that form oxygen-sealing oxide layers.
Figure "Brush seal leakage by backing plate": The gap between the backing plate and rotor (fence height) requires special attention during brush design. A small fence height is vital to ensure a low leakage air rate (top left diagram). However, it must not touch the rotor, which could result in self-increasing rubbing (bottom left diagram). If this cannot be ensured, then the rotor must be protected with a sufficiently thick and wide, insulating, wear- and heat-resistant coating. The coating must not be worn through to the base material under any circumstances (top right diagram).
The thickness/stiffness of the backing plate must be selected so that no plastic deformation occurs due to the pressure differences throughout the designated life span at operating temperatures. The bottom right diagram depicts a variation (Ref. 7.3.3-2) where the backing plate was extended with an abradable coating. In this way, contact should not result in damage to either the rotor or the seal.
Figure "Brush seals for high pressure ratio" (Ref. 7.3.3-7): For a short distance, the leakage flow through the brush is deflected radially inward by the bristles and the backing plate. It then blows the bristles inwards towards the rotor and decreases the leakage gap. This is referred to as “blow down” or “pressure closure” (top left diagram). Friction inside the brush is caused by the bristles laying against one another and against the backing plate. This stiffens the brush and prevents the bristles from springing back after the rotor “moves back” (clearance increase). In order to minimize the friction against the backing plate and, therefore, the brush hysteresis (Fig. "Brush seal leakage behavior"), the backing plate is machined out at the lay-on surface (top right diagram) so that the bristles are free and can spring back due to blow down (LHS brush=Low Hysteresis Seal). In this way, the dependency of the brush stiffness on the differential pressure is also minimized considerably (bottom diagram). The drawback of this brush type is that it is especially sensitive to bristle flutter. Vibrations and tangling of the bristles are cause by whirl in the flow in front of the brush (Fig. "Brush seal problems by turbulence") and around the backing plate. A possible remedy for this is mounting a flow deflector at the front plate (top right diagram).
Figure "Preventing bending of brush seal bristles": Flow deflectors can be made from sufficiently strong perforated metal sheets or porous metal felts. Another possible remedy may be affixing a cover, separate from the brush (Fig. "Designing surrounding of brush seals").
Figure "Design for removal of brush seals": When mounting and removing brush seals, special precautions are necessary in order to avoid the following damaging situations from occurring:
Figure "Designing surrounding of brush seals": In order to prevent brush seal damage due to bristle flutter (see Fig. "Design for removal of brush seals"), unallowable whirl must be suppressed in the space in front of the brush. In the case shown in the top diagram, the brush was rapidly destroyed by uneven wear and entanglement, because the rotating screw heads caused dangerous whirl.
The middle diagram depicts a screw cover fastened onto the rotor to minimize whirl. The bottom diagram shows another configuration which does not prevent whirl, but a ring-shaped cover protects the front side of the brush.
Figure "Brush seals testing design parameters": The complex loads on a brush seal combined with various simultaneously-acting damage mechanisms prohibit satisfactory, purely theoretical procedures in order to ensure that a configuration is sufficiently safe (Fig. "Designing brush seals"). For this reason, developing brush seals requires testing in order to understand the influence of the operating parameters and estimate the operating life span.
Technical literature describes various testing rigs for brush seals. The assemblies are different, depending on the properties that are to be tested (e.g. seal effectiveness, rubbing behavior, maximum pressure differences, sensitivity to flow whirl). If one wants to inspect the realistic seal effectiveness of a brush seal in operation, it must be done in a testing rig in which the entire seal is being tested, in its original size with realistic pressure ratios, leakage gaps, and temperatures. Due to the necessary high pressure (long clearance gap), these tests are very elaborate. This is especially true if the pressurized air must be heated in order to simulate hot part conditions. Testing rigs with only one brush (top left diagram) have a problem of high thrust loads on the rotor. If two brushes are installed opposite one another to compensate for the thrust loads and the air inlet is between them (top right diagram), then the testing rig becomes even more elaborate and requires twice the amount of leakage air.
With long-duration tests for determining the wear behavior of the brush/rotor tribo-system or wear (fretting) and oxidation of the brush, the expense and effort required for testing full-sized brushes is unbearable. Therefore, special testing rigs were designed for these tests, which only inspect a brush segment in a pressure chamber, which can be moved against the rotor along with the brush (bottom diagrams). This configuration has the advantage of needing only small amounts of air, because the gap is closed during rubbing. If one wall of the chamber is made of armored glass, then the brush can be observed during operation, and its behavior can be documented. The rub track on the rotor can be directly observed throughout the entire test, which allows changes to be recognized immediately during operation or after altering testing parameters. If the pressure in the chamber puts too much radial force on one side of the rotor, a balancing force must be applied to the other side (bottom right diagram).
7.3.3-1 G.F. Holle, M.R. Krishnan, “Gas Turbine Engine Brush Seal Applications”, Paper AIAA 90-2142 of the “26th Joint Propulsion Conference” July 16-18, 1990, Orlando,FL, pages 1-9.
7.3.3-2 R.Flower, “Brush Seal Development System”, Paper from 1990, AIAA-90-2143, pages 1-8.
7.3.3-3 J.Derby, R.England, “Tribopair Evaluation of Brush Seal Applications”,Paper AIAA 92-3715 of the AIAA/SAE/ASME/ASEE 28th Joint Propulsion Conference, July 6-8, 1992, Nashville TN, pages 1-14.
7.3.3-4 R.C. Hendricks, M.J. Braun, V. Canacci, R.L. Mullen, “Brush Seals in Vehicle Tribology”, Paper IX (i) of the Symposiums Sept. 1990 Leeds/Lyon, pages 231-242.
7.3.3-5 E. Atkinson, “Effects of Material Choices on Brush Seal Performance”, periodical “Lubrication Engineering”, September 1992, pages 740-745.
7.3.3-6 R.C. Hendricks, S.Schlumberger, M.J. Braun, F.Choy, R.L. Mullen “A Bulk Flow Model of a Brush Seal System”, ASME Paper 91-GT-325 of the “International Gas Turbine and Aeroengine Congress and Exposition” , Orlando,FL, June 3-6, 1991.
7.3.3-7 J.F. Short, P. Basu, A. Datta, R.G. Loewenthal, R.J. Prior, “Advanced Brush Seal Development”, Paper AIAA-96-2907 of the “32th AIAA, ASME, SAE, ASEE Joint Propulsion Conference” July 1-3, 1996, Lake Buena Vista, FL, pages 1-8.