CN111881530B - Vibration reduction optimization design method for aeroengine - Google Patents
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- 239000000446 fuel Substances 0.000 description 7
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Abstract
The invention discloses an aeroengine vibration reduction optimization design method, which comprises the following steps: determining a part needing vibration reduction in the receiver accessory, and establishing a vibration transmission model; selecting one of the components, and taking the parameters of the component as a target for developing an optimal design; transforming the vibration transmission model to a frequency domain, and designing optimal control parameters by utilizing a frequency domain analysis method according to expected performance indexes; and (5) confirming whether the optimal control parameters of the design meet the performance requirements through simulation, and if not, carrying out the optimal design again. The vibration isolator can attenuate the vibration of a plurality of parts of the casing accessory, can obtain the complete vibration isolation of specific parts, and has important practical significance.
Description
Technical Field
The invention belongs to the field of vibration reduction and noise reduction of aeroengines, and particularly relates to an aeroengine vibration reduction optimization design method.
Background
Aero-engines have stringent requirements on reliability and safety, and complete machine vibration is one of the important factors for reducing reliability and safety, so that national and international airworthiness provisions prescribe that aero-engine vibration be monitored and controlled. For example, the national aviation administration has defined vibrations exclusively, i.e. "the design and construction of each type of engine must be such that the engine works properly throughout its stated flight envelope and throughout the operating range of rotational speed and power or thrust, without causing any parts of the engine to be overstressed by the vibrations and without causing excessive vibratory forces to be imparted to the aircraft structure. However, due to the variety of sources that cause the vibration of the engine, it is difficult to control the engine in real time, and the requirement of seaworthiness can be met only by a good structural design. For example, assembly errors are reduced by dynamic balancing to radically reduce periodic vibrations to the engine due to high and low pressure shaft mass imbalance; or through the parameter optimization design of the damper, the vibration is damped along the transmission path.
In practical engineering, another vibration problem is often encountered, that is, vibration caused by pneumatic unbalanced force (pneumatic interference, rotor misalignment, bearing damage, etc.) is transmitted to the casing, so that vibration of a full-weight electronic controller, a fuel regulating mechanism, a fuel and oil pipeline, etc. attached to the casing is excessively large. Because pneumatic unbalance is unavoidable, it is currently common practice to employ vibration isolation and shock resistance technology, i.e., to add vibration isolators at the mounting locations of electronic controllers, combustion tuning mechanisms, etc., to isolate or attenuate vibration transmissions from the engine. The key point of the design method is to select proper vibration isolator parameters, for example Zhao Kui describes the design method of the vibration isolator parameters in the paper of the design method and test study of the vibration isolator for an aeroengine controller; chinese patent application CN103742591a discloses a design method of a rotor adaptive mass damper of a rotary machine.
However, the design method only considers the isolation of a single component, so that the parameter optimization design of the single vibration isolator is only involved; when multiple components need to be isolated, the current approach is complex due to the numerous parameters that need to be designed.
Disclosure of Invention
In order to solve the technical problems mentioned in the background art, the invention provides an aeroengine vibration reduction optimization design method.
In order to achieve the technical purpose, the technical scheme of the invention is as follows:
a vibration reduction optimization design method of an aeroengine comprises the following steps:
(1) Determining a part needing vibration reduction in the receiver accessory, and establishing a vibration transmission model;
(2) Selecting one of the components, and taking the parameters of the component as a target for developing an optimal design;
(3) Transforming the vibration transmission model established in the step (1) to a frequency domain, and designing optimal control parameters by utilizing a frequency domain analysis method according to expected performance indexes;
(4) And (3) confirming whether the optimal control parameters designed in the step (3) meet the performance requirements through simulation, and returning to the step (2) if the optimal control parameters do not meet the performance requirements, and carrying out the optimal design again.
Further, in step (1), a vibration transmission model is established by using a system identification method or a parameter matching method.
Further, for two vibration reduction parts, the established vibration transfer model is as follows:
In the above formula, m u、ku and c u are the mass, stiffness coefficient and damping coefficient of the component 1, respectively; m s、ks and c s are the mass, stiffness coefficient and damping coefficient of component 2; z u (t) is the displacement of part 1, z (t) is the relative displacement of part 1 and part 2, For corresponding first order derivative,/>Deriving for the corresponding second order; /(I)Representing a pneumatic imbalance force; u (t) is the parameter to be optimized.
Further, in step (2), the parameter u (t) to be optimized is selected in the form of:
①u(t)＝kzu(t)；
②u(t)＝kz(t)
where k is the optimal control parameter to be optimized.
Further, the specific process of the step (3) is as follows:
first, the vibration transfer model is transformed from the time domain to the frequency domain by fourier transform:
Wherein ω is frequency;
Z (jω), Z u (jω), U (jω) are frequency response functions of Z (t), Z u (t), U (t), respectively;
D (jω) ≡ω 2Zr (jω) is an unbalanced vibration force;
det[G(jω)]＝(ku-muω2+jcuω)(ks-msω2+jcsω)-(ks+jcsω)msω2;
Then, two circles are drawn on the complex plane, which are respectively called alpha-circle and beta-circle, wherein alpha-circle is a circle with (-1, 0) as a circle center and delta 1 is less than or equal to 1 as a radius, beta-circle is a circle with (-G (j omega)) as a circle center and delta 2 is less than or equal to 1 as a radius, and delta 1 and delta 2 are expected performance indexes;
If the alpha-circle and the beta-circle have intersection, selecting the optimal point of the intersection part as alpha (j omega), and designing an optimal control parameter k:
The beneficial effects brought by adopting the technical scheme are that:
The invention provides a design method for achieving global vibration reduction by local parameter design aiming at the vibration problem of an aircraft engine casing accessory, which can attenuate the vibration of a plurality of parts of the casing accessory, can obtain complete vibration isolation of specific parts and has important value for practical engineering.
Drawings
FIG. 1 is an overall process flow diagram of the present invention;
FIG. 2 is a schematic diagram of α -circle and β -circle in the present invention;
FIG. 3 is a schematic of α -circle and β -circle in an embodiment;
Fig. 4 is a schematic diagram of vibration in an embodiment, in which (a) is a schematic diagram of vibration Z (jω) of the position of the combustion mechanism and (b) is a schematic diagram of vibration Z u (jω) of the position of the electronic controller.
Detailed Description
The technical scheme of the present invention will be described in detail below with reference to the accompanying drawings.
The invention designs an aeroengine vibration reduction optimization design method, which comprises the following steps as shown in figure 1:
step 1: determining a part needing vibration reduction in the receiver accessory, and establishing a vibration transmission model;
step 2: selecting one of the components, and taking the parameters of the component as a target for developing an optimal design;
Step 3: transforming the vibration transmission model established in the step 1 to a frequency domain, and designing optimal control parameters by utilizing a frequency domain analysis method according to expected performance indexes;
Step 4: and (3) confirming whether the optimal control parameters designed in the step (3) meet the performance requirements through simulation, and if not, returning to the step (2) to carry out the optimal design again.
In this embodiment, the above step 1 may be implemented by the following preferred scheme:
The vibration transfer model is typically obtained by a system identification method or by a parameter matching method. Taking a parameter matching method as an example, for two vibration reduction components, a model thereof can be established as follows:
Wherein m u、ku、cu is a structural parameter (mass, stiffness coefficient and damping coefficient) of the component 1; m s、ks、cs is a component 2 structural parameter; z u (t) is the displacement of part 1, z (t) is the relative displacement of part 1 and part 2, For corresponding first order derivative,/>Deriving for the corresponding second order; /(I)Represents a pneumatic unbalanced force, namely a vibration source, and the pneumatic unbalanced force is required to be attenuated and even isolated through an optimal design. That is, it is necessary to make the vibrations at the component 1 and the component 2 damped at the same time by local parameter optimization.
At this point, a mathematical model of the optimization system is available:
Wherein u (t) is a parameter to be optimized.
In the following embodiment, the component 1 is an electronic controller, the component 2 is a combustion mechanism, and the variables in the above model are as follows:
| ms(kg) | mu(kg) | ks(N/m) | ku(N/m) | cs(Ns/m) | cu(Ns/m) |
| 973 | 114 | 42720 | 101115 | 1095 | 14.6 |
in this embodiment, the above step 2 is implemented by adopting the following preferred scheme:
for the above-mentioned vibration transfer mathematical model, the parameter u (t) to be optimized is usually expressed as a function of structural parameters, and the following forms can be selected:
①u(t)＝kzu(t)；
②u(t)＝kz(t)。
Which form is specifically selected is determined by the actual situation (such as space, weight, etc.); but for both, k is the optimal control parameter that needs to be optimized. That is, it is necessary to design the appropriate unknown parameters k such that the vibrational responses of z u (t) and z (t) to z r (t) must be reduced simultaneously. In the following embodiment ② u (t) =kz (t) is selected as the parameter to be optimized, i.e. the vibration amounts of the two parts of the combustion adjusting mechanism and the electronic controller are attenuated by optimizing the parameters of the combustion adjusting mechanism.
In this embodiment, the above step 3 is implemented by adopting the following preferred scheme:
The method is characterized by comprising the following steps:
First, the vibration transfer model is transformed from the time domain to the frequency domain by fourier transform:
Wherein Z (j omega), Z u (j omega) and U (j omega) are frequency response functions of Z (t), Z u (t) and U (t) respectively; d (jω) ≡ω 2Zr (jω) is the unbalanced vibration force (Z r (jω) is the frequency response function of Z r (t), ω is frequency); the common denominator det (G) is defined as:
det(G)＝(ku-muω2+jcuω)(ks-msω2+jcsω)-(ks+jcsω)msω2.
for the above embodiment, the parameters listed in the table in step 1 are brought into the above formula, so as to obtain the required frequency domain model. The corresponding frequency is typically the natural frequency at the pitch mechanism
Next, the definition variable G is as follows:
The calculation in this embodiment can be:
G(jω)＝0.3844+0.1311j
Thirdly, drawing two circles on a complex plane, namely an alpha circle and a beta circle, wherein the alpha circle is a circle taking (-1, 0) as a circle center and delta 1 is less than or equal to 1 as a radius; and beta-circle is a circle with (-G (j omega)) as the center and delta 2 < 1 as the radius, as shown in figure 2.δ 1 and δ 2 are desired performance indexes, i.e. the attenuation amounts of Z (jω) and Z u (jω) are desired to be reduced by 3dB by the optimal design, and δ 1 =0.707 if Z (jω) is required to be reduced; requiring a 6dB decrease in Z u (jω), δ 2 =0.5.
In the present embodiment, δ 1 =0.65 and δ 2 =0.5 are specified as desired performance indexes, that is, it is desired to reduce the vibration Z (jω) at the fuel adjusting mechanism by 4dB (0.65) and attenuate the vibration Z u (jω) at the electronic controller by 6dB (0.5) by an optimal design. Thus, schematic diagrams of α -circle and β -circle in the examples are drawn, as shown in FIG. 3.
Again, for the present embodiment, in fig. 3, α -circle and β -circle have intersections (hatching), so that the vibration Z (jω) at the fuel adjustment mechanism and the vibration Z u (jω) at the electronic controller can be attenuated simultaneously; that is, the optimal design is the intersection part of two circles.
Finally, the optimal point of the intersection is chosen, denoted α (jω), and the optimal control parameters to be designed can be obtained by:
In this embodiment, since the center of β -circle is located in the intersection region of two circles, α (jω) = -G (jω), that is, α (jω) = -0.3844-0.1311j can be selected. At this time, k= -21360-7255.6j is calculated according to the above formula, and it is known that the design reduces the vibration Z (jω) at the fuel adjusting mechanism by 4dB, and the vibration Z u (jω) at the electronic controller is attenuated to zero, i.e., the electronic controller achieves complete vibration isolation.
According to the method provided by the invention, feasibility and performance limit analysis are carried out, and the performance index is confirmed to reach the expected value. For the above embodiment, if the index requires a 4dB reduction in the vibration Z (jω) at the fuel adjustment mechanism and a 20dB reduction in the vibration Z u (jω) at the electronic controller, it is apparent that the design can be confirmed; if the index requirement is 20dB less than the vibration Z (j omega) of the fuel oil regulating mechanism and the vibration Z u (j omega) of the electronic controller, the alpha-circle and the beta-circle are drawn again, and the fact that no intersection exists indicates that the index requirement is too high at the moment, and the performance index requirement must be reduced. The design is now confirmed with a reduction of 4dB in the vibration Z (jω) at the fuel adjustment mechanism and a reduction of at least 20dB in the vibration Z u (jω) at the electronic controller, with the real-time simulation results shown in fig. 4. It can be seen that by adopting the method provided by the invention, the vibration Z (j omega) at the fuel oil adjusting mechanism is truly reduced by 4dB, and meanwhile, the vibration Z u (j omega) at the electronic controller is completely attenuated, namely, the complete vibration isolation of the electronic controller is achieved. The design not only meets the requirements, but also has complete vibration isolation performance, and is the most expected design in practical engineering.
The embodiments are only for illustrating the technical idea of the present invention, and the protection scope of the present invention is not limited by the embodiments, and any modification made on the basis of the technical scheme according to the technical idea of the present invention falls within the protection scope of the present invention.
Claims (4)
1. The vibration reduction optimization design method for the aero-engine is characterized by comprising the following steps of:
(1) Determining a part needing vibration reduction in the receiver accessory, and establishing a vibration transmission model;
(2) Selecting one of the components, and taking the parameters of the component as a target for developing an optimal design;
(3) Transforming the vibration transmission model established in the step (1) to a frequency domain, and designing optimal control parameters by utilizing a frequency domain analysis method according to expected performance indexes;
(4) Confirming whether the optimal control parameters designed in the step (3) meet the performance requirements through simulation, and returning to the step (2) if the optimal control parameters do not meet the performance requirements, and carrying out the optimal design again;
In the step (1), a parameter matching method is adopted, and for two vibration reduction components, a vibration transmission model is established as follows:
In the above formula, m u、ku and c u are the mass, stiffness coefficient and damping coefficient of the component 1, respectively; m s、ks and c s are the mass, stiffness coefficient and damping coefficient of component 2; z u (t) is the displacement of part 1, z (t) is the relative displacement of part 1 and part 2, For corresponding first order derivative,/>Deriving for the corresponding second order; /(I)Representing a pneumatic imbalance force; u (t) is the parameter to be optimized.
2. The method of optimizing vibration damping design for an aircraft engine according to claim 1, wherein in step (1), a vibration transfer model is established by using a system identification method.
3. The method for optimizing vibration damping of an aircraft engine according to claim 1, wherein in step (2), the parameter u (t) to be optimized is selected in the form of:
①u(t)＝kzu(t)；
②u(t)＝kz(t)
where k is the optimal control parameter to be optimized.
4. The method for optimizing vibration damping design of an aeroengine according to claim 1, wherein the specific process of the step (3) is as follows:
first, the vibration transfer model is transformed from the time domain to the frequency domain by fourier transform:
Wherein ω is frequency;
Z (jω), Z u (jω), U (jω) are frequency response functions of Z (t), Z u (t), U (t), respectively;
D (jω) ≡ω 2Zr (jω) is an unbalanced vibration force;
det[G(jω)]＝(ku-muω2+jcuω)(ks-msω2+jcsω)-(ks+jcsω)msω2;
Then, two circles are drawn on the complex plane, which are respectively called alpha-circle and beta-circle, wherein alpha-circle is a circle with (-1, 0) as a circle center and delta 1 is less than or equal to 1 as a radius, beta-circle is a circle with (-G (j omega)) as a circle center and delta 2 is less than or equal to 1 as a radius, and delta 1 and delta 2 are expected performance indexes;
If the alpha-circle and the beta-circle have intersection, selecting the optimal point of the intersection part as alpha (j omega), and designing an optimal control parameter k:
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| CN112881025B (en) * | 2021-01-12 | 2022-06-10 | 南京航空航天大学 | Method for vibration control and energy collection of aircraft engine |
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| CN110362863A (en) * | 2019-06-11 | 2019-10-22 | 南昌大学 | A kind of hub motor wheel vibration insulating system parameter matching and optimization method |
| CN111123705A (en) * | 2019-12-18 | 2020-05-08 | 南京航空航天大学 | Design method for active vibration control of propeller and transmission shaft system |
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| CN110362863A (en) * | 2019-06-11 | 2019-10-22 | 南昌大学 | A kind of hub motor wheel vibration insulating system parameter matching and optimization method |
| CN111123705A (en) * | 2019-12-18 | 2020-05-08 | 南京航空航天大学 | Design method for active vibration control of propeller and transmission shaft system |
Non-Patent Citations (3)
| Title |
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| "Simultaneous vibration suppression and energy harvesting:Damping optimization for performance limit";Jiqiang Wang;《Mechanical Systems and Signal Processing》;第610-611、613-614、616页 * |
| "直升机主减机匣结构振动噪声分析与优化";张琳等;《航空动力学报》;第31卷(第2期);第323-329页 * |
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