mud pump broken bolts brands

Drill-out is a drill-bit and extractor in one tool. Drill-out incorporates a unique interaction between drill bit and extractor resulting in maximum impact and break out ability. Serrated extracting end of tool results in greater torque for hard to remove bolts.

Every mechanic knows the horrible feeling after they have turned a bolt too much and it snaps. The time invested to get your equipment back to work will now take hours or days to fix the problem of a broken bolt.

This is where the Broken Bolt Extractor from Brokenbolt.com comes in. Proven to reduce the time of drilling out a broken bolt to minutes, not hours or days. Without having to send the part out to be remachined, salvaging the original threads.

The inventor of this tool, Chuck Fulgam, has story after story of how this tool has helped many equipment owners recover quickly from a broken bolt repair issue, saving them hundreds or thousands of dollars in repair bills and hours of downtime.

Used by the US Military for many vehicle applications, they immediately saw the potential in both time savings and potentially life threatening situations in the field. Broken bolts can have a disastrous effect in the field if you need to be on the move. The faster repairs are completed, the better. And the Broken Bolt Extractor from Brokenbolt.com helped the military to accomplish this.

Fleet maintenance services have also seen a huge advantage in adding this broken bolt extractor tool in their tool box. From huge time savings advantages to getting the vehicle back on the road. Saving from downtime and increasing the productivity of the vehicle.

Broken Bolts? Not a problem anymore Broken exhaust manifold bolts can be the start of a daunting journey of trying to figure out the best way to remove that broken bolt. Most mechanics cringe at the thought of removing broken exhaust manifold bolts as sometimes the engine will need to be removed in order to […]

You"ve probably heard a horror story about someone losing an engine when the oil pump fell off into the pan because of a broken bolt. Now you can put your mind at ease by using an ARP premium grade oil pump bolt kit. Manufactured from chrome moly steel with a black oxide finish, these bolts are nominally rated at 170,000 psi tensile strength to provide you with plenty of clamping force. The bolts feature a traditional hex style head, and washers are included. These inexpensive fasteners could literally save your engine.

Pump failure can arise from some of the least conspicuous or expensive parts of the pump—the bolts and washers. Pump manufacturers should consider the following six factors when choosing bolts for their rotating assemblies.

Different process applications require specific bolt design, class, strength and Maintenance of Certification. Process application conditions such as speed, pressure, pump driver and other factors have an impact on

Rotary lobe pumps often operate at slower speeds, under 1,000 rpm, and higher viscosities compared to other pump designs such as centrifugal pumps, which are regularly run at 3,600 rpm. They often pump water and other low viscosity materials.

A bolt design that works perfectly in one pump style may fail quickly in another. Pump speed has a drastic impact on the rotodynamic bolt in pumps. The pump’s typical speed range needs to be considered when designing the bolt assembly. Low speeds will place a large amount of torque on the rotating assembly, while high speeds can lead to greater vibration on the bolt. Excess vibration can lead to the failure of pump components, including the rotodynamic bolts. Vibration causes a radial force on the bolt head and shank that can cause the bolt to loosen and shear the bolt (see Image 1).

The process pressure range will also cause failure in a similar manner to speed. The value of the discharge and suction pressure on the pump can cause vibration and destroy a bolt if it is not taken into account. Suction lift can be harsh on rotodynamic bolts. In an application where a pump is expected to regularly see more than 20 feet of suction lift on water, the bolting assembly should be examined. Applications with high vacuum cause excess vibration and radial forces and have been shown to break bolts that succeed in other applications.

The pump driver can also cause this same damaging vibration. The pump driver that can produce the largest amount of vibration is the diesel engine. The diesel engine drive imparts a greater vibration on the pump skid and therefore on the pump and bolt compared to an electric or hydraulic motor drive. A diesel engine drive combined with high suction lift and speed can be akin to a perfect storm for vibration and bolt failure.

The chemicals and temperatures that a bolt encounters are some of the primary considerations when designing the bolting assembly. Many rotodynamic bolts are subject to the process fluid of their respective pump and are subject to the same or more degradation than rotors, housings and other pumping components.

These bolts can be the hottest element in the pump and must often be constructed of a more resilient material than other pump components. These bolts are subject to tensional, axial and radial force in a relatively small cross-sectional area compared to other pump elements. All other decisions in bolt design must follow the material selection since every material has different guidelines.

Different bolt designs and geometries have varying strengths and weaknesses and should be chosen depending on the pump application and service duty. Socket head cap screws (SHCS) are a common bolt design in pumping applications. The SHCS bolt heads have a smaller diameter compared to similar hex head bolts and are therefore easily incorporated into almost any assembly. Due to their widespread use, these bolts are available in a near infinite number of combinations of diameters, lengths, grades and materials.

Hex head screws and bolts, particularly flanged hex head screws (FHHS), can also be used with success in many pumping applications. The FHHS bolts incorporate a flange that distributes the pressure where the bolt head sits. The flange design can prevent loosening in cases where a SHCS bolt will back out of the threaded hole and potentially shear. FHHS bolts are less common and may not be suitable for a pumping application depending on the chemicals and other elements present in the process fluid. The service duty of the pump must also be kept in mind. Reversibility and run time per day impact the bolts needed for rotodynamic pumps.

The application of a threadlocking product to the rotodynamic bolt threads is often necessary to prevent the bolt from rotating separately from the pump components. The manufacturer’s instructions must be followed to fully generate the strength of the threadlocker. These products may require up to 72 hours to cure depending on substrate material, bond gap, temperature or other factors. Threadlockers have been shown to be one of the most effective methods of reducing overall bolt loosening or relaxation.

Most people assume that a stronger—therefore higher grade—rotodynamic bolt will always last longer. But the bolt’s lifespan is not only affected by the strength. Bolt manufacturers rated different grades of bolts at different torques for a reason. The bolt recommended torque value helps the bolt to hold tension properly and ensures that the bolt does not loosen after installation.

The necessary torque value can be very high for some higher grade bolts. The limitations of the pump must be considered when choosing these bolts. It is not beneficial to the overall pump assembly to use a bolt requiring a recommended bolt torque that could cause too much force on the pump assembly. The use of torque that is too high can cause failure in another part of the pump and torque that is too low may not properly maintain the pump assembly.

A circulating water pump is a key equipment of cooling systems in nuclear power plants. Several anchor bolts were broken at the inlet rings of the same type of pumps. The bolts were turned by a special material for seawater corrosion protection. There were obvious turning tool marks at the root of the thread, which was considered as the source of the crack. The fatigue crack extended to the depth of the bolt, causing obvious radiation stripes on the fracture surface, which was a typical fatigue fracture. Obvious overtightening characteristics were found at the head of the broken bolt. Fracture and energy spectrum analysis showed that the bolt was not corroded. The axial vibration of the pump was measured. The static tensile stress along the bolt axis caused by the preload, the axial tensile stress caused by the axial vibration, and the torsional stress were calculated, respectively. According to the fatigue strength theory, the composite safety factor of the bolt fatigue strength was 1.37 when overtightening at 1.2 times the design torque, which was less than the allowable safety factor of 1.5-1.8, so the bolt was not safe, which further verified the conclusion of fracture analysis. The reason for the low safety factor was caused by the overtightening force. The improvement method was to control the bolt preload or increasing the bolt diameter.

A cooling water pump is a very important equipment in nuclear power plants. During overhaul, it was found that the fixing bolts of the embedded parts of four CR1QS1 pumps were broken. The pump is a single-stage, vertical, bottom-suction concrete volute centrifugal pump. The pumps were fixed on the concrete embedded parts with 8 hexagon socket bolts through the mouth ring, as shown in Figure 1. The purpose of the protective cap is to protect the bolt from erosion. The working medium of the pump is sea water.

The common failure modes of bolt fracture are fatigue fracture, stress corrosion cracking, and overload fracture. Due to the large stress concentration of a bolt thread, it is easy for a fatigue source to form at the root, and the possibility of fatigue fracture is high. The bolt fracture studied by González et al. occurred at the second turn of the screw thread, which was caused by hydrogen embrittlement [1]. The bolt studied by Shafiei and Kazempour-Liaisi had M23C6 carbide, which was the source of the fatigue crack. The crack propagates along the grain boundary, and finally, fatigue fracture occurs [2]. Li et al. found that surface decarburization of the bolts and stress concentration at the bolt thread neck decreased the fatigue strength [3]. Wu et al. studied the corrosion fracture mechanism of cable bolts [4]. The fracture had general fatigue fracture characteristics. There were corrosion fatigue crack sources and radial fatigue crack propagation traces. Hydrogen-assisted stress corrosion cracking was the main fracture mechanism of cable bolts failure. The fatigue crack source of the bolt-sphere joint was pitting caused by corrosion [5]. Wen et al. [6] studied the fracture of a 20MnTiB steel high-strength bolt. Microdefects were found near the bottom of the thread. Considerable stress and corrosion accelerated the crack propagation of the bolt. The working capacity of a rock bolt decreased by 25-50% when it worked under the condition of rock and groundwater corrosion [7].

It is generally believed that the fatigue strength of bolts is only related to the stress amplitude. The fatigue strength only studied the stress amplitude of bolt tensile stress [7–10]. For example, the bolt fatigue strength condition was that the allowable stress amplitude was equal to 90 MPa [8], and the fatigue curve studied was the curve [9]. However, in practice, many examples showed that the failure of bolts was related to the average stress (i.e., bolt preload) [11, 12]. The reason for a bolt fracture was that the safety factor is insufficient due to excessive preload [11, 12]. The safety factor of static strength is obtained by preloading, the safety factor of variable stress is obtained by strain, and the safety factor is modified by Goodman’s theory [13].

In this paper, the fracture analysis, mechanical property analysis, and energy spectrum analysis of the broken bolt are carried out. At the same time, the fatigue strength of the bolt is calculated, the failure causes are found out, and the improvement suggestions are put forward. Finally, the calculation method of the bolt fatigue strength is proposed.

The bolts in service are shown in Figure 2, in which Nos. 1 and 2 were the unbroken bolts, Nos. 3-6 were the head of the broken bolts, and Nos. 7-11 were the rest of the broken parts of the broken bolts. Compared with the spares, their surfaces were the same as the serviced bolts, indicating that there was no corrosion.

The fracture of No. 3 bolt in Figure 2 is representative. Take it as an example to illustrate the fracture form of bolts. Figures 3(a) and 3(b) are the overall morphology and local morphology of the No. 3 bolt, respectively. There are obvious radial lines on the edge of the thread teeth, which is the fracture source as the point indicated by the arrow. The fracture source extends to the core, and then the bolt breaks when the crackle reaches the middle. This is the instantaneous fracture zone region, where the section is rough and uneven. The instantaneous breaking zone occupies a relatively large area, indicating that there is a large residual pretightening force when the bolt is broken.

Figure 5(a) is the morphology of the inner hexagon of the head of No. 3 broken bolt. The top of the bolt head is damaged when the sample was taken on site, as shown by the arrow. But the inner hexagon area is damaged during tightening, as shown in the region. Figure 5(b) shows the morphology of the unbroken bolt head, with the inner hexagon of the screw head intact. The comparison shows that the broken bolts have overtightening behavior when they were installed.

Figure 7 shows the macroimages of four unbroken screws through dye penetrant inspection, and no cracks are found on the surface. The metallographic structures of the unbroken and broken bolts are, respectively, shown in Figures 8(a) and 8(b), which show an austenite + ferrite structure. This conforms to the characteristics of dual phase steel, without obvious abnormality.

The bolts were made of a special material for seawater corrosion protection. Due to the small quantity, they were manufactured by turning. The chemical composition meets the ASTM s32760 standard, see Table 1. Using the XHB-3000 Digital Brinell Hardness Tester, the average hardness of bolts is 230-240 HBW, equivalent to grade 8.8 (Chinese national standard GB3098.1), which also meets the requirements of ASTM s32760 of less than 310 HBW.

By the AG100KNG universal testing machine, the tensile properties of sample bolts were tested, as shown in Table 3. The results all meet the requirements of standard values, and the mechanical properties are normal. According to the empirical formula recommended in the mechanical design manual, the symmetrical cycle fatigue limit and torque yield limit are estimated as follows:

Table 4 shows the composition of the fracture surface after cleaning by Energy Disperse Spectroscopy (EDS). The result is the same as the previous conclusion in Section 2.1, that is, as can be seen in Figure 2, the broken bolts were as glossy as the spare parts, and there was obviously no corrosion.

The bolts should be tightened when they are installed; that is, they are subject to the preload (tension) and friction torque. When working, it may be subjected to the variable stress of axial tension. In this paper, the finite element method is used to calculate the tensile stress and torsional stress by ANSYS Workbench 15.0 software.

The pump and the foundation ring are connected by 8 bolts. The finite element model takes 1 bolt and one eighth of the foundation including the ring and concrete, as shown in Figure 9. According to the equipment maintenance manual, the installation torque of the bolt is 40.5 Nm, the torque coefficient is 0.258, and the calculated preload is 13081 N.

The axial tensile stress and torsional stress of the bolt are shown in Figures 10 and 11, respectively. The axial tensile stress is 434.05 MPa, and the torsional stress is 59.29 MPa at design torque. If the overtightening torque reaches 1.2 times the design value, the axial tensile stress is 520.86 MPa, and the torsional stress is 71.41 MPa. The inner hexagon of the broken bolt head has been seriously damaged, and the actual torque is far greater than 1.2 times the design value.

When the pump runs, the impeller will have a working load, acting on the bolt axis direction. The stress is a symmetrical cyclic strain produced by the axial vibration when the pump is running. The axial load was obtained by actual measurement. A speed sensor was installed at the bearing, and the excitation spectrum load was the relationship between the speed and the frequency spectrum, as shown in Figure 12.

(1)There are obvious crack sources at the root of the thread, and there is an obvious fatigue fracture zone and an instantaneous fracture zone at the cross section. The fatigue fracture zone is typically radial and has typical fatigue fracture characteristics(2)The bolt safety factor at 1.2 times the design torque is 1.37, which has been less than the allowable safety factor of 1.5-1.8. Therefore, the fatigue strength of bolts is insufficient, and a bolt fracture is due to fatigue failure when the bolt is overtightened(3)The failure of bolts is not caused by seawater corrosion. The surface of the broken bolt is bright, and there is no trace of corrosion(4)The key cause of a bolt fracture is too much preload. The measure to improve the safety factor is to control the bolt preload or increase the diameter of the bolt